proceeding imat2012 bintan

The 5th IMAT, November 12 – 13th

, 2012

i

THE 5th

IMAT

INTERNATIONAL MEETING ON ADVANCES IN THERMO-FLUIDS

NOVEMBER 12 – 13th

, 2012

BINTAN ISLAND - INDONESIA

DEPARTEMENT OF MECHANICAL ENGINEERING

UNIVERSITAS INDONESIA

VERITAS, PROBITAS, IUSTITIA


, 2012

ii

TABLE OF CONTENT

Welcome......................................................................................................................... iii

Organizing Commitee..................................................................................................... iv

Program Summary.......................................................................................................... vi

List of Paper.................................................................................................................... 1


, 2012

iii

WELCOME

It is our pleasure to welcome you, for the 5th

IMAT 2012, International Meeting on Advances

in Thermo-Fluids, to Bintan Island, Indonesia. The 5th

IMAT 2012 is organized by

Department of Mechanical Engineering, University of Indonesia. The history of this

conference has taken place in the following places:

1. 1st IMAT, 2008, UTM, Malaysia

2. 2nd IMAT, 2009, Bogor, Indonesia

3. 3rd IMAT, 2010, Singapore

4. 4th IMAT, 2011, Melaka, Malaysia

The 5th

IMAT will be coming at November 12-13th

, 2012 in beautiful and exotic island of

Bintan, Indonesia, around 45 minutes from Singapore or Malaysia by Ferry. The 5th

IMAT

aims to provide a technical forum that includes keynote lectures, information of researchs at

UI, NUS and UTM, and oral presentation sessions. In addition to the fundamentals of thermal

phenomena and traditional thermal applications, the 5th

IMAT is expected to address the

emerging domains of thermal transport in fishery, building nano-materials, bio-systems,

power generation, microsystems, and energy conversion devices.

All submitted papers will be peer reviewed, and the accepted paper will be published in the

conference proceeding. Selected papers will be offerred for publication in International

Journal.

We wish to provide the most pleasurable time in meeting all participant to talk and discuss

about new researches and applications related to thermal and fluids engineering, and give an

opportunity for the communication and cooperation between the researchers.

Bintan, November 12 – 13th

, 2012

Organizing Commitee


, 2012

iv

ORGANIZING COMMITTEE The 5

th IMAT 2012 is organized by Department of Mechanical Engineering,

Faculty of Engineering, University of Indonesia.

Kampus UI, Depok 16424, Indonesia

Phone : +62 21 727 0032

Fax : +62 21 727 0033

Email : [email protected]

Advisory Board

UNIVERISITAS INDONESIA (UI)

Prof. Dr. Ir. Bambang Sugiharto, M.Eng.

Prof. Dr. I. Made Kartika, Dipl.Ing.

Prof. Dr. Yanuar, M. Eng., M.Sc.

Prof. Dr. Ir. Budiarso, M.Sc.

Prof. Dr.-Ing. Nandy Setiadi Djaya Putra

Prof. Yulianto Sulistyo Nugroho, M.Sc., Ph.D.

Prof. Dr. Ir. Harinaldi, M.Eng.

Assoc. Prof. Dr. Budihardjo, Dipl.Ing.

Assoc. Prof. Dr. M. Idrus Alhamid

Assoc. Prof. Dr.-Ing Nasruddin, M.Eng.

UNIVERSITI TEKNOLOGI MALAYSIA (UTM)

Prof. Ir. Dr. Azhar Abdul Aziz

Assoc. Prof. Dr. Mazlan Abdul Wahid

Prof. Amer Nordin Darus

Prof. Dr. Farid Nasir Hj. Ani

Prof. Dr. Md. Nor Musa

Dr. Jamaluddin Md. Sheriff

NATIONAL UNIVERSITY OF SINGAPORE (NUS)

Prof. Dr. Kim Choon Ng

Prof. Dr. Christopher Yap

Dr. Kandadai Srinivasan

mailto:[email protected]


, 2012

v

INTERNATIONAL ISLAMIC UNIVERSITY OF MALAYSIA (IIUM)

Prof. Dr. M.N.A. Hawlader

Organising Committee

Chairman:

Dr. Agus S. Pamitran, ST., M.Eng.

Secretary:

Muhamad Yulianto, ST., MT.

Member:

Ir. Senoadi, MT.

Ir. Ruli Nutranta, M.Eng.

Ir. Supryiadi, M.Sc.

Ir. Arief Surachman, MT.


, 2012

vi

PROGRAM SUMMARY

The 5th

International Meeting on Advances in Thermofluids

Universitas Indonesia at Nirwana Garden – Bintan Island Indonesia

12 – 13 November 2012

DAY 0 (11th

November 2012 – Sundayy)

18.30 – 21.00 Welcoming BBQ Dinner

DAY 1 (12th

November 2012 – Monday)

07.00 – 08.30 Registration

08.30 – 08.40 Opening Ceremony

08.40 – 09.00 Research Info from UI (Prof. Harinaldi)

09.00 – 09.20 Research Info from UTM (Prof. Azhar)

09.20 – 09.40 Research Info from NUS (Prof. Ng KC)

09.40 – 10.00 Tea Break

10.00 – 16.00 PARALEL SESSION DAY 1

ROOM A : HVAC &

Fluid Flow

ROOM B : Heat/Mass Transfer &

Combustion

10.00 – 10.15 IMAT-UI 002 IMAT-UI 003

10.15 – 10.30 IMAT-UI 004 IMAT-UI 016

10.30 – 10.45 IMAT-UI 006 IMAT-UI 018

10.45 – 11.00 IMAT-UI 007 IMAT-UI 020

11.00 – 11.15 IMAT-UI 011 IMAT-UI 024

11.15 – 11.30 IMAT-UI 013 IMAT-UI 027

11.30 – 11.45 IMAT-UI 019 IMAT-UI 038

11.45 – 12.00 IMAT-UI 022 IMAT-UI 039

12.00 – 13.00 LUNCH

13.00 – 13.15 IMAT-UI 026 IMAT-UI 041

13.15 – 13.30 IMAT-UI 028 IMAT-UI 001

13.30 – 13.45 IMAT-UI 031 IMAT-UI 005

13.45 – 14.00 IMAT-UI 033 IMAT-UI 008

14.00 – 14.15 IMAT-UI 034 IMAT-UI 009

14.15 – 14.30 IMAT-UI 035 IMAT-UI 010

14.30 – 14.45 IMAT-UI 036 IMAT-UI 014

14.45 – 15.00 IMAT-UI 040 IMAT-UI 015

Serving Tea Break

15.00 – 15.15 IMAT-UI 042 IMAT-UI 021

15.15 – 15.30 IMAT-UI 012 IMAT-UI 023

15.30 – 15.45 IMAT-UI 017 IMAT-UI 029

15.45 – 16.00 IMAT-UI 025 IMAT-UI 030

16.00 – 16.15 IMAT-UI 032

16.15 – 16.30 IMAT-UI 037

19.00 – 20.30 CLOSING CEREMONY and Sea Food Dinner

at Kellong Restaurant

DAY 2 (13th

November 2012 – Tuesday)

09.00 – 11.00 TOUR


, 2012

vii

List Papers of IMAT 2012

Code Topic Name

IMAT-UI 001

Controlled Auto-Ignition Combustion

In A Two-Stroke Cycle Engine Using

Hot Burned Gases

Amin Mahmoudzadeh Andwari, Azhar

Abdul Aziz, M. F. Muhamad Said, Z. Abdul

Latif

IMAT-UI 002

Determination of Motive Nozzle and

Constant-Area Diameters: Numerical

Study of Ejector as an Expansion

Device in Split-type Air Conditioner

Kasni Sumeru, Henry Nasution, Farid Nasir

Ani

IMAT-UI 003 Effect Of External Particles On Brake

Noise Of Disc Braking System

M.A. Nasaruddin, M.K Abdul Hamid, A.R.

Mat Lazim, A.R. Abu Bakar

IMAT-UI 004

Experimental Study On Replacement Of

HFC-134a By Hydrocarbons In

Automotive Air Conditioner

Mohd Rozi Mohd Perang, Henry Nasution,

Zulkarnain Abdul Latif, Azhar Abdul Aziz,

Afiq Aiman Dahlan.

IMAT-UI 005 Effect Of Fuel Droplets During Early

Stage Of Spherical Flame Propagation

Aminuddin Saat, Mazlan Abdul Wahid,

Malcolm Lawes

IMAT-UI 006 The Use Of Mechanical Ventilation

System In An Electric Car

Intan Sabariah Sabri, Haslinda Mohamed

Kamar, Nazri Kamsah, Md Noor Musa

IMAT-UI 007

Retrofitting R-22 Split Type Air

Conditioning With Hydrocarbon (Hcr-

22a) Refrigerant

Henry Nasution, Zulkarnain Abdul Latif,

Azhar Abdul Aziz, Mohd Rozi Mohd Perang

IMAT-UI 008

Characterization Of Generator With

Palm Oil Biodiesel At Different

Compression Ratio

Belyamin, Alias Bin Mohd. Noor, Mohanad

Hamzah Hussein, Mazlan Bin Said, Mohd

Hafidzi

IMAT-UI 009

The Effect Of Fuel Additives On

Gasoline Heating Value And Spark

Ignition Engine Performance

Zulkarnain Abdul Latif, Azhar Abdul Aziz,

Mohd Rozi Mohd Perang, Normaliza

Abdullah

IMAT-UI 010

Design a Four-Stroke Homogeneous

Charge Compression Ignition (HCCI)

Engine

Mohd Rozi Mohd Perang, Zulkarnain Abdul

Latiff, Azhar Abdul Aziz, Mohamad Azzad

Mokhri

IMAT-UI 011

R22 and Various Mixtures of

R290/R600a as its Alternative in

Adiabatic capillary tube Used in split-

type Air-conditioning System

Shodiya Sulaimon, Azhar Abdul Aziz,

Henry Nasution, Amer Nordin Darus

IMAT-UI 012

Flow Pattern at Pipe Bends on

Corrosion Behaviour of Low Carbon

Steel

Muhammadu Masin MUHAMMADU

IMAT-UI 013 Green Refrigerant for Multi-Circuit Air-

Conditioning System Hayati Abdullah and Alif Jalaludin

IMAT-UI 014

Friction Characteristic of Palm Olein at

Different Operating Temperature using

Four-ball Tribometer

S. Syahrullail, C.I Tiong


, 2012

viii

IMAT-UI 015

Pulse Detonation Engine Research

Development at High Speed Reacting

Flow Laboratory, Universiti Teknologi

Malaysia

Mazlan A. Wahid, A. Dairobi G.,

Aminuddin Saat, Mohsin M. Sies, H.A.

Mohammed, A. N. Darus, Mohd Faizal H.,

M. Ibthisham A., Fairus M. Y. and Z. Lazim

IMAT-UI 016 Pool Boiling Of Nanofluids In Vertical

Porous media Nandy Putra, Ridho Irwansyah

IMAT-UI 017 Analysis of Small Bubble

Characteristics in Alum Solution Warjito, Nurrohman

IMAT-UI 018 Effect Of Hot Air Reservoir in The

Development of Vacuum Freeze Drying

M. Idrus Alhamid, Nasruddin, Engkos A.

Kosasih, Muhamad Yulianto

IMAT-UI 019

Experimental of Cascade Refrigeration

System Using Natural Refrigerant

Mixture Ethane and Carbon Dioxide at

Low Temperature Circuit and

Refrigerant Natural Propane at high

temperature circuit

Nasruddin, M. Idrus Alhamid and Arnas

IMAT-UI 020

Performance Analysis of

Thermoacoustic-standing Wave As a

Power Generation Adi Suryo, Sentosa I

IMAT-UI 021 Factors Affecting Performance Of Dual

Fuel Compression Ignition Engines

Mohamed Mustafa Ali, Sabir Mohamed

Salih

IMAT-UI 022

Solar Air-Conditioning System Using

Single-Double Effect Combined

Absorption Ahiller

Hajime Yabase

IMAT-UI 023

Environmental Protection and Fuel

Consumption Reduction Flameless

Combustion Technology : A Review

Seyed Ehsan Hosseini, Saber Salehirad,

Mohsin Mohd Sies, Mazlan Abdul Wahid

IMAT-UI 024

The Effect Of Geometrical Parameters

On Heat Transfer Of Micro-Channels

Heat Sink

Law Wen Zhe, Amer Nordin Darus.

IMAT-UI 025

Investigation of the Velocity Profiles in

a Ninety-Degree Curved Standing Wave

Resonator with PIV

Normah M.G, Irfan Abd. R, Quenet T, Zaki

Ab.M

IMAT-UI 026 MED+AD Desalination Cycle Muhammad Wakil Shahzad, Kim Choon Ng,

Won Gee Chum

IMAT-UI 027 Kinetics Of Propane Adsorption On

Maxsorb III Activated Carbon

Azhar bin Ismail, Loh Wai Soong, Ng Kim

Choon

IMAT-UI 028

Effects of Natural Ventilations on

Indoor Air

of a Double-Storey Residential House

in Malaysia

Haslinda Mohamed Kamara, Nazri Kamsah

& Kam Jia Liq

IMAT-UI 029

Deposit Forming Tendency of Biodiesel

and Diesel Fuel due to High Pressure

Exposure

Muhamad Adlan Abdullah, Arshad Salema

and Farid Nasir Ani


, 2012

ix

IMAT-UI 030

Numerical Analysis of

Elastohydrodynamic Lubrication with

Non-Newtonian Lubricant

Dedi Rosa Putra Cupu, Adli Bahari, Kahar

Osman, Jamaluddin Md Sheriff

IMAT-UI 031

Latest System Simulation Models in

Heating, Refrigerating and Air

Conditioning Field, and Development

of System Simulator

Kiyoshi Saito and Jongsoo Jeong

IMAT-UI 032 Drag Reduction of Bamboo and Abacca

Fiber Suspensions in Circular Pipe

Gunawan, M. Baqi, Yanuar and Sanlaruska

Faternas

IMAT-UI 033

A Study on The Effect of Exhaust

Gases on The Indoor Air Quality

Onboard Naval Ships

Arman Ariffin and Hayati Abdullah

IMAT-UI 034 Application Of Thermal Energy Storage

For A Lowland Farming House

Haslinda Mohamed Kamar, Nazri Kamsah,

Norull Ahmad Azman

IMAT-UI 035

Transient Model And Entropy Analysis

Of A Lithium Bromide – Water

Absorption Chiller

Ang Li and Prof. Kim Choon Ng

IMAT-UI 036 Characteristics of Sea-water Ice Slurry

for Cooling of Fish

A.S. Pamitran, M. Novviali, H.D.

Ardiansyah

IMAT-UI 037

Fluid Flow Characteristic Of Rounded-

Shape FPSO and LNG Carrier During

Off Loading

Mufti Fathonah Muvariz, Jaswar Koto,

Agoes Priyanto

IMAT-UI 038

Comparison Of Simulation Organic

Renkine Cycle (ORC) System Using

Turbocharger and Cycl Tempo V.5

With Environment Friendly Fluid

Ruly Rutranta, M. Idrus Alhamid and

Harinaldi

IMAT-UI 039 Thermophysical Properties of Novel

Zeolite Materials for Sorption Cycles

Kyaw Thu, Young-Deuk Kim, Baojuan Xi,

Azhar Bin Ismail, Kim Choon Ng

IMAT-UI 040

Review Paper: Sea-water Ice Slurry

Generator and Its Application on

Indonesian Traditional Fishing

A.S. Pamitran, H.D. Ardiansyah, M.

Novviali

IMAT-UI 041

Improving Hydrogen Storage Capacity

on Metal Doping Carbon Nanotubes

Using Molecular Dynamics Simulation

Nasruddin, Engkos A. Kosasih, Supriyadi,

Abdul Jabar

IMAT-UI 042 Performance of Thermoelectric and

Heat Pipes Refrigerator Cooling System

Firman Ikhsan, Ali A. Sungkar, M. Afin

Faisol, M. Zilvan Bey, Nandy Putra,

Saripudin

IMAT-UI 043

Preliminary Study on Length of Candle

Filter Surface on the Flow Pattern in

Freeboard of Fluidised Bed Gasifier

A. Farhan Faudzi, Kahar Osman, Nor

Fadzilah Othman, Mohd Hariffin Bosrooh

IMAT-UI 044

Preliminary Study on the Effect of Type

Distributor Plate on Airflow Pattern in

Bubbling Fluidised Bed

Nofrizalidris Darlis, Kahar Osman, Ab

Malik A. Hamat, Nor Fadzilah Othman,

Mohd Hariffin Bosrooh

IMAT-UI 045 Natural Convection in A Differentially

Heated Cavity Using Splitting Method Ubaidullah S., Kahar Osman


, 2012

1

Controlled Auto-Ignition (CAI) Combustion In A Two-Stroke

Cycle Engine Using Hot Burned Gases

Amin Mahmoudzadeh Andwaria, Azhar Abdul Aziz

b, M.F. Muhamad Said

c, Z. Abdul Latiff

d

Automotive Development Center (ADC), Faculty of Mechanical Engineering

Universiti Teknologi Malaysia (UTM), 81310, Johor Bahru, Malaysia

[email protected]

[email protected]

[email protected]

[email protected]

ABSTRACT

A new combustion concept, which is viewed

increasingly as a probable solution to these issues

is Controlled Auto-Ignition (CAI) Combustion. In

such an engine, a homogeneous mixture of air, fuel

and residual gases is compressed until auto-ignition

occurs. Due to its significantly low temperature

combustion, NOx will be dramatically reduced

while the mixture will be under ultra-lean fuel-air

condition, thus able to achieve high efficiency and

low emission. In the case of two-stroke engine,

problem of poor combustion efficiency and

excessive white smoke emission can be addressed

by the incorporation some features that will

ultimately convert a typical two-stroke engine into

an efficient CAI engine demonstrating the best of

both features. Due to its inherent high internal

residual gas rate in partial load operation, the two-

stroke engine has been the first application to

benefit from the unconventional CAI combustion

process. This paper will concisely discuss the

utilization of hot burned gas for induction thus

imposing a CAI combustion feature onto a

reference two-stroke cycle engine. Among the

features incorporated are the increasing in the level

of Internal or External Exhaust gas Recirculation

(In/Ex-EGR) and cycle-by-cycle uniformity of the

air-fuel ratio (AFR) supplied to cylinder which will

be crucial in creating a suitable temperature within

the engine‘s combustion chamber.

Keywords : Two-Stroke cycle engine, Controlled

Auto-Ignition (CAI), hot burned gas,

auto-ignition temperature, ATAC

(Active Thermo-Atmospheric

Combustion), Homogeneous Charge

Compression Ignition (HCCI),

Exhaust gas Recirculation (EGR)

1. INTRODUCTION

Energy conservation and environmental protection

are exerting rigorous demand on internal

combustion engine developers to further improve

fuel economy and emission reduction. In addition,

concerns about the world‘s finite oil reserves and,

more recently, by CO2 emissions brought about

climate change has led to heavy taxation of road

transport, mainly via on duty on fuel. Over the last

30 years, levels of NOx, CO and HC emissions

from vehicles have been dramatically increased.

This has been motivated by a continually tightening

band of legislation related to emission of these

pollutants.

Two-stroke cycle engines are well known owing to

their light weight, simple construction, less

components, cheap to manufacturing and the

potential to pack almost twice the power-density

than that of a four-stroke engine having similar

capacity [1]. For a longtime, the objective of the

different research works on two-stroke engines

optimization was to eliminate its two main

drawbacks leading to high emissions of unburned

hydrocarbons (uHC) and poor fuel efficiency. The

first one is the unstable running operation

combined with incomplete combustion, especially

at light load. The second one is fuel short circuit at

medium and full load. However due to the short-

circuiting of the fuel before combustion, this has

resulted in deterioration in overall performances

especially poor combustion efficiency and high

white smoke emission problem [2].

For that reason, many researchers have begun to

research the new kind of alternative combustion

intensively. One example of these attempts is the

ATAC (Active Thermo-Atmospheric Combustion)

[3], TS Combustion (Toyota-Soken) [4], ARC

(Activated Radical Combustion) [5], Homogeneous

Charge Compression Ignition (HCCI) [6] and CAI

(Controlled Auto-Ignition) combustion concept,

bulk combustion or low temperature combustion, a

combustion process that has conventionally been

used in two-stroke engines. It has been found that

depending on the engine speed, load ratio and level

of Exhaust Gas Recirculation (EGR) applied, it is

possible to induce Auto-Ignited (AI) combustion in

a two-stroke engine as a result of the mixing of

unburned mixtur

IMAT-UI 001


, 2012

2

e gas introduced into the cylinder and hot residual

(burned) gas [5]. These combustion processes can

reduce emissions of unburned HCs and allow stable

engine operation by lower cyclic variation. Owing

to its inherent high internal residual gas rate in

partial load operation, the two-stroke engine has

been the first application to take benefit of the

unconventional CAI combustion process.

Several issues that must be addressed in order to

implement this combustion process in production

engines include control of the ignition timing and

burning rate, which are determined by the

chemical reactions of the unburned mixture, and

expansion of the stable operating region [7]. In

fact, the most recognized original work on CAI

combustion was motivated by some researchers

desire to control the irregular combustion caused

by the auto-ignition of cylinder charge to obtain

stable lean-burn combustion in the conventional

ported two-stroke gasoline engine [8], [3], [4].

It is known that these drawbacks result from the

new mixture shortcut and irregular combustion in

the part load operation. In the light load operating

range, the residual gas in combustion chamber

increases due to the poor scavenging. Normal

flame propagation is disturbed by the large amount

of residual gas, which generates irregular

combustion [5]. The research works during this

period to study the part load lean two-stroke

combustion have led to discover that the

irregularities of the combustion and the auto-

ignition, which are considered as the weak points

of the two-stroke engine, can be effectively

controlled and managed to get a part load stable

two-stroke combustion process for lean mixtures in

which ignition occurs without spark assistance.

Suitably, remarkable improvements in stability,

fuel efficiency, exhaust emissions, noise and

vibration will be achieved [3], [8].

2. ORIGINATION OF CAI

COMBUSTION AND ITS

FUNDAMENTALS

Although it is generally accepted that the first

systematic investigation on the new combustion

process was carried out by Onishi and Noguchi in

1979, the theoretical and practical roots of the CAI

combustion concepts are attributed to the

pioneering work carried out by the Russian

scientist Nikolai Semenov and his colleagues in the

field of ignition in the 1930s. Having established

his chemical or chain theory of ignition, Semenov

sought to exploit a chemical-kinetics controlled

combustion process for IC engines, in order to

overcome the limitations imposed by the physical

dominating processes of SI and CI engines [9].

2.1 Two-Stroke CAI Combustion Engine

The CAI combustion story has been started with

two-stroke engines. Substantial research work was

performed from the end of the 1960s to the end of

the 1970s in order to solve one of the main

problems of the two-stroke engine that was the

unstable, irregular and incomplete part load

combustion responsible for excessive emissions of

unburned hydrocarbons (uHC). Researchers

performed a lot of investigation work during this

period to study the part load lean two-stroke

combustion. They discovered that the irregularities

of the combustion and the auto-ignition that were

considered as the weak points of the two-stroke

engine could be effectively controlled. The

research‘s objective was managed to get a part load

stable two-stroke combustion process for lean

mixtures in which ignition occurs without spark

assistance. The new combustion process occurring

without flame front was called ‗ATAC‘ for (Active

Thermo-Atmosphere Combustion) [3]. During the

same year, another researchers‘ paper concerning

two-stroke auto-ignition was published. They

named this auto-ignition the TS (Toyota-Soken)

combustion process. They also found that such

combustion occurred similarly without flame front

while showing excellent efficiency and emissions

figures. They were the first to suggest that active

radicals in residual gases could play an important

role in the auto-ignition process [4].

2.2 Basic Principle of CAI Combustion

Similar to a conventional SI engine, in a CAI

engine the fuel and air are mixed together either in

the intake system or in the cylinder with direct

injection. The premixed fuel and air mixture is then

compressed. Towards the end of the compression

stroke, combustion is initiated by auto-ignition in a

similar way to the conventional CI engine. The

figure 1 shows ideal representations of both SI and

CAI combustion processes. In the case of SI

combustion, it is the flame front that separates the

burned gases from the fresh unburned gases and its

velocity controls the combustion heat release. In

the case of CAI, the combustion reactions take

place with multiple auto-ignition sites. Even if the

combustion locally can progress slowly, since it

occurs spontaneously and simultaneously at several

locations within the combustion chamber, the

overall heat release can be as fast or even faster

than with the flame front controlled SI without

generating the typically high combustion

temperatures of the flame front. This could

contribute to explain the CAI low NOx emissions

advantage.


, 2012

3

Figure 1: Spark Ignition (SI) combustion (left)

and CAI combustion (right)

The heat release characteristics of the CAI

combustion can be seen with using figure 1. In the

case of SI combustion, a thin reaction zone or

flame front separates the cylinder charge into

burned and unburned regions and the heat release is

confined to the reaction zone. Thus, flame front

velocity controls the combustion heat release. As it

can be seen in equation 1, the cumulative heat

released in a SI engine is therefore the sum of the

heat released by a certain mass, dmi, in the reaction

zone and it can be expressed as

(1)

where q is the heating value per unit mass of fuel

and air mixture, N is the number of reaction zones.

Figure 2: Heat release characteristics of SI (a)

and CAI combustion (b) [3]

In an idealized CAI combustion process,

combustion reactions take place simultaneously in

the cylinder and all the mixture participates in the

heat release process at any instant of the

combustion process. In other word, the combustion

reactions take place with multiple auto-ignition

sites. Even if the combustion locally can progress

slowly, since it occurs spontaneously and

simultaneously at several locations within the

combustion chamber, the overall heat release can

be as fast or even faster than with the flame front

controlled SI without generating the typically high

combustion temperatures of the flame front. This

could contribute to explain the CAI low NOx

emissions advantage that will be described in the

following section. Regarding equation 2, the

cumulative heat release in such an engine is

therefore the sum of the heat released from each

combustion reaction, dqi, of the complete mixture

in the cylinder, m, i.e.

(2)

where K is the total number of heat release

reactions, and qi is the heat released from the ith

heat release reaction involving per unit mass of fuel

and air mixture. Whereas the entire heating value

of each minute parcel of mixture must be released

during the finite duration spend in the reaction zone

in a SI engine, heat release takes place uniformly

across the entire charge in an idealized CAI

combustion. However, in practice, due to in-

homogeneities in the mixture composition and

temperature distributions in a real engine, the heat

release process will not be uniform throughout the

mixture. Faster heat release can take place in the

less diluted mixture and/or high temperature

region, resulting in a non-uniform heat release

pattern as indicated by the dashed lines [3].

3. TWO-STROKE CAI COMBUSTION

CONTROLLING

In spark ignition mode the combustion can be

rather easily directly controlled by the spark

advance. In the case of CAI combustion, there are a

lot of relevant control parameters with, in addition,

complex interactions between some parameters.

Prior to examining in more detail the main relevant

two-stroke CAI control parameters, it is important

to define what has to be controlled: The

Combustion Timing and The Combustion Heat

Release Rate [10]. A correctly controlled CAI

combustion should have the best combustion

timing for the highest combustion efficiency.

3.1 Mixing Between Fresh Charge and Burned

Gases

Inherently in a two-stroke engine, there is a high

amount of internal EGR at part load. But if this

EGR is well mixed with the fresh charge, as is the

case in a conventional two-stroke engine,

especially at low engine speed, it has almost no

effect on the combustion. What is efficient for

getting CAI is to limit as much as possible the

mixing/Stratification of this internal EGR with the

fresh mixture. In such cases, it is possible to

achieve a temperature gradient within the charge

for the same overall amount of in-cylinder EGR,

which means, the same in-cylinder heat

content.[11], [12], [13].

3.2 The Engine Speed

The engine speed is an indirect CAI control

parameter. It has an indirect effect on the

mixing/stratification between the fresh charge and

the EGR. When the engine speed increases the time

for mixing between the internal EGR and the fresh

charge will be shorter, therefore the internal

stratification and the temperature gradient inside

the trapped charge. This has the final consequence


, 2012

4

of advancing and accelerating the CAI combustion

[5].

3.3 The In-cylinder Flow Velocities

To introduce the same amount of fresh mixture in a

much longer time means that the charge is

introduced more smoothly at significantly lower

velocities. This helps to prevent the dilution of the

fresh inlet mixture in the residual gases which is

favorable for getting CAI [14].

3.4 The In-Cylinder Pressure

The easiest solution for controlling the in-cylinder

pressure is to do it through the compression ratio

(CR). A high compression ratio will then be

favorable to extend the CAI range to the low speed

low loads. Nevertheless, the choice of the highest

compression ratio favorable for CAI is always

limited by the fact that the same engine also has to

be able to run in spark ignition at full load without

knock [15].

3.5 The Overall Temperature

Heating the intake charge increases the overall gas

temperature and has the effect of advancing the

CAI combustion timing and therefore of extending

the CAI combustion range in the low load low

speed region. Similarly, the CAI combustion is also

sensitive to the engine liquid cooling temperature,

which indirectly affects the overall gas temperature

[16].

3.6 The Fuel Formulation

Several researchers have studied the effect of the

fuel formulation on two-stroke CAI combustion.

Results show that running an ATAC-CAI two-

stroke engine with methanol allows significant

widening of the auto-ignition range [17]. More

recently, some researchers tried to find some

correlations between the octane number of several

fuels (research octane number RON and motor

octane number MON) and their effect on the auto-

ignition range.

3.7 Changing of Two-Stroke Engine Design

Among the possible control parameters, the in-

cylinder gas temperature effect obtained by

stratifying hot internal EGR (the EGR and fresh

charge mixture/stratification being controlled by

the in-cylinder flow velocities) and the in-cylinder

pressure are the most relevant for practical

application mainly because of their immediate

response time in transient operation (which, for

example, is not the case of intake air heating). Most

of the technologies that have been developed and

applied to obtain CAI combustion on two-stroke

engines were based on the control of these two

internal temperature and pressure effects. For this

purpose, three main technologies and associated

control devices have been developed which

include: Elongated Transfer Duct, Transfer Duct

Throttling and Exhaust Port Throttling [3, 7, 12].

4. EFFECT OF EXHAUST GAS AS

DILUENT

In order to achieve CAI/HCCI combustion, the

temperature of the charge at the beginning of the

compression stroke has to be increased to reach

auto-ignition conditions at the end of the

compression stroke. This can be done by heating

the intake air or by keeping part of the hot

combustion products (charge dilution) in the

cylinder. Both strategies result in a higher gas

temperature throughout the compression process,

which in turn speeds up the chemical reactions that

lead to the start of combustion of homogeneously

mixed fuel and air mixtures. In-cylinder gas

temperature must be sufficiently high to initiate and

sustain the chemical reactions leading to auto-

ignition processes. Substantial charge dilution is

necessary to control runaway rates of the heat

releasing reactions. Both of these requirements can

be realized by recycling and/or trapping the burned

gases within the cylinder, which the former is

represented as External-EGR and the latter is

known as Internal-EGR, respectively.

The presence of the recycled or trapped burned

gases has a number of effects on the CAI

combustion and emission processes within the

cylinder.

4.1 The Charge Heating Effect

If hot burned gases are mixed with cooler inlet

mixture of fuel and air, the temperature of the

intake charge increases owing to the heating effect

of the hot burned gases. This is often the case for

CAI combustion with high-octane fuels, such as

gasoline and alcohols [18].

4.2 The Dilution Effect

The introduction or retention of burned gases in the

cylinder replaces some of the inlet air and hence

causes a substantial reduction in the oxygen

concentration. The reduction of air/oxygen due to

the presence of burned gases is called the dilution

effect [19].

4.3The Heat Capacity Effect

The total heat capacity of the in-cylinder charge

will be higher with burned gases, mainly owing to

the higher specific heat capacity values of carbon

dioxide (CO2) and water vapor (H2O). This rise in


, 2012

5

the heat capacity of the cylinder charge is

responsible for the heat capacity effect of the

burned gases [19].

4.4 The Chemical Effect

Combustion products present in the burned gases

can participate in the chemical reactions leading to

auto-ignition and subsequent combustion. This

potential effect is classified as the chemical effect.

It should be noted that the chemical effects are

influenced by active species or partially oxidized

hydrocarbons or activated radical [20].

It should be noted that the overall effect of hot

burned gases on the CAI combustion process is to

charge heating effect, to advance the Auto-Ignition

timing and to shorten the combustion. By hot

burned gas incorporation, the initial charge

temperature of the total in-cylinder charge will be

increased owing to the heating effect of hot burned

gases, and the relative air fuel ratio λ will be

reduced as burned gases would be replaced some of

the air [21].

The presence of hot burned gases initially causes

the CAI combustion process to accelerate. Both

experiments and analytical studies have shown that

the overall effect of hot burned gases is to advance

the start of CAI combustion due to their charge

heating effect. Ignition is dominated by the charge

heating effect but the combustion duration is

dominated by the dilution and heat capacity effect.

The maximum rate of heat release is equally

affected by the charge heating effect and by the

combined dilution and heat capacity effect. From a

macroscopic point of view of the heat balance, i.e.

the relationship between the calorific value

supplied in a cycle and the total heat capacity of the

in-cylinder gases; a larger heat capacity will take a

longer time to heat up and the maximum

combustion temperature will be lower. Thus,

combustion of a larger heat capacity generates a

slower heat release while that of a smaller heat

capacity permits a quicker heat release. For high-

octane fuels, like gasoline, alcohols, natural gas,

etc., it will be advantageous to retain burned gases

at as high temperature as possible to promote auto-

ignition of fuel/air mixture, particularly at low load

operations.

CONCLUSION

Two-stroke cycle engines can be more efficient and

clean by CAI combustion mode operation. Hot

burned gas utilization in order to induce this unique

combustion has been always interested as result of

some specific advantages, which are included: the

charge heating effect, the dilution effect, the heat

capacity effect and the chemical effect. In general,

hot burned gases that are used in two-stroke engine

either trapped or recycled, have considerable effect

upon combustion phenomenon and its

characteristics which will be led to induce and

control of the CAI combustion as follow:

Hot burned gases are preferred in most cases in

order to increase cylinder charge temperature

without external heating source

Overall effect of hot burned gases is to advance

the start of CAI combustion due to their charge

heating effect.

Start of ignition is dominated by the charge

heating effect

Combustion duration is dominated by the

dilution and heat capacity effect.

Maximum rate of heat release is equally

affected by the charge heating effect and by the

combined dilution and heat capacity effect.

The larger the heat capacity, the slower the heat

release in combustion.

It is most desired for gasoline, alcohols and

natural gas as high-octane fuels to use the hot

burned gases at as high temperature as possible

to promote auto-ignition of mixture, specifically

at low load.

REFRENCE

[1] J. B. Heywood, E. Sher, and S. o. A.

Engineers, The Two-Stroke Cycle Engine: Its

Development, Operation, and Design: Taylor

& Francis, 1999.

[2] G. P. Blair, and S. P. A. S. P. Committee,

Advances in Two-Stroke Cycle Engine

Technology: Society of Automotive

Engineers, 1989.

[3] S. Onishi, S. H. Jo, K. Shoda et al., ―Active

Thermo-Atmosphere Combustion (ATAC) -

A New Combustion Process for Internal

Combustion Engines,‖ 1979.

[4] M. Noguchi, Y. Tanaka, T. Tanaka et al., ―A

Study on Gasoline Engine Combustion by

Observation of Intermediate Reactive

Products during Combustion,‖ 1979.

[5] Y. Ishibashi, and M. Asai, ―Improving the

Exhaust Emissions of Two-Stroke Engines

by Applying the Activated Radical

Combustion,‖ 1996.

[6] R. H. Thring, ―Homogeneous-Charge

Compression-Ignition (HCCI) Engines,‖

1989.

[7] P. Duret, and J.-F. Moreau, ―Reduction of

Pollutant Emissions of the IAPAC Two-

Stroke Engine with Compressed Air Assisted

Fuel Injection,‖ 1990.


, 2012

6

[8] S. H. Jo, P. D. Jo, T. Gomi et al.,

―Development of a Low-Emission and High-

Performance 2-Stroke Gasoline Engine

(NiCE),‖ 1973.

[9] G. P. Blair, The Basic Design of Two-stroke

Engines: Society of Automotive Engineers,

1990.

[10] P. Duret, A New Generation of Engine

Combustion Processes for the Future?:

Proceedings of the International Congress,

Held in Rueil-Malmaison, France,

November, 26-27, 2001: Editions Technip,

2002.

[11] P. Duret, A New Generation of Two-stroke

Engines for the Future?: Proceedings of the

International Seminar Held in Rueil-

Malmaison, France, November 29-30, 1993:

Éditions Technip, 1993.

[12] Y. Ishibashi, ―Basic Understanding of

Activated Radical Combustion and Its Two-

Stroke Engine Application and Benefits,‖

2000.

[13] N. Iida, Y. Yamasaki, S. Sato et al., ―Study

on Auto-Ignition and Combustion

Mechanism of HCCI Engine,‖ 2004.

[14] P. Duret, J.-C. Dabadie, J. Lavy et al., ―The

Air Assisted Direct Injection ELEVATE

Automotive Engine Combustion System,‖

2000.

[15] K. Tsuchiya, Y. Nagai, and T. Gotoh, ―A

Study of Irregular Combustion in 2-Strote

Cycle Gasoline Engines,‖ 1983.

[16] Y. Ishibashi, and M. Asai, ―A Low Pressure

Pneumatic Direct Injection Two-Stroke

Engine by Activated Radical Combustion

Concept,‖ 1998.

[17] N. Iida, ―Combustion Analysis of Methanol-

Fueled Active Thermo-Atmosphere

Combustion (ATAC) Engine Using a

Spectroscopic Observation,‖ 1994.

[18] R. Stone, Introduction to Internal

Combustion Engines, 3rd Edition: Solutions

Manual, 1999.

[19] J. B. Heywood, Internal combustion engine

fundamentals: McGraw-Hill, 1988.

[20] C. F. Taylor, The Internal-combustion

Engine in Theory and Practice:

Thermodynamics, fluid flow, performance:

M.I.T. Press, 1985.

[21] A. Cairns, and H. Blaxill, ―The Effects of

Combined Internal and External Exhaust Gas

Recirculation on Gasoline Controlled Auto-

Ignition,‖ 2005.


2012

7

Determination of Motive Nozzle and Constant-Area Diameters:

Numerical Study of Ejector as an Expansion Device in Split-type

Air Conditioner

Kasni Sumeru

a, Henry Nasution, Farid Nasir Ani

*

aDepartment of Refrigeration and Air Conditioning

Politeknik Negeri Bandung, Indonesia

Email: [email protected]

Department of Thermodynamics and Fluid Mechanics, Faculty of Mechanical Engineering

Universiti Teknologi Malaysia, Skudai 81310 Johor

*Email: [email protected]

ABSTRACT This paper presents a numerical approach for

determining the motive nozzle and constant-area of

an ejector as an expansion device, based on cooling

capacity of the split-type air conditioner using R22

as working fluid. The use of an ejector as an

expansion device in split-type air conditioner can

improve the coefficient of performance (COP).

Typically, the split-type air conditioner may be

installed on the geographical area with moderate or

high outdoor air temperature using capillary tube.

For this reason, the motive nozzle and constant-

area diameters of the ejector must be designed

according to these conditions. The diameters of the

ejector are crucial in improving the COP. Three

equations are applied in developing the numerical

model on the ejector: conservation laws of mass,

momentum and energy equations. The results

showed that the motive nozzle diameter is constant

(1.14 mm) with variations of the condenser

temperature, whereas the constant-area diameter

decreases as the condenser temperature increases.

Keywords: Condenser temperature, COP, ejector,

expansion device, R22.

1. INTRODUCTION

Split-type air conditioner (AC) typically uses a

vapor compression refrigeration cycle (VCRC).

The air-conditioning system uses approximately

50% of the total energy consumption of a building

[1]. As a result, a small improvement on the

performance of the system will generate a

significant impact on energy saving.

Split-type air conditioner is the most widely used

as residential and commercial air conditioners. This

type of AC generally uses a capillary tube as an

expansion device. Shodiya et al. [2] developed a

numerical model to improve the capillary tube

performance prediction using enthalpy equation

and also included metastability phenomenon to

further improve the performance [3]. The use of

capillary tube generates irreversible process on the

throttling and causes energy loss during expansion

from high pressure to low pressure. Replacing

capillary tube by an ejector as an expansion device

in the AC is an alternative way of improving COP.

Theoretically, the pressure drop in the conventional

expansion devices is considered isenthalpic process

(constant enthalpy). Isenthalpic process causes a

decrease in the evaporator cooling capacity because

of energy loss in the throttling process. To recover

this energy loss, isentropic (constant entropy) is

required in the expansion process. An ejector can

be used to generate isentropic condition in the

throttling process.

Based on literature survey, we have not found any

study on the ejector-expansion refrigeration cycle

(EERC) that investigate determination of diameter

of the motive nozzle and the constant-area, based

on the cooling capacity of the air conditioner using

R22 as refrigerant. Hydrochlorofluorocarbons-22

(R22) is the most commonly used refrigerant in

split-type air conditioner. The objective of the

present study is to obtain the main geometric

parameter of an ejector, namely diameters of the

motive nozzle and the constant-area of ejector and

COP improvement on the split-type air conditioner

using R22 as the working fluid.

2. SYSTEM DESCRIPTION

2.1 Ejector Expansion Refrigeration Cycle

In 1931, Gay patented the ejector as an expansion

device in the refrigeration system. He patented the

ejector to minimize throttling losses in expansion

device on the vapor compression refrigeration

cycle. In 1966, Kemper et al. [4] modified the

Gay‘s patent, by using a pump and a heater to

increase pressure and temperature the liquid stream

before entering motive nozzle. In 1972, Newton

(1972) [5-6] proposed patent to improve previous

patent by Kemper [4]. Newton applied a hot gas

IMAT-UI 002



2012

8

from compressor discharge on the liquid stream

before entering to motive nozzle.

In 1950, Keenan et al. [7] carried out experimental

and numerical investigation and concluded that

there are two types of ejector, namely constant-area

mixing ejector and constant-pressure mixing

ejector. Figure 1 shows a constant-area mixing

ejector which has three sections: a nozzle section, a

constant-area mixing section, and a diffuser. In the

constant-area mixing ejector, the mixing between

the primary and secondary flow occurs in the inlet

of the constant-area.

Figure 1: Constant-area mixing ejector

Figure 2 shows a constant-pressure mixing ejector

which has four sections: a nozzle section, a mixing

section, a constant-area and a diffuser. The mixing

between primary and secondary flow occurs in the

mixing section or suction chamber, before

constant-area.

Figure 2: Constant-pressure mixing ejector.

The constant-pressure mixing ejector has a better

performance than that of the constant-area mixing

ejector [4], as a result, a constant-pressure mixing

ejector is generally used in the various refrigeration

applications, especially in the ejector refrigeration

systems. However, Yapici and Ersoy [8] found that

for the same operating temperature, the constant-

area mixing ejector has higher COP than that of the

constant-pressure mixing ejector. Furthermore, in

the last decade, the constant-area mixing ejector is

widely used in numerical and experimental studies

on the EERC [9-13].

Figure 3 depicts the schematic diagram of the

standard cycle and an ejector as an expansion

device. A capillary tube or expansion valve is used

as an expansion device in the standard cycle,

whereas an ejector is used for an expansion device

in the EERC.

Figure 3: Schematic diagram of the vapor

compression refrigeration cycle:

(a) Standard cycle, (b) Ejector-

expansion cycle.

2.2 COP Improvement on the VCRC

The advantages of the ejector as an expansion

device have been demonstrated by several

researchers. Kornhauser [14] was the first to

perform a thermodynamics analysis of vapor

compression refrigeration cycle using an ejector as

an expansion device. He proposed a one-

dimensional model in his study. He found that the

COPimp was up to 21% over the standard cycle. Liu

et al. [15], Li and Groll [16] and Deng et al. [17]

developed a mathematical model and found that the

COPimp was between 6-14%, 7-18% and 22%,

respectively, over the standard vapor compression

refrigeration cycle using CO2 as a refrigerant.

Takeuchi et al. [18] reported an increase of 45-64%

COPimp for a vehicle refrigeration system. Disawas

and Wongwises [19] and Elbel and Hrnjak [20]

carried out experimental investigation on an ejector

as the expansion device using R134a and CO2

respectively, and reported an increment in the COP

over the standard cycle. Nehdi et al. [9] presented a

numerical analysis to determine the effect of the

geometry of ejectors on system performance using

twenty synthetic refrigerants. They found that the

COPimp over the standard cycle is 22%. A

numerical study using natural refrigerant on the

EERC was also performed by Sarkar [13] and

reported that the maximum COPimp for isobutane,

propane and ammonia are 21.6%, 17.9% and

11.9% respectively. Bilir and Ersoy [11] performed

a computational analysis of the performance

improvement of ejector expansion cycle over

standard cycle. Their computational methods are

similar to that of Kornhauser [14]. Using an R134a

refrigerant, the COP improvement of the expansion

cycle over standard cycle is 10.1-22.34%. They

found also that the COP improvement increases

when the condenser temperature increases. This

means that the use of ejector instead of an

expansion valve is more advantageous in the air-

cooled condensers than that of water-cooled

condensers.


2012

9

Figure 4 illustrates the standard cycle and EERC in

the Ph diagram. The isenthalpic throttling process

is from point 3 to 11, while isentropic throttling

process is from point 3 to 4. Refrigerant flow on

the Ph diagram of standard cycle is point 8, 2b, 3,

11 and 8. There are two flow on the EERC,

primary and secondary flow. The primary flow is

circulated by a compressor through condenser,

ejector and separator (point 1, 2, 3, 4, 10, 5 and 1),

whereas the secondary flow circulates in the

capillary tube, evaporator, ejector and separator

(point 6, 7, 8, 9, 10, 5 and 6). The primary and

secondary flow mixes at constant-area and diffuser

(point 10 and 5).

Figure 4: Ph diagram of the ejector-

expansion and standard cycle.

As shown in Figure 4, the pressure at point 1 is

higher than that of suction pressure in the standard

cycle (point 8). This means that the compressor

work of the ejector expansion cycle is lower than

that of the standard cycle. Based on Figure 4, the

COP of standard refrigeration cycle is calculated

as,

comp

bcomp

ecomp

comp

estd

hhm

hhm

W

QCOP

)(

)(

82

118

(1)

since compe mm , equation (1) becomes,

comp

b

stdhh

hhCOP

)(

)(

82

118 (2)

where ηcomp is the isentropic efficiency of the

compressor which is calculated by an empirical

relation by Brunin et al. [21] as,

suct

disccomp

P

P01345.0874.0 (3)

The COP of the EERC is calcylated as,

comp

comp

ecomp

comp

eej

hh

hh

m

m

W

QCOP

)(

)(

12

78

(4)

Furthermore, the COP improvement of the EERC

over the standard cycle can be calculated by,

tds

tdsej

impCOP

COPCOPCOP

)( (5)

Two parameters viz. the entrainment ratio (ω) and

pressure lifting ratio (Plift) are used to investigate

the EERC performance.

c

e

m

m

(6)

8

1

,

,

P

P

P

PP

oute

outdif

lift (7)

Both quantities should be as large as possible to

obtain optimum COPimp. High ejector pressure

lifting ratio decreases the compression ratio of the

compressor. Increasing the mass entrainment ratio

reduces compressor mass flow rate for a given

cooling capacity. Ejector efficiency increases when

mass entrainment ratio and/or pressure lifting ratio

increase. However, the entrainment ratio cannot be

increased as high as possible, because it will cause

the flow of refrigerant on primary flow to reduce.

A high entrainment ratio will cause the primary

flow as a driven-flow to be weak.

The operation principle of ejector as an expansion

device is similar to the ejector function in other

applications, in which the primary flow from high

pressure induces the secondary flow from low

pressure in the suction section of ejector and brings

to a higher pressure at diffuser. Figure 5 illustrates

the refrigerant flow, pressure and velocity profile

inside an ejector. The motive flow or primary flow

from high pressure is accelerated and expands

through motive nozzle, from point 3 to 4. Very

high speed flow at point 4 causes pressure drop.

The low pressure at point 4 induces fluid from

secondary flow (point 8 to 9). The two flows mix in

the outlet of suction chamber and become one

stream in the constant-area. The mixing stream

flows through point 10 to the diffuser. In the

diffuser, the refrigerant experienced deceleration as

a result of pressure increase (point 5). The

refrigerant at point 5 circulates to separator. The

vapor refrigerant in the separator circulates through

compressor (point 1 to 2) as a primary flow,

whereas the liquid refrigerant from separator

circulates through expansion valve and evaporator

(point 6, 7 and 8), as a secondary flow.


2012

10

Figure 5: Refrigerant flow inside an ejector.

3. METHODOLOGY

Three equations viz. conservation of mass,

momentum and energy as shown in equation (8),

(9) and (10), respectively, are used to develop

thermodynamic model on each part of the ejector.

oooiii auau (8)

iiii umaP oooo umaP (9)

)2

()2

(22

o

oo

i

ii

uhm

uhm (10)

The following assumptions are made in the

calculation of each part of the ejector:

1. There are no heat transfer except in the

evaporator and condenser.

2. Properties and velocities are constant over the

cross section (one-dimensional).

3. The refrigerant condition is in thermodynamic

quasi-equilibrium.

4. There is no pressures drop along the evaporator

and condenser.

5. There is no wall friction.

6. The refrigerant conditions at the outlet of the

evaporator and condenser are saturated.

7. The pressure of exit of nozzle and suction

nozzle at the entrance of the constant-area are

assumed to have the same pressure.

8. Deviation from adiabatic reversible processes

for each section of ejector is calculated by

efficiencies.

In the present study, a constant-area mixing ejector

flow model is used as an expansion device. Based

on the manipulation of equations (1)-(3) and

referring to studies performed by Nehdi et al. [9]

and Sarkar [13], the thermodynamics modeling

produces equations (11)-(20), as shown in the

flowchart depicted in Figure 6. This flowchart was

used to determine the diameter of motive nozzle

(d4) and the constant-area of the ejector (d10). The

properties of the refrigerant were obtained from

REFPROP [22]. Using flowchart (Figure 6) and

properties of refrigerant, several parameters such as

the diameter of the motive nozzle and constant-

area, Plift, COPej and COPimp can be calculated.

Set n,d

Determine Te and Tc

At nozzle outlet:

h4 = h3 -n(h3-h4,is) (11)

u4 = [2(h3-h4)]0.5

44

4u

ma c

(12)

(13)

At mixing chamber:

(14)

(15)

(16)

2)(

1

1 2

108310

uhhh

At diffuser outlet:

2

2

10105

uhh d (20)

Determine P4 and ω

Update P4

1)1( 5 x

COPej, COPimp

d4, d10, Plift

Yes

No

1010

10u

mma ec

10410410 )()( ummumaPP ecc

2

10

4

10

42

10

4

2

44

410 )1(225.0

a

a

a

a

u

PP

1

1

1 4

8

4

10

(17)

4101

1uu

(18)

(19)

0)( 410 PP

Yes

Update ω and P4

No

evapTT 5

Figure 6: Flow chart of the calculation

algorithm of motive nozzle and

constant-area diameter.

4. RESULTS AND DISCUSSION

In this study, the cooling capacity and the mass

flow rate of refrigerant R22 in the split-type air

conditioner was taken based on the experimental

result of Zhou and Zhang [23], namely 2.4 kW and

56.95 kg/h, respectively. To start the iteration using

the flowchart in Figure 6, the values of nozzle and

diffuser efficiencies is chosen as 0.9 and 0.8,

respectively [11]. The influence of condenser

temperature on diameters, COPimp and Plift will be

explained in the subsection below.


2012

11

4.1 Influence on the diameters

Split-type AC may be installed on the geographical

area with moderate or high outdoor air temperature.

For this reason, the motive nozzle and constant-

area diameters of the ejector must be designed

according to these conditions. The dimension of

each section of the ejector is crucial in improving

the COP. The diameter of the motive nozzle is

calculated by using equation (13), while the

diameter of the constant-area is iterated by the

flowchart depicted in Figure 6. The iteration result

is shown in Figure 7. It shows that the diameter of

the motive nozzle is constant, at 1.14 mm, except at

the condenser temperature of 40oC, which is 1.15

mm. The size of the diameter of the motive nozzle

based on the numerical results is similar to those

used in the experiment by Caiwongsa and

Wongwises [24]. They tested three different motive

nozzle diameters 0.8, 0.9 and 1.0 mm, using

R134a. The motive nozzle with a diameter of 1.0

mm yields a higher mass flow rate in the condenser

than the other nozzles. Meanwhile, the smallest

diameter (0.8 mm) yields a lower mass flow rate in

the condenser, producing the highest COPimp. There

is slight difference between these numerical

approaches with the experimental data. This

distinction is caused, among other things, by the

differences in the working fluid and cooling

capacity.

The effect of the constant-area diameter with

increase in condenser temperature is shown in Fig.

7. The figure shows that the diameter of constant-

area decreases with increase in condenser

temperature. For example, diameters of the

constant-area are 2.49, 2.51, 2.48 and 2.46 mm

when the condenser temperatures are 40, 45, 50,

and 55oC, respectively. The figure also shows that

the area ratio (AR), the ratio between the cross-

sectional area of constant-area to motive nozzle

(a10/a4), decreases as the condenser temperature

increases. With similar results in the present study,

Sarkar [13] reported that the ejector AR decrease

with increase in temperature of condenser. The

results of the present study showed that the

decrement of the area ratio of ejector is more

influenced by constant-area diameter than the

motive nozzle diameter, as shown in Figure 7.

Figure 7: Variation of the motive nozzle,

constant-area diameters and area

ratio of ejector, versus condenser

temperature. (Te = 5oC, ηn = 0.9

and ηd = 0.8).

4.2 Influence on the COPimp

According to energy analysis, the COP of a

standard refrigeration cycle decreases as the

condenser temperature increases, as shown in

Figure 8. For the EERC, the COP reduction due to

increase in condenser temperature is lower than

that of the standard cycle. As a result, the increase

of the COPimp of EERC is very significant for high

condenser temperature. For example, as seen from

Figure 8, the COPimp is 14.82% at Tc = 50oC and

becomes 23.03% at Tc = 55oC. These results

indicate that the use of an ejector as an expansion

device is effective for geographical areas which

have high outdoor air temperature.

Figure 8: Variation of the COP and COP

improvement, versus condenser

temperature. (Te = 5oC, ηn = 0.9

and ηd = 0.8).


2012

12

4.3 Influence on Plift

To obtain the maximum COPimp, the Plift value

should be as high as possible. However, according

to the flowchart as shown in Figure 6, the condition

of (1+ω).x5 = 1 must be fulfilled for realistic flow.

As depicted in the Ph diagram in Figure 4, to

reduce the compressor work, P1 should be as high

as possible, in order to obtain the minimum

enthalpy difference (h2 - h1). In this study, the Plift is

calculated by equation (7) with the evaporator

temperature of 5oC. In the saturation condition, for

R22 refrigerant, the evaporator temperature of 5oC

is equal to evaporator pressure of 584.11 kPa. The

iteration results for various P1, T5 and Plift with

different condenser temperature is shown in Table

1.

Table 1: Variation P5 and T5 versus the condenser

temperature, where Te = 5oC, ηn= 0.9 and ηd=0.8.

Tc (oC) 40 45 50 55

P5 (kPa) 590.05 601.52 613.59 626.91

T5 (oC)

Plift

5.44

1.01

5.94

1.02

6.58

1.04

7.28

1.06

Pre

ss

ure

(P

)

Specific enthalpy (h)

1

2

3

4

6

7 8

910

Pc=1533.6 kPa Tc = 40oC

T5 = 5.44oC

Te = 5oC

T4 = 3.44oC

P5 =

590.0

5 k

Pa

Figure 9: Ph diagram of the EERC (Te =

5oC, Tc = 40

oC, ηn = 0.9 and ηd =

0.8).

Figure 9 shows a detail of EERC of split-type air

conditioner using R22 as the working fluid, with Tc

= 40oC, Te = 5

oC on Ph diagram. The temperature

and pressure data are obtained from iteration using

flowchart in Figure 6. It can be seen from Table 1

and Figure 9 that since the value of Plift is small, the

different between T5 and evaporator temperature

(Te) is also small.

5. CONCLUSION

It has been found that the motive nozzle diameters

are much closer to the experimental results

obtained by other researchers. The results of this

study showed that the decrement of the area ratio of

ejector is more influenced by constant-area

diameter than the motive nozzle diameter.

Replacement of the capillary tube with an ejector

as the expansion device on split-type residential

AC using R22 as the working fluid can improve the

COP, particularly at condenser temperature above

40oC. This indicates that the use of an ejector as an

expansion device is recommended for geographical

areas which have high outdoor air temperature. The

authors hoped that this present study will

encourage other researches on EERC because there

are still many approaches that have to be explored

to determine the dimensions of the other geometric

parameters of the ejector in accordance with the

cooling capacity of the refrigeration system

ACKNOWLEDGMENT

The present study was supported financially by

ASHRAE and Universiti Teknologi Malaysia: GUP

TIER 2 Fund No. 00J25 from the Ministry of

Higher Education (MOHE) Malaysia.

REFERENCES

[1] L. P. Lombard, J. Ortiz, and C. Pout, ―A

review on buildings energy consumption

information,‖ Energy and Buildings, vol. 40,

pp. 394-398, 2008.

[2] S. Shodiya, A. A. Azhar, and A. N. Darus,

―Improved refrigerant characteristics flow

predictions in adiabtic capillary tube,‖

Research Journal of Applied Sciences,

Engineering and Technology, vol. 4, pp.

1922-1927, 2012.

[3] S. Shodiya, A. A. Azhar, N. Henry, and A. N.

Darus, ―Numerical simulation of refrigerant

flow in adiabatic capilarry tubes including

metastability phenomenon,‖ Proceeding of

11th Asian International Conference on Fluid

Machinery and The 3rd Fluid Power

Technology Exhibition, IIT Madras, Chennai,

India. pp. 1-14, 2011.

[4] G. A. Kemper, G. F. Harper, and G. A.

Brown, Multiple phase ejector refrigeration

system, US Patent Patent No.3,277,660, 1966.

[5] A. B. Newton, Capacity control for

multiphase-phase ejector refrigeration system,

US Patent No. 3,670,519, 1972a.

[6] A. B. Newton, Control for multiphase-phase

ejector refrigeration system, US Patent No.

3,670,519, 1972b.

[7] J. H. Keenan, E. P. Neumann, F. Lustwerk,

―An investigation of ejector design by


2012

13

analysis and experiment,‖ Journal of Applied

Mechanics, vol. 17, pp. 299-309, 1950.

[8] R. Yapıcı, H. K. Ersoy, ―Performance

characteristics of the ejector refrigeration

system based on the constant-area ejector

flow model,‖ Energy Conversion and

Management, vol. 46, pp. 3117-3135, 2005.

[9] E. Nehdi, L. Kairouani, and M. Bouzaina,

―Performance analysis of the vapour

compression cycle using ejector as an

expander,‖ International Journal of Energy

Research, vol. 31, pp. 364-375, 2007.

[10] S. Elbel, P. Hrnjak, ―Experimental validation

of a prototype ejector designed to reduce

throttling losses encountered in transcritical

R744 system operation,‖ International Journal

of Refrigeration, vol. 31, pp. 411-422, 2008.

[11] N. Bilir, H. K. Ersoy, ―Performance

improvement of the vapour compression

refrigeration cycle by a two-phase constant-

area ejector,‖ International Journal of Energy

Research, vol. 33, pp. 469-480, 2009.

[12] H. K. Ersoy, N. Bilir, ‗The influence of

ejector component efficiencies on

performance of Ejector Expander

Refrigeration Cycle and exergy analysis,‖

International Journal of Exergy, vol. 7 pp.

425-438, 2010.

[13] J. Sarkar, ―Geometric parameter optimization

of ejector-expansion refrigeration cycle with

natural refrigerants,‖ International Journal of

Energy Research, vol. 34, pp. 84-94, 2010.

[14] A. A. Kornhauser, ―The use of an ejector as a

refrigerant expander,‖ In: Proceeding of the

USN/IIR-Purdue Refrigeration Conference.

West Lafayette, IN, USA, pp.10-19, 1990.

[15] J. P. Liu, J. P. Chen, and Z. J. Chen,

―Thermodynamic analysis on trans-critical

R744 vapor compression/ejection hybrid

refrigeration cycle,‖ In: Proceeding of the

Fifth IIR Gustav Lorentzen Conference on

Natural Working Fluid. Guangzhou, China,

pp.184-188, 2002.

[16] D. Li, E. A. Groll, ―Transcritical CO2

refrigeration cycle with ejector-expansion

device,‖ International Journal of

Refrigeration, vol. 28, pp. 766-773, 2005,

[17] J. Q. Deng, P. X. Jiang, T. Lu, and W. Lu,

―Particular characteristics of transcritical CO2

refrigeration cycle with an ejector,‖ Applied

Thermal Engineering, vol. 27, pp. 381-388,

2007.

[18] H. Takeuchi, H. Nishijima, and T. Ikemoto,

World's first high efficiency refrigeration

cycle with two-phase ejector: "ejector cycle",

In: SAE World Conggres. Detroit, MI, USA,

Paper 2004-01-0916, 2004.

[19] S. Disawas, S. Wongwises, ―Experimental

investigation on the performance of the

refrigeration cycle using a two-phase ejector

as an expansion device,‖ International Journal


[20] S. Elbel, P. Hrnjak, ―Experimental validation

of a prototype ejector designed to reduce

throttling losses encountered in transcritical

R744 system operation,‖ International Journal


[21] O. Brunin, M. Feidt, and B. Hivet,

―Comparison of the working domains of some

compression heat pumps and a compression-

absorption heat pump,‖ International Journal

of Refrigeration, vol. 20 pp. 308-318, 1997.

[22] E.W. Lemmon, M.L. Huber, and M.Q.

McLinde, REFPROP, Reference Fluid

Thermodynamics and Transport Properties,

NIST Standard Reference Database 23,

Version 9.0., 2009.

[23] G. Zhou, Y. Zhang, ―Performance of a split-

type air conditioner matched with coiled

adiabatic capillary tubes using R22 and

HC290,‖ Applied Energy, vol. 87, pp. 1522-

1528, 2010.

[24] P. Chaiwongsa, S. Wongwises, ―Effect of

throat diameters of the ejector on the

performance of the refrigeration cycle using a

two-phase ejector as an expansion device,‖

International Journal of Refrigeration, vol. 30,

pp. 601-608, 2007.


2012

14

Effects of External Hard Particles on Brake Noise of Disc

Braking System

M. A. Nasaruddina, M. K. Abdul Hamid

a, A.R. Mat Lazim

a, and A.R. Abu

Bakara

aDepartment of Automotive Engineering, Faculty of Mechanical Engineering,

Universiti Teknologi Malaysia 81310 Johor Malaysia.

Tel : (607) 5534667. Fax : (607) 5566159

E-mail : [email protected], [email protected], [email protected], [email protected]

ABSTRACT

The open design and position of disc brake that is

closed to road surfaces enable contaminants to

enter the brake gap and caused noise and

tribological disturbance at the brake interface.

Contaminants such as dirt and soil can be present

and are expected to influence the occurrence of

brake squeal that produce an annoying sound

during braking action. The objective of this study

was to examine the effect of external hard particles

at different disc sliding speed on generation of

brake squeal using a brake dynamometer. Different

rotational speed of disc brake was selected and the

experiments squeal noise data was collected and

analyzed using the Fast Fourier Transformation

(FFT) analyzer. From the experiments, the

presence of external particle and the rotation speed

of disc brake promotes the generation of brake

squeal phenomenon by changing the surface

roughness and effective contact of brake interface.

Results obtained from the experiment also showed

that higher rotating disc generate higher sound

level meter or squeal frequency and increase

numbers of squeal noise generated.

Keywords : External particles, brake squeal, noise

level, surface roughness, effective

contact

1. INTRODUCTION

Higher frictions have significant impact on noise

characteristics. Chen et al. [1] stated that the higher

coefficient of friction (cof) of the brake pad, the

higher squeal tend to occur. However, high cof

does not necessary be the main reason of squeal but

it plays big role in noise characteristics of brake

systems. Surface roughness is another factor that

can affect the squeal noise. Generally smooth

friction surface may provide more stable contact

between sliding surfaces and cause less system

vibration and noise occurrence. Abdul Hamid [2]

states present of external particle give different

value of surface roughness and it tends to reduce

with small external particle such as silica sand and

dust. However, the surface roughness influences

not only the contact pressure distribution but the

stability and noise of the system. According to

Mario et al. [3], number of instability intensity of

unstable modes generated by high friction model

significantly increased and cof is directly

proportional to squeal propensity. In this work, the

external particle influence on surface roughness,

noise level and vibration of brake pad in generating

squeal was studied.

2. METHODOLOGY

Brake dynamometer testing was utilized to study

the effect of external particle on squeal noise of

disc braking system. Present of different size and

shape of external particle are expected to influence

surface roughness and the occurrence of squeal.

2.1 Test Rig

Figure 1 shows the schematic diagram of the test

rig. The drag type brake dynamometer was used in

this experiment in order to validate brake squeal

performance. The test rig was mounted separately

on the two units of I beam. The disc brake was

mounted on the main shaft and the caliper holder

was located below the caliper. The flange was used

to drive the disc brake instead of the main shaft.

To control the applied pressure, a hydraulic unit

was used to apply pressure to disc brake through

brake pad with maximum pressure of 20 bars. Four

transducers used for this experiment are load

transducer to measure the load during shaft

rotation, accelerometer to measure vibration of the

brake pad during squeal, a microphone and a set of

amplifier to record the squeal sound and its

frequency. Data Acquisition System (DAQ) was

used to collect information from the transducer,

and process it to suitable data for display or storage

outcomes. A small plastic tube was used to direct

the external particle from the container to the gap

of the brake disc. A transparent cover was used to

avoid splashing of the external particle particles

during the experiments.

IMAT-UI 003






2012

15

Figure 1: Test rig of brake noise dynamometer

2.2 Testing Procedures

Tests are carried out using brake dynamometer

equipped with actual pad and disc brake of the

conventional car. The arithmetic surface roughness

of the pad (Ra) was measured before and after the

brake test with a surface roughness machine. For

the external particle, to get the required size and

shape, the external particles were filtered by a

shaker machine. There are two sections in squeal

test procedure. First is bedding in process and

second is drag noise procedure. Bedding in process

is the process to make sure that the brake pad and

disc brake are aligned and for the new brake pad to

have maximum contact between disc and brake

pad. A series of tests was conducted at three

different disc rotational speeds of 50 rpm, 60 rpm

and 70 rpm while the pressure of 10 bars. Each

experiment was supplied with 35g of external road

particles.

3. RESULTS AND DISCUSSIONS

Presence of external particles changed the pad

surface roughness and increased the total effective

contact area and resulted in increasing cof value.

Table 1 shows the minimum and maximum value

of mean deviation of the surface height (Ra) or

average roughness of pad samples scan area. The

surface roughness of pad samples result showed a

significant increasing of Ra value of both sides

during squeal. This experiment reveals that surface

roughness values increase with the external

particles and affect the tendency of squeal noise

occurrence.

Table 1: Surface roughness values of brake pad. Brake pad Ra min (μm) Ra max (μm)

Piston_side

(Before test) 9.511 11.546

Finger_side

(Before test) 10.178 17.523

Piston_side

(After test) 3.761 17.397

Finger_side

(After test) 3.743 17.786

Figure 2: Graph of sound level versus squeal

frequency with external particle at

different speed.

Figure 2 shows the result of sound level versus

squeal frequency with the existence of external

particle at different speed of rotating disc. The

number of squeal generated for each speed was 15

at 70 rpm and 10 squeal was recorded for 50 and 60

rpm. The maximum squeal frequency recorded at

speed of 50 rpm was 81 dB at 4255 Hz which is

lower compared to the speed of 60 rpm that has the

reading of 84 dB at 4462Hz. The conclusion from

the graph is that, with present of external particle,

higher sound level and squeal frequency will be

generated at higher rotating disc speed.

Figure 3: Graph of sound level versus

vibration with external particle at

different speed.

Figure 3 shows the relationship between sound

level meter and vibration for three different speeds.

The average vibration for all speed recorded was

around 0.53 ms⁻² to 0.85 ms⁻² while the average

value for sound level meter is from 79 dB to 82 dB.

The maximum vibration occurred at 1.2 ms⁻² for

sliding speed of 70 rpm, while for 60 rpm

maximum vibration occurred at 0.94 ms⁻² and the

maximum vibration for 50 rpm occurred at 0.75

ms⁻². Lowest value of vibration for all the speeds

was almost the same at about 0.31 ms ⁻². Thus,

higher sliding speed will generate higher sound


2012

16

level meter and produce higher vibration value with

the existence of external particle.

4. CONCLUSION

The external particles effect on brake noise of disc

braking system was investigated using a specially

developed brake test rig with the actual size of

external particle below 400 µm. The following

conclusions can be made:

The external particles increase the surface

roughness of the pad interface and affect the

tendency of squeal noise occurrence.

The number of squeal generated increases

with higher rotating disc speed.

The frequency of the squeal also increases

proportionally with the sound level meter

reading.

The vibration of the brake pad increase with

increasing in speed of the rotational disc

brake.

ACKNOWLEDGMENT The authors would like to thank Universiti

Teknologi Malaysia for supporting this

researchwork under the University Grant Project

(GUP-Tier 2 Vot 00J45).

REFERENCES

[1] F. Chen, F. Tan, F. C. Chen, C. A. Tan, and

R.L. Quaglia, ―Disc Brake Squeal: Mechanism,

Analysis, Evaluation and Reduction/Prevention‖,

SAE-Society of Automotive Engineers, 2006

[2] M. K. Abdul Hamid, ‗Study of the grit particle

size and shape effects on the frictional

characteristics of the automotive braking system.

PhD Thesis. University of Western Australia, 2010.

[3] T.J., Mario, N.Y., Samir, and J., Roberto,

―Analysis of brake squeal noise using finite

element method: A parametric study‖, Journal of

Applied Acoustics, 69, pp. 147-162, 2008.

COPYRIGHT The author confirm that this papers is original, and

has not been published or under consideration for

publication elsewhere.


2012

17

Experimental Study on the Replacement of HFC-R134a by

Hydrocarbons in Automotive Air Conditioner

Mohd Rozi Mohd Perang

a, Henry Nasution

a,b,c, Zulkarnain Abdul Latiff

a,b,

Azhar Abdul Aziza,b

, Afiq Aiman Dahlanb

aAutomotive Development Centre, Universiti Teknologi Malaysia

81310 Skudai, Johor, Malaysia.

Phone: +60 75535447, Fax: +60 75535811 bFaculty of Mechanical Engineering, Universiti Teknologi Malaysia

81310 Skudai, Johor, Malaysia.

Phone: +60 7 5534575, Fax: +60 7 5566159 cDepartment of Mechanical Engineering, Bung Hatta University

25134 Padang, Sumatera Barat, Indonesia.

Phone: +62 751 7054657, Fax: +62 751 7051341

E-mail: [email protected], [email protected], [email protected], [email protected]

ABSTRACT Performance characteristics of the current

automotive air conditioning system have been

evaluated in this experimental study which will

evaluate the power consumption, temperature

distribution and coefficient of performance (COP)

at various internal heat loads and engine speed

using hydro-chlorofluorocarbons refrigerant (HFC-

R134a) and hydrocarbon refrigerant (HC-R134a) as

the working fluid of the compressor. Both

refrigerants will be tested on the experimental rig

which simulated the actual cars as an internal cabin

complete with a cooling system component of the

actual car including the blower, evaporator,

condenser, radiator, electric motor, compressor and

alternator. The electric motor acts as a vehicle

engine, and then will drive the compressor using a

belt and pulley system, as well as to the alternator

to recharge the battery. The rig also equipped with

simulation room acting as the passenger

compartment. The tests have been performed by

varying the motor speed; 1000, 1500, 2000, 2500

and 3000 rpm, temperature set-point; 21, 22 and

230C, and internal heat loads; 0, 500, 700 and 1000

W. As the results, the performance characteristics

of the HC-R134a indicate the positive

improvement of the system compared to HFC-

R134a.

Keywords : Air conditioning, automotive, HC-

134a, hydrocarbon refrigerant,

performance, energy saving.

1. INTRODUCTION

The heating, ventilation and air conditioning

(HVAC) of the automotive is designed to provide

thermal comfort level of the driver and passengers.

Thermal comfort is the crucial things to be

fulfilled. Human thermal comfort is defined by the

American Society of Heating, Refrigeration and

Air Conditioning Engineers (ASHRAE) as the

state of mind that expresses satisfaction with the

surrounding environment (ASHRAE Standard 55)

[1].The function of an air conditioning (A/C)

control system is to modulate the A/C system

capacity to match off the design condition, load

variation and climate change, to maintain the

indoor environment within desirable limits at

optimum energy use during the entire drive.

Automotive A/C component have been going

through a steady evolution since the introduction

of A/C in cars in 1940. From the early days to

today, A/C was a very expensive option in luxury

cars when it is standard equipment in many models

and many different types of systems have been

used [2].

Hydrocarbon (HC) is new findings for refrigerants

by the experts to replace the current HFC-R134a

as the refrigerant for A/C system. HC is used as

refrigerants gases at an early stage, which were

accepted before the emergence of CFCs

(chlorofluorocarbons) and HCFCs

(hydrochlorofluorocarbons). After a long period,

HC refrigerant is no longer in use because of the

flammability characteristic. Thus, CFCs and

HCFCs are being used instead of using HC

refrigerant, which is flammable, not practical and

harmful to a user [3]. However, several studies

have shown that the used of hydrocarbons in A/C

system would improve the performance of the

system [4, 5]. Some applications of the HC can be

found in aerosol filler, the heating fuels in gas

ovens, etc. The HC used in the A/C and freezer as

the working fluid is not yet in common.

The research work will perform the results of the

experimental studies on HC mixtures (HC-134a)

as an alternative refrigerant to the automotive A/C

system in Malaysia. HC mixture used for this work

contains propane (R290), butane (R600), and

isobutane (R600a). The mixture is known as HC-

134a and Wongwises et al. [6, 7], Ghodbane [8]

and Tashtoush et al. [9] are one of the researchers

IMAT-UI 004






2012

18

that involved in automotive HC A/C system. Table

1 shows the properties of the refrigerant used in

this experimental work [6, 10].

Table 1: Properties of refrigerant

HC refrigerant is generally considered as

environmental friendly and has been used again

regarding to the noticeable value of the Global

Warming Potential (GWP) and zero Ozone

Depletion Potential (ODP).

2. EXPERIMENTAL APPARATUS AND

PROCEDURE

In this research work, the performance

characteristics of the automotive A/C system have

been performed via experimental analysis. The test

is performed on the automotive A/C experimental

rig in order to evaluate the power consumption,

temperature distribution and coefficient of

performance (COP). The refrigerants used are

using HFC-R134a and (HC-R134a) as the working

fluid of the compressor. The test is done at various

internal heat loads, temperature setting and engine

speed. Figure 1 illustrates the schematic diagram of

automotive A/C system that has been used in this

work.

Figure 1: Automotive A/C system

experimental rig

Figure 2 shows on the experimental rig which

simulated the actual cars as an internal cabin

complete with A/C system component of the actual

car including the blower, evaporator, condenser,

radiator and compressor.

Figure 2: Experimental test rig

Other components involved in this work are the

electric motor and alternator. The electric motor

worked as a vehicle engine, and then will drive the

compressor using a belt and pulley system, as well

as to the alternator to recharge the battery. The

compressor was run by 12 volt original car battery


2012

19

which continuously charged by the alternator to

ensure the whole range of rotating speeds is like the

actual car speed. The rig also equipped with

simulation room acting as the passenger

compartment.

The temperatures and pressures parameters were

measured by fitted the type T thermocouple

(accuracy ±0.1oC) and flow meter (accuracy of ±1

gr/s) respectively, at several locations as shown in

Figure 1. Then, the current and voltage of the

electric system were measured in order to obtain

the compressor energy consumption. The accuracy

of the current meter is ±1%, and voltage meter is

±1.5%.

120 experiments have been done in this work by

varying the motor speed, internal heat load and

cabin temperature using HFC-R134a and HC-

R134a simultaneously, as the refrigerant for the

A/C system. Varied parameters are listed and

shown in Table 2. Table 2: Varied parameters

Parameter Range of variation

Motor speed, N

(rpm)

1000, 1500, 2000, 2500 and

3000

Cabin temperature

(°C) 21, 22 and 23

Type of

refrigerants HFC-R134a and HC-R134a

Internal heat load

(W) 0, 500, 700 and 1000

The experiment was carried out according to the

following procedures: 2.1 The A/C system was evacuated using a

vacuum pump.

2.2 Refrigerant R134a has been charged into the

system.

2.3 The A/C system was started and let the system

running for 15 minutes.

2.4 The cabin temperature was set at 21°C.

2.5 The speed of the motor was set at 1000 rpm.

2.6 The internal heat load was set by switching on

the bulb with 0, 500, 700 and 1000W

simultaneously. Each data was collected after

15 minutes of the A/C system run.

2.7 Procedures 2.7-2.6 were repeated with the

motor speed of 1500, 2000, 2500 and 3000

rpm simultaneously.

2.8 Procedures 4-7 were repeated with the cabin

temperature of 22 and 23°C simultaneously.

2.9 Procedures 1-8 were repeated with the

refrigerant of HC-R134a.

In collecting the data, the thermostat was set to the

maximum cool position. Data was collected when

the A/C system was considered stable after it ran

for 15 minutes for each test condition.

Thermodynamic properties of the HFC-R134a and

HC-R134a used were taken from the REFPROP

database [10] and from the thermodynamic

properties given by the company respectively.


The data from the experiment was analyzed by

varying two parameters and unvarying two other

parameters. The analysis was done on measured

parameters to obtain the performance of the A/C

system. The parameters are the coefficient of

performance (COP), power consumption consumed

by the compressor and compression ratio (Cr). COP

of the system is a relationship between the energy

released from the evaporator (refrigerating effect

(Qe)) and the energy required by the compressor

(Wc). COP was calculated by using equation (1)

[11]:

COP=Qe/Wc= (h1-h4)/(h2-h1) (1)

Where,

Qe= Cooling capacity of evaporator, (kJ/kg)

Wc= Compressor power, (kJ/kg)

h1=Enthalpy on suction compressor, (kJ/kg)

h2= Enthalpy on discharge compressor, (kJ/kg)

h3 = Enthalpy on condenser exit, (kJ/kg)

h4 = Enthalpy on evaporator inlet. (kJ/kg)

In order to get the enthalpy through REFPROP

software, refrigeration cycles necessitate to be

illustrated on the P-h (pressure vs. enthalpy)

diagram as illustrated in Figure 3 [12].

Figure 3: P-h diagram of refrigeration system

3.1 Temperature Distribution of System

Figure 4(a), 4(b) and 4(c) show the graphs of

temperature distribution against the internal heat

load (0, 500, 700 and 1000W) at the compressor

speed of 1000, 1500, 2000, 2500 and 3000 rpm and

the set-point temperature of 21, 22 and 23˚C for

HC-R134a and HFC-R134a.


2012

20

(a) At 21

0C

(b) At 22

0C

(c) At 23

0C

Figure 4: Graph of temperature distribution

against load (at compressor speed of

1000 – 3000 rpm).

From the figures above, the temperature

distribution of HC-R134a is always lower (up to

5%) than HFC-R134a. This is because the cooling

capacity of the HC-R134a is higher than HFC-

R134a. At all temperature set-point, the HC-R134a

is nearer to the set-point temperature at compressor

speed of 1000, 1500 and 2000 rpm. This will

indicate the used of HC-R134a is better than HFC-

R134a because the cooler air is distributed faster to

all compartment of the cabin.

3.2 Coefficient of Performance (COP)

Figure 5(a), 5(b) and 5(c) exhibits the graph of

COP against load at variable compressor speed of

1000 - 3000 rpm with refrigerant of HFC-R134 and

HC-R134a. COP of HC-R134a is higher than HFC-

R134a in which is up to 40%. This is a positive

improvement because HC-R134a produced high

COP which is indicating the better performance

than HFC-R134a.

When the compressor speed increases with the

increment of internal heat load and compressor

speed, the COP will decrease. Thus COP will

decrease when compressor work increases at

constant condition of the evaporator heat

absorption.


2012

21

(a) At 21

0C

(b) At 22

0C

(c) At 23

0C

Figure 5: Graph of COP against load (at

compressor speed of 1000 – 3000

rpm).

There is also a relationship between COP of

different temperature set-point and internal heat

load at compressor speed of 1000 rpm. It can be

observed from the figures above that illustrate the

temperature set-point increased, the COP will

increase. Therefore, COP will be high when

compressor work is constant but evaporator heat

absorption is lower.

3.3 Compressor Ratio (Cr)

Figure 6(a), 6(b) and 6(c) illustrate the graph of

compression ratio against internal heat load which

is showing Cr is decreased with the increasing of

internal heat load. And as the temperature set-point

increased, the compression ratio will be decreased.

This shows that the compression ratio of HC-

R134a is better than HFC-R134a. This is an

encouraging improvement as the compression work

increased with less energy needed to cool the same

set-point temperature. The compression ratio will

affect the energy desired for compressor.


2012

22

(a) At 21

0C

(b) At 22

0C

(c) At 23

0C

Figure 6: Graph of compression ratio against

load (at compressor speed of 1000 –

3000 rpm).

3.4 Compressor Power Consumption

Figure 7(a), 7(b) and 7(c) show the relationship

between power consumption and internal heat load

at various speed of compressor (1000 - 3000 rpm).

The compressor power consumption is increased

proportionally with the increasing of internal heat

load. At the temperature set-point of 21, 22 and

23°C, the graphs show the same trend of rising. At

every temperature set-point temperature and with

the increasing of internal heat load, the compressor

needs to work more often and more power will be

consumed. As a result, the compressor uses more

power to compress the refrigerant to higher

pressure to maintain the temperature in the cabin.

The compressor power consumption of HC-R134a

is lower than HFC-R134a which indicates a

positive improvement up to 25% (an average

decreased is 15%).


2012

23

(a) At 21

0C

(b) At 22

0C

(c) At 23

0C

Figure 7: Graph of power consumption against load

(at compressor speed of 1000 – 3000 rpm).

4. CONCLUSIONS

From the discussions above, the HC-R134a

indicates a positive improvement of performance

characteristics compared to HFC-R134a in term of

COP, temperature distribution and power

consumption. Therefore, the HC-R134a is

suggested to be the replacement of current HFC-

R134a used in the automotive industry.

Energy consumption will vary with the changes of

the A/C compressor speed. Whilst the compressor

speed increases, the room temperature is going to

decreases as well as the COP is decreased. As the

results, the consumption of energy will increase

and less value of the energy can be saved and vice

versa.

ACKNOWLEDGEMENTS

Special thanks to Thermodynamic Laboratory,

Faculty of Mechanical Engineering, Universiti

Teknologi Malaysia for their facilitating support in

this study. Their guidance and assistance are highly

appreciated.


2012

24

REFERENCES

[1] ANSI/ASHRAE Standard 55, Thermal

environmental conditions for human

occupancy. American Society of Heating,

Refrigerating and Air-Conditioning Engineers,

Inc., 2008.

[2] Tom Birch, ―Automotive Heating and Air

Conditioning 2nd

Edition‖, Prentice Hall, 2000.

[3] Granryd, E., Hydrocarbons as refrigerants –

an overview. International Journal

Refrigeration, 24:15-24, 2001.

[4] D. Jung and C.B. Kim., Testing of

propane/isobutene mixture in domestic

refrigerator. International Journal

Refrigeration, 2000.

[5] K. Mani and V. Selladurai, Experimental

Analysis of a new refrigerant mixtures as drop-

in replacement for CFC12 and HFC134a.

International Journal of Thermal Sciences,

2008.

[6] Wongwises, S., Kamboon, A., and Orochon,

B., Experimental investigation of hydrocarbon

mixtures to replace HFC-134a in an automotive

air conditioning system, Energy Conversion &

Management; 47:1644-1659, 2006.

[7] Wongwises, S. and Chimres, N., Experimental

study of hydrocarbon mixtures to replace HFC-

134a in a domestic refrigerator, Energy

Conversion & Management; 46:85-100, 2005.

[8] Ghodbane. M., An Investigation of R152a and

Hydrocarbon Refrigerants in Mobile Air

Conditioning, SAE Paper 1999-01-0874, 1999.

[9] Tashtoush, B., Tahat, M., Shudeifat, M.A.,

Experimental study of new refrigerant mixtures

to replace R12 in domestic refrigerators.

Application Thermal Engineering; 22:495-506,

2002.

[10] REFPROP, Thermodynamic properties of refrigerants and refrigerant mixtures, Version

6.1, Gaittherbeurg, MD, National Institute of

Standards and Technology, 1998.

[11] Y.A. Cengel and M.A. Boles.,

Thermodynamics: An engineering Approach

(6 ed.). McGraw-Hill, 2007.

[12] Henry Nasution, et. al., ―Experimental

Evaluation of Automotive Air-Conditioning

Using HFC-134a and HC-134a‖, American

Institute of Physics, 2012


2012

25

Effect of Fuel Droplets During Early Stage of Flame Propagation

Aminuddin Saata,*

, Malcolm Lawesb, Mazlan Abdul Wahid

a,

Mohd Farid Muhamad Saida

a Faculty of Mechanical Engineering, Universiti Teknologi Malaysia, 81310 UTM Johor, Malaysia.

b School of Mechanical Engineering, University of Leeds, LS2 9JT, Leeds, United Kingdom.

*email: [email protected]

ABSTRACT

There are only few experimental data of a

fundamental nature that clearly demonstrate the

similarities and differences in flame propagation and

burning rates between single phase and two phase

combustion. Such data are essential in order to

establish a better understanding of the spray

combustion phenomena and associated processes. In

the present study, experimental investigations of

combustion of droplet and vapour mixtures under

quiescent condition have been conducted in a closed

combustion vessel. Droplet and vapour mixtures or

aerosol mixtures were generated by expansion of

iso-octane gaseous pre-mixture to produce a

homogeneously distributed suspension of fuel

droplets. The aerosol mixtures were ignited centrally

in the combustion vessel and the flame propagation

was recorded by high-speed schlieren photography.

Flame speed was obtained from the measured flame

radius-time data. The effect of fuel droplets in the

early stage of flame propagation was investigated by

comparing the flame structure and flame speed of

gaseous mixtures at identical conditions.

Comparisons between gaseous and aerosol flame

structure have shown quantitatively that the presence

of fuel droplets causes earlier onset of instabilities

and cellularity than for gaseous flames, particularly

at rich conditions. It is shown that the initial growth

of flame propagation in aerosol mixtures was

different than in gaseous mixtures. This difference

was shown to be a function of droplet size and

overall equivalence ratio. It is suggested that these

factors lead to vary the local equivalence ratio which

increases the initial burning rate of lean aerosols, but

decreases that of rich ones.

Keywords : droplet, aerosol, flame speed,

burning rate

1. INTRODUCTION

The combustion of droplets have been studied

extensively because of their relevance in many

practical combustion devices such as automotive

engines, gas turbines, power generation, boiler

and heating system. However, the multiplicity of

dependant parameters involved and the difficulty

in carrying out well-controlled studies to extract

the effects of individual parameters renders

practical spray combustion systems as some of

the most challenging environments to

investigate. There is theoretical and

experimental evidence [1-4] to suggest that

flame propagation through droplet and vapour

mixture, under certain circumstances, is higher

than that in a fully vaporised homogeneous

mixture. Although this may be advantageous in

giving more rapid burning, its effects on

emissions are uncertain. Conversely, it is a

serious disadvantage in the hazard context.

Thus, an understanding of the influence of the

presence of fuel droplets in flame propagation is

crucial for understanding the practical spray

combustion system.

Since the essential aspects of single-droplet

combustion become clear, investigations on

combustion of droplet and vapour mixtures are

necessary for a further approach to spray

combustion because the majority of droplets

burn as a group and interact with one another in

practical environments [5]. Experimental study

on the effect of fuel droplet on propagating

flame was performed by Hayashi and Kumagai

[1], Nomura et al. [2,6] and Lawes et al. [3].

Based on the principle of Wilson‘s cloud

chamber [7], they investigated flame

propagation in a stagnant pre-mixture in which

fuel droplets were monodispersed. Hayashi and

Kumagai [1] found the highest enhancement in

aerosol burning rate occurred at the slighly rich

mixture when the droplet diameter is 30 m.

Nomura et al. [2,6] investigated the influence of

fuel droplets under microgravity conditions and

they concluded that there are two regions in

which flame speed of aerosol mixtures exceeds

IMAT-UI 005



2012

26

that of premixed gases. Lawes et al. [3] observed

that aerosol flames became more unstable

compared to the gaseous flame, and hence had

faster burning rates. However, the enhancement

was not very significant at which the overall

equivalence ratio is indicated at slightly rich

mixture. In a numerical study, Polymeropoulos

[4] predicted a significant burning rate

enhancement for monodispersed aerosol fuel-air

mixtures comprising droplets in the range of 5-

15 m. He claimed that there is still lack of

experimental data within this range to support

his prediction.

In this paper, spherically expanding flames

following central ignition of globally

homogeneous combustible fuel mixtures at near

atmospheric pressures are employed to quantify

the differences in the flame structure and

burning rates of gaseous flames with aerosol

flames. Iso-octane-air aerosol mixtures with

droplet size ranges of up to 20 m, were

generated by expansion of the gaseous pre-

mixture to produce a homogeneously distributed

suspension of fuel droplets. The effect of fine

fuel droplets in the early stage of flame

propagation was investigated by comparing the

flame structure and flame speed of gaseous

mixtures with aerosol mixtures at several range

of mixture equivalence ratio.

2. EXPERIMENTAL APPARATUS AND

TECHNIQUE

The combustion vessel and auxiliary equipment for

the experimental work are shown schematically in

Figure 1. A full description of the system and aerosol

generation technique is presented in [8]. The

combustion vessel, which essentially resembled a

Wilson cloud chamber [7], was a cylindrical vessel of

305 mm diameter by 305 mm long. Optical access

windows of 150 mm diameter were provided on both

end plates for aerosol characterisation and

photography of flame propagation. Four fans, driven

by electric motors, adjacent to the wall of the vessel,

initially mixed the reactants. Two electrical heaters

were attached to the wall of the vessel to preheat the

vessel and mixture to the desired temperature.

Figure 1: Schematic of aerosol mixtures

generation and combustion

apparatus.

Iso-octane-air aerosol mixtures were prepared by a

condensation technique [7], used elsewhere in

combustion studies by [1-3], to generate well defined

and near mono-dispersed droplet suspensions. This

achieved by controlling the expansion of a gaseous

fuel-air mixture from the combustion vessel into the

expansion tank, which was pre-evacuated to less than

1 kPa. The expansion caused a reduction in mixture

pressure and temperature which took it into the

condensation regime and caused droplets to be

formed. The characteristics of the generated aerosol

were calibrated by in-situ measurements of the

temporal distribution of pressure, temperature,

droplet size and number without combustion, with

reference to the time from start of expansion.

Figure 2 shows a typical variation of pressure, P,

temperature, T, droplet size, D, and number density,

ND, of droplets with time from the start of expansion

for a stoichiometric iso-octane-air mixture expanded

at 200 kPa and 303K. The measurement of droplet

mean diameter, D10, was performed using Phase

Doppler Anemometer system. The estimation of

number density, ND was obtained from calculations

based on laser attenuation and droplet size

measurements using the Beer-Lambert Law

correlation given in [9]. The pressure of the mixture,

P, was set at nearly atmospheric condition for all

experiment. The temperature of all mixtures just

before ignition varies from 265 to 293 K. The overall

equivalence ratio, ov, is varied from lean mixture up

to rich mixture. The mean droplet size of aerosol

mixtures was varied up to 20 m by varying the

initial pressure before expansion and time of ignition.

As shown in Fig. 2, the measured temporal variation

of temperature, Tm, initially exhibited a polytropic

relationship, as shown by the dotted curve in Fig. 2,

in which the polytropix index, n, was found to be

1.35. At the start of droplet nucleation, approximately


2012

27

1.6 seconds after expansion start, the measured

temperature departed from that of the polytropic

expansion, in part due to the latent heat of

condensation. This was also evident by the increase

in ND and D10. It is shown that ND is remained nearly

constant during the condensation. During mixture

expansion, the overall equivalence ratio, ov,

remained constant, however, the gas and liquid phase

equivalence ratios, g and l, varied after nucleation

and during droplet growth. Also shown in Fig. 2 are

the standard deviations of the droplet mean diameter,

D. The low values of D indicate the near mono-

dispersed distribution of droplet size that results in

the combustion vessel. Since the expansion of

mixture took place over a period of several seconds

while combustion took place over less than 100 ms,

the far field values of D10 were assumed to be

constant during combustion.

Figure 2: Typical variation of several parameters

with time for iso-octane-air aerosol

mixtures expanded from initial

condition of 200 kPa and 303 K.

The droplet size varies with time during expansion;

hence the effect of droplet size is investigated by

varying the time of ignition after the start of

expansion. The aerosol mixture was ignited at the

centre of the combustion vessel by an electric spark

of about 400 mJ. The flame front was monitored and

recorded through the vessel windows by high speed

schlieren arrangement at a rate of 1000 frame per

second. The flame image was processed digitally to

obtain the flame radius using image processing

software. The laminar flame speed, Sn, was obtained

from the measured flame front radius against time, t,

by Sn = dr/dt. As the burning rates for aerosol flames

was obtained, measurement of burning rates of

gaseous flames at initial pressure and temperature

closer to those of aerosols could provide a more

accurate comparison between gaseous and aerosol

flames. This was achieved by igniting the mixture

during expansion but before the onset of

condensation.

3. RESULT AND DISCUSSION

Figure 3: Typical spherical flame development

of laminar iso-octane-air flames at

ov = 1.2 for gaseous and aerosol

mixtures.

Shown in Fig. 3 are the typical sequences of schlieren

images showing the growth of two expanding laminar

iso-octane - air flames at ov = 1.2. Figure 3a shows a

gaseous flame ignited at 105 kPa and 282 K while

Fig. 3b shows an aerosol flame with droplet diameter

of 14 µm, ignited at 105 kPa and 279 K. The circular

boundary (150 mm in diameter), represents the size

of the optical access windows. The black horizontal

object in each of images is the spark electrode holder.

Since there were negligible differences in initial

pressure and temperature, it is assumed that the

differences in the flame structure are entirely due to

the presence of droplets. The gaseous flame had a

smooth surface and was nearly spherical throughout

the whole period of observation. However, the flame

with droplets shows unstable surface with cracks and

cells that gradually increased throughout flame

growth. Such cellularity has been observed to

increase with increasing ov and with increasing


2012

28

droplet size. Hence, Fig. 3 demonstrates that flame

with fuel droplets are more unstable than those of

equivalent gaseous flames and droplets can cause the

onset cellularities that would not be present in

gaseous flames at similar conditions.

Figure 4:. Variation of flame speed with time for

gaseous iso-octane-air flame with

and without droplets at ov = 1.2.

Figure 4 shows the variation of flame speed with time

for gaseous and aerosol mixtures at an equivalence

ratio of 1.2 at near atmospheric conditions. For

clarity, each plot shows average values from three

explosions with associated error bars. Initially, after

spark ignition, gaseous flame propagates faster than

aerosol flame within the first 8 ms before attaining an

approximately constant value of about 2.5 m/s at later

stage. The aerosol flame developed slightly slower

than gaseous flame, but attained higher, and still

accelerates at later stage. This acceleration was

associated with instabilities, as shown by unstable

surface in Fig. 3, which triggered by the droplets

presence and further enhance the rate of flame

propagation. The mechanism behind unstable flame

behaviour at the later stage is probably related to the

heat loss from the flame in vaporizing the droplets,

accompanied by the local rapid expansion taking

place in the process. This comparison clearly

demonstrates the differences of flame propagation

rate due to the presence of fuel droplets.

It is clear from Fig. 4, that fuel droplets causes

different burning rate either at the early stage or later

stage of flame development. Further investigation is

focused on the effect of droplets in the early stage of

flame development.

(a)

(b)

Figure 5: Effect of droplets in the early stage of

flame propagation in lean mixtures,

ov=0.9. (a) Schlieren images up to 3

ms after ignition, (b) variation of flame

speed with time. Error bars represent

scatter from up to three explosions.

Figure 5(a) shows sequences of schlieren images of

gaseous and aerosol flame kernels for lean mixtures

at an equivalence ratio of 0.9 during first 3 ms after

ignition, while in Fig. 5(b) is the corresponding flame

speed variation with time. At least a minimum of

three experiments were conducted at each condition

and the error band presented in Fig. 5(b) indicates the

variation in flame speed between the experiments.

The circle symbols represent the size of the flame at

3 ms after ignition as shown by the schlieren images

in Fig. 5(a). All the flames were ignited using the

same spark electrode at identical ignition energy. It is

shown in Fig. 5 that initial size of flame kernel and

the rate of flame development is a function of droplet

size, with the gaseous flame having the slowest initial

rate of development and the aerosol flame for a

droplet diameter of 20 m showing the fastest

development. However, this trend reversed for the

rich mixtures.


2012

29

(a)

(b)

Figure 6: Effect of droplets in the early stage of flame

propagation in rich mixtures, ov=1.4. (a)

Schlieren images for up to 3 ms after

ignition, (b) variation of flame speed with

radius. Error bars represent scatter from up

to three explosions.

Figure 6(a) shows sequences of schlieren images of

gaseous and aerosol flame kernels for rich mixtures

at an equivalence ratio of 1.4 during first 3 ms after

ignition, while in Fig. 6(b) is the corresponding flame

speed variation with radius. In contrast to the trend

presented for leaner mixtures in Fig. 5, it is seen in

Fig. 6 that for rich mixtures, the initial flame

development is fastest in the gaseous flame and

slowest in the aerosol flame with 20 m droplet

diameter. It was also observed that as the droplet size

increased, the ignitibility of aerosol mixtures

improved and the probability of ignition and

subsequent flame propagation in aerosol mixtures

was higher than a gaseous mixture at the same overall

equivalence ratio. This observation was particularly

significant for leaner mixtures. However, a

probabilistic study of ignitability of aerosol mixtures

did not form a part of the present study.

Figures 5 and 6 clearly show that the influence of

fuel droplets in the early stages of flame propagation

is significant. In leaner mixtures, increase in the

droplet diameter causes an increase in the flame

speed. However, in richer mixtures, increase in

droplet diameter causes a decrease in the flame

speed. This reversal in trend during initial flame

development can be explained due to droplet inertia

mechanism [10, 11]. For initially quiescent condition,

gaseous fuel-air mixture is nearly stagnant as well as

the fuel droplets movement. The existence of fuel

droplets in the mixture enriches the fuel vapour

density near the reaction zone, leading to an increase

in the local equivalence ratio. In the case of lean

mixtures, this enrichment in fuel vapour density

shifts the local equivalence ratio closer to the

stoichiometric fuel-air ratio and hence the aerosol

flame burns faster than the equivalent gaseous flame.

For richer mixtures, the same mechanism causes the

local equivalence ratio to become even richer and

affects an attenuation in the growth rates which

implies that gaseous flame burns faster than the

aerosol flame, as evidenced by Fig. 6. Furthermore,

larger the droplet size, higher the inertia and hence

stronger the influence. This explains why this effect

gets more pronounced with an increase in the droplet

size. A similar relationship has been demonstrated

experimentally by Nomura [6] in terms of the

formulation of droplet slip velocity in microgravity

combustion.

4. CONCLUSION

The comparison between flame with and without

droplets under laminar conditions have been

experimentally studied in centrally ignited flames in

which aerosol of fuel droplets was suspended in a

mixture of quiescent air and fuel vapour. Inspection

of schlieren images revealed that aerosol flames were,

in general, more unstable than gaseous flames and

this instability increases as the droplet size increases.

The presence of fuel droplets clearly influences

instabilities by causing earlier onset and more rapid

development of cellularity than for gaseous flames.

During the early stages of flame propagation, droplets

are relatively stagnant with respect to the flame. This

inertia of fuel droplets leads to a local enrichment in

the mixture equivalence ratio. Flame propagation

rates in the early stages were found to be consistent

with the altered equivalence ratio in the sense that for

leaner mixtures, droplets enhanced the burning rate,

while for richer mixtures, droplets slowed down the

burning rate.


2012

30

ACKNOWLEDGMENT

The authors acknowledge the financial support from

Universiti Teknologi Malaysia under a Grant

Research No. R.J130000.7724.4P044. The use of

apparatus in the Thermodynamic Laboratory of

Leeds University throughout the study is gratefully

acknowledged.

REFERENCES

[1] Hayashi, S., Kumagai, S., ―Flame propagation in

fuel droplet-vapor-air mixtures‖, Proc. Comb.

Inst. 15:445-452 (1974).

[2] Nomura, H., Koyama, M., Miyamoto, H., Ujiie,

Y., Sato, J., Kono, M., Yoda, S., ―Microgravity

experiments of flame propagation in ethanol

droplet-vapor-air mixture‖, Proc. Comb. Inst.

28:999-1005 (2000).

[3] Lawes, M., Lee, Y., Marquez, N., ―Comparison

of iso-octane burning rates between single-phase

and two-phase combustion for small droplets‖,

Combustion and Flame 144:515-525 (2006).

[4] Polymeropoulos, C.E., Flame propagation in

aerosols of fuel droplets, fuel vapour and air‖,

Combution Science and Technology 40:217-232

(1984).

[5] Annamalai, K, Ryan, W, ―Interactive processes

in gasification and combustion: Part 1 Liquid

drop array and clouds‖, Progress of Energy

Combustion Science, 18:221-295 (1992).

[6] Nomura, H., Kawasumi, I., Ujiie, Y., Sato, J.,

―Effects of pressure on flame propagation in a

premixture containing fine fuel droplets‖, Proc.

Comb. Inst. 31:2133-2140 (2007).

[7] Wilson, C. T. R., ―Condensation of water vapour

in the presence of dust-free air and other gases‖,

Proc. of the Royal Society of London 189:265-

307 (1897).

[8] Saat, A., ―Fundamental studies of combustion of

droplet and vapour mixtures, PhD Thesis, School

of Mechanical Engineering, University of Leeds

(2010).

[9] Bachalo, W.D., Rosa, A. B., Sankar, S.V.,

―Diagnostic for fuel spray characterisation, in

Combustion measurements‖, (Ed. Chigier N.),

Hemisphere, p.229-278 (1991).

[10] Atzler, F., ―Fundamental studies of aerosol

combustion‖, PhD Thesis, School of Mechanical

Engineering, University of Leeds (1999).

[11] Atzler, F., Demoulin, F., Lawes, M., lee, Y.,

―Oscillations in the flame speed of globally

homogeneous two phase mixtures‖, 18th

International Colloquium on the Dynamics of

Explosions and Reactive Systems, paper 83

(2001).


2012

31

The Use of Mechanical Ventilation System in an Electric Car

Intan Sabariah Sabria, Haslinda Mohamed Kamar

b, Nazri Kamsah

c, Md Nor Musa

d

aFaculty of Mechanical Engineering,

Universiti Teknologi Malaysia, 81310 Skudai, Johor, Malaysia


bFaculty of Mechanical Engineering,



cFaculty of Mechanical Engineering,



dFaculty of Mechanical Engineering,



ABSTRACT

The heart of an electric car is its rechargeable

battery pack, which supplies the electric motor with

the energy to move the vehicle. When parked in the

sun, the soak air temperature inside a passenger

compartment can rise up to 60°C. Reducing the air

temperature inside the passenger compartment is

very important not only for passenger comfort but

also to reduce the power consumed by the air-

conditioning (AC) system. This leads to saving in

the battery power. This paper presents the use of

mechanical ventilator to lowering the air

temperature inside the passenger compartment

during parking in a sunny day condition. The

commercial software FLUENT 6.3 is used to

simulate three-dimensional (3-D) air temperature

distributions and air flow field inside the passenger

compartment. The simulated result without

mechanical ventilator is compared with the

experimental data. They show a good agreement

with average deviation of 1.8°C in general. This

study found that the number of mechanical

ventilator has a weak influence on the air

temperature inside the passenger compartment

when they are installed on the rear dashboard.

Keywords : Computational Fluid Dynamics (CFD)

analysis, mechanical ventilation

system, passenger compartment,

electric car, air temperature field, air

flow field.

1. INTRODUCTION

The heart of an electric car is its battery [1]. The

battery of an electric car runs everything such as

operate the lighting, accessories and AC system.

The AC system consumes a lot of energy from the

battery to cool down the air temperature inside the

passenger compartment. The battery power

consumption can be reduced by saving the energy

required by the AC system. During the sunny day,

about 80% of the air temperature inside the

passenger compartment rises during the first 30

minutes

[2]. A car parked in the sun with its windows

closed experiences a greenhouse effect, where the

passenger compartment becomes extremely hot [3].

The greenhouse effect can arise the air temperature

inside the passenger compartment up to 60 - 70°C

[4]. The windows glazing allows about 50 - 70% of

thermal energy entering the passenger compartment

during parking condition [3]. The air and materials

affected by solar radiation reach considerable

temperatures such as about 70°C for the dashboard

when a car parked facing the sun [5]. This will

make a driver feel very uncomfortable when he

enters the car. The AC system is unable to cool

down the air temperature inside the passenger

compartment within a short period of time.

Reducing the air temperature inside the passenger

compartment is very significant not only for

passenger comfort but also to reduce the power

consumed by the AC system. This could lead to

saving the battery power consumption of the

electric car for a better driving range [6]. The air

temperature inside the passenger compartment can

be reduced in many ways such as solar-reflective

glazing, solar-reflective coatings and parked car

ventilation [3]. Reducing the size and the glass

transmissivity of the window are able to decrease

the air temperature inside the passenger

compartment [7]. Solar-reflective glazing reflects

the incidence solar radiation and helps in reducing

the air temperature inside the passenger

compartment. Study found that the solar reflective

glazing reduces the air temperature inside the

passenger compartment by 2.7°C as well as

lowered the instrumental panel temperatures by

IMAT-UI 006






2012

32

7.6°C. This leads to reduce about 11% of AC

system power consumption [8]. The use of solar-

reflective glass in all locations reduced the average

air temperature by 34% of the maximum possible

[3]. The use of solar reflective coatings for opaque

surfaces can also decrease the air temperature

inside the passenger compartment when a car is

parked in the sun [3]. Study found that increasing

the solar reflectance (ρ) of the vehicle‘s shell about

0.5 reduces the air temperature inside the passenger

compartment by about 5 - 6°C [9]. Increasing of

each 0.1 solar reflectance of the shell decrease the

air temperature inside the passenger compartment

by about 1°C [10]. Solar energy that enters a

vehicle heats the interior mass as well as the air of

the passenger compartment. Venting the warm air

and pulling in cooler ambient air by using car

ventilation system can reduce the soak air

temperature inside the passenger compartment [3].

The use of ventilation system on a parked car is

able to reduce the average air temperature by 5.6°C

and the seat temperatures by 5 - 6°C [11].

Study found that by using parked car ventilation

system the average air temperature inside the

passenger compartment is reduced by about 9.3%

[12].

This research aims to quantify the effects of

mechanical ventilator on the soak air temperature

inside the passenger compartment of the electric

car by using a numerical simulation technique.

Three dimension steady state simulation results

inside the passenger compartment are obtained by

using the commercial software FLUENT 6.3. The

numerical simulation results will be compared with

experimental data for the validation purpose. In

addition, the effects of mechanical ventilator

positions inside the passenger compartment are

investigated in terms of air temperature

distributions as well as air flow field.

2. NUMERICAL SIMULATION MODEL,

METHODS AND VALIDATION

2.1 Computational domain and boundary

conditions

The entire inner space of the vehicle is considered

as consists the computational domain and the inner

surfaces of the passenger compartment are defined

as the boundaries of the domain. The numerical

simulation is in a steady state operating condition.

Figure 1 indicates the computational domain based

on the compartment construction of Saga BLM

model. Rear passenger compartment model is

located on the origin position. The passenger

compartment was modelled by using dimensions of

a real car of Proton Saga BLM model where length,

width and height are 2523.30 mm, 1080 mm

and 1240 mm respectively. There are two

separated seats in the front passenger compartment

and a long bench at the rear passenger

compartment. There are four circles representing

the air inlets on the front dashboard and a

mechanical ventilator is located on the center of

rear dashboard. The four inlets are situated

symmetrically about the z-coordinate as shown in

Figure 1.

Both the inlets and mechanical ventilator are

defined as the air-flow inlets and outlet. The

mechanical ventilator is treated as the outlet. The

air velocity at the ventilator is 2.84 m/s and the

direction of the resultant velocity vector was

normal to the surface boundary. Since the air

velocity of the outlet is defined, the air velocity of

the inlets can be calculated by using the

conservation of mass principle. The air velocity at

the four inlets are fixed at 0.83 m/s and turbulent

intensities are 10% [7,15]. The boundary condition

at the bottom surface is set up as a temperature at

300 K. The convection and radiation boundary

conditions are applied to glass and roof surfaces.

Convective heat transfer coefficient for glass

surfaces are fixed at 15 W/m2°C and the thickness

of the glazing is set as 5 mm. Roof is assumed as

opaque wall and the convective heat transfer

coefficient is set as 15 W/m2°C with 12 mm

thickness [14]. The passenger compartment is

assumed to be well sealed, therefore no additional

air-flow inlet as well as outlet should be considered

[7]. No-slip solid wall is applied to the whole car

surfaces. The air flow in the passenger

compartment is assumed to be incompressible with

constant thermophysical properties [7,13].

Figure 1 : Geometric model of Proton Saga

BLM model


2012

33

2.2 Grid generation and numerical methods

Meshing of the computational domain of the model

consists of 955437 tetrahedral cells. The tetrahedral

cell are used due to the complexity of the model

[7,15].

The CFD software FLUENT 6.3 is used for the

simulation to solve continuum, energy and

transport equations numerically with natural

convection effects. The convection term is

discretized by the second-order upwind difference

[7,14,15]. The energy governing equations are

discretized by the finite volume method [7,17]. The

Semi-Implicit Method for Pressure-Linked

Equations (SIMPLE) algorithm [4,7,13-17] is used

for handling the coupling between pressure and

velocity. The standard turbulence model is

adopted in conjunction with the standard wall

function for the near wall region treatment [4,7,13-

17].

2.3 Validation of the numerical model

The numerical model is validated by comparing the

result of simulation with experimental data. During

the experiment, a metallic white color of Proton

Saga BLM model was parked in an open space

under the sunlight from 12 pm – 3 pm. An interior

air temperature and both internal and external

surface temperatures were monitored. The

experiment was conducted under the conditions

where the ambient temperature is 35°C and

incidence solar radiation is 1 kW/m2. The steady

state 3-D simulation result is shown in Figure 2,

where the experimental data are also presented for

comparison. The predicted and measured air

temperature values inside the passenger

compartment from 12 pm – 3 pm are listed in Table

1.

Figure 2 indicates the average air temperature

differences between the predicted data and

measured data are about 1.8°C. The measured data

were affected by many factors such as solar

radiation, material properties of the passenger

compartment and heat losses from the outer

surfaces of the passenger compartment.

Table 1: Predicted and measured air temperature values.

Figure 2 : Predicted air temperature and the

comparison with measured data

inside the passenger compartment.


3.1 Ventilation System in the Passenger

Compartment

The inlet air temperature is fixed at 36°C. The

ventilator is located on the center of rear dashboard

where the velocity of the ventilator is 2.84 m/s.

Figure 3 shows the steady state of air temperature

distributions inside the passenger compartment

with mechanical ventilation system installed in the

car. At the beginning of the simulation, the average


is about 48°C. With the existing of mechanical

ventilator in the passenger compartment, the steady

state average air temperature inside the passenger

compartment is reduced to 44°C. From the figure,

it can be seen that the distributions of the air

temperature inside the passenger compartment is

fairly uniform.

Figure 4 shows the air velocity vectors on a vertical

middle of z-plane. It can be observed that the air

flow field inside the passenger compartment is not

uniform. The air flow field concentrates more at

rear compartment rather than at the front. The

mechanical ventilator may not be at the best

position. The mechanical ventilator should be

placed at different position to obtain more uniform

air flow field inside the passenger compartment.


2012

34

Figure 3 : Steady state of air temperature

distributions at vertical middle of z-

plane (z = 0.54m)

Figure 4 : Velocity vectors distributions at

vertical middle of z-plane (z = 0.54

m).

3.2 Two mechanical ventilators on the rear

dashboard

The steady states air temperature distributions and

velocity vectors inside the passenger compartment

with the two mechanical ventilators are shown in

Figure 5 and 6. The ventilators are located

symmetrically 0.27 m away from the sidewalls on

the rear dashboard respectively. The air velocity at

the ventilators are fixed as 2.84 m/s. The average


is 43°C. It can be observed from figure 5 that the

air temperature distribution inside the passenger

compartment is uniform. The uses of two

mechanical ventilators are able to reduce the air

temperature inside the passenger compartment by

4°C.

Figure 6 shows the air velocity vectors of vertical

plane at z = 0.27 m from the sidewall. It can be

observed that the air flow field inside the passenger

compartment is uniform. The results suggest that

the number of mechanical ventilator has strong

influence on the air temperature inside the

passenger compartment.

Figure 5: Steady state of air temperature

distributions at vertical middle of z-

plane (z = 0.54 m).

Air velocity at v =1.11 m/s


2012

35

Figure 6 : Velocity vectors distributions at

vertical plane, z = 0.27 m from

sidewall.

4. CONCLUSIONS

The use of mechanical ventilator is identified as an

efficient way to lowering down the air temperature

inside the passenger compartment. Introducing the

mechanical ventilator located at the middle of rear

dashboard reduces the air temperature inside the

passenger compartment by 3°C. Introducing two

mechanical ventilators located 0.27 m away from

the sidewall of the rear dashboard reduces the air

temperature by 4°C. The simulation results suggest

that the number of mechanical ventilator has a

weak influence on the air temperature inside the

passenger compartment when they are installed on

the rear dashboard.

ACKNOWLEDGMENT

The authors gratefully acknowledge the Universiti

Teknologi Malaysia (UTM) for funding this study

under the project number 00G41 and also UTM-

PROTON Future Drive Laboratory for giving a

space to conduct the experimental work.

REFERENCES

[1] Christoper Lampton, ―How Electric Car

Batteries Work‖,

http://auto.howstuffworks.com/fuel

efficiency/vehicles/electric- car-battery.htm

(July, 2012).

[2] Catherine McLaren, Jan Null & James Quinn

(2005), ―Heat Stress From Vehicles:

Moderate Ambient Temperatures Cause

Significant Temperature Rise in Enclosed

Vehicles‖, Pediatrics.116;e109.

[3] Desikan Bharathan, Larry Chaney, Robert B.

Farrington, Jason Lustbader, Matthew Keyser

and John P. Rugh (2007), ―Fuel Use

Reduction Project Close-Out Report‖,

Technical Report National Renewable Energy

Laboratory (NREL)/CP-540-41155. June

2007.

[4] K.David Huang, Sheng-Chung Tzeng, Wei-

Ping Ma and Ming-Fung Wu, ―Intelligent

Solar-Powered Automobile-Ventilation

System‖, Applied Energy 80 (2005) 141-154.

[5] Mezrhab, A. and Bouzidi, M., ―Computation

of thermal comfort inside a passenger car

compartment‖, Applied Thermal Engineering,

26 (2006), pp. 1697–1704.

[6] Nebojsa I., Jaksic & Cem Salahifar (2003),

―A Feasibility Study Of Electrochromic

Windows In Vehicles‖, Solar Energy

Materials and Solar Cells. 79: 409-423.

[7] Huajun Zhang, Lan Dai, Guoquan Xu, Yong

Li, Wei Chen, Wen-Quan Tao, ―Numerical Of

Air-flow and Temperature Fields Inside A

Passenger Compartment For Improving

Thermal Comfort and Saving Energy. Part I:

Test/Numerical Model and Validation‖,

Applied Thermal Engineering 29 (2009)

2022-2027.

[8] John P., Rugh, Terry J., Hendricks & Kwaku

Koram (2001), ―Effect of Solar Reflective

Glazing on Ford Explorer Climate Control,

Fuel Economy, and Emissions‖, Society of

Automotive Engineers (SAE) Technical Paper

Series. Paper No. 2001-01-3077.

[9] Ronnen Levinson, Heng Pan, George Ban-

Weiss, Pablo Rosado, Riccardo Paolini &

Hashem Akbari (2011), ―Potential Benefits Of

Solar Reflective Car Shells: Cooler Cabins,

Fuel Savings And Emission Reductions‖,

Applied Energy. 88: 4343-4357.

[10] Hoke, P., B. & Greiner, C. (2005), ―Vehicle

Paint Radiation Properties And Effect On

Vehicle Soak Temperature, Climate Control

System Load, And Fuel Economy‖, Society of

Automotive Engineers (SAE) Technical Paper

Series. Paper No. 2005-01-1880.

[11] John P., Rugh and Robert B., Farrington

(2008), ―Vehicle Ancillary Load Reduction

Project Close-Out Report‖, Technical Report

http://auto.howstuffworks.com/fuel


2012

36

National Renewable Energy Laboratory

(NREL)/TP-540-42454. January 2008.

[12] R.Saidur, H.H Masjuki and M.

Hasanuzzaman, ―Performance Of An

Improved Solar Car Ventilator‖, International

Journal of Mechanical and Materials

Engineering (IJMME), Vol. 4 (2009), No. 1,

24-34.

[13] Dengchun Zhang, Peifen Weng, ―Numerical

Simulation and Experiment Research of Air

Organization in Air-Conditioned Passenger

Car‖, Building Simulation 2007.

[14] Sevilgen, G. and Kilic, M., ―Investigation of

Transient Cooling of an Automobile Cabin

with a Virtual Manikin under Solar

Radiation‖

[15] Sevilgen, G. and Kilic, M., ―Transient

numerical analysis of airflow and heat transfer

in a vehicle cabin during heating period‖,

International Journal of Vehicle Design, 52

(2010), 1-4, pp. 144-159.

[16] A.Aroussi, A.Hassan, Y.S Morsi, ―Numerical

Simulation Of The Airflow Over And Heat

Transfer Through A Vehicle Windshield

Defrosting And Demisting System‖, Heat and

Mass Transfer 39 (2003) 401-405.

[17] Jalal M. Jalil and Haider Qassim Alwan,

―CFD Simulation for a Road Vehicle Cabin‖,

Engineering Science, Vol. 18 No. 2, pp: 123-

142 (2007 A.D. /1428 A.H.)


2012

37

Retrofitting R-22 Split Type Air Conditioning With

Hydrocarbon (HCR-22) Refrigerant

Henry Nasutiona,b,c

, Zulkarnain Abdul Latiffa,b

, Azhar Abdul Aziza,b

, Mohd Rozi Mohd

Peranga


81310 Skudai, Johor, Malaysia, Phone: +60 75535447, Fax: +60 75535811 bFaculty of Mechanical Engineering, Universiti Teknologi Malaysia

81310 Skudai, Johor, Malaysia, Phone: +60 7 5534575, Fax: +60 7 5566159

email: [email protected]

cDepartment of Mechanical Engineering, Bung Hatta University

email: [email protected]

25132 Padang, Sumatera Barat, Indonesia, Phone: +62 751 7054657, Fax: +62 751 7051341

ABSTRACT

An experimental study to evaluate the energy

consumption of a split type air conditioning is

presented. The compressor works with the fluids R-

22 and HCR-22a and has been tested varying the

internal heat load 0, 500, 700 and 1000 W. The

measurements taken during the one hour

experimental periods at 10-minutes interval times

for temperature setpoint of 20oC. The performance

data considered where the evaporator cooling load,

the condenser heat rejection, the electrical energy

consumption, the refrigeration system temperatures,

and the room temperature. And hence the

Coefficient of Performance (COP) could be

determined. The final results of this study show an

overall better energy consumption of the HFC-22a

compared with the R-22.

Keywords: split air conditioning, R-22 and HCR-

22a, energy saving

1. INTRODUCTION

Hydrochlorofluorocarbons-22 (HCFC22) is the most

commonly used on the split-type air conditioner as a

refrigerant. However, because of issues of

environmental damage, causing the experts try to

use environmentally friendly refrigerants, namely

hydrocarbons. Several studies have shown that the

use of hydrocarbon refrigerants can improve the

performance of air-conditioning systems [1-5].

The negative impact of HCFCs is ozone layer

destruction and global warming. The HCFCs such as

R-22 and R-123 have the impact of the greenhouse

effect or global warming potential (GWP) that is

still relatively high [6]. The positive properties of

hydrocarbon refrigerant are zero ozone depletion

potential (ODP), very low global warming potential,

high miscibility with mineral oil, and non-toxicity.

The main negative of refrigerant hydrocarbons is

their flammability [7]. Although it has been known

that hydrocarbon refrigerants to replace CFC

refrigerants, HCFC, and HFC, but due to economic

reasons and flammability properties, which cause

the development of hydrocarbon refrigerant is

relatively stagnant. Flammability of hydrocarbon

refrigerants becomes which most prominent reason

to weaken the use of this refrigerant. Vehicle fuel is

much more flammable than hydrocarbon

refrigerants, but still safe to be used.

The Montreal Protocol requires the US (United State

of America) to reduce its consumption of HCFCs

by 75% below the US baseline. Allowance holders

may only produce or import HCFC22 to service

existing equipment. Virgin R-22 may not be used in

new equipment. As a result, heating, ventilation and

air-conditioning (HVAC) system manufacturers may

not produce new air conditioners and heat pumps

containing R22. January 1, 2015: The Montreal

Protocol requires the U.S. to reduce its consumption

of HCFCs by 90% below the U.S. baseline. January

1, 2020: The Montreal Protocol requires the U.S. to

reduce its consumption of HCFCs by 99.5% below

the U.S. baseline. A refrigerant that has been

recovered and recycled/reclaimed will be allowed

beyond 2020 to service existing systems, but

chemical manufacturers will no longer be able to

produce R22 to service existing air conditioners and

heat pumps [8].

Hydrocarbons are new findings for refrigerants by

the experts to replace the current HCFC22 as the

refrigerants for air conditioning system. Many

investigators have used a mixture of hydrocarbons

to improve the performance of air-conditioning

system. For example: Wongwises et al. [9] using a

mixture of hydrocarbons (propane-R290, butane-

R600 and isobutene-R600a). Wongwises and

Chimres [10] using a mixture of propane, butane,

and isobutene to replace HFC134a in freezer

applications. The results showed that a mixture of

propane (60%) and butane (40%) as the most

appropriate some other refrigerant. Hammad and

Alsaad [11] using LPG (24.4% propane, butane,

56.4%, and 17.2% isobutane) to replace CFC12

IMAT-UI 007



2012

38

refrigerant for refrigerator. Jung et al. [12] replace

CFC12 with a mixture of propane/isobutane

(R290/R600a) for the refrigerator. Han et al. [13]

study of azeotropic non R32/R125/R161 of the

vapor compression system applications, the results

showed that the cooling capacity and COP are better

than R407c. Park et al. [14] using a mixture of

propylene, propane, HFC152a, and dimethylether as

an alternative refrigerant to replace HCFC22

refrigerant in cooling systems and heat pump

split. The results show that the COP of the mixture

is 5.7% higher than HCFC22. Mani and Selladurai

[5] using the mixture as a replacement refrigerant

CFC12 R290/R600a and HFC134a vapor

compression refrigeration system. According to

experimental results, R290/R600a refrigerant

cooling capacity 19.9% to 50.1% higher than R12

and 28.6% to 87.2% compared to

R134a. R290/R600a mixture COP increased by 3.9-

25.1% compared to R12. Refrigerant R134a has a

slightly lower coefficient of performance of the R12.

With respect to these opportunities, current research

is focused on the temperature distribution, cooling

capacity and coefficient of performance, COP at

various internal heat loads using experimental result.

The experiment is also conducted using two

different refrigerants that are R-22 and HCR-22 to

evaluate the energy consumption of the current split

type air conditioning system.

2. SYSTEM DESCRIPTION AND

PROCEDURE

A schematic diagram of the experimental apparatus

is shown in Figure 1. The air-conditioning was

originally to work with refrigerant R-22 and with

compressor capacity, is 1860 W. Flow meter was

installed on the system to measure the flow of

refrigerant. The objective of the research was to

compare the refrigeration performance of different

refrigerants in terms of a coefficient of performance

(COP), cooling capacity, and power consumption by

the compressor.

The temperatures and pressures of the refrigerant are

measured at various locations in the experimental

set-up as shown in Figure 1. The refrigerant

temperature was measured by type T thermocouple

with an accuracy of ±0.1oC. The pressure was

measured through a tap with a small hole drilled into

the tube in which the refrigerant flows. The flow of

the refrigerants was measured by flow meter with an

accuracy of ±1 gr/s. The compressor energy

consumption was measured by current and voltage

of the electricity in the system. The accuracy of the

current meter is ±1%, and voltage meter is ±1.5%.

The refrigerants R-22 and HCR-22 were charged

after the system have been evacuated by a vacuum

pump. Drop-in experiments were carried out without

any modification on the system. The experiment on

the air-conditioning system was started with

HCFC22. Cooling capacity was obtained by varying

the load on the output of the evaporator. Data were

measured every 10 minutes during one hour. A

commercial refrigeration company prepared the

refrigerant mixtures. Composition of refrigerant

mixtures could not be presented in this research

because of company confidentiality. The data shown

in this study were the result of a mixture of the best.

In collecting data, the thermostat was set to the

maximum cool position. Data was collected on

when the conditions are considered stable, after the

air-conditioning was running 15 minutes for each

test condition. Thermodynamic properties of the

refrigerants were taken from the REFPROP database

[15]. The room setpoint temperatures during the

experiments were 20oC. In each experiment, the AC

system responded to the actual cooling load that

prevailed during the experimental period.

Figure 1: A schematic diagram of the

experimental set-up.

Caption:

1. Inlet compressor temperature, T1 (⁰C).

2. Exit compressor temperature, T2 (⁰C).

3. Exit condenser temperature, T3 (⁰C).

4. Inlet TEV temperature, T4 (⁰C).

5. Inlet evaporator temperature, T5 (⁰C).

6. Supply air temperature, T6 (⁰C).

7. Cabin temperature, T7 (⁰C).

8. Ambient temperature, T8 (⁰C).

9. Inlet compressor pressure, P1 (MPa).

10. Exit compressor pressure, P2 (MPa).

11. Exit condenser pressure, P3 (MPa).

12. Inlet evaporator pressure, P4 (MPa).

To obtain the differences that stand out about the

performance of the air conditioning due to drop-in,

analysis of several parameters is conducted. These

parameters are a coefficient of performance, cooling

capacity of the evaporator, and power consumption

of the compressor. The coefficient of performance

of refrigeration machine is the ratio of the energy

extracted by the evaporator (refrigerating effect) to

the energy supplied to the compressor. The


2012

39

coefficient of performance was calculated with

equation (1).

)(

)(

12

15

hh

hh

W

QCOP

c

e

(1)

where h1, h2 (kJ/kg) are the enthalpy at the

compressor inlet and outlet respectively, h5 (kJ/kg)

is the enthalpy at the evaporator inlet and outlet

respectively, Qe (kJ/kg) is the cooling capacity of

the evaporator, and Wc (kJ/kg) is the compression

work.

The energy saving calculated is expressed in terms

of saving in percentage unit, based on the difference

between energy consumed using R-22 and energy

consumed using HCR-22. The energy consumption

is calculated by multiplying the power consumption

of the compressor by the actual operating hours. The

equations are given as:

100Energy) 22-(R

Energy) 22-(HCR - Energy) 22-(R SavingEnergy

(2)


3.1 Compressor Pressure

Figure 2 shows the compressor absolute pressure of

R-22 compared with HCR-22 refrigerants at internal

heat loads 0 - 1000 W. The graph illustrates the inlet

(Pi) and outlet (Po) absolute pressure of the

compressor. The result shows that the HCR-22

indicates the lower absolute pressure at the inlet and

outlet of the compressor compared with R-22.

Figure 2: Compressor absolute pressure of R-22

compared with HCR-22.

3.2 Mass Flow Rate

Figure 3 exhibits the mass flow rate against internal

heat load variation. The mass flow rate of the HCR-

22 refrigeration system is constantly lower

compared to the R-22. The result shows that the

HCR-22 refrigerant is higher than the R-22, the

average difference is 20%.

Figure 3: Mass flow rate of R-22 compared

with HCR-22.

3.3 Room Temperature

Figures 4 shows the relationship between room

temperature and time at various internal heat loads

with R-22 and HCR-22. For conditioned room at a

temperature of 20oC at internal heat load's variation

of 0, 500, 700, 1000 Watts obtained that the room-

temperature distribution at 16-19oC and 18-20

oC for

refrigerant R-22 and HCR-22 respectively. HCR-22

is superior when compared with HCFC22 due to

refrigerant mass flow rate using HCR-22 a lot more

because of the refrigerant density lower than R-22,

so that the ability to cool the room is faster.

Figure 4: Room temperature of R-22 compared

with HCR-22.

3.4 COP

Having obtained the values of enthalpy, the COP

was obtained as in Figure 5. It shows that the COP

for R-22 was lower than that of HCR-22. At the

time of the beginning data collection, this was 15

minutes after the air conditioner was operated, the

COP of R-22 lower than that of HCR-22. The COPs

increase of about 2.84% to 11.98%. The increase of

this COP was lower than that of the study conducted

by Mani and Selladurai [5], i.e. 11.8-17.6%, where

the evaporator temperature was about -8oC, for

freezer purpose. Besides, the drop-in was performed

on a refrigeration system with refrigerant CFC12


2012

40

replaced with hydrocarbon mixture, propane (R290)

and isobutene (R600a). The differences of

evaporator temperature and refrigerants presumably

is caused different the results.

Figure 5: COP of R-22 compared with HCR-

22. 3.5 Cooling Capacity

Increasing COP can be caused by increasing in

cooling capacity at the evaporator and/or by

decreasing the power consumption of the

compressor. To find a more dominant effect,

whether the addition of cooling capacity or decrease

power consumption in the compressor, one can see

it in Figures 6. The result shows that the cooling

capacity of HCR-22 was higher than that of R-22.

Due to the fact that the COP of R-22 lower than that

of HCR-22.

Figure 6: Cooling capacity of R-22 compared

with HCR-22.

3.6 Electrical Energy Consumption and Energy

Saving

Figure 7 shows the relationship between compressor

power consumption and internal heat load. Based on

the data described in the previous section, it was

indicated that, a drop-in from HCFC22 to the HC22

may reduce power consumption of the compressor

and increasing of cooling capacity. However, in

general it can be said that the drop-in from HCHC22

to the HC22 may cause saving of power

consumption. According to of the COP, the average

saving was 9%, and according to of the compressor

power consumption, savings were 15.14%. The

Figure 8 explains the occurrence of saving on the

compressor power consumption significantly.

Figure 7: Compressor power of R-22 compared

with HCR-22.

Figure 8: Energy saving of R-22 compared with

HCR-22.

4. CONCLUSION

Replacement of the R-22 with a HCR-22 as the

working fluid refrigerant of the split-type air

conditioning has been investigated to show the

performance of the system. This indicated that with

hydrocarbon mixtures can further improve the COP

and energy saving. During the experimental test,

HCR-22 mixtures were found to be safe. However,

care should be taken when using R-290/R-600/R-

600a mixture in an air-conditioning system.

Based on this study, the following conclusions were

drawn.

1. The absolute pressure of the HCR-22 at the inlet

and outlet of the compressor is always lower

compared to the R-22. The average difference of

the inlet and outlet are 5.77% and 18.08%

respectively.

2. The mass flow rate of the HCR-22 claims the

average difference is 20% compared to the R-22.

The result shows that the HCR-22 is lighter and

lower in specific gravity than the R-22. This is

the positive improvement for the HCR-22

refrigeration system by reducing the mass flow

of the system as this is relevant to the reduction


2012

41

of the inlet and outlet absolute pressure of the

compressor.

3. The refrigeration system work with both

refrigerants show the room temperature was

increased as the internal heat loads increased.

The average gap of the HCR-22 and R-22 is

8.7%. The room temperature of the HCR-22

exhibits the positive improvement as its

temperature is close to to the temperature setting

(200C) compared to the R-22. The average

difference of room temperature between the

HCR-22 to the temperature setting is 1.95% and

for the R-22 is 9.8%. The less percentage reveals

the nearest temperature to the temperature

setting, and it indicates the best temperature of

the refrigeration system.

4. The evaporator cooling load of the HCR-22

indicates more heat was absorbed into the system

compared to the R-22. The average difference is

50.2%. This is a good improvement of heat

absorption.

5. The COP of the HCR-22 refrigeration system

shows an encouraging improvement compared to

the R-22. The average difference is ranging from

2.84 to 11.98%.

6. The electrical energy consumption of the HCR-

22 is less compared to the R-22 and it explains

the positive improvement of electrical energy

saving of the system. The average difference is

15.14%.

ACKNOWLEDGMENT

The present research was supported financially by

Redto Green (M) Sdn. Bhd., Automotive

Development Centre – UTM, Universiti Teknologi

Malaysia : Fundamental Research Grant Scheme

(FRGS) No.78686 from the Ministry of Higher

Education (MOHE) Malaysia and appreciation to

Refrigeration Laboratory, Faculty of Mechanical

Engineering, Universiti Teknologi Malaysia for

facilitating support for this research. Their guidance

and assistance are gratefully acknowledged.

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domestic refrigerators,‖ Applied Thermal

Engineering, vol. 19, pp. 1181–1189, 1999.

[12] D. Jung, C.B. Kim, B.H. Lim, and H.W. Lee,

―Testing of a hydrocarbon mixture in domestic

refrigerators,‖ ASHRAE Transactions, vol. 3,

pp. 1077–1084, 1996.

[13] X.H. Han, Q. Wang, Z.W. Zhu, and G.M. Chen,

―Cycle performance study on R-32/R-125/R-

161 as alternative refrigerant to R-407C,‖

Applied Thermal Engineering, vol. 27, pp.

2559-2565, 2007.

[14] K.J. Park, T. Seo, and D. Jung, ―Performance of

alternative refrigerants for residential air

conditioning applications,‖ Applied Energy, vol.

84, pp. 985–991, 2007.

[15] E.W. Lemmon, M.L. Huber, and M.Q.

McLinde, REFPROP, Reference Fluid

Thermodynamics and Transport Properties,

NIST Standard Reference Database 23, Version

9.0., 2009.


2012

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Characterization of Generator with Palm Oil Biodiesel

at Different Compression Ratio

Belyamina,b

, Alias bin Mohd. Noorc, Mohanad Hamzah Hussein

d,

Mazlan bin Saide, Mohd Haffidzi

f

aTransportation Research Alliance

Universiti Teknologi Malaysia, 81300 Skudai, Johor Bahru Tel : +60137675038

E-mail : [email protected]

bMechanical Engineering Department

Politeknik Negeri Jakarta, Depok 16425 Tel : +60137675038


cTransportation Research Alliance



dMechanical Engineering Faculty



eTransportation Research Alliance



fTransportation Research Alliance



ABSTRACT

Experiment to determine exhaust gas emission and

combustion characteristics of a compression

ignition generator was carried out. The experiment

used single cylinder four strokes direct injection

engine which was fueled with diesel and palm oil

methyl ester of B2 (blends 2% palm oil methyl

ester with 98% diesel on a volume basis), B5, B7

and B10. The experiment was conducted at a fixed

engine speed of 3000 rpm and 50% load with

variety compression ratios of 16:1, 18:1, and

20:1and 22:1. Optimum compression ratio,

influence of compression ratio on specific fuel

consumption and thermal efficiency were

examined. Palm oil methyl ester produce better

output when the engine operate with variable

compression ratio. Specific fuel consumption and

NOx decrease and thermal efficiency increase when

using highest compression ratio.

Keywords : palm oil methyl ester , variable

compression ratio, four stroke, direct

injection

1. INTRODUCTION There is a global increase in the investigation on

the application of alternative fuel sources for daily

use, such as biodiesel and alcohol. This is due to

the fact that petroleum products are becoming very

scarce and expensive and also the price of

petroleum products is always on the high side.

There is also an awareness of air pollution caused

by the extensive use of conventional fuel in an

internal-combustion engine.

In the past two decades, vegetable oil such as

mahua oil, sun flower, seed oil, waste cooking oil,

and palm oil have been used as a substitute to

diesel in an internal-combustion engine [1, 2, 3].

These papers study the performance and emission

characteristic of generator fueled with biodiesel or

its blend. It shows that biodiesel can substitute

fossil fuel in an internal-combustion engine with or

without engine modification. It is also economical

and competitive compare to pure diesel. When the

fuel is waste palm oil, no engine modification is

required.

IMAT-UI 008


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43

In addition, biodiesel has lower sulphur, aromatics

contents, and net carbon dioxide (CO2) emission

[4]. It has better lubricity and biodegradability and

less toxic relative to fossil diesel [5]. Bio diesel can

be used readily since it can be mixed at any

proportion with diesel. This enables it to be applied

immediately in diesel power generator without

much modification. Considering exhaust emissions,

reference [6] reported that the use of bio diesel

results in lower emissions of unburnt hydrocarbons,

carbon monoxide, smoke and particulate matter

beside some increase in emissions of NOx. A

number of researchers have investigated vegetable

oil-based fuels [2, 3, 7, and 8]. Reference [3] had

concluded that vegetable oil can be safely burnt for

a short period of time in a diesel engine. However,

the use of raw vegetable oil for extended period of

time may result in severe engine deposits, piston

ring sticking, injectors choking, and thickening of

the lubricating oil.

This experiment is to determine exhaust gas

emission and combustion characteristics of

a compression ignition generator using B2, B5, B7

and B10.

2. METHODOLOGY

Experiment to examine combustion characteristic

and exhaust gas emission was conducted on a

single-cylinder diesel engine four stroke for variety

CR (Compression ratio) and biodiesel formulation,

B2, B5, B7 and B10. Variety of CR is achieved by

changing the cylinder head gasket thickness. CR

increase when gasket thickness is reduced.

The exhaust gases emissions such as NOx, CO, CO2

and exhaust smoke density was measured by using

the emission analyzer. All these results were

discussed to find the effect of changing the CR and

fuel formula.

Power generated was loaded by set of lamps.

Ampere and volt of lamps were measured by

ampere meter and voltmeter as alternative to torque

measurement.

The generator set specification is as shown in Table

1. whereas fuel property is shown in Table 2.

Table 1. Engine test specification

Type Yanmar l70N6-MTRIYJ made in Italia 4 stroke, vertical cylinder diesel

generator engine

No. of cylinder 1

Bore x stroke 78 x 67mm

Displacement 0.320L

Combustion system Direct injection

Cooling system Forced air by flywheel fan

Maximum engine speed 3600(rpm)

Starting system Electric start/Recoil start

Max Rated output(KW] @3600rpm 4.9

Compression ratio 20:1

Table 2. Fuel test properties

Specification Pure diesel B2 B5 B7 B10

Specific gravity

@60f0

0.8448 0.8448 0.8453 0.8458 0.8473

Density @15c0kg/m

3 844.3 844.3 844.8 845.3 846.9

A p I 36.0 36.0 35.9 35.8 35.7

Viscosity 28c0 7.123 6.897 6.893 6.881 6.573

Viscosity 40c0 5.17 5.131 5.063 5.048 5.019

Viscosity 100c0 1.781 1.671 1.667 1.638 1.579

Viscosity index 243.35 250.57 263.73 266.74 272.66

Pour point c0 3 3 3 3 3

Flash point c0 84.0 83 86 86 87

gross calorific

value(kJ/kg)

45652.0 45801.3

45589.0

45235.9

45011.9


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3.1 Specific Fuel Consumption

Fuel consumption, FC, is calculated by Equation 1.

FC = V / t (1)

Specific Fuel Consumption is then calculated by

Equation 2

SFC = FC/ P (2)

The variation of Specific Fuel Consumption (SFC) with

Compression Ratio (CR) is given in Figure 1. It can be

observed that the SFC is a clear indication of efficiency

with which the engine develops power. The smaller SFC

indicates the more effective use of fuel to generate power.

For all fuels tested, the SFC decreased with an increased of

CR. This is due to increase of temperature in combustion

chamber, leading to complete combustion. It has been

observed that the maximum SFC of B2 reduced by 0.96%

at CR 22:1 relative to CR 20:1 and the other fuels. This

can be because B2 has high caloric value compare to other.

At CR 18 and 16, SFC B5 and B7 have the highest rise of

SFC by 1.1%, 1.26% respectively. Minimum SFC mean

efficient use of fuel. This happen in CR 20 to 22. B2

always have the lowest vaue of SFC which mean the best

one in term of fuel efficiency. At higher percentage of

blends, the SFC increases due to decrease in calorific

value.

Figure 1.Variation of specific fuel consumption with compression ratio for different fuel blends

3.2 Thermal Efficiency.

Thermal efficiency is calculated by Equation 3

ɳth = P/ Qin (3)

Qin is calculated by Equation 4

Qin = m CV (4)

The thermal efficiency (ɳth) of the engine is considered one

of the most important criteria for evaluating the

performance of the engine. It indicates the combustion

effectiveness of the engine. The ɳth is defined as the actual

work per cycle divided by fuel chemical energy (fuel

calorific value). Figure 2 shows the variation in ɳth with

CR for blends fuel tested .The ɳth was found to be lower

for biodiesel at all blends than diesel for all specified CR.

This might be due to lower fuel heat value and higher fuel

consumption of the bio diesel blends to produce the same

power.

In this figure, it appears that the optimum ɳth of B2 occur

at CR 22:1. This may be due to the fuel calorific value and

low SFC. At CR18 and CR16 the ɳth reduce by 11%, 17%

for B5, B7 respectively compare to CR20, it was also

observed that increasing the CR more beneficial when

biodiesel is used rather than diesel. Due to their low

volatility and high viscosity, biodiesel perform relatively

better at higher compression ratios [9].

3.3 Exhaust Gas Emission

3.3.1 Nitrogen Oxide

The variation of Nitrogen Oxide (NOx) with respect to CR

for different blends and constant load is shown in Figure 3.

NOx emission for diesel and other blends increase when

the CR is increased. The augmentation in the biodiesel

ratio in the fuel blend increased NOx emissions by 1.07%,

1.12%, 1.16% and 1.18% for B2, B5, B7 and B10,

respectively, the reason for higher NOx emission for

blends is higher peak temperatures. This figure shows an

increase in NOx by 1.3% at CR 22 while the NOx

decreased by 17.7% and 30% at CR 18:1 and 16:1

respectively as compared to the CR 20. The changes in

NOx resemble up to some extent to exhaust temperature

which is related to an increase in CR.


2012

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Figure 2.Variation of thermal efficiency with compression ratio for different fuel blends

Figure 3.Variation of Oxide Nitrogen with compression ratio for different fuel blends

3.3.2 Carbon Monoxide

Figure 4 illustrate the variation of CO for variety fuel

blends with respect to VCR. From this figure it seen that

the specified blends produce less CO emission than diesel

for every CR at applied load, it might due to increase the

cylinder temperature. Therefore engine temperature lead to

better combustion process and might cause less CO

emission. The CO decrease by 29.1%, 10.87%, 4.2% and

0.5%

when the compression ratio increase from 20:1 to 22:1 for

diesel,B2,B5,B7 and B10 respectively, This could be

because biodiesel provide more oxygen to the combustion

chamber. This lead to the more complete combustion. The

other reason is that the percentage of CO decreases due to

rising temperature in the combustion chamber. physical

and chemical properties of the fuel, air–fuel ratio, the

effects of fuel viscosity on spray quality will be expected

to cause CO emission increase with vegetable oil

fuels [10].


2012

46

Figure 4.Variation of Carbon Monoxide with compression ratio for different fuel blends

3.3.3 Carbon Dioxide

The variation of CO2 with VCR is shown in Figure 5.

From this figure, it can be observed that the CO2 increases

by around 18% than CR20. In decreasing CR, blends of

fuel increase its CO2 emission by 4%, 4.5%, 8% and 11%

for the B2, B5, B7 and B10 respectively. This is due to the

high oxygen content of blends. Higher amounts of CO2 is

an indication of complete combustion of fuel in the

combustion chamber. It also relates to the exhaust gas

temperature. CO2 emissions of the fuel blends slightly

increase by increasing the load for specified compression

ratios due to complete combustion.

3.3.4 Smoke Density

The variation of Smoke density emission with VCR at

constant load is shown in Figure 6. In this figure, it was

shown that the smoke density increased by 35% and 60%

when the CR decreased to 18:1 and 16:1 respectively. At

lower CR the temperature is lower. Incomplete

combustion in the combustion chamber then lead to more

smoke exhausted from the engine. The smoke decreased

when the blends percentage is increased. These figures

show that the smoke was reduced significantly by around

9%, 12%, 17% and 22% for B2, B5, B7 and B10 than

diesel. In addition, it was found that the maximum

reductions were around 20% at CR 22 for all fuel blends

than CR20, this is due to the increase of inside temperature

of the combustion chamber and because palm oil contains

more oxygen which improves the combustion process. At

the end this will decrease the smoke. The Smoke is emitted

from diesel engines because of the incomplete combustion

in the combustion chamber.

Figure 5. Variation of Carbon Dioxide with compression ratio for different fuel blends


2012

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Figure 6. Variation of Smoke with compression ratio for different fuel blends

4. CONCLUSION

An experimental was conducted on direct injection diesel

engine generator evaluate the performance, combustion

and exhaust emission at different blends and compression

ratio. Thermal efficiency, ɳth of the generator and CO2

emission increases with an increase in the CR whereas

SFC decreased. The use of biodiesel in general reduce the

CO emission and increase CO2 which indicate better

combustion performance. Performance of the B2 is

superior compare with Diesel. Optimum CR to provide

optimum SFC and Thermal efficiency is 20 and above

ACKNOWLEDGMENT

The authors are grateful to the Ministry of Higher

Education and Universiti Teknologi Malaysia for grant

GUP project under Vot Q.J130000.2609.00J35

NOMENCLATURE

B2 mixture of 2% biodiesel and 98% diesel




CO carbon monoxide

CO2 carbon dioxide

CR compression ratio

CV Calorific Value of fuel

D 100% diesel

FC Fuel consumption

m mass flow rate of fuel

NOx nitrogen oxides

O2 oxygen

P Power generated

Qin Heat input

t time taken to consume the fuel

SBFC specific brake fuel consumption

VCR variable compression ratio V volume of fuel used

ɳth brake thermal efficiency

REFERENCES

[1]. S.C.A. ALMIEDA, C.R. BELCHIOR, M.V.G. NASCIMENTO,

L.S.R. VIEIRA AND G. FLEURY, ―PERFORMANCE OF A DIESEL

GENERATOR FUELLED WITH PALM OIL‖, FUEL, VOL 81,

PP.2097-2102, 2002.

[2] A.SRIVASTAVA, R.PRASAD, ―TRIGLYCERIDES-BASED DIESEL

FUELS‖, RENEWABLE SUSTAINABLE ENERGY REV., VOL 4,

PP.111–33, 2000.

[3] F.KARAOSMANOGLU, G.O.KURT, ¨ ZAKTAS T .LONGTERM CI

ENGINE TEST OF SUNFLOWER OIL‖, RENEWABLE ENERGY,

VOL.19, PP.219–21, 2000.

[4]. M.Z.SULAIMAN, F.M.ISA, ―THE EFFECT OF DIFFERENT

GASOLINE BLENDS DOPED WITH USED ENGINE OIL ON THE

FORMING TENDENCY OF SIMULATED IN TAKE VALVE

DEPOSITS‖, PROCEEDINGS OF THE INSTITUTION OF MECHANICAL

ENGINEERS, PP.213, PART D, 1999.

[5] H.Z.GOETLLEN, M.ZIEJEWSKI, K.R.KAUFMAN, G.L. PRATT,

―FUEL INJECTION ANOMALIES OBSERVED DURING LONG-BURN

ENGINE PERFORMANCE TEST ON ALTERNATE FUELS‖, SAE

TECHNICAL PAPER SERIES. SOCIETY OF AUTOMOTIVE

ENGINEERS, NO.852089, 1985.

[6] M.S.GRABOSKI AND R.L.MCCORNIMIK, ―COMBUSTION OFF

AT AND VEGETABLE OIL DERIVED FUELS IN DIESEL ENGINES‖,

PROG ENERGY COMBUST SCI ., VOL.24, PP.125–64, 1998.

[7] A.ISIGIGUR, F.KARAOSMANOGLU, H.A.AKSOY,

F.HAMDULLAHPUR, L.O.GULDER, ‖PERFORMANCE AND

EMISSION CHARACTERISTICS OF A DIESEL ENGINE OPERATING

ON SUNFLOWER SEED OIL METHYLESTER‖, APP BIOCHEM

BIOTECHNOL, VOL.NO.45/46, PP.93–102, 1994.

[8] R.ALTIN, S.CETINKAYA AND H.S.YUCESU, ―THE POTENTIAL

OF USING VEGETABLE OIL FUELS AS FUEL FOR DIESEL

ENGINES‖, ENERGY CONVERSION MANAGEMENT . VOL.42,

PP.529–38, 2001.


2012

48

[9] H.RAHEMAN AND S.V. GHADGE, ―PERFORMANCE OF DIESEL

ENGINE WITH BIODIESEL AT VARYING COMPRESSION RATIO AND

IGNITION TIMING‖, FUEL, VOL.87, PP. 2659–2666, 2008.

[10].K.MURALIDHARAN AND D.VASUDEVAN, ―PERFORMANCE,

EMISSION AND COMBUSTION CHARACTERISTICS OF A VARIABLE

COMPRESSION


2012

49

The Effect of Fuel Additives on Gasoline Heating Value and

Spark Ignition Engine Performance: Case Study Z.A. Latiff

a, Azhar Abdul Aziz

b, M.R. Mohd Perang

c, N. Abdullah

d

Automotive Development Centre (ADC)

Faculty of Mechanical Engineering

Universiti Teknologi Malaysia (UTM)\

81310 Johor Bahru, Malaysia

[email protected]

[email protected] [email protected]

[email protected]

ABSTRACT

Today fuel additives had been used widely for the

enhancement of fuel economy and engine

performance. Fuel additives are substance that acts

as catalysts for the completeness combustion of

fuel in order to increase the heat released and hence

the work output will be improved. The purpose of

this paper is to investigate the effect of the

additives on fuel heating value and engine

performance. In this study, three different additives

available in the market have been chosen to

determine the effect on heating value and engine

performance when mixed with fuel. Two types of

test were conducted, namely the calorific value and

engine performance test. The first test was

conducted using a bomb calorimeter with test

method in accordance with the DIN 51900 and

ASTM D240. The later test was done using engine

test bed and with the agreement of BS 5514 (Parts

1 to 6), Reciprocating Internal Combustion

Engines: Performance, and SAE 1349 Standard

Engine Power Test Code. The study shows that fuel

additives can cause a standard fuel to have higher

heating value up to 5%. As for the engine

performance, the engine brake thermal efficiency

and brake mean effective pressure were increased

up to 8% and 10% respectively. The specific fuel

consumption can be reduced up to 9%.

Keywords : Engine performance, Fuel additive,

Fuel economy

1. INTRODUCTION

Fuel additives are compounds formulated to

enhance the quality and efficiency of the fuels used

in motor vehicles. This additive can be

incorporated in the fuel itself as needed. The fuel

additive is sold as a separate product and

consumers may use to improve or maintain the

performance of their engines.

One of the main advantages of the fuel additives is

to improve the engine performance. With some fuel

additives added, the product is claimed to boost the

octane level of the fuel, providing the engine with

more power with the same of fuel used. Thus with

this improvement one have the ability to travel

more in other words better mileage.

Along with preventing carbon build-up in the

combustion chamber, fuel additives are claimed to

also enhance proper lubrication of working

components. This particular benefit means less

wear and tear due to less surface friction on the

moving parts, which translates into lower and less

frequent repairs during the life of the vehicles.

2. IMPORTANCE OF FUEL ADDITIVE

Most people agree that some of the basic fuel

additives found in many gasoline products today

are of some benefit, there is some issues regarding

the use of over the counter additives. Proponents

claim the product boost the protection offered by

the gasoline products and make a significant

difference in how well a vehicle performs.

Opponents claim that over-the-counter additives for

fuel provide no extra benefits and in fact could

damage the engines if nor used properly.

A number of experimental investigations have been

reported with a wide variety of metal additives to

improve the fuel properties and the engine

performance, as well as to reduce emissions. The

effect of calcium, barium, iron, and nickel

naphthenates have been studied, concluding that

calcium and barium most efficiently reduce soot,

by both suppressing soot formation and enhancing

soot oxidation [1]. Based on experimental

investigations, Gürü et al. [2] concluded that

manganese, as a fuel additive, has a greater effect

in the reduction of the freezing point of the fuel,

than copper, magnesium, or calcium. Emission

measurements with manganese as a fuel additive

demonstrated that O2 and CO could be decreased

by 0.2% and 14.3%, respectively, SO2 emission

could be reduced, and the overall impact of all

these effects was found to lead to an increase of

0.8% in the net operating efficiency.

Valentine et al. [3] experimentally observed that

bimetallic platinum and cerium diesel fuel borne

catalyst reduces the engine emissions and improves

IMAT-UI 009



2012

50

the performance of the diesel particulate filter.

Shi et al. [4] reported that the particulate matter

emission decreases with increasing oxygenate

content in the fuels, but nitrogen oxides emissions

increase. De et al. [5] experimentally observed that

the presence of ethanol and ethyl ter-butyl ether

(ETBE) significantly alters the characteristics of

volatility and reduces the cetane number, impairing

the fuel‘s performance in engine tests. The effect of

methanol-containing additive (MCA) on the

emission of carbonyl compounds generated from

the diesel engine was studied by Chao et al. [6] and

it was observed that the emission factors for some

of the carbonyl compounds with the use of MCA

are higher than the values for those without the use

of MCA.

Metal oxides such as those of copper, iron, cerium,

and cobalt have been extensively used as fuel

additives. The effect of cerium on the size

distribution and composition of diesel particulate

matter has been studied by Skillas et al. [7],

indicating a reduction in the accumulation mode,

but an increase in ultrafines. Lahaye et al. [8]

studied the effect of cerium oxide on soot

formation and postoxidation and observed that the

soot yield is not affected significantly by the

presence of cerium oxide in the fuel for given

oxygen content. Based on experiments,

Jung et al. [9] observed that the addition of cerium

to diesel fuel causes significant changes in the

number concentration of particles in the

accumulation mode, light off temperature, and the

kinetics of oxidation. Even though the 4 oxidation

rate increased significantly with the addition of

cerium to the fuel, the dosing level was found not

to have much influence [10, 11].

3. HEATING VALUE

The heating value or calorific value of a substance,

usually a fuel or food, is the amount of heat

released during the combustion of a specified

amount of it. The calorific value is a characteristic

for each substance. It is measured in units of energy

per unit of the substance.. Heating value is

commonly determined by the use of a bomb

calorimeter.

The quantity known as higher heating value (HHV)

(or gross calorific value or gross energy or upper

heating value) is determined by bringing all the

products of combustion back to the original pre-

combustion temperature, and in particular

condensing any vapor produced. This is the same

as the thermodynamic heat of combustion since the

enthalpy change for the reaction assumes a

common temperature of the compounds before and

after combustion, in which case the water produced

by combustion is liquid.

The quantity known as lower heating value (LHV)

(or net calorific value) is determined by subtracting

the heat of vaporization of the water vapor from the

higher heating value. This treats any H2O formed as

a vapor. The energy required to vaporize the water

therefore is not realized as heat. Gross heating

value accounts for water in the exhaust leaving as

vapor, and includes liquid water in the fuel prior to

combustion. This value is important for fuels like

wood or coal, which will usually contain some

amount of water prior to burning.

The water vapor produced by combustion,

recovering heat which would otherwise be wasted.

Most applications which burn fuel produce water

vapor which is not used and thus wasting its heat

content. In such applications, the lower heating

value is the applicable measure. This is particularly

relevant for natural gas, whose high hydrogen

content produces much water. The gross calorific

value is relevant for gas burnt in condensing boilers

and power plants with flue gas condensation which

condense.

3.1 Bomb Calorimeter Test

A calorimeter is a device used for calorimetry, the

science of measuring the heat of chemical reactions

or physical changes as well as heat capacity. The

word calorimeter is derived from the Latin word

calor, meaning heat. Differential scanning

calorimeters, isothermal microcalorimeters,

titration calorimeters and accelerated rate

calorimeters are among the most common types. A

simple calorimeter just consists of a thermometer

attached to a metal container full of water

suspended above a combustion chamber.

A bomb calorimeter is a type of constant-volume

calorimeter used in measuring the heat of

combustion of a particular reaction. Bomb

calorimeters have to withstand the large pressure

within the calorimeter as the reaction is being

measured. Electrical energy is used to ignite the

fuel; as the fuel is burning, it will heat up the

surrounding air, which expands and escapes

through a tube that leads the air out of the

calorimeter. When the air is escaping through the

copper tube it will also heat up the water outside

the tube. The temperature of the water allows for

calculating calorie content of the fuel.

Fuels such as coal and oil, are traded based on the

calorific value of the material. Regulations have

been established with regards to the total calorific

content of the coal, the quality or purity of the coal,

and the classification of the coal. The gross

calorific value of the coal is also used to evaluate

the effectiveness of the beneficiation process in use

at the plant. ISO 1928, ASTMD 5865, BS1016, and

DIN 51900 are the most common methods for

determining the gross calorific value of coal and

coke.

Liquid fuels such as gasoline, kerosene, diesel, and

gas turbine fuels are also tested by bomb


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calorimeter. The heat of combustion (HOC) of the

fuel will provide a measure of the energy available

from a fuel. The mass heat of combustion (essential

for airplanes and hydrofoils) and the volumetric

heat of combustion (essential for automobiles and

ships) can also be determined with the HOC value.

ASTM D240 and D4809 are common methods for

this sort of work.

4. ENGINE PERFORMANCE

An internal combustion engine is a work producing

device that converts chemical energy into heat

followed by conversion into mechanical energy.

Chemical energy contained in the fuel is released as

heat by burning the fuel inside the combustion

chamber to produce gas at high temperature and

pressure. The high pressure gas acts to move the

piston and other related mechanisms inside the

engine to produce mechanical work.

Some parameters have to be obtained to evaluate

the behavior and performance of an internal

combustion engine such as speed, brake load and

fuel consumption. The experimental data have to be

analyzed to determine brake power, brake mean

effective pressure, specific fuel consumption and

brake thermal efficiency.

4.1 Engine Performance Test

The test was done using engine test bed and with

the agreement of BS 5514 (Parts 1 to 6),

Reciprocating Internal Combustion Engines:

Performance, and SAE 1349 Standard Engine

Power Test Code.

5. GASOLINE AND ADDITIVES

MIXTURE

Three type of additives were selected in this study

and the details are given as follows:

Additive A – Power additive which contains

mixture of oil, dimenthyl heptanes, trimethyl

etc.

Additive B – Cleaner additive with high purity

of polyetheramine (PEA) detergent; and

Additive C – Power additive with concentrated

biodegradable formula.

The mixture preparation of gasoline and

additive was down according to the proportion

suggested by the additive manufacturer.


Table 1 shows the energy value of fuel mixed with

different additive concentration for each sample,

established by using bomb calorimeter test.

Table 1: Fuel energy for different additive concentration

Additive Additive

Concentration (%) A B C

0 42146.5 40714.4 39983.4

5 42313.2 41425.4 42473.0

10 43101.7 42416.3 42710.8

15 43680.0 43078.0 42826.6

20 43861.8 44066.2 43804.4

Figure 1 shows the profile of energy content with

different additive concentrations. The trend shows

increasing in energy posed by the fuel mixture as

the concentration increases. It also shows additive

A have the highest increased in energy value then

followed by C and B. In term of percentage

increased, additive A, B and C shows 5.42, 3.93

and 3.88% respectively when compared to the

standard fuel energy value.

For engine performance test the preparation of fuel

used was based on the standard mixing procedure

recommended by the additive manufacturers.

Figure 2 shows the brake power against engine

speed. The test was done at part throttle opening.

The graph show the increased in brake power at

low speed and high load compared to standard fuel

used. While at high speed, low load the parameter

shows no significant difference. Generally additive

B has the highest average brake power increment of

9.51% and followed by additive A 8.44%. Additive

C in contrast shows a negative improvement except

at low speed.

Figure 1: Energy content at different additive

concentration

Figure 2: Brake power against engine speed


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Specific fuel consumption shows the economic

point of view for fuel usage and shown in Figure 3.

Based on the same power output, additive B shows

attractive saving followed by additive A and C.

Average saving made are 9.04, 9.06 and 5.9%

respectively for additive A, B and C.

Figure 3: Specific fuel consumption against

engine speed

Brake mean effective pressure is a measure of an

engine's capacity to do work that is independent of

engine displacement. From the experiment done the

measures is shown in Figure 4. Conclusively there

are no major changes made by the additives as far

as brake mean effective pressure is concerned.

Figure 4: Brake mean efective pressure against

engine speed

The ability of an engine to convert thermal energy

to available work can be shown in Figure 5 as

brake thermal eficiency. The thermal energy

obtained from the combustion of air-fuel mixture

represent 100% of energy input to the engine. The

experimental results exhibit additive B posed the

most potential of poducing work, moderately by

additive A and least work produced by additive C

compared to the standard fuel used.

Figure 5: Brake thermal efficiency against

engine speed

7. CONCLUSION

From this experimental study the following

conclusions are made:

Additives mixed with fuel increased the

energy content of the fuel.

The additives used have the effect on

engine performances as discussed.

The value of energy content has

significant effect on the engine

performance.

REFRENCES [1] N. Miyamoto, H. Zhixin, A. Harada, H.

Ogawa, and T. Murayama, Characteristics of

Diesel Soot Suppression with Soluble Fuel

Additives. SAE Technical Paper 871612,

1987.

[2] M. Gürü, U. Karakaya, D. Altiparmak, and

A. Alicilar, Improvement of Diesel Fuel

Properties By Using Additives, Energy

Conversion And Management. Volume. 43,

no. 8, pp. 1021–1025, 2002.

[3] J. M. Valentine, J. D. Peter-Hoblyn, and G.

K. Acress, Emissions Reduction and

Improved Fuel Economy Performance From

A Bimetallic Platinum/Cerium Diesel Fuel

Additive At Ultra-Low Dose Rates. SAE

Technical Paper 2000-01-1934, 2000.

[4] X. Shi, Y. Yu, H. He, S. Shuai, J. Wang, and

R. Li, Emission Characteristics Using

Methyl Soyate—Ethanol—Diesel Fuel Blends

On A Diesel Engine. Fuel, Vol. 84, no. 12-

13, pp. 1543–1549, 2005.

[5] E. W. De Menezes, R. Da Silva, R. Cataluña,

and R. J. C. Ortega, Effect Of Ethers And

Ether/Ethanol Additives On The

Physicochemical Properties Of Diesel Fuel


2012

53

And On Engine Tests Fuel. Vol. 85, no. 5-6,

pp. 815–822, 2006.

[6] H.-R. Chao, T.-C. Lin, M.-R. Chao, F.-H.

Chang, C.-I. Huang, and C.-B. Chen, Effect

Of Methanol-Containing Additive On The

Emission Of Carbonyl Compounds From A

Heavy-Duty Diesel Engine, Journal of

Hazardous Materials B, 73, pp. 39–54, 2000.

[7] G. Skillas, Z. Qian, U. Baltensperger, U.

Matter, and H. Burtscher,. Influence of

Additives on the Size Distribution and

Composition of Particles Produced By Diesel

Engines. Combustion Science and

Technology, 154,. 259–273, 2000.

[8] J. Lahaye, S. Boehm, P. H. Chambrion, and

P. Ehrburger, Influence Of Cerium Oxide On

The Formation And Oxidation Of Soot,

Combustion and Flame, 104,199–207, 1996.

[9] H. Jung, D. B. Kittelson, and M. R.

Zachariah, The Influence Of A Cerium

Additive On Ultrafine Diesel Particle

Emissions And Kinetics Of Oxidation.

Combustion and Flame, 142, 276–288, 2005.

[10] B. Stanmore, J. F. Brilhac, and P. Gilot, The

Ignition And Combustion Of Cerium Doped

Diesel Soot, SAE Technical Paper 01-0115,

1999.

[11] K. Pramanik, Properties and Use of Jatropha

Curcas Oil and Diesel Fuel Blends In

Compression Ignition Engine. Renewable

Energy 28, 239–248, 2003.


2012

54

Design a Four-Stroke Homogeneous Charge Compression

Ignition (HCCI) Engine

Mohd Rozi Mohd Perang

a, Zulkarnain Abdul Latiff

a,b, Azhar Abdul Aziz

a,b, Mohamad

Azzad Mokhrib


81310 Skudai, Johor, Malaysia, Phone: +60 75535447, Fax: +60 75535811

email: [email protected], [email protected], [email protected] bFaculty of Mechanical Engineering, Universiti Teknologi Malaysia

81310 Skudai, Johor, Malaysia, Phone: +60 7 5534575, Fax: +60 7 5566159

ABSTRACT

This research is to study the operation of the four-

stroke HCCI engine. The design and analysis works

have been performed using computer software

which is GT-Power and Solidwork to study on the

engine performance simulation work and 3-D

modelling on the combustion chamber designed

respectively. The design is based on 4-cylinder

passenger car, 2000 cc and a four-stroke cycle

engine. The compression ratio used is 10. The fuel

used is ethanol in which the air-fuel ratio (AFR) is

9. The parameters selected have typical range of

value based on the previous study and research

done. With the use of GT-power, the analysis will

consider two parameters which are the cam timing

angle and the injection timing angle to get the

optimum result for the HCCI engine. The typical

angle of cam timing angle is between 2600 – 270

0

since this is the moment of the compression cycle

of the engine. For the injection timing angles, the

angles that will be studied for this project are 50, 0

0,

-50, -10

0,-15

0 and -20

0 relative to Top Dead Centre

(TDC). The objective is to obtain the maximum

torque and brake power when the engine speed is in

between 4000 rpm to 5000 rpm and 6000 rpm to

7000 rpm respectively. Finally, the optimum

conditions for the engine to perform better are at

2640 of cam timing angle for the valve and at -5

0

before TDC for the injection timing angle. The

maximum torque and brake power achieved is

37.60 Nm at 4000 rpm and 23.46 kW at 7000 rpm.

Keyword : Homogeneous Charge Compression

Ignition (HCCI); combustion

characteristics; ethanol fuel; gt-power

1. INTRODUCTION

The Internal Combustion Engine (ICE) is the

pioneer of the engine maker in the automotive

industry closely related to spark ignition (SI) and

compression ignition (CI) engine. Each of them

posses their own advantages and disadvantages.

Two-stroke engine is known for higher weight-to-

power ratio and four-stroke engine provide better

combustion and less pollution. The industry is

aimed to enhance the combustion behaviour of the

ICE by applying the HCCI engine mode that can

improve the engine operator. Thus, all big

automotive companies in Europe and Japan such as

General Motor (GM) and Toyota invest in a

research on HCCI engine that can be implemented

in car for future.

It is a challenge for the engineers nowadays to

develop an engine that have better performance

without neglecting the environmental issue. For the

current research technology, HCCI engine is

leading the new innovative technology in ICE.

HCCI is one of the ICE process as an alternative to

the conventional SI and CI engine combustion

process. The HCCI phenomenon is a combination

of the best features of both SI and CI engine which

is to produce an engine operating with very low

emission and high efficiency respectively. HCCI is

similar with SI in term of premixed charge usage

and also performs auto ignition to initiate the

combustion as practiced by CI [1]. In addition,

another advantage of HCCI engine is it can

consume any type of fuel. HCCI engine is the new

trend under extensive research nowadays. The

reason why HCCI engine is under the spotlight

today is the ability of the engine to satisfy better

performance without having to sacrifice the

environmental issues.

In the HCCI engine, the fuel and air should be

blended homogenously before the combustion

starts and that mixture will be auto-ignited due to

the increase in engine cylinder temperature and

pressure from compression stroke [2]. The

combustion process of HCCI is totally different

with SI and CI engines. HCCI combustion

capability is performed by controlling the

temperature, pressure, and composition of air and

fuel mixing to spontaneously ignite in the engine.

Since ignition occurs in HCCI engine is by auto-

ignition of the air/fuel mixture, the selection of fuel

will have an important impact on both engine

design and control system. All of the HCCI engine

concepts will be developed around the requirement

of making the fuel auto-ignition at the correct

timing to achieve good combustion phasing, and a

IMAT-UI 010





2012

55

rapid bulk burn to obtain high fuel economy and

minimize emissions.

In this study, alcohol based-fuel, ethanol was used

as the alternative fuel to operate the HCCI 4-stroke

engine mode. Currently, ethanol is one of the

renewable energy which widely used at Brazil and

United States [3]. Ethanol was created from the

natural agriculture feedstock. It is produced from

subsistent crops such as sugar cane, potato and corn

[4]. Besides that, ethanol can be produced through

cellulosic waste. This source offers promise

because cellulose fibres, a major component in

plant cells walls, can be used to produce ethanol in

large quantities.

The paper will focus on the design of combustion

chamber by identifying the cam and injection

timing angle in order to obtain the optimum result

for the HCCI engine. The research is based on the

simulation analysis via Gt-Power. The objective is

to obtain the maximum torque and brake power

when the engine speed is in between 4000 rpm to

5000 rpm and 6000 rpm to 7000 rpm respectively.

2. SIMULATION ANALYSIS

The simulation involved throughout this research is

GT-Power as to study on the engine performance.

The 3-D modeling of combustion chamber design

will be performed by CAD drawing, Solidwork.

Figure 1 shows the flow of the research activity.

Figure 1: Research activity flow process

Initially, some parameters of the combustion

chamber were calculated manually. Then, those

values were computed in the GT-Power, which was

modelled earlier according to the scope.

Subsequently, the software will solve the problem

and execute the analysis. The results were

presented in the form of tables and graphs.

Finally, the optimum condition of the design will

be decided based on the results by providing

relevant reasons and evidences. In addition, the

schematic diagram of ECU is also included since it

is very important and has significant value for this

research.

For manual calculation, it focused on the

parameters of the combustion chamber such as the

size of the bore, the length of the stroke and the

swept volume. Before that, some specifications

were established in order to limit the scope and

guide the analysis process. The specifications for

this study are based on the spark ignition engine,

passenger car and four stroke cycles.

Hence all the parameter can be summarized as the

following;

Table 1: Summary of the parameters for the engine

Parameters Symbol Value

Bore B 0.086m

Stroke S 0.086m

Swept volume Vd 4.996 x 10-4

m3

Total swept volume

(4 cylinder)

Vd,total 1.998 x 10-3

m3

Clearance volume Vc 5.55 x 10-5

m3

Compression ratio rc 10

Crack radius/offset a 0.043m

Connecting rod r 0.15m

r/a ratio R 3.5

Air-fuel ratio AFR 8.99

Engine cycle - 4-stroke

Crack angle ɵ 00-720

0

Intake valve - 1

Exhaust valve - 1

Fuelling system - IDI (indirect

injection)

Start of injection

angle

- 50, 0

0, -5

0, -10

0,

-150, -20

0

Cam timing angle - 2600 - 270

0

2.1 Method of using the GT Power

The method using GT-Power software were,

2.1.1 Run the software. Make sure that the interface

of the program as shown in Figure 2. The format of

the file is *.gtm. The red circle is the Template

Library and the blue circle is the Project Area. The

Template Library contains all the available

templates that can be used in GT-Power

applications whilst the Project Area model can be

created by our virtual engine according to our

design.

No

No

Yes

Yes

Literature review

Parameter setting

(using Gt-Power)

Satisfy?

3D Modeling using

Solidword

Satisfy?

Final engine design


2012

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Figure 2: The interface of GT Power for modelling

the designed engine

2.1.2 Some of the templates in the Template

Library will be used in this project. Drag the

template into the Project Area according to the

project requirement. The templates that needed are

as shown in Table 2.

Table 2: Templates used for the project

Flow Folder

EndEnvironment

EngCylinder

Pipe

InjAF-RatioConn

InjProfileConn

OrificeConn – def(object)

ValveConn

Ethanol – Vapc (object)

N2 – vapc (object)

O2 – vapc (object)

Air (object)

Mech Folder

EngineCrankTrain

2.1.3 Next, fill in the template parameter with

appropriate value according to the project

objectives. The default value can be applied to

some of the parameters as it is compatible for all

engine designs. However, some parameters such as

size of the bore, length of the stroke, compression

ratio and others as shown in Table 2 are required to

be filled in the templates. Those parameters need to

be computed accordingly to the manual calculation

earlier from the design. Table 2 exhibits the value

that needed to define according to the design. The

start of the injection angle and cam timing angle

are manipulated parameters which will be adjusted

in order to get the optimum result.

2.1.4 Then components can be arranged according

to its sequence. The procedure is to drag the

template into the Project Area. Automatically it

will convert the value into an object. The sequence

is as illustrated in Figure 3.

Environment – intake runner – intake port – intake

valve – cylinder – IDI injection (above cylinder) –

exhaust valve – exhaust port – exhaust runner –

environment

Figure 3: The sequence of template in the project

Area

Next, make the connection between the objects.

There is a button named Create Link at the toolbar

which can do the linkage (Figure 4).

Figure 4: The location of the icons and its

purpose

2.1.5 Finally, the schematic diagram as Figure 5

will be obtained;

Figure 5: The schematic diagram of the

designed engine in the GT Power

2.1.6 Then, the study case for this project can be

set up and Run/Case Setup was selected to run the

simulation program. The engine speed varies from

0 rpm to 10 000 rpm. Next, the analysis can be

started by clicking on Run/Start Simulation. The

result can be viewed in GT-Post. This software can

show the data in the form of tables and also graphs.

Create Link


2012

57

3. ANALYSIS SETUP

Two important parameters will be manipulated in

this project to generate a new parameter (Table 1)

that will be fixed in order to obtain the optimum

result for the engine. The first parameter is the cam

timing angle and the second is the injection timing

angle.

The cam timing angle is the angle when the intake

valve will be lifted to permit fresh air get into the

combustion chamber. The typical angle is in

between 260⁰ - 270⁰ since this is the moment of the

compression cycle of the engine.

The injection timing angle is the moment of the

crank angle when the injector will spray the fuel

into the combustion chamber. The fuel and air are

assumed to have enough time to mix

homogeneously inside the combustion chamber

before the charge combust to produce power. The

injection timing angles that will be studied for this

project are 50, 0

0, -5

0, -10

0, -15

0 and 200

relative to

Top Dead Centre (TDC).

The objective of this work is to get the highest

torque when the engine speed runs between 4000

rpm to 5000 rpm. This is because it will help the

vehicle to accelerate faster for the short period at

the earlier time, for instance to overtake another

vehicle in front. In addition, the maximum brake

power desires to be achieved during the engine

speed operates between 6000 rpm to 7000 rpm. The

reason is the vehicle will reach the maximum speed

when the load is constant with stable driving

condition. As reference, some important data such

as maximum brake power and maximum torque for

several types of engines are collected in order to

compare and validate our result. Table 3 shows the

data.

Table 3: The maximum brake power and torque for

several engines

Parameter

Engine

4B11

(Mitsubi

shi

Lancer)

N74

(Ford)

N34

(Ford)

R4 FSI

(Volkswa

gen Golf)

Capacity (cc) 1998 1998 1998 1984

Bore (mm) 86.0 86.0 86.0 82.5

Stroke (mm) 86.0 86.0 86.0 92.8

Compression

ratio 10:1 10.3:1 9.8:1 10.5:1

Maximum

power (kW) 114 @

6000 pm

110

@630

0 rpm

101 @

6300

rpm

147 @

6600 pm

Maximum

Torque (Nm)

198 @

4250

rpm

190 @

4500

rpm

175 @

4200

rpm

280 @

4700 rpm


Table 4 exhibits the result on the cam timing angle

analysis. It shows that the maximum torque and

power decreased when the engine speed and cam

angle increased from 2500 - 5000 rpm and 2600 -

2700

respectively. However the engine speed values

for all maximum brake power are not in the

targeted range, which is between 6000 rpm and

7000 rpm. The targeted engine speed for maximum

torque is in between 4000 - 5000 rpm. Finally the

cam timing angle selected is 2640 since the angle

produced the highest maximum torque at 4000 rpm,

which is within the targeted range. Table 4: The result of various cam timing angles

Cam

timing

angle

(o)

Maximum torque

(Nm) @ Engine

speed (rpm)

Maximum brake

power (kW) @

Engine speed (rpm)

260 41.52 @ 2500 24.38 @ 8000

261 40.81 @ 3000 24.31@ 8500

262 40.35@ 3000 24.23 @ 8500

263 39.26 @ 3000 24.17 @ 8500

264 37.60 @ 4000 24.10 @ 8500

265 36.81 @ 4500 24.03 @ 8500

266 36.39 @ 4500 23.96 @ 8500

267 35.93 @ 4500 23.88 @ 8500

268 35.45 @ 4500 23.81 @ 9000

269 34.76 @ 5000 23.72 @ 9000

270 34.32 @ 5000 23.64 @ 9000

Table 5 shows the data for the injection timing

angle. Obviously the maximum torque and

maximum brake power are increased as the fuel

was injected earlier before reaching the Top Dead

Centre. There is no problem for the maximum

torque produced by each injection timing angle

since the engine speed is within the targeted range

except for 5⁰. However, there is a problem to

define the optimum condition since the maximum

brake power at required engine speed for all

injection timing angle are not in targeted range

Thus, more details analysis should be done on it.

Table 5: The result of various injection timing angles

Injection

timing

angle (⁰)

Maximum torque

(Nm) @ Engine

speed (rpm)

Maximum brake

power (kW) @

Engine speed (rpm)

5 33.11 @ 3000 19.61 @ 7500

0 35.46 @ 4000 22.01 @ 8000

-5 37.60 @ 4000 24.10 @ 8500

-10 39.29 @ 4000 26.07@ 9500

-15 40.38 @ 4000 27.57 @ 9500

-20 40.82 @ 4000 28.43 @ 9500


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Table 6 exhibits the detail result of brake power for

various injection timing angles(ITA).

ITA (o)

Max

BP

(kW)

@

rpm

BP (kW) @ rpm Differe

nce

(kW)

6000 6500 7000

0

22.01

@

8000

20.67 21.70 21.87 0.14

-5

24.10

@

8500

22.14 22.97 23.46 0.64

-10

26.07

@

9500

23.33 24.38 24.89 1.18

-15

27.57

@

9500

24.17 25.22 25.92 1.65

-20

28.43

@

9500

24.60 25.71 26.48 1.95

The difference between maximum brake power at

8000 – 9500 rpm and brake power at 7000 rpm

were increased as the injection timing angle gets

earlier. The angle of 00 and -5

0 will be considered

since the difference is lower and can be accepted.

On the other hand, the results for the angle from -

100 to -20

0 need to be eliminated from the

consideration because the difference is too big,

which is more than 1kW. Finally, between 00 and -

50, the injection timing angle of -5

0 is chosen as the

best angle since it produces more brake power

compared to 00.

As the conclusion, the optimum condition for the

engine operation to perform the best output is at

2640 of cam timing angle and at -5

0 of injection

timing angle.

4.1 Brake Torque Result

Figure 6 illustrates the graph of brake torque

against engine speed. The maximum brake torque

occurred at 37.6 N.m when the engine speed runs at

4000 rpm. This is for one cylinder operation. For

four cylinder engine, the total brake torque is 150.4

N.m. The value is not too far compare to engine

N34 (175 N.m) since it is a basic engine. However,

the difference with other engines such as

Mitsubishi Lancer (198 N.m) and Golf

Volkswagon (280 N.m), it is too far as they are

high performance engine.

Figure 6: Graph of brake torque against engine

speed.

4.2 Brake Power Result

Figure 7 shows the graph of brake power of the

engine during the optimum condition for various

engine speeds. The observation can be made that

the brake power achieved the maximum around

8000 rpm and sustain until 10000 rpm. From the

previous analysis, the maximum brake power of the

engine is 23.46 kW at 7000 rpm. Thus, the power

produced by the whole engine, which consist of

four cylinders is 93.84kW. The value is acceptable

since this is just a basic engine and does not equip

with other features such as turbocharged (Golf

Volkswagen) and MIVEC (Mitsubishi Lancer).

Figure 7: Graph of brake power against engine

speed (at optimum condition)

4.3 The Indicated Specific Fuel Consumption

(ISFC) Result

Figure 8 exhibits the graph of ISFC against engine

speed. Typically, a car consumes less fuel during

cruising condition (4000 to 6000 rpm). This is

because the vehicle moves under steady-state

condition.

The slope of the curve decreased and then starting

to increase from 6250 rpm to 10000 rpm. The

minimum ISFC that can be achieved for the engine

is 0.51 kg/kWh at 5500 rpm. Initially, the ISFC is

higher in order to overcome the inertia effect of the

vehicle. Besides that, it is not recommended to

drive a vehicle with high rpm as it is consuming

more fuel and not economic.


2012

59

Figure 8: Graph of ISFC against engine speed

(at the optimum condition).

4.4 The P-V Diagram

P-V diagram is the curve to show the relationship

between pressure and volume in the engine

cylinder [5]. Basically, the diagram consists of two

loops, which are upper and lower loop [5]. The

upper loop represents the power loop produced

during compression and expansion cycle.

Meanwhile, the below loop is the pumping loop,

which come from the intake and exhaust cycle.

Figure 9 shows the net power produced by

subtracting the power loop to pumping loop. From

the P-V diagram of the engine, the net power

produced is positive, since the area of power loop is

bigger than pumping loop. The positive net power

shows that the work is transfer out from the

cylinder to the crank shaft. The general idea is to

get higher net power by having bigger power loop

and minimized the pumping loop.

Figure 8: P-V diagram of the engine (at

optimum condition)

4.5 The P-Ө Diagram

The P-Ө diagram shows the relationship between

the pressure rises in the engine cylinder relative to

the crank angle [5]. It is very significant because

the data provide the highest pressure exerted on the

piston at specific crank angle. Figure 9 shows the

maximum pressure at momentary crank angle.

Figure 9: P-Ө diagram of the engine (at the

optimum condition)

In the engine, the piston is among the most critical

part that needs to be focused. For this engine, the

highest pressure exerted on the piston is 6 MPa at

170. Most of the pressure rises during the

compression and power cycle. The peak occurred

after the combustion take place.

4.6 Cycle Event

Cycle event is another important diagram for the

engine cycle that gives significant value. It

represents the process position at specific crank

angle for the engine such as the moment of valve

lift during the intake and exhaust cycle, injection

timing and etc [5]. Figure 10 illustrates the moment

for the intake valve, exhaust valve and fuel injector

to operate.

Figure 10: the event cycle diagram of the

engine

The intake valve will start open at 410 after TDC

and close at -460

before TDC in the first cycle. At

this moment, the fresh air will be induced into the

combustion chamber. Then, at -50 before TDC, the

fuel injector will start to spray the fuel. Lastly, the

exhaust valve will start open at 1250 after TDC and

close at 3980 at the next cycle.

4.7 ECU

Electronic Control Unit (ECU) is an integrated part

in the engine consists of sensors, microcontroller


2012

60

and actuators in order to assist the engine to

perform better. The basic idea is a sensor will

detect particular characteristic and send a signal to

the microcontroller. The microcontroller will give a

response according to the programmed set up

earlier. At the end, the signal will send to the

actuator to execute.

Figure 11 shows the map of the ECU. The first

sensor is the air-temperature sensor. This

component is installed in the air-intake track and

sense the air intake temperature. If the temperature

is not hot enough for the charge combust

efficiently, the EGR will be executed [6]. Thus, it

will increase the incoming air.

Figure 11: The map of the ECU of the engine

The second sensor consists of several components,

which is Manifold Absolute Pressure (MAP),

Throttle Position Sensor (TPS) and crack angle

encoder. These components will cooperate with

each other to determine the amount of fuel supplied

and injection timing [6]. So, it will give accurate

fuel needed and injection timing according to the

condition of the vehicle.

If a driver desire to accelerate, he/she will step

more on the acceleration pedal. Then, the butterfly

valve of the throttle plate (gasoline engine) and

injection pump (diesel engine) will operate since its

influence the acceleration of the vehicle.

Conventionally, the connection between pedal and

that particular component is by mechanical system,

which is Bowden cable or linkage.

5. CONCLUSIONS Finally, the optimum conditions for the engine to

perform better are at 2640

of cam timing angle for

the valve and at -50 before TDC for the injection

timing angle. The maximum torque and brake

power achieved is 37.60 Nm at 4000 rpm and 23.46

kW at 7000 rpm.

REFERENCES

[1] Mingfa Yao, Zhaolei Zheng, Haifeng Liu.,

―Progress and Recent Trends in Homogeneous

Charge Compression Ignition (HCCI)

Engines‖. Energy and combustion science 35,

398-437.

[2] T. Aoyama, et. al, ―An Experimental Study on

Premixed-Charge Compression Ignition

Gasoline Engine‖, Society of Automotive

Engineering, SAE 960081, 1996.

[3] Kohtaro Hashimoto, ―Effect of Ethanol on

HCCI Combustion‖, SAE of Japan, JSAE

20077106, 2007.

[4] Lu Xingcai, et. al, ―Experimental Study on the

Auto-Ignition and Combustion Characteristics

in the Homogeneous Charge Compression

Ignition (HCCI) Combustion Operation with

Ethanol/n-heptanes Blend Fuels by Port

Injection‖, Fuel 852622-2631, 2006.

[5] Heywood, John B., ―Internal Combustion

Engine Fundamentals‖, Massanchusetts : Mc-

Graw Hill, 1998.

[6] Hirschlieb, et. al., ―Engine Control. In Ronald

K. Jurgen. (ed) Automotive Electronics

Handbook‖. New York : McGraw Hill, 1999.

[7] Yunus A. Cengel, Michael A. Boles.(2007).

Thermodynamics : An Engineering Approach

Sixth Edition (SI Unit). s.l. : McGraw-Hill

Higher Publication.

[8] Mack, J. Hunter , Salvador M. Aceves, Robert

W. Dibble., ―Demonstrating Direct Use of Wet

Ethanol in a Homogeneous Charge

Compression Ignition (HCCI) Engine‖.

Energy 34, 782-787, 2008.


2012

61

R22 and Various Mixtures of R290/R600a as its Alternative in

Adiabatic capillary tube Used in split-type Air-conditioning

System

Shodiya Sulaimona,b

, Azhar Abdul Aziza, Henry Nasution

a,c, Amer Nordin Darus

a

aAutomotive Development Centre (ADC), Faculty of Mechanical Engineering

Universiti Teknologi Malaysia (UTM),Skudai, Johor, Malaysia.

Email:[email protected], [email protected], [email protected] bDepartment of Mechanical Engineering, Faculty of Engineering

University of Maiduguri (UNIMAID), Maiduguri, Borno, Nigeria. cDepartment of Mechanical Engineering

Bung Hatta University

[email protected]

ABSTRACT

Conventionally, (hydro chlorofluorocarbon)

HCFC22 is used as a working fluid in small vapor

compression refrigeration system with capillary

tube as expansion device. According to Montreal

Protocol, HCFC22 must be phased out owing to its

high ozone depleting potential (ODP). Several

natural substances including ammonia, carbon

dioxide, water and hydrocarbon (HC) such as

propane (HC290) and iso-butane (HC600a) and

their mixtures have immerged as close substitute.

Literature showed that pure HC refrigerant may not

be suitable enough because of the difference in

operating pressure and volumetric cooling capacity

when compared with HCFC22. The main objective

of this study is to theoretically investigate different

ratios of HC refrigerants HC290/HC600a mixtures

flowing through adiabatic capillary tube using

homogenous model. In this study, the percentage of

HC600a was varied from 0 to50 % in a step of 5%.

The pressure at the two extreme ends and

temperature along the capillary tube, using

HCFC22 refrigerant, which was used as

benchmark, was experimentally determined in the

air-conditioning (AC) system. Comparing the

model results with the experimental data showed

that HC refrigerants HC290/HC600a in ratio

70%/30% gave 2.65% minimum error and thus it

can be used as a substitute to HCFC22 in the split-

type AC system.

Keywords:Capillary tube, Split-type air

conditioner, HCFC22, HC290/HC600a,

Homogenous model

1. INTRODUCTION As a result of the increase in public concern about

the depletion of ozone layer and global warming of

some refrigerants such as chlorofluorocarbon

(CFC)‘s and HCFC‘s, a lot of refrigerants testing is

going on in search of the best alternative for these

non-eco-friendly refrigerants. For decades now, the

hydro fluorocarbon (HFC) refrigerants have been

chosen as an alternative to HCFC‘s and CFC‘s

because of their zero ODP despite their high global

warming potential (GWP). In addition, some of the

advantages associated with the HFC refrigerants‘

include the vapor pressure similarity with the

CFC‘s and HCFC‘s, their stability and non-

flammability. However, the problem of the GWP of

the HFC refrigerants has forced the scientists to

look for a more eco-friendly refrigerants such as

ammonia, carbon dioxide, water and hydrocarbons

(HC) like propane, butane, iso-butane and their

mixtures. The HC refrigerants are preferred

because of their advantages such as their zero ODP,

low GWP, non-toxic, highly miscible with mineral

oil and their higher performance when compared

with the CFC‘s and HCFC‘s refrigerants. As a

result, many refrigeration and air-conditioning

systems are using HC refrigerants and safety

precautions are made concerning the leakage of the

refrigerant from the system.

All investigations conducted using HC refrigerants

mixtures showed that these refrigerants have higher

coefficient of performance (COP) and energy

efficiency compared with the CFC‘s and HCFC‘s.

Wongwises et al. [1] investigated the behavior of

HC refrigerant mixtures of HC290, HC600 and

HC600a in automotive air-conditioning system

which was formerly using refrigerant R134a. They

concluded that all the HC mixtures yielded the

COP higher than R134a. Nasution, H [2]

conducted an experiment using a mixture of HC

refrigerants HC290/HC600/HC600a on split-type

air-conditioning system. The AC system is

formerly designed to use HCFC22 as a working

fluid. The result showed that the COP of the HC

mixture is higher than the HCFC22 by 9% and

saving of energy consumption of about 16%. Akash

and Said [3] replaced R12 refrigerant with liquefied

petroleum gas (LPG) composing of 30% R290,

55% R600 and 15% R600a in a household

refrigerator. Their result showed that the LPG

refrigerant gives the best performance compared

with R12. Likewise, Jung et al. [4] replaced R12

IMAT-UI 011



2012

62

with mixture of R290 and R600a having mass

fraction of 60% and 40% respectively. They

reported that the energy efficiency and the COP

increase by 4% and 2.3% respectively.

The expansion device used in a split-type air-

conditioner is capillary tube. This expansion device

is the heart of the system [5] because the proper

design of the tube will improve the performance of

the system and thus the behavior of the alternative

refrigerants in capillary tube needs special

attention. Many investigation have been reported,

both experimental [6-8] and theoretical [9-12] on

the characteristic flow of alternative refrigerants in

capillary tube. Undoubtedly, most of the

experimental and theoretical studies on capillary

tube used hydro fluorocarbon (HFC) like R407C

and R410A as alternative to R22 and HC

refrigerants are rarely used. As a result of the

excellent properties of HC mentioned above, the

world attention is now highly concentrated on HC

as a refrigerant.

Literature showed that pure HC refrigerant may not

be suitable as substitute to the non-eco-friendly

refrigerants due to different in operating pressure

and volumetric cooling capacity between the

HCFC‘s and the pure HC [13]. Pure HC290 has the

highest saturation pressure and pure HC600a has

the lowest saturation pressure as shown in Figure 1.

Though, the saturated pressure of pure refrigerant

HC290 is closer to refrigerant HCFC22 (Figure 1),

however, when considering their specific heat

capacity (cooling capacity), it can be seen that they

are farther apart as shown in Figure 2. Therefore,

appropriate composition mixture of HC290 and

HC600a will be required to substitute HCFC22 in

the split-type air-conditioner. The objective of this

study is therefore, to theoretically explore the flow

behavior of different composition mixtures of

HC290 and HC600a in the capillary tube using an

improved homogenous model which has not been

used by other researchers and determine a possible

composition that can replace HCFC22 refrigerant

in a split-type AC system.

-30 -20 -10 0 10 20 30 40 50 60 70

0

0.5

1

1.5

2

2.5

3

Temperature (oC)

Sa

tura

tio

n P

re

ssu

re

(M

pa

)

HC600a

HCFC22

HC290

Figure 1: Saturation pressure against

Temperature

-30 -20 -10 0 10 20 30 40 50 60 70

0.8

1

1.2

1.4

1.6

1.8

2

Tempearture (oC)

Sp

ec

ific h

eat c

apa

city (k

j/kg

-k

)

HC600a

HCFC22

HC290

Figure 2: Specific heat capacity against

Temperature

2. MATERIAL AND METHODS

2.1 Experimental work

The main aim of the experimental work is to collect

experimental data from the capillary tube in the

commercially purchased split-type air-conditioning

system which is used to compare the theoretical

model results. The air-conditioner worked with

HCFC22 as working fluid having cooling capacity

of 2.47kW. Flow meter is installed on the system to

measure the refrigerants flow rate with an accuracy

of ±0.1

g/s. Likewise, the temperature of the

refrigerant are measured at equal interval along the

capillary tube by connecting eight temperature

sensors (K-type thermocouple with an accuracy of

±0.1oC) to a computer through a data logger. The

pressure of the refrigerant is measured at the two

extreme ends of the capillary tube using pressure

transducer with an accuracy of ±0.1Mpa. The

capillary tube of length 1.30 m and diameter

1.327mm which is from the manufacturer of the

AC was used.

To collect the experimental data, the refrigerant

HCFC22 was charged into the system after the air

in the system has been evacuated using vacuum

pump. When the system is on, it takes about 30 to

45mins before the system could attain steady state.

The pressure of refrigerant at the two extreme ends

of the capillary tube and temperature along the

capillary tube is displayed on the computer. The

experiment is repeated five times and the average

value is taken.

2.2 Model Description

The detailed description of the two-phase

homogenous model is presented in [14]. As shown

in Figure 3, the capillary tube is connected between

the condenser and evaporator. The flow in the

capillary tube is modeled by dividing the flow into

subcooled liquid single phase, metastable phase


2012

63

and two-phase liquid-vapor regions. The flow in

the capillary tube is based on one dimensional

homogenous two-phase flow assumptions. The

model is also based on the fundamental equations

of mass, momentum and energy and that of

Wongwises et al. [10] with modifications.

1 2 3 4 5Condenser Evaporator

Subcooled liquid

region

Metastable

region

Two-phase

region

Capillary tube

Figure 3: Adiabatic capillary tube with three

regions

2.2.1 Single-phase flow region

The liquid single-phase length is given

in eq. (1)

Where, p1 and p3 are pressures at points 1 and 3

respectively, G is mass flow per unit area , D is

inner diameter of capillary tube, k is coefficient of

entrance loss and its value is 1.5 as given by Zhou

and Zhang, [15]. Single-phase friction factor is

calculated from Colebrook and Churchill formula,

given in eq. (2) and (3) respectively.

Where

and are wall roughness and refrigerant viscosity

respectively

2.2.2 Metastable flow region

The inception of vaporization does not take place at

the real saturated vapor pressure Ps, but actually

takes place at a later point Pv, downstream from

thermodynamic saturated position. This process is

called metastability. Chen et al. [16], quantified this

metastable flow and propose a correlation, given in

eq. (5), to predict this delay in vaporization.

Where =

is level of subcooling, is critical

temperature, are pressure at points 3 and

4 respectively and is temperature at point 3.

and are specific volumes of liquid and

liquid/vapor phases respectively.

The length of metastable is given by

Where metastable friction factor, was

evaluated using Colebrook formulation as shown in

eq. (2).

2.2.3 Two-phase region

The differential equation of the two-phase length of

the capillary tube is given in eq. (7)

The two-phase friction factor, can be

conveniently evaluated using Bittle and Pates [17]

correlation given in eq. (8)

Where

The viscosity models used in calculating the two-

phase viscosity are given by Cicchitti et al.[18],

Dukler et al.[19], McAdam et al. [20] and Lin et al.

respectively as follow:

To evaluate the elemental two-phase length of the

capillary tube, eq. (7) is discretized given in eq.

(14)

Summation of the elemental lengths results in the

two-phase length given in eq. (15)

Thus, the total length of the capillary tube, L, can

be written as

All thermophysical and thermodynamics properties

are taken from the REFPROP [21] computer

program, version 8 (2007) which are developed in

the function of pressure.


2012

64


In this section, model verification is made and

using the model, various ratios of the HC mixtures

are compared with the standard refrigerant,

HCFC22. The model operating parameters and

their ranges are as follows: pressure inlet, 0.4 – 2.0

Mpa, mass flow rate, 0.005 – 0.018 g/s, capillary

tube diameter, 0.20 – 1.8 mm, degree of

subcooling, 1.0 – 12.0oC.

The present model results gave a good agreement

when compared with the experimental data of

Fiorelli et al. [22] with respect to the shape of

pressure distribution as shown in Figure 4.

Although, Colebrook‘s and Churchill‘s

formulations were used to determine the single-

phase friction factor and the three viscosity models

mentioned in this study was also used in the

simulation, however, the Colebrook and Dukler et

al. correlations‘ combination gave the best

prediction with an average error of 1.75%.

0 0.5 1 1.50

0.2

0.4

0.6

0.8

1

1.2

1.4

1.6

1.8

2x 10

6

D i s t a n c e f r o m C a p i l l a r y I n l e t ( m )

C a

p i l l a

r y

T

u b

e P

r e

s s

u r

e (

P a

)

N u m e r i c a l r e s u l t

E x p e r i m e n t a l d a t a o f F i o r e l l i e t a l .

Figure 4: Comparison of the numerical results

with experimental data of Fiorelli et

al. [22].

3.1 Comparison of Different ratios of the

HC with the Model The developed model was used to simulate

different ratios of HC290 and HC600a. The

percentage of HC600a was varied from 0 to 50% in

a step of 5% and comparison with experimental

data using refrigerant HCFC22, the benchmark

refrigerant, is reported. From the results obtained,

the best composition lies between 75% propane,

25% iso-butane (HC1), 70% propane, 30%

isobutene (HC2) and 65% propane, 35% iso-butane

(HC1).

The sample of these HC mixtures and the

experimental data with respect to pressure

distribution along the capillary tube, up to choked

condition, is shown in Figure 5. From the Figure, it

can be seen that the rate of decrease in pressure of

HC1 is higher than HC2 and HC3. This could be

attributed to the fact that the specific volume and

velocity of HC1 is higher as a result of lesser

amount of HC600a present in HC1 compared with

HC2 and HC3. Consequently, the mach number of

HC1 is increased as a result of its high velocity,

leading to its earlier choke flow compared with

HC2 and HC3. From the Figure, refrigerant HC1,

HC2 and HC3 at pressure 2.2, 2.5 and 2.8 bar have

their choked flow at length 1.30, 1.32 and 1.34 m in

the capillary tube respectively. Therefore, pressure

drop per unit length for HC1, HC2 and HC3 are

1.62, 1.92 and 2.10 bar/m respectively for a given

mass flow rate and capillary tube inner diameter.

Thus, HC2 refrigerant is the best composition,

having pressure drop per unit length of 1.92 and the

minimum error of 2.78% when compared with the

benchmark refrigerant, HCFC22.

0 0.2 0.4 0.6 0.8 1 1.2 1.4

0

2

4

6

8

10

12

14

16

18

x 10

5

Pin = 1.636 Mpa

Tsub. = 4.5oC

m = 25g/sec.

D = 1.752 mm

E=2.592*10-4

D i s t a n c e f r o m C a p i l l a r y I n l e t ( m )

C a

p i l l a

r y T

u

b e

P

r e s s u r e

( P

a

)

HC1

Experimental data

HC2

HC3

Figure 5: Comparison of various HC mixtures

with the experimental data

The variation of mass flow rate of HC1, HC2, HC3

and HCFC22 with respect to pressure inlet for

diameter 2.00 mm and inlet subcooling of 10.5oC

under choked flow condition is shown in Figure 6.

It can be seen from the Figure that HC1 has the

lowest mass flow rate for the flow condition and

HC3 has the highest mass flow rate. However, HC2

and HCFC22 have almost the same mass flow rate

confirming that HC2 can be substituted for

HCFC22.

4.92 4.94 4.96 4.98 5 5.02 5.04 5.06 5.08 5.1 5.12

x 10

5

0.009

0.01

0.011

0.012

0.013

0.014

0.015

0.016

0.017

L = 1 . 3 0 m

T s u b . = 1 0 . 5 o C

D = 2 . 0 0 mm

E = 2 . 5 9 2 * 1 0 - 4 m m

P r e s s u r e I n l e t ( P a )

M a

s s f l o

w

r a

t e

( K

g

/ s )

HC1

HC2

HC3

HCFC22

Figure 6: Comparison of refrigerant mass flow

rate of HC1, HC2, HC3 and

HCFC22 with pressure inlet at inlet

subcooling 10.5oC and diameter 2.00

mm

The comparison of mass flow rate of HC1, HC2,

HC3 and HCFC22 with respect to pressure inlet for

diameter 1.742 mm and inlet subcooling of 4.5oC

under choked flow condition is shown in Figure 7.


2012

65

Figure 6 and 7 are similar in respect of their graph

profile, however, the respective mass flow rates of

Figure 6 are higher than Figure 7. This can be

attributed to the increase in their subcooling and

diameter of capillary tube. A higher subcooling will

have more refrigerant in liquid phase in the

capillary tube and liquid phase refrigerant offers

lesser resistance to flow. Figure 7 also reveals that

the mass flow rate ratio of HCFC22 to HC1 is

about 0.51, with HC2 is about 0.92 and with HC3

is about 0.49.

4.92 4.94 4.96 4.98 5 5.02 5.04 5.06 5.08 5.1 5.12

x 10

5

0.006

0.008

0.01

0.012

0.014

0.016

0.018

P r e s s u r e I n l e t ( P a )

M a

s s f l o

w

r a

t e

( K

g

/ s )

L = 1 . 3 0 m

T s u b . = 4 . 5 o C

D = 1.752 mm

E=2.592*10-4

HC1

HC2

HC3

HCFC22

Figure 7: Comparison of refrigerant mass flow

rate of HC1, HC2, HC3 and

HCFC22 with pressure inlet at inlet

subcooling 4.5oC and diameter 1.742

mm

4. CONCLUSION

In this present work, a theoritical model that was

developed was used to explore the behavior of HC

HC290 and HC600a refrigerant mixtures, in order

to determine the possible HC composition ratio that

can replace HCFC22 refrigerant in the split-type air

conditioner. Different ratios of HC HC290 and

HC600a was simulated and the results compared

with the experimentally determined data using

refrigerant HCFC22, the bench mark. The result

showed that HC2 (70% HC290 and HC600a 30%)

refrigerant is the best composition, having pressure

drop per unit length of 1.92 and the minimum error

of 2.78% when compared with refrigerant

HCFC22. The mass flow rate ratio of HCFC22

with HC2 is about 0.92 signifying their closeness.

With the result obtained, it can be concluded that

HC2 refrigerant is a close substitute to HCFC22 in

the split-type air conditioner.

ACKNOWLEDGMENT

The present study was financially assisted by

research university grant (RUG) program: Tier 2

Cost Centre Code: QJ130000.7124.01J85,

Universiti Teknologi, Malaysia. The financial

support of Educational Trust Fund (ETF), Nigeria

is also acknowledged.

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Goncalves and M. M. Mezavila, "An

experimental analysis of adiabatic capillary

tubes," Applied thermal engineering, vol. 19,

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[6] H. Wijaya, "Adiabatic capillary tube test data

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[7] S. M. Sami and H. Maltais, "Experimental

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Mohan Lal, "Flow characteristics of

HFC407C/HC600a/HC290 refrigerant

mixture in adiabatic capillary tubes," Applied

Thermal Engineering, vol. 26, pp. 1621–1628,

2006.

[9] S. Wongwises and W. Pirompak, "Flow

characteristics of pure refrigerants and

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[10] S. Wongwises and S. Chingulpitak, "Effects

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and Mass Transfer, vol. 37, pp. 1305-1311,

Nov 2010.

[11] S. M. Sami and H. Maltais, "Numerical

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1359–1371., 2000.


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[12] S. Shodiya, A. A. Azhar, N. Henry and A. N.

Darus, "New Empirical Correlations for

Sizing Adiabatic Capillary Tubes in

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propane/commercial butane mixtures as

possible alternatives to R134a in domestic

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[14] S. Shodiya, A. A. Azhar and A. N. Darus,

"Improved Refrigerant Characteristics Flow

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"A correlation for metastable flow of

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Saldaini and R. Zavattarelli, "Two-phase

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Review paper on:

Flow Pattern at Pipe Bends on Corrosion Behaviour of Low

Carbon Steel Muhammadu Masin Muhammadu

1 , Jamaluddin Muhammad Sheriff

2 and Esah Binti Hamzah

3

1Department of Mechanical Engineering, Federal University of Technology, Minna

2,3Faculty of Mechanical Engineering, Universiti Technologi, Malaysia

[email protected], [email protected] and [email protected]

ABSTRACT

Most importantly the identification of positions or

sites, within the internal surface contact areas

where the maximum corrosion stimulus may be

expected to occur, thereby allowing better

understanding, mitigation, monitoring and

corrosion control over the life cycle. Some case

histories have been reviewed in this context, and

the interaction between corrosion mechanisms and

flow patterns closely determined, and in some cases

correlated . Since the actual relationships are

complex, it was determined that a risk based

decision making process using selected ‗what‘ if

corrosion analyses linked to ‗what if‘ flow

assurance analyses was the best way forward.

Using this in methodology, and pertinent field data

exchange, it is postulated that significant

improvements in corrosion prediction can be made.

This paper outlines the approach used and shows

how related corrosion modelling software data such

as that available from corrosion models Norsok

M5006, and Cassandra to parallel computational

flow modelling in a targeted manner can generate

very noteworthy results, and considerably more

viable trends for corrosion control guidance. It is

postulated that the normally associated lack of

agreement between corrosion modelling and field

experience, is more likely due to inadequate

consideration of corrosion stimulating flow regime

data, rather than limitations of the corrosion

modelling.

Keywords: ALARP (As Low As Reasonably

Practicable) Co2 corrosion,

corrosion resistant alloy (CRA),

decision gates, erosion-corrosion,

life cycle performance, risk basis

1. INTRODUCTION

The situation of flow separation, for example inside

a sudden expansion in the pipe, turbulence is

moved downstream from the objective of

separation (1). There is no simple relation

involving the bulk flow parameters as well as the

local near-wall, hydrodynamic, mass transfer and

erosion-corrosion conditions as well as the latter

ought to be determined either experimentally [1,8,

13] or by record simulation both by [1,8,13]. This

paper focuses the bond between flow pattern and

corrosion behaviour at pipe bends as well as the

advances in using turbulence models for the record

simulation of erosion-corrosion throughout

yesteryear decade.

The development of water pipeline infrastructure in

a few part of World remains rapid growth within

the last few years. If you have been new projects

plus much more tie backs into existing systems.

This helps it be crucial that you acknowledge,

identify, and develop, the critical associations

between corrosion conjecture and flow systems at

pipe bends. Most substantially the identification of

web sites where maximum corrosion stimulus can

be expected thus enabling better understanding,

minimization, monitoring and corrosion remedies

for the whole existence cycle. Some situation

histories are actually examined in this particular

context, as well as the interaction between

corrosion systems and flow designs carefully

determined for a number of design campaigns. The

specific associations are complex, too for basic

reasons some risk based making choices procedure

using ―what if‖ corrosion analyses connected with

―what if‖ flow assurance analyses was considered

our advice [15-23].

The experience based on formerly looked into work

by [24] and [25] has generated an instantaneous

final results of corrosion rate of low carbon steel

loss and fluid flow designs. A modification of flow

pattern can lead to significant improvements in

tube existence by remaining from impinging flow.

The mechanism in the internal corrosion rates aren't

fully understood but appears being linked strongly

towards the particulate matter inside the flow [24-

29]. Even if pollutants are simply inside the

micrometre size range, they could still deviate from

fluid streamlines to make sure that glancing flow

can lead to abrasion in the protective surface layer

[31-37]. Rapid positioned on by corrosion-erosion

will result when the layer is soft in compliance

while using particulate matter.

By using this methodology, and area data

exchange, it's thought significant enhancements in

IMAT-UI 012




2012

69

corrosion behaviour are achievable. This paper

outlines the approach used and shows how relevant

corrosion modelling software data for example that

supplied by corrosion models available Norsok

M5006, and Cassandra Software [15,30] to parallel

computational flow modelling within the specific

manner can generate very significant results, and

sometimes unforeseen trends for corrosion control

guidance. The methodology offers good semi-

quantitative arguments for justification and a

variety of references are really used.

Having considered the basic characteristics of flow

in curved open channels are fairly understood,

theoretically and numerically the configuration of

the side wall has significant effects on the flow

structures [29].

Since the streamlines ought to be curved inside the

same sense since the pipe itself there is a radial

pressure gradient, so pressure is bigger round the

outer wall in the pipe compared to the related point

round the inner wall.

The mechanism which supplies birth for the

stagnation pressure gradient is not considered

incorporated within the secondary flow

phenomena, although there's evidence of mutual

interaction involving the two inside the turning

passage to ensure that as result causes turbulent

which in turns involve some unwanted effects

round the pipe bend.

This paper is tried to give consideration towards the

approaches and techniques familiar with identify,

develop, and verify critical associations between

corrosion behavior and flow programs systems.

Used this frequently reduces with a practical

interpretation in the outcomes of flow on corrosion

turbulence systems.

2. Turbulence Models and methods

The task of turbulence models is always to provide

equations that will enable calculation in the

Reynold stresses,liuj, as well as the turbulent

diffusion fluxes,lm5uj, which arise when the time-

averaged equations for turbulent flow and mass

transport are acquired within the immediate

equation [14]. The k- , turbulence models [15]

which are currently [16] broadly useful for the

computation of business flows are eddy viscosity

model which be a consequence of the concept

recommended by Boussinesq and assumes caused

by turbulence round the mean flow can be

considered through viscosity. The turbulent

viscosity, µ,t is made the decision within the kinetic

energy of turbulence, k, which is rate of

dissipation,

µt = Cµ ƒµ ( k2)/ ) (1)

The effective viscosity is provided by

Deff = µ + µt (2)

similarly the effective diffusivity is given by

Deff = µ/ t + µt/ (3)

Where is the turbulent S chmidt number

[13].The conservation equations for mass,

momentum, kinertic energy of turbulence which is

dissipation, and species, m, might be witten in the

general form. For axisymmetrical flow in 2D round

co-ordinates [8]:

( ) + ( ) = ( ) + (r

) + (4)

Where

The values of , general diffusion coeffients and

, the source terms.

Flow separation with recirculation and

reattachment for example inside a sudden

expansion and understanding in the concentration

area enables the rate of mass transfer being

calculated [1,13] with the expansion. The

calculation in the concentration area near the wall

requires utilizing a low Reynolds number [LRN] k-

model since the mass transfer boundary layer is

deeply embedded within the viscous sublayer. The

concentration is required at y ~ .1, insidewithin all

the viscous sublayer where the mass transport is

diffusion controlled, to have the ability to calculate

the wall mass transfer rate. The 2nd requirement is

because of the bit of turbulence inside the vjscous

sublayer obtaining a significant effect on mass

transfer within the high Schmidt amounts Sc ~

1000 frequently familiar with aqueous mass

transfer. Low reynold number (LRN) models

utilize turbulence damping functions [13,32] as the

easiest way of modelling wall-boounded flow with

warmth/mass transfer under separated flow

conditions [9]. The option wall function (WF)

closure places the initial computation mode inside

the logarithmic law region (30 < y+ < 150) and

bridges over the important viscous sub-layer.

Furthermore, the motion of a dispersed particulate

phase within a turbulent flow field can be modelled

by either a Lagrangian or Eulerrian approach [34].

He further lamented that in Lagrangian models a

large number of individual particle trajectories are

calculated in the flow domain whereas in the

Eulerian approach the particles are treated as a

second fluid. From an erosion modelling standpoint

the direct calculation of particle/wall interaction

statistics, impact frequency, angle and velocity with

the Lagrangian approach is an advantage: at least in

dilute particulate suspensions. As pointed out by

Nesic [8], the Eulerian approach would be more

appropriate in concentrateed suspensions.

Also [10] review and revealed that particle

breakage may cause quality control problems,

whereas wall erosion increases equipment

maintence costs and environmental burden, and

causes loss of productivity and a requirement to

replace damaged components. Under normal

operating conditions, erosion rates in pipe bends

are much higher than those in straight pipe sections

due to local turbulence and unsteady flow


2012

70

behaviour. Lamented that therefore, the time for

failure piping systems is often dependent on the life

of pipe fittings (valves, bends, elbows, flow

metres), flow pumps, turbines, and compressors

[56-59]. It is, therefore imperative to improve the

protection of these components from solid particle

erosion by better understanding the physics of

particulate flows.

Many investigators [58-61] have carried out both

physical and numerical modelling of the erosion of

pipe bends, elbows, tees and related geometries.

Since the early 1990‘s, computational fluid

dynamics (CFD) has been widely used for solid

particle erosion prediction in curved pipes and

ducts, with various analytical, semi-empirical and

empirical models having been developed. In 1979

[67] provided a critical review of some of the

erosion models that had been developed since

Finnie (1960) proposed the first analytical

approach, and found 28 models that were

specifically for solid particle-wall erosion. The

authors [66,67] reported that 33 parameters were

used in these models, with an average of five

parameters per model. These parameters influence

the amount of of material eroded from a target

surface and the mechanism of erosion. The review

revealed that each model equation was the result of

a very specific and individual approach, hence it is

clear that no single equation exists that can be used

to predict wear from all known standard material or

particle parameters, and that some reliance on

experimental measurement will always be required

to provide empirical constants necessary in the

various erosion models. The following review is

limited to models that have been used in CFD-

based erosion modelling and which have received

wide usage in applications to erosion in pipes bends

and pipe fittings.

In 1987 [69] developed an empirical erosion model

for AISI 1018 steel. Mclaury (1993-1996) extended

this model for aluminium and used it to predict

particle erosion resulting from both direct and

random impingements in two-dimensional

geometries. This erosion model was developed at

the Erosion-corrosion Research Centers. In 1996

[71] predict erosion in elbows, plugs, tees, sudden

contractions, and sudden expansions for pipes with

circular cross-sections, and was subsequently

referred to as erosion -corrosion Research Centres

models. Also in 1973 [52] applied this model to

investigate the effects of elbow radius of curvature

on erosion rates in circulate pipe. Meanwhile,

[60] developed mechanistic models for predicting

erosion in elbows based on the E/CRC model. In

1998 and 2001, [67] and [68] used a commercial

CFD code to model fluid-solid flow and added

routine to predict erosion on particle impact using

the E/CRC models. The [67] and [68] also

modelled the erosion in oilfield control valves

using a commercial CFD code, accouting for both

deformation and cutting erosion. They obtained a

good agreement for the predicted wear rates and

wear locations in pipe bends with the experimental

gas-solid erosion results of Bourgoyne, added [63

and 64] the stochanstic rebound model of [61 and

62] and the E/CRC model [56] to a commercially

available CFD code and investigated the relative

erosion severity in elbows and plug tees found in

oilfield geometrics. Numerical simulations showed

that particle rebound behaviour played an important

role in determing the motion of the particle [65 and

71].

The performed their erosion research on 90 degree

elbows and bends of circular cross-section. The

fluid phase was modelled using a simple modefied

mixing-length model. The predicted fluid axial

velocity was validated against the experimental

data of [56], with erosion modelled using the

E/CRC model [68]. They compared their predicted

penetration rates with the experimental data of [65],

obtaining good qualitative agreement but poor

qualitative agreement between the predictions and

data. The poor agreement occurred because most of

the data available were from erosion experiments

with high particle rates. The authors also found that

erosion in long radius bends was reduced when the

carrier phase was changed from liquid to gas. They

further reported that the effect of the squeeze film,

secondary flows and turbulent flow fluctuations

may all play important roles in erosion prediction

when the carrier fluid is a liquid.

In 1991 studied [4]the local erosion in chokes and

determining the local fluid velocity and particle

impingement. Also, in 2006 [21] used a

commercial CFD code coupled with an in-house

particle tracker to predict fluid particle flow in a

full 180 bend. He further implemented two erosion

models, namely those of finite (1961) and [ 1 and

65] investigated erosion-corrosion problems in U-

bends. However, there was no comparison was

made with experimental data ; hence, the validity

of the model could not be ascertained. Attempted to

account for the shape of wear scars in predicting

the life of pneumatic conveyor bends undergoing

erosive Wear. However, these Authors did not use

the shape of the scar to alter the computational

mesh used in the fluid phase calculations.

Fan and co-workers [55] used Eulerian-Lagrangian

approaches with the empirical restitution

coefficient of [63] to model particle rebound

velocities, and subsequently employed the semi-

empirical erosion equation of [64] to study erosion

in a vertical-to-horizontal bend. The authors used

this technique to predict the erosion of tube banks

in heat exchangers (65 and 66), to study protection

techniques against tube erosion (67), and to

investigate anti-erosion in 900 bends (68, 69, 70

and 71). In all the studies, the authors used the

standard k- turbulence model, except in [62]

where the authors used large eddy simulation LES


2012

71

to obtained flow and turbulence field predictions.

The LES solution was not, however, validated

against previous experimental or numerical results.

In 1994 and 2004 [69 and 70] also used the semi-

empirical erosion equation of [63] in 1998 to

estimate the erosion rate in tube banks.

In 2007 [70] conducted a number of CFD-based

erosion modelling investigations using a

commercial CFD code. The erosion equation

developed at the University of Tulsa (Albert, 1994;

Edwards et al, 2001; McLaury 1993, 1996;Mclaury

and [49] in 2000; [50] in 1995; [64] in 1995 and

[54] in 2006, were included in the through user-

defined-function. Particle trajectories were

validated against the authors‘ experimental data for

liquid-solid flows obtained using laser diagnostics.

The entire CFD -based erosion authors‘ modelling

procedure was then validated by comparison with

the authors‘ experimental data obtained in 900

standard elbows with air flows, measured using a

sensitive electrical-resistance probe. In 2009 [66]

investigated particle motion in the near-wall region

using a commercially available CFD code,

modifying the code to account for particle size

effects in this region before and after particle

impact. For turbulent flow in a 900 bend, their

results showed that the near-wall modifications and

turbulent particle interactions significantly affect

simulation results when compared with

experimental data.

In 2009 [66] applied an Eulerian-Lagrangian

approach with particle-particle interaction and a

erosion model to simulate solid particle movement

as well as the particle erosion characteristics of a

solid-liquid two-phase flow in a choke. The authors

used the standard k- model to treat the turbulence,

the discrete particle hard sphere model to

accommodate inter-particle collisions and the semi-

empirical correlations of [65] in 1979 to study anti-

erosion effects.

Despite all this work, there is continued interest in

pipe wall erosion modelling because the prediction

of erosion, in particular, is of value in estimating

the service life of pipe bends systems, as well as in

the identification of those locations in a particular

pipe geometry most prone to erosion. In this

review, [55] a study on three-dimensional

computational fluid dynamic model of erosion is

developed to investigate the erosion of both the

concave and the convex walls of cross-sectioned

ducts of different bend geometries and orientations

due to particle collisions with the wall surfaces.

Results are discussed in terms of eroded depth and

the location of primary and secondary wear, and

are compared with available experimental data

[67].

In practice after corrosion risk appraisal, this can be

reduced to:

Total CA = Uniform CA + erosion allowance +

localized (pitting) allowance. The

predictive corrosion rates can be based on the

predicted temperature profiles and the published

top of line corrosion correlations for example, the

Nyborg-Dogstad correlation [69]. There are few

published such [equation 5] can be used on an

iterative basis, to determine top of line activity

[50,67].

Corrosion rate = 0.004*Rc*Cfe*(12.5 - 0.09*7) (5)

Where:

Corrosion rate: given in mm/y, Rc: Condensation

rate g/m2/s, Cfe: Saturation iron levels in

condensation (difficult variable, 50 - 200ppm often

select), and T: Temerature in degree centigrade (0c)

respectively.

3. Flow Accelerated Corrosion

The key factor factor response to determine flow

faster corrosion (FAC) might be the oxide film on

structure surfaces, which evolves consequently of

corrosion and, simultaneously, controls the

corrosion rate within the role like a protective film

[47]. The primary parameters to discover FAC they

fit into material parameters, flow dynamics

parameters and atmosphere parameters (48).

Metallic ions, mainly ferrous ions (Fe2 ), are

released for the water within the boundary layer

where a number of within the supersaturated Fe2

ions become oxide pollutants and in addition they

deposit over the metal surface being magnetic

oxide layer [48]. The oxide layer plays an

important role in preventing further relieve Fe2

(corrosion reaction). The thickness inside the

boundary layer is impacted by flow dynamics of

people processes, oxygen concentration (O2) inside

the boundary layer plays an important role for

oxidizing magnetite to hematite, which contributes

to achieving much greater corrosion resistance

[49].

Generally, subsea pipelines flow assurance covers

all multiphase transport phenomena. Diligent

design techniques, understanding and capabilities

are very important to make sure safe, continuity of

fluids transport from reservoir to topsides

processing plant. The main areas involve steady

condition and transient multiphase flow hydrates,

sand, oil, emulsions, wax, scale and corrosion

phenomena. The interaction between

corrosion/scaling and flow assurance can therefore

be instrumental in determining true production

rates, and software packages such as the Scand

energy or Non-destructive test types codes can

certainly give reliable predictive modelling.

Delivering an exact and reliable advantages of flow

assurance and corrosion modelling is regarded as as

as one of the primary challenges facing the today.

The overriding performance fingerprint is often

best known to as tub curve and also to help

facilitate this better, corrosion needs to be

recognized to love an operating hazard. Once that's


2012

72

recognized the benefits of a soundly planned and

faithfully applied corrosion management strategy

becomes self apparent along with a pronounced

reliance upon important water subsea infrastructure

and tie back [25,27, and 30].

The beginning of issues with corrosion integrity

throughout early existence is important, though mid

existence is often better handled since area

existence is extremely frequently well below design

existence and for your reason options and time can

be expected to favour planned retrofit as needed.

Existence extension beyond the situated on out

zone is generally more tightly related to older

researches whereupon original design life‘s are

actually exceeded and continuing production

needed [28,31].

4. Enabling Presumptions

Uncover the dominant corrosive species usually

CO2, with defined or understood water chemistry

content.

Review, verify, and prioritize or assume the

dominant flow programs, e.g. single or multiphase

flow, stratified, slug, annular flows etc. forecasted

while using existence cycle. This might require

review of the steady condition and non-steady

activities or transient flow situations. Frequently

the flow assurance report examines the likely

situations and phone connection, and with this

judgment [37].

Validate the H2S souring inclination because this

greatly impacts cracking and corrosion behaviour.

As being, this is often less influenced by flow

programs provided the most effective scales remain

intact.

Corroborate dissolved chloride and oxygen levels,

and the existence of aggressive organic species for

example acetic/formic chemicals additionally for

their types.

Assume the bottom corrosion is uniform but

validated by real-time inspection and monitoring

and via analysis of coups/probes, and deposits.

However anticipate to witness localized corrosion

when unsteady conditions occur.

Confirm threat and chance of biofilm formation

assuming pattern of growth follows laminar or fluid

stagnation sites, identify microbe species and MIC

within the existence cycle.

Corroborate the extent of sand production (steady

and episodic), along with the effect on materials

degradation. Determine sand concentration, and

particulate dimensions, and review interactions

with inhibitor performance [53].

Determine the role of small-stagnant zone

corrosion under primary flow conditions per small

locations where pollutants and biofilms may

proliferate [37].

Establish extent of pigging and inhibitors to both

create sustainable inhibitor films and repair such

inhibitor films even at high flow rates under

sanding nd erosion conditions [54].

5. Impact of Temperature and Flow

Temperature and flow regime are carefully linked

since CO2 corrosion is dynamics and very mindful

to electro-chemical and physical fluctuations (for

example changing P, T, V). Generally steady

condition (P, T, V) Conditions frequently promote

protective film compaction as well as for your

reason passivation, and low corrosion rates. Lower

temps < 1200F (50

0C) tend to promote patchy

corrosion with softer multi-layer iron carbonate

scales providing some barrier protection increasing

up to 140 to 1600F (60-70

0C). Above these

temperatures damaging localized corrosion is

observed as films lose stability and spall off giving

rise to galvanic ‗mesa‘ attack. Though there is

evidence of a down turn in the plateau after 800C

for certain cases. In reality the project design basis

usually insists on a maximum value for

temperature, as it does for other critical parameters

such as pressure, materials characterization (yield

stress, hardness, toughness, etc.). Regarding flow,

the production rates can be influenced by flow

regimes such as slug flows and annular gas flows.

This can prove critical for vertical risers connecting

the flowline to the offshore structure or topsides.

The geometry can act almost like a ‗specification

break‘ whereupon the flow regime shifts largely

due to the effects of gravity as the flowline

transforms from a horizontal to a vertical member.

The sag bend at the touch down zone can become a

high risk corrosion component warranting greater

degree of corrosion control, such as a thicker

section, increased local CA, internal coating epoxy

or increased monitoring routines etc.

Furthermore, as lamented earlier most likely

probably the most challenging conjecture is less

whether localized corrosion will occur but much

more likely at this point you request , generally

where? An thorough study on the expected flow

regime curves and maximum flow velocities might

help help with that judgment [25, 26]. It's also

likely that whenever flow line localized corrosion

starts, then it is more probably being self

propagating (auto catalytic) and is less influenced

by modifications inside the majority conditions

though more jobs are needed in this region (42,45

50).

For operating profiles examined in that way, for

example stratified flow, annular flow, slug flow

and bubble flow, it should be assumed that essence

turbulent as with line using the typical flow line

Reynolds number, but realize that within each flow

regime you will observe complex inter-facial

behavior, including laminar, stratification and

turbulence. The level of smoothness of people

activities isn't necessarily expected, however


2012

73

reliable online monitoring results might help

minimization and control, given to seize control of

the feelings at a great choice.

6. Internal Corrosion Direct Assessment

Harmful internal corrosion in addition to failures

have happened on pipelines carry gas /liquid

specified being dry [37]. The process referred to as

internal corrosion direct assessment (ICDA)

remains designed to look at the corrosion impact of

short-term upsets on pipeline integrity. The process

is anticipated to boost pipeline integrity, reliability,

and public safety. ICDA might come to terms with

enhance the assessment of internal corrosion in gas

transmission pipeline and help ensure pipeline

integrity. The process is primarily positioned on

gas transmission lines that normally carry dry

gas/liquid but they're affected from temporary

upsets of wet gas/liquid water (or electrolyte). The

understanding basis must be readily transferrable

together with other media liquids.

7. Single phase flow

An low Reynolds number (LRN), k-€ model has

lately been positioned on the calculation of pipe-

wall mass transfer rates in the sudden change or

expansion [5], an immediate constriction [1] and

flow round the groove [8]. They pointed out, in

addition mass transfer rates are really calculated for

that pipe wall in the sudden bend where small

patches in the protective ‗rust‘ film were assumed

to possess been removed. The above mentioned

pointed out stated results indicate the mass transfer

regions of erosion- corrosion processes in flow

pattern conditions may be satisfactorily laboured

with through turbulence models.

A much more intractable issue is an chance to

calculate protective film removal under single

phase aqueous flow conditions. The partial

elimination of a protective surface film is

frequently the precursor to rapid corrosion and

component failure. For instance failures in copper

piping in apartment structures frequently connect to

rapid corrosion in the sudden difference in the

geometry in which the normally protective film

remains broken using the enhanced turbulence.

Recent findings have proven that although point

about this kind of corrosion happens near the outlet

of 900 bends the film breakdown and subsequent

failure began inside the sudden step in which the

downstream pipe was soldered towards the elbow.

The idea of an essential shear stress for eliminating

protective layers remains asked for because the

small stresses involved wouldn't be sufficient to

robotically remove a surface oxide film. As pointed

out by Launder, B. E., 1998 pressure fluctuations

unlike velocity fluctuations don't vanish inside the

wall. In 1991, [4] has recommended that pressure

fluctuations inside the wall, in flow pattern, raise

the overall shear stress and could produce

mechanical damage.

8. Liquid/Solid flow

The presence of solid particles enhances the

destruction of protective films giving rise to

increased corrosion rates and may add to the

overall metal loss by the mechanical erosion of the

underlying metal. As with single phase flow these

destructive effects are more pronounced under flow

pattern condition.

In 1991 [4] have developed a predictive model for

localized erosion- corrosion under flow pattern

conditions based on the application of a two phase

flow version of an low Reynolds number (LRN), k-

model of turbulence. The motion of the particles

was predicted by means of a Lagrangian

Stochastic-Deterministic (LSD) model. The model

which was applied to various pipe geometries

including a sudden expansion, constriction and a

groove was based on an oxygen-mass-transfer

controlled corrosion model with the assumption

that the particles removed the protective rust film,

and an erosion model based on the cutting wear

erosion equations [58].

The successfully applied a two phase k- model to

the numerical simulation of uniform CO2 erosion-

corrosion under separated flow conditions at a

sudden expansion. The particles were modeled by

an Eulerian approach [62].


2012

74

9. Summary of previous works

Year(s) Researcher(s) Method(s) Flow Pattern

Reynolds

number

Size(D=depth/L=length)

1966 Burgraff O. Numerical:

modified

relaxation

Up to 400 Square bend (D/L = 1)

1979 Benjamin A., et

al

Numerical:

ADI Scheme

for linear

system and

Newton-like

method for

non-linear

system

Up to 10000 D/L =1 and 2

2004 Yao H., et al Numerical:

Finite

difference for

unsteady 3D

incompressible

Nevier-Stoke

equation

1000 to 10000 D/L = > 0.1

2005 Cheng M., et al Numerical:

Lattice

B0ltzmann

Range of 0.01

to 5000

D/L range from 0.1 to 7

2006 Thierry M., et

al

Experimental:

Doppler

Velocimetry

1150 < Re <

10670

0.5 < D/L = 2

2009 Ozalp A., et al Experimental:

PIV

1230, 1460, and

1700

Rectangular, Triangular

and Semi-circular with

D/L = 2

2011 Muhammad R.

M.

Experimental:

PIV ( tape as

fluid) and

Numerical: 2-

D fluent

23144, 32963

and 39275

D/L =extend surface,

X = 0, 5, 10, 15, 20cm

2012 Muhammadu

M. M.

Experiment:

PIV (Seawater

as fluid) and

Numerical: 3-

D fluent

On going On going

Corrosion Behaviour

1960 Bitter J. G.A. 2-D model the result could not ascertained

1993 to

1996

Fan J., et al E/CRC and

Eulerian-

Lagrangian 3-

D models

Invesgate Concave and Convex walls of

square duct, and no significant results

1990 Nesic S., et al K- model Determined gas flow pattern near-wall

turbulence intensity


2012

75

(under gas

flow)

1991 Nesic S. K- model to

low Reynolds

number

Determined wall-pipe mass transfer rates

1993 Postlethwatte

S., et al

2-D (under

disturbed flow)

Determined local mass rates and s

Products at sudden expansion

2009 Binder S., et al Norsok M506

(Multiphase

gas /liquid

flow)

Generate noteworthy results but lack

agreement b/w modelling and field data

2011 Mamat M. F. Immersion and

Salt spray tests

Determined corrosion rate of the welded

and unwelded joint

2012 Muhammadu

M. M.

Non-

destruction test

(NDT) and X-

ray test

To determine the effects of fluid flow at

pipe bends on the corrosion behaviour of

low carbon steel

Further Work

Further work is required to clarify the stability of

corrosion behaviour under the condition of

turbulence fluid flow at pipe bend where are

amenable to both experiment observation and

numerical simulation in relation to Reynolds

number or particle image velocimetry (PIV).

Also, identification of positions or sites, within the

internal surface contact areas where the maximum

corrosion stimulus may likely to occur, thereby

allowing better understanding, mitigation,

monitoring and corrosion control over the life

cycle.

Conclusion

Corrosion and its interactions with the flow

phenomena is a complex discipline, neither one

dominates the other. Corrosion modelling results

are often found not to agree with field data. The

discrepancy is often blamed on the inadequacy of

the models. However, this article or review has

concluded that the differences are better explained

by the rationalization of the effects of flow pattern

on the base and localized corrosion rates. The core

postulate is that the lack of agreement between

corrosion modelling data and field experience is

due more to inadequate account of corrosion

stimulating flow regimes, rather than limitations of

the modelling: thus reverting the onus for corrosion

prediction away from corrosion modelling to the

flow regime side.

The arguments are new and on going and a

paradigm shift in thinking is needed to explore

further the links between flow pattern and

corrosion behaviour. In other word, a

computational fluid dynamic model coupled to a

Lagrangian particle tracking routine and a number

of erosion models have been used to predict the

solid particle erosion in square cross-section bends

for dilute particle-laden flow. The results obtained

were clearly affected by uncertainties in the

empirical restitution coefficients. However, the

results obtained do demonstrate the ability of the

CFD techniques employed to predict erosion

location and the depth of material eroded, and

hence their usefulness in providing estimate of the

service life of pipe systems as well as in the design

of mitigation measure.

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Friction Characteristic of Palm Olein at Different Operating

Temperature using Four-ball Tribom S. Syahrullail, C.I Tiong


Universiti Teknologi Malaysia, UTM Skudai

81310 Skudai, Johor, Malaysia

Tel : (607) 5534661. Fax : (607) 5566159


ABSTRACT

One of the main disadvantages of vegetable oil is its

poor performance at high temperature. In this

experimental work, the performance of refined,

bleached and deodorized (RBD) palm olein was tested

at different operating temperatures using four-ball

tribometer, following the procedure of ASTM D 4172.

The result produced by RBD palm olein was

compared with the result by additive free paraffinic

mineral oil. The result showed that the RBD palm

olein had lower coefficient of friction compared to the

paraffinic mineral oil. However, the wear scars on the

ball bearings surface lubricated with RBD palm olein

were larger compared to those lubricated with

paraffinic mineral oil.

Keywords : Four-ball tribometer, friction coefficient,

wear scar diameter.

1. INTRODUCTION Lubrication is a technique applied to reduce the wear

caused by contacting surfaces either in close proximity

or moving relative to each other. A substance called

lubricant is interposed in between the contacting

surfaces to help it to lessen the friction caused by the

pressure due to the weight of the the load. Proper

lubrication is the most effective method to control

friction and wear.

Tribology is science of friction, wear and lubrication.

Tribology is very important in modern machinery

since it consists of sliding and rolling surfaces. In fact,

tons of lubricants are used every year in the industry.

In 2004, 37.4 million tons of lubricants were used

worldwide. The percentages included 53% automotive

lubricant, 32% industrial lubricant; including related

specialities, 5% marine oil and 10% process oil [1].

The total consumption of industrial lubricant included

37% of hydraulics

oil, 7% industrial gear oils, 31% other industrial oils,

16% metal working fluids and 9% greases. Researches

in this field can ensure greater efficiency, better

performance, less breakdown and significant saving

[2]. Due to high demand, the researchers all over the

world try to come out with the alternative for mineral

oil based lubricant, such as vegetable oil. The

vegetable oil is so far the best choice because it is

made from vegetables and the most environments

friendly. Public awareness about the importance of the

conservation of the environment also encourages the

use of vegetable oil as lubricant. The public has been

made aware that the lubricants used in machines can

contaminate the soil and underground water supply.

Therefore, the role of vegetable oil as a lubricant is

crucial and the potential is very high.

Nowadays, lubricants based on vegetable oil have

started to replace the role of mineral based oil as

industrial lubricants. This is due to the fact that the

mineral oil lubricants are very dangerous in many

applications because they are not readily

biodegradable and are toxic. Global environment

awareness also encourages researchers to produce

environmental-friendly lubricants. Non-toxic and

biodegradable lubricants have become a major issue

especially when the lubricants involve contacts with

soil, crops and ground water. Biodegradability is the

ability of a substance to be decomposed by the action

of the bacteria into CO2, water, mineral compound

and bacterial bodies. There are some factors affecting

the biodegradability such as molecular structure,

chemical properties and environmental conditions [3].

Some other properties of vegetable oil such as high

viscosity index, good lubricity, high flash point and

low evaporative loss [4] are measurable characteristics

to choose the best lubricant. As a result, there has been

a major interest in developing all sorts of lubricants,

including greases and hydraulic fluids, based on

vegetables oils such as rapeseed oil, castor oil and

palm oil [5]. Rapeseed oils have excellent lubricating

properties, load carrying capacity, and corrosion

protection properties compared to mineral oil.

Vegetable oils or natural oils are not only limited for

use in the food industry, but there are also various

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applications of vegetable oil, such as the use as

lubricants. The examples of the fields of application

for environmental-friendly lubricants are [6]: outboard

two strokes engine oils, chain saw and saw frame oils,

hydraulic oils for forest and agricultural equipment,

lubricants for sewage treatment plant, water weir plant

and lock gate, lubricant for food machinery, metal

working and metal forming processes and internal

combustion engine and hydraulic system.

However, vegetable oils have poor performances at

high temperature and poor oxidation properties. In

general, vegetable oils have a structure of triglyceride

with three different fatty acids. In this structure, there

are saturated and unsaturated fatty acids. The higher

the concentration of unsaturation, the poorer the

oxidation stability will be. These fatty acids are

effective in improving the capability of vegetable oil

to create the boundary lubrication condition along the

process. When two surfaces move towards each other,

the movement will generate heat. The increment of

heat would break the double bond (unsaturated bond)

in vegetable oils and create a possibility to other

elements to have chemical reaction with the vegetable

oils structure. At the same time, the existence of

oxygen molecules in vegetables oils chain would react

with other metals and lead to oxidization of the metal.

In this experimental work, a refined, bleached and

deodorized (RBD) palm olein was tested using a four-

ball tribometer at different operating temperatures

following the procedure of ASTM D 4172. The results

produced by RBD palm olein were compared with

results by the additive free paraffinic mineral oil. It

was found that RBD palm olein showed a lower

coefficient of friction.

2. EXPERIMENTAL METHOD

The experimental work was done using a four-ball

tribometer as shown in Figure 1, following the

procedure of ASTM D 4172. The normal load was

fixed to 40kg and the speed was set at 1200rpm. The

test was conducted for one hour. Test lubricants were

heated up to 75°C before starting the experiments. The

details of the test procedure had been described in

previous publications [7]. The standard steel balls used

in this experiment are made from AISI E-52100

chrome alloy steel, with the diameter of 12.7 mm,

extra polish (EP) grade 25, hardness 64 to 66 HRC

(Rockwell C Hardness). Four new balls were used for

each test. Each time before starting a new test, the

balls were cleaned with acetone and wiped dry using a

fresh lint-free industrial wipe.

The test lubricant used was RBD palm olein. RBD is

the abbreviation of refined, bleached and deodorized.

Palm olein is the liquid fraction obtained from the

fractionation of palm oil after crystallization at a

controlled temperature. In this research, a Malaysian

Standard of MS 816:1991 of RBD palm olein was

used. This type of palm olein has been refined and

contains less free fatty acid. The results obtained from

the experiments using RBD palm olein were compared

with the results from the experiment which used

paraffinic mineral oil. Each test used 10ml of

lubricant.


3.1. Coefficient of Friction

Antifriction ability is one of the important

characteristics of lubricants in order for the

mechanical system to run smoothly and reduce the

maintenance cost. In this experiment, the performance

for RBD palm olein (PO) and paraffinic mineral oil

(PMO) were evaluated in term of coefficient of

friction (COF) using a four-ball tribometer. The results

in Figure 1 shows the COF for RBD palm olein and

paraffinic mineral oil at different operating

temperatures. In this experiment, the operating

temperatures were set from 55˚C to 85˚C with gradual

10˚C increment. It was observed that the trend of COF

for RBD palm olein and paraffinic mineral oil

increased as the operating temperature increased. At

the operating temperature of 55˚C, the COF for RBD

palm olein and paraffinic mineral oil were similar,

with 0.066 and 0.068 respectively. At the operating

temperature of 85˚C, the COF for RBD palm olein and

paraffinic mineral oil increased to 0.08 and 0.086

respectively. The result showed that the coefficient of

friction for both oils increased with the increase of the

operating temperature. The coefficient of friction for

RBD palm olein (PO) was lower compared to

paraffinic mineral oil (PMO). It was due to the

presence of fatty acid molecules in vegetable oil that

stuck well on the surface and created a boundary

lubrication condition [8,9]. The increment of operating

temperature might have lead the boundary lubricant

breakdown due to lower viscosity [10].


2012

80

0.04

0.05

0.06

0.07

0.08

0.09

55 65 75 85

Temperature (Celcius)

Co

eff

icie

nt

of

fric

tio

n

PO

PMO

Figure 1: Coefficient of friction for RBD palm

olein and paraffinic mineral oil under

different operating temperature.

3.2. Wear Scar Diameter

The wear scar diameter (WSD) of three stationary

bearing balls was measured using a CCD microscope

for each lubricant in this experiment, and plotted as

Figure 2.

0.5

0.55

0.6

0.65

0.7

0.75

0.8

0.85

0.9

55 65 75 85

Temperature (Celcius)

Wear

scar

dia

mete

r

PO

PMO

Figure 2 : Wear scar diameter measured for

bearing ball lubricated with RBD

palm olein and paraffinic mineral oil

under different operating

temperature.

The result showed that the RBD palm olein had almost

similar or slightly higher value of WSD compared to

paraffinic mineral oil. This could have been attributed

by the chemical attack on the rubbing surfaces of fatty

acid in RBD palm olein [11]. Both test oils showed a

slight reduction of WSD value with the increase of

operating temperature.

3.3. Wear Scar Surface Roughness

The surface roughness profile of wear scar surface of

bearing balls lubricated with RBD palm olein and

paraffinic mineral oil were measured using a Mitutoyo

surface roughness profiler. Figure 3 shows the surface

roughness, Ra of wear surface for RBD palm olein and

paraffinic mineral oil under different operating

temperatures. From Figure 3, it can be seen clearly

that the wear surface roughness of bearing ball

lubricated with RBD palm olein became smoother;

however, those lubricated with paraffinic mineral oil

had more wear surface roughness with the increase of

the operating temperature. For RBD palm olein

lubricant, the highest surface roughness value was

2.313µm, at the operating temperature of 55°C. For

paraffinic mineral oil, the highest surface roughness

value was 3.250µm, at the operating temperature of

85°C. From the results in Figure 3, it can be

understood that the wear surface roughness for RBD

palm olein after experiments had become smoother

with the increase of the operating temperature but the

wear surface profile paraffinic mineral oil became

worse with the increase of the operating temperature.

The change in surface roughness value mainly resulted

from the change of wear mechanisms as suggested by

Hiroyuki [12] and the reduction of lubricant viscosity.

0

0.5

1

1.5

2

2.5

3

3.5

55 65 75 85

PO

P2

Figure 3 : show the surface roughness for the

wear surface of bearing ball lubricated

with RBD palm olein and paraffinic

mineral oil respectively after the

experiment.


2012

81

(a) PO at 55°C

(b) PMO at 55°C

(c) PO at 65°C

(d) PMO at 65°C

(e) PO at 75°C

(f) PMO at 75°C

(g) PO at 85°C

(h) PMO at 85°C

Figure 4 : wear worn surface lubricated with PO and PMO under different temperature.


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3.4. Worn Surface Observation

The worn surfaces of balls bearing were observed and

captured using a CCD microscope. All the worn

surfaces are shown in Figure 4 with lubrication of PO

and PMO respectively, under different operating

temperatures. PO and PMO represent the RBD palm

olein and additive free paraffinic mineral oil in this

experiment. The worn surfaces of PO and PMO were

compared mutually at specific operating temperature.

For the wear surface with lubrication of PO at

temperature of 55°C until 85°C, we could clearly see

the scar or grooves on the worn area. The scars or

grooves became narrower with the increase of

operating temperature. The largest wear groove with

lubrication of PO was found at the temperature of

55°C. This also indicated that the wear surface was

the roughest compared to others. The surface became

smoother with the increase of operating temperature

by showing narrower grooves on the surface.

However, at the temperature of 85°C, the wear

surface with lubrication of PO became rough and an

adhesive wear could be observed just like others. The

best wear worn surface was at the temperature of

75°C for PO by giving the smoothest surface. On the

other hand, the wear surface with lubrication of PMO

showed the opposite trend of result compared to PO.

The wear worn surface became worse with the

increase of operating temperature in this experiment.

At the operating temperature of 55°C until 65°C, the

wear groove could clearly be seen on the worn

surface lubricated with PMO. The wear scar or

grooves became unclear with the increase of

operating temperature, thus it could be observed that

material transfer or remover occurred on the worn

surfaces at the operating temperature of 75°C onward.

4. CONCLUSION

The tribological properties of RBD palm olein have

been investigated using four-ball tribometer at

different operating temperatures. The results of RBD

palm olein are compared with the result by additive

free paraffinic mineral oil. It can be concluded that

RBD palm olein gives better lubrication

performances based on the lower coefficient of

friction compared to paraffinic mineral oil. However,

the wear with the use of RBD palm olein is slightly

higher compared to the wear with the use of

paraffinic mineral oil. This problem could be solved

by the addition of anti-oxidant agent. The wear scar

on the ball bearings with lubrication of RBD palm

olein are smooth, showing that less metal-to-metal

contact occurred.

ACKNOWLEDGMENT

The authors wish to thank the Faculty of Mechanical

Engineering at the Universiti Teknologi Malaysia for

their support and cooperation during this study. The

authors also wish to thank Research Management

Centre (RMC) for the Research University Grant

(GUP) from the Universiti Teknologi Malaysia,

Fundamental Research Grant Scheme (FRGS) from

the Ministry of Higher Education (MOHE) and E-

Science Grant and ERGS from the Ministry of

Science, Technology and Innovation (MOSTI) of

Malaysia for their financial support.

REFERENCES

[1] T. Mang and W. Dresel, Lubricants and

Lubrication, 2nd Edition, WILEY-VCH Verlag

GmbH & Co. KGaA, Weinheim, 2007.

[2] B. Bhusan, Principles and Application of

Tribology, John Wiley and Sons, 1999.

[3] B. Krzan and J. Vizintin, ―Vegetable-Based Oil

as a Gear Lubricant‖, Gear Tech., July/August

2003, pp.28-33.

[4] N.H. Jayadas, K. Prabhakaran Nair, and

Ajithkumar G, ―Tribological Evaluation of

Coconut Oil as an Environment-friendly

Lubricant‖, Trib. Int., vol.40, pp.350-354, 2007.

[5] A.R. Lansdown, Lubrication and Lubricant

Selection, Professional Engineering Publishing

Limited, London and Bury St Edmund, UK,

2004.

[6] W. J. Bartz, ―Lubricants and the Environment‖,

Trib. Int., vol.3, no.1-3, pp.35-47, 1998.

[7] C.I. Tiong, A.K. Mohammed Rafiq, Y. Azli and

S. Syahrullail, ―The Effect of Temperature on the

Tribological Behavior of RBD Palm Stearin‖,

Trib. Trans., vol.55, no.5, pp.539-548, 2012.

[8] S. Syahrullail, K. Nakanishi, S. Kamitani,

―Investigation of the effects of frictional

constraint with application of palm olein oil

lubricant and paraffin mineral oil lubricant on

plastic deformation by plane strain extrusion‖, J.

Jap. Soc. Trib., vol.50, no.12, pp.877-885, 2005.

[9] S. Syahrullail, B.M. Zubil, C.S.N. Azwadi and

M.J.M. Ridzuan, ―Experimental Evaluation of

Palm Oil as Lubricant in Cold Forward

Extrusion‖, Int. J. Mech. Sci., vol.53, pp.549-

555, 2011.

[10] H.H. Masjuki, M.A Maleque, A. Kubo, T.

Nonaka, ―Palm oil and mineral oil based

lubricant- their tribological and emission

performance‖, Trib. Int., vol.32, pp.304-314,

1999.


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[11] F.P. Bowden and D. Tabor D, The Nature of

Metallic Wear. The Friction and Lubrication of

Solids, Oxford Classic Texts. New York: Oxford

University Press, 2001.

[12] K. Hiroyuki and K. Hokkirigawa, ―Transitions of

microscopic wear mode of silicon carbide

coatings by chemical vapor deposition during

repeated sliding observed in a scanning electron

microscope tribosystems‖, Wear, vol.185, pp.9-

15, 1995.


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Pulse Detonation Engine Research Development at High Speed

Reacting Flow Laboratory, Universiti Teknologi Malaysia Mazlan A. Wahid, A. Dairobi G., Aminuddin Saat, Mohsin M. Sies, H.A. Mohammed,

A. N. Darus, Mohd Faizal H., M. Ibthisham A., Fairus M. Y. and Z. Lazim

High-Speed Reacting Flow Laboratory - HiREF

Department of Thermofluids


Universiti Teknologi Malaysia

81310 Skudai, Johor Darul Takzim

MALAYSIA


ABSTRACT

Pulse detonation is a propulsion technology that

involves detonation of fuel to produce thrust more

efficiently with mechanical simplicity than currently

available engine systems. Detonation combustion

modes in pulse detonation engines (PDE) is known to

be more efficient in creating high power simply

because the energy release in detonation modes of gas

combustion are several magnitude higher than in

deflagration combustion modes. PDE research

program has been started at High Speed Reacting

Flow Laboratory (HiREF), Universiti Teknologi

Malaysia (UTM) since 2005. The studies began with

single pulse detonation study for detailed investigation

of detonation characteristics of various fuels and

followed by development of repetitive PDE engine,

performance study and augmentation of thrust using

various types of ejectors. This paper summarizes the

laboratory facilities, research activities conducted and

the output from all of the related research programs.

Keywords : High Speed Reacting Flow Laboratory

(HiREF), pulse detonation engine,

combustion, high speed reacting flow

1. INTRODUCTION

Pulse detonation engine (PDE) is a new type of

propulsion technology with advantages on thermal

efficiency, higher thrust impulsive and simplicity

mechanical design [1-4]. Researchers from various

institutions also suggest that the PDE have the

potential to power aircraft on subsonic, supersonic and

hypersonic speed [5-7]. PDEs could also be used to

power tactical aircraft, air and ship-launched missiles,

unmanned aerial vehicles, power generation, and a

wide range of stand-off munitions [8,9]. However

extensive research and development are indeed

essential before PDE can be successfully applied on

those practical applications.

High Speed Reacting Flow Laboratory or in short

HiREF, has been established to focus on the study of

various discipline of sustainable combustion, high

speed reacting flow and heat transfer. In 2005 HiREF

embarked on pulse deflagration and detonation

combustion research studies and since then the group

has earned various invention awards as well as several

patents on pulse combustion technology. The

advantages and potential inherited by PDE as the new

propulsion engine concept has driven HiREF members

to focus on such promising area especially in the era

when fuel efficiency and sustainability has been a real

concern in today's depleting fossil fuel scenario. Since

then HiREF laboratory has been evolved stage by

stage to meet the PDE research requirement.

The research programs on PDE initially began with

fundamental study on detonation characteristic using

shock tube to study the characteristic of various fuels.

The studies continued with the development of the

repetitive PDE and some applications of PDE. The

team successfully operate the PDE at a reasonable

repetitive rate with substantial thrust. HiREF team is

currently busy in upgrading the fuel-oxidizer injection

system, higher energy ignition system, Deflagration to

Detonation Transition (DDT) mechanism as well as

improving the firing sequencing system.

Augmentation of PDE thrust also been studied by

employing ejectors of various geometry placed at the

open end of the PDE tube.

2. HIREF LABORATORY AND

COMPUTING FACILITIES

HiREF laboratory has been designed, modified and

equipped in meeting the PDE research requirement

and precise data collection purpose with safety as the

priority. The HiREF research facility is comprised of a

research office furnished with computing facility and a

laboratory that is equipped with soundproof room,

damping chamber, exhaust system, control room, fuel

and oxidizer supply system. Fuel is stored at the

IMAT-UI 015



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85

outside of the laboratory and delivered through

certified stainless steel pipeline and brass tubing.

Figure 1 shows the schematic diagrams of the HiREF

laboratory. Description of the HiREF laboratory

facilities is discussed next.

Figure 1: Schematic diagram of HiREF

laboratory arrangement

2.1 Soundproof Room

The dimension of the soundproof room is 8‘ x 11‘ x

14'. The room is made of double layers wooden walls

where polyurethane foam is fitted in between them.

Soundproof room is equipped with damping chamber,

exhaust fan, and fuel pipeline. The PDE setup with all

the necessary instrumentation is placed in this room

during experiment. Most of the experiments are

conducted inside this room as a confine area to

prevent detonation shock wave, noise, heat or any

unwanted event from harming the researcher.

Researcher can observe, operate and gather all PDE

experimental data from a control room that is

connected to the sound proof room through a blast

proof observation window.

2.2 Damping Chamber

Pulse detonation engine producing noise level with

more than 100 db [10]. The sound need to be damped

and this is achieved with the design and installation of

damping chamber in the soundproof room. This

chamber is also function to absorb shock waves and to

extract the exhaust gas from the chamber. It also helps

prevent any unburned mixture from spreading to the

open space in the laboratory in order to prevent

unexpected accident. The damping chamber is made

of large stainless steel silencer housed in a concrete

cylinder. The outer concrete cylinder formed as a rigid

structure for the silencer and act as an absorber to the

silencer due to the impact from the shock wave. Steel

silencer is a perforated metal plate rolled into

cylindrical shape. Polyurethane foam in the inner liner

of the concrete cylinder is function to further absorbs

the sound produce during the detonation event. Figure

2 shows the assembly of damping chamber concrete

cylinder with the silencer.

Figure 2: Damping chamber design

2.3 Control Room

Researcher control and gather performance data for

the PDE during operation using two respective

computers located in a control room. A dedicated

control room served to protect the operator from PDE

shock wave and noise produced by PDE during

operation. Figure 3 shows the schematic diagrams of

control room in the HiREF laboratory.

Figure 3: Schematic of control room

2.4 Measurement System and Devices

Measurements are an essential part of the PDE

experiments. Typical data captured for PDE analysis

are pressure, temperature, force, and acceleration. All

digital data acquired from various transducers such as

9” wall

Observation

windows

Sound

absorption

layer

4” wall


2012

86

pressure transducers, thermocouples, load cell and

accelerometers are fed through data capturing systems

comprises of a software and acquisition card. Labview

7.1 software is used to control and command the PDE

operation. Separate computer using the same

programming language is used to write, store and read

back the data for analysis. PCI-6133 acquisition card

is used to read the signal from all of the transducers

for each data type in terms of voltage before it will be

stored in the hard disc. This card then connected to

connector block and the signal conditioner. Signal

conditioner receives the signal from a transducer in

analog signal type that is in voltage. This signal

converted to digital signal that later sent to the

connector block that communicated with date

acquisition card. Figure 4 shows the schematic

diagrams of the data-acquisition connections. In the

figure, it also shows the location of pressure

transducer 1 and 2 (PT1 and PT2), load cell and

accelerometer.

Figure 4: Schematic diagram of PDE data

acquisition system

In the previous study, accelerometer and load cell

were used to measure the acceleration and thrust

respectively. The study on the suitability of using the

accelerometer on PDE thrust measurement have been

performed to fulfill the data validation requirement

[11, 12].

2.5 Cooling System

Experiments on supersonic combustion at a high

repetitive rate will generate heat sometimes up to

thousands of degree Celsius. The heat produces from

the combustion process need to be managed properly

since excessive heating will cause damage to the

transducers. The maximum operating

temperature for pressure transducer is only around

250°C whereas temperature during detonation

combustion may easily reach more than 1000°C.

Temperature higher than allowable operating

temperature will most likely cause damage to the

electronic part of the transducers.

The PDE instrumentation cooling system, such as

shown in Figure 5, includes a cooling station and

transducers cooling adaptors. This station functions

as liquids storage, supply, circulations and control.

The cooling station consists of water storage tank,

water pump, radiator, fan and also the control panel.

Cooling adapter is served also as the mounting for

transducer to provide cooling to the transducer. This

adaptor was custom made from stainless steel material

for the corrosion resistant. In the system cooling

fluid is circulated by 0.5 horsepower water pump with

40 milliliters per minute of water flow rate.

Figure 5: Picture of cooling system station.

3. SINGLE PULSE DETONATION

SETUP

Single pulse detonation setup was constructed in order

to study the details characteristic of high-speed

reacting shock waves for various fuels [13]. The

detonation tube is made of stainless steel with internal

diameter of 100 mm, damping chamber, deflagration

to detonation transition (DDT) section, soot film

section, gas filling system, a data acquisition system

and ignition control unit. The tube was designed to be

able to withstand static pressure loading of 100 bars

with safety factor of 2.49.

Water

tank

Water pump

Control panel

Radiator


2012

87

Figure 6: Schematic of single pulse detonation

setup

Figure 7: Single Pulse Detonation

The single pulse detonation tube was constructed in

three modular sections such as shown in Figure 6 and

7. Each section is 0.5 m long and has an internal

diameter of 100 mm and an outer diameter of 112 mm.

Section A and C has ports for plumbing and

instrumentation while section B only has ports for

instrumentation. Each section was connected to each

other using flanges. The flanges are 164 mm in

diameter and 5 mm thick and are held together by

eight M10 bolts. Silicon glue and gasket paper were

used to seal small gap between flanges and

connections when they are connected. The close end

part of the tube, which is located at the section A, is

attached with a plumbing valve, transducers and spark

plug. The open end where the detonation wave

exhaust into the atmosphere was inserted inside a

damping chamber to reduce sound that is created by

the transmitted wave. The damping chamber was

placed onto a detachable support. The detachable

support was tightened to the main structure using four

M12 bolts.

This facility was also used to study the characteristics

of reacting shock waves for biogas in comparison to

several other gaseous fuels [13] and experimental

study of confined biogas pulse detonation combustion

[14]. Biogas was comprised of 65% methane with

35% carbon dioxide. The oxygen concentration in

the oxidizer mixture was diluted with nitrogen gas

at various percentage of dilution. Computational as

well as the experimental studies of biogas and natural

gas fuel characteristics has also been performed in

HiREF [15, 16]. From both research programs it can

be concluded that the biogas and natural gas were not

sensitive to detonation propagation compared to other

gaseous like propane due to the lower calorific value.

4. REPETITIVE PULSE DETONATION

ENGINE SETUP

Repetitive Pulse Detonation Engine been developed

by HiREF team and the setup is shown in Figure 8.

The dimension of PDE tube was 50 mm in inner

diameter and 600 mm in length. Obstacles with

blockage ratio of 43 percent was used as an flame

accelerator inside the tube. Such tube size was chosen

since it will provide enough space for the inlet device

to be mounted on the tube surface. Propane and

oxygen was used as fuel and oxidizer. In the

experiments the cell size (λ) for propane-oxygen was

determined to be 1.3 mm [17].

Purging system uses solenoid valves with the

capabilities to supply compressed air up to a pressure

of 800 kPa. The solenoid coil was powered by 24Volt

DC power supply circuit and controlled by transistor-

transistor logic (TTL) signal from a PDE control

circuit. This PDE been designed to operate by using

gaseous propane as fuel and pure oxygen as oxidizer.

LO-Gas injector is the manufacturer for injector and

pressure regulator. This injector was designed to work

on 12Volt DC supply voltage with minimum pressure

as low as 100kPa to maximum pressure of 300kPa.

The injector‘s assembly came with 4 in line units for

oxidizer injection and 2 in line units for the fuel

injection. The injectors are mounted not directly to the

detonation tube to prevent overheating and failure due

detonation pressure wave and heat generated. These

injectors are attached to a custom-made mounting and

a 6 mm stainless steel pipe is used to deliver the

injected mixture to the detonation tube. One-way

valves are used to prevent the flame from flashing

back to the fuel-system tank. 12 VDC MSD Digital

DIS- 4 was chosen as an ignition system. This device

can provide four sparks of 105-115 mJ each. The

mechanism of filling, purging and ignition was

controlled by using LabView. Figure 9 shows the

sequences of the filling, ignition and purging. The

injection process was set for 50% duty cycle, 15%

duty cycle for ignition process and 20% for purging

process.


2012

88

Figure 8: Repetitive Pulse Detonation Engine In order to determine the characteristics and

performance of the PDE, the complete system has to

be mounted on a strong support structure. The support

structure must be rigid enough to be able to withstand

the vibration and force exerted by the impulse

generate during the test running. The tube mounting

also designed to have free movement on a railing so

that the amount of thrust generated can be measured.

Figure 9: PDE sequence

The PDE support structure was built from mild steel

angle iron with a dimension of 50mm x 50mm with

the thickness of 5mm combine with rectangular

hollow with a dimension of 50mm x 25mm and the

thickness of 3mm. The main support structure is

assembled by using arc welding to get rigid structure

to resist shock by the impulse cause by the detonation

event. The detonation tube is mounted on a freely

moving structure that is installed on a railing. This

structure has a sliding mechanism at the base so that it

will reduce the friction to the movement. This

requirement is needed to measure thrust by using load

cell. During the test, this structure will move backward

and compressed a load cell located between these

structures to the thrust wall.

Figure 10 shows the thrust recorded by the load cell at

frequency 5 Hz for 5 seconds. It can be seen that the

average values of the thrust spike by the PDE was

between 480 N and 520 N at these operating

frequencies. The average thrust spike calculated from

these results is 512.7 N. Figure 11 below shows the

pressure profile at frequency 5 Hz. The average values

of pressure spike were around 8 to 14 bars. The rise

time of the pressure is within the order of 10-4

s. In

order to further understanding the phenomena occur

inside of the PDE chamber, simulation had be

conducted detail in the literature [18].

Figure 10: Repetitive PDE thrust signal

Figure 11: Repetitive PDE pressure signal

5. THRUST AUGMENTATION

Installation of ejectors has shown some improvement

on the thrust generated by the PDE [19]. Experimental

test was conducted at HiREF to study thrust

performance on 5 Hz PDE and the ejector setup is

shown in Figure 12. This research utilized propane as

a fuel, oxygen as oxidizer and air as purge gas. The

effects of four different ejectors dimension, shape and

axial location on augmentations were investigated and

the ejector photo is shown in Figure 13. All ejectors

have similar length and inlet diameter which are 400

mm and 130 mm respectively. The length chosen for

all ejector configurations was at ejector length-to-

ejector diameter, LAugmentor/DAugmentor= 3.08. 5Hz

operation frequency of PDE shown the wave velocity


2012

89

was in range of 1500 m/s to 1800 m/s (~5-6 Mach)

and the thrust is about 9-14 bar is produced. The

ejector was found to be very sensitive to the axial

position. All the ejectors give improve performance of

thrust and the upstream position provided the best

performance due to the thrust augmentation.

Figure 12: Ejector arrangement on PDE

Figure 13: Ejector geometries; from left diverge,

converge, convergent-divergent and

straight ejector

6. FUTURE WORK

HiREF research group is currently becoming more

active in acquiring research grant in various discipline

of sustainable combustion, high speed reacting flow

and heat transfer. In 2012 the group has secured about

six new research grants and hence HiREF laboratory

will undergo few more upgrading in term of facilities.

One of the proposed research aims is to understand

fundamentally on the process of high speed reactive

flows especially for safety and accidents prevention.

The transition from deflagration to detonation will be

observed in unconfined congested environments. Such

environments are analogous to that in factories,

chemical process plants, forests, and warehouses. The

congestions (i.e. obstacles) will cause high-speed

turbulent deflagrations to form shock waves at the

flame front, which is the main characteristic of

supersonic combustion. In contrast, detonation also

occurs spontaneously when the ignition energy is

sufficient to enforce a shock wave at the onset of

combustion. In year 2012 the group has successfully

increases the number of publications in high rank

journals with cumulative impact factor of 30 [20-32].

ACKNOWLEDGMENT

The authors would like to acknowledge Universiti

Teknologi Malaysia for providing laboratory facility

and fund granted under Research University Grant

(RUG) scheme. The financial support provided by the

Malaysian Ministry of Higher Education (MOHE)

under Vote 79299 for PDE reasearch and laboratory is

highly acknowledged. Special thanks also to Ministry

of Science and Technology of Malaysia (MOSTE) for

providing various type of funding to support various

research programs at HiREF. We also want to

acknowledge the combustion and thermodynamic

laboartory staff for their help and advice on

experimental research program at HiREF laboratory.

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B.H., The effect of nanofluids flow on mixed

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A.S., 2012, Heat transfer enhancement of

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M.Z., Saidur, R., 2012, Buoyancy-assisted

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Pool Boiling of Nanofluids in Vertical Porous Media

Ridho Irwansyah, Nandy Putra

Heat Transfer Laboratory

Department of Mechanical Engineering

Unviersity of Indonesia, Depok, 16424

E-mail: [email protected]

Abstract:

The development of electronic components such as

microprocessor requires a better thermal

management system to overcome the high heat flux

produce by the component. The method to absorb

the heat produce by the microprocessor is still use

the conduction or either natural or free convection

which still in a single phase heat transfer. One of

heat transfer method that suitable for a high heat

flux application is pool boiling which has a two

order of magnitude higher than of a single phase

heat transfer and does not require a pump to move

the fluid. In this study has been conducted the pool

boiling experiment with four different porous

media surface which are sintered copper 300 µm

and 400 µm, copper screen mesh and stainless steel

screen mesh with four different fluid which are

H2O-Al2O3 1%, 3% and 5%. The sintered copper

400 µm has shown a better heat transfer

performance compared to the other porous media.

The H2O, H2O-Al2O3 5% has shown a performance

no better than H2O-Al2O3 1% and 3%.

Keywords: Nanofluids, Sintered Copper, Screen

Mesh, Pool Boiling

1. Introduction Technological developments towards

miniaturization of electronic components require

methods for better thermal management. The

microprocessor is one example of electronic

components with rapid growth, the development of

which was followed by an increase in waste heat

generated when the microprocessor works. The

typical cooling system for electronic components

based on conduction and single phase forced or

natural convection already inadequate to handle

such a high performance electronic components

[1].

One method of heat transfer that is often used in

the cooling system on the electronic components is

pool boiling. This is due to the high heat transfer

capabilities and the process does not require a

pump to move the working fluid [2]. Pool boiling

chosen due to its ability to move heat two times

better compared to the single-phase heat transfer in

conventional cooling methods [3]. There are

several method to enhance the heat transfer ability

of pool boiling, one of the common methods is

additive for liquids. In this case the addition of

nano size solid particle was required (10-100 nm).

The addition of nanoparticles to the base fluid

tends to increase the thermal conductivity of

working fluids [4, 5]. Several researchers have

done the research about boiling heat transfer, they

found that the boiling of nanofluids increase the

critical heat flux (CHF) of the boiling process

compared to the base fluid. Hyungdae Kim et.al

found that the using of Al2O3 and TiO2 nanofluids

increase the CHF of the boiling process into 170%

[6]. S.M You et.al found that the using of Al2O3

nanofluids at 0-0.05 g/l concentration increase the

CHF into 200% [7]

The other method to enhance the heat transfer

ability in pool boiling is the modification of heater

surface. Y. Takata et.al found that the coating of

TiO2 nanoparticles on the surface of the heater

improve the critical heat flux (CHF) compared to

the heater without nanoparticle coating [8]. Other

researchers have done another method to modified

the heater surface with coating with additional

nanoparticle layer on the heater surface [9, 10]

Shoji Mori and Kunito Okuyama found that the

using of porous media on the surface of the heater

increase the CHF value 2.5 times compared to the

heater without porous media[2].

The purposes of this research are to conduct the

pool boiling experiment of H2O-Al2O3 nanofluids

in vertical porous media and compare the result of

the nanofluids to the base fluid.

2. Experimental Setup The H2O-Al2O3 nanofluids in 1%, 3% and 5% of

volume concentration were used in this research.

The average size of the particle was 20 nm.

Distilled water (H2O) was used as the base fluid.

There are two steps to produce the nanofluids; the

first step was the stirring between the mixture of

nanoparticles and base fluid by utilize the magnetic

stirrer for 15 minute, and later the mixture were

added to the ultrasonic processor for one hour in a

high frequency process.

The porous media that being used in this research

were made from copper powder (particle size 300

µm 400 µm) and screen mesh (copper and stainless

steel) which has 40 mesh number, the porous media

is depicted in fig. 1. The SEM (Scanning Electron

Microscope) images of 300 μm copper powder are

depicted in fig.2. The particle was magnified 100

times and 300 times.

IMAT-UI 016


2012

93

Fig. 1. Copper screen mesh, stainless steel screen mesh and sintered copper porous media

(a) (b)

Fig.2. SEM images of 300μm copper powder (a) 100 times magnifying (b) 300 times magnifying

1

2

8

4

6

5

3

9

7

(a) Experimental setup

150 mm

130 mm

20 mm40 mm40 mm

(b) Thermocouple position on the heater surface

1. Circulating thermostatic bath

2. DC power supply

3. Condenser

4. Main heater

5. Auxiliary heater

6. Wall thermocouple

7. Fluid thermocouple

8. Computer

9. Data acquisition

Fig.2. The experimental setup and detail for thermocouple position


2012

94

300 W DC electric heater was used in this research.

The length and diameter of the heater was 150 mm

and 6 mm respectively and made of stainless steel.

The heater was connected to the adjustable DC

power supply. The auxiliary heater was used during

the early stage of the experiment. The temperature

measurement was done at 5 points which are 3

point at the heater surface and 2 points in the fluid.

5 type K thermocouples were connected to the NI

9201 data acquisition system and the error of the

thermocouple was 0.1 %. The boiling vessel was

made of glass. The thickness, height and diameter

of the glass were 6 mm, 200 mm and 115 mm

respectively. At the top of the boiling vessel a

condenser made from copper connected to the

circulating thermostatic bath. The temperature of

the bath were maintained at 25 0C. The detail of

experimental setup and thermocouple position of

the heater is depicted in fig. 3.

The thermal conductivity of the nanofluids was

measured using the KD2 Decagon method, the

same method was also use by [11, 12, 13]. The

thermal conductivity of the fluids from the

measurement result was shown in table 1.

Table 1. Thermal conductivity of fluids

Fluids

Thermal

Conductivity

[W/m.K]

Thermal

Conductivity

Improvement

[%]

H2O 0.56 -

H2O-Al2O3 1% 0.67 19.64

H2O-Al2O3 3% 0.69 23.21

H2O-Al2O3 5% 0.72 28.57

16 experiments were done during this research. 4

different fluids were tested for every porous media.

The surface roughness of the porous media that

made from sintered copper was measured before

and after the boiling process. The matrix of the

experiment was shown in table 2.

Table 2. Matrix of the experiment

Porous Media Fluid

Sintered Copper 300 µm H2O

Sintered Copper 400 µm H2O+ Al2O3 1%

Screen Mesh Copper H2O+ Al2O3 3%

Screen Mesh Stainless Steel H2O+ Al2O3 5%

3. Result and Discussion

3.1 Boiling heat transfer in sintered copper

porous media

The boiling curve of 300 μm sintered copper under

the variation of fluids is depicted in fig. 4. It can be

observed that boiling with the H2O-Al2O3 1%

volume concentration provides higher heat transfer

compared to other fluids. It can be seen that the

H2O-Al2O3 1% produce the lowest temperature

difference between the heater wall and the fluid

which was 3.74 K when the highest heat flux was

applied. It can be noted that the range of the heat

flux was between 1.8-36 kW/m2. The temperature

difference between the heater wall and the fluid of

water, 3% and 5 % nanofluids were 3.95 K, 4.42 K

and 4.2 K respectively.

Fig.4. Boiling curve of 300 μm sintered

copper under the variation of fluids

The boiling curve of 400 μm sintered copper under

the variation of fluids is depicted in fig. 5. The

additions of 3% nanoparticles to the base fluid

provide a better heat transfer compared to the other

fluids with the temperature difference between the

heater wall and the fluid was 0.22 K. In this

experiment it was found that the 1% and 3%

nanofluids provide a better heat transfer compared

to base fluid (H2O). The temperature difference

between heater wall and fluid for water, 1% and

5% nanofluids were 0.49 K, 0.47 K and 0.52 K

respectively.

Fig.5. Boiling curve of 400 μm sintered

copper under the variation of fluids


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3.2 Boiling heat transfer in screen mesh porous

media

The boiling curve of stainless steel screen mesh

porous media is depicted in fig. 6. The 1% H2O-

Al2O3 nanofluids consistently shows a better heat

transfer compared to other fluid until the heat flux

reach 30 kW/m2 with the temperature difference

was 3.67 0C. When the highest heat flux 26 kW/m

2

applied, the base fluid shows the better heat

transfer compared to other fluid. When the highest

heat flux applied, the temperature difference

between heater wall and fluid for water, 1%, 3%

and 5% H2O-Al2O3 nanofluids were 1.89 K, 2.62

K, 2.99 K and 3.16 K respectively. It means that

the base fluid provide a better heat transfer

compared to the nanofluids.

Fig.6. Boiling curve of stainless steel screen

mesh

The same phenomenon also observed by Sarit K.

Das et.al [14, 15] where they found that the

addition of nanoparticle to the base fluid deteriorate

the heat transfer.

The boiling curve of copper screen mesh was

depicted in fig. 7. The 1% and 3% H2O-Al2O3

consistently shows a better heat transfer compared

to the 5% H2O-Al2O3 and base fluid. The 1% and

3% H2O-Al2O3 have almost a similar heat transfer.

The temperature difference between the heater wall

and the fluid of The 1% and 3% H2O-Al2O3 were

0.57 K and 0.54 K.

Fig.7. Boiling curve of copper screen mesh

3.3 Surface roughness measurement

Table 3 shows the surface roughness of 300 μm

sintered copper surface before and after the boiling

process with nanofluids. The measurement results

show the changes in the surface roughness after

boiling of nanofluids. The surface characteristic of

the porous media were measured using the

profilometer.

Table 3. The surface roughness of sintered copper

porous media before and after boiling Measurement

Point

Ra Before

Boiling [µm]

Ra After

Boiling[µm]

1 14.38 5.46

2 11.7 7.36

3 12.18 7.1

4 13.18 11.22

The addition of nanoparticles deposit on the surface

of the porous media can be one of the causes in the

decrease of the surface roughness, since the

nanoparticles has a smaller particle diameter

compared to the powder that being used as the

porous media.

4. Conclusion

From the result of the experiment, it can be

concluded that :

The 5% H2O-Al2O3 shows a heat transfer

that no better than the other fluids. The 5%

H2O-Al2O3 always shows a higher

temperature difference of the heater wall and

fluid compared to other fluids.

The 400 µm sintered copper consistently

shows a better heat transfer compared to

other porous media.

The decreased of the surface roughness of

the porous media was found after the boiling

process. It can be said that the additional

layer of nanoparticles appear on the surface

of the porous media. This phenomena was

possible due to the smaller size of the

nanoparticle compared to the porous media.

5. Acknowledgement

The author would like to thank DRPM UI for

funding this research. The first author would like to

thank the Department of Mechanical Engineering,

University of Indonesia through the IMHERE

(Indonesian-Managing Higher Education for

Relevance and Efficiency) for the scholarship.


2012

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6. References

[1] B. Pulvirenti, A. Matalone and U. Barucca,

"Boiling heat transfer in narrow channels with

offset strip fins: Application to electronic

chipsets cooling," Applied Thermal

Engineering, vol. 30, pp. 2138-2145, 2010.

[2] S. Mori and K. Okuyama, "Enhancement of

the critical heat flux in saturated pool boiling

using honeycomb porous media," International

Journal of Multiphase Flow, vol. 35, pp. 946-

951, 2009.

[3] Z. Xu, Z. Qu, C. Zhao and W. Tao, "Pool

boiling heat transfer on open-celled metallic

foam sintered surface under saturation

condition," International Journal of Heat and

Mass Transfer, vol. 54, pp. 3856-3867, 2011.

[4] S. S. Murshed, C. N. d. Castro, M. Lourenco,

M. Lopes and F. Santos, "A review of boiling

and convective heat transfer with nanofluids,"

Renewable and Sustainable Energy Reviews,

vol. 15, pp. 2342-2354, 2011.

[5] R. A. Taylor and P. E. Phealan, "Pool boiling

of nanoflids: Comprehensive review of

existing data and limited new data,"


Transfer, vol. 52, pp. 5339-5347, 2009.

[6] H. KIM, J. Kim and M. H. Kim, "Effect of

nanoparticles on CHF enhancement in pool

boiling of nano-fluids," International Journal

of Heat and Mass Transfer, vol. 49, pp. 5070-

5074, 2006.

[7] S. You, J. Kim and K. Kim, "Effect of

nanoparticles on critical heat flux of water in

pool boiling heat transfer," Applied Physics

Letters, vol. 83, pp. 3374-3376, 2003.

[8] Y. Takata, S.Hidaka, J. Cao, T. Nakamura, H.

Yamamoto, M. Masuda and T. Ito, "Effect of

surface wettability on boiling and

evaporation," Energy, vol. 30, pp. 209-220,

2005.

[9] B. Stutz, C. H. S. M. M. d. F. d. Silva, S.

Cioulachtijan and J. Bonjour, "Influence of

nanoparticle surface coating on pool boiling,"

Experimental Thermal and Fluid Science, vol.

35, pp. 1239-1249, 2011.

[10] E. Williamson, E. Forrest, J. Buongiorno, L.-

W. Hu, M. Rubner and R. Cohen,

"Augmentation of nucleate boiling heat

transfer and critical heat flux using

nanoparticle thin-film coatings," International

Journal of Heat and Mass Transfer, vol. 53,

pp. 58-67, 2010.

[11] W. N. S. H. R. R. I. Nandy Putra, "Thermal

performance of screen mesh wikc heat pipes

with nanofluids," Experimental Thermal and

Fluid Science, vol. 40, pp. 10-17, 2012.

[12] M. Kao, C. Lo, T. Tsung, Y. Wu, C. Jwo and

H. Lin, "Copper-oxide brake nanofluid

manufactured using arc-submerged

nanoparticle synthesis system," Journal of

Alloys and Compounds, Vols. 434-435, pp.

672-674, 2007.

[13] X. Wei, H. Zhu, T. Kong and L. Wang,

"Synthesis and thermal conductivity of Cu2O

naofluids," International Journal of Heat and

Mass Transfer, vol. 52, no. 19-20, pp. 4371-

4374, 2009.

[14] S. K. Das, N. Putra and W. Roetzel, "Pool

boiling of nano-fluids on horizontal narrow

tubes," International Journal of Multiphase

Flow, vol. 29, pp. 1237-1247, 2003.

[15] S. K. Das, N. Putra and W. Roetzel, "Pool

boiling characteristics of nano-fluids,"


Transfer, vol. 46, pp. 851-862, 2003.


2012

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Analysis of Small Bubble Characteristics in Alum Solution

Warjitoa, Nurrohman

b

aMechanical Engineering Departement,

University of Indonesia

Kampus Baru UI Depok, Indonesia 16424

[email protected] bMechanical Engineering Departement,

University of Indonesia

Kampus Baru UI Depok, Indonesia 16424

[email protected]

ABSTRACT

Waste of batik processing can raise Total

Suspended Solid (TSS) that exceeds water quality

standard. One technique that can be used to reduce

TSS is bubbles flotation. The effectiveness of

bubbles flotation depends on three parameters, i.e.

the probability collision between bubbles with

particles, the particles stick to the bubbles surface

and the particles carried by bubbles. These

probabilities are greatly influenced by the bubble

characteristics, i.e. diameter, rise velocity along

column, and terminal velocity. Understanding the

dynamics of bubbles is necessary in order to

increase the effectiveness of separation of the

flotation process. The purpose of this research was

to study characteristics of small bubbles (bubbles

with diameter of 0.2-1 mm), which rise in a liquid

column. Experimental set up was a column made of

an acrylic pipe with inner diameter of 90 mm and

length of 2000 mm. Small bubbles were generated

by copper cathode. The dynamics of bubbles were

observed using a video camera. Videos images

were processed using image processing software.

The results showed that at height of 500 mm from

cathode tip bubbles in average have reached its

terminal velocity. It has been proven that effect of

alum surfactant can reduce the bubbles terminal

velocity.

Keywords : bubbles, batik, waste, flotation,

velocity, alum

1. INTRODUCTION

Water is essential for living thing. There will be no

life on earth if there is no water. Industries that

develop rapidly cause reduction in availability of

clean water. These reduction is caused by pollution

of industries‘ waste. One of the pollution is waste

of batik processing.

Batik is one of spesific characteristics of

Indonesian. People in the world have admitted that

batik is one of Indonesian culture heritages.. This

has been proven with authentication by UNESCO.

Day after day batik industries develop. This

development makes waste of batik processing

increases. The waste of batik processing causes

increasing Total Suspended Solid (TSS) of water,

therefore exceed the water quality standard [1]. A

technique for decreasing TSS is needed therefore,

at least, the TSS does not exceed the water quality

standard. One of the techniques that can be used is

bubbles flotation.

Microbubble is bubble which is less than 200 µm in

diameter. Bubbles with more than 200 µm but less

than 1 mm in diameter are called ―small bubble‖,

whereas bubbles with more than 1 mm in diameter

are called ―large bubble‖. Bubble flotation is a

process conducted by producing bubbles from

bottom of water column contaminated by waste

particles. Bubbles will rise because of Bouyancy

force and collide with the particles. These particles

stick to the bubble and will be carried by the

bubbles to water column surface. When these

particles gather in the surface column, it can be

separated from the water easily.

The efficiency of the bubbles flotation depends on

three parameters, i.e. the probability of collision

between bubbles with particles, the particles stick

to the surface of the bubbles and the particles

carried by the bubbles [2]. The higher value of

these probablities, the higher the efficiency of the

flotation process. Microbubble prefers to be used

because it has larger surface area therefore the

probability of collision between bubbles and

particles can be increased.

Today, the aplication of bubbles in flotation is still

troubled by the understanding of the bubbles

characteristics. Therefore, it is necessary to do a

research that studies the characteristics of bubbles

rise in liquid column. These charactersitics include

diameter, velocity profile along liquid column, and

terminal velocity of the bubbles. This research was

to study the characteristics of ―small bubbles‖ rise

in liquid column with addition of surfactant, i.e.

alum or aluminum sulfate.

IMAT-UI 017




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2. THEORITICAL ASPECT

2.1 Bubble

Bubble is a particle whose dispersed phase is gas.

Bubble shape can be divided by three, i.e.

spherical, ellipsoidal and cap. Bubble motion can

be rectilinear, zigzag and spiral [3].

2.2 Dimensionless Numbers

The movement of the bubbles is usually in

terms of dimensionless numbers, i.e.:

(1) Reynods Number

Reynods number is ratio of inertia force to

viscous force.

Re = ρ.v.D/µ (1)

(2) Eotvos Number

Eotvos number is ratio of gravity force to

surface tension force.

Eo = g.Δρ.D/σ (2)

(3) Weber Number

Weber number is ratio of inertia force to

surface tension force.

We = ρf.vb2.D/σ (3)

(4) Morton Number

Morton is ratio of viscous force to surface

tension force.

Mo = g.µ4.Δρ/ρ

2.σ (4)

Where ρ, v, D, µ, Δρ and σ is liquid density, bubble

velocity, bubble equivalent diameter, liquid

dynamic viscocity, density difference between gas

inside bubble and liquid where bubble moves and

bubble surface tension [3].

2.3 Surface Tension

Surface tension works on surface plane, normal or

perpendicular to each line that works on surface

and its magnitude is the same at every point.

Surface tension will decrease at a certain

temperature on the surface of two substances and

when temperature rises [4]. Bubble surface tension

is defined as:

σ = pσ.R/2 (5)

where pσ and R are the surface tension pressure

and bubble radius [5]

2.4 Terminal Velocity

At the first, bubble will accelerate after leaving the

tip of the cathode. At some point, it will experience

terminal velocity. It is when the speed constant

when the influence of body force (gravity) and drag

force equal to the buoyant force. Sam et al. (1996)

has conducted research on single bubble velocity

and characterize a three-stage velocity as shown in

Figure 1 [6].

Figure 1: Velocity stages of bubbles rise in

liquid column

Hadamard-Rybczynski have formulated terminal

velocity in creep flow as follows:

u∞ = (6)

where u∞, g, r, Δρ, μ, and κ are the terminal

velocity, the acceleration of gravity, the bubble

radius, the difference in density bubble where the

bubble moves with the fluid, dynamic viscosity and

dynamic viscosity ratio of the particles to fluid. The

above formula applies to bubble with mobile

surface [7].

In addition to the above formula, Stokes (1880)

formulated an equation for bubble terminal velocity

where the surface does not move (immobile) as:

(7)

with u∞, de, ρl, ρg, and μl are the terminal velocity,

the equivalent diameter of the bubble, liquid

density where the bubble moves, the density of the

gas inside the bubbles and the dynamic viscosity of

the liquid where the bubbles move [8]. Davies-

Taylor (1950) have formulated an equation for

large bubble where the dynamic surface tension

and viscosity can be neglected:

(8)

with u∞ and de are the terminal velocity and the

equivalent diameter of the bubble [8].

2.5 Effect of Surfactant

Researchs have proven that the properties and

behavior of bubbles will change significantly in the

contaminated fluid. Effect of surfactant used to be


2012

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compared in the term of terminal velocity of the

bubble. Surfactants tend to inhibit movement of the

bubble internal circulation thereby reducing the

terminal velocity. With a low terminal velocity, the

probability of attachment of particles can be

increased so that the flotation process efficeiency

increases.

2.6 Electrolysis

Hydrogen bubbles can be produced by separating

hydrogen and oxygen from water molecules by

electrolysis method. Electrical energy can be used

to separate hydrogen. It is made by supplying

electric current to an electrolitic cell. Thus,

electrolysis is the course of events of a chemical

reaction by an electric current. The tool consists of

electrolytic cell containing electrolyte (solution or

melt), and two electrodes (anode and cathode). In

the anode, oxidation happens while in the cathode

reduction happens. In an electrolysis experiment,

the reaction at the cathode depends on the tendency

of reduction.

2.7 Flotation

Bubbles attached to the hydrophobic particles and

carry the particles to the surface of the liquid, in

which particles are removed by using skimming

equipment. There are two classifications based on

the size of the bubble flotation. In the microbubble

flotation, bubbles are usually used 10-70 µm in

diameter. The bubbles are often generated by

depressurization of the dissolved air of liquid

(water dissolved liquid). This process is called

dissolved air flotation (DAF). In dispersed air

flotation, bubbles used was 1 mm.

The main key of flotation process is in bubbles-

particles capture which commonly known as a

series of three subprocesses. Total capture

efficiency is usually known as the product of three

successive steps, i.e. the collision efficiency, the

sticking efficiency, and the efficiency of stability

the bubble-particle aggregates [9].

3. EXPERIMENTAL METHOD

Experimental scheme is shown by Figure 2.

Figure 2: Experimental equipment scheme

3.1 Flotation Column

Column flotation was made of acrylic pipe with

inner diameter 8.4 cm, thickness 0.6 cm and height

of 200 cm. Water jacket was used to overcome the

optical distortions with dimension of 26.2 mm x

26.2 cm x 200 cm.

3.2 Surfactant and Water

The surfactant used was alum or aluminum sulfate

to the level of 100 grams per liter of water. Water

which was used to fill the acrylic pipe was drinking

water from AQUA. Meanwhile, water used to fill

the water jacket was tap water.

3.3 Bubbles and Surfactant Producing

Equipment

The bubbles were produced by electrolysis. The

electrodes are copper wire. For cathode, the wire

diameter was 0.1 mm, whereas for the anode the

wire diameter was 0.2 mm. DC Power Supply was

KPS3030DA of ATTEN instrument.

3.4 Camera Mechanisms

The camera was a Nikon D5000 with a Nikor 60

mm lens AF f/28D. Guide ways was used to trace

the bubbles as shown in Figure 3.


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3.5 Image Processing

Data processing was done using ImageJ software to

measure the size of the bubble and Frame Shot

software to determine the time required for the

bubble to move.

3.6 Lighting

For the lighting in order that the produced bubbles

images are good, the back lighting technique by

placing diffusive paper on the water jacket on the

opposite side of the installed camera was used.

3.7 Tools Set Up

At first, both edges of the electrodes were sanded to

remove the insulation. One edge was connected to

a DC Power Suply whereas the other edge was

inserted into the column flotation. The cathode is

inserted from the bottom of the column to generate

bubbles and the anode inserted from the top of the

column. Then the water from the AQUA mixed

with surfactant and acrylic was inserted into the

acrlylic pipe. Whereas tap water put into the water

jacket. Cameras mounted on one side of the water

jacket on the guide ways. To determine the size of

the bubble, some lines were given on the diffusive

paper. To determine the position of the bubble,

water jacket was marked on certain points. To

determine the distribution of bubbles, the camera

was placed at 12.5 cm, 25 cm, 75 cm, 125 cm, and

160 cm from the tip of the cathode to capture the

bubble at these points. The focus of the camera lens

was set manually whereas the exposure and shutter

speed was adjusted automatically.

Figure 3: Guide ways

3.8 Data Collection Procedures

To get the data distribution of bubbles, the camera

was placed at 12.5 cm, 25 cm, 75 cm, 125 cm and

160 cm. Once the camera was mounted on one

point, DC Power Supply switch was turned on for

about a second and then turned off quickly. The

generated bubbles were quite a lot and the picture

was taken for the image bubbles at the top, middle

and bottom. Then the stored image data were saved

to be processed using ImageJ software.

To get the bubble velocity data, the camera was

placed at the level of a 12.5 cm and was set to be

ready to record. Then DC Power Supply switch

was turned on then immediately shutted down. The

traced bubbles was the top one. It is because the

bubbles are more likely affected by the other

bubble on top of it. Then the data were stored to be

processed using the Frame Shot software.


4.1 Bubble Size Distribution

From the data processing, for the distribution

of bubbles, it was obtained result as shown in

Figure 4.

Figure 4: Bubble size distribution graph

From the graph of Figure 4, it can be seen that

the higher the position of the bubble, the greater

diameter. This is because the hydrostatic pressure

decreases. The resulting bubbles ranging from 0.27

mm to 0.4 mm. All of them are small bubbles.

4.2 Bubble Rise Velocity

Bubble rise velocity was varied with voltage 3 and

7.5 volts. The result for voltage of 3 volts was

shown in Figure 5. As for the voltage of 7.5 volts,

the result was shown in Figure 6. From these data it

can be seen that most of the line tends to be

straight. This means that the bubble has reached

terminal velocity. For some graphs, the velocity at

a height of 25 cm is steeper than most of the other

points. This shows that the bubbles on a height of

25 cm still deaccelerate to reach terminal velocity

or by Sam et al. (1996) are in stage 2 to reach the

stage 3. For the graph with a voltage of 7.5 volts

bubble looks more in line with the results shown by

Sam et al. (1996). This is because the diameter of

the bubble bigger so it is easier to visualize it. From

the data processing the velocity of the average

bubble either to the voltage of 3 volts and 7.5 volts


2012

101

at a height of 25 cm were in the second stage, in

which the declaration occurs. While at a height of

50 cm, the average bubble has reached its terminal

velocity (stage 3).

Figure 5: Bubble rise velocity with voltage of

3 volts

Figure 6: Bubble rise velocity with voltage of

7.5 volts

4.3 Terminal Velocity

4.3.1 Comparison Between Generated Bubbles

From the data available, it can be determined the

estimated terminal velocity of the bubble

generated, i.e. by averaging velocity values at the

points that tend to be straight. From the data

processing, Figure 7 shows a graph of the results of

the bubble terminal velocity with voltage of 3 and

7.5 volts. It can be seen that the average terminal

velocity of bubbles generated by voltage of 3 volts

are smaller than those indicated by voltage of 7.5

volts. This is because the bubbles produced by a

voltage of 7.5 volts were bigger than the bubbles

generated by the 3 volts.

Figure 7: Graph of terminal velocity vs

diameter voltage of 3 volts and 7.5

volts 4.3.2 Effect of Surfactant

Graph of Figure 8 is the result of experiments

conducted by Huang et al. (2011) for a comparison

of the value of the terminal velocity of small

bubble in pure water to the results of this study [9].

The data experimental results by Huang et al. are

indicated by the circle. From experimental result by

Huang et al., the terminal velocity for bubbles with

the same size in this study was lower than those in

the result by Huang et al.. This shows the effect of

alum or aluminum sulfate contaminant in this

study. As can be seen from graph of Figure 8, for

pure water with bubbles of 0.5 mm in diameter, it

has terminal velocity about 10 cm/s, whereas the

results from this study as shown in the graph of

Figure 7 only ranged from 2 cm/s to 2.5 cm/s.


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Figure 8: Graph of voltage terminal velocity vs

diameter ratio

4.3.3 Comparison between Experimental Result

and Some Equations

Terminal velocity occurs when the drag and

bouyancy force are equal. Therefore, it can be

derived a simple equation of terminal velocity of

the bubble in such conditions by incorporating the

existing parameters. The equation in this question

is:

FB = Fd (9)

With the bouyancy force (FB) is formulated as:

FB = ρ.V.g (10)

with ρ is the density of the fluid where the bubble

moves, V is the volume of the bubble and g is the

acceleration of gravity. Whereas the drag force Fd

is formulated as:

Fd = CD.1/2.ρ.UT2.A (11)

where CD is drag coefficient and A is the surface

area in this case the value is equal to πR2, where R

is the radius of the bubble. By incorporating the

existing values into the equation, the equation of

terminal velocity of the bubble obtained is:

u∞ = 0.258 (12)

with a value judgment of CD based on graph CD vs

Re for solid sphere as shown in graph of Figure 9

[10].

Figure 9: Graph of CD vs Re for solid sphere.

From the graph of Figure 9, CD value that was

taken was 20 due to Re obtained from the results of

the study has value of 22 in average. By inserting

the CD values, Equation (12) was obtained. The

graph of Figure 10 is a comparison between the

experimental results, equation (7), (8) and (12) for

a voltage of 3 volts and the graph of Figure 11 for a

voltage of 7.5 volts.

Figure 10: Graph of comparison between

experimental result and some

equations for voltage of 3 volts.

Figure 11: Graph of comparison between

experimental result and some

equations for voltage of 7.5 volts.


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From graph of the Figure 10 and 11, either bubbles

produced by 3 volts or 7.5 volts, both of them are

between graph of equation (8) (Davies-Taylor) and

equation (12). However, for a voltage of 3 volts the

graph is more likely to approach the equation (12)

whereas for a voltage of 7.5 volts tends to be right

in between them. This is because bubbles generated

with a voltage of 3 volts are smaller than that

produced by a voltage of 7.5 volts. This is

reasonable because the Davies-Taylor equation is

for the application of a large bubble. Bubble graph

experimental results are all above the graph of

equation (12). This suggests that there may be

some motion on the surface of the bubble (mobile

surface) so that the value of its terminal velocity

greater than equation (8) where equation (8) is for a

solid sphere where there is no motion on the

surface (immobile surface) or maybe it was mobile

but the value of Re taken was too small for a

voltage of 7.5 volts so that the graph of

experimental result a little bit away from the graph

of equation (12). Stokes equations are far away

from the experimental results. It has been

mentioned that Stokes equation accurates for very

small bubbles, such as microbubbles.

5. CONCLUSION

A study on the characteristics of small bubble

in a solution of alum has been performed and

analyzed. From the analysis and data processing it

was shown that the bigger bubble the bigger its

terminal velocity. It was known that the average

bubble has reached its terminal velocity at a height

of 50 cm above the cathode. It has also been proven

that the effect of surfactant alum can reduce the

terminal velocity of bubbles. With the low value of

the terminal velocity, the efficiency of flotation

process can be improved.

REFERENCES

[1] F. Astuti, ―Pengolahan Limbah Cair Industri

Batik dengan Koagulan dan Penyaringan

(Studi Kasus di CV. Batik Indah Rara

Djonggrang)‖, Postgraduate thesis

Environtmental Science Program

Interdisciplineary Group, Gadjah Mada

University, 2004.

[2] K. A. Matis, Flotation Science and

Engineering, CRC Press, Inc., 1995.

[3] R. Clift, J. R. Grace, and M.E. Weber,

Bubbles, Drops, and Particles. New York:

Academic Press, Inc., 1978.

[4] C. Brücker, ―Structer and dynamics of the

wake of bubbles and its relevance for bubble

interaction‖, Phys. Fluids, 11, pp. 1781-179,

1999.

[5] T. G. Leighton, The Acoustic Bubble.

California: Academic Press, Inc., 1997.

[6] A. A. R. Mehrabadi, ―Effects of frother type on

single bubble‖, Postgraduate Thesis Department of

Mining, Metals and Materials Engineering,

McGill University, Montreal, Canada, 2009.

[7] L. Parkinson, R. Sedev, D. Fornasiero, and J.

Ralston, ―The terminal rise velocity of 10–100

µm diameter bubbles in water‖, Journal of

Colloid and Interface Science 322, pp. 168–

172, 2008.

[8] A. R. M. Talaia, ―Terminal velocity of a

bubble rise in a liquid column‖, World

Academy of Science, Engineering and

Technology, 2007.

[9] Z. Huang, D. Legendre, and P. Guiraud, ―A

new experimental method for determining

particle capture efficiency in flotation‖, Journal

of Chemical Engineering Science 66, pp. 982 –

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2012

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Effect of Hot Air Reservoir in the development of Vacuum freeze

drying M. Idrus Alhamid

1, Nasruddin

2, Muhamad yulianto

3

Mechanical Engineering Department University of Indonesia

Kampus UI Depok 16424, INDONESIA

Ph. +62 21 7270032, Fax. +62 21 727003

Email: [email protected] 1, [email protected]

2,[email protected]

3

ABSTRACT The Objective of this work is to know effect of

inserting hot air from reservoir to the process of

vacuum freeze drying. Tentacle of jelly fish as

sample with constant weight 50 g and placed at

container which isolated, the samples were freeze

dried with condition at experiment varying between

inserting and without inserting hot air at

temperature 27oC. The result of experiment shows

that while inserting hot air into vacuum freeze

drying make pressure rise in until pressure reach 40

mbar. And this phenomena make material

evaporation and this event cant be done in vacuum

freeze drying. And when without hot air reservoir

the pressure can reach 3.5 mbar and the subimation

can be done in this process. Vacuum freeze drying

process without hot air reservoir need time 12.5

hours and for vacuum freeze drying with hot air

reservoir need time 10.5 hour to drying 50 g of

jelly fish tentacle. From this experiment can be

concluded that for vacuum freeze drying with

inserting hot air need more ability of vacuum pump

specially in flowrate and ultimate vacuum.

Keywords : Vacuum Freeze Drying, Hot Air

Reservoir, Tentacle of Jelly Fish.

1. INTRODUCTION

Freeze vacuum drying is a dehydration process in

which water removed by sublimation of ice from

frozen materials directly at low pressure (Vacuum

pressure). Freezing and sublimation process starts

from outside surface and then to the material

recedes[1]. The others researcher also describes

that Vacuum Freeze drying (VFD) is an optimal

drying technology method because it maintains the

structure, nutrients, and color of the original

substance [2, 3, and 4]. Vacuum freeze drying

process consists of three processes that are freezing

process, primary drying and secondary drying [5].

In freezing process, actually use vacuum freezing

method. It is a freezing process based on the rapid

evaporation of moisture from the surface and

within the products due to the low surrounding

pressure below saturation pressure of the product.

This methode have a problem because of Vacuum

freezing caused high evaporation and mass loss,

this phenomenon for any product make a serious

damage [6]. In the other side, to reduce energy

consumption in vacuum freeze drying process,

manny of researcher has been declare the

innovation to solve this problem, that are : Adding

energy from electrical heating at lower and top

possition [7], adding energy from micowave to the

system [8, 9, 10], adding energy from infra red

radiatio [11], and the news one is adding energy

from heater from condenser‘s heat loss [12]. About

combination vacuum freezing and internal freezing

to reduce mass loss due to evaporation and also

using heater from condenser heat loss also have

been described to solve the problem in vacuum

freeze drying, and have a diagram can be seen at

figure 1 [13].

Figure 1. Diagram P-T of vacuum freeze

drying using internal freezing and

heater from condenser heat loss

There is one methode which is never done by other

research, using hot air reservoir to reduce energy

consumption. Fress air with high temperature

which entering the drying chamber will be

increasing the molar mass of air at drying chamber

and due to this phenomena create differences

mollar mass between material and air at drying

chamber and the moisture content the material will

be evaporate.

Due to the problems above, the objective of this

research is to analyze the efeect of hot air reservoir

at vacuum freeze drying process. The material in

this research using jelly fish (scyphomedusae) as

basic ingredient of medicine.

IMAT-UI 018





2012

105

2. MATERIAL AND METHODE

2.1. Sample Preparation

Before tested using vacuum freeze drying, jelly fish

is first stirred as seen in figure 2 and checking the

initial moisture content and weighed the Initial and

final mass measured using digital balance with

range 0–600 g and accuracy of 0.4%. For each

experiment, 50 g of samples was used on vacuum

freeze drying process.

Figure 2. Initial condition of jellyfish tentacles

2.2. Experimental Set Up

A compact vacuum freeze drying with internal

cooling and hot air reservoir from condenser heat

loss was designed, built and installed at

Department of Mechanical Engineering, University

of Indonesia (Fig 2 and 3). It consist of 3 Insulated

cylindrical, drying, cold trap and reservoir chamber

with a diameter and a length of 25 cm. Cold trap is

a cascade refrigeration with refrigerant at High

stage R22 and at low stage mixture between

HCR22 80% and CO2 20%. Cold coil installed at

drying, cold trap and reservoir chamber. Heat from

condenser‘s heat loss also installed at drying

chamber with same coil. To shut on and shut off

cold and heat coil using shut off valve. A pressure

transmitter PTX 1400 with an accuracy of 0.4%

and measure range 0-1600mbar was used to

measure pressure in the drying chamber. A tray of

material used from Teflon and insulated to keep the

heat transfer during process. The temperature at

drying chamber, cold trap, material, cascade

refrigeration system and reservoir were monitored

by thermocouple type K with accuracy 0.4%. Dial

pressure used to monitor pressure in cascade

refrigeration system. The thermocouple and

pressure transmitter connected to Data Acquisition

which is have number of slot 4, total power 15 w

and operating temperature -20oC until 55

oC.

Figure 3. Experimental Set Up

Figure 4. Schematic of experiment 1. Vacuum

pump , 2. Coldtrap, 3. Drying

Chamber, 4. Compressor at LS stage,

5. PHE, 6. Expansion valve (Needle

Valve), 7. Check Valve, 8.

Evaporator HS, 9. compressor HS,

10. Evaporator HS, 11. Expansion

valve HS (Needle Valve), 12.

Material tray, 13. Capilarry tube, 14.

Flow Meter, 15. Hot air reservoir,

16-26. Thermocouple, 27-28.

Pressure Transmitter


3.1. Temperature characteristics

The characteristic of product temperature during

vacuum freeze drying process can be seen at Figure

4. Product temperature devided into 3 region, that

is freezing region, sublimation region dan

secondary drying region as the other research

mentioned [14]. At process vacuum freeze drying

with internal cooling has longer time to take

material freeze than the vacuum freezing and this is

also make the energy consumption more higher

than vacuum freezing. As can be seen at the figure

1, primary drying occurs at constant pressure and

temperature. For vacuum freeze drying with

internal freezing and heating at room temperature

40oC, primary drying occurs at product temperature

-10oC. For Vacuum freeze drying with hot air

reservoir, primary drying occurs at product

temperature 1oC. At figure also can be seen that

vacuum freeze drying with internal cooling and


2012

106

heat from condenser heat loss take time longer than

vacuum freeze drying with hot air reservoir.

Figure 4. Product temperature during vacuum

freeze drying

3.2. Phase change During Process

At figure 5 can be seen the change phase of product

during vacuum freeze drying process. For process

with internal cooling and heat from condenser heat

loss product temperature reach he solid phase until

product temperature -20oC without mass loss due to

evaporation and after that change phase to gas

region by the sublimation process at temperature -

20oC and pressure of 3 mbar. This process is

occurs in all vacuum freeze drying prosess as

general. For process with hot air reservoir the

product temperature can not reach the ice / solid

region due to the pressure rise in after entering hot

air from reservoir. After that the product

temperature change phase from liquid to gas phase

at temperatur 1oC and pressure of 20 mbar, and

this process is evaporation not sublimation. This

process can not called as vacuum freeze drying but

vacuum drying.

Figure 5. Phase change diagram of product

during process

3.3. Cold trap Temperature

At figure 6 can be seen the temperature of coldtrap

at vacuum freeze drying process with internal

cooling and heating and also hot air reservoir. For

vacuum freeze drying process with internal cooling

the cold trap temperature has fluctuate graphing at

begining due to the refrigerant devided in to 2 part

for cold trap and drying chamber (double

evaporator) and also process of flash point. The

temperatur in this variation can rech of -25oC. For

vacuum freeze drying process with hot air reservoir

the temperature of cold trap can reach temperatur -

45oC, becuse in this process without internal

freezing and the function of coldtrap still on single

evaporator. All refrigeration condition at this

experiment is same condition, that is for High

Stage using refrigernt R22 and for Low Stage

mixture between HCR 22 (80% mass) and CO2

(20% masss)

Figure 6. Profile of cold trap temperature

3.4. Final Product and moisture content

Figure 7 and table 1 can be seen the final product

of vacuum freeze drying with variation internal

freezing and heater (A) and also hot air reservoir

(B). Product of vacuum freeze drying with internal

cooling and heater has color very white and smooth

this result due to the process of vacuum freeze

drying at sublimation process . For misture content

of material is approximately 0%. For process with

hot air reservoir the final product has color more

dark and moisture content is 0.15%

Figure 7. Final Product jelly fish


2012

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Table 1. Mositure Content Final Product

No Variation Initial

mass

(g)

Final

Mass

(g)

Mositure

Content

1 Vacuum

freeze

drying with

internal

freezing

and heating

50 2.289 0.18%

2 Vacuum

freeze

drying with

Hot air

reservoir

50 3.25 2.1%

CONCLUSION :

1. Vacuum freeze drying at the process

devided 3 region, freezing, primary

drying / sublimation and secondary

drying

2. Adding hot air to the system can be

reducing drying time but the process

become evaporation not sublimation

3. Adding hot air to the system make the

moisture contaent higher than vacuum

freeze drying with internal cooling and

heating

ACKNOWLEDGMENTS

The authors acknowledge the financial support

from Ministry of Higher Education Indonesia

(DIKTI) by Strategis Nasional Grant with no

contract : 3398 / H2.R12/HKP.05.00/2012

Reference :

[1] A.B. Edinara, R.M. Filho, E.C.V. De Toledo,

Freeze drying process : real time model and

optimization, Elsavier International Journal

Chemical Engineering and Process 43 (2004)

1475-1485

[2] George.J, P., A,K,Datta., 2002. Development

and validation of heat and mass transfer

models for freeze-drying of vegetable slices.

Elsevier Journal of Food Engineering (52),

89-93

[3] Chakraborty,R,.,A,K,Saha.,P, Bhattacharya.

2006. Modeling and simulation of parametric

sensitivity in primary freeze-drying of

foodstuffs. Elsevier Separation and

Purification Technology (49), 258-263

[4] Ghio, S., A,A, Barresi, G, Rovero.2000. A

comparison of evaporative and conventional

freezing prior to freeze-drying of fruits and

vegetables. IChem Journal., 0960-3085

[5] George-Wilhelm Oetjen., Peter Haseley.,

2004. Freeze Drying Second, Completely

Revised and Extended Edition. WILEY-VCH

Verlag GmbH&Co. KGaA, Weinheim ISBN:

978-3-527-30620-6

[6] Jackman, Patrick., Da-wen Sun., Jivun

Zheng., 2007. Effect of combined vacuum

cooling and air blast cooling on processing

time and cooling loss of large cooked beef

joints. Elsevir International Journal of Food

Engineering. (32). 266-271.

[7] Belyamin, Tambunan, A.H., Hadi, K.,

Purwadaria, & M.I. Alhamid, 2007, ―The

application of freeze vacuum and heating

from top and bottom on vacuum freeze

drying‖, Journal of Agricultural Engineering,

Association of Agricultural Engineering

Indonesia, Vol. 21, 235-248

[8] Wang, Rui., Min, Zhang.,Mujumdar. 2010.

Effects of vacuum and microwave freeze

drying on microstructure and quality of potato

slices. Elsevier Journal of Food Engineering

(101), 131-139

[9] Duan, X., Zhang, M., Li,X., Mujumdar, AS.,

2008b. Ultrasonically enhanced osmotic

pretreatment of sea cucumber prior to

microwave freeze drying. Drying Technology

26 (4), 420-426

[10] Huang, Lue-lue., Min, zhang., M,Mujumdar.,

Rui, X,L. 2011. Comparison of four drying

methods for re-structured mixed potato with

apple chips. Elsevier Journal of Food

Engineering (103), 279-284

[11] Chakraborty, R., M. Bera., P.

Mukhopadhyay., P. Bhattacharya. 2011.

Prediction of optimal condition of infrared

assisted freeze-drying of aloe vera (Aloe

barbadensis) using response surface

methodology. Separation and purification

technology. 80 (2011) 375-384.

[12] Nasruddin., M. Idrus Alhamid., Engkos A

Kosasih., M. Yulianto. 2011. Effect of Freeze

Vacuum Drying and Heating from

Condenser‘s Heat Loss on Drying Rate and

Microstructure of Aloevera. Research Journal

of Applied Sciences 6 (5) : 335 - 343

[13] M. Idrus Alhamid, Muhamad Yulianto,

Nasruddin, Engkos A. Kosasih. 2012

Development of a Compact Vacuum Freeze

Drying for Jelly Fish (Schypomedusae).

Jurnal Teknologi UTM-Malaysia 58 (2012)

Supl 1, 25 – 32. ISSN 0127-9696

[14] May, JC.. 2004. Freeze-Drying /

Lyophilization of Pharmaceutical and

Biological Products (Second Edition, Revised

and Expanded). Marcel Dekker, Inc. ISBN :

0-8247-4868-9


2012

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Experimental of Cascade Refrigeration System Using Natural

Refrigerant Mixture Ethane and Carbon Dioxide at Low

Temperature Circuit and Natural Refrigerant Propane at High

Temperature Circuit

Nasruddin*, M. Idrus Alhamid, Darwin R.B. Syaka and Arnas

Refrigeration and Air-Conditioning Laboratory, Mechanical Engineering Department –

Faculty of Engineering - University of Indonesia, Kampus UI Depok, 16424, Indonesia *E-mail: [email protected]

ABSTRACT Medicine and biomedical research activities require

cold storage (cold storage) to store biomedical

specimens such as, for example, stem cells (stem

cells), sperm, blood and other organs. During

storage, to prevent the specimen from damage

required a special cold storage reaches -80oC [1].

Using single cycle refrigeration machine can only

reach -40oC, and performance deteriorates below -

35oC drop in pressure associated with evaporation.

Thus, to reach lower temperatures, use cascade

refrigeration machine [2]. During this low-

temperature circuit cascade refrigeration systems still

use refrigerants that contain ozone-depleting or

global warming (CFCs and HCFCs). To overcome

this, a mixture of carbon dioxide and ethane

azeotropis a promising alternative refrigerants.

Simulation studies and experiments indicate a mixture

of carbon dioxide and ethane were able to achieve the

minimum temperature to -80oC [4-7]. With the mass

ratio 70% R170 and 30% R744 circuit at low

temperature refrigeration systems and uses a

capillary tube expansion device 0.054 inch diameter

with a length of 6 meters and 3 meters then use an

electric heater as the cooling load. Cooling load is

given by the variation of 90 W, 120 W and 150 W at a

cabin in the low temperature circuit. From the

experiment will be known characteristics of cascade

refrigeration system with refrigerant mixture and will

get the parameter data to make cascade refrigeration

machine.

Keywords : Cascade, Refrigerant, Ethane, Co2,

Capillary tube

1. INTRODUCTION

Cascade refrigeration system consists of at least two

refrigeration systems that work independently. Two

refrigeration systems are connected in cascade heat

exchanger where the heat is released in the condenser

circuit low temperature (low temperature circuit /

LTC) is absorbed from the evaporator temperature

circuit (high temperature circuit / HTC) [3]. During

this time at a low temperature circuit used CFC

refrigerants such as R13 or R503 banned for

damaging the ozone layer. Meanwhile, HFC

refrigerants such as R23 alternatives although it does

not contain ozone-depleting substances, but the cause

of global warming. So, look for alternative

refrigerants directed on natural refrigerants and one of

which is carbon dioxide [5].

Carbon dioxide has an advantage because it is not

toxic, not flammable (non-flamable), easy to get, no

ozone depletion potential and very low global

warming [5]. However, the high pressure and

temperature triple preclude the use of carbon dioxide

when used for low-temperature circuit [6]. The

solution to overcome this shortcoming is to mix

carbon dioxide with other natural refrigerants are

hydrocarbons. To overcome this, a mixture of carbon

dioxide and ethane azeotropis a promising alternative

refrigerants. Simulation studies and experiments

indicate a mixture of carbon dioxide and ethane were

able to achieve the minimum temperature to-80oC [4-

7].

If the alternative refrigerant is used in a refrigeration

system, each component of the system must be

redesigned for reliability and high efficiency. In this

particular cascade refrigeration system at low

temperature circuit using the capillary tube expansion

device. The capillary tube is a tool used in the

expansion generally makes a small cooler as air

conditioning, refrigeration and cold storage, because it

is cheap, simple and reliable [8].

This study aims to develop a low temperature cold

storage for applications in the biomedical field using a

mixture of carbon dioxide and ethane refrigerants that

have high energy efficiency and safe that has a low

flammability and non-toxic for use in low-temperature

circuit in cascade refrigeration system.

2. METHODS

Cascade refrigeration machine consists of two

refrigeration circuits, a circuit called the refrigeration

circuits of high temperature (high temperature circuit /

HTC) that will contain environmentally friendly

refrigerant propane (R290) and the low temperature

circuit (low temperature circuit / LTC). which will be

filled with the refrigerant mixture R744/R170 the

azeotropic composition.

IMAT-UI 019


2012

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Refrigerant is a closed cycle of cascade refrigeration

system comprising a compressor, oil separator,

kondesnser, PHE, filter dryer, expansion valve,

evaporator, and accumulator, can be seen in Figure 2.

At the position of the high temperature refrigerant

discharge circuit and oil through the oil separator and

only high-pressure refrigerant gas phase can exit to

get to the condenser. Refrigerant is cooled by the

ambient temperature so that the refrigerant out of the

condenser has a temperature equal to the ambient

temperature. After going through the condenser

refrigerant dryer filter before heading into the

expansion valve. Refrigerant is expanded according to

the temperature needed to cool the cascade condenser

or heat exchanger between the high-temperature

circuit with low temperature circuit. Heat exchanger

that is used is a type of PHE (plate heat exchanger).

Out of PHE refrigerant back to the compressor.

Temperatures were measured at high temperature

circuit only four positions are out compressor,

condenser exit, entry and entry kompreser PHE.

Temperature measurement using a thermocouple type

K with the value of reading ± 0.14% accuracy. While

pressure is measured only on the position of the

compressor discharge and suction. Measurement of

pressure using a pressure transmitter type Druck PTX

1400 readings with an accuracy of ± 0.15%.

Figure 1. Scheme diagram of refrigeration

cascade system At the position of the low temperature circuit

refrigerant compressor exit directly to the condenser

to be cooled to ambient temperature, it aims to ensure

that oil enters the oil separator in the liquid phase so

as not to be drawn into the whole system. Once out of

the oil separator directly cooled condenser refrigerant

cascade (PHE) so that the temperature is low. Then

the refrigerant through the filter dryer before heading

capillary tube to be expanded according to the desired

temperature. At this low temperature circuit system

desired temperature is -80oC. Very low temperature

refrigerant goes to the evaporator. Evaporator is inside

the cabin where the cabin is made from a fan to blow

air into the evaporator. Then the refrigerant flow to

the accumulator before getting back into the

compressor to be pressed. Temperature measurement

at low temperature circuit is positioned in the

compressor exit, entry PHE, PHE exit, before the

capillary tube, evaporator entry, exit evaporator,

inside the cabin and into the compressor. Temperature

measurement using a thermocouple type K with the

value of reading ± 0.14% accuracy. For the

measurement of the pressure placed on four parts:

compressor exit, enter the capillary tube, out of the

capillary tube and into the compressor. Measurement

of pressure using a pressure transmitter type Druck

PTX 1400 readings with an accuracy of ± 0.15%.

The air inside the cabin is heated by a 500 watt heater

is blow by a fan. Heater is controlled by a dimmer to

create a stable electrical current into the heater that

will be used as the cooling load. Pictures of the cabin

can be seen in Figure 3. The flow of air in the cabin is

measured by anemometer. The temperature is

measured with a thermocouple type K. The data

results from performance refrigeration system are

pressure, temperature and mass flow rate. The data

obtained will be used as the design and operating

parameters of cascade refrigeration system.

T

TEvaporator

Heater

Figure 2. Scheme of cabin


The results of this data collection illustrate the

character of the cascade refrigeration system circuit

high temperature and low temperature circuits.

Refrigeration system has a small capacity so it can be

put in the room. Refrigerant substitute experiments

were carried out only with the modification to the

capillary. The experiment was started with Ethane in

LTC to set up the base reference for further

comparisons with new mixture under identical

working conditions while HTC was kept at same

condensing and evaporating pressure. The diameter of

capilary tube is 0.054 inch and its length is 3 meter

dan 6 meter.


2012

110

Figure 3. Variation of cooling load (heater) with

the refrigerant temperature at the

evaporator inlet in LTC.

In Figure 3. It can be seen the influence of the cooling

load on the evaporator entrance temperature low

circuit temperatures. The addition of cooling load

resulting evaporator temperature inlet increases.

Using a capillary tube 0.054 inch and a length of 6

meters evaporator inlet temperature is -80.6oC when

given cooling load by 90 W and the temperature

continues to increase to-78.4oC when the cooling load

is added to 150 W. The same capillary with a length

of 3 meters and temperatures higher evaporatornya-

70.5oC when the cooling load is given at 90 W and

continue to increase with the cooling load is added.

the evaporator inlet temperature by using a capillary

tube 6 meters is lower than the capillary length 3

meters and it is shown that the longer of capillary tube

the lower evaporator inlet temperature. The

evaporator inlet temperature affects the temperature in

the storage room or cabin and can be seen in Figure 4

below.

Figure 4. Relation of evaporator inlet

temperature with room storage

temperature.

The temperature of the storage room have the same

tendency to the evaporator inlet temperature at low

circuit temperatures when increasing the evaporator

inlet temperatures rise due to the cooling load

increases. At LTC using a capillary tube with a length

of 6 meters shown in figure 4. temperature of storage

room when load 90 W is -75.7oC and the evaporator

inlet temperature is -80oC and then the cooling load

was increased to 150 W resulted in a storage room

temperature is 71.9oC and the evaporator inlet

temperature is -78.4oC. The capillary tube length 3

meters has a trend similar to the capillary tube length

of 6 meters where the temperature difference between

the evaporator inlet temperature and the storage room

temperature increasing when the cooling load is

increases.

In figure 5 seen influences of the cooling load at the

evaporator inlet pressure that has the same tendency

to influence the cooling load on the evaporator inlet

temperature continues to rise when the cooling load is

increases. The capillary tube length 6 meters has a

expanssion pressure lower than the capillary tube 3

meters.


the refrigerant pressure at the

evaporator inlet in LTC.

Figure 6 below shows the effect of the cooling load

against the mass flow rate refrigerant in the LTC.

Where the mass flow rate of refrigerant adjusts

cooling load that increases with the mass flow rate

refrigerant increase in both LTC using a capillary tube

with a length of 6 meters and a length of 3 meters. In

Figure 7 below shows the relationship between the

evaporator inlet temperature and the evaporator inlet

pressure at LTC.


the refrigerant mass flow rate in LTC.

Figure 7. Relation of evaporator inlet

temperature with evaporator inlet

pressure.


2012

111

In Figure 8 below shows the effect of the cooling load

of the discharge temperature. Where the trend is

continues to increase when the cooling load is

increases.



discharge in LTC.

The relationship between the discharge temperature

and the discharge pressure can be seen in Figure 9

below, the capillary tube between 6 meter and 3 meter

delta temperature difference seen when the cooling

load was increase very different where delta

temperature capillary tube with a length of 6 meters

which is approximately 1oC while capillary length 3

meter has delta temperature is 0.1oC.

Figure 9. Relation of discharge temperature with

discharge pressure.

At high temperature circuit has a tendency similar to

the low circuit temperature which increases the

temperature when the cooling load in the low circuit

temperature increases. Graph influences of the cooling

load at HTC can be seen in figure 10 and 11 below.



cascade condenser in HTC.



discharge in HTC.

4. CONCLUSION

The experimental results using mixture refrigerant of

ethane and carbon dioxide at low temperature circuits

and propane at high temperature circuit and then using

0.054 inch diameter capillary tube and by varying the

cooling load using an electric heater and getting the

cascade refrigeration system characteristics are:

• Pressure, temperature and mass flow rate at low

temperature circuit will increase when the length of

capillary tube is lower and enhanced cooling load.

• By using a capillary tube 0.054 inch and a length of

6 meters at LTC it will produce 1.9 bar expansion

pressure and the evaporator inlet temperature is -

80.6°C, the cabin or storage room temperature is -

75.7oC and mass flow rate is 6.2 x 10

-4 kg/s with

cooling load is 90 W.

ACKNOWLEDGMENTS

This research was supported by Hibah Bersaing 2011

and Hibah Kompetensi Tahun 2012, Direktorat

Jenderal Pendidikan Tinggi, Kementerian Pendidikan

Nasioal, Republik Indonesia.

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Condenser In CO2/NH3 Cascade Refrigeration

Systems, International Jurnal Of Lee. Tzong.

Shing, Liu. Cheng-Hao, Chen. Tung-Wei,

2006, Thermodynamic Analysis Of Optimal

Condensing Temperature Of Cascade-

Condenser In CO2/NH3 Cascade Refrigeration

Systems.C. Zhang, 2005, Generalized

correlation of refrigerant mass flow rate

through adiabatic capillary tubes using

artificial neural network, International Journal

of Refrigeration 28 (2005) 506–514

[16] ASHRAE Handbook, 2006, Refrigeration

System and Applications (SI), American

Society of Heating, Refrigerating, and Air-

Conditioning Engineer, Atlanta, Georgia.

[17] Nasruddin, Edi Hamdi, 2003, Natural

Refrigerants in Indonesia : Challenge and

Opportunity, Presented in ISSM Delft, The

Netherland.

[18] http://www.Wikipedia.org/Biomedical.htm, 12

Februari 2009

[19] Best. Ben, 2007, Physical Parameter of

Cooling in Cryonics,

www.benbest.com/cryonics/cooling.html. 11

Februari 2009.

[20] Cengel. Yunus.A, Boles. Michael,

1998,‖Themodynamics An Enfineering

Approach‖, Third Edition, Mcgraw-Hill,

International Edition

[21] Reinholdt. Lars, Andreasen. Marcin. Blazniak,

2007, Industrial Freezers For Food Utilizing

CO2Part 2: Development And Testing Of A

CO2 Cascade System, Spiral Freezer And Ice-

Cream Freezer, International Congress of

Refrigeration, ICR07-B2-454, Beijing, 2007

[22] Christensen. Kim.G, Bertilsen.P, 2003,

Refrigeration systems in supermarkets with

propane and CO2 – energy consumption and

economy, Proceedings International Congress

Of Refrigeration, ICR0131, Washington, DC ,

USA, 2003

[23] Nasruddin, Donni Redford, 2008, Pengujian

Mesin Refrigerasi Cascade Menggunakan R22-

R404A Dengan Variasi Tekanan Pada Low-

Stage, Departemen Teknik Mesin Fakultas

Teknik Universitas Indonesia, Depok.

http://www.benbest.com/cryonics/cooling.html.%2011%20Februari%202009

http://www.benbest.com/cryonics/cooling.html.%2011%20Februari%202009


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Performance Analysis of Thermoacoustic-Standing Wave As a Power Generator

Surjosatyo A, Sentosa I Mechanical Engineering Department, Universitas Indonesia, Kampus Baru UI, Depok, Indonesia


ABSTRACT

This study is relating to analyze performance of

thermoacoustic-standing wave. Stirling cycle

thermoacoustic engine is developed conventional

stirling engine. This system is more efficient than

ordinary stirling engine because does not use a

moving piston[7]

. The engine uses thermal power to

generate acoustic power. It consists mainly of

three parts: a thermodynamic part consisting of a

stack, two heat exchangers, and a thermal buffer

tube; an acoustic network consisting of an acoustic

compliance and an inertance; and a resonator.

When thermodynamic part heated, it will generate

sounds. The sounds will flow along cylinder tube.

Some aspects can be analiyzed to determine

performance of tharmoacoustic-standing wave.

The effect of temperature difference, stack

geometry, stack position determine performance of

the thermoacoustic-standing wave. Some research

show that acoustic power will increase with

increasing of temperature at hot heat exchanger.

And optimal position and geometry of stack will

generated optimal acoustic power.

Keywords: Thermoacoustic-standing wave, stack,

sound, acoustic power

1. INTRODUCTION

Thermoacoustics is a field that studies the

conversion of heat to acoustic energy. Research on

thermoacoustics can be dated back to the late 19th

century. Modern research of thermoacoustic

systems is largely based on the work of Rott,

Steven Garrett and Greg Swift, in which linear

thermoacoustic models were developed to form a

basic quantitative understanding and numeric

models for calculation. The thermoacoustic engines

contains no moving parts yet the acoustic

stimulation of heat flux and the generation of

acoustic work, point to some type of timed phasing

of thermodynamic process. This phasing in

thermoacoustic engines is due to the presence of

two thermodynamic media fluid and stack plate.

the temperature and pressure oscillations induce

sound waves. The combination of all such process

produces an affluent ‗‗thermoacoustic‘‘ effects.

The applications of thermoacoustics are widely

spread in the gas mixture separation, natural gas

liquefaction, heat pumps, pulse tube,

thermoacoustic regenerator and thermoacoustic as a

electrical generator.

The experiment about performance of

thermoacoustic standing-wave engine has been

studied. Hariharan, P. Sivashanmugam, S.

Kasthurirengan Influence of stack geometry and

resonator length on the performance of

thermoacoustic engine. The results obtained from

the experiments are in good agreement with the

theoretical results from DeltaEc.

2. THERMOACOUSTIC COMPONENT

The open end standing wave thermoacoustic engine

consists of parts such as heat exchangers, stack,

resonator, and working fluid.

2.1. Heat exchangers

The function of heat exchangers in a

thermoacoustic engine is to transfer heat from an

external source to the working fluid in the sealed

resonator chamber and they are used to maintain

the temperature gradient across the stack. The

active heat exchange takes place between the

working fluid and a series of closely spaced,

parallel plates with their surfaces aligned with the

direction of the wave propagation and positioned at

either end of the stack. The heat exchanger should

provide high heat transfer coefficient and low

acoustic power dissipation to the thermoacoustic

side. The hot heat exchanger supplies heat to hot

end of the stack and ambient heat exchanger

extracts heat from other end of stack. The blockage

ratio is considered as same as that of stack so plate

size and spacing used for heat exchanger is

identical to that of stack for the present system.

This allows the gas parcels to move freely from

heat exchanger to stack. With the assumption of

same heat transfer coefficient and temperature

difference between solid plate and the working

fluid, the hot heat exchanger requires more heat

transfer area compared to ambient heat exchanger.

So the length of hot heat exchanger is chosen as

twice the length of ambient heat exchanger. The

optimum PS and length of heat exchanger, which is

equal to the peak to peak displacement of the

working gas is given by the following expression:

y0 = 2la/A

lc = sin(kl)

2.2. Stack

Stack is the heart of standing wave engines, where

the thermoacoustic cycle is generated. It provides

solid heat capacity and large cross sectional area to

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maintain a good thermal contact between gas and

solid stacks. They are finely divided in small

parallel channels with hydraulic radius comparable

to thermal penetration depth. The stack is placed at

a certain location in the resonator, where the

magnitude of local acoustic impedance |Z| is larger

than qga/A, because the magnitude of gas velocity

amplitude is relatively small to reduce the viscous

dissipation of viscous power. Acoustic impedance

is defined as the ratio of complex pressure to

volumetric flow rate. A good stack should

minimize the ordinary heat conduction along

temperature gradient and viscous dissipation of

acoustic power. The minimum thickness of the

stack plate should be 8ds, where ds is solid thermal

penetration depth, which is defined as

𝛿s =

the above equation is used to calculate the plate

thickness used for the present system which states

that the thermoacoustic effects are optimal if the

plate thickness is in the range of 6–8ds. The

thermal penetration depth is defined as the layer

around the stack plate where the thermoacoustic

phenomenon occurs. It is measured perpendicular

to the direction to the motion of gas and it gives

approximately the distance that the heat can diffuse

through the gas. Gas thermal penetration depth is

𝛿s =

by placing heat exchangers at either side of the

stack, heat can be moved so that the temperature

difference across the stack is created. As a result

sound wave can be induced. In order to maximize

the hydrodynamic heat flow, the stack material

with large Ksqscs in comparison to Kgqgcp is

favorable. Successful operation of a standing wave

engine requires an imperfect thermal contact

between the gas and the stack which is obtained

when the spacing between the plates is roughly two

to four times of thermal penetration depth of gas.

Viscous penetration depth is defined as the

thickness of the layer of fluid around the stack plate

that is restrained in its movement under the

influence of viscous forces. Within this layer,

viscous dissipation is responsible for the loss of

kinetic energy, so that the fluid layer of thickness

dv in the vicinity of each stack plate contributes

less to the thermoacoustic effect.

2.3. Resonator

The resonance tube is one of the key components

of a thermoacoustic engine. A smooth, linear

cylindrical resonator pipe without steps,

misalignments and abrupt transitions should be

used to avoid unwanted eddying or non-linear

pressure variations that would greatly complicate

the analysis. Resonance frequencies are mainly

determined by the length of the resonator.

Prolongation of resonance tube may leads to

decrease of working frequency and increase of

stacks hot end temperature with the same heating

power. The velocity amplitude increases from the

heater to the water cooler with a certain length of

the resonance tube, because the heater is closer to

the velocity node. On the other hand, when the

resonance tube is prolonged, the relative location of

the thermoacoustic core shifts nearer to the velocity

node so the velocity amplitude in the

thermoacoustic core decreases [18]. For lowest

dissipation, resonator should provide sufficient

inertance and compliance, thereby maintaining

resonance frequency, while simultaneously

minimizing the acoustic power dissipation. For

thermoacoustic engine the resonance frequency can

be estimated from

|U1|2 - |p1|

2

The above expression gives the acoustic power

dissipation per unit length of the channel, due to

thermal and viscous processes at the channel walls.

To avoid the thermal relaxation losses, tube

material with the combination of smallest possible

Ksqscs and gas with largest possible combination

of Kgqgcp should be selected. Easiest way to

decrease the acoustic power losses is to decrease

the surface area of resonant tube walls. In order to

decrease the acoustic power loss, k/4 wavelength

resonator has been chosen for the present study.

The equation for normalized acoustic power

dissipated in the quarter wavelength resonator is

estimated

ΔĖ2n,r ≈ -

The energy dissipated in the resonator is

proportional to wall surface area of resonator.

2.4. Working fluid

The choice of working fluid especially gas for a

thermoacoustic engine is an important aspect to be

considered as it affects power and efficiency. Gas

properties play an important role in determining the

onset temperature difference. The lightest gases

have highest sound speeds and high thermal

conductivity which will give highest powers due to

high thermal penetration depth, since heavier gases

condense or freeze at low temperatures or exhibit

non-ideal behavior. Gases with high ratios of

specific heats and low Prandtl numbers are well

suitable for thermoacoustic devices. These


2012

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properties can be optimized by the use of a mixture

of light and heavy noble gases. The optimum

mixture of gases in a thermoacoustic engine

depends upon the application and design goals. For

thermoacoustic engines, gases with Prandtl nearer

to one must be used to attain the minimum onset

temperature difference [21]. As the present

experimental setup is an open end thermoacoustic

primemover, air is chosen as working fluid.

3.a. Experimental Set-up

Figure 3 shows the experimental set-up of

thermoacoustic-standing wave. The experiments

are perform the influence of temperature difference

toward acoustic power generated.

Subject Description

Tube Diameter

Tube Lenght

Tube Material

Stack(Regenerator)

Material

Stack Lenght

Amplifier Diameter

Amplifier Lenght

12.5mm

150 mm

Glasses

Steel

35mm

100mm

280 mm

Heater

Sound

Meter

Ambient Heat

Exchanger

Figure 3. Thermoacoustic Devices set-up for

measurment

The sound oscilation will generate due to

temperature differences. This oscilation will be

measured with sound meter and temperature in

both hot side and cold side are measured using

termocouple by national intsrument.


3.a. Acoustic Power

Acoustic power is measured using sound meter.

The data which will converted into acoustic power

is sound intencity (I). The formula is:

β = 10 log

where β is intencity degree(dB), I is sound

intensity(W/m2) and Io is ambient sound

intensity(10-12

W/m2). Further an acoustic power P

(Watt) will obtained by multiplication between an

is an area which through by sound A (m2) and

sound intencity I. Given by equation:

P= I.A

The acoustic power could to ceonverted into

electrical energi using piezoelectricity. ‗Piezo‘ is a

Greek term meaning to apply pressure to, or to

press. Piezoelectric, therefore, refers to the way in

which certain materials can generate a current

when pressure is applied to them. Piezoelectric

charge coefficient values were in the range 1-100

pico coloumb / Newton. Natural Piezoelectric

materials such as: quartz (Quartz, SiO2), berlinite,

tourmaline and salt Rossel. Made of piezoelectric

material are: Barium titanate (BaTiO3), Lead

zirconium titanate (PZT), Lead titanate (PbTiO3)

and so on. The phenomenon of the piezoelectric

effect can be described as follows:

Figure 4. The phenomenon of the piezoelectric

effect (A) before gived pressure or

electrical field. B) giving electrical

field, lenght will increase. (C) giving

reverse field , lenght will decrease.

(D) giving pressure , induksi

polarisation and out tension

happened.

3.b. Figures and Tables Until present, the correlation between acoustic

power obtained and time as shown in figure 5

a)


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b)

Figure 5.a) Intencity degree versus time and b)

Acoustic power output wersus time

Figure 5 shows that intencity degree is a parameter

of sound generated. Intencity degree increases with

after hot side of tube flamed and stable after 75

seconds with about 78 dB. Power output will also

increase after flamed with average of 0.225mW .

The increasing of intencity degree and power

output because of temperature diferences. Within 1

until 75 seconds, temperature differences between

hot side and cold side was unstable. After 75

seconds, the temperature difference was obtained

stable so that the degree of intensity will also be

stable at that time.

Further axperiment was focused influencing of

temperature differences toward acoustic power

obtained. The temperature difference will

influences sound intencity generated. Figure 6

shows the temperature difference in every passing

time. Figure 6 shows also that temperature

differences increases which each passing seconds

and will stable after about 150 seconds. This

research will be adjusted with the thermoacoustic

device to be built.

Figure 6. Temperature hot side and cold side at

stack1]

4. CONCLUSION

A simplified theory gives a satisfactory estimation

for the intencity degrees and power acoustic

obtained. With possible applications of more

optimal stack materials and design, the

thermoacoustic efficiency of the engine can be

further increased. For more accurate determination

of acoustic parameters a pressure measurement

inside the resonator is desirable. Other directions

for the system improvement include a reduction of

the heat leak through the stack holder. The

integration of the engine with compact and efficient

combustors using electroacoustic transformers can

open a possibility to develope power systems pf

thermoacoustic.

5. ACKNOWLEDGEMENT

This experiment was supported by IMHERE

program, Mechanical Engineering of Universitas

Indonesia from 2011 to 2013. The author also

appreciate all students for their help with this

experiment.

5. REFERENCES

1] N.M. Hariharan, P. Sivashanmugam, S.

Kasthurirengan.2011.Influence of stack

geometry and resonator length on the

performance of thermoacoustic engine

2] Matveev, I,K., Najmeddin, S.T, Richards,

C.D.2008. Small Scale Thermoacoustic

Demonstrator. Proceedings of PowerMEMS

2008+ microEMS2008, Sendai, Japan.

3] Swift, GW. 2002. Thermoacoustics: A

Unifying Perspective for Some Engines and

Refrigerators. Sewickley, PA, Acoustical

Society of America

4] Trapp, C. Andrew., Zink, Florian. 2011.

Thermoacoustic heat engine modeling and

design optimization. Applied Thermal

Engineering;2518e2528


2012

117

5] Garrett, SL. 2005. Acoustic laser kit

instruction

6] Backhaus, Scott., Swift, Greg. 2002. New

Verieties Of Thermoacoustic. LA-UR-02-

2721, 9th International Congress on Sound

and Vibration.

7] Gardner, Catherine., Lawn, Chris. 2009.

Design of a Standing-Wave Thermoacoustic.

The sixteenth International Congress on

Sound and Vibration.

8] Garrett, S.L., Backhaus, Scott.. 2000. The

Power of Sound. The Scientific Research

Society.

9] Trapp, Andrew C., Zink, Florian. 2011.

Thermoacoustic heat engine modeling and

design optimization. Applied Thermal

Engineerin: 2518e2528


2012

118

Factors Affecting Performance of Dual Fuel Compression Ignition

Engines

Mohamed Mustafa Ali a, Sabir Mohamed Salih

b

a Mechanical Engineering Dep., Faculty of Engineering Sudan University of Science and

Technology – Khartoum - Sudan

Tel : (00966)509370959. Tel : (00249) 911270089

E-mail : [email protected] b Mechanical Engineering Dep., Faculty of Engineering Sudan University of Science and

Technology – Khartoum - Sudan

Tel : (00249)912133761


ABSTRACT

Compression Ignition Diesel Engine use Diesel as

conventional fuel. This has proven to be the most

economical source of prime mover in medium and

heavy duty loads for both stationary and mobile

applications. Performance enhancements have been

implemented to optimize fuel consumption and

increase thermal efficiency as well as lowering

exhaust emissions on these engines.

Recently dual fueling of Diesel engine has been found

one of the means to achieve these goals. Different

types of fuels are tried to displace some of the diesel

fuel consumption.

This study is made to identify the most favorable

conditions for dual fuel mode of operation using

Diesel as main fuel and Gasoline as a combustion

improver. A single cylinder naturally aspirated air

cooled 0.4 liter direct injection diesel engine is used.

Diesel is injected by the normal fuel injection system,

while Gasoline is carbureted with air using a simple

single jet carburetor mounted at the air intake. The

engine has been operated at constant speed of 3000

rpm and the load was varied.

Different Gasoline to air mixture strengths

investigated, and diesel injection timing is also varied.

The optimum setting of the engine has been defined

which increased the thermal efficiency, reduced the

NOx % and HC%.

Keywords : Dual Fuel Combustion; Thermal

Efficiency; Exhaust Emissions;

Mixture Strength; Injection Timing.

1. INTRODUCTION

Different combustion strategies were used to improve

thermal efficiency and reduce exhaust emissions of

diesel engines. It was started by optimizing the

combustion of diesel fuel alone through turbocharging

and high fuel injection pressures with electronically

controlled fuel quantity and injection timing. But still

this created unwanted exhaust emissions which was

attempted to control using EGR or through exhaust

after treatment including catalytic converters and urea.

That is to reduce NOx, HC, and CO as well as

particulate matter.

Recently the approach has been changed and more

thinking is towards changing combustion strategy by

fuel design, thus changing of fuel characteristics and

control of start of ignition [1]. This will lead to

introduction of another fuel into the engine either by

direct blending of diesel fuel or by partial pre-mixing

of fuel into air induction. In all cases this will require

different ignition delay than that of diesel fuel and

change the combustion pressure and temperature as

well as end gas properties. These approaches are

summarized as follows:

1- Partially Premixed Combustion (PPC): In-cylinder

fuel blending of gasoline with diesel fuel with

parameter sweeps included gasoline-to-diesel fuel

ratio, intake charge mixture temperature, in-cylinder

swirl level, and diesel start-of-injection timing. This

resulted in improved thermal efficiency and reduced

NOx and particulate matter (PM).[2,6 &8].

2- Use of fuel additives in a Reactivity Controlled

Compression Ignition (RCCI) combustion through

addition of the cetane improver di-tert-butyl peroxide

(DTBP) to pump gasoline. Unlike previous

diesel/gasoline dual-fuel operation of RCCI

combustion, this used a single fuel stock (gasoline) as

the basis for both high reactivity and low reactivity

fuels. The strategy consisted of port fuel injection of

gasoline and direct injection of the same gasoline

doped with a small volume percent addition of DTBP.

This resulted in higher thermal efficiency and NOx

and PM within the emission limits [3].

3- Fumigation of diesel engine using alcohol. Here

alcohol is either port injected or just introduced using

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a simple carburetor by modifying the air intake. The

burning of alcohol with injected diesel fuel replaced

diesel and at the same time reduced combustion

temperature which resulted in less NOx and PM

[4,5&7].

Gasoline is also used to fumigate diesel engines to

improve their efficiency and reduce emissions. In this

study gasoline will be introduced in the air intake by

adding a single jet carburetor specially designed to

provide extra lean mixture. Different jet sizes were

used to vary the mixture strength, and the effect on

efficiency and emissions is evaluated.

Another control factor affecting the performance is the

diesel start of injection timing. This is also

investigated in order to define the trends and rules

which will lead to better performance of diesel engines

under dual fuel mode of operation.

2. ENGINE MODIFICATION AND DUAL

FUELING

A single cylinder diesel engine with specifications

mentioned in table 1, has been modified by adding a

carburetor into its air intake. The engine is directly coupled

to AC alternator which generates electricity at constant

rotational speed of 3000rpm.

Table 1. Diesel Engine Specification

Type Single Cyl. Direct Injection

Cooling Air cooled engine

Bore x Stroke 86x72mm

Compression ratio 19:1

Max Output 4.5kW

Speed 3000rpm

The carburetor was designed to suit the engine

displacement and at the same time, the geometry of

the air intake. Different fuel jet sizes have been tried

and the most suitable sizes with regards to emissions

were found, fuel jets of 0.25mm and 0.50mm. Both

were found to give lean mixture to ensure enough

oxygen remain for burning diesel fuel.

Fuels are supplied through two different tanks and

gasoline flow rate has been measured using a rota-

meter while diesel consumption was measured with a

digital weigh machine. A (Kane) 5-gas analyzer was

used to measure exhaust gas analysis (NOx, CO, CO2,

O2), and HC was measured using Horiba gas analyzer.

An electric load has been applied and varied to cover

all engine operating output range. The test rig (Fig.1)

was used to test the engine performance using diesel

fuel alone for baseline performance, and also the

performance when introducing gasoline with different

mixture strengths. The start of injection timing was

also varied to obtain a trend for the tendency of diesel

engine emissions.

Figure 1: Description of test rig setup.

Four tests at constant speed and variable load are

carried out on the engine as follows:

1. Baseline performance using diesel fuel alone.

2. Dual fueling with gasoline using 0.25mm jet.

3. Dual fueling with gasoline using 0.50mm jet.

4. Dual fueling with gasoline using 0.50mm jet

and with retarded start of diesel injection timing.

The constant engine speed is always maintained by the

speed governor which controls the diesel fuel quantity. The engine rotational speed has been set to 3000rpm

throughout the test period.

3. RESULTS

The test has covered all the load range, but for purpose

of comparison for this study, point of maximum load

will be investigated for comparison. Figure 2 shows

variation of brake thermal efficiency with different

operating mode. Using diesel fuel only, efficiency at

maximum load is 24.5% while for 0.25mm and

0.50mm fuel jets, efficiencies are 24% and 27%

respectively. With 0.50mm jet and injection timing

retarded, the efficiency is 26%. Emission gases of

NOx and HC are shown in figures 3 & 4.

Figure 2. Comparison of Brake Thermal

Efficiency for different fuel modes

and injection settings


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D: Diesel fuel alone.

D+G-I: Diesel and gasoline with 0.25mm jet.

D+G-II: Diesel and gasoline with 0.50mm jet.

D+G-III: Diesel and gasoline with 0.50mm jet and

injection timing retarded.

Figure 3, shows comparison of NOx emission for

different modes of operation.

At maximum load NOx was highest with diesel alone

(800ppm), while decreased to lowest level (620ppm)

when operating with 0.25mm gasoline jet. ). 0.50mm

jet gave moderate decrease in NOx with its two

different injection timing settings.

Figure 4. Shows trend for HC emission using

different operating modes.

Significant change has been recorded, not only in the

amount of HC emission, but also the trend of variation

with load as shown in figure 4. That was obtained by

injection timing retardation. HC decreased from

130ppm with diesel to 90ppm using 0.50mm jet and

injection retarded. The HC is decreasing with load

increase. This is a reverse trend to that of diesel

operation.

Figure 5. Variation of CO with different fuel and

injection settings.

As in figure 5, CO% in exhaust gases is highly

reduced throughout the operating loads for gasoline-

diesel mode of fueling using the 0.25mm jet. Level

became as low as 0.01% by volume at maximum load.

0.50mm jet with retarded injection showed also

decrease of CO% but at upper half of load range,

while 0.50mm jet without retardation gave the highest

CO%.

4. DISCUSSION

Results are showing comparison of performance and

emission parameters with varying operating settings.

Super lean gasoline-air mixture using 0.25mm jet gave

best results in terms of CO (0.01%) and

NOx(620ppm) emissions, but at a sacrifice of decrease

in efficiency (4%less). Extra lean mixture using

0.50mm jet gave the best increase in thermal

efficiency (10%more than diesel) and a decrease in

NOx to 720ppm. HC and CO are above diesel ones.

Retarding the injection timing helped to decrease (HC)

throughout the operating load range, with a negative

gradient. HC decreased to only 90ppm at maximum

load. Super lean mixture helped to reduce combustion

temperature and hence reduced NOx and CO, while

increased HC due to quenching effect. Extra lean

mixture have better tendency to ignition by diesel

injection, and therefore resulted in increased thermal

efficiency and lower NOx, but again it quenching

effect produced more HC. The strategy used to retard

the injection timing seem to be favorable for the

combustion of the mixed charge, and resulted in lower

HC levels keeping at the same time higher thermal

efficiency and lower NOx in comparison with diesel

fuel operation.


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5. CONCLUSION

From these results, factors affecting performance of

dual fuel operation of diesel engines can be stated as

follows:

1- Mixture strength of the added gasoline fuel should

be optimized. Super lean mixtures will result in lower

efficiency and higher HC levels, while richer mixtures

can improve efficiency but increase HC

2- To have control over HC with higher mixture

strengths (still lean), the injection timing retardation

showed a positive effect and reversed the curve trend

for HC emission keeping at the same time the gain in

both thermal efficiency and NOx and CO reductions.

6. ACKNOWLEDGMENT The authors would like to acknowledge the support of

Sudan University of Science and Technology (SUST),

department of mechanical engineering for facilitating

this research.

7. REFERENCES

[1] Gautam Kalghatgi, Leif Hildingsson, Bengt

Johansson ―Low NOx and Low Smoke

Operation of a Diesel Engine Using

Gasolinelike Fuels‖ - Journal of Engineering for

Gas Turbines and Power-Transactions of the

ASME 2010 Volume 132 Issue 9 –

[2] Scott Curran, Vitaly Prikhodko, Kukwon Cho,

Charles Sluder, James Parks, Robert Wagner

- Oak Ridge National Laboratory

Sage Kokjohn, Rolf Reitz - Univ of Wisconsin

―In-Cylinder Fuel Blending of

Gasoline/Diesel for Improved Efficiency and

Lowest Possible Emissions on a Multi-

Cylinder Light-Duty Diesel Engine‖SAE

paper number 2010-01-2206

[3] Derek Splitter, Rolf Reitz, Reed Hanson -

Univ of Wisconsin Madison ―High Efficiency,

Low Emissions RCCI Combustion by Use of

a Fuel Additive‖ SAE paper number 2010-

01-2167

[4] C. Sundar Raj*,1, S. Arul2 and S.

Senthilvelan3- University, Chennai‖ Some

Comparative Performance and Emission

Studies on DI Diesel…‖ The Open Fuels &

Energy Science Journal, 2008, 1, 74-78

[5] Kent Ekholm, Maria Karlsson, Per Tunestål,

Rolf Johansson, Bengt Johansson, Petter

Strandh,‖ Ethanol-Diesel Fumigation in a

Multi-Cylinder Engine‖SAE International

Journal of Fuels and Lubricants, 1:1, pp. 26-

36, April 2009.

[6] Leif Hildingsson, Bengt Johansson - Lund

Univ., Gautam T. Kalghatgi ,Andrew J.

Harrison - Shell Global Solutions UK ―Some

Effects of Fuel Autoignition Quality and

Volatility in Premixed Compression Ignition

Engines‖SAE paper 2010-01-0607

[7] M. Abu-Qudais, O. Haddad, M. Qudaisat-

Jordan University of Science and Technology

―The effect of alcohol fumigation on diesel

engine performance and emissions‖ Energy

Conversion & Management 41 (2000)

[8] Vittorio Manente, Bengt Johansson- Lund

University Faculty of Engineering ―Gasoline

Partially Premixed Combustion -…….‖

ISBN: 978-628-8144-3


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Figure.1： Energy Use Composition in

Buildings

For Air Conditioning: 43%

Air Cnonditioning

Machinery

Others

Water

Transport

Others

Air

Transport

Others

Hot

Water

Lighting

Electric

Device

Ventilation

Water Supply

& Drainage

Elevator

By HP of The Energy Conservation Center, Japan

Solar Air-conditioning System Using Single-Double Effect

Combined Absorption Chiller

Hajime Yabase

Kawasaki Thermal Engineering Co.,Ltd., Engineering Office,

Kusatsu, Shiga, Japan

Tel ::+81-77-563-1111, Fax:+81-77-564-4353

E-mail: [email protected]

ABSTRACT

Since further energy saving for global environmental

protection becomes a matter of urgency, promotion of

introduction of renewable energy sources is required

for realization of low-carbon society. We developed a single-double effect combined

absorption chiller for "Solar air-conditioning system"

in 2010. This chiller is composed of a highly-efficient

gas absorption chiller as a main machine which are

equipped with a solar heat recovery unit comprising a

heat recovery heat exchanger and special condenser. It

enables low temp. solar hot water at 75ºC under

operation at the cooling rating of load factor: 100%. And we constructed the demonstration plant in Japan.

We confirmed that the solar heat priority usage

function and gas-based backup function operate

properly and overall system functions normally. In

summer, fuel gas reduction by 10% could be achieved

and the results as estimated were obtained. Keywords : Absorption chiller, Solar heat, Solar air-

conditioning system, solar collector, gas-based

backup function, Demonstration plant

1. INTRODUCTION

Absorption chillers are units to supply chilled water using gas and oil as fuel. In Japan, absorption chillers

have been widely used for industrial and commercial

central air-conditioning because they contribute to

electric-load leveling in summer because of capable of

cooling using little power, and use water having zero

ozone depletion potential (ODP) as refrigerant.

Meanwhile, the global warming issue has worsened

markedly in recent years, which causes us to be

confronted with the urgent task of realization of low-

carbon society. As shown in Figure 1, in case of

Japan, power for air-conditioning accounts for 43% of

total power consumption used for office buildings and

absorption chillers are also strongly required saving-

energy.

Under these situations, a solar cooling system which

performs cooling by introducing hot water obtained

from solar heat into absorption chillers using thermal

energy as driving source has received increasing

attention and undergone promotion of development

toward practical use recently. This is because this

cooling system is capable of using solar heat whose

reserve amount is much abundant and whose energy

conversion efficiency is higher among renewable

energy for air-conditioning application with high

power consumption rate in industrial and commercial

fields.

We has developed a single-double effect combined

absorption chiller exclusively designed for the solar

cooling system and launched in August 2010[1][2],

and we constructed the demonstration plant of this

system in Japan. we report the outline and the

performance of the chillers and demonstration plant.

2. SINGLE-DOUBLE EFFECT COMB-

INED ABSORPTION CHILLERS

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2.1 Use and Problem of Solar Heat

Among renewable energy, solar energy is rich in

reserve, which undergoes promotion of application,

however, it is almost applicable to solar battery

(photovoltaic energy) but solar energy has been not

picked up as the method for using as heat source so

much. However, as shown in Figure 2, photovoltaic

energy generation is as low as approx. 10% in

generating efficiency and in case that solar energy is

picked up as hot water at approx. 90ºC, the energy

conversion efficiency is as high as 40% and the high-

end evacuated tubular type reaches 50%.The system

using solar hot water as driving source is applicable

only to absorption chillers practically. The

conventional air-conditioning system is shown in

Figure 3. However, solar heat is unstable heat source which is

easily influenced by weather and it is difficult to use it

according to fluctuating air-conditioning loads. Solar

thermal air-conditioning system has been tried to be

diffused since 1980‘s, however, they have been

familiarized fully. The reasons are shown as follows:

(1) As shown in Figure 3, in the solar thermal air-

conditioning system, in addition to the absorption

chiller, the backup boiler and accompanying

machines are required, which causes the system

composition to be complicated and the investment

efficiency using renewable energy is not

expected.

(2) It is necessary to control the solar hot water and

backup system according to fluctuations in solar

heat and air-conditioning loads, however, it is

difficult to control and establish an optimal

control system to use solar heat efficiently, and it

is necessary to familiarize local operators to learn

as well.

(3) The double-effect type which is mainly used as an

absorption chiller, which requires heat at 120ºC

or more as driving source. In case that solar heat

at approx. 90ºC is used, only the single-effect

system functions. In case that the backup system

functions, even if fuel is used, the efficiency is

low because of single effect system, the effect of

introduction by renewable energy is not expected

so much.

2.2 Improvement from Conventional System

In consideration with the problems in the above-

mentioned solar thermal air conditioning, we

developed single-double effect combined absorption

chillers for using solar hot water preferentially

in combination with backup heat source such as

gas, oil, etc. (hereinafter referred to as the Solar

Absorption Chillers). The aspect is shown in

Figure 4.

The features are shown as follows:

Figure 2: Solar energy conversion

efficiency

From NEDO homepage

Figure 3 ventional Solar Thermal Air-Conditioning

System

Figure 4: Aspect of Solar Absorption

Chiller


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(1) Since these solar absorption chillers are equipped

with generators driven by solar hot water based on

direct-fired absorption chillers, a backup system

are unnecessary to be prepared, No. of

composition elements are reduced, which

simplifies the air-conditioning system

(2) These solar absorption chillers control so as to use

solar hot water preferentially based on driving by

fuel such as gas, oil, etc. In addition, control of

loading is performed according to fluctuations in

air-conditioning loading.

(3)When driving by fuel, double-effect operation is

performed, the same efficiency as absorption

chillers which are currently diffused is obtained,

which allows saving energy operation because

renewable energy is used.

2.3 Outline of Solar Absorption Chillers

Solar Absorption Chillers are composed of highly-

efficient gas absorption chillers with COP1.3 (gross

calorific value) as main machines which are equipped

with a solar heat recovery unit comprising a heat

recovery heat exchanger and special condenser.

Figure 5: Cycle flow-diagram of Solar Absorption Chiller

Figure 6: Principle of Gene-Links

Figure 7: Principle of Solar Absorption Chillers


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As shown in the cycle-flow diagram in Figure 5, solar

heat hot water is used for heating and regenerating

absorbing solution at the heat recovery heat

exchanger. Using refrigerant generated during this

regeneration process for cooling enables the amount of

fuel used for the high temp. generator to be reduced.

Generally, solar energy collectors have characteristics

in which the smaller the difference in temperature

between the collection temperature and outside air

temperature is, the higher the collection efficiency is,

therefore, it is necessary to allow solar absorption

chillers to use even low temp. hot water to increase the

efficiency of the overall system.

As chillers which are capable of reducing fuel

consumption by introducing hot water, exhaust heat

introduction absorption chillers (Gene-Link) can be

considered, however, Gene-Links are products

designed to use exhaust heat hot water at stably high

temperature (83 to 90ºC) obtained by cogeneration

systems, etc. and cannot use low temp. hot water.

The reason is that Gene-Links are composed as shown

in the principle drawing in Figure 6, in which

refrigerant vapor generated at the heat recovery heat

exchanger and refrigerant vapor generated at the low

temp. regenerator are condensed in the condenser of

the base absorption chillers. Therefore, the saturated

temperature of the heat recovery heat exchanger is

restricted by the saturated temperature of the

condenser of this absorption chiller body, which

prevents the log-mean temperature difference to

collect low temp. hot water from being maintained.

Consequently, as shown in the principle drawing in

Figure 7, in Solar Absorption Chillers, a condenser

exclusive for the heat recovery heat exchanger is

newly provided to separate the heat recovery unit and

the base absorption chiller and a structure to initially

introduce cooling water to the special condenser is

employed, which reduces the pressure in the heato

recvery unit and maintains the log-mean temperature

difference to collect low temp. hot water.

This chillers enable low temp. hot water at 75ºC under

operation at the cooling rating (load factor: 100%,

cooling water temp: 32 ºC) or even lower temp. hot

water depending on loading conditions and cooling

water conditions to be used.

2.4 Performance of Solar Absorption Chillers

The performance of Solar Absorption Chillers is

shown in Figure 8 and 9. Figure 8 shows the heat

recovery amount in each cooling load factor and

Figure 9 shows the combustion gas consumption

amount in this case. The cooling water inlet

temperature is set to 32ºC at 100% load and 27 ºC at

0%, and proportional values at 0 to 100%.

Figure 8 shows that the heat recovery amount in case

of hot water at 75ºC is 0.45kW/RT when the cooling

load factor is 100 %, which increases as the load

factor decreases, and reaches the maximum amount of

1.37kW/RT when the load factor is approx 30%. In

this case, in a loading area with approx. 30% load

factor where the heat recovery reaches the largest

amount, cooling operation only by hot water without

use of combustion gas is possible, therefore, the heat

recovery amount in proportion to load is obtained in a

loading area with approx. 30% and lower load factor.

Figure 8: Heat recovery rate of Solar

Absorption Chillers

chiller-heater

Table 1: Specifications of system

Figure 9: Fuel gas consumption of Solar

Absorption Chillers

chiller-heater


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2.5 Performance of Solar Absorption Chillers

The performance of Solar Absorption Chillers is

shown in Figure 8 and 9. Figure 8 shows the heat

recovery amount in each cooling load factor and

Figure 9 shows the combustion gas consumption

amount in this case. The cooling water inlet

temperature is set to 32ºC at 100% load and 27 ºC at

0%, and proportional values at 0 to 100%.

Figure 8 shows that the heat recovery amount in case

of hot water at 75ºC is 0.45kW/RT when the cooling

load factor is 100 %, which increases as the load

factor decreases, and reaches the maximum amount of

1.37kW/RT when the load factor is approx 30%. In

this case, in a loading area with approx. 30% load

factor where the heat recovery reaches the largest

amount, cooling operation only by hot water without

use of combustion gas is possible, therefore, the heat

recovery amount in proportion to load is obtained in a

loading area with approx. 30% and lower load factor. Further, in case where the hot water is 90ºC, the heat

recovery amount becomes 1.69kW/RT at 100%

cooling load factor and reaches 2.42kW/RT at the

most and the cooling load area operatable only by hot

water is expanded to approx. 57%, therefore, it is

found that heat amount can be more effectively used

even if the temperature of hot water increases.

Figure 9 shows that in case where the hot water is

75ºC, the combustion gas consumption amount can be

reduced by approx. 9% at 100% cooling load factor

compared to the case where hot water is not

introduced. Since the above-mentioned decrease in

combustion gas consumption amount has a

proportional relationship with the heat recovery

amount shown in Figure 8, it becomes larger as the

load factor decreases, as mentioned above,

combustion gas is not required in a area with 30% or

less of load factor.

Further, it is found that in case where the hot water is

90ºC, the combustion gas consumption amount can be

reduced by approx. 32% at 100% cooling load factor.

3. SOLAR COOLING SYSTEM 3.1 Outline of Solar Cooling System

The demonstration plants are installed in our factory

located in Kusatsu City of Shiga Prefecture, Japan.

This system was completed in Dec. 2010 and started

to undergo full-sized verification test in Feb. 2011.

The flow diagram of this system is shown in Figure

10.

Solar heat (hot water at 75ºC to 90ºC) is introduced

into the Solar Absorption Chiller. In addition, if solar

heat is insufficient, the backup system to compensate

for the energy through gas is available.

Evacuated glass tube type solar energy collectors

which are highly efficient in a high-temp area at 75ºC

to 90ºC is used for the solar energy collector. 160

sheets of collectors (260m2) which satisfy the exhaust

heat recovery amount (0.6kW/RT, 126kW *cooling

water at 31℃) during rated operation in case of solar

heat hot water of Solar Absorption Chiller at 75ºC

were installed on the roof of the office.

The hot water storage tank is provided to absorb the

difference of flow rate between the solar energy

collector and Solar Absorption Chiller and serves as a

temporal cushion if solar radiation fluctuates

suddenly.

The radiator is provided to prevent hot water from

boiling by excessive heat collection by operating when

collected solar heat cannot be used on holidays, etc.

3.2 Feature of Solar cooling system

collector

Table 1: Specifications of system

Heatin

g

Coolin

g

Hot water storage

tank (1m3)

Radiator

Heat

exchanger

for heating

Cooling

→

absorption

chiller/heater

pum

p

pump

Gas

inpu

t

↓ Heating

Figure 10: Schematic diagram of system

Figure 11: Aspect of Collector


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3.2.1 Lowering the temperature of solar heat

usable area

Generally, solar energy collectors have characteristics

in which the smaller the difference in temperature

between the collection temperature and outside air

temperature is, the higher the collection efficiency is,

therefore, it is necessary to allow solar absorption

chillers to use even low temp. hot water to increase the

efficiency of the overall system.

As section 2 already described, in the Solar

Absorption Chiller in this system, solar heat at 75ºC

(rated) and approx. 60 ºC (under partial loading) can

be used by employing a hot water heat exchanger

optimized for use of solar heat and improving the flow

of cooling water.

3.2.2 Simplifying and downsizing the solar system

The auxiliaries (Figure 12) such as pumps, etc. to

supply hot water obtained from the solar energy

collector to the Solar Absorption Chiller are required,

however, those items are simplified as much as

possible while considering packaging with the Solar

Absorption Chiller.

Packaging after reflecting these verification results can

reduce the details of work at site and costs for building

up the system as one of targets.

3.2.3 Controlling and Monitoring System

When introducing the solar cooling system, it was

necessary to specially build up the control functions

such as starting/stopping the heat collecting facility

and excessive solar heat collection of the solar energy

collector, however, these control functions are

assembled into the Solar Absorption Chiller in this

system and control of the overall system is enabled.

Assembling the control functions into the Solar

Absorption Chiller maximizes the saving energy effect

based on the use of solar heat by strengthening the

linkage with the control functions of the Solar

Absorption Chiller in addition to eliminating the needs

of the special control equipment.

Some examples of linkage control with the Solar

Absorption Chiller added to this system are shown as

follows:

(1) Backup linkage at the Solar Absorption Chiller

(2) Interlocking control in response to change in the

temperature setting

Figure.13: Monitoring System

Figure 12: Collection of heat facility


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Further, the system is equipped also with the

information collection function to monitor

the system operation status and energy-

saving effect.

Figure 13 shows a part of the monitor system

to display the collection data. The monitor

system enables daily and hourly collection

amount and efficiency to be displayed, which

contributes to the grasping of operation

conditions and review for improvement of

system.

As mentioned above, when building up the

solar cooling system, cost reduction is a big

issue. Assembling the control system and

monitor system into the Solar Absorption

Chiller greatly serves to reduce the cost for

building up the system.

3.3 Evaluation status and results

3.3.1 Operation conditions

Figure 14 shows the operation data on May

20 and Figure 15 shows the operation data on

June 28. From the data, it was confirmed that

the solar heat priority usage function and gas-

based backup function operate properly and

overall system functions normally.

Because of operations with comparatively-

low loads on the conditions where the

maximum temperature was 28.4ºC and the

air-conditioning loading factor was 23% on

May 20, the gas amount could be reduced by

25%. Meanwhile, the maximum temperature

was 34.5 ºC and the air-conditioning loading

factor was as high as 60% on June 28,

however, the gas amount could be reduced by

11%.

Cooling operation starts in late May,

therefore, the monthly reduction rate shows

the data only in June, however, reduction by

10% could be achieved and the results as

estimated were obtained.

3.3.1 Effect of Solar Absorption Chiller

It was confirmed that hot water obtained

from the solar energy collector is constantly used at

75ºC or less and can be used even at approx. 60 ºC

during low-load operations.

In the actual system, the effect could not be quantified

because fluctuation in solar radiation and load should

be considered, however, use of Solar Absorption

Chiller developed exclusively for use of solar heat can

reduce the hot water temperature from the solar energy

collector more than use of conventional exhaust heat

introduction type absorption chiller（Gene-Link）,

therefore, it was confirmed that this system increased

the collection efficiency of the solar energy collector

and improves the efficiency of overall system.

3.3.3 Improvement points

When changing to the low-load operation mode where

the refrigerant pump of the chiller activates the

start/stop control, introduction of hot water is turned

Figure 14: Operation data of cooling (20th

May)

Figure 15: Operation data of cooling (28th July)


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on and off interlinking with start/stop of the refrigerant

pump, however, even in operable load only by solar

heat hot water, it is confirmed that the pick-up

temperature increases due to output delay when

turning on and backup control by combustion

activates. We plan to review the control to minimize

the delay and add it to the system.

4. CONCLUSION

(1) The Combination of the heat recovery exchanger

and special condenser of Solar Absorption

Chiller enables the machine to be operatable even

if the temperature decreases up to 75ºC and does

not require the backup by combustion gas in an

area with 30% or less of load factor

(2) In case where the temperature of hot water is 90

ºC, the reduction rate of combustion gas at 100%

load factor becomes 32%, therefore, the

performance increased in comparison with 26%

of the conventional Gene-Link. In addition, the

cooling load area operatable only by hot water

was expanded to approx. 57%.

(3) From the verification of the demonstration plant

of solar cooling system, the usefulness of the

Solar Absorption Chiller developed exclusively

for use of solar heat was confirmed.

(4) Further, we make sure that our built-up system is

useful to make it easier to introduce a solar

cooling system. By commercializing the system

into which improvement points during

verification were fed back in the future, we aim

to make the solar cooling system to be

recognized as a useful solution tool for global

warming problem and promote them. .

REFERENCES

[1] Hyodo, Y., 2011, Solar Absorption Chillers using

solar heat for cooling of Kawasaki Thermal

Engineering Co., Ltd, Clean Energy, vol.20, no.3,

pp.5-9.

[2] R. Kajii, H. Yabase, M. Ohta, 2011, Development of

Solar Absorption Chillers-Heaters, Trans. Of the

JSRAE, vol.28, no.3,not require the backup by

combustion gas in an area with 30% or less of

load factor


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Environmental Protection and Fuel Consumption Reduction by

Flameless Combustion Technology: A Review

Seyed Ehsan Hosseini, Saber Salehirad, Mazlan Abdul Wahid, Mohsin Mohd Sies,

Abuelnuor Abdeen Ali Abuelnuor

Faculty of Mechanical Engineering, Universiti Teknologi Malaysia, 81310 UTM Skudai,

Johor, Malaysia, January 2012 Tel: +60176830504 Email: [email protected]

ABSTRACT In recent years global fuel consumption has

increased in the world due to modernization and

progress in the standard of living. The

conspicuous rate of carbon dioxide and nitrogen

oxide released to the environment and fuel

resources are depleted day by day due to

inconsiderate fuel consumption. Requirement for

efficient use of any kinds of fuel has become the

other concern due to the oil crisis and limitation

of fuel resources. In combustion process, the

abatement of pollutants often associates with

efficiency loss. In the other word, high efficiency

and low pollutant which are the main

requirements of combustion are not fulfilled by

the existing combustion. During the development

process of new combustion technology, a

particular

focus was on low NOx burners and engines.

Today, flameless combustion has received more

attention because of its low NOx emission and

significant energy saving.

Generally, compatibility between high

performance and low NOx emission has been

observed by preheated air application and

changing the combustion characteristics from

traditional flame to flameless mode. Although, in

flameless mode the oxidizer is diluted and low

concentration of oxygen can be seen, combustion

is still sustained if the air is preheated higher than

the fuel self-ignition temperature. This aims to

review the concepts and the applications of

flameless combustion and gathers useful

information to understand the necessity of

transient from traditional flame mode to

flameless combustion.

4. Introduction

1.1 General concept of flameless oxidation

Heat generation and power production by

combustion of hydrocarbon container fuels are

the main goals of combustion [1, 2]. One of the

best ways to achieve the higher efficiency and

low pollution in combustion is using regenerative

burners with flameless combustion. Recycled

burned gases have been applied in this

technology to make the preheated air lean in

order to achieve low- NOx emissions and a

reasonable thermal efficiency [3]. This

technology, emerged from 1990, and has been

successfully applied, specially, in metallurgy and

steel industries of some developed countries.

Flameless combustion which is known as

Flameless Oxidation (FLOX) in Germany [4],

also known as High Temperature Air

Combustion (HiTAC) In Japan [5], Moderate and

Intensive Low oxygen Dilution (MILD)

combustion in Italy [6, 7] or Colorless

Distributed Combustion (CDC) [8], Low NOx

Emission Injection in the US is a new

combustion system which accomplishes low NOx

emissions and high efficiency among several

techniques. Postponed mixing of fuel and air and

flue gas utilization in the flame zone are the

fundamentals of FLOX [9]. The application of

high temperature air combustion has been

investigated experimentally [10-15] and

numerically [16-22]. Fundamentally, flameless

combustion is identified by aspects of turbulence

and chemistry strongly [23]. The characteristics

of flameless oxidation for various gas type fuels

like methane or ethane [24], also for mixtures of

gaseous hydrocarbons and hydrogen [25-27], and

biogas [28-30] has been analyzed. In addition,

this method has been applied for liquid fuels [31-

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34]. Moreover, solid fuels flameless combustion

has been experimented successfully [35, 36].

1.2. Industrial applications of flameless

combustion

Flameless combustion is suitable for different

industrial procedures that need a uniform high

temperature profile inside the furnace [37]. The main

industrial applications of flameless combustion now

concern the metallurgy area for which the major issue

is energy efficiency. For the other industrial sectors,

the issues are sometimes different, but for such as

glass-making and cement industry [38, 39], waste

treatment [40], petrochemicals, gas turbines [41, 42]

or industrial boilers [43], it is very likely that this new

combustion mode will find its place, in the short or

medium term. The main reasons for development of

this technology in industries can be cited as decreasing

the NOx emissions, increasing the heat transfers, and

rising the duration of the equipment, which are mostly

damaged by very high heat flux.

2. Flameless formation

Preheating of the reaction air and burnt recycled gases

inside the chamber are the basic terms of flameless

combustion. The impacts of recycled flue gases under

highly preheated air conditions (from1200 to 1600 K)

were investigated by Katsuki and Hesegawa [44]

where they found that the flameless combustion

generates by high velocity of reaction air. Dally et al.

[45] stipulated that the flame structure starts to change

when the level of oxygen decreases and it happens at

high Reynolds number for air jet and low oxygen

concentration. In order to understand the transition

from conventional combustion to flameless

combustion, distribution of the axial temperature in

the chamber should be measured regularly during the

experiment versus time [1]. The peaks of temperature

are located in the center of the furnace at the burner

level, far away from the burner position [53]. Recycle

ratio (Kv) of the combustion chamber is the most

important factor which describes the efficiency of a

flameless burner. This ratio can be described as

Kv=Me / (Ma+Mf ) [44]. In the recent equation Me is

the exhaust gases flow rate which is recirculated into

fuel and air before reaction, Mf is the fuel flow rate,

and Ma is the combustion air flow rate. In HiTAC

burners, several combustion modes exist, which are

strongly depend on the average temperature of the

chamber. In addition main factor for determining the

regime of combustion is the volume of recycled flue

gases back into the inlet air jets as shown in Fig.1.In

this figure, temperature is plotted versus Kv for

hydrocarbon fuels. It shows the conversion process of

traditional flame to the flameless combustion and the

conception of the flameless oxidation procedure.

Wünning &Wünning experimentally achieved the

relationship between Kv and temperature of the

furnace for different modes of combustion [4].

Fig. 1 –T- diagram for the conversion of

conventional flame to colorless oxidation mode:

A, traditional regime; B, conversion; C, MILD

combustion; D, no reaction zone.

(i), (ii), (a), (b) lines are showing the boarders of

MILD combustion were achieved experimentally

and represent the path of air heating, increasing

the recycle ratio, path of cooling; and diluting

path respectively.

Zone A. At this zone because of the low jet velocity

(low Kv level) the combustion chamber is working in

conventional flame mode and flame is stable. In the

flameless combustor, the preheating section is

provided to avoid the reaction from quenching. Thus,

the temperature of the chamber is raised to the

amounts greater than fuels‘ self-ignition temperature

by traditional flame (normally up to1000 K).

Zone B. When the temperature inside the combustion

chamber is high enough and more than the auto-

ignition value for the fuel, by enhancing the entrance

momentum of the reactants, the amount of Kv is

increased. Therefore, concentration of the Oxygen in

the air declines and fuel velocity in the reactant side

increases. As a result, the flame pales and the average

temperature of the furnace decreases, this region also

can be called as instability zone.

Zone C. The rate of flue gas is much more than the

incoming air for reaction with the fuel, as a result, the

flame becomes invisible and inaudible and the

reaction zone spreads to the downstream regions of the

combustion chamber. It is possible for much higher


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recycle rates (the high Kv) which causes dilution of

oxygen concentration in reaction [46].These

specifications lead to clean flameless combustion [47].

But, in conditions that Kv is greater than 19 the clean

flameless combustion cannot be obtained, particularly

for the coke oven gas. In the flameless combustion

region, the recycle ratio (Kv) is larger than 2.5, and the

chamber temperature reaches the values up to 1100 K

[1].

Zone D. As it mentioned, in zone C the temperature of

the furnace decreases during flameless mode, but in

zone D the furnace temperature reaches near to a

critical value by intense heat transfer inside the

combustion chamber. Therefore, the flame lifts off,

and eventually, if the temperature is not sufficient,

blows out. As an instance, for methane, the limitation

is less than 1300 K, and for coke oven gas (H2/CH4

60/40 % by vol.) it is around 1180 K [46].

For preventing of the flame front formation, great

values of Kv are required. The jet velocity should be

greater than the velocity of flame propagation. For

example, adding hydrogen gives better range of

stability of flame, and causes an increase in the speed

of the laminar flame propagation about six times with

respect to the methane/air flame. It means when the

fuel is supplied from a single nozzle, larger jet

momentum needs at the constant heat input [48].

Some investigations declared initial jet speed of the

fuel, amount of Oxygen in oxidant air, and the ratio of

density of the fuel inlet to density of the ambient gas

are three basic factors which define flame volume

[49]. In flameless mode reactants are fuel and air

which are highly diluted by an amount of inert flue

gases, and the temperatures within the furnace are

higher than the auto-ignition temperature of the fuel.

In these circumstances conventional flame is not

stable and the flame lifts off due to strong shear

motion caused by gas recirculation (Kv).

Consequently, uniform temperature distribution

appears along the combustion chamber and flux of the

net radiation increases by around 30% [50]. Therefore,

temperature uniformity and the chemical type fields

are the main aspects of the flameless combustion

method. For non-premixed fuel and air jets, there is a

critical rate for Kv which the flameless combustion

does not occur below this amount of recirculation

ratio. Experiments also confirmed this claim [51].

Recirculation of flue gas inside the combustion

chamber makes the reaction oxidizer becomes diluted

and decreases the concentration of oxygen.

Consequently, high efficiency and low thermal NOx

formation are achieved by flameless combustion

method [52]. Fig.2 illustrates the shape of a FLOX

furnace working in traditional flame and flameless

modes.

Fig.2 -Schematic diagram of a FLOX burner

firing in flame and flameless condition.

3. Air preheating process

3.1 Required equipments

In the combustion furnace when air and fuel mixed

together as reactants, it requires some heat to occur

combustion. In order to stabilize flames, combustion

products flow should be recirculated behind a pilot

flame. In these conditions combustion takes place

anywhere in the furnace, therefore, this kind of

combustion is called as highly preheated air

combustion and it is different with conventional

preheated air combustion where there is no self-

ignition.For instance auto-ignition of normal air and

natural gas happens when the temperature of

preheated air goes up around 1100 K. Injection of fuel

and air is done by special nozzle whose configuration

plays crucial role in flameless combustion furnace,

because the burner geometry sets the intensity of

turbulence and exhausting recirculation to reach

flameless combustion. In flameless combustion, air

preheating is the main key to achieve higher

efficiency. As a result, the energy of exhaust gases

which have high temperature is transmitted to the

reaction air in regenerative and recuperative heat

exchangers [4]. Regenerators and recuperators imbibe

excess heat from the product gases and use it again by

increasing the temperature of inlet air. Highly

preheated air increases peak temperature of the flame

in conventional combustion with a great impact on the

formation of NOx , however in flameless oxidation

higher air preheat temperature is desirable. Using


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recuperative and regenerative burners are two main

ways for saving energy. They can be applied by

considering the heat exchanger area and the required

preheating rate. In recuperators and regenerators,

energy transmitted from the exhaust gases back to the

inlet reaction air. Normally, recuperators and

regenerators are heat exchangers and located outside

of the combustion chamber to absorb part of enthalpy

of the hot flue gases for preheating the combustible air

which is called as the secondary air [53, 54].

3.2 Recuperator

In the recuperative burners the air can be preheated

near to 1000 K but the regenerator can increase the

temperature of combustion air up to 1300 K. However,

in traditional combustion burners, NOx formation will

certainly enhance due to rise of the preheating

temperature, although combustion intensity maybe

enhances. S.E Hosseini et al [55] calculated NOx

formation in methane traditional and flameless

combustion computationally. It has been stated that

the rate of NOx formation declines in diluted and

preheated oxidizer conditions in flames combustion,

however the rate of NOx constitution increases in

conventional combustion when applying preheated

oxidizer. Recuperative systems normally are used in

the steel industries because they can be used in

industrial burners for direct heating also for indirect

heating in combination of radiant tubes. The following

burner arrangements could be used for different

applications as shown in fig.3 [4]. Empirical

outcomes on a 300-kW setup showed that using

recuperative flameless combustion burners instead of

conventional burners causes drastically reduction of

NOx production, by 1400 ppm, from 1500 ppm to the

amounts less than 100 ppm [39].

Fig.3 Recuperative burner for use in radiant tubes

3.3 Regenerator

After 1990, high efficient regenerators were applied in

order to increase the temperature of combustion air to

very high values, also large amounts of flue gases

recirculation back to the flame was performed for

creating proper ambient with diluted Oxygen. High

temperature amount of oxidant (between 873 and 1273

K) was achieved by a honeycomb type regenerative

preheater which had become very hot in means of flue

gases from natural gas conventional combustion

before the experiment [56]. Basically, regenerative

burners are using the exhaust gases heat to preheat the

combustion air which would be evacuated to the

ambient in the traditional combustion mode. In the

burners which are working in flameless combustion

systems, air inlets and fuel gas outlets are around the

fuel nozzle. Each one of these inlets or outlets contains

a honeycomb regenerator which is made of ceramic

and during the exhaust cycle imbibes heat from flue

gases and relegates this heat during the firing cycle

into the combustion air. Fig.4 depicts the shape of the

burner and air and fuel inlets and outlets placements.

Fig. 4-Schematic of the HiTAC furnace: (a)

furnace Dimension,(b) Configuration of

fuel and air inlets and burned gas outlets.

These regenerators consist of in 6 pairs and divided

into two groups which are specified by intervals. First

category (three pairs) preheats the combustion air with

a switching time of 10 seconds, and the next group

works as an exhaust-gas storage bed and heat

extractor. The temperature of the inlet air increases to

the values of recirculated flue gas temperature when it

goes through the honeycomb regenerator before


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reaction occurs. Fuel is entered persistently by the

same nozzle, while the air nozzles change from inlets

to outlets and repeat during a defined time space. This

performance causes the single flame to be constituted.

Since all regenerators are located around the fuel-jet

nozzle position of the flame, the temprature is nearly

constant during defined time interval. Around 80% of

the flue gases are recirculated inside of the burner, the

other 20%passes through the chimney which is located

on the wall of the furnace [57]. The honeycomb heat

regenerators highly affect (nearly 88%) the process of

combustion and recovers energy which is existed in

the burnt gases by amounts of about 72%.

Experiments also accept these values and endorse

these facts. Moreover, honeycomb regenerator reaches

a periodic steady state performance very fast in

compare to the other kinds of heat regenerators like a

randomly packed bed of solid storing materials.

Generally, the amount of temperature which is needed

for reaching to flameless combustion circumstances

can be steadily supplied only after two minutes. Fig.5

shows a heat regenerator burner [58, 59].

Fig.5-High temperature air combustion system

In oxyfuels flameless combustion [60], air can be

replaced by pure oxygen and shows very good

performance in the steel industries; also it covers all

the specifications of flameless air combustion. Kumar

et al. [61] demonstrated experimentally that flameless

combustion can be constituted without preheated air

when high rates of Kv (larger than three) is used and

great values of volumetric heat load ( greater than 10

MW/M3). Same outcomes were reported by

Krishnamurthy et al. [62] even when using Oxygen

instead of the air for reaction. Their experiments,

performed by using high amounts of velocities (near to

the speed of the sound) in order to obtain a high rates

of Kv [63].

4. Utilization of different fuels in flameless

combustion

4.1 Natural gas and biogas

A.F. Colorado at el. [30] compared the performance of

a flameless combustion furnace using burner run with

natural gas and biogas at 20KW. Also 20% by volume

excess air during flameless combustion of biogas and

natural gas, 74% and 85% of the total flow rate of flue

gas was conducted to the chamber through

regenerators respectively. To ensure the auto-ignition

temperature of biogas and natural gas is fulfilled,

walls mid temperature sustained up to 870°C. During

the operation with natural gas and biogas the

preheated air was 680°C and 537°C respectively. They

concluded that the system performance was the same

in both experiments. Also, produced CO and NOx

were less than 16 ppm and 3 ppm, respectively. The

efficiency of biogas was 2% lower. They also

mentioned that for the same gases value passed via the

regenerators, the biogas combustion products showed

the high density values of CO2, so specifications of

radiation, capability of imbibitions, and the amount of

heat capacity increased. These aspects lead to higher

rates of heat exchange from the produced flue gases to

honeycomb regenerators. Fig.6 depicts their

experiment in detail.

Fig.6- Experimental setup for natural gas and

biogas flameless combustion

Fuel dilution with N2 or CO2 decreases the NOx

production and causes the flame to extinction.

This proofs that premixing of the fuel flow with

flue gases shows better results on the flameless

combustion regime formation without applying


2012

135

of higher fuel jet momentum [45]. Effuggi at el.

[24] performed some experimental tests using

biogas as an equimolar mixture of CH4/N2.The

conclusions declared that the probability of

applying flameless combustion with low calorific

value fuels without jeopardizing its proficiency

for NOx reduction. The characteristics of biogas

show that biogas is a kind of a diluted fuel itself;

therefore it‘s lower heating value (LHV) declines

when the amount of N2 or CO2 rises. Burner

thermal efficiency decreases due to increase the

volume of inert gas which should be heated up by

the fuel inflammation. Consequently, biogas

composition is a main item in designing of the

burner to determine the value of thermal

efficiency approximately. Flameless oxidation is

impressive in abatement of NOx production

(amounts less than 15 at 3% O2), and

prevents constitution of soot in rich fuels

utilization. Therefore in flameless combustion the

type of fuels and inside temperature of the

chamber do not play conspicuous effects on

production of pollutants. CO emission in rich fuel

condition is inevitable in every types of

combustion [24]. Derudi et al. [25] stipulated the

low calorific value biogas flameless combustion

can be sustained by reaching to high recirculation

values ( more than 5) and chamber temperature

larger than 800°C.

4.2 Solid fuels

Stadler at el. [36] utilized a high velocity of the inlet

oxidizer to make the needed recirculation of the

product gases which are then entrained into the fresh

gases and the coal. Schematic design of the burner is

illustrated in Fig. 7. Air flow transfers the coal into the

furnace by the velocity around 10 m/s. Back recycled

high temperature flue gases makes the combustion air

lean and permeate straightly among the air entrances

to the coal jet due to the configuration of the nozzles,

as indicated in Fig.7 [36].

Fig. 7.Sketch of burner design and flow scheme

5. Environmentally sound characteristics of

flameless combustion

In recent decade, more stringent laws have been

ordained to cope with environmental issues and global

warming. . In combustion process the reaction occurs

between the fuel and the oxidizer to release heat

(thermal energy) as the required factor for electricity

generation. Also, a lot of emissions such as unburned

hydrocarbon (UHC), dioxide carbon (CO2), mono

oxide carbon (CO), nitride oxide (NOx), soot,

particulate matter (PM) are usually released to

atmosphere during combustion process. These

undesirable pollutants can jeopardize the environment

while the rate of their production increases due to

rapid industrialization. NOx (NO2 + NO) is usually

formed in presence of nitrogen and oxygen in very

high temperature conditions. Atmosphere can be

compromised by raising NOx formation in industrial

sectors. Particularly, acid rain, ozone depletion and

smog are the main consequences of more NOx

constitution [64]. Thermal NOx, prompt NOx and N2O

intermediate NOx formation are mentioned as the most

important NOx formation mechanisms. Nitrogen and

oxygen can react inside the combustion furnace in

extremely high temperature according to the reactions

which are called Zeldivich formulation [65]. This so-

call thermal NOx formation can be accelerated

exponentially at temperatures more than 1500oC [66].

Thermal NOx is suppressed in flameless combustion

due to low resident time, low oxygen concentration

and moderate temperature inside the chamber. In

conventional combustion the efficiency of chamber

increases significantly by using preheated air in

combustion process. However, NOx formation

augments drastically. In flameless NOx formation is

kept at very low level due to exhaust gas recirculation

(EGR) application [67]. These exhausted products are


2012

136

conducted into the fresh reactants inside the furnace;

therefore high peak temperature is omitted. As a

result, thermal NOx mechanism is suppressed, and

other insignificant NOx formations methods are

remained. N2O-intermediate NOx formation

mechanism is dominance method for NOx formation

in Hitac due to its moderate temperature and lean fuel

condition. Fig 8 shows a summary of flameless

combustion specifications and NOx formation

mechanisms

Fig8. Flameless combustion specifications

and NOx formation mechanisms

6. Conclusion

Clean combustion and energy efficiency are important

aspects of biofuel and fossil fuel consumption. Since

NOx is an important factor in the constitution of acid

rains and photo chemical smog, control of NOx

production is considered in combustion burners

designing. It has been proven that flameless

combustion systems can be applied vastly in different

industries. Energy savings by the amount of around

30%, also significant reduction of CO230%, reduction

in dimension of the furnace in addition of reduction in

pollutant emissions by 25%in compare of

conventional combustion can be cited as some benefits

of the flameless combustion. The aspects of flameless

combustion introduce it as a new technology to cover

the break among the seemingly paradoxical necessity

of hoarding energy

with respect to the production lower rates of NOx.

Flameless combustion method has been

considered as the optimum combustion technology

because it really saves energy and produces low level

of harmful emissions.

Furthermore, flameless oxidation has more salient

privileges as follow:

1.

2. Concusion

3.

1. The combustion region is extended over the whole

furnace; therefore, thermal gradients can be easily

controlled by applying flameless oxidation which

leads to prevent the formation of hot spots in the

combustion chamber. Therefore the temperature is

uniform in flameless combustion.

2. In combustion of low calorific value (poor quality)

fuels, flameless combustion shows better

performance. Because FLOX technique is able to

reduce emissions and energy consumption

significantly, this method can be considered as one

of the best applicable and useful combustion technologies in the international combustion

community.

3. As a result of reaction air preheating more energy

saving is expected.

4. The geometry of burner is simple; also thermal

efficiency of the burner in flameless operation is

improved. Outer surface temperature of furnace is

more uniform, which shows homogeneous

circumstances compare to the conventional mode,

this uniformity conditions are so useful for many

industrial applications particularly, steel factories.

5. The reaction happens in farther distances from the

face of the burner; therefore, the nozzle of burner

is not damaged. There is no ignition for flameless

combustion, peak of the temperature is low and the

flame is not audible. Moreover, the safety factor of

the colorless combustion system is better than

conventional flame combustion.

6. FLOX can be used for different types of fuels like

gaseous, liquid, and solid fuels in a variety of

conditions, with or without preheating air and fuel,

for stoichiometric, lean, or rich reactions also for

diffusion, partially premixed, or nonpremixed

combustion.

7. The chamber dimensions will be smaller in the

same capacity in compare of conventional

combustion.

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Studies on low-intensity oxy-fuel burner. Proc

Combust Inst, 2009, 32(2): 3139–3146.

[62] Anna Ponzioa,Sivalingam Senthoorselvan,

Weihong Yang, Wlodzmierz Blasiak, Ola Eriksson,

Ignition of single coal particles in high-temperature

oxidizers with various oxygen concentrations , 2008 J

Fuel 87 (6) , pp. 974-987.

[63] Abbas Khoshhal, Masoud Rahimi, Ammar

Abdulaziz Alsairafi, CFD study on influence of fuel

temperature on NOx emission in a HiTAC furnace,

International Communications in Heat and Mass

Transfer 38 (2011) 1421–1427.

[64] Bowman C. Kinetics of pollutant formation

and destruction in combustion. Progress in Energy

and Combustion Science 1975; 1:33-45.

[65] Lavoie G, Heywood J, Keck J. Experimental and theoretical study of nitric oxide formation in internal combustion

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doi:http://dx.doi.org/10.1063/1.4704352.



2012

140

THE EFFECT OF GEOMETRICAL PARAMETERS ON

HEAT TRANSFER OF MICRO-CHANNELS HEAT SINK

Law Wen Zhea, Amer Nordin Darus

b


Universiti Teknologi Malaysia, 81310 Skudai, Malaysia [email protected];

[email protected]

ABSTRACT

A three-dimensional numerical simulation on heat

transfer of trapezoidal microchannel was

conducted. Three sets of hydraulic diameter with 4

types of angles each were simulated using

commercial CFD package, FLUENT. The

numerical model was validated using available

literature. The effects of geometrical parameters

(hydraulic diameter, height, width) were

comprehensively studied. The best Reynolds

number was found by considering the Nusselt

number and pressure difference in channel.

Channel hydraulic size optimization was carried

out using thermal resistance as objective function

and pumping power as constraint. For channels of

the same hydraulic diameter, lower bottom-to-top

width ratio yields higher Nusselt number and lower

pressure difference. In addition, for channels of the

same hydraulic diameter, lower height-to-top width

ratio yields higher heat transfer performance.

Channels with 0.17 bottom-to-top width ratio and

0.41 height-to-top width ratio are found to be the

best in terms of Nusselt number and pressure

difference. By evaluating the Nusselt number and

pressure difference, the best Reynolds ranges from

140-180 for 100micron channel, 140-170 for

200micron channel and 141-167 for 300micron

channel. Finally, by evaluating the thermal

resistance and channel pumping power, the optimal

channel size is 300micron.

Keyword : Microchannel; Trapezoidal;

Optimization; Pressure Drop; Heat

Transfer.

1. INTRODUCTION

With the advent of advanced Integrated Circuits

and MicroElectroMechanical Systems, the

demands on heat removal and temperature control

in modern devices require new techniques for

providing high cooling rates. As shown by

Tuckerman and Pease [1], decreasing liquid

cooling channel dimensions to micro scale will lead

to increase of heat transfer rate.

Pfahler et al. [2] tested rectangular three

microchannels. The results showed that for

relatively

large cross section, friction coefficient increased

with increasing Reynolds number. Peng and

Peterson [3] investigated single-phase convective

heat transfer and flow characteristic of water in

small rectangular

channels. The experimental data revealed that

geometric configuration significantly affect

convective heat transfer as well as flow

characteristic. Qu et al. [4] studied heat transfer

characteristic in trapezoidal silicon microchannels.

The result showed that experimentally determined

Nusselt number is much lower compared to

numerical analysis. The authors attributed this

difference to the effects of surface roughness and

proposed a modified roughness-viscosity model to

interpret the experimental data. Qu and Mudawar

[5] investigated pressure drop and heat transfer

characteristic of a single-phase microchannel heat

sink. It was found that higher Reynolds number

reduces outlet temperature and temperature within

heat sink but at the expense of greater pressure

drop. Li et al. [6] performed three-dimensional

numerical simulations on the laminar convective

heat transfer in microchannel with non-circular

cross-section. It was found that heat transfer

intensity in trapezoidal micronchannel was higher

than compared to triangular microchannel.

Wu and Cheng [7] performed experiments on 13

different trapezoidal silicon microchannels. The

Nusselt number increases almost linearly at low

Reynolds number.

Chen et al. [8] presented a paper on optimum

thermal design of microchannel heat sinks using

the annealing method. The paper found that larger

flow power and smaller substrate thickness

provides lower thermal resistance. Tonomura et al.

[9] conducted a shape optimization of

microchannels using CFD and adjoint method. The

study concluded that the adjoint method can be

used to formulate computationally feasible

procedures for channel shape optimization. Wang

et al. [10] conducted a multi-parameters

optimization for microchannel heat sink using the

inverse problem method. The study concluded that

increased pumping power reduced the overall

thermal resistance of the design.

By referring to literature, most of the optimization

process are complicated and requires extensive

mathematical work. As far as the author is concern,

IMAT-UI 024




2012

141

few or none of the published paper considers

pressure difference and the best Reynolds number

that give minimum thermal resistance in their

optimization process.

2. OBJECTIVES The present work is dedicated to study the effects

of geometrical parameters on heat transfer of

microchannel heat sink and to determine the

optimal cross sectional area of microchannel in

terms of heat transfer performance and pressure

difference.

3. MATHEMATICAL FORMULATION

3.1 Governing Equations

To focus on the effect of the geometrical

parameters on the heat sink performance, the

following assumptions are made:

1) The governing equations based on Navier-

Stokes can be used to describe the physical

processes.

2) The process is steady and fluid is

incompressible.

3) The flow is laminar.

4) The left and right sides and the top of the

channel are assumed to be adiabatic.

5) The thermal properties of solid and water are

constants.

6) The effect of viscous heating is negligible.

7) The effect of buoyancy is negligible.

8) The effect of radiation heat transfer is

negligible.

The continuity, momentum and energy equations

for the current problem can be written as Li et al.

[6]:

Continuity:

Momentum:

Energy:

(a.) (b.)

Figure. 1: (a.) Schematic of microchannel (Li et al.

[6]). (b.) Cross section of modified channel model.

Nomenclature

cross sectional area

heated surface area

constant pressure specific heat of the flow

Dh , hydraulic diameter

DP pressure difference

apparent friction factor

height

effective heat transfer coefficient

k surface roughness

thermal conductivity

channel length

Normalized

, Nu Nusselt number

pressure

pumping power

q heat flux

combined convective and radiative heat

flux

Reynolds number

thermal resistance

temperature

reference temperature

wall temperature

velocity

x,y,z cartesian coordinates

mean fluid velocity

bottom width

top width

Greek Letters

pressure difference

mean temperature difference

thermal conductivity

viscosity

density

angle in degrees

Subscripts

fluid

glass

inlet

maximum

minimum

outlet

solid


2012

142

3.2 Boundary Conditions

From the model used in Li et al.[6], the inlet and

outlet velocity is given as:

(7)

The temperature for inlet and outlet is given as:

Some modifications were made to Li et al. [6]

model in order to simplify it. The thermal boundary

conditions are given as follows. The boundary

conditions of the left and right sides of the

computational domain are adiabatic:

At the bottom position, the heat flux is a given

value whereas the top is adiabatic:

A constant heat flux of is applied

to all the channels bottom used in this study. The

properties of water used are ,

, and

. The properties of silicon

used are , and

. The properties of pyrex glass

used are , and

.


2012

143

Table 1: Dimensions of channels used in the study.

4. METHODOLOGY

In order to investigate the heat transfer capability, a

unit cell of the complete heat sink is considered.

This is to simplify the investigation so as to find

out solely the effects of geometrical parameters on

the heat transfer of microchannel heat sink. The

analysis is confined to the domain where the

coolant enters the channel and leaves the channel.

The channel is created using GAMBIT. The heat

sink itself is not included in the analysis because

the effect of heat sink material and heat sink

conduction is not within the scope of this study.

The numerical computation was carried out using

commercial CFD package FLUENT by solving the

governing conservation equations and the boundary

conditions. The discretization of governing

equations in the fluid and solid regions was done

using finite-volume method (FVM) with second

order upwind method. The flow field was solved

using the SIMPLE algorithm.

Grid independence test is conducted to determine

optimum grid meshing size for the fluid region

inside the channel to minimize computation time

and also memory required for the modeling.

Computational cells with 51440 grids, 77160 grids

and 102880 grids were compared. Results shown

that 51440 grids with 20 edge interval count and 2

face interval size has almost the same result as

77160 grids. Thus, 20 edge interval count and 2

face interval size were used throughout the whole

study.


5.1 Validation

Using optimum grid system, the present numerical

model is validated with available experimental

result by Li et al. [6]. In Fig. 2., there is an

appreciable difference between the numerical

simulation and experimental data. This deviation

may come from the surface roughness effect. The

effect of surface roughness on the flow and heat

transfer in tube becomes apparent as channel size

decreases to the order of microns. In the case of

numerical simulation, FLUENT does not have the

capability to account for roughness effect.

Figure 2: Nusselt number validation

Channel

Wt

(micron)

Wb

(micron) H (micron) Wb/Wt H/Wt k/Dh L/Dh

Surface

material

100micron

45degrees 241.6 41.6 100 0.17218543 0.413907 0 300 silicon

100micron

55degrees 192.2 52.158 100 0.271373569 0.520291 0 300 silicon

100micron

65degrees 157 63.738 100 0.405974522 0.636943 0 300 silicon

100micron

75degrees 130.4 76.81 100 0.589033742 0.766871 0 300 silicon

200micron

45degrees 483 83 200 0.17184265 0.414079 0 150 silicon

200micron

55degrees 385 104.916 200 0.272509091 0.519481 0 150 silicon

200micron

65degrees 314 127.476 200 0.405974522 0.636943 0 150 silicon

200micron

75degrees 260.8 153.62 200 0.589033742 0.766871 0 150 silicon

300micron

45degrees 725 125 300 0.172413793 0.413793 0 100 silicon

300micron

55degrees 577 156.86 300 0.271854419 0.519931 0 100 silicon

300micron

65degrees 471 191.214 300 0.405974522 0.636943 0 100 silicon

300micron

75degrees 391 230.23 300 0.588823529 0.767263 0 100 silicon


2012

144

Table 2: Dimension of channels used in validation

Data

types

Wt

(micr

on)

Wb

(micr

on)

H

(micr

on)

Wb/

Wt H/Wt L/Dh

Surfa

ce

mater

ial

Heat

Flux

Exper

iment

al Li

et al. 770.5 672.6 56.34

0.872

94

0.073

12

299.0

2

silico

n

1.00E+0

6

Prese

nt

work

(2012

) 770.5 672.6 56.34

0.872

94

0.073

12

294.1

17

silico

n

1.00E+0

6

5.2 Nusselt Number

A comparison with literature was made using

almost the same channel dimension. From Fig. 3.,

the present numerical model underestimates at

Reynolds number lower than 240 and

overestimates at Reynolds number higher than 240.

This may due to the constant viscosity model of the

current study. The equation for Nusselt number

taken from (Mahdi, [11]):

Viscosity, μ, has some appreciable effect on the

Nu. But, as an overall, constant properties model

can be used as the trend of Nusselt number against

Reynolds number correlates well with literature.

Gunnasegaran et al. [12] employed constant

property model in their numerical studies to study

the effects of geometrical parameters as well. From

Fig. 3., it is shown that at higher Reynolds number,

the Nusselt number tends to approach a constant.

This correlates well with literature as Mala and Li

[13] show that transition flow regime started at

Re=650.

Figure 3: Nusselt number comparison.

Table 3: Dimension of channels used in

comparison

Data

types

Wt

(micr

on)

Wb

(micr

on)

H

(micr

on)

W

b/

Wt

H/

Wt

L/

Dh

Surfa

ce

mate

rial

Heat

Flux

Experi

mental

Li et al. 770.5 672.6 56.34

0.8

72

0.0

73

29

9.0

silico

n

1.00E

+06

Experi

mental

Wu et

al.

770.4

8

672.6

3 56.34

0.8

73

0.0

73

29

8.6

silico

n

1.00E

+06

Numeri

cal Li et

al. 770.5 672.6 56.34

0.8

72

0.0

73

29

9.0

silico

n

1.00E

+06

Present

work

(2012) 770.5 672.6 56.34

0.8

72

0.0

73

29

4.1

silico

n

1.00E

+06

In order to find out the effects of geometrical

parameters, three sets of channel sizes were used,

that is: 100micron, 200micron and 300micron

hydraulic diameter. Next, with each hydraulic

diameter, 4 types of channel each of different

angles were used, that is: 45degrees, 55degrees,

65degrees and 75degrees. In order to easily

compare with data available in literature, the values

of Wt/Wb, H/Wt, k/Dh and L/Dh were given for

easy reference. These values are important

parameters with which the effects of geometrical

parameters are to be investigated. Surface heat

transfer coefficient, , is given by the following

equation:

Surface Nusselt number is calculated using the

following equation:

From Fig. 4., Fig. 5. and Fig. 6., it is shown that

channels with 45degrees exhibit the highest Nusselt

number for the same given hydraulic diameter. It is

clear that as and increased, Nusselt number

decreased for channels of the same sizes. This is

consistent with the study made by Wong [14] who

found out that reduced distance between side walls

in channel will increase the velocity gradient at the

wall boundaries. This in turn lowers thermal

resistance and increases Nusselt number. In the

case of trapezoidal channel, the smaller heated

surface at channel bottom meant smaller distance

between channel walls. Velocity gradient between

channel wall boundaries at channel bottom

increased due to the smaller distance. This

phenomenon likewise caused the increment of

Nusselt number for 45degrees channel, followed by

55degrees channel, 65degrees channel and finally

75degrees channel. When the value increased,

channel top area decreased in proportion to channel


2012

145

height, H. The decrement of channel top area may

cause the fluid convection between channel top and

channel bottom to decrease due to fact that fluid at

the upper volume remains unheated or not fully

heated compared to the fluid at the channel bottom.

Larger top area means higher volume of unheated

fluid can be carried and effective fluid convection

can occur.

As channel height, H, increased, the Nusselt

number increased as well. Larger channel height

means higher mass flow rate and heat removing

capacity. The present result correlates well with

literature as shown by Cheong [15] who found out

that for channels of same hydraulic diameter, the

size of heated surface does not necessarily increase

heat transfer capabilities of microchannels. The

study showed that channel height also affects heat

transfer performance of a channel.

The result from present study is compared with

available literature. In the experiment conducted by

Wu et al. [7], it was found that as decreased,

Nusselt number decreased.

Figure 4: Nu against Re for 100miron



5.3 Pressure Difference

The pressure difference of present numerical study

was compared with experimental results from

literature. From Fig. 7., it was shown that

appreciable difference occurred between numerical

model and experimental data. This is due to the

different dimension of channels used in both cases.

The trend of the pressure difference correlates well

with literature as shown in Fig.7.

Pressure difference is one of the major issues in

optimization of channel. Higher pressure difference

increases pumping power requirements. The

equation for pressure difference taken from Wu et

al. [7] is given as:

Figure 7: Pressure difference comparison


2012

146

Table 4: Dimension of channel used in PD

comparison

Data types

Wt

(micron)

Wb

(micron)

H

(micron) Dh Wb/Wt H/Wt L/Dh

Surface

material

Heat

Flux

Experimental

Qu et al. 237.01 66.11 109.77 115 0.27893 0.463 262 silicon zero

Present study

(2012) 770.5 672.6 56.34 102 0.87294 0.073 294 silicon 1.00E+06

Fig. 8., Fig. 9. and Fig. 10. showed that pressure

difference is the lowest for 45degrees channel,

followed by 55degrees, 65degrees and finally

75degrees. This may due to the bigger channel

bottom surface area as the trapezoidal angles

changes from 45dgrees to 75degrees. The pressure

difference decreases as hydraulic diameter

increases. Since the same channel depth is being

tested for the same hydraulic diameter, ie:

100miron height for 100micron hydraulic diameter

and so forth, the results also show that pressure

difference is inversely proportional to channel

depth. The result correlates well with literature as

Harms et al. [16] also indicated that pressure drop

is inversely proportional to depth of channel.

Upon close inspection, the magnitudes of the

velocity at channel bottom for both channels are

found to be relatively the same. This means that,

although the velocity contour is different when the

fluid is developing, it does not affect the velocity at

the channel bottom for both channels whilst

developing and after fully developed. Therefore,

the effect of velocity can be ruled out in explaining

the difference of pressure drop between the

45degrees channel and 75degrees channel. The

author proposed that the difference in pressure drop

may be attributed to the difference in surface area

of channel bottom.

Figure 8: PD against Re for 100micron




2012

147

5.4 Reynolds Number

Analysis was done to find out the best Reynolds in

which channel exhibits the most optimal heat

transfer performance. Both normalized Nusselt

number and normalized pressure difference were

calculated in order to find the best range. Equation

of normalized Nusselt number and normalized

pressure difference are as follows:

The best Reynolds number can be found by finding

the intersecting point between the normalized

Nusselt number and the normalized pressure

difference. This intersecting point corresponds to

the highest Nusselt number and lowest pressure

difference possible. From Fig. 11., Fig. 12. and Fig.

13., it is shown that the best Reynolds ranges from

140-180 for 100micron, 140-170 for 200micron

and 141-167 for 300micron. These ranges of

Reynolds were then used as a reference in the

calculations of channel size optimization in the

next section.

Figure 11: Normalized Nu against Normalized

PD 100micron


PD 200micron


PD 300micron

5.5 Channel Size Optimization

or pumping power is the power required to

drive the flow inside a microchannel. Pumping

power is equal to the product of volumetric flow

rate and pressure drop as shown by the equation:

Thermal resistance is given by the following

equation:

Optimization of channel sized was done by first

calculating the normalized channel thermal

resistance and normalized pumping power. Then

these two parameters were plotted against channel

diameter in order to find the optimal size for heat

transfer. The calculation for normalized channel

thermal resistance and normalized pumping power

were given as follows:

Lower pumping power and thermal resistance is

desirable in order to optimize the channel size. Fig.

14. showed that the optimal size is 300micron. This

may due to the fact that for larger hydraulic

diameter, pressure difference between inlet and

outlet is lower. The thermal resistance for larger

hydraulic diameter is also lower. Hence, 300micron

channel exhibits the best heat transfer. The result of

this channel optimization is in agreement with Ong

[17] and Chan [18]. Both Ong [17] and Chan [18]


2012

148

also found out that higher hydraulic diameter

performs better in heat transfer capability.

Figure 14: Normalized Thermal Resistance

against Normalized Pumping

Power

6. CONCLUSION

Numerical simulation on the fluid flow and heat

transfer characteristics in trapezoidal microchannel

was conducted in this study. The effects of

geometrical parameters (hydraulic diameter, height,

width) were extensively studied. Based on the

results, the following conclusions can be made:

1) The Nusselt number in the fully developed

region increases as the Reynolds number

increases. This is in disagreement with

conventional theory of flow in ducts where the

fully developed Nusselt number is a constant.

2) The bottom-to-top width ratio, the height-to-

top width ratio and the hydraulic diameter of

trapezoidal channel were found to have great

effect on the laminar Nusselt number.

3) For channels of the same hydraulic diameter,

lower bottom-to-top width ratio yields higher

Nusselt number. This is in disagreement with

the study made by Wu et al. [7].


lower bottom-to-top width yields lower

pressure difference.


lower height-to-top width ratio yields higher

heat transfer performance.

6) For each of the hydraulic diameters, channels

with 0.17 bottom-to-top width ratio and 0.41

height-to-top width ratio are found to be the

best in terms of Nusselt number and pressure

difference.

7) By evaluating the Nusselt number and pressure

difference, the best Reynolds ranges from 140-

180 for 100micron channel, 140-170 for

200micron channel and 141-167 for

300micron channel.

8) By evaluating the thermal resistance and

channel pumping power, the optimal channel

size is 300micron.

References

[1] Tuckerman, D.B. and Pease, R.F. (1981).

High Performance Heat Sinking for VLSI.

IEEE Electron Device Letters. Vol. EDL-2,

No.5

[2] Pfalher, J., Harley, J., Bau, H.H., and Zemel,

J.N. (1990). Liquid transport in micron and

submicron channels. Sensors Actuators. A21–

A23 431–434.

[3] Peng, X.F. and Peterson, G.P. (1996).

Convective heat transfer and flow friction for

water flow in microchannel structures. Int. J.

Heat Mass Transfer. 39 (12) 2599-2608.

[4] Qu, W., Mala, G.M., and Li, D. (2000).

Pressure-driven water flows in trapezoidal

silicon microchannels. Internat. J. Heat Mass

Transfer. 43 353– 364.

[5] Qu, W. and Mudawar, I. (2002). Experimental

and numerical study of pressure drop and heat

transfer in a single-phase micro-channel heat

sink. Internat. J. Heat Mass Transfer. 45

2549–2565.

[6] Li, Z., Tao, W.Q., and He, Y.L. (2006). A

numerical study of laminar convective heat

transfer in microchannel with non-circular

cross-section. International Journal of

Thermal Sciences. 45 1140–1148.

[7] Wu, H.Y. and Cheng, P. (2003). An

experimental study of convective heat transfer

in silicon microchannels with different surface

conditions. Internat. J. Heat Mass Transfer.

46 2547–2556.

[8] Chen, C.W., Lee, J.J., and Kou, H.S. (2008).

Optimum thermal design of microchannel

heat sinks by the simulated annealing method.

International Communications in Heat and

Mass Transfer. 35 980–984.

[9] Tonomura, O., Kano, M., and Hasebe, S.

(2010). Shape Optimization of Microchannels

Using CFD and Adjoint Method. 20th

European Symposium on Computer Aided

Process Engineering – ESCAPE20. 37-42.

[10] Wang, Z.H., Wang, X.D., Yan, W.M., Duan,

Y.Y., Lee, D.J., and Xu, J.L. (2011). Multi-

parameters optimization for microchannel

heat sink using inverse problem method.


Transfer. 54 2811–2819.

[11] Mahdi Zhaleh Rafati (2010). Numerical

simulation of fluid flow and heat transfer in a

trapezoidal microchannel. Master of

Engineering, Universiti Teknologi Malaysia,

Skudai.


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[12] Gunnasegaran, P., Mohammed, H.A., Shuaib,

N.H. and Saidur, R. (2010). The effect of

geometrical parameters on heat transfer

characteristics of microchannels heat sink

with different shapes. International

Communications in Heat and Mass Transfer.

37 1078–1086.

[13] Mala, G.M. and Li, D. (1999). Flow

characteristics of water in microtubes.

Internat. J. Heat Fluid Flow. 20 142–148.

[14] Wong Wai Hing (2005). Numerical

simulation of a microchannel. Bachelor of


Skudai.

[15] Cheong Tuck Meng (2006). A micro-channel

heat transfer with a heat source. Bachelor of


Skudai.

[16] Harms, T.M., Kazmierczak, M.J., Gerner,

F.M., Holke, A., Henderson, H.T.,

Pilchowski, J., and Baker, K. (1997).

Experimental investigation of heat transfer

and pressure drop through deep

microchannels in a 110 silicon substrate.

Proceedings of ASME Heat Transfer Division,

in: ASME HTD. vol. 351-1, pp. 347–357.

[17] Ong Jiun Shyong (2012). The effect of Prandtl

number to the performance of microchannel

heat sink. Bachelor of Engineering Thesis,

Universiti Teknologi Malaysia, Skudai.

[18] C

han Zaolon (2012). Second law analysis of

microchannel. Bachelor of Engineering

Thesis, Universiti Teknologi Malaysia, Skuda


2012

150

Investigation of the Velocity Profiles in a Ninety-Degree

Curved Standing Wave Resonator with PIV

Normah M. G.a, Irfan Abd. R.

b, Quenet T.

c, and Zaki Ab.M.

b

aFaculty of Mechanical Engineering

University Teknologi Malaysia (UTM) Skudai, Johor

Tel : (021) 7270011 ext 51. Fax : (021) 7270077

E-mail : [email protected] bSchool of Mechanical Engineering

University Malaysia Perlis (UniMAP), Kangar, Perlis

Tel : (04) 9885035 Fax : (04) 9885034

E-mail : [email protected] cUniversite de Rennes 1, IUT de Saint Malo, France

[email protected]

ABSTRACT

Travelling wave thermoacoustic heat engines have

been reported to have a higher efficiency than the

standing wave ones. The former are generally large

systems which consist of toroidal shape resonators.

While standing wave heat engines are inherently

smaller, a reduction in size could be considered

which may involve curvatures as compared to the

straight tube conventional systems. However, as with

the streaming losses in the travelling wave

resonators, losses due to the curvature may be

generated. This study involves preliminary

experimental measurements using the Particle Image

Velocimetry (PIV) method to analyze the velocity

profiles in a standing wave resonator before and after

a ninety degree curvature. This design can reduce the

space generally occupied by the straight standing

wave resonator. The overall length of the resonator

fits a quarter wavelength wave based on the straight

closed-end tube type. The working gas is air at 1

atmospheric pressure. Results have shown that the

velocity profiles after the stack but before the

curvature exhibit clear straight paths up just as

reported elsewhere. Signs of disordered motion could

be observed just before the bend and the pattern

continues until after the curvature. The results are

obtained before one periodic cycle and before the

acoustic wave front hit the tube end. The trend is

expected to affect the overall thermoacoustic

performance of the engine as returning gas particles

interact with the oncoming particles that pass by the

curvature.

Keywords : Standing wave resonator,

thermoacoustic heat engines,

Particle Image Velocimetry,

Curvature

1. INTRODUCTION

Thermoacoustic heat engines and refrigerators are

devices that utilize acoustic waves to drive a

thermodynamic process. A thermoacoustic

refrigerator generates cooling from solid-fluid

interactions. Acoustic waves establish a temperature

gradient, transferring heat from one to another, as

fluid particles oscillate over solid boundaries.

Connected to the ambient heat exchanger and a

cooling load, a thermoacoustic refrigerator poses an

attractive alternative to the conventional system.

However, thermoacoustic systems have yet to hit the

commercial market due to the high cost associated

and the selected research community that fully

understand the concept. Studies are being done to

explore possibilities of potential practical applications

as well as towards the better understanding of the

related theories.

There are two types of thermoacoustic systems; the

travelling wave and the standing wave. The former

generally involves a large curved tube with a

regenerator placed inside for the solid-fluid

thermoacoustic effects to take place whilst the

simplest standing wave system consists of a long

straight resonator with a stack located somewhere

between the maximum pressure and displacement of

the oscillating fluid particles. Inspired by the

curvatures associated with the generally large

travelling wave systems, this study looked into the

possibilities of reducing the long tube of a standing

wave resonator by having it bent at a ninety degree

angle. Flows around bend may initiate disorderly flow

at certain velocities and this is being investigated

here. Unlike the common flow around curvatures,

however, the present work involves acoustic waves

that have passed through a thermoacoustic stack

before proceeding by a ninety-degree bend.

IMAT-UI 025




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The solid-fluid interactions that generate and

maintain the thermoacoustic effects are of importance

and the heart of thermoacoustic systems.

Experimental research into the thermoacoustic

phenomena include optical methods that are used to

measure the flow activities within such systems

because they are non-intrusive. Visualization

techniques of flow field that have been used in

pulsating/oscillating flow experiments include

Holographic Interferometry [1, 2], Laser Doppler

Anemometry (LDA) [3] and Particle Image

Velocimetry (PIV) [4-7]. The latter gives velocity

data over a large area which is favorable in

thermoacoustic experiments. Fig 1 shows a schematic

of a standing wave thermoacoustic refrigerator

(without the cold and ambient heat exchangers)

consisting of a closed-end long tube with an acoustic

driver on one end.

Figure 1: A schematic of a basic thermoacoustic

refrigerator

2. Experimental Set-Up

The resonator is made from acrylic for transparency,

a necessity in the PIV experiment. It has a square

cross-section of 80x80 mm2 with a center-line length

of 86 meters which follows that of Blanc-Benon et al

[5]. The length was based on a previously completed

design with the same PIV apparatus [8] which was

configured on a half-wavelength straight resonator.

Standing wave resonators are made to fit a quarter or

half wavelength wave within the tube, the shorter one

having lesser losses. However, the present study

followed that of previous reported thermoacoustic

experiments with PIV which utilized a half-

wavelength tube. Fig 2 shows the resonator

fabricated with a ninety-degree curvature.

Figure 2: The ninety-degree acrylic resonator

The experimental set-up is shown in Figure 3. The

resonator was filled with air at atmospheric pressure

with the thermoacoustic stack made of glass plates.

The stack consisted of parallel non-conducting plates

of thickness 2mm and separated by a distance of 1

mm. Each plate is 25 mm long and 79 mm wide.

The stack was positioned 21.5 cm away from the

driver where the amplitude for both pressure and

velocity is high enough for the thermoascoustic

effects to take place within the stack.

The resonator was seeded with smoke particles

located near the acoustic driver. An Nd-YAG laser

and a CCD camera were used to detect and capture

the particles motion. The driver generated acoustics at

200 Hz, which is the theoretical resonant frequency, f,

of this particular resonator based on the simple

relationship,

nL

RTf

(1)

where , R, T, and L are the ratio of specific heats,

gas constant, operating tempearture and resonator

length respectively. Since the tube was designed for a

half-wavelength resonator, n is equal to 2 in this case.

A function generator produced sinusoidal waves

which was amplified by a power amplifier before

passing through a loud-speaker which acted as the

acoustic driver.

Figure 3: The experimental set-up for the PIV


2012

152

3. Results and Discussion The PIV measurements are taken 5 seconds after the

acoustic driver was turned on. Twenty sets of data

were collected each for before and after the curvature

0.5 seconds apart. Since the domain covered by the

PIV is comparatively large, selected regions are

shown here to draw attention to those affected by the

curvature. Fig 4 shows the velocity profiles before

the curvature at locations a, b, c, and d, as described

in Fig 3.

Since the images were captured closed to the ninety-

degree bend, the curved wall effects are clearly seen

here, particularly on the right hand side of (a) and (d),

the top and bottom of the resonator, closer to the

walls. Previous work with PIV on profiles near the

stack showed that profiles are laminar away from the

wall effects with vortices forming close to the stack

walls and significant ones behind think plates, soon

after start-up as well as after steady-state is established [4, 7]. Fig 5 shows the velocity profiles

after the ninety-degree bend.

Figure 4: Velocity profiles after the stack before the curvature at positions (a); (b); (c); (d)

Figure 5: Velocity profiles after the stack after the curvature at positions (a); (b); (c); (d)


2012

153

This time, the curvature effects are observed on the

left hand side of the figures, in all the figures, (a)

through (d). The trend after the bend also exhibits a

tendency towards a disorderly flow as seen

throughout Fig 5. Based on the quarter-wavelength

standing wave resonator calculations, the first wave

front has not hit the tube end yet. This non-linear

behavior is expected to intensify as returning

particles meet with oncoming ones over several

periodic cycles. The consequences are undesirable

since there will be energy lost within the already low

performing standing wave thermoacoustic system.

Thus, a ninety-degree curvature is probably not a

favorable possibility in a standing wave resonator.

7. Conclusion

A preliminary investigation with PIV on a ninety-

degree curved half-wavelength standing wave

thermoacoustic resonator was completed. Although

images were obtained before one periodic cycle was

completed or even before the first wave front hit the

tube end, results showed that a ninety-degree bend is

not a favorable design. This is because the velocity

profiles obtained indicate that the particles did not

maintain its straight paths, moving towards disorderly

behavior as particles approached the bend, worsening

after the bend. The situation is expected to aggravate

as this is a standing wave resonator.

ACKNOWLEDGMENT

The authors wish to thank Ministry of Education

FRGS-KETTHA(9003-00352) for the research grant,

Universiti Teknologi Malaysia (UTM) and Universiti

Malaysia Perlis (UniMAP) for the facilities to do the

research. The authors also appreciate the help from

Mr Johari, the technician at the PIV laboratory who

has assisted in the series of experiments.

REFERENCES

[1] Majid N, and Kamran S. A Critical review on

Advanced Velocity Measurement Technique in

Pulsating Flows. Meas. Sci. Tecnology

2010(21);042002-19pp.

[2] Herman C, Kang E, Wetzel M. Expanding the

Applications of Holographic Interferometry to

the Quantitative Visualization of

Oscillatory Thermofluid Processes Using

Temperature as Tracer. Experimental Fluid1998

(24); 431-446

[3] Bailet H, Lotton P, Bruneau M, Gusev V,

Valiere JC, and Gazembel B. Acoustic

PowerFlow Measurement in Thermoacoustic

Resonator by Means of Lasser Doppler

Anemometry (L.D.A) and Microphonic

Measurement. Appl. Accoustic 2000, 60(1);1-

11

[4] Blance-Benon P, Besnoin E, and Knio O.

Experimental and Computational Visualization

of the Flow Field in a Thermoacoustic Stack. C.

R. Mecanique 2003. 331;17-24

[5] Argenthael B, Marc M, and Blanc-Benon P.

Measurement of Acoustic velocity in the Stack

of a Thermoacoustic refrigerator using

[6] Particle Image Velocimetry. J Heat Mass

Transfer 2007. DOI 10.1007/s00231-007-0316-

x

[7] Shi L, Yu Z, Jaworski AJ, and Abduljalil AS.

Vortex Shedding at the End of Parallel-Plate

Thermoacoustic Stack in the Oscillatory Flow

Condition. World Academy of Science,

Engineering and Technology 2005; 49

[8] Mao X, Marx D, and Jaworski AJ. PIV

Measurement of Coherent Structures and

[9] Turbulence Created by an Oscillating Flow at

the End of Thermoacoustic Stack. Proceeding of

the ITI Conference in Turbulence 2005. 25-28

September. Bad-Zwishenahn, German.

[10] Irwan SA and Mohd-Ghazali N. Stack geometry

Effects on Flow Pattern with Particle Image

Velocimetry (PIV). Jurnal Mekanikal 2011; 33


2012

154

MED+AD Desalination Cycle

Muhammad Wakil Shahzad a, Kim Choon Ng

a,b* , Wai Soong Loh

a, Won Gee Chun

c

a Department of Mechanical Engineering,

National University of Singapore,

9 Engineering Drive 1, Singapore 117576, Singapore b Visiting professor(sabbatical), Water Desalination and Reuse Centre,

King Abdullah University of Science & Technology, Thuwal, 23955-6900, Saudi Arabia. c Department of Nuclear and Energy Engineering,

Cheju National University, 66 Jejudaehakno, Jejusi, South Korea

*[email protected]

ABSTRACT

The MED+AD cycle is a hybrid MED+AD. It has

the potential to increase the water production rate of

traditional MED plant by extracting vapor from the

last MED effect and injecting the desorbed vapor into

first stage of MED. It lowers the operational

temperature of MED+AD as compared to traditional

MED. The lower temperature operation typically

ranges from 5-50oC which reduces the corrosion and

fouling chances as well as scavenging the energy

from ambient in last stages of MED. Simulation

results are presented for MED+AD using FORTRAN

linked with IMSL. It can be seen from results that the

top brine temperature can be as low as 50C and the

concentration can be as high as 120,000ppm.

Keywords : Desalination, MEDAD, MED with

thermal compressor, Low

temperature MED.

1. INTRODUCTION

The fresh water demand is increasing with population

increase. The total water on earth covers almost

three-fourth of its surface. However, the seawater is

almost 97% of total water and only a small amount

about 3% is fresh water. The useable water is less

than 1% of fresh water available [1-7]. Global water

consumption is doubling every 20 years, more than

twice the rate of human population growth. If current

trends persist, in 2025, about 67% of population will

be under water stress as shown in Figure1.

1.1- Seawater Desalination

To address these looming crises of fresh water, the

sea is only unlimited source of water that can be use

to fuel the world population in future. The main

bottleneck in direct use of sea water is its salinity

level (≥35,000ppm). The sea water can be desalinated

by desalination methods to reduce the salt

concentration (≈500ppm) to make it useable.

Seawater desalination is being applied at 58% of

installed capacity worldwide, followed by brackish

water desalination accounting for 23% of installed

capacity [2]. The Figure 2 shows the worldwide

installations on the basis of feed water.

In thermal desalination systems, the MED process

has great potential because of high recovery ratio and

high thermal efficiency. Many researchers like; A. E.

Al-Nashar et al. [8], M.Al Shammiri et al. [9], A.O.

Bin Amer [10], J. Blanco et al. [11] and H.K.

Sadhukhan [12] provided the extensive detail on

MED design and operation. The top brine

temperature (TBT) of traditional MED is ranges from

100~ 120C. This high TBT accelerate the corrosion

and fouling of evaporator at high salt concentration.

High corrosion and fouling rate increase the capital

IMAT-UI 026

Figure 1: World population under water stress in

future

(Source: http://experience.sika.com/innovations)

Figure 2: Worldwide desalination

installations on the basis of

feed water [2].

mailto:*[email protected]


2012

155

and operational cost of plant and also reduce the plant

life.

In conclusion, if above mentioned problem (high

TBT) can be solved; MED can be an attractive

process for desalination in future. To solve this, the

traditional MED is combined with an AD beds and

this novel system called as MED+AD desalination

cycle. The AD cycle shift the whole desalination

cycle toward lower operating temperature (5~50C)

and also increase the performance of plant.

MED+AD cycle simulation is completed by using

FORTRAN linked with IMSL. It is found that by combining the MED with AD, the production rate

increases by 35-40%.

2. MED+AD MODELING

MED+AD is the combination of traditional MED and

adsorption cycle AD. The vapors from last stage from

MED are adsorbed by AD beds, and the desorbed

vapors from AD bed injected back between steam

generator and the first stage of MED. To overcome

the pressure losses due to low pressure desorbed

vapor mixing, steam jet ejector is used. This jet pump

helps to recover the pressure loss due to mixing of

low pressure desorbed vapors. The detailed

model of three stage MED coupled with AD is shown

in Figure 3.

The modeling of the system is completed by using

1) mass conversation, 2) energy conservation and,

3) salt conservation equations. The overall heat

transfer coefficient is calculated by using falling

film correlation developed by M.W. Shahzad et al.

[13] for low pressure evaporation. This novel

correlation is given in equation 1. For the ejector,

in addition to these three equations, momentum

conservation is applied to calculate the throat

velocity and throat diameter. The throat diameter is

designed to make sure that the primary steam

pressure must drop below the desorption pressure

to pull the desorbed vapors.

47.084.0

89.038.0

85.345.0

16.0

32

2

.65.2

1exp.2

PrRe

...00143.0

ref

g

ref

sat

o

ll

l

nevaporatio

v

v

T

q

T

T

S

S

kg

h

(1)


MED and AD modeling equations are written in

separate user defined sub-routine of FORTRAN. The

IMSL is used to solve the equations simultaneously.

The tolerance 1x10-7

is used to converge the solution.

Figure 4 shows the adsorber bed and desorber bed

temperatures profiles. The temperature of bed is

increasing during desorption due to heat supply by

hot water for desorption. On other hand, the

temperature of bed is dropping during absorption due

to heat taken by cold water. The switching time helps

the beds (pre-heating & pre-cooling) to prepare for

next operation.

The Figure 5 shows the distillate production from

traditional MED and an advanced MED+AD plant.

The dotted lines represent the typical MED

production. It can be seen that when MED coupled

with AD machine the production increases from 35-

40%. The desorbed vapour pulled by primary vapour

produced in SG in steam jet ejector and mixture of

these two streams introduced into the 1st stage of

MED. There is sharp increase in distillate production

Figure 3: MED+AD detailed

model

Figure 4: AD beds temperatures

profile


2012

156

and then drop gradually due to desorption rate

decrease. It can also be seen that during switching

when no desorption vapour are available the

production is just by traditional MED plant and it

catching the dotted lines of respective effects.

The Figure 6 shows the temperature profiles of AD+

MED stages. The SG and two stages have feed pre-

heater loop inside the chamber to recover the heat

from brine. The dotted line shows the feed

temperature increase from ambient (30C) in feed pre-

heaters. The pressure drop due to mixing of desorbed

vapour is recovered by jet pump. It is observed that

temperature difference between stages varies from 2-

3C.

The salt concentration increase in each effect is

shown in Figure 7. The brine is cascaded to get the

flash affect in next lower pressure stages.

4. CONCLUSION

To overcome the limitations of conventional MED

plants, an advanced desalination cycle is proposed.

The proposed cycle is combination of traditional

MED and AD cycle. In MED+AD cycle, top brine

temperature (TBT) reduced to 50C as compared to

120C in practical traditional MED plants. The low

grade waste heat or solar energy can operate the

system. The distillate water production rate is 35-

40% higher than traditional MED plants as well as

ambient energy is scavenging in last stages of MED.

It is also found that the highest concentration

(120,000ppm) of brine exposed to lowest

temperatures (5C) that reduces the chances of

corrosion and fouling of MED evaporators.

NOMENCLUTURE

l = Liquid viscosity (kg/m-sec)

l = Liquid density (kg/m3)

lk= Liquid conductivity (W/m-K)

Re= Film Reynolds number

Pr = Prandtl number

S = Feed water salinity (ppm)

oS= Reference sea water salinity (30000ppm)

satT=Evaporator saturation temperature (K)

refT= Reference saturation temperature (K) ( refT

=

322.15K )

q = input heat flux (W/m2)

gv= vapor specific volume

ΔT= Tch,out – Tevap

Figure 5: MED+AD production

profile temperatures profile

Figure 6: Temperature profiles of MED+AD

Figure 7: MED+AD salt concentration

profile


2012

157

ABBREVIATION

MED = Multi effect desalination

AD = Adsorption desalination

IMSL =International math and state library

TBT = Top brine temperature

SG = Steam generator

ACKNOWLEDGEMENT

The authors wish to thanks to Dr. Aung Myat for help

and guidance during simulation programming.

REFERENCES [1]. S. A. Kalogirou, Seawater desalination using

renewable energy sources, Progress in Energy

and Combustion Science 31 (2005) 242–281.

[2]. M. A. Eltawil, Z. Zhengming, L. Yuan, A review

of renewable energy technologies integrated

with desalination systems, Renewable and

Sustainable Energy Reviews 13 (2009) 2245–

2262

[3]. J. Buff, P. Re, R. Glynn, Benfield, Silvio

Fischer, Desalination Plants – Technological

development, Risks affecting Engineering

Insurers and Claims Experience, IMIA - WGP

57 (08) Conference Gleneagles, 16th Sep. 2008.

[4]. P. Dickie, Desalination: option or distraction for

a thirsty world? , WWF‘s Global Freshwater

Programme, June 2007,

(www.melaleucamedia.com).

[5]. A. D. Khawaji, I. K. Kutubkhanah, J. m. Wie,

Advances in seawater desalination technologies ,

Desalination 221 (2008) 47–69

[6]. http://environment.nationalgeographic.com/envir

onment/freshwater/freshwater-crisis/

[7]. http://www.globalchange.umich.edu/globalchang

e2/current/lectures/freshwater_supply/freshwater

.html.

[8]. A. M. El-Nashar, A. A. Qamhiyeh, Simulation of

the steady state operation of a multi effect stack

seawater distillation plant, Desalination 101

(1995) 231-243.

[9]. M. A1-Shammiri, M. Safar, Multi-effect

distillation plants: state of the art, Desalination

126 (1999) 45-59.

[10].A.O. Bin Amer, Development and optimization

of ME-TVC desalination system, Desalination

249 (2009) 1315–1331.

[11].J. Blanco, E. Zarza, D. Alarcón, S. Malato, J.

León, Advanced Multi-Effect Solar Desalination

Technology: The PSA Experience , CIEMAT -

PSA, P.O. Box 22, 04200 Tabernas (Almería),

Spain.

[12]. H. K. Sadhukhan and P. K. Tewari, Small

Desalination Plants (SDPS), Thermal

Desalination Processes – Vol. II - Small

Desalination Plants (SDPs), Bhabha Atomic

Research Centre, Mumbai 400085, India.

[13]. M. W. Shahzad, A. Myat, C. W. Gee and K. C.

Ng, Bubble-assisted film evaporation correlation

for saline water at sub-atmospheric pressures in

horizontal-tube evaporator, Applied Thermal

Engineering 50 (2013) 670-676.

Figure 7: MED+AD salt concentration profile

http://environment.nationalgeographic.com/environment/freshwater/freshwater-crisis/

http://environment.nationalgeographic.com/environment/freshwater/freshwater-crisis/

http://www.globalchange.umich.edu/globalchange2/current/lectures/freshwater_supply/freshwater.html




2012

158

Kinetics of Propane Adsorption on Maxsorb III Activated Carbon

Azhar Bin Ismaila, Loh Wai Soong

a, Ng Kim Choon

a*

aDepartment of Mechanical Engineering,


9 Engineering Drive 1, Singapore 117576

*E-mail : [email protected]

ABSTRACT

Experimental kinetics results of propane in Maxsorb

III activated carbon is obtained at temperatures of

10°C and 30°C, and pressures up to 800kPa using a

magnetic suspension balance. A multi-gradient linear

driving force (LDF) approximation is used for

adsorbate uptake as a function of time. The LDF

mass-transfer-rate coefficients were thus determined.

Using this approach, the experimentally derived LDF

coefficients based on independently measured kinetic

parameters for propane in the activated-carbon bed

agree very well with experimental results. The

computational efficiency is gained by adopting this

extended LDF model.

Keywords : Adsorption, Adsorption Chiller,

Adsorption Kinetics

1. INTRODUCTION

Interest in adsorption refrigeration (AD) has grown

due to its advantages related to its direct utilization of

thermal energy sources such as low grade waste heat

from various industrial sources, solar hot water as

well as geothermal sources. The study of adsorption

in the National University of Singapore is a long

standing and continuous project aimed to achieve

higher refrigeration capacity, better Coefficient of

Performance (COP) and an exploration of diverse

applications of the thermal heat pump system [1-2].

In this work, the kinetics of adsorption of propane at

various temperatures and pressures are presented as

an ongoing study of utilizing alternative refrigerants

as adsorbate in an AD system. The Linear Driving

Force (LDF) approach has been adopted to represent

the uptake curves as a function of time. However, the

constant k in this model has been shown to be

subjected to the effects of both temperature and

pressure differences. Due to the sudden compression

effects during charging as well as the isosteric heat

released, the kinetics experiment is non-isothermal as

stipulated by the LDF model. As such, Loh et al

(2012) [3] introduced a model to take into account

the non-isothermal effects on the adsorption kinetics

of assorted adsorbates on Maxsorb III activated

carbon during a Constant volume variable pressure

(CVVP) experiment. He et al [4] on the other hand

identified the effects of pressure differences. In this

paper, a simplified LDF model is implemented to

fully describe the non-isothermal adsorption kinetics

of propane on Maxsorb III.

2. EXPERIMENTAL

2.3 Materials

Figure 1: Schematics Diagram for the magnetic

Suspension Balance unit (Rubotherm)

A magnetic suspension balance (Rubotherm) is used

to measure the instantaneous uptake of the adsorbent

as shown in Figure 1. This balance measures the

weight of the sample with a reproducibility of ±0.03

mg. The advantages of this balance are the high

accuracy and long-time stability due to the balance

being outside the measuring cell and having no

contact to any solvent vapor. Furthermore,

continuous data readout via PC is possible. Maxsorb

III (by Kansai Coke Company, Japan) is utilized with

pure propane, purity 99.5% is utilized. The values of

derived quantities are extracted from NIST

(REFPROP).

2.4 Adsorption Experiments

2.4.1 Buoyancy Correction

Buoyancy forces are taken into account to correct the

influence of gas density on the measured apparent

IMAT-UI 027


2012

159

weight of the sample. The displacements of gas by

the sample holder, solid adsorbent, and adsorbed

phase are taken into consideration. The correction

due to the sample holder is obtained with blank

experiments performed at different pressures with the

empty holder. The buoyancy due to the solid matrix

of the adsorbent, which results in an apparent weight

loss, is estimated as the product of the skeletal

volume of the adsorbent and the gas density. Finally

the buoyancy effect exerted on the adsorbed phase is

corrected to obtain the absolute adsorption isotherm

(P,T) .

The weight, m, displayed by the balance results from

the net force exerted on the sample:

m = mh (1-ρg⁄ρh ) + ms [1- ρg ⁄ρs +q (1-ρg ⁄ρa)] (1)

Here, mh and ρh are the mass and density of the

sample holder respectively, ms and ρs are the mass

and density of the adsorbent sample, respectively, ρg

and ρa is the density of the bulk gas and adsorbed

phase respectively at the equilibrium pressure and

temperature.

The blank experiments with an empty holder give the

mass and density of the holder from the intercept and

slope of the linear decrease of apparent weight with

gas density:

m = mh – (mh/ρh) × ρg (2)

The density (ρg) of Nitrogen gas is obtained using

values from REFPROP. Adsorption experiments

using a non-adsorbing gas such as helium at a high

temperature of 120°C, provide the mass (ms) and

density (ρs) of the carbon sample:

m - mh (1-ρg ⁄ρh ) = ms - (ms⁄ρs) × ρg (3)

Here, it is assumed that helium acts as an inert gas

that penetrates into all the accessible pore volume of

the carbon without being adsorbed.

Finally, the experiments with the carbon sample

provide the uptake q:

msq(1-ρg⁄ρa) = m - mh(1-ρg⁄ρh) - ms(1-ρg⁄ρs ) (4)

The value of ρa is estimated using the approximation

by Ogawa (1975) [5], given by

ρa = ρa* ⁄ exp [αe (T-Tb ) ] (5)

where ρa indicates the density of the liquid at the

normal boiling point Tb, and αe indicates the thermal

expansion of the superheated liquid. The pressure

dependencies of ρa, ρa*

and αe is negligibly small in

the pressure range of the present work and are thus

neglected. The normal boiling point and the density

of the liquid at the normal boiling point are taken

from Miyamato and Watanabe (2000) [6]. Further,

the value of the thermal expansion was assumed to be

independent of the species of the adsorbate, and the

mean value of the thermal expansion of liquefied

gases (αe = 2.5×10-3

K-1

) was used in the numerical

calculation.

2.4.2 Measurement Procedure

The pressure controller is set to the desired pressure,

and water from the water bath enters the jacket to

maintain the desired temperature. When equilibrium

is reached, the valve is opened to allow the propane

gas to enter the chamber, and the pressure,

temperature and weight changes are logged in the

data logger.

3. SUPPORTING THEORY

3.1 Linear Driving Force (LDF) Model

The LDF model or Lumped Parameter Model (LPM)

[7] has been used to describe adsorption kinetics at

isothermal conditions. It is advantageous in that it is

easy to incorporate in simulation programs given the

computational ease. The LDF model describes the

kinetics of adsorption well due to averaging of the

kinetic properties at the particle, the column, and the

overall cyclic steady state levels [8]. The

characteristics of the models describing the local

rates of adsorption at the particle level are also lost

during these integration processes. In this model, the

mass transfer equation is described by a driving

force, defined as the difference between the

equilibrium uptake (q*) and the instantaneous uptake

(q) [9].

dq/dt = k [q* - q(t)] (6)

where k is the effective particle-phase transfer

coefficient as a function of adsorbate concentration.

The heat transfer equation is then given by

cp· dT/dt = QST· dq/dt – ha(T-To) (7)

The initial and final conditions are

at t = 0, q=qo, T=To (8)

at t = ∞, q=q∞, T=To (9)

qo=qo*(po,To), q∞=q∞

*(p∞,T∞), q=q

*(p∞,T) (10)

q is the adsorbate loading per kg of adsorbent at time

t while qo and q∞ are the equilibrium loading at the

initial and final conditions respectively. T is the

adsorbate temperature at time t while To is the initial

and final temperature of the adsorbate. cp is the heat

capacity of the adsorbent, while h is the external heat

transfer coefficient while a is the external heat

transfer area. QST on the other hand is the isosteric

heat of adsorption. For a differential test where the

changes in the adsorbate loading and temperature are

small, q* can be written as

(q∞- q*)= (∂q

*/∂T)q=q∞, T=To (To-T) (11)

Equation (6), (7) and (11) may be solved

simultaneously to give [10]


2012

160

(q-qo)/( q∞- qo) = (βα2/(1- βα

2))exp(rt)

- exp[-k(1-αβ)t]/ βα2 (12)

where

α = k/(k+r) (13)

β = (QST/cp) (dq*/dT) q=q∞, T=To (14)

r = (-k/2){(1-β+λ)-[(1-β+λ)2-4λ]

1/2} (15)

λ = (ha)/(cpk) (16)


4.1 Blank measurements and Buoyancy

Corrections

The blank measurements with Nitrogen gas at

different densities gave a good straight line fit as

shown in Figure 2 to obtain an empty cylinder mass

of 4.4989g and density of 8282kg/m3. The buoyancy

measurements on the other hand were carried out

with inert Helium gas at a high temperature of 120°C

with different pressures to achieve the desired

densities. These buoyancy measurements as shown in

Figure 3 gave a density of 2.2g/cm3 for the Maxsorb

III activated carbon and a mass of 0.1547g of solid

adsorbent in the testing chamber. These results are

tabulated and summarized in the following Table 1.

Table 1: Mass and Densities of Empty Cell and Adsorbent

Mass (g) Density (g/cm3)

Empty Cell 4.4989 8.3

Maxsorb III Adsorbent 0.1547 2.2

Figure 2: Blank Measurements of the empty cylinder

with Nitrogen Gas.

Figure 3: Buoyancy Measurements of the empty

cylinder with Helium gas at high

temperatures of 120°C.

4.2 Adsorption Kinetics of Propane on

Maxsorb III at 10°C and 30°C

The Adsorption kinetics of Propane on Maxsorb III at

temperatures of approximately 10°C and 30°C are

measured and evaluated for the first 300s as

presented in the following Figures 4 and 5

respectively. The LDF model for non-isothermal

adsorption presented earlier are utilized to curve-fit

the data.

Figures 3 and 4 show the experimental temperature

profiles of the adsorbate during the experiment. The

temperature of the adsorbent, which is Maxsorb III

follows the profile when the adsorbent-adsorbate

system is at thermal equilibrium at the beginning of

the experiment and as t approaches infinity,

Figure 3: Temperature Curves of Maxsorb III-

Propane: experimental data at ▲-

To=10.96°C, P∞=497 kPa,

experimental data at ●- To=9.15°C,

P∞=192kPa

time (s)

tem

pe

ratu

re (

°C)

density (kg/m3)

mas

s (k

g)

density (kg/m3)

mas

s (k

g)


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Figure 4: Temperature Curves of Maxsorb III-

Propane: experimental data at ▲-

To=28.60°C, P∞=700 kPa,

experimental data at ●- To=28.89°C,

P∞=497 kPa, ■- To=29.27°C,

P∞=195kPa

From equation (12), it may be shown that as time

approaches infinity,

(q-qo)/( q∞- qo) = 1 + (βα2/(1- βα

2))exp(rt) (12)

Hence, using the data from the buoyancy

measurements and the mass recorded by the

MessproTM

software, the uptake of the Masxorb III –

Propane at time t is calculated and curve 1-(q-qo)/(q∞-

qo) can be plotted, and the gradient obtained to give

the value of r and the cut at the y axis gives the value

of -ln(-βα2/(1-βα

2)). The 1-(q-qo)/(q∞-qo) profiles for

To=28.89°C, P∞=497 kPa and To=28.60°C, P∞=700

kPa are presented in Figure 5 as examples.

Figure 5: 1-F(t) profiles for ∆-To=28.89°C,

P∞=497 kPa, experimental data at

○-To=28.60°C, P∞=700 kPa, ---

straight lines to obtain the gradient

which gives the value of r and

intersection at y axis giving the

value of -ln(-βα2/(1-βα

2)).

The parameters q∞, k and the respective errors were

evaluated from 10s onwards and presented in the

following Table 2. The reason for this is in the first

10s, the pressure has not stabilized, and the zero error

correction cannot be determined. β was evaluated to

be 2.0 during these experiments.

Table 2: Parameters of Kinetics Model

T

(°C)

P

(kPa)

q∞

(g/cm3)

k

(1/s)

RMS

error

(%)

9.15 192 0.876 0.00304 1.25

10.96 497 0.729 0.00784 0.65

29.27 195 0.592 0.00495 1.57

28.89 497 0.792 0.00542 1.19

28.60 700 0.835 0.00841 0.75

The fitted models were plotted alongside the

experimental data as shown in Figure 6 and 7.

Figure 6: Uptake kinetics of Maxsorb III-

Propane: experimental data at ○-

To=9.15°C, P∞=192kPa,

experimental data at ∆-To=10.77°C,

P∞=300 kPa, □-To=10.96°C,

P∞=497 kPa --- fitted curves from

the non-isothermal adsorption

kinetics model.

Ln[1

-F(t

)]

time (s)

time (s)

tem

pe

ratu

re (

°C)

up

take

(kg

/kg)

time (s)


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Figure 7: Uptake kinetics of Maxsorb III-

Propane: experimental data at □-

To=29.27°C, P∞=195kPa,

experimental data at ∆-To=28.89°C,

P∞=497 kPa, experimental data at ○-

To=28.60°C, P∞=700 kPa, --- fitted

curves from the non-isothermal

adsorption kinetics model.

The value of k as expected increases with pressure at

a given temperature, since the value of q* will be

higher at every point in the kinetics curve resulting in

a higher driving force resulting in larger k value.

7. CONCLUSION

The regressed value fit the experimental data very

well for use of simulation in determining the

performance of an AD chiller that runs with propane.

ACKNOWLEDGMENT

The researcher, Azhar Bin Ismail is supported by the

National Research Foundation Singapore under its

National Research Foundation (NRF) Environmental

and Water Technologies (EWT) PhD Scholarship

Programme and administered by the Environment

and Water Industry Programme Office (EWI).

REFERENCES

[1] K.C. Ng, X. Wang, Y.S. Lima, B.B. Saha, A.

Chakarborty, S. Koyama, A. Akisawa and T.

Kashiwagi ―Experimental study on performance

improvement of a four-bed adsorption chiller by

using heat and mass recovery‖ International

Journal of Heat and Mass Transfer, vol. 49, no.

19-20, pp. 3343–3348, September 2006.

[2] B.B. Saha, A. Chakarborty, S. Koyama, K.

Srinivasan, K.C. Ng, T. Kashiwagi, P. Dutta,

―Thermodynamic formalism of minimum heat

source temperature for driving advanced

adsorption cooling device‖ Applied Physics

Letters, vol. 91, no. 11, pp. 111902 - 111902-3,

September 2007.

[3] W.S. Loh, A. Chakraborty, B. B. Saha , and K.

C. Ng ―Experimental and Theoretical Insight of

Nonisothermal Adsorption Kinetics for a Single

Component Adsorbent–Adsorbate System‖ J.

Chem. Eng. Data, vol. 57, no. 4, pp. 1174–1185,

April 2012.

[4] I.I. El-Sharkawy ., J.M. He, K.C. Ng, C. Yap and

B.B. Saha ―Adsorption Equilibrium and Kinetics

of Gasoline Vapors onto Carbon-Based

Adsorbents‖ J. Chem. Eng. Data, vol. 53, no. 1,

pp. 41–47, December 2007.

[5] S. Ozawa, S. Kusumi and Y. Ogino ―Physical

adsorption of gases at high pressure. IV‖ An

improvement of the Dubinin--Astakhov

adsorption equation. Journal of Colloid and

Interface Science, vol. 56 no.1, pp. 83-91, 1976.

[6] H. Miyamoto and K. Watanabe ―A

Thermodynamic Property Model for Fluid-Phase

Propane‖ International Journal of

Thermophysics, vol. 21, no. 5, pp. 1045-1072,

September 2000.

[7] D.D. Do, Adsorption analysis: equilibria and

kinetics. Imperial College Press London Vol. 2.

1998.

[8] S. Sircar and J.R. Hufton ―Why Does the Linear

Driving Force Model for Adsorption Kinetics

Work?‖ Adsorption, vol. 6, no. 2, pp. 137-147,

June 2000.

[9] J. I. Coates and E. Glueckauf ―Theory of

chromatography. Part III. Experimental

separation of two solutes and comparison with

theory‖ J. Chem. Soc, no. 0, pp. 1308-1314,

1947.

[10] S. Sircar ―Linear-driving-force model for non-

isothermal gas adsorption kinetics‖ J. Chem. Soc,

vol. 79, pp. 785-796, 198

up

take

(kg

/kg)

time (s)


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ABSTRACT

This paper presents an investigation on the effects

of using solar chimney, gable vent, and the

combination of the two natural ventilations on the

average air temperature and air-flow condition

inside a double-storey house in Malaysia, using

computational fluid dynamics (CFD) method. The

representative model of the house comprises of a

main hall, a kitchen and an upper hall. Both

temperature and air velocity boundary conditions

were prescribed on the model. Results of the

simulation indicates that the average temperature of

the air in the house at 1 pm closely matched the

measured values. It was found that the average

temperature of the air in the house is not so

significantly affected by the types of natural

ventilation used. Opening the kitchen door

causesthe air to flow from the main and upper

halls towards the kitchen and causing a bottle neck

at the pathways. A more uniform air flow is

obtained when solar chimneys are used. When

gable vents are used, high intensity air flow occurs

in the main hall and it spreads uniformly towards

the kitchen and upper hall. The air-flow intensity

becomes even higher in the main and upper halls

when a combination of solar chimney and gable

vents are incorporated into the CFD model.

Keywords : Natural ventilations, Average air

temperature, Air-flow conditions,

CFD simulation, Solar chimney,

Gable vents.

1. INTRODUCTION

Natural ventilation has attracted a strong growing

interest in building sectors because of its potential

advantages over mechanical ventilation systems, in

terms of energy requirement, economic and

environmental benefits. Mechanical ventilation

systems have undesirable energy implication since

they require more electricity to run [1]. Earlier

work on natural ventilation mainly concerned with

aerodynamic loading [2] and they were carried out

in wind tunnel. But with the advancement of

Effects of Natural Ventilations on Indoor Air

of a Double-Storey Residential House in Malaysia


b & Kam Jia Liq


Universiti Teknologi Malaysia, Skudai, Johor

Tel : (+607) 5534748. Fax : (+607) 5566159


bFaculty of Mechanical Engineering


Tel : (+607) 5534749. Fax : (+607) 5566159


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computing technology, more complicated studies

have been conducted using computational fluid

dynamics (CFD) techniques. Nikas et al. [3]

showed that it is possible to get information about

induced velocity and pressure fields for natural

cross ventilation using CFD modelling which

otherwise are quite difficult to extrapolate from

experimental methods.

Various strategies have been proposed in the

literature to enhance buoyancy effect so that

adequate air flow rate and a desired level of

thermal comfort can be achieved inside a building.

One good example is a solar chimney, which is

designed to maximize ventilation effect by

maximizing solar gain [4]. This creates a sufficient

temperature difference between the inside and

outside of the building to drives an adequate air

flow rate. Solar chimney is a thermo-syphoning air

channel in which the principal driving mechanism

of air flow is through thermal buoyancy [5]. One

can find different variations in solar chimney

design, which is affected by a number of factors

such as the location, climate, orientation, size of the

space to be ventilated and the internal heat gains

[6].

Computational methods based on CFD technique

have been used by many to predict flow pattern

inside the chimney as well as in the space (room)

adjoining the solar chimney. The existing CFD

models are able to predict velocity and temperature

profiles along with other flow characteristics

accurately. However, they usually do not consider

the thermal energy storage in the walls of the

building [7]. Nevertheless, the use of CFD

modelling in solar chimney study has been

increasing. These studies have greatly contributed

to the present understanding of the solar chimney.

In this study, we used the CFD method to

investigate the effect of using natural ventilations

in a double-storey residential terrace house in

Malaysia. The natural ventilations considered are

solar chimney, gable vents and the combination of

the two. The focus of this study is not on the types

of ventilation. The main goal of this study is to find

out the effect of using these ventilations on the

thermal and flow conditions of the air inside the

house. For that purpose, the solar chimneys are

represented only as simple square openings located

on the roof of several sections of the house. The

gable vents on the other hand are represented by

long rectangular openings on the upper part of

several walls of the house. During the simulation,

both air velocity and temperature are prescribed on

these openings to model the outward air flow from

the house.

2. METHODOLOGY

2.1 Computational Domain

Figure 1 shows a representative model of the

house and a computational domain for the CFD

simulations. It consists of three sections namely the

main hall, the upper hall and the kitchen. There are


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only four walls that are considered to be exposed to

solar radiation. These are the eastern and southern

walls of the main hall, northern wall of the kitchen

wall, and the eastern wall of the upper halls. Other

walls are considered to be insulated and are at the

same temperature as the air in the house, which is

at 29C (302K).

2.2 Actual Average Air Temperature &

Humidity

The actual average dry-bulb temperature, wet-bulb

temperature and relative humidity of the interior air

were determined in the three sections of the house:

the main hall, the upper hall and the kitchen, using

a sling psychrometer, for every hour beginning

from 9 am until 4 pm. In each section, all the data

were measured at several locations and then the

average values were computed, for every hour. The

complete hourly data are shown in Table 1. It is

observed that the average dry-bulb temperature of

the air is about 30C and relative humidity is

around 73%.

Figure 1: A representative model of the house

considered for the CFD computational domain

(rear view).

Table 1: Hourly data for the air inside the house.

2.3 Validation of CFD Simulation Procedure

To validate the numerical simulation procedure, a

CFD simulation was performed on the model of the

house to represent a condition when there are no

ventilations. However, a door on the rear wall of

the kitchen was left fully opened to generate some

air flow within the house. We call this as a ―base

case‖ condition. The goal of this simulation is to

estimate the average temperature of the air in the

various sections of the house and compare them

with the actual temperatures measured at 1 pm,

when the kitchen door was opened.

Both temperature and air velocity boundary

conditions were used. A uniform temperature of

47C (320K) was prescribed on the wall of the

main hall facing south and the wall of the kitchen

facing north. A uniform temperature of 29C

(302K) was prescribed on the wall of the main hall

KITCHEN MAIN HALL

SOLAR CHIMNEY

UPPER HALL

GABLE VENT


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facing east and the wall of the upper hall facing

east. A constant inlet air velocity of 0.1 m/s, at a

temperature of 29C (302K), was prescribed on all

the door seams. These door seams represent the

gaps between the doors and the slabs and the

clearance between the doors and the walls. The air

velocity boundary conditions allow a turbulent

analysis to be performed on the computational

domain. Turbulent flow analysis using a k- model

with 10% turbulent intensity was performed on the

CFD model until an acceptable convergence was

attained.

2.4 Modeling of the Natural Ventilations

Three natural ventilation systems were considered

in this study, namely a solar chimney, gable vents

and the combination of the two. The solar chimney

is a natural-draft device that is used in many

passive cooling applications for residential houses.

Density of air decreases with increasing

temperature. It means that air with higher

temperature than ambient air is driven upwards by

the buoyancy force. A solar chimney exploits this

physical phenomenon and uses solar energy to heat

air up. Gable vents are usually placed at the top of

the gable on the end of the house. This is to create a

draft through the space by having both intake and

exhaust vents. While gable vents do increase

ventilation they do not offer uniform air flow

within the space and their ability to move large

amounts of air are limited.

The solar chimneys were incorporated into the

CFD model by adding square-shaped openings on

the roof (at the middle) of the main hall, kitchen

and the upper hall. A constant air outlet velocity of

0.3 m/s, at 29C (302K) was prescribed on all the

solar chimneys while a constant inlet air velocity of

similar magnitude was prescribed on all the door

seams as the boundary conditions. The same

temperature boundary conditions as in the base

case were employed in this CFD simulation.

The gable vents were incorporated into the CFD

model by introducing thin rectangular-shaped

openings on the walls of the house. The width of

these openings was made nearly the same as the

width of the walls. An inlet gable vent was placed

on the eastern wall of the main hall, while outlet

gable vents were placed on the southern wall of the

main hall, the northern wall of the kitchen and the

eastern wall of the upper hall. A constant air outlet

velocity of 0.3 m/s, at 29C (302K) was prescribed

on all the solar chimneys while a constant inlet air

velocity of similar magnitude was prescribed on all

the door seams as the boundary conditions. The

same temperature boundary conditions as in the

base case were used in this CFD simulation.


3.1 Base Case Conditions

Results of the CFD simulation for the ―base case‖

condition give an average air temperature of about


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30.4C in both the main hall and the kitchen, and

about 29.7C in the upper hall. These values are

superimposed on the plots of measured temperature

vs. time (within a circle) for the three sections on

the house, shown in Figure 2. It can be seen that the

average temperatures obtained from the CFD

simulation fall within the acceptable range of the

measured temperature range. Thus it is safe to say

that the CFD model, the boundary conditions used

and the turbulent analysis model employed in the

simulation are valid and can be further used in the

proceeding simulations. The air flow distribution in

the house obtained from the CFD simulation for the

―base case‖ conditions is shown in Figure 3. It can

be seen that, with the door on the eastern door of

the kitchen left opened, the air tends to flow from

the main and upper halls towards the kitchen,

producing a bottle neck at the pathway connecting

the main hall and the kitchen. The air flow is seen

fairly uniform in both halls.

Figure 2: Comparison between the average air

temperature obtained from the CFD simulation

and the measured values.

Figure 3: Air-flow distribution (m/s) inside the

house when the door on eastern wall

of the kitchen is left opened.


house when solar chimney

ventilation is used.

3.2 The Effect of Solar Chimney Ventilation

Results of the CFD simulation when solar

chimneys are incorporated into the model give an

average air temperature of 302.8 K in the main hall,

302.9 K in the kitchen, and 302.3 K in the upper

Kitchen

Main Hall

Upper Hall

CFD

N

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W


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hall. These are slightly lower than the average air

temperature for the base case condition [303.4 K

(hall); 303.3 K (kitchen); 302.7 K (upper hall)]. On

average, the CFD simulation results indicate that

the average air temperature in the house is reduced

by about 0.6C when three solar chimneys were

incorporated into the model. This is considered as

an insignificant improvement on the average

temperature of the air inside the house. The air

flow condition in the house when solar chimneys

are used is shown in Figure 4. It can be seen that

the air flow is fairly uniform in all three sections of

the house. Slightly higher air velocity occurs at all

the door seams (inward flow) and the solar

chimneys (outward flow). A swirling air flow

condition can be seen near the northern wall of the

kitchen and the southern wall of the main hall.

3.3 The Effect of Gable Vents

Results of the CFD simulation when gable vents

are incorporated into the model give the average air

temperature of 302.4 K in the main hall and in the

kitchen, and 302.2 K in the upper hall. These are

slightly lower than the average air temperature for

the base case condition [303.4 K (hall); 303.3 K

(kitchen); 302.7 K (upper hall)]. On average, the

CFD simulation results indicates that the average

air temperature in the house is reduced by about

0.8C when the gable vents were incorporated into

the CFD model. This can also be considered as an

insignificant improvement on the average

temperature of the air inside the house.

The air flow condition in the house when gable

vents are used is shown in Figure 5. It can be seen

that the air flow is fairly uniform in the main hall

and the kitchen but it is less intense in the upper

hall section. Higher air velocity condition can be

seen at the vicinity of all the gable vents, especially

at the inward flow gable vent on the eastern wall of

the main hall. No swirling air flow condition can be

seen in the figure.


house when gable vents are used.

3.4 The Effect of Combined Solar Chimney

Ventilation & Gable Vents

Results of the CFD simulation when a combination

of solar chimneys and gable vents are incorporated

into the model give an average air temperature of

302.2 K in the main hall, 302.3 K in the kitchen

and 302.1 K in the upper hall. When compared to

the base case conditions, it is found that the

average air temperature is dropped by about 1.1C,

1.0C and 0.6C in the main hall, kitchen and the

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upper hall, respectively. These are considered a

mild reduction in the average air temperature inside

the house.


house when a combination of solar

chimneys and gable vents are used.

Figure 6 shows the air flow distribution inside the

house when a combination of solar chimneys and

gable vents are used. It is seen that high intensity

air flow occurs in the main hall and the air appears

to move towards the kitchen. The air flow in the

upper hall and the kitchen appears to be less

intense. Higher air velocity is seen at both the inlet

and outlet gable vents. No swirling air flow can be

seen from the figure. Also, the air tends to flow

toward the solar chimneys located on the ceiling of

each section of the house.

4. CONCLUSION

A CFD simulation method has been used to

investigate the effects of several natural

ventilations, namely solar chimney, gable vent and

the combination of both, on the conditions of the

air inside a double-storey residential house in

Malaysia. It was found that the average

temperatures of the air at various sections of the

house, obtained from the CFD simulation for the

base case condition, agree quite well with the

measured values at 1 pm. The average air

temperature drops by about 0.6C when solar

chimneys are used and about 0.8C when gable

vents are incorporated into the CFD analysis. The

temperature drops by about 1C when the

combination of both ventilations are included in the

analysis. When the kitchen door is left opened, the

air tend to flow from the main hall and upper hall

towards the kitchen. Using solar chimney

ventilation results in a more uniform air-flow inside

the house. High intensity air flow occurs in the

main hall and it spreads uniformly towards the

kitchen and upper hall when inlet and outlet gable

vents are used. The air-flow intensity becomes

even higher in the main and upper halls when a

combination of solar chimney and gable vents are

incorporated into the computational model.

REFERENCES

[1] Rakesh Khanal & Chengwang Lei, Solar

chimney - A passive strategy for natural

ventilation, Energy and Buildings 43 (2011)

1811–1819.

[2] P.F. Linden, The fluid mechanics of natural

ventilation, Annual Review on Fluid

Mechanics 31 (1999) 201–238.

N

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[3] K.-S. Nikas, N. Nikolopoulos, & A.

Nikolopoulos, Numerical study of a naturally

cross-ventilated building, Energy and

Buildings 42 (2010) 422–434.

[4] N.K. Bansal, R. Mathur, M.S. Bhandari, Solar

chimney for enhanced stack ventilation,

Building and Environment 28 (3) (1993) 373–

377.

[5] G. Gan, Simulation of buoyancy-induced flow

in open cavities for natural ventilation, Energy

and Buildings 38 (5) (2006) 410–420.

[6] D.J. Harris, N. Helwig, Solar chimney and

building ventilation, Applied Energy 84 (2)

(2007) 135–146.

[7] A. Dimoudi, Solar chimneys in buildings – the

state of the art, Advances in Building Energy

Research 3 (2009) 21–44.


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Deposit Forming Tendency of Biodiesel and Diesel Fuel due to

High Pressure Exposure

Muhamad Adlan Abdullah, Arshad Salema and Farid Nasir Ani

Faculty of Mechanical Engineering, Universiti Teknologi Malaysia, Skudai, Johor D.T., Malaysia

ABSTRACT

Fuel deposit issues in common rail diesel fuel

system require attention as it will affect the

operation of the closely designed equipment

resulting in increased emissions and poor

performance. However, the fuel propensity in

forming deposit in high pressure common rail

system is inadequately understood. This paper

recounts a test program designed to investigate

the effects of the exposure to high pressure and

temperatures on the diesel and biodiesel fuel

tendency to form deposit. A test rig was built

allowing the fuel to be pre-stressed under

conditions of common rail system and the deposit

forming tendency was determined by using

modified Jet Fuel Thermal Oxidation Tester

(JFTOT) procedure as well as deposition on hot

surface. The result showed that the exposure to

high pressure and temperature such as in common

rail system increases the tendency of deposit

formation.

Keywords : diesel, biodiesel, fuel deposit,

common rail, high pressure diesel

1. INTRODUCTION

Typically, deposit in diesel fuel injection system

lies on the injector nozzles. This deposit blocks

the flow of the fuel, which affects the delivery.

Recently, new type of injector deposits in

common rail diesel fuel systems were reported [1]

to be on the internal of injectors. The new type of

deposit was hypothesized to be a result of

exposure to the high pressure in the common rail

system [2]. Fuels with known storage stability

may be instable after being exposed to operation

in the common rail system. Presence of some fuel

additive components was also demonstrated to

contribute to this deposit [1, 3].

Efforts to quantify the diesel fuel system deposit

covers both on engine tests [4, 5] and laboratory

rigs [6, 7]. Engine tests are often very lengthy and

costly. Hence, some research efforts focus on

using test rigs for the purpose of screening tests

for the deposit forming tendency.

This paper describes a test program to measure

the deposit forming tendency of diesel and

biodiesel after exposure to high pressure of a

commonrail engine using in-house bench test. It

attempts to investigate the tendency of diesel and

biodiesel to form both types of injector deposits –

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the injector tip (nozzle deposit) as well as the

internal injector deposit – after the exposure to

conditions experienced in common rail system.

2. EXPERIMENTAL SETUP AND

PROCEDURES

The experiments were conducted as per the

following sequence:

1. The diesel fuel is pre-stressed under high

pressure common rail system

2. The pre-stressed fuel is then subjected to the

deposit forming test using modified JFTOT as

well as the hot surface deposition test

The common rail test rig was built using

BOSCH CP3 common rail pump system as shown

in Figure 1. The fuel pressure was measured by

the pressure sensor located at the fuel rail.

Temperatures of the fuel in the tank, Ti, and at the

return line, To, were measured by means of

thermocouples. The fuel is circulated through the

system simulating the conditions in an engine

operation, through high pressure and returned to

tank at atmospheric pressure. In order to maintain

the fuel temperature, a water cooled heat

exchanger was installed at the fuel return line

from the rail. The test fuel is pre-stressed under

these conditions for a specified duration.

The pre-stressed fuel is then subjected to the Jet

Fuel Thermal Oxidation JFTOT III tester for the

deposit forming study as in Figure 2. Detail of the

procedure was described elsewhere [8].

In principle, the JFTOT exposes the fuel to high

temperatures that represents the conditions found

inside injector nozzles where deposits typically

occur. Thus, this test program attempts to emulate

the conditions where fuel is first exposed to high

pressure and temperature in the common rail

system, and subsequently delivered to the

injectors where it is heated and deposit is formed.

In this case, no fuel evaporation occurs.

The hot surface deposition test was adopted from

the works of Yusmadi [9] as shown in Figure 3.

The heater was used to supply heat to the

aluminium cylinder block until a desired

temperature was achieved i.e. 200°C, 225°C and

250°C. Fuel was dropped onto the surface of the

alumimium block during the deposition test. K-

Type thermocouple was used and was located at

the heater in the cylinder block. In the deposition

experiment, the fuel was dropped from a burette.

Fuel was filled in burette and then released as

droplets to the surface of the aluminium block at

specific rate and the resultant deposit was

observed.

The hot surface deposition test emulates the

deposit formation in the combustion chamber of

diesel engine such as on the injector tips. In this

case, fuel evaporation occurs and the deposit

formation is different from the JFTOT test.

Both the deposit formed under JFTOT and hot

surface deposition tests were quantified by using

JPI Varnish rating as in Figure 4.


2012

173

The test fuel used were typical diesel fuel

obtainable from the market and a biodiesel

produced from waste cooking oil, supplied by

Xtrac Tech Sdn. Bhd.

Figure 1: The common rail test rig

(a)

(b)

Figure 2: (a) The Jet Fuel Thermal Oxidation

Tester and (b) the heater tube

Figure 3: The hot surface deposition test

Figure 4: The JFTOT rating and JPI varnish rating


3.1 Effects of exposure to high pressure on

internal deposit

Initial work [8] had demonstrated that diesel fuel

exposure to high pressure and temperature of

common rail system can affect the deposit

formation. Figure 5 shows the JFTOT heater tube

rating for fuels after 4 hours of pre-stressing in the

common rail rig at different pressures and

temperatures. It is shown that for similar inlet

JFTOT rating

JPI Varnish rating


2012

174

temperatures Ti (fuels marked as B, D and E),

higher pre-stressing pressure tends to produce

more deposit (i.e. lower deposit rating). It was

also shown that the temperature at which the fuel

is exposed has significant effects on the deposit

formation.

It was also demonstrated that exposure to high

pressure and temperatures for as short a duration

as 30 minutes may affect the deposit forming

tendency. This is shown in Figure 6 which shows

the deposit forms at lower temperatures as the

pre-stressing time is increased (lower JPI rating

indicates higher deposit).

However, due to some constraints, only diesel

fuel was tested in this study. Biodiesel fuel was

not evaluated for its tendency to form deposit in

the condition of this test

0

10

20

30

40

50

60

70

80

0

2

4

6

8

10

12

unaged 910bar 95C

910bar 82C

700bar 83C

500bar 77C

Inle

t Te

mp

era

ture

He

ate

r tu

be

ra

tin

g

JPI rating inlet temp, °C

A

B

C

D

E

JFTOT temperature :200C

Figure 5: The effects of pressure and

temperatures

0

2

4

6

8

10

12

210 220 230 240 250 260

He

ate

r tu

be

rat

ing

JFTOT temperatures, degC

fresh fuel

30 minutes

4 hours

pressure=900bar

To= 80°C

Figure 6: The effects of pre-stressing

duration

3.2 Effects of exposure to high pressure on

deposit on hot surface

Diesel and biodiesel fuel were tested on the hot

surface tests after exposure to 10 minutes of high

pressure common rail system. The hot surface

temperature was set at 200°C, 225°C and 250°C.

As expected, at higher surface temperature, more

deposit is formed. The amount of fuel dropped (as

given by the longer test duration) also increases

the deposit formation. Figure 7 shows the

photograph of typical deposit of biodiesel formed

at different surface temperature and duration.

Figure 8 to 10 shows the deposit for diesel and

biodiesel for hot surface temperature of 200°C,

225°C and 250°C respectively. The pre-stressing

pressure in the common rail was done at pressures

of 200 bar, 400 bar, 600 bar and 800 bar.

It was shown that for diesel, the pre-stressed fuel

has significantly more deposit than the fresh fuel.

However, increasing the pre-stressing pressure

from 200-800 bar in the common rail system did


2012

175

not increase the deposit formation further. This is

true for all hot surface temperatures tested.

In contrast, for biodiesel, there is some difference

in the quantity and the shape of the deposit

formed as the pre-stressing pressure is increased.

It seems that as the fuel is pre-stressed at higher

pressure, the deposit formation concentrated more

on the fringes. This is probably due to changes in

the fuel‘s surface tension when it has undergone

high pressure and high shear operation in the

common rail.

Note also that at surface temperature of 200C, the

deposit formation was shown to increase with

higher pre-stressing pressure. This is, however,

not evident at other surface temperatures.

4. CONCLUSIONS

From this study, the following conclusions can be

drawn.

1. The deposit forming tendency of diesel and

biodiesel when exposed to high pressure and

temperature operation of a common rail fuel

system was studied.

2. It was demonstrated that fuel exposed to

high pressure and temperatures seen in

common rail system significantly increased

the deposit forming tendency even after a

short duration.

3. Diesel fuel was shown to increase its deposit

formation in JFTOT tests with higher pre-

stressing pressures. However, it did not

show the same trend in hot surface

deposition test. On the other hand, biodiesel

showed increased tendency to form deposit

in hot surface deposition test with increasing

pre-stressing pressure.

ACKNOWLEDGEMENT

The authors are grateful to the Research

University Grant, Universiti Teknologi Malaysia,

Vot 01H03 for the financial support and Research

Management Centre, UTM for the management

support.

REFERENCES

[1] Ullmann, J. Geduldig M., Stutzenberger H.,

Caprotti R., Balfour G., (2008), Investigation

into the Formation and Prevention of Internal

Diesel Injector Deposits, SAE 2008-01-0926

[2] Steve Cook and Paul Richards (2009),

Possible Influence of High Injection Pressure

on Diesel Fuel Stability: A Review and

Preliminary Study, SAE2009-01-1878

[3] Leedham A, Caprotti R, Graupner O, Klaua

T, (2004), Impact of Fuel Additives on Diesel

Injector Deposits, SAE2004-01-2935

[4] Rod Williams (2002), Development of a

Nozzle Fouling Test for Additive Rating in

Heavy Duty DI Diesel Engines, SAE 2002-

01-2721


2012

176

[5] Graupner O, Klaua T, Caprotti R, Breakspear

A, Schik A, Rouff C, (2005), Injector

Deposit Test For Modern Diesel Engines,

www.infineum.com/Documents/.../TAE/Esslin

gen%202005.pdf , accessed on 19th

October

2010.

[6] Chintoo Sudhiesh Kumar (2009),Modelling

Deposit Formation in Diesel Injector Nozzle,

MSc. Thesis, Massachusets Institute of

Technology, June 2009

[7] Stavinoha LL, Barbie JG, Yost DM, (1986),

Thermal Oxidation Stability of Diesel Fuels,

Interim Report BFLRF No 25,. 1986,

Southwest Research Institute

[8] Yusmadi Mohd Arifin (2009), Diesel and

Biodiesel Fuel Deposit on a Hot Wall

Surface, PhD Thesis, Gunma University,

Japan, August 2009

[9] Muhamad Adlan Abdullah and Farid Nasir

Ani, (2012), The Effects of Diesel Fuel

Exposure to High Pressure Common Rail

System on its Deposit Forming Tendency,

SEATUC Conference, Bangkok, March 2012

.


2012

177

Time

(min)

200 ◦C 225 ◦C 250 ◦C

0

5

10

15

20

25

30

Figure 7: The biodiesel deposit formation on the hot surface at different temperature and duration

(fuel without pre-stressing)


2012

178

Figure 8: The effect of pre-stressing on deposit formation for diesel and biodiesel at surface temperature of

200°C


225°C


250°C


2012

179

Numerical Analysis of Elastohydrodynamic Lubrication with

Non-Newtonian Lubricant

Dedi Rosa Putra Cupu1, Adli Bahari

2, Kahar Osman

3, Jamaluddin Md Sheriff

3

1Mechanical Engineering Department of Engineering Faculty,

University of Riau, Pekanbaru, Riau, Indonesia

Email1 : [email protected]

2Automation & Mechatronics Section, Industrial Electronics Department,

German Malaysian Institute, Malaysia

3Faculty of Mechanical Engineering,

Universiti Teknologi Malaysia, 81310 UTM Skudai, Malaysia

ABSTRACT

Elastohydrodynamic lubrication is a form of

hydrodynamic lubrication involving physical

interaction between two contacting surfaces and

liquid where elastic deformation of the contacting

surfaces due to heavily loading applied will affect the

elastohydrodynamic pressure and fluid film thickness

significantly. In this paper, a line contact EHL is

modeled through the cylinder contact to a flat surface

to represent the application of roller bearing. This

solution is limited to two dimensional line contact

problem only, an infinite length of cylinder will be

used as physical modeling. The behavior of non-

Newtonian fluid also was investigated using power

law fluid model. Bearing speed is to be assumed in

steady state and temperature is assumed constant. The

bearing performance parameters such as pressure,

film thickness and friction coefficient of lubricated

contacts are calculated using Newton-Raphson

method.

The results show that the peak pressure increases as

the parameters such as velocity, load, material

parameter and power law index were increases and

the spike was found to shift to the center of roller.

The film was almost flat at contact region and formed

a dimple shape near the outlet flow. The coefficient

of friction is reduced as the power law index and

slide to roll ratio were decreased. The value of

pressure spike and minimum film thickness were

smaller at lower speed and were increased during

raising speed then the peak point was found to be

shifted to center of roller.

Keywords : Elastohydrodynamic lubrication,

Newton-Raphson, Pressure profile,

film thickness.

1. INTRODUCTION

The whole idea of this study is to write a

programming code that can provide a solution for

fluid film lubrication problem related to

elastohydrodynamic lubrication. This can further be

applied to investigate the effect of parameters on

bearing design and performance such as pressure and

film thickness.

Detailed analysis of gaseous or liquid films is usually

termed hydrodynamics lubrication (HD), while

lubrication by solid is termed solid lubrication. A

specialized form of hydrodynamics lubrication

involving physical interaction between the contacting

bodies and the liquid lubricant is termed

elastohydrodynamics (EHD) lubrication (EHL) and is

considerable practical significance. Another form of

lubrication involves the chemical interactions

between contacting bodies and the liquid lubricant is

termed boundary and extreme pressure lubrication. A

form of lubrication that operates involving the

external force is termed hydrostatic lubrication where

liquid or gaseous lubricant is forced into the space

between contacting bodies.

The most commonly encountered forms of contacts,

commonly known as conjunction, are point and line

contacts. When a sphere comes into contact with a

flat surface, it initially forms a point contact with a

circular shape and the size of conjunction grows as a

function of load. When a cylinder comes into contact

with a flat surface, it forms a line contact and it

grows into a rectangular conjunction as the load is

increased. Incidentally point contact between a ball

and raceway develops into an elliptical conjunction.

In elastohydrodynamic lubrication contact, the

deformed surfaces in lubricated contact are almost

similar to Hertzian contacts with an interposed

lubrication film. A minimum film thickness occurs

near to the outlet region. The film thickness is

IMAT-UI 030



2012

180

important because it is the key parameter to ensure

the protection of mating surfaces for bearing

components.

Dien and Elrod [1] derived the generalized steady

state Reynolds Equation for non-Newtonian fluids

using power law fluid with application to journal

bearing. A. Elsharkawy [2] then applied the formula

into magnetic head-rigid disk interface

hydrodynamically lubricated. The result was shown

that the power law exponent has a significant effect

on the hydrodynamic pressure profile. An early study

of soft elastohydrodynamic in a rolling contact for a

power law fluid by Lim et al. [3] was used in the

exploration of a printing application in which near

pure rolling takes place. The importance of both

power law coefficient and exponent was quantified.

This showed significant impacts for both parameters,

with increases in each resulting in increased film

thickness and maximum pressure. Bohan et al. [4],

explore the application of numerical simulation to

coating applications that involve combined sliding

and rolling mechanisms. This was extend previous

work done by Carvalho and Scriven, [5] and Lim et

al. [6], through the incorporation of actual fluid

properties that exhibit shear thinning. The effect of

power law coefficient, power law exponent and

sliding on the nip performance in terms of pressure

distribution, film thickness profile, strain rate and

viscosity variation through the nip section and flow

rate was investigated.

H.M Chu et al. [7] derived a one-dimensional

modified Reynolds equation for power law fluid from

the viscous adsorption theory for thin film

elastohydrodynamics lubrication (TFEHL). The

lubricating film between solid surfaces was modeled

as three fixed layers, which are two absorption layers

on each surface and a middle layer between them.

The comparison between classical non-Newtonian

EHL and non-Newtonian TFEHL was done. The

result was showing that the TFEHL model can

reasonably calculate the pressure distribution, the

film thickness, the velocity distribution and the

average viscosity. Another result showing that the

greater the thickness and viscosity of the adsorption

layer and the flow index, the greater the deviation in

central film thickness versus speed between EHL

model and TFEHL model produced in the very thin

film regime.

A thermal and non-Newtonian fluid model under

thermal elastohydrodynamics conditions was

proposed by A. Campos et al. [8]. The concept of

apparent viscosity was used to introduce the non-

Newtonian behavior of the lubricant and the thermal

behavior of the contact. The Newton-Raphson

technique was used to obtain the lubricant film

geometry and the pressure distribution inside the

elastohydrodynamics contact. The model was applied

to the analysis of experimental traction curves of a

traction fluid measured in a twin disc machine,

obtained for a significant ranges of the operating

condition. The comparison between numerical and

experimental traction curves showed a very good

correlation.

In 1986, Houpert and Hamrock [9] presented a fast

method to solve the problem facing by Hamrock and

Jacobson (1983). The Reynolds equation is solved

using Newton Raphson iterative procedure and can

solve problem related to higher dimensionless load in

shorter computing time. However this method is not

suitable to be used for point contact problem (three

dimensional) due to memory storage problem. Lin

and Lin in 1990, using the method proposed by

Houpert and Hamrock (1986), derived the Reynolds

equation embedded with power law fluid. Assuming

the flow were compressible and viscous, Lin and Lin

plotted the graph showing the effects of power law

index on pressure profile, film thickness and

coefficient of friction. However their work was only

focusing on Barus equation to represent the pressure-

viscosity model.

2. PHYSICAL PROBLEM DEFINE

In this work, a roller bearing is adapted to the

physical problem for the numerical calculation to be

done. A roller bearing is a device used to support a

rotating shaft to the bearing housing and at the same

time used to reduce the friction between contacting

surfaces. Figure 1 shows a typical roller bearing

commonly used in industrial machine. Since the

scope of this study is limited to two dimensional line

contact problem only, an infinite length of cylinder is

model to be contacted with a flat surface in order to

represent the roller bearing application.

Figure 1: A typical roller bearing

An infinite length of steel roller with diameter, R =

11.4 mm is used for analysis. The cylinder is loaded

against the flat surface. The roller and the flat surface

are then been driven independently to create a mixed


2012

181

rolling/ sliding contact. The temperature is assumed

to be constant. Figure 2 shows some physical

parameters that are applied to the system of roller

bearing. A load, w is applied to the cylinder, which

both the roller and the flat surface are driven with the

different speed, ua and ub. Some of the parameters are

listed in the Table 1 that has been used in the

programming code throughout the study.

Figure 2: An infinite length of cylinder contact

with flat surface

Table 1: Some of the physical parameters involve in

the calculation.

Material Steel: AISI 52100

Elastic Modulus Eb = Ed = 210 GPa

Poisson‘s ratio b = d = 0.3

Ball radius R = 11.4 mm

Dimensionless Applied Load W = 2.0452x10-5

Lubricant Absolute Viscosity 0 = 6.6 Pa.s

Dimensionless Speed U = 1x10-11

Temperature T = 35C (constant)

3. NUMERICAL SOLUTION

The fluid film between two solid surfaces shown in

Figure 3 is considered. Reynolds equation is an

equation to obtain the pressure generated in a fluids

film when two such surfaces undergo relative motion.

However, the fluid film must be sufficiently thin so

that Reynolds‘ assumptions described below will

hold. For simplicity, the lower surface is assumed to

be a plane. The velocity of the fluid in the directions

x, y, and z are denoted by u, v, and w, respectively,

and the velocity of the lower surface is similarly

described by u1, v1, and w1 and that the upper surface

by u2, v2, and w2. In many practical cases, the lower

surface and the upper surface perform a straight

translational motion relative to each other. In this

case, if the x axis is in the translational direction, then

we have w1 = w2 = 0 and so the equations can be

simplified. The gap between the two surfaces of the

liquids film, be donated by h(x,z,t), with t being time.

The coefficient of viscosity of the fluid is donated

by. In deriving Reynolds‘ equation with Newtonian

fluid, the following assumptions are made as follows

[5]:

(i) The flow is laminar.

(ii) The gravity and inertia forces acting on the

fluid can be ignored compared with the

viscous force.

(iii) Compressibility of the fluid is negligible.

(iv) The fluid is Newtonian and the coefficient

of viscosity is constant.

(v) Fluids pressure does not change across the

film thickness.

(vi) The rate of change of the velocity u and w in

the direction and z direction is negligible

compared with the rate of change in the y

direction.

(vii) There is no slip between the fluid and the

solid surface.

Figure 3: Fluid film between two solid surfaces.

Figure 4: A small element of fluid.

The balance of forces acting on a small volume

element in the fluid is considered as shown in Figure

4.

Neglecting the gravity and inertia forces, we obtain

the following equation:


2012

182

0

dxdydxdz

dxdydxdydzz

dxdzy

dydzdxx

zxyx

x

zx

zx

yx

yx

x

x

(1)

Where x is the normal stress acting on the plane

normal to the x axis and yx and zx are the shear

stress acting on the plane normal to the y axis and z

axis, respectively, in the direction of the x axis.

Equation (1) can be rearranged as follows:

0zyx

zxyxx

(2)

Let the fluid pressure be p, laminar flow of

Newtonian fluid is considered here, and the ―side

leakage pressure‖ zp is neglected, so equation

(2) can be written as follows:

y

u

yx

p

(3)

On the assumption that the rate of change of the flow

velocity u in the z direction is sufficiently small

compared with that in the y direction, the second term

of the right-hand side of the above equation can be

disregarded compared with the first term and

assumption that is constant, the equation of the

balance of forces in the x direction is finally obtained

as follows:

2

2

y

u

x

p

(4)

Integrating equation (4) twice gives the flow velocity

u and w respectively. The boundary conditions for

the velocities are (from the Reynolds assumption: no

slip condition) as follow:

0yatww,uu 11

Then the fluid velocities will be as follows:

211

2

1u

h

yu

h

yyhy

x

pu

(5)

The continuity equation for a small volume element

in an incompressible fluid can be written as follow:

0z

w

y

v

x

u

(6)

Equation (6) can be written as follows in terms of the

surface velocities from the boundary conditions:

0122

200

vv

z

hw

x

huwdy

zudy

x

hh

(7)

21

3h

0uu

2

h

x

p

12

hudy

(8)

21

3h

0ww

2

h

z

p

12

hwdy

(9)

Where and p are assumed to be constant in the y

direction. Substituting these integral (8) and (9) into

Equation (7) gives the following equation:

2121

21

2121

33

2

6

vvwwx

hh

x

hww

uux

hh

x

huu

z

ph

zx

ph

x

(10)

In many practical cases, the x axis can be taken as the

direction of the relative motion of the two surfaces,

so 0wwv 211 . And if the motion of the body

is only in one-dimensional, 0v2 . Assuming that

the viscosity, and the velocity, u1 and u2 are

constant, thus Equation (10) can be reduced to:

x

huu

z

ph

zx

ph

x

21

33

6

(11)

For two-dimensional flow, the ―side leakage

pressure‖ is neglected. Thus Equation (11) finally can

be written as:

x

hu12

x

ph

x

3

(12)

where the entrainment speed, 2

uuu 21


2012

183

Using these dimensionless

parameters,

00

0

2

;;p

pP;

b

xX

;'EG;R'E

uU;

b

hRH;

R'E

wW

H

Reynolds equation (12) can be written as:

X

H

W

U

dX

dPH

X

2

23

4

3

(13)

4. SOLUTION OF REYNOLDS EQUATION

Newton-Raphson formula was used for root finding

of non-linear equation of function, f(x). To find a root

where f(x) = 0, let xi is initial guess for of root.

1ii

i

xx

0xftanx'f

(14)

This equation could be simplified to:

old

i

new

i

old

i x'fxxf

(15)

The Reynolds equation (13) could be written as:

i

eeii

3

ii

HHK

dX

dPHf

(16)

The boundary condition is:

0X

PP,XXat

0P,XXat

out

in

Reynolds Equation (16) is solved by Newton-

Raphson iterative method as describe by Houpert and

Hamrock [21]. The unknowns eeH , jP and 0H are

achieved by two successive iterations represent the

new and old value after iteration.

new

ee

old

ee

new

ee HHH (17)

new

j

old

j

new

j PPP

(18)

new

0

old

0

new

0 HHH

(19)

From the Newton-Raphson definition in Equation

(16), one can write:

old

i

N

jnew

old

j

inew

j

old

ee

inew

ee

old

i

H

f

H

P

fP

H

fHf

02

0

(20)

Equation (20) can be treated as solution of

simultaneous linear equation. Since the number of

equations and the number of unknowns must be

equal, therefore N+1 number of equations is required

to solve the unknowns N+1. An additional load

condition is required, such as:

N

2j

newnew

jj WPC

(21)

A convergence criterion to stop the iteration of

Equation (17), (18) and (19) is chosen to be:

0.00001<P

PPnew

j

old

j

new

j

(22)

A linear system of N+1 equation is therefore to be

solved as below matrix:

old

N

new

N

ee

old

N

N

N

NN

ee

N

Nee

Nee

W

f

f

f

H

P

P

H

CC

H

f

P

f

P

f

H

f

H

f

P

f

P

f

H

f

H

f

P

f

P

f

H

f

2

1

0

2

2

02

0

22

2

22

0

11

2

11

00

(23)

In this simulation, the film thickness is calculated

based on fast approach of published paper by

Houpert and Hamrock [9], as this following formula:

i

2

i0i

2

XHH

(24)


2012

184

The elastic deformation of contacted surface is:

WRln

'dX'XXln'XX'dX

dPend

in

X

Xi

8

4

1

22

1

2

2

(25)

Applied load in dimensionless form can be calculated

from:

2PdXout

in

X

X

(26)

Viscosity of lubricant is able to be calculated from

viscosity-pressure relationship. Barus and Roelands

have proposed their equation to obtain the viscosity.

In this simulation, Barus expression will be used to

calculate the viscosity. p

0e (27)

And the density distribution of lubricant is obtained

from Dowson and Higginson.

iH

9

iH

9

iPp10x7.11

Pp10x6.01

(28)


Few graphs have been plotted to show the result from

the programming code in Figure 5 to 9. Mineral Oil

of PAO 800 was used for simulation with the

operational viscosity was 6.60 Pa.s and pressure-

viscosity coefficient was810x276.2 m

2/N. The

dimensionless parameters were used in this

simulation, U = 1x10-11

, W = 2.0452x10-5

, and G =

4500.

Figure 5 shows the plot of EHD pressure using 321

nodes and Barus pressure-viscosity model. The inlet

boundary was set at X = 4. The outlet boundary

was calculated at X = 1.17051. Initial condition of

pressure used Hertzian pressure. Using personal

computer, it took about 20 seconds to converge after

19 iterations with absolute relative error, = 0.00001.

The EHD pressure was increasing from the inlet flow

until reach a higher value at center. The pressure was

then drops a bit and increased until reach to the peak

value. The tremendous dropping of pressure is found

after the peak point. The ratio of peak pressure to

maximum Hertzian pressure was 1.377.

Figure 6 shows the relevant elastohydrodynamic film

thickness. The initial guess for dimensionless central

film thickness was chosen to be H0 = 0.6. The shape

is almost flat at contact region and form a dimple

shape near the outlet flow.

The effects of velocity on Elastohydrodynamic

lubrication were shown in Figure 7 and Figure 8.

Figure 7 shows the effect of velocity on EHD

Pressure. At lower the speed, the pressure spike value

is smaller. As the speed is increased the value of

spike also increased and the peak point is found to be

shifted to center of roller. From figure 8, it can be

seen that the effect of velocity on film thickness. At

lower speed, the value of minimum film thickness is

smaller. As the speed is increased the minimum film

thickness also increased and the film dimple shifted

to the center of roller.

Figure 9 and Figure 10 show the effect of load on

EHD Pressure. At higher load (W = 3x10-5

) the value

of pressure spike is smaller and nearly same with the

value of pressure at center of roller, and as the load is

decreased (W = 0.8x10-5

) the value of spike is

increased and the peak point is found to be shifted

near to roller center (Figure 9). Figure 10 shows the

effect of load on film thickness. At lower load (W =

0.8x10-5

) the value of minimum film thickness is

bigger. As the load is increased the value of

minimum film thickness became decreased and the

film shape is more flat along contact region. The film

dimple is found to be shifted from roller center to

outlet region.

Figure 5: EHD Pressure and Hertzian Pressure


2012

185

Figure 6: Elastohydrodynamic Film Thickness

Figure 7: Effect of Velocity on EHD Pressure

Figure 8: Effect of velocity on film thickness

Figure 9: Effect of load on EHD Pressure

Figure 10: Effect of load on film thickness

6. CONCLUSION

The simulation shows that the peak pressure

increases with the parameter such as velocity, and

applied load. In case of velocity, the pressure

increases and approaching its load bearing capacity.

The film thickness is also increases as the velocity

increases and the film dimple approaching to the

center of roller.

REFERENCES

[1] K Dien and H.G Elrod, A Generalized Steady-

State Reynolds Equation for Non-Newtonian

Fluids, With Application to Journal Bearings,

Trans. ASME Journal of Lubrication

Technology vol.105, page 385-390, 1983.

[2] Abdallah A. Elsharkawy, Magnetic head-rigid

disk interface hydrodynamically lubricated with


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a power-law fluid, Journal of Wear vol.213 page

47-53, 1997.

[3] Lim, C.H., Bohan M.F.J., Claypole, T.C.,

Gethin, D.T. and Roylance, B.J., A finite

element investigation into a soft rolling contact

supplied by a non-newtonian ink, Journal of

Appl. Phsy, vol.29, page 1894-1903, 1996.

[4] M.F.J. Bohan, I.J Fox, T.C Claypole and D.T

Gethin, Numerical Modelling of

Elastohydrodynamic Lubrication in Soft

Contacts using non-Newtonian Fluids.

International Journal of Numerical Methods for

Heat & Fluid Flow, vol. 12, no. 4, 2003.

[5] Carvalho, M.S. and Scriven, L.E., Deformable

roller coating flows: steady state and linear

pertubation analysis. Journal of Fluid

Mechanics, vol.339, pp. 143-172, 1997.

[6] Lim, C.H., Bohan M.F.J., Claypole, T.C.,

Gethin, D.T. and Roylance, B.J., A finite

element investigation into a soft rolling contact

supplied by a non-newtonian ink, Journal of

Appl. Phsy, vol.29, page 1894-1903, 1996.

[7] H.M. Chu, W.L. Li and Y.P. Cheng. Thin film

elastohydrodynamics – a power law fluid

model. Tribology International, vol. 39, page

1474-1481, 2006.

[8] A.Campos, A.Sottomayor and J.Seabra, Non-

Newtonian and Thermal Elastohydrodynamics.

Journal of Mechanica Experimental, vol. 13,

page 81-93, 2006.

[9] Houpert, L.G. and Hamrock, B.J., Fast approach

for calculating thickness and pressures in

elastohydrodynamically lubricated at high loads.

ASME Journal of Tribology, vol.108, page 411-

420, 1986.


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Latest System Simulation Models in Field of Heating, Refrigeration,

and Air-conditioning, and Development of System Simulator

KiyoshiSa, JongsooJ

b

a,bSchool of Fundamental Science and Engineering

Waseda University, 3-4-1, Okubo, Sinjuku, Tokyo 169-8555, Japan

Tel : 81 (3)52863259. Fax : 81 (3)52863259 aE-mail : [email protected]

bE-mail : [email protected]

ABSTRACT

The energy consumption of heating, refrigeration,

and air-conditioning systems is steadily increasing. It

is, however, not easy to reduce the energy

consumption of such thermal systems because they

have already been greatly improved to save energy.

To meet the demands of the global energy saving

policy, we need to determine the best combination

and total energy management scheme for heating,

refrigeration, and air-conditioning systems.

Simulation is a promising technology for such

investigations because it is not feasible to carry out

experiments with large-scale energy systems. High-

precision simulation models are used for these

investigations, and we are developing such models of

the heat pump, room air-conditioner, variable

refrigerant flow (VRF) system, desiccant

dehumidifier, indirect evaporative cooler, fuel cell,

solar panel, solar collector, etc This paper introduces

highly accurate models of a VRF system and an

absorption heat transformer. The simulator that we

are presently developing is also introduced. Named

‗Energy Flow +M‘, the simulator is very easy to

handle because of its user-friendly graphical user

interface (GUI). It has already been unveiled to the

world through the Internet and is expected to be used

for energy saving in heating, refrigeration, and air-

conditioning systems.

Keywords: Total energy management, Energy Flow

+M, VRF system, Heat transformer,

energy saving

Nomenclature A area m2

COP coefficient of performance -

D mass diffusivity m2s-1

d diameter m

G mass flow rate kgs-1

g gravitational acceleration ms-2

gm mass flow rate per unit length kgm-1s-1

h specific enthalpy Jkg-1

j mass flux kgm-2s1-

K overall heat transfer coefficient kWm-2K-1

L tube length m

l distance between droplets m

p pressure Pa

Q heat transfer rate kW

q heat flux kWm-2

r radius m

T temperature K

t time s

u Specific internal energy kJ kg-1

v velocity ms-1

W compressor power kW

X concentration -

x refrigerant flow direction axis m

Greek symbols

α heat transfer coefficient kWm-2K-1

β mass transfer coefficient ms-1

δ liquid film thickness m

Г mass flow rate per unit length kgm-1s-1

θ angle rad

λ thermal conductivity kWm-1K-1

μ viscosity Pas

ρ density kgm-3

Subscripts

A air

AH high temperature absorber

b bulk

COM compressor

c concentration boundary

EL low temperature evaporator

EVA evaporator

f film

GL low temperature generator

I inlet

in interface, inside

j perpendicular direction axis

O outlet

out outside

R refrigerant

S solution

sh superheat

V refrigerant vapour

W water

1. INTRODUCTION

Recently, the governments of many countries

required industries to further reduce their energy

IMAT-UI 031



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188

consumption. In the field of heating, refrigeration and

air-conditioning, the amount of energy consumed is

very high that further reduction is imperative. Hence,

investigations of optimum system combinations and

their total energy management schemes are

important. However, the electrical, air-conditioning,

and water heating loads involved in this field change

greatly as a result of environmental conditions and

human lifestyles. This makes it more difficult to use

only experiments to optimise system compositions

and operation methods. Simulation has thus been a

useful and powerful tool in investigating the

optimization and energy management of combined

systems. We have discussed the efficient and detailed

simulation of a complicated system [1] and also

developed high-precision simulation models of a

compression-type heat pump, dehumidification

system, solar panel, fuel cell, indirect evaporator,

thermal transportation system, etc. As examples, this

paper introduces our latest simulation models of a

multi-type compression air conditioner (variable

refrigerant flow, or VRF system), a multi-stage

absorption heat transformer, and a general-purpose

energy system simulator that can calculate the

characteristics of each of these systems. The

simulator, called ‗Energy flow +M‘, has been

unveiled to the public through the Internet.

2. SIMULATION OF UNSTEADY STATE

OF VRF SYSTEM

A VRF system is an air conditioning system that has

many indoor units connected to a single outdoor unit.

Examples are multi-type air conditioners used in

buildings. Compared to the single-type air

conditioner, the system performs better and saves

space and energy, even though there are several

indoor units. But because there are several indoor

units each placed in different rooms, driving

conditions for each unit differs by its usage

conditions. Thus, when one indoor unit is operated at

maximum load and another at a light load, or

switched off, an imbalance occurs in the system.

Therefore, to improve the performance and efficiency

of the system without reducing its reliability and

usability, accurate simulation is vital. This simulation

must include the various operation conditions of the

indoor units and the unsteady state of the system

during mode change. In this research, we developed a

simulation model of a system with four indoor units

and validated it by running performance tests.

2.5 System Description

The refrigerator used in our research was a

compression-type VRF system composed of one

outdoor unit and four indoor units, as shown in Fig.

1. The outdoor unit incorporated two compressors

and a heat exchanger for subcooling. The indoor

units were cassette-type, set in the ceiling. The

cooling performance of each indoor unit was

controlled by expansion valves, which were set in the

indoor units. The rated cooling capacity was 28.0

kW, and the rated input power of the compressor was

7.19 kW. The refrigerant was R410A. Fig. 1 shows

the system flow.

2.6 Mathematical Model

Fig. 2 shows the heat exchanger model adopted in the

evaporator and condenser. The heat exchanger model

equations are given below. The refrigerant-side

continuity, energy, and pressure drop were calculated

using the following equations:

R RR

GA

t x

(2.1)

_R R R R

R M in

u G hA d q

t x

(2.2)

Figure 1: Schematic flow of VRF system

Expansion valve

Compressor

Accumulator

Condenser

Evaporator

Expansion

valveHeat exchanger Expansion

valve

(a) Schematic flow of VRF system

(b) Outdoor unit

(c)

Indoor unit


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189

0RP

x

(2.3)

The air-side energy equation is as follows:

_ _M M

M M out M M in

uA d q d q

t

(2.4)

The tube-side energy equation is as below:

_m O I M out Mg h h d q (2.5)

The following equations were also used to calculate

the heat transfer rates:

M Rq T T (2.6)

M M A Mq T T (2.7)

/EVA COMCOP Q W (2.8)

2.7 Simulation and Experimental Results

The outdoor and the indoor temperatures of the

simulation and experiment were 30 °C and 27 °C,

respectively. In the PH diagram of Fig. 3, the

operation processes of the experiment and the

simulation at the rated capacity are indicated by the

dots and the line, respectively. The two results are in

good agreement. As shown in Fig. 4, we considered

the unsteady condition of the system when the

number of operating indoor units changed from one

to four. We also considered the reverse situation

when the number of operating units changed from

four to one. Further details about the latter are

available in [2]. In the former case, the operation of

the system changed when the cooling capacities and

compressor input power were in good agreement at

the rated capacity, as shown in Fig. 4. Despite the

fact that the heat transfer coefficient was considered

constant and the pressure drop was not considered,

we concluded that it was possible to accurately

predict the performances. A simulation model that

takes the detailed heat transfer into consideration can

also be developed to investigate the characteristics of

the system. However, to ensure an easy evaluation of

the very complicated unsteady state of the VRF

system, a simple simulation model with a good

accuracy is required. This is because it takes a very

long time to obtain results with complicated

simulation models. We are searching for a simulation

method that will optimally combine accurate

simulation results with calculation speed.

200 250 300 350 400 450 5000.4

1

2

4

10

Pre

ssu

re

MP

a

Enthalpy kJ/kg

T =

20

oC 40

60

80

100

120 140

s =

1.0

kJ/

kg

K

1.1

1.2

1.3

1.4

1.5

1.6

1.7

1.8

1.9

2.0

2.1

Figure 3: PH diagram of experiment and

simulation

_

_

_

R I

R I

R I

G

P

h

_

_

_

R O

R O

R O

G

P

h

_ _ _ _A I A I A I A IG P h X

_ _ _ _A O A O A O A OG P h X

x

Air inlet

Air outlet

Refrigerant

outlet

Refrigerant

inlet

(b) Control volume

Figure 2: Heat exchanger model

(a) Heat exchanger of evaporator and condenser


2012

190

3. CHARACTERISTICS OF DOUBLE-

STAGE ABSORPTION HEAT

TRANSFORMER

A considerable amount of steam is used in industry,

most of which is produced by boilers that burn fossil

fuels. To respond to increasing energy saving

demands, a more efficient method of steam

production is vital. There is the need for heat pump

technology to develop a waste heat–driven steam

generator. However, it is difficult for compression-

type heat pumps to generate steam of temperatures

above 120 °C because of the problem it poses to the

stability of the refrigerant and lubricant. In recent

years, there has been more focus on absorption heat

transformers because of their ability to produce high-

temperature steam from virtually only low-grade

waste heat.

20

40

60

80

100

120

Rota

tional

spee

dn

CO

M01 r

ps

0

200

400

600

Val

ve1

open

ing

puls

e

0

200

400

600

Val

ve2

open

ing

puls

e

0

200

400

600

Val

ve3

open

ing

puls

e

0

200

400

600

Val

ve4

open

ing

puls

e

0

5

10

Cooling

capac

ity

QE

VA

1 k

W

0

5

10

Cooling

capac

ity

QE

VA

2 k

W

0

5

10

Cooling

capac

ity

QE

VA

3 k

W

0

5

10

Cooling

capac

ity

QE

VA

4 k

W

0

2

4

6

8

Com

pre

ssor

input

kW

0

10

20

30

Cooling

capac

ity

kW

-200 0 200 400 600 800 10000

5

10

CO

P

Time t s

2000

2500

3000

Pre

ssure

PC

OM

O k

Pa

-200 0 200 400 600 800 1000500

1000

1500

Time t s

Pre

ssure

PC

OM

I k

Pa

Figure 4: Unsteady Simulation Results; operated

units changed from one to four

Absorber

GeneratorEvaporator

Condenser

Steam

Hot

Water

80-90oC

Cooling

Tower

40oC

Tem

pera

ture

Refrigerant

vapor

Refrigerant

lift

temperature

Refrigerant

vapor

Strong

Solution

Solution Heat

Exchanger

Weak

Solution

Feed

Water

temperature

difference

120oC

Figure 5: Concept of absorption heat transformer

3.1 System Description

Fig. 5 shows the concept of a basic absorption heat

transformer. The system comprises five main

components: absorber, generator, evaporator,

condenser, and solution heat exchanger. The lithium

bromide–water pair is used as the working fluid.

Lithium bromide brine is condensed in the generator

by using the temperature difference between the

generator and the condenser. High-temperature steam

is produced by elevating the boiling point of lithium

bromide brine while it is in the absorber. Figs. 6(a)

and 6(b) show the commercialised system, which can

produce 180 °C steam from 80–90 °C hot water. The

system consists of a generator, condenser, evaporator,

absorption evaporator, refrigerant separator, high-

temperature absorber, and steam separator. The

Duhring diagram of the commercialised system is

shown in Fig. 6(c). This commercialised system is

driven by a series-flow double-lift cycle, a test type

of which we fabricated. In this paper, we discuss the

characteristics of the objective heat transformer. A

comparison of the test data and the results of the

analysis is also used to validate the simulation model.

3.2 Mathematical Model

The mathematical model of each component mainly

consisted of equations of continuity, pressure drop,

and energy. We will explain the model of the

absorber as an example. Fig. 7 shows the absorber

model, which is separated into a falling liquid film

and inside tube flow model, liquid drop formation

model, distributoer model, refrigerant vapour flow

model, bifurcation model, mixture model, and pool

model. The falling liquid film and droplet formation

models are representatively shown in this paper.


2012

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3.2.1 Falling film regime

Fig. 8 shows the falling film regime model. The

mathematical models formulated by Jeong and

Garimella [3] are as follows:

1 S S

f

v j

r

(3.1)

10S S Sv X

r

(3.2)

1 jS S S

f

jh qv h

r

(3.3)

0

, 0

V

j

sh in in

h jh

f X T j

(3.4)

The Nusselt liquid film theory respectively gives the

liquid film thickness and velocity as follows: 1 3

2

3

sin

S S

fS

S g

(3.5)

2

sin2

SS f

S

g xv x

(3.6)

The parameters of the heat transfer are given by the

following equations:

S Wq K T T (3.7)

1 1 1 1ln out

in in W pipe in o S

rK

r r r r

(3.8)

0S inT T

(3.9)

,S T S ST f X h (3.10)

,W T W WT f P h (3.11)

Assuming that the temperature distribution in the

direction of the liquid film thickness is linear, the heat

transfer coefficient is given by the following equation:

8

5

S

S

f

(3.12)

AAHH

AALL

EELL

GGLL

CC

EEHH SS

1100

2000

(a) Objective heat transformer system

(unit: mm)

Condenser

Evaporator

Refrigerant Separator HT Absorber

Generator

Absorbing Evaporator

Steam Separator

C

EL

EH

GL

AL

AH

S

Cooling

Water

Feed Water

Hot Water

Hot Water

Steam

SPRP

SV

T

P

G

Gρ

T

T

T

T

T

T

T

Gρ

T

T

T

T G

TTGρ

P

P

GT

T

T

T

G

T

T

G

GT

TT

T

T

G

P

(b) Schematic flow of system

Q Q

QQ

QQ

C

EL

EH

AL

AH

GL

Hot Water SteamCooling Water

Saturatedtemperature

Solution temperature

Intermediate steam

(c) Duhring Diagram

Figure 6: Objective absorption heat transformer

Figure 7: Absorber model


2012

192

The mass transfer flow rate is also calculated from

the following equation:

( )film S b inj C X X (3.13)

, = 2.0film

c

mDm

(3.14)

3,

2

b in c

S b

f

X XX X

m

(3.15)

0b SX X

(3.16)

C is the correction factor for the effect of the

surfactant. The heat transfer coefficient inside the

tube adopts the Dittus-Boelter correlation [4] for the

single-phase flow, and the Thome correlation [5] for

the two-phase flow.

3.2.2 Droplet formation regime

Fig. 9 shows the droplet formation regime

model. The mathematical models formulated by

Jeong and Garimella [3] are as below.

_ _ 0form formS I S O VG G G (3.17)

_ _

0form form

V jS I S OGh Gh G h (3.18)

_ _

0form formS I S O

GX GX (3.19)

_ _ 0form formS I S OP P (3.20)

The mass transfer flow rate is determined with the

following equation:

2

_2 formV S form a S I in

nLG d X X

l

(3.21)

1 2

24

7form

form

D

t

(3.22)

1 2

24

7form

form

D

t

(3.23)

tform and da are the formation time and the diameter of

the droplet, respectively. The details of the droplet

formation regime model are presented in [3]. The

system performance is defined as follows:

/( )AH GL ELCOP Q Q Q (3.24)

Figure 8: Falling liquid film regime model

Figure 9: Droplet formation regime model

0 5 10 15 20 251

510

50100

500

Pre

ssu

re P

kP

a

AH EL C

Solution mass flow rate GS kg/min

0.1

0.2

0.3

0.4

CO

P

0 5 10 15 20 25

56

58

60

62

64

Co

nce

ntr

atio

n X

%

Strong solution Middle solution Weak solution

Solution mass flow rate GS kg/min

0

10

20

30

40

Ste

am G

ener

atio

n R

ate

Gst

eam

kg

/h

(b) The effect of solution mass flow rate

Figure 10: Characteristics of the double-stage

absorption heat transformer


2012

193

3.3 Simulation and Experimental Results

Table 1 shows the simulation and experimental

conditions. The simulation and experimental results

are shown in Figs. 10(a) and 10(b). Figs. 10(a) and

10(b) respectively show the effects of the hot water

inlet temperature and the solution mass flow rate on

the characteristics of the system. As can be seen, the

experimental results agreed with those of the

simulation, thereby validating the simulation models.

We also investigated the characteristics of the

double-stage absorption heat transformer and

confirmed that the system could efficiently generate

high-temperature steam from low-grade heat without

losing stability.

4. DEVELOPMENT OF SIMULATOR

The detailed models of the sample thermal systems

were shown above. The simulation with these models

was not easy owing to the complexity of the models.

We therefore developed a general-purpose analysis

simulator by adding a graphical user interface (GUI).

Fig. 11(a) shows the main screen of the simulator,

called ‗Energy flow +M‘. By pushing the ‗Start

button‘ at the upper right, the pallet is opened on the

display, as shown in Fig. 11(b). The modules are

selected in the pallet using the appropriate icons. One

feature of this simulator is its ability to carry out a

simulation by just connecting modules together. Even

when the flow of some elements changes, the entire

system analysis can easily be carried out again.

Moreover, the energy of a large-scale system and the

simulation of a complicated unsteady state can also

be easily analysed as described in sections 2 and 3.

Furthermore, the calculation results can be obtained

as an Excel document in Fig. 11(c). Fig. 11(d) shows

the simulator that can be also expressed on pallet

with EXCEL interface. This simulator by EXCEL

interface has good features that make simulation to

be easily carried out. As shown in Fig. 11(b) and Fig.

11(d), we could develop two types of simulator by

adopting the modular analysis method into thermal

system. We are also making various calculation

modules including the transient characteristics of

thermal systems for simulator with GUI and EXCEL

interface. The simulator we have developed greatly

reduces the burden on the user in developing

simulation codes, and the simulator by GUI interface

is available on the Internet and can be accessed

anytime.

Figure 12: System flow on EF+M (comparison of

experimental and simulation results)

75 80 85 90 951

510

50100

500

Pre

ssu

re P

kP

a

AH EL C

Hot water inlet temperature THW oC

0.1

0.2

0.3

0.4

CO

P

75 80 85 90 9558

59

60

61

62

63

64

Co

nce

ntr

atio

n X

%

Strong solution Middle solution Weak solution

Hot water inlet temperature THW oC

0

10

20

30

40S

team

Gen

erat

ion

Rat

eG

steam

kg

/h

(a) Effect of hot water inlet temperature

(d) Elements on pallet of EXCEL interface

Figure 11: Simulators for VRF system


2012

194

5. CONCLUSION

In our study, we developed mathematical models of

the multi-type compression air conditioner (VRF

system) and the multi-stage absorption heat

transformer, which were used as examples of heating,

refrigeration, and air-conditioning systems.

Considering the complexity of these models, we also

discussed a simulator that we developed. The

mathematical models were fully validated by the

agreement of the experimental and simulation results.

Anyone in the world can easily use the simulator on

the Internet. We are expanding it for use with other

systems, which will be discussed in a future paper.

REFERENCES [1] K. Saito and J.S. Jeong, ―Latest system simulation

models for heating, refrigeration, and air-conditioning

systems, and Their applications,‖ IJACR , vol. 20, no.

1, pp.13, 2012.

[2] K. Ohno, K. Saito, H. Nakamura, H. Murata, Y. Jinno,

K. Konishi, and Y. Nakaso, ―Unsteady State

Simulation of VRF Systems,‖ 10thIEA Heat Pump

Conference 2011, Tokyo, Japan, 3.33, 2011.

[3] S. Jeong and S. Garimella, ―Falling-film and droplet

mode heat and mass transfer in a horizontal tube

LiBr/water absorber,‖ Int. J. Heat Mass Transfer, vol.

45, pp. 1445–1458, 2002.

[4] F. W. Dittus and L. M. K. Boelter, ―Heat transfer in

automobile radiators of the Tubular Type,‖

Publications in Engineering, 2, 443, Univ. of

California, Berkeley, 1930.

[5] N. Kattan, J. R. Thome, and D. Favrat, ―Flow boiling

in horizontal tubes: part 3–development of a new heat

transfer model based on flow pattern,‖ Trans. ASME,

vol. 120, pp. 156–165, 1998.

(a) Main screen of pallet

(b) Elements on pallet of GUI interface

(c) Simulation results on pallet by EXCEL file


2012

195

Drag Reduction of Bamboo and Abaca Fiber Suspensions

in Circular Pipe

Gunawana, M. Baqi

b, S. Fathernas

c and Yanuar

d

aDepartment of Mechanical Engineering, Faculty of Engineering

University of Indonesia, Depok 16424

Tel : (021) 7270032. Fax : (021) 7270033


bDepartment of Mechanical Engineering, Faculty of Engineering


Tel : (021) 7270032. Fax : (021) 7270033


cUnder Graduate Student, Department of Mechanical Engineering


Tel : (021) 7270032. Fax : (021) 7270033

dDepartment of Mechanical Engineering, Faculty of Engineering


Tel : (021) 7270032. Fax : (021) 7270033


ABSTRACT

The drag reduction of dispersions of fibers in aqueous

solutions of was studied as a function of concentration

with a circular pipe apparatus. Experiments were

carried out by measuring the pressure drop. The

purpose of this research is to investigate the reduction

of pressure drop in a circular pipe with the addition

fiber in aqueous solution. Circular pipe with 4 mm of

diameter is used in this study. Concentration of

bamboo and abaca fibers solutions are 200 ppm and

300 ppm. It was found that fibers solutions give rise to

drag reduction in turbulent flow range. Experimental

was conducted from low to high Reynolds number up

to 55,000. We observed a maximum drag reduction

ratio of 7 % at Reynolds number about 35,000 and

found that increased by increasing a concentration of

fiber solution.

Keywords: drag reduction, bamboo and abaca fibers

solution, pressure drop, turbulent flow.

1. INTRODUCTION

Environmental issues are a major topic of interest

studied mainly in energy efficiency. One topic of

particular interest is drag reduction in fluid transport

systems. The addition of a small amount of additives

suspension such as polymers, surfactans and fibers to

a turbulent Newtonian fluid flow can result in a drag

reduction, which appears in a number of flow fields,

and has received considerable attention. This

phenomena is reach to investigate sice initial

publication of Toms [1]. Using surfactans [2,3] to

obtain drag reduction in turbulent flow is very

effective and low mehanical degradation. However,

surfactans are contain as syntetic chemical so very

dangerous in environment. Although polymers are

safe in environment, they are not practical due to their

significant mechanical degradation. Yanuar et al [4,5]

also investigated the influence of biopolymer solutions

for drag reduction in internal and external flow. His

research show that biopolymer can reduce frictional

drag up to 30% but the mechanical degradaton

occured fastly.

Ogata, Numakawa and Kubo [6] reported that fiber

solutions from bacterial cellulose undergoing a

turbulent flow in a pipe thereby require a lower

pressure drop to maintain the same volumetric flow

rate. The addition of small amounts of fiber to the

flowing fluids can show significant effects on a lot of

flow types. It was found that bacterial cellulose

suspensions give rise to drag reduction in the turbulent

flow range. The maximum drag reduction ratio of 11%

and found that it increased with the concentration of

the fibre suspensions from bacterial cellulose.

The other fibres suspensions also investigated by

reserchers [7,8,9] such as asbestos or nylon fibers.

This research obtained that nylon and asbestos fibers

are effective to reduce drag but requires high

concentation and have a disadvantages with regard to

environmental load.

IMAT-UI 032


2012

196

Bamboo and abaca fibres are considered to have a low

environmental load and it is naturally derived from

bamboo and abaca plants. The purpose of this research

is to investigate the reduction of pressure drop in a

circular pipe with the addition fiber in aqueous

solution. Circular pipe with 4 mm of diameter is used

in this study. Concentration of bamboo and abaca

fibers solutions are 200 ppm and 300 ppm. It was

found that fibers solutions give rise to drag reduction

in turbulent flow range. Experimental was conducted

from low to high Reynolds number up to 55,000. We

observed a maximum drag reduction ratio of 7 % at

Reynolds number about 35,000 and found that

increased by increasing a concentration of fiber

solution.

2. EXPERIMENTAL SETUP

Figure 1: Experimental setup

The experimental set up is shown in figure 1. Figure 1

shows the test of rheological properties. The fibers

suspensions are circulated by piston pump. The

pressure drop gradient is measured at 1000 mm length

between each pressure tap by pressure transducer. The

diameter of pressure tap is 2 mm. The inner diameter

of test circular pipe d is 4 mm. The shear stress and

the shear rate can be obtained by measuring the

pressure drop gradient and the gradient of velocity,

respectively. The concentrations of fibers solution in

form of aqueous suspensions are 200 ppm and 300

ppm. The temperature is kept at 25 oC.

The bamboo and abaca fibers were taken from

bamboo and abaca plant. The size of this fibers are

homogen and have length about 0.5 mm. Bamboo and

abaca tree was gently wiped across the surface of a

special smooth metal table. After the bamboo and

abaca are wiped, fibers will separate with the other.

Then the fibers are dried and made the cutting process.

3. RHEOLOGICAL MODELS

The shear stress, τ is proportional to the velocity

gradient, (shear rate), can be described by

Newtonian model:

du

dy

(1)

Where is constant for the particular fluid that

is viscosity. The Newtonian viscosity depends on

the temperature and the pressure and is

independent of the shear rate. The viscosity is

defined as the ratio of shear stress to shear rate.

Several rheological models or rheological

equations of state have been proposed in order to

describe the nonlinear flow curves of non-

Newtonian fluids. Non-Newtonian fluids

Bingham, pseudo plastics, and dilatants are those

for which the flow curve is not linear. The

viscosity of a non-Newtonian fluid is not constant

at a given temperature and pressure but depends

on other factors such as the rate of shear in the

fluids.

Thus, the relationship shear stress and shear rate

may be described by measuring the pressure drop

gradient and the volumetric flow rate in circular

pipe flow is given by:

8

4

D P u

L D

(2)

Where: D is the inner pipe diameter, P is

pressure drop, L is the length of pipe (test

section), and u is the avarage velocity.

Coefficient of friction, f, can be obtained by

Darcy Equation:

2

2D gf h

L u

(3)

Where: f is the coefficient of friction, h is the

head gradient over the considered pipe length,

and g is the gravity acceleration.

Drag reduction in pipe can obtain by equation:


2012

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100%fiberf f

DR xf

(4)


Figure 2: Flow curve of fibers suspensions

Figure 2 shows the flow curves of the fibers

suspensions. The wall stress τ and sheer rate were

calculated from the experimental data from laminer to

turbulent flow regime. The solid line in figure 2

indicates the value obtained by the viscosity of water.

The data of bamboo fiber and abaca fiber are shown

linear relationship between wall stress and flow rate. It

is incated that the fibers solutions are Newtonia fluid.

The viscosity is seen to increase with concentration.

The value of figure 2 is used to obtain the Reynolds

number and friction factor.

103

104

105

10-1

Pure water

Abaca Fibre 200 ppm

Abaca Fibre 300 ppm

Bamboo Fibre 300 ppm


T = 25o C

f = 64/Re*

f = 0.3164*Re* (̂1/4)

f

Re

Figure 3: Flow curve of fibers suspensions

Figure 3 shows the relationship between Reynolds

number and friction factor coefficient based on the

measured pressure drop for 2 suspensions with 2

variation of concentration. The data will be compared

with Hagen Pouiselle equation in laminar flow and the

Blasius equation in turbulent flow. The data of water

also shown in this figure. The coefficient of friction of

fibers suspensions fit with the coefficient of friction of

water for circular pipe in laminar flow. In turbulent

flow, up to Reynolds number about 25.000, the

coefficient of friction also fit with coefficient friction

of water and Blasius equation. The data show that

coefficient of friction fibers at Re > 25.000 is lower

that water data and Blasius equation. The drag

reduction is increase with increasing of fiber

concentration. The data of bamboo fiber is lower that

abaca fiber in same concentration and Reynolds

number.

102

103

3

4

5

6

7

Re.f 1/2

f -1

/2


Abacca Fibre 300 ppm


Abacca Fibre 200 ppm

C

B

A

Figure 4: Characteristic of drag reduction

Figure 4 shows the relationship between f -1/2

and

Re.f1/2

where f is denote the Fanning friction factor. It

can be seen that in the Laminar flow regime where

Re.f1/2

is small, the data well fitted by a Newtonian

laminar flow curve (C). In contrast the data in

turbulent flow regime is alligned parallel in the curve

(A). Drag reduction occured if the data is higher than

curve A. The figure shows that increasing the

concentration, can increase the data from the curve A.

Data of bamboo and abaca fibers at high Re.f1/2

are

parallel with curve A buat not fitted. The data is

greater that curve A. The drag reduction that occured

based on this graph is indicates a Type B drag

reduction, which can be seen in fiber and polymer

suspensions. Generally, a type B dra- reducing

mechanism is associated with suppression of vortices.

For fiber suspensions the flow fields influence the

fluid resistance, and the fiber suppress vortices when

they are uniformly distributed in the flow direction,

thus resulting the drga reduction. The friction factors

of high-concentration solutions data still far to the

Virk‘s line (B) according to the increase in the

Reynolds number, Re.


2012

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2,0x104

3,0x104

4,0x104

5,0x104

2

4

6

8

10

12


Abaca Fibre 300 ppm


Abaca Fibre 200 ppm

DR

(%

)

Re

Figure 4: Ratio drag reduction

Figure 4 shows the ratio drag reduction of fiber

suspensions. Based on figure, it can be seen from

these results that drag reduction for a given bammbo

and abaca fibers concentration only occurs above a

critical value Reynolds number. The value of critical

Reynolds number is about 25.000. Below this critical

value the fluid exhibits normal Newtonian viscous

behavior, although the flow is turbulent flow.

The maximum drag reduction occured at the Reynolds

number about 35.000. The drag reduction increases

start from Reynolds number 25.00 up to 35.000. After

Reynolds number about 35.000, the data shows

constant. The drag reduction increased slightly with

increasing concentrations. Drag recution of bamboo

fiber is greater then abaca fiber. The reported value for

bamboo fiber suspensions of 300 ppm and 2000 ppm

in the turbulent flow range in circular pipe has

maximum drag reduction about 7% and 5%. For same

concentration and Reynolds number, drag reduction

for abaca fiber is 6% and 4% respectively.

5. CONCLUSION

Pressure drop measurements for bamboo fiber and

abaca fiber suspensions flowing in circular pipe were

performed and the following result is obtained . The

drag-reduction effect of the bamboo fiber and abaca

fiber were verified. The effect occurred only above

some critical Reynolds number which was affected by

the concentration of the fiber suspensions. Drag

reduction is significantly affected by the type of fiber

and the concentration of fiber. For bamboo fiber, the

range drag reduction is about 5% through 7% depend

on concentration. For abaca fiber, the drag reduction

accured about 4% through 6% respectively. The

maximum drag reduction is 7% at Reynolds number

about 35.000. Drag reduction of bamboo fiber and

abaca fiber is type B drag reduction same as polymer

type drag reduction.

ACKNOWLEDGMENT

This work is supported by the Directorate for Research

and Community Service, University of Indonesia

(RUUI).

REFERENCES

[1] Toms. B. A. ―Somle observations onl the flow

of linear polymer solutions through straight tubes

at large Reynolds numbers," International

Congress onl Rhecology,I Holland. 1948.

Amsterdlam. North I lolh.aid, 1949, Part 11, pp.

135-141

[2] F.-C. Li, Y. Kawaguchi, K. Hishida, and M.

Oshima, ―Investigation of turbulent structures in

a drag-reduced turbulrnt channel flow with

syrfactant additive by stereoscopic particle image

velocimetry‖, Experiments in Fluids, vol. 40, no.

2, pp. 218-230, 2006.

[3] H. W. Bewersdorff, ―Rheology of drag reducing

surfactat solutions‖, in Proceedings of the ASME

Fluids Engineering Division Summer Meeting

(FED‘96), vol. 237, pp. 25-29, San Diego, Calif,

USA, 1996.

[4] Yanuar and Watanabe K. ―Tom‘s effect of guar

gum additive for crude oil in flow through square

ducts.‖ The 14th

International symposium on

transport phenomena. Bali Indonesia. Elsevier

2004. P.599 – 603.

[5] Yanuar, Gunawan and M. Baqi, ―Characteristics

of Drag Reduction by Guar Gum in Spiral Pipes‖

Journal Teknologi. Vol.58 2012, pp. 95–99.

[6] Satoshi Ogata, Tetsuya Numakawa and Takuya

Kubo. ―Drag reduction of bacterial cellulose

suspensions. Advanced in Mechanical

Engineering. 2011. Pp 1-6.

[7] P.S. Virk and R.H. Chen, ―Type B drag reduction

by aqueous and saline solutions of two

biopolymers at high Reynolds number‖, in

Preceedings of the 2nd International Symposium

on Seawater Drag Reduction, pp. 545-558,

Busan, Korea, May 2005.

[8] A.A. Robertson and S.G. Mason, ―The

characteristics of dilute fiber suspensions‖,

TAPPI, vol. 40, pp. 326-334, 1957.


2012

199

[9] W. Mih and J. Parker, ―Velocity profile

measurements and phenomenological description

of turbulent fiber suspension pipe flow‖, TAPPI,

vol. 50, pp. 237-246, 1967.

[10] Yanuar and Watanabe K. ― Drag Reduction of

Guar Gum in Crude oil‖. The 13th

International

Symposium on Trannsport Phenomena. Victoria

Canada. Elsevier 2002. P. 833 – 836.

[11] Yanuar, et al. ―Hydraulics conveyances of mud

slurry by a spiral pipe‖ Journal of Mechanical

Science and Technology 23 (2009) 1835 – 1839.

Springer.


2012

200

A Study on the Effect of Exhaust Gases on the Indoor Air Quality

Onboard Ships

Arman Ariffin

a and Hayati Abdullah

b

aPlan Department, Royal Malaysian Navy Headquarters, Ministry of Defence,

Jalan Padang Tembak, 50634 Kuala Lumpur, Malaysia

Email:[email protected]

bFaculty of Mechanical Engineering, Unversiti Teknologi Malaysia

81310 UTM Johor Bahru, Johor, Malaysia

Email:[email protected]

ABSTRACT

The understanding of the exhaust gas behaviour from

ship plume is necessary in order to avoid serious

operational problems onboard modern ships. The

interference of exhaust gas on the air intake for

ventilation system can result in poor indoor air quality

(IAQ) and can adversely affect the performance of

human and equipment onboard ships. This paper

presents the results of an indoor air quality study

carried out onboard ship and focuses on parameters

such as temperature, humidity and major air

pollutants. An initial study of the velocity ratio K

which represents the ratio of exhaust velocity to

relative wind velocity will also be presented to

investigate the effect of the velocity ratio K on the

indoor air quality for three locations of the air intake

for the ventilation system and two different conditions

of the ship namely alongside and cruising at an

economical speed.

Keywords : Indoor Air Quality, Ship’s Exhaust

Gases, Velocity Ratio

1. INTRODUCTION

The understanding of exhaust gas behavior is

important in ship design and smoke nuisance

onboard ship has long been studied since the

evolution of ship construction. The downwash of

exhaust smoke, especially on modern naval ship

with lower stack in order to reduce the infrared

signature, can cause negative effects such as

suction of hot plume to intake of gas turbine or

HVAC onboard, high temperature and

contamination of topside electric equipment and

interference of exhaust smoke with the flight

deck operation. Problem of the exhaust gas

behavior was reported by Nolan [1]. In the study,

various types of wind tunnel testing were

conducted with the cooperation of Maritime

Commision‘s smoke test program and Langley

Memorial Aeronautical Laboratory. It started

with the actual model S.S America with a scale of

1:96. It was selected as the first model in the test

because of smoke trouble. Some modifications

were made by raising the stack 15 feet (4.57m).

Equipment used in the test includes an

anemometer to measure the wind velocity and an

orifice in the air supply to identify the smoke

velocities. The test was also conducted with

smoke temperatures of 300 oF (149

oC) and 500

°F (260 oC). On the first test, it was concluded

that the height of the stack was affecting the flow

of the smoke. Higher stack will create turbulence

away from the aft stern. The downwash does not

take the smoke down far enough to reach the

turbulence zone, so it will float clear of the ship

and it is an advantage for all on board. Nolan

introduced the S/W ratio or velocity ratio K

where S is the smoke velocity and is evaluated as

the stack gas volume per second divided by the

discharge stack area. W is the wind velocity

relative to the ship. He found that with an S/W

ratio or K below 2, if the wind was at an angle of

even 5 degrees, the smoke could travel all the

way to the base of the stack. Other stacks were

tested with this model but satisfactory results

were only obtained if the S/W ratio was at least 2.

Heated smoke was also used in the study. The

temperature of the smoke was previously about

130 oF (54.4

oC). The temperatures were then

increased to 300 oF (149

oC) and 500 °F (260

oC).

It was concluded that heated smoke floats clear of

the ship better than unheated smoke. The

experiments carried out included different

characteristics of the funnel from the velocity

IMAT-UI 033


2012

201

ratio K, height of stack, stack design, and exhaust

temperature. It was concluded that the K value

shall be at least 2 to keep the smoke above the

turbulence zone. The stack design should also be

as small as possible to allow high speed of smoke

velocity. The heated smoke floats clear of the

ship because the buoyancy of the smoke exerts

considerable influence on the performance of a

ship‘s stack.

A flow visualization study of exhaust smoke-

superstructure interaction onboard naval ships

was studied by P.R. Kulkarni et al [2]. Four

variants of superstructure arrangements with 1:50

model were studied to gain an understanding of

the typical flow field around the topside of naval

ships and the interaction between bluff body air

wake and the ship exhaust. They noted that most

of the problems arise from the fact that the funnel

superstructure creates a low-pressure zone behind

the exhaust stack or lee side and it will naturally

suck in wind into that area. The flows that will be

sucked in are the flows from the funnel top and

side as well as the exhaust gases from the funnel.

As in the study by Nolan, they concluded that a

velocity ratio of at least 2 is a requirement for

satisfactory operation and to prevent the sucking

of exhaust into the gas turbine intakes.

Apart from wind tunnel studies, numerical

investigations also gave similar results on the

importance of the velocity ratio K on the exhaust

gas behavior. S. Ergin et al [3] carried out

numerical studies on the exhaust smoke-

superstructure interaction on a naval ship and

demonstrated that computational fluid dynamics

can be a powerful tool to study the problem of

exhaust smoke-superstructure interaction. They

investigated 3 main elements that affect the

smoke print onboard which are the yaw angles,

velocity ratios and the exhaust smoke

temperature. The values of the velocity ratio K

studied are 1,2, 3 and 4. The yaw angle, ψ from

Port side included 0°, 5°, 10°, 15°, 20° and 30°,

and exhaust temperature studied are for 15°C,

200°C, 300°C and 400°C. The numerical results

concluded that to minimize the effect of

downwash, the yaw angle, ψ should be more than

10° and velocity ratio K equals to the value of 2

should be maintained. They also note that the

effect of buoyancy forces on the plume rise when

compared to momentum is not as significant.

A numerical study on the effect of air quality

using a very large eddy simulation program was

presented by F.Camelli et. al [4]. They studied

the coupling of the ship topside flow to the

thermal transport and diffusion of the exhaust gas

for 0o and 30

o of angle of attack of the inflow.

They computed the temperature, NOx and SO2

levels and their results compared well with

experimental data from wind tunnel testing.

P.R. Kulkarni et al [5](Kulkarni, Singh, &

Seshadri, 2005) also conducted experimental

study of the flow field around a simplified

superstructure with two funnels. Two

configurations were studied in which the first

configuration is when the 2 funnels are aligned in

the centerline and in the second configuration, the

2 funnels are not aligned. They observed that

when two funnels are aligned with the wind,

there is a momentum shielding effect by the

upstream plume on the downward plume but

when the funnels are offset with respect to the

incident wind, there is no shielding effect of the

forward plume on the aft plume.

Huang J. et al [6] carried out a CFD study on the

temperature and NOx concentration levels. Their

results showed that operation with lower smoke

speed and larger head wind speed will result in a

lower exhaust plume that can move close enough

to the superstructure of the ship to be entrapped

in the down flow downstream of the

superstructure and into the flight deck. This poses

a risk to potentially harm equipment and

ventilation intakes.

2. PROBLEM DESCRIPTION

The tests presented in this paper were carried out

in the South China Sea in the range of 50 nautical

miles close to coastal. The cruising condition is

considered a normal activity without any

additional requirements involved.

The ship has complex Heating, Ventilating and

Air Conditioning (HVAC) System onboard. It is

complete with the capability to be operated in

normal environmental condition up to the

Chemical, Biological, Radiological and Nuclear

(CBRN) contaminated area. In the general

arrangement, it has 3 independent Self

Confinement Zone (SCZ). Each SCZ zone has

its own intake, air handling and air conditioning

system with total independence configuration.


2012

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In the main machinery configuration, the ship is

powered by two marine diesel engines with a

total of 12 Megawatt and four marine diesel

generators with a total of 2 Megawatt. To

increase the stealthy design of this ship in

infrared signature, it was constructed with 6

points of discharge exhaust at sea water level on

both sides of the ship. The points of exhaust

discharge are depicted in Figure 1.

Figure 1: Points of exhaust discharge

It was reported that the indoor air quality was

affected by the exhaust system. Studies were

then conducted to identify the possible cause of

this problem.

3. METHODOLOGY FOR INDOOR

AIR QUALITY TEST

Indoor Air Quality (IAQ) Test was conducted to

evaluate the IAQ onboard ship. It is used to

confirm the level of concentration at several

measuring points and would be the baseline for

future studies. An initial walkthrough of the area

during normal activity provides information on

all four basic factors influencing IAQ (occupants,

HVAC system, pollutant pathway and

contamination sources). All studied areas were

visited for initial investigation and several

measurements were taken. Appropriate

instruments and measurements were selected for

data collection process.

The sampling probes were located between 75

and 120 cm from the floor of the sampling

position and sampling was carried out when the

ship was alongside and cruising. Typical

measurements using direct reading devices were

employed for measurements of temperature,

relative humidity (RH), carbon dioxide, sulfur

dioxide, nitrogen dioxide and nitrogen oxide.

Measurements using air pumps and collection

media were sent for laboratory analysis.

Air test for temperature, RH, carbon dioxide,

sulfur dioxide, nitrogen dioxide and nitrogen

oxide were taken with the portable gas analyzer

Model Testo 350XL. Air tests for diesel dust,

HCI, H2SO4, HN3, HN2, VOC compounds were

taken with GilAir-5 & Low Flow Module

Constant Flow air pumps. The air sampler was

calibrated using the Gilian Gilibrator 2

Calibration System.

The analysis was conducted at the laboratory

facility of Australia Laboratory Services (ALS).

ALS laboratories operate in compliance with ISO

17025 (General requirements for the competence

of testing and calibration laboratories). The

diesel dust, inorganics acid and VOC are

analyzed using methods with reference to NIOSH

5040, NIOSH 7903 and NIOSH 1500.

4. ENVIRONMENTAL CONDITIONS

The IAQ test was conducted in March 2010 with

two different conditions; alongside and cruising.

The daily weather summary recorded is given in

Table 1. Table 1: Daily Weather Summary.

Temperature:

Mean 28

Max 32

Min 25

Moisture:

Dew Point 25

Average RH 80

Max RH 89

Min RH 63

Sea Level Pressure 1007.77 hPa

Wind:

Average Speed 6 km/h

Max speed 13 km/h

Visibility 8.9 kilometers

Event Rain

5. SAMPLING PROCESS

The IAQ test was conducted in two different

conditions. Each case uses the same method of

sampling process:

a. Case 1: Ship alongside.

b. Case 2: Ship cruising.

There are 3 locations identified to be

significant in the effect to the indoor air quality


2012

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as indicated in Figures 2, 3 and 4. They are the

intake positions for each SCZ and marked with

―star‖ :

a. Location 1: Intake for SCZ 1.

Position at forward upper deck

starboard side.

b. Location 2: Intake for SCZ 2.

Position at amidships flag deck

starboard side

c. Loaction 3: Intake for SCZ 3.

Position at aft boat deck starboard

side.

Figure 2: Location 1.



The direct reading measurements were

taken using the portable gas analyzer Model

Testo 350XL for temperature, RH, carbon

dioxide, sulfur dioxide, nitrogen dioxide and

nitrogen oxide.

For the diesel dust, HCI, H2SO4, HN3,

HN2, and VOC compounds, measurements were

taken with the GilAir-5 and Low Flow Module

Constant Flow air pumps using the sampling

bottles. The duration for each position is 1 hour

and accumulates 12 liters of air. The collected air

in the sampling bottles were placed in a sealed

box and sent to ALS for laboratory analysis.


For both cases, the result of the concentration on

sulfuric acid is shown in Table 2.

Table 2: Sulfuric acid concentration (mg/m3) at air intake

area.

Area Case 1 Case 2

Intake SCZ 1 3.42 47.96

Intake SCZ 2 8.09 9.3

Intake SCZ 3 2.42 22.28

The trial for case 1 has the K value of 2.5 and the

yaw angle, ψ = 60° relative to portside. It can be

seen that the intake for all SCZ was slightly

affected by smoke with a higher concentration

level at SCZ 2. The trial for case 2 has the K

value of 0.2 and the yaw angle, ψ = 45° relative

to portside. It can be seen that the intake for all

SCZ was significantly affected by the exhaust

gas. The results indicate a similar trend to the

results obtained in the literature where higher K

values is required for satisfactory ship operation

in terms of exhaust gas behavior and the

downwash phenomena. However, further studies

need to be carried out with more sample data.

Wind tunnel testing for the side exhaust

configuration is planned for the future in order to

further understand the interaction of exhaust

smoke and superstructure interaction for this

configuration.

7. SUMMARY

The study on indoor air quality has indicated

similar trend to the published research results. It

is important to ensure that the concentration

levels of species such SO2 and NOx do not rise

above healthy levels onboard ships so as not to

harm the crew and ship. The results of future

studies in this area will lead to a better

understanding of the recirculation zones and how

they affect the concentration levels of air

pollutants and will be able to assist in the

understanding of the actual effect of exhaust

smoke to the internal environment onboard ships.


2012

204

Acknowledgment

The authors wish to thank Royal Malaysian Navy

and Universiti Teknologi Malaysia for the

support in carrying out this project.

REFERENCES

[1] Nolan, R. W. (1946). Design of stacks to

minimise smoke nuisance. Trans

SNAME, 54, 42-82.

[2] Kulkarni, P. R., Singh, S. N., & Seshadri,

V. (2005). Flow visualization studies of

exhaust smoke-superstructure interaction

on naval ships. Naval Engineers Journal,

117(1), 41-56.

[3] Ergin, S., Parah, Y., & Dobrucali, E.

(2012). A numerical investigation of

exhaust smoke-superstructure interaction

on a naval ship

Sustainable Maritime Transportation and

Exploitation of Sea Resources.

[4] Camelli, F., Sandberg, W. C., &

Ramamurti, R. (2004). VLES Study of

Ship Stack Gas Dynamics. The 42nd

AIAA Aerospace Science Meeting and

Exhibition.

[5] Kulkarni, P. R., Singh, S. N., & Seshadri,

V. (2005). Experimental study of the

flow field over simplified superstructure

of a ship. Int J Maritime Eng, IJME Part

A3, 147, 19-42.

[6] Huang, J., Carrica, P. M., & Stern, F.

(2012). A method to compute ship

exhaust plumes with waves and wind.

International Journal for Numerical

Methods in Fluids, 68(2), 160-180.


2012

205

ABSTRACT

This article presents a study to estimate the potential

saving in annual operating cost of a hypothetical

greenhouse used for planting strawberry, in Johor

Bahru, Malaysia. The greenhouse needs to be

maintained at a constant temperature of 20°C at all

time. The goal of this study is to select a suitable TES

system that can save the annual cost of electricity

usage to meet the cooling load requirement of the

greenhouse, based on a 24 hours operating duration

and local electricity tariff. Comparison is made with

the annual cost for running a conventional air-

conditioning (AC) system to meet the cooling

requirement. The cooling load requirement of the

greenhouse dictates the capacity and size of the

potential TES systems, which was estimated based on

the highest total annual cooling load. Three TES

system operating arrangements were considered in this

study: TES full storage combined with AC systems,

TES full storage and TES partial storage. Among

these three arrangements, the TES full storage was

found to have the highest an annual cost saving of

about RM 58,990 compared to the cost of using the

conventional AC system alone. This represents about

68 % of annual operating cost saving, which is

considered very significant.

Keywords : Thermal Energy Storage (TES), Lowland

Farming House, Cooling Load

Estimation, Greenhouse Air-

conditioning System.

1. INTRODUCTION

Thermal energy storage (TES) can be considered as

the temporary storage of energy for later use when

cooling or heating is needed. For cooling applications,

energy is stored at low temperatures while for heating

applications, the energy is stored at high temperatures

[1]. The interest in cool storage for commercial

applications grew significantly especially for countries

in hot and humid regions where a very high on-peak

demand load occurred in the midday but persisted for

a short period of time [2]. TES technology is seen as

one of the primary solutions to the electrical power

imbalance between its production and continuous

demand. The fundamentals, case studies, design and

history of the TES be found in various literatures

[3,4]. TES may also be a potential cost-saving solution

in countries where the electricity rate is on time-based.

This technology can shift cooling energy usage time

from on-peak periods to off-peak periods and hence

avoids peak demand electricity charges.

In non-residential buildings the TES technology may

become an attractive alternative if one or more of the

following conditions exist [5]: short period of HVAC

demand, frequently varying HVAC loads, infrequent

or cyclical loads, HVAC demand and supply do not

match, economic incentives are provided for using off-

peak energy, energy supply is limited by the utility

company, hence making it impossible to satisfy the

maximum load directly, and the capacity of an

existing chiller is too low to meet the peak load

demand. TES technology can be promoted because it

can substantially reduce the total energy consumption,

conserving fossil fuels and reducing costly imports of

oil and other energy resources. With TES, one can

adjust the time-discrepancy or rate variance between

energy supply and energy demand, thereby playing a

vital role in the conservation of energy [6,7].

TES systems are usually operated in two modes: full

storage and partial storage. The partial storage TES

can further be categorized into load leveling and

demand limiting storage systems [8]. The Full storage

TES systems, also known as load shifting systems are

typically designed to shift all building cooling load

demands from the on-peak period to the off-peak

IMAT-UI 034

Application of Thermal Energy Storage System For a

Lowland Greenhouse


b & Norull Ahmad Norull Azman



Tel : (+607) 5534748. Fax : (+607) 5566159


bFaculty of Mechanical Engineering


Tel : (+607) 5534749. Fax : (+607) 5566159



2012

206

period of a day. In this system, the chiller runs at its

full capacity during the off-peak period and night time

when the building cooling load demand is low and

electricity tariff is cheaper. In those periods, the chiller

charges the storage and meets the building cooling

load requirements simultaneously. Since the full

storage TES systems meet all the building cooling

loads during the day time, it will result in larger and

therefore more expensive chillers and storage units

compared to the partial storage systems. The full

storage TES systems are likely to be attractive under

the following conditions [7]: spikes in the peak load

curve are of short duration, time of use energy rates

based on short-duration peak periods, there are short

overlaps between peak loads and peak energy periods,

high peak demand charges apply, and some utility

companies offer incentives for using TES.

The partial storage TES system provides the best

mode for reducing demand charge and saving

electricity cost. Therefore it represents more that 50%

of the thermal storage installations worldwide.

However, these systems are not capable of shifting as

much load on a day as the full storage TES systems. In

the partial storage system, the chiller operates to meet

part of the cooling load demand and the rest is met by

the storage tank during the day time. Usually in this

system, the chiller is sized at a capacity smaller than

the design load. Partial storage TES systems can be

further classified based on the selected operation

strategies, load leveling, or demand limiting

operations. In a load leveling system the chiller

operates at full capacity for 24 hours of the design

day. When the building cooling load demand is less

than the chiller capacity, the excess cooling is stored

in the storage tank until the tank is full. When the load

exceeds the chiller capacity, the additional cooling is

supplied from the storage tank.

This paper presents a study on the use of TES system

in a lowland greenhouse for planting strawberry. In

Malaysia, Cameron Highland is one of the suitable

areas to plant strawberry in the open due to its suitable

ambient temperature. Strawberry grows healthy in a

temperature range of 17ºC - 20°C [9]. However due to

insufficient land area in Cameron Highland, building

greenhouses at a lowland area is seen as one of the

possible solutions to this issue. Most lowland areas in

Malaysia experiences temperature around 32°C all

year. Conventional air conditioning (AC) systems will

be required to maintain the greenhouse at the

temperature needed for planting strawberry. However,

such systems will consume a lot of electricity to

operate continuously. This will results in high

operating cost for the greenhouses. Thermal energy

storage (TES) systems can therefore be considered as

the way to achieve this goal since they are able to shift

cooling energy use to from peak time to non-peak

times. They can chill storage media such as water, ice,

or a phase-change material during the periods of low

cooling demand for use later to meet the air-

conditioning loads.

The goal of this study is to select a suitable TES

system that can help save the cost of electricity needed

to meet the cooling load requirement of the

greenhouse, based on a 24 hours operating duration.

The cooling load requirement of the greenhouse was

estimated to determine the capacity and size of the

TES systems. This was done based on the highest total

annual cooling load. Three TES system operating

arrangements were considered in this study: TES full

storage combined with AC systems, TES full storage

and TES partial storage. The operating cost of the

these TES systems were compared in term of the total

amount of electricity usage during a 24 hours

operation, based on the local electricity tariff.

2. METHODOLOGY

The effects of operation strategy on electricity

consumption of TES systems for strawberry

production were estimated for three different operating

strategies. These are TES full storage combined with a

conventional AC system, TES full storage and TES

partial storage. For the TES full storage combined

with AC systems, the TES system was designed to

meet all on-peak cooling loads, while the AC system

meets all the off-peak cooling loads. The TES full

storage was designed to meet all on-peak cooling

loads from storage and the TES partial storage was

designed to meet part of the cooling load requirement

from the storage and the other part directly from the

chiller during the on-peak period. The major

components of the TES system consists of an

evaporator, a condenser, a cooling tower, storage tank

and water pumps, as shown in Figure 1. The

evaporator is used to generate chilled water and later

stored in the storage tank. The chilled water is used to

cool the hypothetical green house by discharging it

through the secondary chilled water pump. The heat

from the green house is carried by the water to the

storage tank before it is removed in the evaporator.

Primary chilled water pump is used to transport the

water from the storage tank to the evaporator. The

condenser water pump directs cool water from the

cooling tower into the condenser to absorb heat from

the evaporator. The heat absorbed by the condenser is

rejected to the cooling tower by the cooling water and

rejected it to the surroundings. The TES system was

designed based on the specification given in Table 1.


2012

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Figure 1: Schematic diagram of the TES system.

Table 1: Design specification for the TES system.

Parameter Value

Indoor temperature 20℃

Chilled water supply temperature 5℃

Chilled water return temperature 15℃

Storage water supply temperature 7°C

Storage water return temperature 17°C

Condenser water supply temperature 30℃

Condenser water return temperature 35℃

The hypothetical greenhouse is a single gable type

with a single-peaked roof, measuring 60 x 10 x 4 m

and is covered with polyethylene material for both the

walls and roof. The greenhouse is to be located in

Johor Bahru, Malaysia, in which the location is 1.3°

North and 103.7° East. Figure 2 illustrates the layout

of the strawberry planting arrangement in the

greenhouse.

Figure 2 Layout of strawberry plants (in mm) in

the hypothetical greenhouse.

The electricity cost for operating the TES systems was

estimated by first estimating the cooling load of the

hypothetical greenhouse for strawberry production.

This information is then used to determine the

capacity and size of the TES systems. The total

cooling load for the greenhouse consists of external as

well as internal thermal loads. The external thermal

load is due to heat transfer by conduction through the

walls, roof, floor and doors. The internal thermal loads

are due to the sensible and latent heat transfer from the

occupants, appliances and the strawberry plants. The

cooling load calculation for the hypothetical

greenhouse was performed using a TROPICA

software [10] that was developed based on a weighting

factor method. Figure 3 shows the flow chart of this

software.

Figure 3 Flow chart of the TROPICA software

[10]


2012

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2.1 TES Full Storage Combined with AC

System

The TES full storage combined with AC systems

consist of a TES system (chiller and storage) and a

conventional AC system. The AC system includes a

cooling tower and water pumps. The chiller system

was selected based on the highest total cooling load

during the on-peak hours. The highest total cooling

load during on-peak occurs between the month of May

and June which is about 822 kW. The chiller operating

schedule was then determined to estimate the hourly

storage balance. It is done based on the total cooling

load during the on-peak hours and the cooling capacity

of the selected chiller system. The combined TES and

AC systems are designed to operate for 24 hours. The

conventional AC system was selected based on the

highest cooling load during the off-peak hours which

is about 47 kW. In this arrangement, the AC system

will operate to meet the cooling load demand of the

greenhouse during off-peak hours. The TES system

will be in a charging mode during this period. During

the on-peak hours, the AC system will be shut off and

the cooling load demand of the greenhouse is met by

the fully charged TES system.

2.2 TES Full Storage System

In this TES system arrangement, the chiller will

operate at its full capacity during the off-peak hours,

i.e. from 10:00 pm until 8:00 am, to charge the system

and at the same time meet the cooling loads demand of

the greenhouse. During the on-peak hours which is

from 9 am until 10 pm, the charged TES system will

be used to meet all cooling requirements by the

greenhouse. The chiller for this TES system was

selected based on the highest total cooling load in a

day, which is about 1056 kW. This highest cooling

load occurs between the month of May and June. The

full storage system is also designed for 24 hours

operation.

2.3 TES Partial Storage System

The TES partial storage system is designed to meet

part of the cooling load from its storage during the on-

peak hours. The other part is supplied directly from its

chiller system. The chiller system will be in the

charging mode when the cooling load requirement is

less than the output of the chiller. The chiller will be in

the discharging mode when the cooling load

requirement of the greenhouse is greater than the

output of the chiller. For this TES arrangement, the

chiller system was selected based on the highest total

cooling load during the on-peak hours, which is about

822 kW. This highest total cooling load occurs

between the month of May and June. This TES system

is also designed for the chiller to operate at full

capacity for 24 hours.

2.4 Operating Cost Analysis The total cost of electricity usage by the TES systems

depends on the power rating of the electric motor of

each equipment in the systems, the operating duration

of the equipments and the local electric tariff, during

both the on-peak and off-peak hours. Table 2 shows

the electric tariff of the on-peak and off-peak hours in

Johor Bahru, Malaysia.

Table 2: Electricity tariff in Johor Bahru, Malaysia

On-peak 8:00 am - 10:00 pm RM 0.312 /kWh

Off-peak 10:00 pm - 8:00 am RM 0.192 /kWh

Table 3 shows the power rating of the electric motor

for all the three TES system arrangements. The

equipment operating schedule during the off- and on-

peak hours is also shown in the table. The electric

motor ratings were obtained from the corresponding

manufacturers.

Table 3: Electric motor rating and operating schedule of the

TES systems Equipment Motor

Rating

(KW)

Off-Peak

Hours

On-

Peak

Hours

TES full storage with

AC systems:

Chiller

Cooling tower

Chilled water pump

Distribution water

pump

Condenser water pump

AC system

31.19

0.7456

0.75

0.75

1.1

19.3

9

0

9 0

9

0

9

10

0

14

0

0

TES full storage

system:

Chiller

Cooling tower

Chilled water pump

Distribution water

pump


37.95

1.12

0.75

0.75

1.1

9

0

9 0

9

1

9

0

14

0

TES partial storage

system:

Chiller

Cooling tower

Chilled water pump

Distribution water

pump


19.71

0.37

0.55

0.55

0.75

10

8

10 8

10

0

10

8

14

8


2012

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3.1 Cooling Loads of Hypothetical Green

House

Figure 4 shows the hourly cooling loads profile of the

hypothetical green house in a year. The profile of the

cooling load reflects the local weather data which

dictates the amount of sensible heat conduction into

the green house and the magnitude of solar heat gain.

From 10 am to 1 pm, conduction heat gain and solar

load increases, resulting in the increase of cooling

loads. The peak cooling loads occur at about 1 hour

past noon time. As the external heat gains drop past

the noon time, the cabin air temperature exhibits a

similar decreasing trend. The maximum cooling load

occurs between the month of January and February,

which is about 100 kW.

Figure 4: Hourly cooling load profile for the

hypothetical green house in a year.

3.2 Combined TES Full Storage and AC

Systems


hypothetical green house when the combined TES full

storage and AC systems are employed to meet the

cooling load demand. In this arrangement, the TES

system is designed to meet all the on-peak cooling

loads, while the AC system meets all the off-peak

cooling loads. As seen from the figure, the chiller of

the TES system is charged from 10 pm until 8 am

during the off-peak hours. During this period, the AC

system is used to meet the cooling load demand of the

green house. It can be seen that the chiller charging

capacity is close to 100 kW.

Figure 5: Hourly cooling load profile for

combined TES full storage and AC

systems.

The AC system is operating at a capacity of about 50

kW when the cooling load requirement of the green

house is about 20 kW. The AC system is running with

cooling capacity much higher than the cooling load

requirement. This is obviously not economical.

However, since this occurs during an off-peak period

in which the electric tariff is lower, a potential saving

of cooling cost can still be achieved. The figure also

shows that from 8 am until 10 pm, which is during the

on-peak hours where the electric tariff is higher, the

cooling load requirement of the green house is met by

the energy stored by the TES system. During this time,

the AC system is turned off to further save electricity

consumption. The same cycle will be repeated for the

next 24 hours period.

3.3 TES Full Storage System Only


green house when only the TES full storage system is

employed. In this case the TES system is designed to

solely meet all the cooling load demand, during both

the off-peak and on-peak period. As before, the TES

system is charged from 10 pm to 8 am, i.e. during the

off-peak period. The chiller capacity of this system is

now about 120 kW. The extra 20 kW of cooling

capacity is used to meet the cooling load requirement

of the green house.


2012

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Figure 6: Hourly cooling load profile for TES full

storage system.

3.4 TES Partial Storage System


green house when the TES partial storage system is

used. With this system, the TES is charged from 10

pm to 5 pm. The chiller runs at a capacity of about 61

kW during the charging period. This is the lowest

charging capacity compared to the previous

arrangements. In this arrangement, the chill water

produced during charging is directly used to meet the

cooling load demand of the green house. At same

time, the excess chill water is stored in the storage

tank. When the cooling load exceeds the chiller

capacity, the additional chill water will be discharged

from the storage tank. This happens from the period of

9 am to 4 pm which is the on-peak hours.

Figure 7: Hourly cooling load profiles for TES

partial storage system.

3.5 Estimation of Electricity Cost

An economic evaluation of the cooling systems

requires estimation of the annual operating cost of the

system. The operating costs are affected by several

factors such as cooling system capacities, the

operating time period and the usage of the

conventional AC system. Higher chiller capacity

obviously results in higher electricity consumption,

which increases the operating cost. Longer period of

time taken to charge the chiller will also increase the

electricity consumption. Therefore, the operating cost

will also increase. Whether the systems operate only

during off-peak or during both the off-peak and the

on-peak period, it will also has an impact on the

operating costs because the electric tariff is higher

during the on-peak period. The use of conventional

AC system will further increase the electricity

consumption and thus the operating cost. Table 4

shows the cost estimation of the electricity usage for

the three TES system operating arrangements as

described above.

Table 4: Cost estimation of various TES system operating

arrangements

Time/Cost

(RM)

Type of System

AC

System

Full

Storage

With AC

System

Full

Storage

Partial

Storage

A Day 235.27 98.66 75.13 96.83

A Month 7,058.10 2,959.80 2,253.90 2,904.90

A Year 84,697.20 35,517.60 27,046.80 34,858.80

As seen from Table 4, if the conventional AC system

is used to meet the cooling load demand of the

greenhouse, it would cost about RM 84,697 to operate

it in a year. It can also be seen that the operating cost

is significantly reduced when TES systems are

employed to meet the cooling load demand of the

greenhouse. Although the TES full storage system has

the highest of chiller charging capacity of 121.5 kW,

the system actually has the lowest annual operating

cost of about RM 27,047 compared with the other two

arrangements. This is because the system used

electricity only for charging and this is done during

the off-peak hours when the electrical tariff is low.

When the system is used to meet the cooling load

demand of the greenhouse, no electricity is consumed

by the system. Although the TES partial storage

system has the lowest chiller charging capacity of 61

kW, the system costs about RM 34,859 to run in a

year, which is much higher than the TES full storage

system. This is because this system does not only

operate during off-peak hour, but also during on-peak

hour, when electricity tariff is high. It is also seen


2012

211

from the table that the combination of TES full storage

system with AC system has a higher operating cost

among all TES systems. In this arrangement, the AC

system is used to meet the cooling load demand of the

greenhouse during off-peak hours, while the chiller is

charging the TES system. However, the AC system is

operating with capacity higher than what is required

by the greenhouse, thus consumes more electricity

than required. Although the off-peak hour electricity

tariff is low, the operating cost is high because the

power consumption of the AC system is generally

high.

Table 5 shows the possible amount annual saving on

the operating cost of the TES systems when compared

to the conventional AC system. It is seen that the TES

full storage system can give the highest cost saving of

about 68 % compared to the conventional AC system.

The TES partial storage system offers the second

highest cost saving of about 59 %, followed by the

TES full storage combined with the AC system, which

offers a cost saving of 58 %.

Table 5: Annual saving on the operating cost of the cooling

systems using TES compared to the conventional AC system.

System Saving (RM) Percentage (%)

TES Full Storage With

AC System 50,518.80 58.06

TES Full Storage 58,989.60 68.07

TES Partial Storage 51,177.60 58.84

4. CONCLUSION

This study investigates potential saving in electricity

cost to operate a hypothetical greenhouse for planting

strawberry in Johor Bahru, Malaysia. It was found that

if a conventional air-conditioning (AC) system is used

to meet the cooling load demand of the greenhouse the

annual operating cost in term of electricity usage

would amount to RM 84,697, based on the local off-

peak and on-peak hour electricity tariff. The use of

thermal energy storage (TES) system has a potential to

help save the operating cost of the greenhouse. Three

operating arrangements have been considered:

combine TES full storage with AC system, TES full

storage system and TES partial storage system.

Among these three arrangements, the TES full storage

was found to have the highest an annual cost saving of

about RM 58,990 compared to the cost of using the

conventional AC system alone. This represents about

68 % of annual operating cost saving, which is

considered very significant.

ACKNOWLEDGEMENT

The authors would like to acknowledge the supports

from Universiti Teknologi Malaysia and fund

provided by the Ministry of Higher Education,

Malaysia throughout this study under the ERGS Vot

No. 4L404.

REFERENCES

[1] Rakesh Khanal & Chengwang Lei, Solar chimney - A

passive strategy for natural ventilation, Energy and

Buildings 43 (2011) 1811–1819.

[2] P.F. Linden, The fluid mechanics of natural ventilation,

Annual Review on Fluid Mechanics 31 (1999) 201–

238.

[3] K.-S. Nikas, N. Nikolopoulos, & A. Nikolopoulos,

Numerical study of a naturally cross-ventilated

building, Energy and Buildings 42 (2010) 422–434.

[4] N.K. Bansal, R. Mathur, M.S. Bhandari, Solar chimney

for enhanced stack ventilation, Building and

Environment 28 (3) (1993) 373–377.

[5] G. Gan, Simulation of buoyancy-induced flow in open

cavities for natural ventilation, Energy and Buildings 38

(5) (2006) 410–420.

[6] D.J. Harris, N. Helwig, Solar chimney and building

ventilation, Applied Energy 84 (2) (2007) 135–146.

[7] A. Dimoudi, Solar chimneys in buildings – the state of

the art, Advances in Building Energy Research 3

(2009) 21–44.

[8] Pacific Northwest Laboratory, Thermal energy storage

for space cooling. Federal Energy Management

Program (FEMP), Federal Technology Alert (FTA),

Richland, Washington, 2000.

[9] Anita Sønsteby and Ola M. Heide (2006). Dormancy

Relations and Flowering of the Strawberry Cultivars

Korona and Elsanta as Influenced by Photoperiod and

Temperature. Scientia Horticulturae. Volume 110, Issue

1: 57-67.

[10] Mohd Yusoff Senawi (2000). Development of a

Building Energy Analysis Package and its Application

to Analysis of Cool Thermal Energy Storage Systems.

Ph.D. Thesis. Universiti Teknologi Malaysia.


2012

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Transient Modeling of a Lithium Bromide – Water Absorption

Chiller

Ang Lia, Wai Soong Loh

a, Kim Choon Ng

b



9 Engineering Drive 1, Singapore 117576,

bProfessor,

Department of Mechanical Engineering,


9 Engineering Drive 1, Singapore 117576,

[email protected]

ABSTRACT

This article presents a thermodynamic framework

for a lithium bromide – water absorption chiller, in

which a transient model is developed to simulate

the operation process. Local energy and mass

balance within the main components like absorber,

regenerator, condenser, evaporator and solution

heat exchanger is respected to investigate the

behavior of the chiller. Experimental correlations

are used to predict heat transfer of the related

working fluids. The cooling water is set to typical

cooling tower conditions of tropical countries such

as Singapore. The coefficient of performance

(COP) is evaluated against a range of heat source

temperatures from 75oC to 100

oC. The results

indicate the operation conditions of the chiller at its

maximum COP is 95oC to 100

oC.

Keywords : Absorption chiller, lithium

bromide, entropy, modeling

1. INTRODUCTION

Low-grade heat driven and maintenance free are

two attractive features of absorption chillers. As

such, they are favored to produce cooling in

cogeneration and solar systems. This article

presents a thermodynamic framework for a lithium

bromide – water absorption chiller. A transient

model is developed to simulate the operation

process of main components like absorber,

regenerator, condenser, evaporator and solution

heat exchanger. The effect of heat source

temperature to the performance of the chiller is

evaluated.

2. MODELING OF ABSORPTION

COOLING CYCLE

2.1. Description of Operation Process

The transient modeling developed in this work is

based on YAZAKI WFC-900S single effect lithium

bromide – water absorption chiller. A schematic

presentation of the chiller is shown in Figure 1. The

chiller has a rated cooling capacity of 3 tons of

refrigeration (rton). It mainly comprises an

absorber, a regenerator, a condenser, an evaporator

and a solution heat exchanger. The first two

components are the reactors where the absorbent

and absorbate interact to perform thermal

compression. In the absorber, the water, also

known as the absorbate which plays a role of

refrigerant, is attracted by the concentrated LiBr

solution that holds a strong affinity to the former.

This absorption process produces heat as the water

changes its state from gaseous phase to absorbed

phase, and the solution becomes diluted. To keep

the absorption ability of the solution, cooling water

is supplied to carry away the heat produced and

maintain the absorber temperature. In the

regenerator, heat source is provided; the diluted

solution that comes from the absorber is warmed

up and saturated. Water molecules are released,

leading to an increase in the concentration of the

lithium bromide solution. This concentrated

solution flows back to the absorber, and the

absorption / desorption process continues. The

solution heat exchanger is located in between the

two reactors to effectively reuse the energy of two

solution streams, whereby the diluted flow is

preheated prior entering the generator and the

concentrated solution is pre-cooled before going

into the absorber. The water vapor released from

the saturation process of the regenerator flows to

the condenser and condenses to liquid phase. The

cooling water that branched out from the absorber

cooling water is used to remove the heat generated

in the condensation process. The cooling effect of

the chiller is provided by the condensate flowing

into the evaporator where the liquid vaporizes. The

evaporation process lowers down the incoming

chilled water which is circulated between the load

and the chiller. A U-tube is applied to connect the

condenser and the evaporator, and maintain the

pressure difference of the two components. The

absorption cooling cycle is completed when the

refrigerant vapor is absorbed by the concentrated

lithium bromide solution in the absorber, and next

IMAT-UI 035



2012

213

cycle of cooling starts. As the water is used as the

refrigerant, the chiller components operate in sub-

atmospheric conditions.

Figure 1: Schematic presentation of a single-effect

LiBr-water absorption chiller

2.2. Transient Modeling

The transient model developed below simulates the

operation process of the absorption chiller. The key

of the model are the energy and mass balance

equations of each major component which are

differential form and to be solved simultaneously

with respect to time. This computation is

performed in FORTRAN 90 Developer Studio

software with the help of DIVPAG subroutine of

IMSL Library. The equations are solved iteratively

by Gear‘s BDF method, employing the starting

conditions of the chiller components to initialize

the computation. The results that converge in the

tolerance of 10-6

are accepted. Properties of the

lithium bromide solution are calculated using Yuan

and Herold [1] correlation functions except the

vapor pressure for which McNeely [2] correlation

is used. The IAPWS Formulations [3-5] are

implemented for the properties of water. Prior to

model the system, some assumptions are made. 1)

Each component of the chiller is properly insulated.

Heat loss by conduction to the insulation and

radiation to the surrounding are ignored. 2) The

heat exchanger material inside each component has

the same temperature as the content of the

component. 3) The content of the two reactors are

well mixed.

2.2.1. Absorber

The thermodynamic properties of the lithium

bromide solution are dependent on temperature,

pressure as well as concentration. Defining as the

mass ratio of the salt to the solution, the

concentration, X, is expressed as follows:

, /

/

/

100%LiBr ds cs

ds cs

ds cs

mX

m (1)

where the subscription ds and cs denote diluted

solution and concentrated solution, respectively.

In the absorber, water vapor is continuously

absorbed by the solution. To achieve this, the

solution must remain sub-cooled throughout the

process and even upon leaving the absorber [6].

Choose a control volume to be the space taken up

by the solution and heat exchanger in the absorber,

the energy and mass conservation of the absorption

process is given as:

,

,

, ,

ab

hx sol ab

cs sol cs ab cs ds sol ab ds

ab

o i ve g e losscw ab

dTMCp MCp

dt

m h T X m h T X

Cp T T m h T Qm

(2)

And,

ds cs vem m m (3)

In Equation (2), the term on the left hand side

represents the sensible heat required to change the

temperature of the heat exchanger material as well

as the solution content in the absorber chamber. On

the other side, the first two terms denotes the

energy that is brought in or taken out by the

concentrated or diluted solution, respectively. The

third term is the heat removal by the cooling water,

whereas vapor energy flowing into the absorber is

given by the last term.

2.2.2. Regenerator

Attributing to the heat source, the content solution

in the regenerator boils in high temperature and

pressure at which the water vapor molecules are

released. As a result, both temperature and pressure

of the vapor are boosted as compared to its state

upon leaving the evaporator. Another use of vapor

is that it helps to pump the solution out of the

generator. This is achieved by configuring the

generator tube inner diameter to be approximately

equal to the size of bubbles. The mixture of the

bubble and liquid forms the slug flow regimen in

which each bubble lifts a small amount of liquid

upwards. A vapor liquid separator is place on top

of the generator tubes to split the two phases. Such

configuration is known as the ‗air bubble pump‘[7].

Similar to the absorber, considering the heat

exchanger and the solution enclosed space in the

regenerator as the control volume, the energy

balance and mass conservation equation can be

expressed as:


2012

214

,, ,

,

rg

hx sol rg

ds sol ds rg ds cs sol rg cs

rg

o i vc g rg c losshw

dTMCp MCp

dt

m h T X m h T X

mCp T T m h T P Q

(4)

And,

cs ds vcm m m (5)

The amount of water vapor desorbed from the

lithium bromide solution can be obtained from the

equation below that is modified from a correlation

given by Chua et al. [8]. The modification takes

into account the influence of regenerator pressure

to the desorption rate.

w w

w rg rg

rg rg

m mdm dT dP

T P

(6)

Where,

And,

Where θrg is 0 if the solution is in sub-cooled state,

and otherwise, it assumes to be 1. Tdp represents the

dew point temperature, and AD and BD denotes

Duhring constants founded in McNeely [2]

correlation.

2.2.3. Condenser

In the condenser, thermal phenomena involved are

phase change of water and heat transfer to the

cooling water. The conservation of energy can be

described as:

,

,c

vc g rg chx f c

c

fc f c o i losscw c

dTMCp MCp m h T P

dt

m h T mCp T T Q

(7)

The terms on the left hand side of the equation

again represent the sensible heat required to change

the temperature of the heat exchanger material as

well as the liquid content remained in the

condenser chamber. The first two terms on the right

hand side, on the other hand, denotes the energy of

vapor incoming to the condenser and energy of

liquid leaving the chamber. The last term of the

equation gives the heat removal by the cooling

water.

2.2.4. Evaporator

Assume the control volume of the evaporator is the

space enclosed by the U-tube that connects the

condenser with the evaporator, the evaporator heat

exchanger and liquid water. Similar to the

condenser, the evaporator can be modeled as:

e

fc f chx f e

e

ve g e i o losschi

dTMCp MCp m h T

dt

m h T Cp T T Qm

(8)

2.2.5. Solution Heat Exchanger

In the solution heat exchanger, the decrease of

thermal energy of high temperature concentrated

solution is the increase of the low temperature

diluted solution. It can be modeling as following.

, ,rg cs ab ds rg abcs dsCp T T Cp T Tm m (9)

The above equation is solved by ε-NTU method to

determine the heat exchanger output, Tcs,ab and

Tds,rg, with Trg and Tab simultaneously solved by

Equation (2) and (4) as inputs.

2.2.6. External water sources

The above energy balance and mass conservation

equations are solved with coupling to the

calculation of external water sources. For hot

water, cooling water and chilled water of respective

component, the outlet temperature is determined

from UA-LMTD method, and is given by:

, / / , / / / / / , / /

/ / /

/ /

1 exp

o hw cw chi i hw cw chi rg ab c e i hw cw chi

rg ab c e

hw cw chi

T T T T

UA

mCp

(10)

The overall heat transfer coefficient of regenerator,

absorber, condenser and evaporator heat exchanger

U is discussed in the following section.

2.2.7. Heat transfer coefficients

The overall heat transfer coefficient encountered in

the modeling of each component, like the absorber,

regenerator, condenser and evaporator are

determined by the following equation:

1

, ,

ln1

2

i o i i o

t i t o

D D D D DU

h k h

(11)

The local heat transfer coefficient of the tube side,

ht,i, where external water sources for each of the

component heat exchanger, such as cooling water

2

100

- 273.15

w LiBr

rg ds

rg

D D

dp rg

ds ds

m m

T X

dA dBT P

dX dX

(6a

)

2

100

273.15

w LiBr D

rgrg ds

dp

rg

D D

dp rg

ds ds

m m A

dPP X

dT

dA dBT P

dX dX

(6b)


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for the absorber and condenser, hot water for the

regenerator as well as chilled water for the

evaporator is determined by Dittus-Boelter

correlation for pipe water. For the shall side local

heat transfer coefficients, ht,o, the following

correlations are used. 1) The absorber is

constructed in a way that the lithium bromide

solution falls from the top of the tube bundles. Park

[9] correlation is applied in this case. 2) Pooling

boiling of the LiBr solution is taking place in the

regenerator where Charters [10] experimental

results are used to predict the local heat transfer

coefficient. 3) Nusselt condensation correlation for

horizontal tube bundles counts for the phase change

phenomenon in the condenser. 4) Assuming the

salinity to be zero, the local heat transfer

coefficient of refrigerant water in the evaporator is

calculated by extrapolating from Shahzad‘s [11]

correlation for sub-atmospheric pressures. The heat

transfer area and key simulation parameters are

given in Table 1.

Table 1: Key simulation parameters.

Heat transfer area

Aab 2.98 m2 Ae 1.95 m

2

Arg 2.95 m2 Ac 0.96 m

2

(UA)hx 0.734 kW/k

Flow rate of external water sources

mhw

0.68 kg/s ,

mcw c

0.54 kg/s

,

mcw ab

0.79 kg/s mchi

0.5 kg/s

Initial conditions

Trg,ini = Ti,hw Tcw,ab,ini = Ti,cw

Te,ini = Ti,chi Tcw,c,ini = Ti,cw

2.2.8. Evaluation parameters

The absorption chiller is evaluated by its cooling

capacity QC and the Coefficient of Performance

(COP). Both parameters are defined as follows:

C i o chi

Q Cp T Tm (12)

C

H

QCOP

Q

(13)

Where QH is the total heat input to the system and

calculated from the hot water behavior.

H i o hw

Q Cp T Tm (14)

4. RESULTS AND DISCUSSION Figure 2 shows simulated temperature profile of

chiller major components and water sources during

operation at hot water inlet temperature Ti,hw 90oC,

cooling water inlet temperature Ti,cw 29.5oC and

chilled water requirement To,chi of 9oC. The heat

loss to the surrounding encountered in regenerator

and evaporator is assumed to be 10% of heat input

and cooling effect, respectively. The same quantity

for absorber and condenser is set at 2% of their

total heat rejection. The absorption chiller

experiences a starting period and stabilizes

afterwards. In the formar period, regenerator

content is sub-cooled with no water molecules

being desorbed. Cooling takes into effect when

saturation of concentrated LiBr solution is reached.

In addition, the largest temperature difference of

chiller components and respective external water

sources happens in the regenerator. A deviation of

8.6 o

C is found between regenerator temperature

and outlet temperature of heat source. This is well

explained by the inefficient boiling heat transfer of

lithium bromide solution.

Figure 2: Simulated temperature profile of

chiller components and water

sources during operation at hot water

inlet temperature Ti,hw 90oC, cooling

water inlet temperature Ti,cw 29.5oC

and chilled water requirement To,chi

of 9oC

The effect of heat source inlet temperature Ti,hw to

cooling capacity and overall thermal input is

displayed in Figure 3. The results are calculated at

cooling water inlet temperature Ti,cw 29.5oC and

chilled water requirement To,chi of 9oC. Both

computed quantities unveils linear upward trend

with respect to the temperature of hot water input,

while the amount of thermal input increases faster.

The phenomena are due to the fact that the raising

of heat source temperature results in an increase of

temperature of regenerator content and hence the

amount of refrigerant water released. The addition

of quantity in the refrigerant circulation promotes

the need to the thermal energy, and delivers more

cooling at the same time.

Figure 3: Effect of heat source temperature to

cooling capacity and thermal input at

cooling water inlet temperature Ti,cw

29.5oC and chilled water requirement

To,chi of 9oC


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Despite a slightly faster increase of thermal input

than the cooling effect with respect to heat source

temperature, it is still wealth of operating the

chiller at the higher end. As implied from Figure 4,

in conditions as above, the chiller‘s coefficient of

performance is improved by higher heat source

temperature. The COP reaches its maximum 0.64 at

95oC to 100

oC, while at lower end of 75

oC the

useful effect is only equivalent to 48% of the total

input. The relation, however, is not linear. The

potential of improvement drops as hot water

temperature increases. The results indicate that at

typical cooling tower conditions (29.5oC) of

tropical countries, the higher end heat source

temperature produces cooling most efficiently for

the lithium bromide – water absorption chiller

presented in this work.

Figure 4: Effect of heat source temperature to the

coefficient of performance (COP) at

cooling water inlet temperature Ti,cw

29.5oC and chilled water requirement

To,chi of 9oC.

5. CONCLUSION

In the current work, a transient model was

developed for a lithium bromide – water absorption

chiller to simulate its operation process. The

cooling capacity, heat input and the COP was

evaluated with respect to heat source temperature.

The results shows that at typical cooling tower

conditions of tropical countries, the chiller operates

at its best COP the hot water temperature of 95oC

to 100oC.

NOMENCLATURE Alphabets Description

A Area, m2

AD Duhring constant

BD Duhring constant

COP Coefficient of performance

Cp Specific heat capacity, kJ/kg.K

D Diameter, m

h Specific enthalpy, kJ/kg

ht Local heat transfer coefficient,

kW/m2.K

k Thermal conductivity, kW/m.K

LiBr Lithium bromide solution

M Mass, kg

m Mass flow rate, kg/s

P Pressure, Pa

Qh Total heat input, kW

Qc Cooling capacity, kW

rton Ton of refrigeration

T Temperature, K

t Time, s

U Overall heat transfer coefficient,

kW/m2.K

X Concentration, wt%

θ State indicator in Eqn (6a) and (6b)

Subscripts Description

ab Absorber

c Condenser

chi Chilled water

cs Concentrated solution

cw Cooling water

dp Dew point

ds Diluted solution

e Evaporator

hw Hot water

i Inlet

ini Initial conditions

loss Heat dissipation

o Outlet

rg Regenerator

sol Lithium bromide solution

REFERENCES

[1] Z. Yuan and K. E. Herold, "Thermodynamic

properties of aqueous lithium bromide using a

multiproperty free energy correlation," HVAC and R

Research, vol. 11, pp. 377-393, 2005.

[2] L. A. McNeely, "Thermodynamic properties of

aqueous solutions of lithium bromide," ASHRAE

Trans vol. 85, pp. 413-434, 1979.

[3] W. Wagner, J. R. Cooper, A. Dittmann, J. Kijima,

H. J. Kretzschmar, A. Kruse, R. Mareš, K. Oguchi,

H. Sato, I. Stöcker, O. Šifner, Y. Takaishi, I.

Tanishita, J. Trübenbach, and T. Willkommen, "The

IAPWS industrial formulation 1997 for the

thermodynamic properties of water and steam,"

Journal of Engineering for Gas Turbines and

Power, vol. 122, pp. 150-180, 2000.

[4] IAPWS. Release on the IAPWS Formulation 2008

for the Viscosity of Ordinary Water Substance

[Online]. Available: http://www.iapws.org

[5] IAPWS. Release on the IAPWS Formulation 2011

for the Thermodynamic Conductivity of Ordinar

Water Substance [Online]. Available:

http://www.iapws.org

[6] M. J. Kirby and H. Perez-Blanco, "Design model for

horizontal tube water/lithium bromide absorbers,"

1994, pp. 1-10.

[7] H.I. Abu-Mulaweh, D.W.Mueller, B.Wegmann,

K.Speith, and B. Beohne, "Design of a Bubble

http://www.iapws.org/

http://www.iapws.org/


2012

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Pump Cooling System Demonstration Unit," Int. J.

of Thermal & Environmental Engineering, vol. 2,

pp. 1-8, 2011.

[8] H. T. Chua, H. K. Toh, A. Malek, K. C. Ng, and K.

Srinivasan, "A general thermodynamic framework

for understanding the behaviour of absorption

chillers," International Journal of Refrigeration,

vol. 23, pp. 491-507, 2000.

[9] C. W. Park, S. S. Kim, H. C. Cho, and Y. T. Kang,

"Experimental correlation of falling film absorption

heat transfer in micro-scale hatched tubes,"

International Journal of Refrigeration, vol. 26, pp.

758-763, 2003.

[10] W. W. S. Charters, V. R. Megler, W. D. Chen, and

Y. F. Wang, "Atmospheric and sub-atmospheric

boiling of H2O and LiBr/H2O solutions,"

International Journal of Refrigeration, vol. 5, pp.

107-114, 1982.

[11] M. W. Shahzad, A. Myat, W. G. Chun, and K. C.

Ng, "Bubble-assisted film evaporation correlation

for saline water at sub-atmospheric pressures in

horizontal-tube evaporator," Applied Thermal

Engineering, vol. 50, pp. 670-676, 2013.


2012

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Characteristics of Sea-water Ice Slurry for Cooling of Fish

A.S. Pamitrana, M. Novviali

b, H.D. Ardiansyah

b


Universitas Indonesia, Kampus UI Depok 16424, Indonesia

Tel : (021) 7270032, Fax : (021) 7270033

E-mail : [email protected] bGraduate Student, Department of Mechanical Engineering,


ABSTRACT

A more effective of cooling method is necessary

for fish storage to get high quality and long

freshness of fish. Ice block is not sufficient for fish

storage because of its hard-solid surface that can

damage the fish. Moreover for some remote area it

is difficult to find ice block in good time with

reasonable/low price. One solution for this problem

is the using of sea-water ice slurry for fish cooling.

Ice slurry is formed when the sea-water

temperature goes down to its freezing point, when

the early nucleation is formed. At this moment

there is a chemical potential againts the saturation

condition. Crystal ice can be formed when

chemical equilibrium is occured. The purpose of

this present study is to observe the characteristics

of ice slurry generation using scraper blade

evaportor and orbital rod evaporator. The

experiment is done under some experimental

conditions.

Keywords : Ice slurry, cooling, sea-water, salinity,

evaporator

1. INTRODUCTION Indonesia is an archipelago nation and is located at

tropical area where many kind fish species come

and growth. It has total area of Exclusive Economic

Zone of around 3 million km2, coastal line length of

around 104,000 km, number of fisherman of

around 1.7 million people who can supply fish of

around 9 million Ton in 2010. However, in fact this

figure is not linear with the fisherman welfare. This

situation could be caused by limited quality and

quantity of equipments used for fishing and

storaging the fish. Most of Indonesian fishermans

still use block ice for fish storaging. This way is not

sufficient for the fish because the block ice can

damage the fish and then the value of fish become

lower. Moreover for some remote area in Indonesia

it is difficult to find block ice with reasonable

price. Using outboard engine boat, or using no-

engine boat, traditional fisherman can only do

fishing around coatline, and can only effectively

work around 7-9 months a year. This situation

makes low productivity in fishing. One effective

way for cooling is replacing the ice block cooling

by ice slurry. Because sea-water ice slurry content

natural preservative, it is good for fish storaging

after fishing. This study is devoted to observe the

characteristics of sea-water ice slurry under

experimental conditions variation of salinity, room

tempature, shaft rpm, and sea-water volume in two

developed ice slurry generator types of scraper

blades evaporator and orbital rod evaporator.

E. Stamatiou et al., 2005, defines ice slurry consists

of water solution and ice crystal. Another definition

by Peter W Egolf et al., 2003, ice slurry is an ice

particle with average diameter equal as or less than

1 mm. Ice slurry forming consist of three steps of

supersaturation, nucleation, and propagation.

Supersaturation is when the freezing point

temperature of the fluid is reached. Lower than the

freezing temperature, nucleation is then formed.

Propagation is the phase when the ice crystal

formed.

2. EXPERIMENT AND PROCEDURE

The experiment is done in Refrigeration

Laboratory, Faculty of Engineering University of

Indonesia. As well, under Community Social

Responsibility project supported by the University

of Indonesia, one developed ice slurry generator is

implemented in fisherman community, in

Balongan, Indramayu, Indonesia. The experimental

apparatus is illustrated schematically in Figure 1. It

consists of two systems viz. ice slurry generator

system and cooling system. In ice slurry generator,

sea-water rejects some heat to refrigerant in the

evaporator to generate ice slurry. The present study

develops two kinds of evaporator viz. orbital rod

evaporator and scraper blade evaporator. The

cooling system consists of compressor, condenser,

liquid receiver, filter dryer, sight glass, expansion

valve, and accumulator. The developed test

apparatus is shown in Figure 2.

Ice slurry generator with orbital rod evaporator has

a rod scrapper inside the evaporator. The rod

rotates on the shaft line to scrap the formed ice

slurry on the evaporator inner surface. The present

ice slurry generator with orbital rod evaporator has

63.5 mm inner diameter, 76.2 mm outer diameter,

and 1500 mm length. Figure 3 illustrates developed

shaft of the orbital rod evaporator. Many industries

use scraper blade evaporator because it produces

more ice fraction than others system, as mentioned

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Figure 1: Schematic experimental apparatus

Figure 2: Developed ice slurry generators

Figure 3: Shaft of the orbital rod evaporator (mm) Figure 4: Shaft of the scraper blade evaporator (mm)


2012

221

by E. Stamatioua et al. 2005 and T. A. Mouneer et

al.2011. The present developed ice slurry generator

with scraper blade evaporator has 2600 mm

evaporator diameter, 2800 mm total length of

evaporator pipe with 9.525 mm evaporator pipe

diameter. Shaft of the scraper blade evaporator is

illustrated in Figure 4.

Temperature of the ice slurry is measured by

thermocouples and recorded by data aquisition.

Some properties of ice, as follow, are calculated by

referring A. Melinder, 2010.

(1)

(2)

(3)

The ice fraction is calculated by referring Jean-

Pierre Be´de´carrats et al., 2009, and Cecilia Hägg,

2005.

(4)

D.G. Thomas, 1965, and Jacques Guilpart et al.,

2006 use the following equation to calculate

viscosity.

(5)

Enthalpy is calculated by referring T. Kousksou et

al., 2010.

(6)

By referring Taret, 1940, the thermal conductivity

of ice slurry is calculated.

(7)

Parameter of COP (Coefficient of Performance) is

obtained using the following equation.

(8)

The experiment was runned under experimental

condition shown in Table 1.

3. RESULTS AND ANALYSIS

Figure 5 shows ice slurry formation for salinity 26

ppt. The temperature decreases in supersaturation

process from point A to point B. Ice fraction is

initially formed at point B. The freezing point of

sea-water is lower than of water due to its higher

salinity. Referring Melinder et al., 2008, point B is

starting point of nucleation. Ice slurry is initially

Table 1: Experimental condition

Salinity 18, 22, 26, 30 ppt

Shaft rpm 70, 90, 110, 130 rpm

Room temperature 22, 26, 30, 34ºC

Volume 3, 3.5, 4, 4.5 liter

Figure 5: Ice slurry formation

Figure 6: Effect of salinity on ice slurry

formation

formed when the temperature reach freezing point.

Water in the solution is partially freezed, forming

ice fraction.

The effect of salinity on ice slurry formation for

evaporative heat of 2.947 kW is depected in Figure

6. The salinity of sea-water in Indonesia is around

30 ppt. The results show that higher salinity needs

longer time and lower temperature for ice fraction

formation. It means that sea-water with higher

salinity takes more energy for ice slurry formation.

As well, it means that sea-water with higher

salinity can absorb more heat from fish at low

temperature. Figure 6 shows freezing temperatures

for salinities of 18, 22, 26 and 30 ppt are -1.29ºC, -

1.588ºC, -1.667ºC and -1.94ºC, respectively.

In order to validate the measurement data of

freezing temperature, the results are compared to a

reference as shown in Figure 7 with error of around

40%.

Diameter of ice fraction is obtained using IMAGEJ

software as shown in Figure 8. The effect of

Time (s)


2012

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salinity on ice fraction diameter is shown in Table

2. At the same experimental condition, higher

Figure 7: Freezing temperature comparison

Figure 8: Diameter of ice fraction

Figure 9: Salinity effect on ice slurry

formation

Table 2: Diameter of ice slurry

Figure 10: Freezing temperature comparison

salinity results in smaller ice fraction diameter. The

effect of salinity on ice slurry formation is depected

in Figure 9, as well.

Figure 10 shows the enthalpy of ice slurry is

summation of ice enthalpy and sea-water enthalpy.

For salinity of 18 ppt, the ice slurry enthalpy is low

because it has more ice fraction than for the higher

salinity. The effect of cooling capacity on enthalpy

is shown in Figure 10, as well. The enthalpy is

higher for the lower cooling capacity due to

existing of ice fraction. Higher ice fraction may

results in higher pressure drop in ice slurry flow,

therefore it should be avoided.

4. CONCLUSION

The freezing point of sea-water is lower than of

water due to its higher salinity. Higher salinity

needs longer time and lower temperature for ice

fraction formation. Higher salinity results in

smaller ice fraction diameter and higher enthalpy.

ACKNOWLEDGMENT

The work described in this paper was supported by

grants of Hibah Madya 2012 from DRPM

Universitas Indonesia and Hibah Pengabdian

Masyarakat (Community Engagement Grant) 2012

from DRPM Universitas Indonesia.

REFERENCES

[1] E. Stamatioua, J.W. Meewisseb, M.

Kawajia.2004. Ice slurry generation involving

moving parts.International Journal of

Refrigeration 28 (2005) 60–72

[2] Peter W. Egolf, Michael Kauffeld. 2004. From

physical properties of ice slurries to industrial

ice slurry applications.International journal of

refrigeration 33 (2010) 1491-1505

[3] A°. Melinder*, Properties and other aspects of

aqueous solutions used for single phase and ice

Salinity (ppt)

QEvap 1.718 kW

QEvap 2.947 kW

Cooling time 1 hour 3 minutes

Salinity (ppt)

Ice

fra

ctio

n (

%)

*Cooling time 48 minutes

Cooling time 1 hour 3 minutes

Salinity (ppt)

En

tha

lpy

(k

J/k

g)


2012

223

slurry applications, international journal of

refrigeration 33 (2010 ) 1506e1512

[4] T.A. Mouneer *, M.S. El-Morsi, M.A. Nosier,

N.A. Mahmoud, Heat transfer performance of

a newly developed ice slurry generator: A

comparative study

[5] Jean-Pierre, Thermal and hydrodynamic

considerations of ice slurry in heat exchangers,

Be´de´carrats*, Franc¸oise Strub, Christophe

Peuvrel

[6] Cecilia Hägg ,2005,Ice Slurry as Secondary

Fluid in Refrigeration Systems, Fundamentals

and Applications in Supermarkets,School of

Industrial Engineering and Management,KTH

[7] D.G. Thomas, Transport characteristics of

suspension. VIII. A note on the viscosity of

Newtonian suspensions of uniform spherical

particles, Journal of Colloid Science 20 (1965)

267–277.

[8] Jacques Guilpart*,1, Evangelos Stamatiou,

Anthony Delahaye, Laurence Fournaison

Comparison of the performance of different ice

slurry types depending on the application

temperature, International Journal of



2012

224

FLUID FLOW CHARACTERISTIC OF ROUNDED-SHAPE FPSO AND

LNG CARRIER DURING OFFLOADING

Mufti F. M., Jaswar, A. Priyanto, and Efi Afrizal

Department of Marine Technology, Faculty of Mechanical Engineering

Universiti Teknologi Malaysia

Skudai, 81310 Johor, E-mail : [email protected]

ABSTRACT

The design concept of FPSO begins with a shape like a ship.

Nowadays some researcher proposed a different shape for

design concept of FPSO called rounded-shape. The rounded-

shape has advantage in the seakeeping and the construction.

One of the common activity performed by the FPSO is

transfer process of the product which is called offloading

process. There are a wide range of phenomena that occur

during offloading one of which is fluid flow. Fluid flow that

occurs in structure will be different depend on the several

factors, such as shape of structure, Reynolds numbers and

Froude numbers. From the characteristic of fluid flow that

occurs in the structure can be determined the effect of the

fluid flow on the structure. Current research discuss the

concept design of fluid flow around a round-shaped FPSO

LNG interacted with a LNG carrier during side-by-side

offloading condition by CFD method based on the Reynolds

Average Navier-Stokes (RANS) equations.

Keywords: Fluid flow; Offloading; RANS; CFD.

1. INTRODUCTION

The development of oil and gas exploration industry,

particularly exploration in the deep ocean, in recent decades

has increased. Correspondingly with the increase in oil and

gas exploration in deep ocean waters, facilities and

infrastructures required to support exploration activities are

also necessary. One of the supporting facilities in the

activities of oil and gas exploration in deep water is the FPSO

LNG (Floating Production Storage Offloading Liquid Natural

Gas) and LNG Carriers. The FPSO LNG is enabled to

receive and process gas products from the field and save the

LNG into Cargo Containment System (CCS) tank before the

LNG transferred to the LNG Carrier to be distributed to the

market or destination.

Development of design concepts to the FPSO start with

a shape like a ship. The FPSO is designed to accommodate

the construction of the module production process to be the

product oil. The new concept design of the FPSO is currently

a cylinder shaped being developed by Sevan Marine [1] and

SSP Offshore [2]. Wang, Zhang, and Liu [3] study new

design concept of FPSO which is propose a non-ship-shape

FPSO called inverted fillet quadrangular frustum pyramid-

shaped FPSO (IQFP). Another design concept of circular

FPSO for Arctic Deepwater proposed by Srinivasan and

Sreedhar [4].

LNG products that have been produced on the FPSO LNG

will be transferred to the LNG Carrier. The transfer process is

called offloading process, which is a common activity

performed by the FPSO LNG and the LNG Carrier. There are

two methods of offloading process for oil and gas transfer

which is tandem and side-by-side. Tandem configuration is

done when a LNG carrier is moored in tandem with the

FPSO LNG. The hoses or hawsers are connected between the

stern off-loading stations on FPSO LNG to the cross over

manifold of the LNG tanker. While, the side-by-side

configuration is done by moored the LNG Carrier parallel

with the FPSO LNG and offloading is carried out via a

flexible hose between the cross over manifold of the FPSO

LNG and LNG Carrier as shown in Figure 1.

Figure 1. Side-by-side configuration [1]

Environmental loads generally occur in the structure at time

of FPSO LNG and a LNG tanker offloading conditions.

These environmental loads are classified into three types

which area wind, wave and currents load. Wind load exerts a

force on the part of the structure exposed to the air. Wave

load and current load exerts force on the part of the structure

exposed to the water. These three loads will determine

the external forces on structures, stability, and motion of

floating structures and patterns of fluid flow.

Fluid flow around a structure can significantly alter the

structure‘s loading characteristics. Waves and currents is a

fluid that has the most impact on the structure. Influence of

fluid flow generated from waves and currents on floating

structure is seen in the form of motion and

pressure distribution. These motion and pressure distribution

that occurs in the structure depend on the characteristic of the

IMAT-UI 037



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fluid flow. Characteristic of the fluid flow is affected by

several factors such as shape of the structure and Froude

number.

The study of fluid flow around hull in side-by-side offloading

condition have been performed by Arslan, Pettersen, and

Andersson [5]. The calculations of three dimensional (3D)

unsteady cross flow past a pair of ship sections in close

proximity and behavior of the vortex-shedding around the

two bluff bodies is investigated numerically by using the

software FLUENT.

The studies of fluid flow around hull of rounded-shape FPSO

has been performed by Lamport and Josefsson [6]. The

studies is shows the current induced velocity fields on the

leeward side of the vessel between round-shaped versus a

traditional ship-shaped FPSO.

Current research is discuss on analysis of flow around a

round-shaped FPSO LNG interacted with a LNG carrier

during offloading. The fluid flow characteristic around the

round-shaped FPSO LNG and LNG carrier during offloading

conducted using Computational Fluid Dynamics (CFD) based

on the governing Reynolds Average Navier-Stokes (RANS)

equations.

2. LITERATURE REVIEW

Studies of the fluid flow on the hull of ship have been

performed by many researchers in the last decade. For the

viscous flow around ship hull field, in 2002 Zhang, Zhao, and

Li [7] performed numerical simulation free surface flow

around Wigley hull. Numerical simulation using RANS

method and SST k-w turbulence model was performed by

Zhao, F., Zhang, Z.-R [8] with a complex modern ship model

DTMB 5415. In 2005, Schweighofer et. al. [9] focuses on the

applicability of different RANS methods to full-scale viscous

flow computations. Alexe [10] study about the effects of

dimensional and movement parameters of the ship on the

pressure distributions in surrounding sea water. Zhao, Zhu,

and Zhang [11] study the flow around Wigley and DTMB

5415 hull using RANS, Visonneau [12] predicting the full-

scale viscous flow field around a ship including the

evaluation of the free surface, the wake field, the

hull/propeller interaction, the resistance and the power.

In 2006, Kinnas, Yu, and Vinayan [13] study the unsteady

viscous flow over the bilge keels of an FPSO hull subject to

roll motions. Wang，Zou，and Tian [14] study the viscous

flow field around a KVLCC2 model moving obliquely in

shallow water using a general purpose computational fluid

dynamics (CFD) package FLUENT. Wang et al [15] study

the wake field in viscous flow and resistance prediction of a

full ship-KVLCC2M by using FLUENT. For coupling the 3D

incompressible RANS equations with level set method was

performed by Wan , Shen, and Ma [16] with numerical

simulation. In 2011, Wackers et. al. [17] reviewed the surface

descretisation methods with different code.

For the turbulence flow around ship hull field, many

researchers have been studied. Kim [18] studies the three-

dimensional turbulent flow using RANS equations. In 2002,

Kim, Kim, and Van Suak [19] developed an efficient and

robust numerical method for turbulent flow calculation.

Ciortan et al [20] investigate the free surface incompressible

turbulent flow around the hull by the numerical solution of

the unsteady Navier-Stokes equations for slightly

compressible flows. Deng, Queutey, and Visonneau [21]

study the simulation of two appended hull configurations

using all hexahedral unstructured grids. The three-

dimensional turbulent flow around a Wigley hull using

slightly compressible flow formulation performed by Ciortan,

Wanderley, and Soares [22]. Lungu [23] presented a

methodology for computing the 3D turbulent free-surface

flow.

Ciortan, Soares, and Wanderley [24] study the turbulent and

laminar free-surface flow around ship hulls using slightly

compressible flow formulation. Ahmed, Fonfach, and Soares

[25] investigated the flow pattern around the DTMB 5415

hull at two speeds. In 2011, Ahmed [26] uses Volume of

Fluid method (VOF) to simulate the flow pattern around the

DTMB 5415 hull at two speeds. Ciortan, Wanderley, and

Soares [27] study the simulation of flow around a Wigley

hull using the slightly compressible flow formulation.

In 2006, Tahara et. al. [28] evaluates the computational fluid

dynamics (CFD) as a tool for hull form design along with

application of state-of-the-art technology in the flow

simulations. Two Reynolds-averaged Navier-Stokes (RANS)

equation solvers were employed, namely CFDShip-Iowa

version 4 and Flowpack version 2004e, for the towing and

self-propulsion cases, respectively. An accurate, efficient

algorithm for solving free surface flows around ship hulls

using a compressive advection discretization which maintains

a sharp free surface interface representation without relying

on a small time step [29].

The application of the Fluent code to the numerical

simulation of the free-surface flow around a model naval

ship; the DTMB 5415. Simulations were performed using

both a structured hexahedral mesh and an unstructured

tetrahedral mesh of lower resolution. The results show that

Fluent is able to accurately simulate the total ship resistance,

near-field wave shapes, and the velocity field in the propeller

plane [30].

Tahara et al [31] conducted research on high-speed multi

hull. Multi hull which is used in that research is catamaran

hull with forward speed. In 2011, Broglia, Zaghi and Di

Mascio [32] study about the simulations of the flow around a

high speed vessel in both catamaran and monohull are carried

out by the numerical solution of the Reynold averaged

Navier–Stokes (RANS) equations.

3. METHODOLOGY

Current research about fluid flow characterization of round-

shaped FPSO LNG and LNG Carrier during offloading. The

steps for the working of this research as follows (Figure 2.):

a) Data Collection

Data collection for environmental condition and dimension of

the rounded-shape FPSO LNG and LNG carrier using a

previous study.

b) Side-by-side Offloading Arrangement

The majority of planned offshore LNG transfer systems are

designed for side-by-side configuration loading procedures,

although offshore transfer operations are limited to

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2012

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significant waves heights of Hs = 3 m [33]. Safety distance

between a rounded-shape FPSO LNG and a LNG carrier in

side-by-side offloading conditions was from 5 [34] to 10

[33] meters.

c) Design RANS Code for CFD

The simulation code created in Fortran software, result of the

calculation exported to Visual Basic for visualization. The

code from the Fortran software and the environmental data

will be used to analyze. From the result of software, then

determine the characteristic of fluid flow around hull of

rounded-shape FPSO LNG and LNG carrier during

offloading. Visualize the fluid flow around hull of rounded-

shape FPSO LNG and LNG carrier during offloding by using

Visual Basic software.

START

Data Collection

RANSE Round-shaped FLNG

and LNG Carrier

Dimension

Input Dimension and

Environmental Condition Data

into RANS Code

Running the Software to get

fluid flow of Round-shaped

FLNG and LNG Carrier

Environmental

Condition

Governing Equations

of RANS

Design Code of RANS

Equation

Analysis the Out Put from

Software of Fluid Flow

Determine the

Characteristic of Fluid

Flow

Finish

Conclusion

Figure 2. Flowchart of Methodology


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4. ROUND-SHAPE FPSO

According to Paik and Thayamballi [35], the first

floating, production, storage and offloading vessel

(FPSO), Shell‘s Castellon, was installed in 1977.

Since the industry has seen a large and diverse suite of

different FPSO solutions, from converted tankers to

purpose-built barge shaped vessels. Until recently,

most FPSOs had one thing in common that the design

philosophy was based on classic ship-shaped vessels.

While a ship has beneficial characteristics for

transporting cargo from one location to another, with

great maneuverability and little water resistance, it‘s

slender and non-ax symmetrical shape presents major

disadvantages when permanently moored in one

location.

One such disadvantage is that ship-shape vessels must

be able to align themselves with predominate sea state

to minimize their motions and vessel stresses. The oil

and gas industry mitigated this problem by developing

turrets and swivels, which allowed the ship shaped

vessels to weathervane into the predominant sea state.

Though swivels and turrets allow ship-shaped vessels

to weathervane, they are costly, have long lead times

and are typically available from only few specialized

designers and fabricators. Swivels and turrets also

have associated maintenance requirements and

potential downtime (from leaking seals, for example).

According to Lamport and Josefsson [6], slender ship-

shapes are subjected to significant bending loads due

to hogging and sagging and, as a result, are subject to

fatigue damage. In the case of converted hulls, the

fatigue problem is exasperated when using hulls built

after 1985 where high tensile strength steel was used

extensively to reduce weight. Ship-shapes are also less

efficient in storage volume per plated area than more

compact shapes of the next generation round-shaped

FPSOs.

To overcome short comings associated with using

traditional ship-shaped vessels for FPSOs, the industry

is now developing fit-for-purpose FPSOs. Unlike

traditional ship shaped FPSOs, which must

weathervane into the predominate sea state to

minimize water resistance and motions, the next

generation FPSOs are being designed to have similar

motion characteristics from all directions and to

eliminate yaw excitation. This eliminates the need for

a costly turret and swivels, minimizes the bending

loads and fatigue and increases the storage capacity

per plated area.

Round-shaped FPSOs also have the advantage of

being more easily approachable by service and

installation vessels with minimum collision risk.

According to Lamport and Josefsson [6] Round-shape

have the several advantages, as follow:

a) More efficient storage shape and the smaller

bending load.

b) The motions are similar from all directions with

little to no yaw excitation.

c) More efficient storage shape and the smaller

bending load.

d) The pie-shaped tanks in round-shaped units create

smaller sloshing forces.

e) Providing additional savings in structural

reinforcement

f) Allows for larger freeboard, decreases the risk of

green water on the deck

g) Simple block construction and repeatable

fabrication.

Figure 3. Design Concept Round-Shape FPSO

5. FLOW AROUND THE SHIP SECTION

Computational fluid dynamics, usually abbreviated

as CFD, is a branch of fluid mechanics that uses

numerical methods and algorithms to solve and

analyze problems that involve fluid flows. Computers

are used to perform the calculations required to

simulate the interaction of liquids and gases with

surfaces defined by boundary conditions. With high-

speed supercomputers, better solutions can be

achieved. In this section discuss the theory of the fluid

flow for CFD simulation.

5.1. Incompressible Potential Flow

Incompressible flow is constant density flow, i.e.

. Visualize a fluid element of fixed

mass moving along a streamline in an incompressible

flow. Because its density is constant, then the volume

of the fluid element is also constant. Determine to

the time rate of change of the volume of a fluid

element, per unit volume. Since the volume is constant

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for a fluid element in incompressible flow, the

equation becomes:

Futhermore, if the fluid element does not rotate as it

moves along the streamline, i.e. if its motion is

translational only, then the flow is called irrotational

flow. For such flow, the velocity can be expressed as

the gradient of a scalar function called the velocity

potential, denoted by .

Combining Eqs. (1) and (2),

or,

Equation (3) is Laplace‘s equation – one of the most

famous and extensively studied equations in

mathematical physics. From Eq. (3), can be determine

that inviscid, irrotational, incompressible flow

(sometimes called potential flow) is governed by

Laplace‘s equation.

5.2. Reynolds-Averaged Navier-Stokes (RANS)

Equation

The non-dimensional RANS equations for unsteady,

three-dimensional incompressible flow can be written

in Cartesian tensor notation as,

where

and are the Cartesian

components of mean and fluctuating velocities,

respectively, normalized by the reference velocity U0,

is the dimensionless

coordinates normalized by a characteristic length ,

is the Reynolds number, is the kinematic

viscosity, the barred quantities

Reynolds stresses normalized by , and is the

upressure normalized by are related

to the corresponding mean rate of strain through an

isotropic eddy viscosity, , i.e.

Where is the turbulent kinetic

energy, Equation (4) becomes

where 1/Rø =1/Re+vt, and ø=Ui (i=1,2,3). Equations

(5) and (7) can be solved for Ui and p when a suitable

turbulence model is employed to calculate the eddy-

viscosity distribution.

5.3. Hess-smith Method

A.M.O. Smith at Douglas Aircraft directed an

incredibly productive aerodynamics development

group in the late ‘50s through the early ‘70s. In this

section we describe the implementation of the theory

given above that originated in his group. *Our

derivation follows Moran‘s description6 of the Hess

and Smith method quite closely. The approach is to i)

break up the surface into straight line segments, i i)

assume the source strength is constant over each line

segment (panel) but has a different value for each

panel, and i i i) the vortex strength is constant and

equal over each panel.

Roughly, think of the constant vortices as adding up to

the circulation to satisfy the Kutta condition. The

sources are required to satisfy flow tangency on the

surface (thickness).

Figure 4 illustrates the representation of a smooth

surface by a series of line segments. The numbering

system starts at the lower surface trailing edge and

proceeds forward, around the leading edge and aft to

the upper surface trailing edge. N+1 points define N

panels.


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Figure 4. Representation of a smooth airfoil with

straight line segments.

The computer program based on the Hess-Smith panel

method (HSPM) approximates the body surface by a

collection of panels and expresses the flow field in

terms of velocity potentials based on sources and

vortices in the presence of an onset flow.

(8)

where, is the total potential function and its three

components are the potentials corresponding to the

free stream, the source distribution, and the vortex

distribution. These last two distributions have

potentially locally varying strengths and ,

where is an arc-length coordinate which spans the

complete surface of the airfoil.

The potentials created by the distribution of

sources/sinks and vortices are given by:

Combining Eqs. (8), (9) and (10)

The potential relation given above in Eq. (4-22) can

then be evaluated by breaking the integral up into

segments along each panel:

Since Eq. (12) involves integrations over each discrete

panel on the surface of the airfoil, we must somehow

parameterize the variation of source and vortex

strength within each of the panels. Since the vortex

strength was considered to be a constant, we only need

worry about the source strength distribution within

each panel.

This is the major approximation of the panel method.

However, you can see how the importance of this

approximation should decrease as the number of

panels, (of course this will increase the cost of

the computation considerably, so there are more

efficient alternatives.)

Hess and Smith decided to take the simplest possible

approximation, that is, to take the source strength to be

constant on each of the panels

Therefore, we have unknowns to solve for in

our problem: the panel source strengths qi and the

constant vortex strength . Consequently, we will need

independent equations which can be obtained

by formulating the flow tangency boundary condition

at each of the panels, and by enforcing the Kutta

condition discussed previously. The solution of the

problem will require the inversion of a matrix of size

.

5.4. Cubic Spline

Cubic spline interpolation is a useful technique to

interpolate between known data points due to its stable

and smooth characteristics. The cubic spline has been

utilized within the grid generation procedure to

accurately model curves that may be found in

engineering situations. An example of such a curve is

a hull ship section which, using cubic splines, can be

regenerated using relatively few data points. The cubic

spline fits a cubic polynomial between each set of

defining data points. The cubic spline is equal at the

data points and the spline is thus continuous. If the

gradient and the curvature are also assumed to be

continuous then the spline can be derived.

The fundamental idea behind cubic spline

interpolation is based on the engineer‘s tool used to

draw smooth curves through a number of points. This

spline consists of weights attached to a flat surface at

the points to be connected. A flexible strip is then bent

across each of these weights, resulting in a pleasingly

smooth curve. The mathematical spline is similar in

principle. The points, in this case, are numerical data.

The weights are the coefficients on the cubic

polynomials used to interpolate the data. These

coefficients ‘bend‘ the line so that it passes through

each of the data points without any erratic behavior or

breaks in continuity.


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Cubic Splines Derivation

Consider a collection of known points ,

, ... , , , ...

. To interpolate between these data points

using traditional cubic splines, a third degree

polynomial is constructed between each point. The

equation to the left of point is indicated as

with a value of at point . Similarly, the

equation to the right of point is indicated as

with a value of at point .

If a set of data points is defined a unique cubic

polynomial can be defined between each set of points.

where is a third degree polynomial defined by

for

The first and second derivatives of these

equations are fundamental to this process, and they are

for

6. CONCLUSION

The new concept design of the FPSO is currently a

cylinder shaped or non ship-shape proposed by several

researchers which is providing a better design. LNG

products that have been produced on the FPSO LNG

will be transferred to the LNG Carrier. The transfer

process is called offloading process. There are two

methods of offloading process for oil and gas transfer

which is tandem and side-by-side. Current research

focuses on for side-by-side configuration.

Fluid flow around a structure can significantly alter

the structure‘s loading characteristics. Waves and

currents is a fluid that has the most impact on the

structure. Current research discuss the concept design

of fluid flow around a round-shaped FPSO LNG

interacted with a LNG carrier during side-by-side

offloading condition by CFD method based on the

Reynolds Average Navier-Stokes (RANS) equations.

ACKNOWLEDGMENT

Special thank to Pengajian Tinggi Malaysia (MOHE)

and Universiti Teknologi Malaysia for supporting this

research.

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[13] Kinnas, S. A., Yu, Y. H., and Vinayan, V., 2006,

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[14] Wang, H.-M，Zou, Z.-J., and Tian X.-M., 2009,

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The Hull of The DTMB Model 5415 at Different

Speeds, International Review of Mechanical

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[26] Ahmed, Y.M., 2011, Numerical simulation for the

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[28] Tahara, Y. et al., 2006, RANS Simulation of a

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[32] Broglia R., Zaghi, S. and Mascio, A. D., 2011,

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[33] Clauss, G. F. et al., 2011, Hydrodynamic

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Simulation of Organic Rankine Cycle System Using Turbocharger with

Cycle Tempo and Environmentally Friendly Fluid

Ruli Nutranta a, Idrus Al Hamid

a, Nasruddin

a, Harinaldi

a

aFaculty of Engineering


Tel : (021) 7270011 ext 51. Fax : (021) 7270077


Abstract

Organic Rankine cycle (ORC) is a modified rankine cycle

with working fluids, of organic material (Refrigerant).

Refrigeran pentane has low boiling point, therefore ORC can

be used in power plant which uses low temperature resources,

such as solar thermal exhausted gases and geothermal wells.

Organic Rankine Cycle (ORC) is used to convert heat energy

into mechanical energy or electricity generated by a low

temperature of the hot sun. The working fluid used is R12,

R22, R134a and Pentane. Simulations performed with an

organic Rankine cycle temperature and pressure with cycle

tempo program. By programming the simulation cycle tempo

and got the result on the maximum power a turbine to the

conditions of the working fluid Pentane to the input turbine T

= 700C and pressure = 2 bar can generate 2.07 kW.

Turbocharger is one of the alternatives in the energy

conversion of the energy of motion into electrical energy.

Turbocharger rotation will be used to turn a generator and

converts the energy of motion into electrical energy.

Keyword : organic rankine cycle, energy, working fluid,

turbocharger

1. Introduction

Power energy in Indonesia has increase rapidly in national

daily consumsion and has developed renewable energy

include solar energy. Solar energy plays an important role in

the utilization of electrical energy in Indonesia. On average

the sun shines in Indonesia about 12 hours per day. If it takes

three hours on average for the utilization of solar thermal

then it becomes extremely beneficial to the interests of

society, especially in rural areas. To meet the demand for

electricity has not come into the house (by 33.4% in 2010),

then the solar power into one of the renewable alternative

energy that should be developed [1]. In this solar energy

utilization, solar thermal energy is one of Indonesia are very

rarely used. Utilization of solar thermal by finding a suitable

working fluid and meets the latest technology, is one reason

researchers simulate materials such as R12, R22, R134a and

Pentane.

2. Study Literature The majority of electrical generating plants are variations of

vapor power plants where water is the working fluid. The

basic components of a simplified Rankine cycle are shown in

Figure 1

Figure 1. Basic Organic Rankine Cycle

The Rankine cycle is the thermodynamic cycle that models

of Figure 1. In analyzing this cycle, one neglects the stray

heat that takes place between the plant components and their

surroundings. Further, kinetic and potential energy effects are

ignored. Finally, each component is considered to be

operating at steady state

Table 1. Equation of Basic ORC

The Rankine cycle differs from the Carnot cycle in that the

heat trans-fer processes take place at constant pressure

instead of constant tem-perature. On the other hand, much of

the Rankine cycle cooling and heating processes include

phase changes, and thus, they also preserve the isothermal

character while the phase change is in progress. For this

reason the efficiency is very good, but still less than the

Carnot efficiency. For example, the isothermal character is

not preserved for the first part of the heating process, when

the liquid is in the compressed form, and or the last part of

the heating process, when the fluid is in the superheated

vapor region. The steam must be in a superheated state before

entering the turbine so that the liquid state inside the turbine

can be avoided. Condensation in the turbine can cause blade

erosion.

Subsistem Equation

Pump (1) =

Evaporator

(2)

Turbine (3) Condenser

(4)

IMAT-UI 038


2012

234

Many investigation is conducted the simulations and

experiments to find an effective working fluid used in the

ORC. Yamamoto [2] investigated on the estimation of

operating conditions using a ORC working fluid of water and

HCFC-123. Saleh [3] examine 31 working fluid on the

condition of the sub-critical and Supercritical of geothermal

power generation on the ORC and Tchanche et.al.[4] using

organic solar Rankine cycle system to investigated 20

working fluid that has the best work in low-temperature.

In organic Rankine cycle, Mills [5] examines several

technologies including solar-powered ORC is capable of

generating electricity in several countries. Research on a

combination of fuel cell engine, gas turbine and ORC has

been done by Sanchez [6]. This study used the work of some

of the refrigerant fluid in the ORC machines. To R245fa

found to have the best advantage to produce electricity.

Vankeirsblick [7] also found that the ORC is the most

efficient engine to run a small generator instead of steam

power plants. In this study, ORC with simulated regenerator

to produce power 368.2 kw generator with an efficiency of

95%. Research on solar-powered ORC was also carried out

by Nasri [8] in the Sahara desert region in the form of

modeling and simulation. Although many studies examine

both ORC working fluid and the system, but not many

researcher is substituted turbocharger as turbine in Organic

Rankine Cycle

The objective of this study is to investigate the performance

working fluid R12, R22, R134a and pentane to electricity in

solar low-temperature ORC with turbocharger as a turbine

with cycle tempo.

3. Research Methodology The research is to compare the simulation and solar-

powered ORC engine. Activity of research is taken to create

a prototype solar-powered ORC with turbocharger. The ORC

will be connected with solar thermal collectors consisting of

flat and/or parabolic collectors. Research methodology in this

study can be seen in Figure 2 below. Starting from the

previous test, which has a solar thermal flat and parabolic

solar collectors, the research will investigated the

characteristics of water as a medium. The results obtained are

solar radiation and the efficiency of collectors, Heat Removal

Factor and Heat Loss Coefficient for some combination of

flat and parabolic

Figure 2. Reasearch Methodology

Setup experiment was conduct to small ORC with

temperature below 800C.

Figure 3. Organic Rankine Cycle with Turbocharger

The consideration of simulation are listed as follow:

a. The condensation temperature is 350C (308 K)

b. The evaporator temperature was set to 700C (343 K)

c. The reference state temperature was set to 300C (303 K)

d. The efficiency of the pump was set to 0.75

e. The efficiency of the turbocharger was set to 0.75


2012

235

Working fluid were investigated in this study is a working

fluid circulating in Indonesia and environmentally friendly.

Working fluid mixture of from R12, R22, RC134a and

Pentana. For Pentane, Indonesia still imports from abroad.

But working is required characteristics and perfomanya for

further research in the UI. The selected working fluid can be

seen in Table 2 below

Table 2. Physical properties working fluid

Molecular

weight

(g/mol)

(oC)

(MP

a)

Std 34

Safety

group

ODP GWP

(100

year)

R12 120.91 112 4.114 A1 1.000 10,890

R22 86.47 96.2 4.99 A1 0.055 1810

R134a 102.03 101.1 4.06 A1 0 1300

Pentane 72.15 196.6 3.37 A3 0 20

The turbocharger turbine, which consists of a turbine wheel

and a turbine housing, converts the engine exhaust gas into

mechanical energy to drive the compressor.The gas, which is

restricted by the turbine's flow cross-sectional area, results in

a pressure and temperature drop between the inlet and outlet.

This pressure drop is converted by the turbine into kinetic

energy to drive the turbine wheel.

Figure 4. Gasoline Turbocharger 1300 cc

As the radial-flow turbine is the most popular type for

automotive applications, thefollowing description is limited

to the design and function of this turbine type. In the volute

of such radial or centripetal turbines, exhaust gas pressure is

converted into kinetic energy and the exhaust gas at the

wheel circumference is directed at constant velocity to the

turbine wheel. Energy transfer from kinetic energy into shaft

power takes place in the turbine wheel, which is designed so

that nearly all the kinetic energy is converted by the time the

gas reaches the wheel outlet.[9]

4. Result and discussion

Research has done with electricity with condesor temperature

of 313K, 348K and evaporator temperature ambient

temperature of 300K. 0.85 pump efficiency, turbine

efficiency 0.7 and 5 kW of electric power [10]. Thus the

calculation of this system can be seen in Table 3 below.

Figure 5. Efficiency - Power 5 kW

In this research, thermal efficiency of Pentane also has a

number greater than most other working fluid. It is 8.9626 at

348K evaporator temperature. In Figure 4 can be seen five

working fluid thermal efficiency. Working fluid which has

the smallest thermal efficiency is HRC 12 with a value of

6.5233 is almost equal to R134a the value of 6.5975

Table 3. Summary of working fluid with Cycle tempo

No Fluid

System

Subsystem pin

(Bar)

Ppump

(W)

m

(kg/s)

P

(kW)

1 R 12

Pump 9

30 0.121 2.15 Evaporation 10

Turbine 10

Condenser 8.462

2 R 22

Pump 14

20 0.107 0.16 Evaporation 15

Turbine 15

Condenser 13.55

3 R134a

Pump 12

20 0.11 0.52 Evaporation 13

Turbine 13

Condenser 8.87

4 Pentane

Pump 1

20 0.121 2.07 Evaporation 2

Turbine 2

Condenser 0.977

This research, simulation cycle tempo conducted to

determine how the power generated in the turbine. R12, R22,

R134a and pentane are included as a working fluid

temperature at 35 ° C minimum and 80 ° C conditions at its

maximum. Simulation in cycletempo at table 3 shown the

highest value for the power is in R12 (2.15 kW). However, to

generate power, its required pump 30W and pressure up to 10

bar. Comparing with pentane which only requires power 20

W and pressure pump 2 bar, it is clear that for pentane

working fluid lighter than R12


2012

236

Figure 6. Pentane Simulation

The results of pentane is obtained power 2.07 kW at inlet

turbine pressure 2 bar and 0.977 bar at outlet pressure in

figure 6. The data can be explained:

- Pump: pump power 20 W with 75% efficiency entropy and

mechanical efficiency of 65%.

- Evaporator: subtitud by PHE with inlet temperature 35 ° C

and outlet temperature 80°C. A pressure of 2 bar for water

entering the evaporator temperature 80°C entry pressure of

1 bar, the incoming flow 0123 kg / s

- Turbine: from the inlet temperature 70 C to 40°C outlet

temperature. Pressure of 2 bar entry to 0.977 bar 2.07 kW

turbine generates power

- Condenser: the incoming pressure 0.977 bar and

temperature 35°C.

5. Conclusion

Conclusion can be drawn from the cycle tempo simulation by

using a turbocharger and environmentally friendly working

fluid:

a. The turbocharger can be used by decrease flow rate. This

adjustment can be made by used inlet nozzle

turbocharger.

b. Pentane have good conditions for development, especially

at low pressure and temperatures below 100 ° C.

Therefore the use of pentane substance is suitable for use

in Indonesia because of its nature.

References

[1] Sumiarso L.Regulasi dan pengembangan energi baru

terbarukan dalam rangka energi bersih. UMB 2011.

[2] Yamamoto T, Furuhata T, Arai N, Mori K. Design and

testing of the organic Rankine cycle. Energy

2001;26:239-51.

[3] Saleh B, Kogibauer G, Wedland M, Fischer J. Working

for low temperature organic Rankine cycle. Energy

2007;32:1210-21

[4] Tchanche BF, Papandakis G, Lambrios G, Frangoudakis.

A Fluid selection for low temperature solar organic

Rankine cycle. Appl Therm Eng 2009;29:2168-76

[5] D. Mills, Advanced in Solar thermal electricity

technology, Solar Energy (2004), vol 76 0hal 19-31

[6] D Sancez, JM Munoz de Escalona, B Monje, R

Chacartegui, T Sanchez, Preeliminary analysis of

compound system based on high temperatur fuel cell, gas

turbine and organic Rankine Cycle, Journal of Power

Source (2011), vol 196 hal 4355-4363

[7] Vankeirsblick, Vanslambrouck, Gusev, De Paepe,

Organic Rankine cycle as efficient alternative to steam

cycle for small scape power generation, International

Conference of Heat Transfer, Fluid Mechanic and

Thermodynamic, Mauritius (2011)

[8] Faozi Nasri, Chouki Ali, Habib Ben Bacha, Electricity

production system form solar heated rankine cycle :

modelling and simulation, IJRRAS (2011) vol 8 hal 176-

183

[9] Brown S, The Turbomustangs.com : Complete

Turbocharging guide, Underpsi of utahstang.com, 2003

[10] Nutranta. R, AlHamid. MIA, Nasrudin, Harinaldi. Studi

karakteristik fluida kerja hydrokarbon ramah lingkungan

pada siklus rankine (SRO) bertenaga surya. SNTTM X

Universitas Brawijaya Malang, 2011


2012

237

Thermophysical Properties of Novel Zeolite Materials for Sorption Cycles

Kyaw Thua, Young-Deuk Kim

a, Baojuan Xi

b, Azhar Bin Ismail

b, Kim Choon Ng

a,b,*

aWater Desalination and Reuse Center

King Abdullah University of Science and Technology, Thuwal 23955-6900

Saudi Arabia Tel : (966) 2-8084969.

E-mail : [email protected]; [email protected]

bDepartment of Mechanical Engineering

National University of Singapore, Singapore 117576

Tel : (065) 65162214. Fax : (065) 6779-1459

E-mail : [email protected]; [email protected]; [email protected]

*Corresponding Author

ABSTRACT

This article discusses the thermophysical properties of

zeolite-based adsorbents. Three types of zeolite (Z-01, Z-02

and Z-05) with different chemical compositions developed by

Mitsubishi Plastics, Inc. are analyzed for possible

applications in adsorption chillers and desalination cycles.

Static volumetric method is adopted with N2 gas sorption at

77 K. Thermophysical properties such as pore surface area,

micropore volume and pore size distribution are evaluated

using standard multi-point Brunauer-Emmett-Teller (BET)

and Non-Local Density Functional Theory (NLDFT)

methods. It is observed that Aluminosilicate functionalized

Z-02 exhibits the highest surface area with huge micropore

volume.

Keywords : Adsorption, Zeolite, Thermophysical properties

1. INTRODUCTION

Frequent and severe natural disasters such as super storms

and earthquakes are claimed to be attributed to the

environmental issues such as Global warming and

greenhouse gas emissions. The burning of hydrocarbons for

various applications i.e., the industrial processes, the

transportation and the life comfort, inevitably contributes

damages to the environment. Thus, secondary fuels

(photovoltaic, winder turbine and fuel cells) gain much

attention as clean and environmental-friendly alternates to

depleting oil [1-3]. However, the primary systems powered

by secondary fuels conventionally operate at low efficiencies,

typically below 60%. With the introduction of the

cogeneration concept such as Combined Heat and Power

(CHP) systems, the overall system efficiency can be realized

as high as 80% [4]. The combine systems normally include

waste heat-driven absorption and adsorption cycles producing

useful effects such as cooling and potable water [5].

The adsorption cycles employed sorption principles between

the solid adsorbent and the vapor phase adsorbate. Common

adsorbent materials for adsorption chillers are silica gel,

zeolite and activated carbon whilst water, ethanol and

methanol are used as adsorbate. Poor coefficient of

performance (typically less than 0.7) of adsorption cycles

calls for the development of new adsorbent materials and

improved heat and mass recovery schemes [6-11]. The

adsorbent selection depends on the quality of the available

waste heat and type of refrigerant. Silica gel is commonly

employed for low-temperature waste heat application whilst

zeolite and activated carbon are used where the waste-heat is

higher than 100 °C [12-15]. With recent development in the

novel zeolite materials, it is possible to regenerate the

adsorbate from the zeolite using low-temperature heat

sources, typically as low as 60 °C [16-18].

This article presents the thermophysical properties of novel

zeolite materials developed by Mitsubishi Plastics, Inc. These

materials are developed for possible applications where

regeneration temperature is as low as 55 °C with water vapor

adsorption. Three types of powdered-zeolite materials with

different chemical compositions are investigated for their

thermophysical properties such as pore surface area, pore

volume and pore size distribution.

2. MATERIAL AND METHOD

Three types of zeolite materials code names (Z-01, Z-02 and

Z-05) are investigated. The chemical composition and

Scanning Electron Microscope (SEM) pictures of these

samples are given in Table 1.

Table 1: Chemical composition and SEM pictures of the

samples.

Z-01 Z-02 Z-05

Composition FeAlPO2.nH2O SiAlPO2.nH2O AlPO2.nH2O

SEM image

(Particle

description)

SEM image (high

resolution)

Static volumetric method is used with N2 gas adsorption at 77

K. The AutoSorb-1 analyzer manufactured by Quantachrome

Corporation is employed to investigate thermophysical

properties. The minimum relative pressure available by this

IMAT-UI 039







2012

238

analyzer is 1.0 x 10-7

with N2 sorption whilst the applicable

lowest surface area is 0.01 m2/g without upper limit. The

AutoSorb-1 has the capability of measuring adsorbed or

desorbed volumes of nitrogen at relative pressures in the

range 0.001 to slightly lower than 1.0. The supplied software

(ASWIN) facilitates the analyses of the surface

characteristics using various methods such as BET surface

area (single and/or multipoint), Langmuir surface area,

adsorption and/or desorption isotherms, pore size and surface

area distributions, micropore volume and surface area using

an extensive set of built-in data reduction procedures. All the

samples are performed degasification at 140 °C under

vacuum for 12 hours prior to the experiments.


2.8 N2 Isotherm

The N2 gas adsorption by three types of selected zeolite

materials is given in Figure 1. It is observed that all the

adsorbent and adsorbate pairs exhibit Type II isotherms and

Aluminosilicate-based Z-02 has the highest uptake followed

by Z-01 and Z-05. The amount of uptake by Z-02 is

significantly higher than those of Z-01 and Z-05 with average

uptake of 170 cm3/g.

Figure 1: N2 gas adsorption by selected adsorbent

2.9 Micropore Analysis

Using the N2 adsorption data, the present of micropore and

the external surface area are evaluated using the t-method of

de-Boer [19]. A t-plot is a plot of the volume of gas adsorbed

versus t, the statistical thickness of an adsorbed film and

Figure 2 gives the t-plot of the zeolite materials whilst Table

2 provides the summary of the micropore analysis by t-

method.

The results show that all the samples present micropore.

However, Z-02 exhibits huge amount of micropore and the

micropore volume and micropore surface area of Z-02 are

found to be one order than those of Z-01 and Z-05.

Figure 2: Micropore analysis of the selected Zeolite

materials using t-plot method

Table 2: The summary of the micropore analysis of the selected

Zeolite materials by t-method.

Parameter Z-01 Z-02 Z-05

Slope 6.386 1.891 4.183

Intercept 18.551 172.00 13.90

Correlation coefficient, r 0.999994 0.998 1.000

Micropore volume (cm3/g) 0.029 0.266 0.021

Micropore area (m2/g) 67.387 737.000 41.032

External surface area (m2/g) 98.781 29.246 64.696

2.10 Surface Area Analysis

Using the Brunauer-Emmett-Teller (BET) method, the

surface area of the adsorbent can be determined as,

1 1 1

0 01

C P

P W C W C Pm mW

P

(1)

where W is the weight of the adsorbed gas at a relative

pressure (P/P0), Wm is the weight of adsorbate at a monolayer

coverage and C is the BET constant. This is related to the

adsorption energy of the first adsorbed layer, indicating the

magnitude of the adsorbent/adsorbate interactions. Table 3

summarizes the surface area analysis results. It is noted due

to the present of huge micropore, Z-02 possesses huge

surface area of 766 m2/g.

Table 3: BET analysis summary of Zeolite materials.


Slope 20.936 4.544 30.053

Y-intercept, i 2.21E-02 6.96E-06 3.26E-02

Correlation coefficient, r 1 1 0.999995


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BET constant, C 946.833 653308.334 922.153

Surface area (m2/g) 166.168 766.374 115.752

2.11 Pore Size Distribution

According to the IUPAC recommendation, the classification

of pores present in an adsorbent is such that micropores (<

2nm), mesopores (between 2nm and 50nm), and macropores

(larger than 50nm) as to their pore widths. Pores in

adsorbents represent higher surface area as well as high

selectivity in reaction and adsorption. Pore Size Distribution

(PSD) analysis is the most essential information of the pore

and it is used to evaluate the population of pores as a function

of the pore width [20]. The pore-size distribution analyses of

the mentioned Zeolite materials are conducted using the Non-

Local Density Functional Theory (NLDFT) method with the

application of the provided software package by the

AutoSorb-1. Here, the Equilibrium Model is adopted to

determine the PSD of the aforesaid adsorbents. Figure 3

shows the cumulated pore volume of the selected adsorbent

materials. It is observed that Z-02 type adsorbent exhibits the

highest total pore volume followed by Z-01 and Z-05.

Figure 3: Cumulated pore volume of the selected Zeolite

materials by NLDFT method

Figure 4 gives the PSD comparison of three zeolite materials

whilst Table 4 shows the analyses summary. The pore width

of all the adsorbent is found to be between 2 and 5 nm. It is

noted that Z-05 exhibits two maxima distribution or bimodal

type.

Dubinin-Astakhov (DA) analysis of the adsorption isotherm

of the Zeolite materials for N2 gas adsorption is depicted in

Figure 5 where the DA equation is given as,

0

0

lnn

RT P PW W Exp

E

(2)

here W is the weight of adsorbed amount at relative pressure,

P/P0 and T, W0 is the total adsorbed weight, E is the

characteristic energy and n is the DA constant. The

significantly higher pore volume of Z-02 is detected here.

Finally, Table 5 gives the analysis summary using DA

method. DA analysis reveals that Z-02 possesses the highest

micropore volume (0.29 cm3/g) as well as the highest

characteristic energy (71 kJ/mol).

Figure 4: Pore Size Distribution summary of different

types of Zeolite materials Table 4: Summary of NLDFT (Equilibrium Model) for PSD analysis

of Zeolite materials.

Pore Volume

(cm3/g)

Pore width

(Å)

Fitting Error

(%)

Z-01 0.108 34.18 1.414

Z-02 0.320 45.7 0.310

Z-05 0.084 11.14 0.83

Figure 5: DA plot of the Zeolite materials for N2 gas

adsorption Table 5: DA analysis summary of the Zeolite materials.


Characteristic energy, E (kJ/mol) 6.122 70.566 5.988

DA constant, n 1.0 1.0 1.0

DA Micropore Volume (cm3/g) 0.100 0.288 0.069

Pore Radius (Å) 7.10 3.20 7.20


2012

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4. CONCLUSION

Three types of novel zeolite materials with different chemical

compositions provided by Mitsubishi Plastics, Inc. have been

investigated for their possible application in adsorption

refrigeration and desalination cycles. The samples are

analyzed using the static volumetric method with N2 gas as

adsorbate at 77 K. It is observed that Z-02 type adsorbent

with Aluminosilicate composition exhibits superiority in

terms of surface area (766 m2/g) and micropore volume (0.32

by NLDFT method and 0.29 by DA method). It is also noted

that the pore diameter of all the selected adsorbents is

between 2 and 5 nm with Z-05 showing bimodal type or two

maxima distribution. Based on these analyses, it is suggested

that Z-02 type material is suitable for adsorption cycle

application with superior surface characteristics.

ACKNOWLEDGMENT

The authors gratefully acknowledge Mayekawa

Manufacturing Co., Ltd. for the supply of zeolite materials.

REFERENCES

[1] S. Giddey, S.P.S. Badwal, A. Kulkarni, C. Munnings, A

comprehensive review of direct carbon fuel cell

technology, Prog Energy Combust Sci, 38(3) (2012)

360-399.

[2] A. Kirubakaran, S. Jain, R.K. Nema, A review on fuel

cell technologies and power electronic interface,

Renewable Sustainable Energy Rev, 13(9) (2009) 2430-

2440.

[3] G. Zhang, S.G. Kandlikar, A critical review of cooling

techniques in proton exchange membrane fuel cell

stacks, International Journal of Hydrogen Energy, 37(3)

(2012) 2412-2429.

[4] S. Mekhilef, R. Saidur, A. Safari, Comparative study of

different fuel cell technologies, Renewable Sustainable

Energy Rev, 16(1) (2012) 981-989.

[5] A. Myat, K. Thu, Y.D. Kim, B.B. Saha, K. Choon Ng,

Entropy generation minimization: A practical approach

for performance evaluation of temperature cascaded co-

generation plants, Energy, 46(1) (2012) 493-521.

[6] K.C. Ng, X. Wang, Y.S. Lim, B.B. Saha, A.

Chakarborty, S. Koyama, A. Akisawa, T. Kashiwagi,

Experimental study on performance improvement of a

four-bed adsorption chiller by using heat and mass

recovery, Int J Heat Mass Transfer, 49(19–20) (2006)

3343-3348.

[7] X. Wang, H.T. Chua, K.C. Ng, Experimental

investigation of silica gel–water adsorption chillers with

and without a passive heat recovery scheme, Int J Refrig,

28(5) (2005) 756-765.

[8] J. Di, J.Y. Wu, Z.Z. Xia, R.Z. Wang, Theoretical and

experimental study on characteristics of a novel silica

gel-water chiller under the conditions of variable heat

source temperature, Int J Refrig, 30(3) (2007) 515-526.

[9] H. Luo, R. Wang, Y. Dai, The effects of operation

parameter on the performance of a solar-powered

adsorption chiller, Appl Energy, 87(10) (2010) 3018-

3022.

[10] T. Miyazaki, A. Akisawa, B.B. Saha, I.I. El-Sharkawy,

A. Chakraborty, A new cycle time allocation for

enhancing the performance of two-bed adsorption

chillers, Int J Refrig, 32(5) (2009) 846-853.

[11] K.C. Ng, Recent developments in heat-driven silica gel-

water adsorption chillers, Heat Transfer Eng, 24(3)

(2003) 1-3.

[12] C. Chen, R. Wang, Z. Xia, R.G. Oliveira, J. Hu,

Performance analysis of silica gel and its composite

adsorbent-water adsorption refrigeration working pairs,

Huagong Xuebao/CIESC Journal, 59(SUPPL.) (2008)

43-48.

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Kashiwagi, A. Akisawa, B.B. Saha, Experimental

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characteristics, Appl Therm Eng, 21(16) (2001) 1631-

1642.

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silica gel-water system, Int J Heat Mass Transfer, 52(1-

2) (2009) 516-524.

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zeolite crystallites and coatings on metal surfaces,

Chemistry of Materials, 2(6) (1990) 712-719.

[17] A. Freni, L. Bonaccorsi, E. Proverbio, G. Maggio, G.

Restuccia, Zeolite synthesised on copper foam for

adsorption chillers: A mathematical model,

Microporous Mesoporous Mater, 120(3) (2009) 402-

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Broekhoff, A. van den Heuvel, T.J. Osinga, Thet-curve

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2012

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Review Paper: Sea-water Ice Slurry Generator and Its Application on

Indonesian Traditional Fishing

A.S. Pamitrana, H.D. Ardiansyah

b, M. Novviali

b



Tel : (021) 7270032, Fax : (021) 7270033

E-mail : [email protected] bGraduate Student, Department of Mechanical Engineering,


ABSTRACT

Sea-water ice slurry generator is aimed to produce ice-

slurry using sea-water while fishing on boat. Some models

of ice slurry generator have been researched and developed.

Indonesia as an archipelago and maritime nation has big

potential in fishery. Appropriate model for Indonesian

traditional fishing boat is necessary. The most important

part of ice slurry generator is evaporator, including auger

and scrapper. Some studies on ice slurry are presented in

this review paper in order to get larger view on design of

ice slurry generator.

Keywords: Ice slurry, cooling, sea-water, evaporator,

fishing

1. INTRODUCTION

Nowadays international agreement has been applied to

protect environment from using refrigerant which contain

chlorine. Every chlorine refrigerant has effect for

environment destructive where the parameter is said Ozone

Depletion Potential. Refrigerant contains hydrogen was

developed to replace CFCs, such as NH3. However, the

toxicity of NH3 must be considered for a refrigeration

system. The properties of refrigerants are very important to

be considered in a system due to their adhered effects.

Therefore, secondary refrigerant such as water solution can

be largely used to minimize accident and leak effect of

refrigerants.

Ice slurry is useful secondary refrigerant for many

applications. Low temperature of primary refrigerant can

absorb more heat and change phase of fluid from liquid

become mixture of ice-liquid. Fluid for ice slurry can be

pure water or solution has freezing point depressant such as

Sodium Chloride, Ethanol, Ethylene Glycol, Propylene

Glycol (Kauffeld et al., 2005) and sea water (A.S Pamitran

et al., 2012). Figure 1 illustrates a schematic diagram for ice

slurry as a secondary refrigerant published in Meewise,

2004.

The characteristic and advantage of ice slurry has been

researched before 1975. One of interesting topic of ice

slurry study for some researchers are heat transfer and

pressure drop of ice slurry even

Figure 1: Secondary refrigerant scheme (Meewise,

2004)

some topic presented ASHRAE Meeting at June 1998

Toronto, Canada (Kirby P. Nelson et al., 1998). Energy

storage of ice slurry is higher than others secondary

refrigerants because ice slurry contain ice particle which

has high latent heat in solution. Moreover, ice slurry is fast

and effective for cooling because of large contact surface

for heat transfer between ice particle and product. Ice slurry

can reduce dimension of tank, pipe, chiller, and can reach

economic efficiency by reducing more than 70% power of

pump compare with water (Kasza et al., 1988). Nowadays,

some researchers concern on ice slurry because of its large

benefit. For example application in industry (Wang and

Kusumoto,2001; Rivet, 2009), medical and direct cooling

for food or fish (Wang and Goldstein, 2003; Pineiro et al.,

2004).

Some published papers study on thermofluids

characteristics of ice slurry. Gupla dan Frazer (1990)

explained ice slurry in heat exchanger with 6% ethylene

glycol, ice fraction of 0%-20%, flow rates of 1.18 m3/hour

and 2.16 m3/hour, diameters of ice slurry of 0.125 mm and

0.625 mm. The result showed total heat transfer coefficient

proportional to flow rate and opposite to increasing ice

fraction. Pressure drop is constant until ice fraction of 20%

and then rapidly increase more than 20%. Kauffeld (1999)

compared ethanol solution and potassium carbonate

solution as fluids to produce ice slurry. Ethanol solution

result small ice particle and has heat transfer coefficient

increase along with ice fraction increase. The opposite

result was found for potassium carbonate solution with big

ice particle. The heat transfer coefficient decreases along

with ice fraction increase. Knodel (2000) concluded heat

transfer coefficient decrease with ice fraction increase. As

well Gupla and Frazer (1990) showed the same result.

Knodel (2000) explained heat transfer decrease because

IMAT-UI 040


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moving of ice slurry flow change from turbulent to laminar

due to increasing of the fraction ice. Different result was

reported by Bellas J et al. (2002) that measured ice slurry of

5% propylene glycol solution in plate heat exchanger with

0-25 % ice fraction and 1-3.7 m3/hour flow rate. Bellas et

al. (2002) explained that increase of ice fraction 0-20%

make pressure drop about 15% and total heat transfer

coefficient rapidly increase along with increase of flow rate.

Variation of ice fraction has insignificantly effect on heat

transfer coefficient.

N. Putra et al. (2004) used ice breaker to make ice slurry.

Heat transfer coefficient increases whilst flow rate and ice

fraction increase in plate heat exchanger. They assumed

heat transfer coefficient and pressure drop are functions of

viscosity, Reynolds number, diameter and ice fraction of ice

slurry. Stamatiou dan Kawaji (2005) studied heat transfer

coefficient in vertical rectangular channel with given heat

flux on wall surface. Nusselt number increase along with

increase of ice fraction and heat flux. The fusion of ice

particle formed on pipe result in higher convection heat

transfer than conduction heat transfer along of the pipe. The

profile of velocity on wall surface for ice slurry and liquid

is different. Liquid velocity was not function of heat, but ice

slurry velocity was function of fusion of ice particle. Lee

D.W. et al. (2006) deeply researched heat transfer with

6.5% ethylene glycol in diameter of 13.84 mm and length

of 1500 mm, ice slurry mass flux of 800-3500 kg/m2s and

ice fraction of 0-25%. The result was heat transfer

coefficient increase along with increase of flow rate and ice

fraction, but effect of ice fraction was insignificant in

higher flow rate and in lower flow rate region. Niezgoda-

Zelasko (2006), Niezgoda et al. (2006) and Grozdek (2009)

researched on heat transfer and pressure drop with ice

slurry of 10% ethanol in horizontal pipe. High ice fraction

and high velocity give high heat transfer coefficient and

pressure drop. Heat flux has small effect on heat transfer.

Ice fraction of 10-15 % increases heat transfer coefficient in

laminar flow, but not for turbulent flow under same heat

flux. Beyond the value has high heat transfer coefficient.

J.P. Nedecarrats et al. (2009) used corrugated and smooth

pipe with given heat flux on pipe wall, velocity of 0.3-1.9

m/s and ice fraction of 0-30%. Pressure drop and heat

transfer coefficient increase along with fraction ice and

velocity. They founded critical point of pressure drop and

heat transfer coefficient. Comparison of corrugated and

smooth pipe was reported with result of heat transfer and

pressure drop are 2.5 times higher for corrugated pipe.

Report from some researchers can be concluded that ice

slurry characteristic depend on the solution, flow rate, ice

fraction, and diameter of ice particle. Although many

researchers have offered some methods and correlations

regarding ice slurry characteristic, but their results could

not used largely to predict heat transfer and pressure drop in

heat exchanger (Ayel et al., 2003). Therefore, study on ice

slurry is still opened in investigating of their thermofluid

characteristic. The other interesting topics on ice slurry are

its application, ice slurry generator, melting of ice slurry,

ice slurry formation, microscopic observation, measurement

and control, modeling and simulation.

2. ICE SLURRY DEFINITION

Ice slurry consists of liquid and ice particle (E Stamatiou et

al.,2003). Ice slurry is defined as fine-crystalline ice slurry

or liquid has ice particle with average diameter of equal or

less than 1 mm (Peter W Egolf et al., 2003). Nandy P. et al.,

2006, mentioned general definition for ice slurry:

a. Solution and solid with temperature up to -15 oC.

b. Ice slurry can be produced from brine solution with

freezing temperature lower than the freezing temperature

of water up to -50 oC.

c. Ice slurry has different characteristic and behavior

compared to brine solution.

d. Ice slurry is 2 phase fluid non Newtonian in high ice

fraction.

e. Ice slurry needs different calculation and prediction pipe,

pump, heat exchanger even storage tank.

3. ICE SLURRY FORMATION

Ice slurry formation has been explained by E Stamatiou

(2003). Generally, process of ice slurry formation consists

of supersaturation, nucleation, and growth. Moreover, there

are other processes of attrition, agglomeration and ripening

can happen on ice slurry formation.

Supersaturation occurs when driven force complete.

Therefore, supersaturation needs conditions of instability

and difference of chemical potential between solution and

solid crystal.

(1)

Chemical potential difference occurs because of

temperature or pressure driven force. Rate of crystallization

is influenced by grade of supersaturation solution.

Considering Raoult Law, when a liquid mixture with

methanol, ethylene glycol, propylene glycol, sodium

chloride, magnesium chloride, potassium chloride, etc, has

mixture pressure between vapor partial every component

and freezing temperature is lower than water.

Nucleation is formed when molecule getting stable. There

are two kind of nucleation viz. homogeneous and

heterogeneous. Nucleation can be recognized from existing

ice fraction in solution. Equilibrium of chemical potential

from driven force makes separation of solution. Pure water

separated from solution is partially freezing become ice

fraction.

Growing of crystal occurs in three steps viz. mass transfer

with molecule diffusion in bulk liquid through boundary

layer near nucleus, molecule merger to wall area and heat

transfer simultaneously from crystal to bulk area include

change of phase. Part of growing ice is supported by

rotation of shaft auger to accelerate transfer of mass and

heat. Some points of shaft auger break slug of ice crystal on

evaporator inner wall, and ice crystal is revolved in center

of generator. Interaction between nucleation and growth of

crystal determine crystal characteristic such as diameter,

distribution, and morphology of crystal (Mullin, J. W.

2001).


2012

243

Figure 2 shows that point t0 to t1 is supercooling or

supersaturation, then t1 to t2 is area where ice fraction

forming, t2 to tf is sensible heat until complete of nucleation

(T.A. Mouneer et al,.2011).

Figure 3 present a similar illustration as Figure 2 but with

different result. Nucleation is marked with increasing bulk

temperature. Nucleation starts from forming ice particle.

Figure 2: Time dependence curve for generator and

forming of volumetric ice concentration

(scraper) (T.A. Mouneer et al,.2011)

Figure 3: Time dependence curve and torque of shaft auger

(scraper) (Frank Qin et al., 2006)

4. ICE SLURRY GENERATOR

Every ice slurry generator has similar process with different

evaporator and scraper system. Scraper of ice slurry

generator consists of scraper itself and shaft auger as

illustrated in Figure 4. The purpose of scraper is to avoid

lump of ice on evaporator inner wall. There is a clearance

between scraper and the wall surface. Thermal resistance

makes ineffective heat transfer when ice attach on the wall.

Many industries use scraper system because it is cheaper

than others and produce high ice fraction (E. Stamatioua et

al., 2005, T.A. Mouneer et al., 2011). Heat transfer

equipment can be made with shell and tube or flooded

evaporator.

M. Miguel Leon (2006) modified scraper model with

helical scraper and flooded evaporator as shown in Figure

5.

There are some systems of ice slurry generator such as

falling liquid film ice slurry generator, orbital rod ice slurry

generator, fluidized bed,

Figure 4: Scraper of Ice Slurry Generator (T A

Mouneer et al,. 2011)

(a) (b)

Figure 5: (a) Flooded Evaporator, (b) Helical Scaper

(M. Miguel Leon et al., 2006)

(a) (b)

Figure 6: (a) Falling liquid system (b) Refrigeration

system


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Figure 7: Orbital Rod Ice Slurry Generator (E.

Stamatioua et al., 2005)

Figure 8: Vacuum Freezing (Michel Barth)

vacuum freezing, supercooled slurry ice production,

supercooled water jet, direct contact heat transfer. The

system is called falling film because ice slurry flow with

gravitation. Falling film is depicted in Figure 6. Falling

liquid system works without motor but needs pump for

circulation of ice slurry. Falling film system works similar

as scraper system, shown in Figure 7, but the motor for

auger has less friction. Falling film has higher speed of

rotary than scraper but has less power (E. Stamatioua et al.,

2005).

Vacuum freezing, as shown in Figure 8, use vacuum

pressure in evaporator to get low temperature for ice

crystal. This method needs vacuum pump. Asaoka et al.

2006 researched ice slurry using ethanol solution with

vacuum freezing. Hasegawa et al. 2002 used pure water

with this method (Hasegawa et al., 2002).

Fluidized bed, as shown in Figure 9, used bed heat

exchanger mechanism. Refrigerant flow through evaporator

in small pipes with flooded pipe, then

Figure 9: Fluidized Bed (Michel Barth)

Figure 10: Supercooled Slurry Ice Production

(SlurryICE TM

Manual Book)

Figure 11: Super Cooled Water Jet (T A Mouneer et

al., 2011)

ice slurry will separate on certain diameter as mesh

gravitation system. Ice fraction in small pipe is pushed by

pump (Pronk et al., 2005).

Supercooled Slurry Ice Production, as illustrated in Figure

10, is ice slurry producing method with supercooling.

Water flow with low velocity can be cooled to bellow of

freezing temperature without ice forming. Before leave

evaporator flow of supercooled water is teased with ice

crystal forming. Formed ice fraction depends on

supercooling solution leave evaporator.

Figure 11 illustrates a method develop by T.A. Mouneer et

al. (2011). Supercooled water jet applies water jet method

to improve velocity of solution. Pump is a main component

for circulation. The model works without shaft auger.

Different than others, in Direct Contact Heat Transfer

model refrigerant is blended with solution using coolant

nozzle. In application, the system can produce 40% ice

fraction (N.E. Wijeysundera et al., 2004).

Figure 12: Coolant Nozzle (N.E. Wijeysundera et al.,

2004)


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245

Figure 13: Ice slurry installation for Fishing (M. J Wang et

al., 1999)

5. ICE SLURRY FOR FISH COOLING

There are three aspects for maintenance fish quality, viz.

cooling, handling and cleaning. Handling is effort to avoid

bacteria with cutting broken area. The second is cleaning

bacteria source with hygienic equipment after get fish from

sea, and the last is cooling system. However from all steps,

cooling has dominant effect for fish quality.

Traditional fishing still uses block ice, commonly, rather

than ice flake. Schematic diagram for ice slurry installation

is shown in Figure 13. Today, ice slurry become popular

applied to keep fish quality in developing countries. Ice

slurry can avoid air between fish and ice slurry, so fish

cooling become faster because of larger contact surface,

and slowing growth of bacteria (M.J Wang et al. 1999).

Ice slurry with depressant temperature made active protein

function and protect probiotics good from heat risk

(T.Vajda, 1999). Ice slurry has three times cooling faster

than ice flake to get 2oC (J Paul, 2002). Ice slurry does not

injure the fish because of soft ice texture (Pineiro et al.,

2004). Ice generator works as heat exchanger for water and

refrigerant commonly called harvest tank where ice particle

is formed. Harvest tank consists of scraper and auger shaft.

Sea water is injected to push ice fraction above liquid with

different density. Ice fraction

Table 1. Time for good quality of fish with storage

temperature (Masyamsir, 2001).

Storage temperature Time for good quality

16 oC 1-2 days

11 oC 3 days

5 oC 5 days

0 oC 14-15 days

with low density goes to third tank. Ice fraction in

third tank is added by sea water because of technical reason

while still stirred by scraper. Wang et al., 1999, reported ice

slurry application in fisheries.

Table 1 lists limitation of time when fish kept in a certain

low temperature. Lower temperature is better in getting

longer time for fish quality.

6. ICE SLURRY FOR INDONESIAN FISHIERY

Indonesia as archipelago country has 18,306 islands which

are united by ocean with length of coast line of 81,000 km.

Indonesia is the 4th

largest fishery county has fishery

potential of around 6.4 million/years (Dahuri et al., 2002).

Although Indonesia has huge potential in sea, many

fishermen are still far away from their prosperity. Indonesia

Fishery Minister and BPS in 2010 reported amount 7.87

million destitute people spread in coastal area. It is means

that totality fishermen are poor. Therefore, using

advantages ice slurry for Indonesian fishermen are

supposed can improve their economic income.

In 2010, from 590,352 fishing boats, just 6,370 units (less

than 2%) can be classified as modern boats, and they are

more than 30 GT. Inboard motor boats are 155,922 units

(26%), outboard motor boats are 238,430 units (40%), and

189,630 units (32%) are sail boats (KKP, 2010). Outboard

motor boat with simple equipments for fishing of traditional

fishermen just can catch fish in coastal area during 7 to 9

months/year. It is low productivity and there is reduction of

209 kg/month. They use unfair profit sharing system

between owner and workers. This condition makes number

of new fishermen utilizing fishery source is just 69.68%

(Yonvitner, 2007). New technology such as ice slurry

generator portable suppose can repair fish cold chain of

Indonesian fishermen.

In several survey for community engagement program

supported by grant of Hibah Pengabdian Masyarakat 2012

from DPRM University of Indonesia, authors concluded

some considerations must be taken for ice slurry application

in outboard motor boat, such as:

a. Ship stability

b. Increasing of draft (capacity)

(a)

(b)


2012

246

(c)

Figure 14: (a) Fishing boat 20 GT with inboard motor, in

Muara Angke, North Jakarta, (b) Fishing boat

with outboard motor, in Tidung Island,

Kepulauan Seribu, Jakarta, (c) Fishing boat with

outboard motor, in Balongan, Indramayu

c. Average production of fish

d. Duration and distance for fishing

e. Wide of close area and open area in board to place

equipments and activity area

f. Economic level

g. Local wisdom of fishermen

Figure 14 shows sample of some local boats used by

Indonesia fishermen.

7. SEA-WATER ICE SLURRY

Salinity is main parameter denote salt content in water.

Salinity unit is ppt (part per ton). NaCl (gram) contain 1000

gram sea water (Wibisono, 2004). Salinity may depend on

tidal, rainfall, distillation, and topography. Figure 15 shows

chart

Figure 15: Salinity chart (Calor M Lalli,.2006)

Table 2. Ion in sea water salinity of 35 ppt (Calor M

Lalli,.2006)

Ion

Concentration

(g/kg) Weight (%)

Chloride (Cl-) 18.98 55.04

Sodium (Na+) 10.56 30.61

Sulphate (SO42-) 2.65 7.68

Magnesium (Mg2+) 1.27 3.69

Calcium (Ca2+) 0.4 1.16

Potassium (K+) 0.38 1.1

Bicabonate (HCO3-) 0.16 0.41

Bromide (Br-) 0.07 0.19

Borate (H3BO3) 0.03 0.07

Strontium (Sr2+) 0.01 0.04

Figure 16: Effect of concentration of NaCL with

Freezing Point (K. S. Hilderbrand., 1998)

of salinity for location around the world.

The highest component in sea water is NaCl. Table 2

explain composition of ion in sea water. NaCl has number

of 85.65% of weight gram sea-water. Therefore, calculating

of ice slurry can use properties of NaCl.

K. S. Hilderbrand (1999) presented correlation of NaCl in

solution with freezing temperature, as shown in Figure 16.

The figure is discovered eutectic point at -21.1 oC. Eutectic

temperature is limit temperature for NaCl solution that

direct change from solution to solid salt and ice. The

mixture percentage is called mixture eutectic.

Table 3. Effect of salinity with freezing temperature (Feistel

et al., 2008)


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Salinity (ppt) Freezing temperature (oC)

5 -0.269

10 -0.536

15 -0.803

20 -1.074

25 -1.348

30 -1.625

35 -1.908

40 -2.195

45 -2.487

50 -2.784

55 -3.087

60 -3.396

65 -3.711

70 -4.033

Table 4. Effect of salinity with freezing temperature (The

Practical Salinity Scale 1978)

Salinity (ppt) Freezing temperature (oC)

0 0

5 -0.274

10 -0.542

15 -0.812

20 -1.083

25 -1.358

30 -1.638

35 -1.922

40 -2.212

Feistel et al. (2008) have been discovered relation salinity

with freezing temperature, as listed in Table 3. The

Practical Salinity Scale 1978 and the International Equation

of State of Seawater 1980, Unesco Technical Papers in

Marine Science No.36 determine relation of salinity with

freezing temperature, as listed in Table 4.

Ice slurry is mixture of ice particle and solution, so it‘s

important to know sea water properties for calculate

parameters of ice slurry thermofluids. Mustafa H et al.

(2010) reported about sea water thermofluid. Figure 17

shows (a) effect of temperature on density, (b) effect of

temperature on dynamic viscosity, (c) effect of temperature

on thermal conductivity, (d) effect of temperature on

specific heat, and (e) effect of temperature on specific

enthalpy.

(a)

(b)

(c)

(d)

(e)

Figure 17: Experimental results of Mustafa H. et al.,

2010


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248

A. Melinder (2010) proposed some correlation of ice slurry

properties as function of temperature.

(2)

(3)

(4)

(5)

Ice fraction can be calculated using correlation proposed by

Jean-Pierre Be´de´carrats et al., 2009 and Cecilia Hägg,

2005.

(6)

Total enthalpy can be calculated using correlation proposed

by T. Kousksou et al. (2010) and A. Melinder et al. (2005).

(7)

Thermal conductivity can be calculated using correlation

proposed by Taref (1940).

(8)

D.G. Thomas (1965) dan Jacques Guilpart et al. (2006)

proposed correlation of ice slurry viscosity.

(9)

Moreover, Kasza and Hayashi (2001) have reported micro

scale of shape and surface of ice slurry. Kasza and Hayashi

(2001) and Cecilia Hägg (2005) explained that ice slurry

with big diameter and rough surface had bigger pressure

drop.

8. DISCUSSION

Research on sea-water ice slurry characteristics, including

the formation, flow, heat transfer, and others related topics

is necessary in order to develop ice slurry generator. A

compact and cheap product of ice slurry generator is

supposed can improve fishing productivity, especially for

traditional fishermen in Indonesia. Using appropriate

product, fishermen can work at more far location from coast

line with longer time because they can produce ice slurry

while fishing. Research and implementation of sea-water

ice slurry generator have been doing by Lab. of

Refrigeration Engineering, Department of Mechanical

Engineering, Universitas Indonesia. Optimizing in some

important components and topics, such as evaporator,

auger, scraper, flow, and energy source, is necessary in

developing product.

ACKNOWLEDGMENT

The work described in this paper was supported by grants

of Hibah Madya 2012 from DRPM University of Indonesia

and Hibah Pengabdian Masyarakat (Community

Engagement Grant) 2012 from DRPM University of

Indonesia.

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[48] Wang, M.J., Goldstein, V., 2003. Ice slurry: advanced

fish chilling and preservation technology. Am. Fis

Soc. Symp. 38, 379e386

[49] Wibisono, M.S. Pengantar Ilmu Kelautan. Jakarta: PT.

Grasindo.2004

[50] Wang, M.J., Kusumoto, N., 2001. Ice slurry based

thermal storage in multifunctional buildings. Heat

Mass Transfer 37, 597e604.

[51] Kementrian Kelautan dan Perikanan (www.kkp.go.id)

[52] Badan Pusat Statistik (2008). Jumlah Kapal/Perahu

Berdasarkan Daerah Tangkapan

[53] .


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Improving Hydrogen Storage Capacity on Lithium-Doped Carbon

Nanotubes Using Molecular Dynamics Simulation

Nasruddina, Engkos A. Kosasih

a, Supriyadi

b and Abdul Jabbar

a

aFaculty of Engineering


Tel : (021) 7270011 ext 51. Fax : (021) 7270077

e-mail : [email protected] bFaculty of Engineering

Trisakti University of Indonesia, Jakarta 11440

Tel : (021) 5663232 ext 8431. Fax : (021) 5665840

e-mail : [email protected]

ABSTRACT

The technical challenge for the coming years is to found new

alternative energy sources that are environmentally friendly

and continuously renewable. Hydrogen has been under

investigation for its potential use as an alternative energy,

because of non-polluting and high thermal efficiency.

Hydrogen has a wide range of applications, as both stationary

and mobile use. The use of hydrogen require development of

efficient storage method. Adsorption method is considered as

a very efficient and safety. For this purpose, to improve the

performance of carbon nanotubes have attracted a lot of

interest due to their high volume and pore size distribution.

Experimental research on carbon nanotubes generally still too

expensive, it is necessary to be supported by another method

ie Molecular Dynamics Simulation. The use of metal doping

to increase the capacity of hydrogen storage has been done,

and in this study will be conducted using lithium. The aim of

this work is to build and to make two simulation models

using LAMMPS to study hydrogen storage capacity on

lithium-doped and without lithium-doped carbon nanotubes.

Doping metal presentation will focus on weight, position, and

distribution of elements on the carbon nanotubes substitution

and its influence on adsorption of hydrogen. Result of two

models simulation can be compared, for the simulation with

temperature range 253 to 293 K in the range pressure from 1

to 12 atm, hydrogen storage capacity of li-doped SWNT can

be enhanced significantly.

Keywords : CNT, hydrogen, adsorption, metal doping,

molecular dynamics, LAMMPS

1. INTRODUCTION

The generation and utilization of clean and renewable energy

is one of major worldwide concern. Currently, energy

consumption is intimately linked with CO2 emissions, a

significant human contributor to climate change [1].

Hydrogen is regarded as a pollution-free energy carrier.

Many investigations have been therefore carried out in recent

years for the utilization of hydrogen as an effective and cheap

storage system. Hydrogen adsorption in porous solids is one

of the alternatives that has been investigated [2]. Different

types of carbon nanostructures were investigated for their

suitability as hydrogen storage material. In addition to

compressed hydrogen gas and liquid H2 vessels, hydrides,

chemical hydrides or light element for hydrogen storage [3].

To obtain the maximum storage capacity of carbon nanotubes

is still interested for many researchers. Dilon, 1997,

published on the SWNT hydrogen storage capacity is not

pure, at room temperature and moderate pressure. By using

thermal desorption spectroscopy is obtained 12.01 wt%.

Based on the results obtained Dilon, Hirscher predict if the

SWNT was purified, the hydrogen storage capacity will

increase to about 5-10 wt%. Measurement of hydrogen

storage capacity in this type of material, which has been

published in various journals ranging from 0.1 wt% to 67

wt%. Spectacular results up to 67 wt%, far exceeding the

target set by DOE in the amount of 6.5 wt% or by 62 kgH2.

M-3

, performed by Baker and Rodriguez on the type of

carbon nanofiber at a pressure of 110 atm, at room

temperature and monitored for 24 hours [4].

Callejas and colleagues (2004) conducted a study to improve

the hydrogen storage capacity of SWNT pasa through the

reduction of the sample. SWNT produced by arc-discharge

modified using Ni/Y as a catalyst with a different

percentage. SWNT are synthesized by using electric-arc-

discharge method has a ratio of metal elements and catalyst

2/0, 5 and 4/1 (Ni / Y). Next examined the metal content and

value BETnya and obtained: Ni / Y 2/0, 5 (266), Ni / Y 2/0, 5

+ 3500C / l at (728), Ni / Y 4/1 (207), Ni / Y 4/1 + 3500C / l

at (585) [5].

The effects of thermal treatments and palladium loading on

sorption characteristics of single-walled carbon nanotube

(SWCNT) samples were investigated by Kocabas et al

(2008). The thermal treatment experiments were carried out

in a temperature range of 300–8000C. The sorption

characteristics of nitrogen and hydrogen on the original, heat

treated and the palladium loaded samples were investigated.

The highest surface area samples obtained at 5750C were

loaded by 3.3, 6.3 and 10.1 wt% palladium and hydrogen

adsorption isotherms on these samples were obtained at 77.4

K. The hydrogen sorption capacities of the original and the

10.1 wt% palladium loaded samples were found to be 0.76

and 1.66 wt%, respectively [6].

Boron substitution in carbon nanotubes is investigated by

Sankaran et al (2008), maximum storage capacity of 2 wt% at

80 bar pressure is obtained for BCNT1, whereas pure carbon

nanotubes shows 0.6 wt% [2]. Shevlin and Guo (2008)

performed ab initio density functional theory simulations on

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2012

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titanium-atom dopants adsorbed on the native defects of an

(8,0) nanotube. Adsorption on a vacancy strongly binds

titanium, preventing nanoparticle coalescence (a major issue

for atomic dopants). The defect-modulated Ti adsorbs five H2

molecules with H2 binding energies in the range from -0.2 to

-0.7 eV/H22 , desirable for practical applications. From their

simulations is obtained indicate that this complex is stable at

room temperature, and simulation of a C112Ti16H160 unit cell

finds that a structure with 7.1 wt % hydrogen storage is stable

[7, 8].

Zubizarreta et al (2009) investigated the effect of nickel

distribution and content in Ni-doped carbon nanospheres on

hydrogen storage capacity under conditions of moderate

temperature and pressure. It was found that the nickel

distribution, obtained by using different doping techniques

and conditions, has a noticeable influence on hydrogen

storage capacity. It was found a higher storage capacity in

samples containing 5 wt.% of Ni. This is due to the greater

interactions between the nickel and the support that produce a

higher activation of the solid through a spillover effect [9].

Hydrogen storage by chemisorption on MWCNTs was

studied by Zuttela et al (2010). By oxidation treatment to

produce defects and subsequent loading with a Pd-Ni catalyst

significantly increased the hydrogen storage capacity up to

6.6 wt%. In the same manner Meiyan et al (2010)

investigated Li-doped charged SWNTs [10].

2. POTENTIAL AND INTERACTION MODELS Molecular dynamics simulation based on statistic mechanics

and statistic thermodynamics to simulate the particles‘

interactions and consists of several processing methods, such

as trajectories, position updates, the cut-off radius, and the

initial condition. Hydrogen-hydrogen and hydrogen-carbon

interactions are both modeled with Lennard-Jones potential.

For a pair of particles i and j separated by the distance r , the

interaction between them is given by [11, 12]:

(1)

where i and j denote hydrogen, carbon or lithium particles,

ε/k and σ are the energy and size potential parameter, are

obtained from the literature., which are 0.3158 K and 2.915 Å

for hydrogen, 0.026 K and 2.27 Å for lithium and 0.2327 K

and 3.4 Å for carbon, respectively. The cross interaction

parameters σij and εij are obtained from Lorentz-Berthelot

mixing rules [12].

(2)

Diameter of nanotubes can be calculated using simple

formula,

(3)

The number of hydrogen molecules, besides of simulation

box volume, its depends on the external temperature and

pressure. Hydrogen is real gaseous, relation between

volume, pressure and temperature can be expressed by

equation real gas:

(4)

Where P is the absolute pressure of gas (atm), V is volume

(m3), n is number of moles (mol), R is the ideal gas constant

equal to 8.314 (J/mol.K), T is temperature gas (K), a and b is

van der waals constant. For example, using this formula,

simulation at 5 atm and 253 K , the number of hydrogen

molecules is 160.

3. SIMULATION METHOD

We investigated two models of simulation. The first model,

simulation in (10,10) armchair SWNT with diameter about

1.375 nm and 1.122 nm length. The SWNT was simulated

using molecular dynamics at 253 – 293 K temperature for

pressure ranging 1 atmosphere to 18 atmospheres. In the

simulation, the nanotubes are placed in a 5.5385 x 6.413 x

12.243 nm simulation box.

In the second model, we build (8,8) armchair SWNT with

diameter about 1.155 nm and 3 nm length. The SWNT was

simulated using molecular dynamics at 253 – 453 K

temperature for pressure ranging 1 atmosphere to 11

atmospheres. Figure 3.1 shows the initial condition of atoms.

In the simulation, the nanotubes are placed in a 6.9625 x

6.9625 x 22.6915 nm simulation box. The lithium atoms are

placed surrounding the nanotubes. Proportion number of

atoms is 415 atoms carbon, 35 atoms lithium and 16 – 352

atoms hydrogen, respectively.

Periodic boundary conditions on the position of the atoms

are use in all directions to eliminate surface effect [6].

Figure 1. Initial position of the simulated system, atoms

C (cyan), Li (purple) and H (green).

In this study molecular dynamics simulations performed on

the with a variety chirality, diameter and length of CNT,

while the doping element used Lithium. The results are then

used to construct models of the CNTs. Interactions between

atoms in the simulation is computed using the Lennard-Jones

potential, the Coulombic and van der Waals forces. Lennard-

Jones potential dominates at the short distances so that the

distance between atoms in excess of the cut-off radius will be

ignored, while for the coulomb force is more dominant at

longer distances, so long as the charge between atoms is still

significant value calculated despite the relatively large

distance between the atoms.


2012

253

Figure 2. Final position of the simulated system, atoms

C (purple), Li (green) and H (cyan).

Purpose molecular dynamics simulation is to update the

position, direction and speed at any time, due to hit each

other, push each other, due to the forces acting on each

particle. The condition is most difficult is determining the

initial conditions that simulated the entire atom. Position,

distance and tolerance should be adjusted to the conditions of

system that exists. In this simulation to determine the initial

conditions through several stages. To specify one or several

groups of molecules made in the program Avogadro, the

coordinates of all atoms are stored in the form file.pdb, while

to build a CNT can use a variety of software including

Generator CNT, CNT Wrapping and VMD. In this simulation

using the existing facilities Nano Builder in VMD.

Coordinates of all atoms obtained is stored in the form of

file.pdb. Furthermore, to determine how many molecules are

desired, the position where and how much tolerance between

molecules, used packmol program.

To update the position of each atom is run by software

LAMMPS (Large-Scale Atomic / Molecular Massively

Parallel Simulator). Input from this program in the form of

the main program and each of the data presented in notepad

++ format. LAMMPS program is only doing calculations to

update the position, whereas longer required to display the

results of visualization programs such as Atom Eyes, Pizza,

and VMD. In this work VMD program was used as a tool for

visualization


We make two simulation models of hydrogen adsorption in

SWNT. System conditions is emphasized on the effect of

temperature and pressure in the simulation space. For the first

model, system temperatures ranged from 253 K to 293 K

with a pressure of 1 to 18 atm. In this variation to change the

pressure or temperature is done by varying the amount of

hydrogen or can also be done by changing the volume of the

simulation box. For the second model, system temperatures

ranged from 253 K to 453 K with a pressure of 1 to 12 atm.

In this variation to change the pressure or temperature is done

by varying the amount of hydrogen or can also be done by

changing the volume of the simulation box. Hydrogen

storage capacity of SWNTs were observed for 700,000 until

1,000,000 running steps.

Table 1. Storage capacity (% weight) at first model simulation

Storage (% weight)

Pressure Temperature (K)

(atm) 253 273 293

1 0.14 0.05 0.05

2 0.14 0.05 0.18

4 0.32 0.23 0.18

6 0.46 0.28 0.18

8 0.42 0.46 0.28

10 0.60 0.51 0.46

12 0.87 0.78 0.69

14 1.01 0.78 0.55

16 1.05 0.64 0.69

18 1.14 0.87 0.78

Table 1 shows the hydrogen storage in SWNT Li-doped at

pressure in the range between 1 atmosphere to 12

atmospheres and temperature between 253 to 453 K.Table 2

shows the hydrogen storage in Li-doped SWNT at pressure in

the range between 1 atm to 12 atm and temperature between

253 to 453 K.

Table 2. Storage capacity (% weight) at second model simulation.

Storage (% weight)


(atm) 253 273 293 313 333

1 0.35 0.31 0.27 0.27 0.24

2 0.74 0.70 0.66 0.47 0.43

4 0.97 0.82 0.66 0.63 0.59

6 1.43 1.20 1.17 1.17 1.05

8 2.49 2.19 2.08 1.89 1.74

10 2.75 2.42 2.19 2.16 2.16

12 3.38 3.09 3.42 3.16 2.72

Table 2. (continued)

Storage (% weight)


(atm) 353 373 393 413 433

1 0.24 0.20 0.20 0.20 0.16

2 0.43 0.43 0.39 0.39 0.35

4 0.59 0.51 0.51 0.47 0.35

6 0.97 1.05 0.97 0.86 0.86

8 1.63 1.63 1.51 1.36 1.32

10 2.12 2.31 1.89 2.04 1.51

12 2.68 2.34 2.04 1.93 2.16


2012

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Figure 3 presents results for hydrogen storage in CNT

without Li-doped in the range from 253 K until 293 K and

pressures in the range from 1 to 18 atm.

Figure 3. The hydrogen adsorption in various temperature

and pressure (model 1)

Figure 4 and Figure 5 presents results for hydrogen storage in

Li-doped CNTs in the range from 253 K until 453 K and

pressures in the range from 1 to 12 atm. From three tables

above, it can be seen that the hydrogen adsorption increased

with a decreasing temperature at constant pressure, and at

constant temperature the greater pressure the greater

hydrogen storage capacity will be obtained. In order to

increase the hydrogen storage at a constant temperature the

pressure should be increased, and similarly, to increase the

hydrogen storage at a constant pressure, the temperature

should be decreased.

Figure 4. The hydrogen adsorption in various

temperature (model 2).

Figure 5. Variation of hydrogen storage at various pressures

(a)

(b)

(c)

Figure 6. Storage capacity (% wt) in Li-/Non-Li-Doped

SWNT. (a). Simulation at temperature 253 K, (b).

Simulation at temperature 273 K and (c).

Simulation at temperature 293 K

Figure 6. (a), (b) and (c), show that hydrogen storage

capacity of Li-doped SWNT can be enhanced significantly. It

can be said that doping effect of the second model is better

than the first model.

Besides that, several parameters , diameter, length and metal

doping are the dominant parameters that influence the

hydrogen storage capacity.

4. CONCLUSION

Generally the second model, Li-doped SWNT, in the range

from 253 K until 453 K and pressures in the range from 1 to

12 atm, has better storage capacity than the model without

Li-Doped SWNT. Diameter, length and metal doping are the

dominant parameters that influence the hydrogen storage

capacity.


2012

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REFERENCES

[1] S.A.Shevlin and Z.X.Guo, High-Capacity Room-

Temperature Hydrogen Storage in Carbon Nanotubes

via defect-Modulated Titanium Doping, J. Phys. Chem. C

2008, 112, 17456-17464.

[2] A. Toprak and T. Kopac, Surface and Hydrogen Sorption

Characteristics of Various Activated Carbons Developed

from Rat Coal Mine (Zonguldak) and Antrachite,

Separation Science and Engineering. Chinese Journal of

Chemical Engineering, 19(6) 931-937 (2011).

[3] M. Hentsche, H.Hermann, D.Lindackers, and G. Seifert,

Microstructure and low-temperature hydrogen storage

capacty of ball-milled graphite, International Journal of

Hydrogen Energy 32 (2007) 1530 - 1536.

[4] Sang-Hun Nam, Seong H. J, Soon-Bo Lee, and Jin-Hyo

Boo, Investigation of hydrogen adsorption on single wall

carbon nanotubes, Physics Procedia 32 (2012) 279 – 284

[5]. Callejas, M.A., et al, Hydrogen adsorption studies on

single wall carbon nanotubes, Carbon 42 (2004) 1243 –

1248

[6]. Kocabas, S., Kopac, T., Dogu G., Effect of thermal

treatments and palladium loading on hydrogen sorption

characteristics of single-walled carbon nanotubes,

International Journals of Hydrogen Energy 33 (2008)

1693 – 1699.

[7]. M. Sankaran, B. Viswanathan, S. Srinivasa Murthy,

Boron substituted Carbon Nanotubes - How appropriate

are they for hydrogen storage?, International Journal of

Hydrogen Energy 33 (2008), 393 – 403.

[8]. S. A. Shevlin and Z. X. Guo, High-Capacity Room-

Temperature Hydrogen Storage in Carbon Nanotubes

via Defect-Modulated Titanium Doping, J. Phys. Chem.

C 2008, 112, 17456 – 17464

[9]. L. Zubizarreta, J.A. Menéndez, J.J. Pis, A. Arenillas,

Improving hydrogen storage in Ni-doped carbon

nanospheres, International Journal of Hydrogen Energy

34 (2009) 3070 – 3076

[10]. A. Zuttela, Ch. Nutzenadela, P. Sudana, Ph. Maurona,

Hydrogen sorption by CNT and other carbon

nanostructures, Journal of Alloys and Compounds 330–

332 (2002) 676–682

[11]. C. Gu, G.H. Gao, Y.X. Yu, and Z.Q. Mao, Simulation study of

hydrogen storage in single walled carbon nanotubes,

International Journal of Hydrogen Energy 26 (2001) 691 -

696.

[12]. Banerjee, S., Molecular Simulation of Nanoscale

Transport Phenomena, Desertasi, Virginia Polytechnic

Institute and State, 2008.


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Performance of Thermoelectric and Heat Pipe Refrigerator Cooling

System

Ali A. Sungkara, Firman Ikhsana, Saripudinb, M.Afin Faisola, M. Zilvan Beya, Nandy Putraa

aDepartment of Mechanical Engineering, bDepartment of Electrical Engineering

Faculty of Engineering Universitas Indonesia, Depok 16424

Tel : (021) 7270011 ext 51. Fax : (021) 7270077


ABSTRACT

Most of the refrigerators commonly use the conventional

refrigeration system known as Vapor Compression

Refrigeration System becoming a big issue lately due to

ozone depleting substance it uses as the refrigerant. This

paper will shows step by step of an experiment with the

objective of constructing a refrigeration system based on

thermoelectric which is reliable and compete able with the

Vapor Compression Refrigeration System. The designing of

this refrigeration system shows attention to the environment

that is combined with the knowledge so the environmental

friendly technology can be applied. The performance of

thermoelectric refrigerator was conducted in variation

input power (40Watt, 72Watt, and 120Watt) and operated

in ambient temperature and cooling load of water 1000mL

to investigate the characteristic of system, the performance,

and also the COP values. The COP values is decrease

increasing of cooling load, QL. The best actual COP is

0.182 reached when the refrigerator operated at input power

40W. The result, it showed that decreasing of temperature

ambient affects the decreasing of cabin temperature.

Thermoelectric and heatpipe refrigerator cooling system

can reach cabin temperature with power 120 Watt (8.73A,

14V) produces temperature of compartment is 10.63˚C

indicates effective performance work-based thermoelectric

applications. Keywords: heat pipes, thermoelectric, thermoelectric

refrigerator, COP, environmental friendly.

1. INTRODUCTION

In engineering and applications, we have known cooling

system using vapor compression refrigeration system and

absorption refrigeration system. Vapor compression

refrigeration system, which is the most common used

refrigerating system has the advantages due to its high

cooling capacities and COP values, but this system has an

environment issues due to the ozone depletion causes by it

refrigerant, and also its unstable working temperature[1].

Absorption system has the advantage of its un noisy

working condition because this system is a-Heat driven

refrigerating system. Absorption system can use an exhaust

heat from other system as the heat generator so it can

increase its efficiency. Absorption system has some

shortages due to its low COP values and bulky design[2].

Thermoelectric system has the advantages that it is a solid

state device which is very practical use and there is no work

fluid applied that makes it environmentally friendly, the

working temperature of this system is stable, easy to

control, and the reliability of system is high so it has a long

life time. Nevertheless, this system shows a low COP

values and expensive[1]. Thermoelectric technology has

been widely used for both cooling and power generation[2].

Thermoelectric cooling have been applied in wide range of

application, from the portable vaccine carrier box, surgical

device, cooling system for electronic equipment, food

processing equipment, military & aerospace instruments[3-

11]. Thermoelectric device has a given operating

temperature range beyond which its operation may cease.

For this reason, all thermoelectric coolers (TECs) require

heat sinks in order to dissipate the energy generated or

absorbed at the two junctions. Design and selection of a

heat sink is crucial to the overall operation of a

thermoelectric system, the heat sink should be design to

minimize the thermal resistance. The heat transfer between

thermoelectric module and heat dissipation device may be

further improved by use of the heat pipe [12]. Heat pipe is a

device with a very high thermal conductivity and typically

consist of a sealed tube with an internal wick. The heat pipe

is charged with refrigerant, such as water, ethanol or

methanol and nano fluids. Heat pipes are widely adopted

for their high efficiency, cooling capability, reliability and

shape flexibility [13-17].

II. PRINCIPAL OF WORK AND

CALCULATION

2.1 Thermoelectric

Each thermoelectric module consists of two or more

semiconductor which is connected electrically in series and

thermally in parallel. All of the thermoelectric elements are

attached to a pair of ceramic substances. This ceramic

substances act as the body that mechanically holds the

structure of the connections and as the insulator electrically.

The semiconductors used are the ―N‖ and ―P‖ type made of

Bismuth Telluride. The arrangement of these

semiconductors causes the heat transfer when a current

flows between upper and lower ceramic trough every

semiconductor element P and N. The ―N‖ type

semiconductor that has been doted becomes surplus of

electrons while the ―P‖ type semiconductor that has been

doted becomes lack of electrons. The surplus of electron at

the ―N‖ type and the hole produced by the lack of electron

at the ―P‖ type semiconductors become the way of

IMAT-UI 042



2012

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transferring heat in the thermoelectric materials. Figure 1

shows the heat transfer caused by the electrical current (I)

that

Figure 1. Working scheme of Thermoelectric (Phillip C.

Watts, 2011)

NOMENCLATURE

Qc = Heat pumping capacity (J)

P = Input power (Watt)

N = Number of termocouple

α = Seebeck coefficient (V/K)

G = Geometry factor-area/length of thermoelectric

(cm)

ρ = Resisitivity (ohm cm)

K = Thermal conductivity (W/m.K)

I = Electrical current (A)

Tc = Cold side temperature (°C)

Th = Hot side temperature (°C)

Tm = Average temperature (°C), where

Tm = (°C)

(1)

Z = Figure of merit (K-1

)

∆T = Temperature difference of cold side temperature

and hot side temperature (°C)

applied to the thermoelectric module. Most of

thermoelectric made with an equal number of ―P‖ and ―N‖

type semiconductors in couple.

.

Figure 2. Electric current scheme in thermoelectric

(source: http://www.tec-microsystems.com)

Figure 2 shows the electron flow from the ―P‖ type

semiconductors that lacks of energy, absorbing heat from

cooled space and then electron flows to the ―N‖ type

semiconductors. The ―N‖ type semiconductor became

surplus of energy and dissipates the excess energy to the

environment. The amount of heat flux (heat pumped by

thermoelectric) will equal to DC current flow which is

applied. Controlling the amount of DC current, we can

control the heat flow and temperature as well.

2.2 Heat Pipes

Heat pipes generally consisting of a tube are a sealed tube

at both ends. The tube is made of metal which can absorb

and deliver heat (thermo conductive metal) very well; such

as aluminum or copper. Inside this tube contains a liquid

coolant (such as water, ethanol, or mercury) and a number

of gas from the liquid. On the side of the inner tube there is

an axis with the character of capillary walls that serves to

drain the steam produced by cooling liquid that evaporates

due to receive the amount of heat. Heat pipe consists of

three parts: the evaporator is located at one end, where the

heat is absorbed and the liquid is evaporated, then

condenser at the other end, where the vapor condensed and

the heat is released; final adiabatic section located between

the two. Adiabatic is a state where there is no heat transfer

to or from the surrounding environment. Adiabatic can

occur under two possibilities: the system perfectly isolated,

or the temperature inside and outside the same.

Figure 3. Working scheme of heat pipe

(source : http://heatpipe.nl)

2.3 COP Calculation

COP is ratio of work or useful output to the amount of work

or energy . There are two equation used to determine the

COP values for thermoelectric.

(2)

.....(3)

(4)

III. EXPERIMENTAL SET UP

Determining the characteristic of thermoelectric

refrigeration system, the experiment was conducted to

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http://www.businessdictionary.com/definition/input.html


2012

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collect the data from every important spot from system

Testing was conducted to obtain performance data to

determine the performance of refrigeration systems work

performance. The test procedure is the thermocouples are

placed on each side of the thermoelectric, cabins, and

outside the cabin as a temperature sensor. The whole

thermocouple is connected to the module is used as a

National Instrument data acquisition. Later, it connected to

a computer. Perform set-up and testing of the software

labview each sensor to ensure that the sensors are working

normally. Monitor the temperature and wait until all sensors

temperatures approaching ambient temperature. Connecting

the power cable thermoelectric cooling system with power

supply as the source of power used. After all the above

procedures have been implemented, data collection is done

with the initial conditions the system off for five minutes.

Furthermore, the power supply is turned on so that the

system works. Data is collected for four hours, including

five minutes of the initial conditions. Figure 3 is the scheme

of experiment.

Figure 4. Design of experiment performance of

Thermoelectric and Heat Pipes Refrigerator

Figure5. Construction of Cooling System

From figure 5, can be seen that thermoelectric system

consists of several components that have different

functions. The following will explain the components used

in thermoelectric systems are:

1. Fan brushless DC 12 V

2. Heat Pipes PC Cooler

3. Thermoelectrics-Peltier element

4. Heat sink

Source of six thermoelectric cooling uses the Peltier

elements are arranged in series. Heat removal process is

done by pumping heat in the form of a fan with power

13Watt for 6 fans. Heat flows through the air as

intermediaries and excreted through the bulk head under the

system. The end result made the development of tools that

can be seen in Figure 6.

(a)

(b)

Figure 6. The experiment of thermoelectric and heat pipes

refrigerator. (a) in front of view, (b) behind view.

1. Thermoelectric and heat pipes refrigerator

2. Power supply

3. Thermocouple (sensor of temperature)

4. Module DAQ data acquisition National

Instrument

5. PC (data display)


2012

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IV. RESULT

4.1 Performance of thermoelectric cooling system

temperature cold and hot side

Figure 7. Measurement temperature cooling systems cold

and hot side at variation of input power by Flir i50

thermograph

Figure 7 shows the thermographic (temperature

distribution) of thermoelectric placed on heatsink and

heatpipe with a variety of voltage and current supplied

power supply.

1. Figure 1.1 at 8V, 1.3A shows the temperature, Thotside =

60˚C and Tcoldside = -11.4˚C,

2. Figure 1.2 at 12V, 2.93A shows the temperature, Thotside

= 40.10˚C and Tcoldside = -8.50˚C, and

3. Figure 1.3 at 14V, 3.36A shows the temperature, Thotside

= 48.70 C and Tcoldside = -13.60˚C

It explains that the thermoelectric works by controlling the

temperature difference. In applications as a cooling system

is needed both in terms of performance so that will be

utilized thermoelectric the lower temperature. Higher

temperature will be controlled so that the lower temperature

can reach the minimum temperatur side. Temperature

control is optimized to move higher into the environment

using the principles of forced convection heat transfer by

fans.

4.2 Experimental Result of Thermoelectric and Heat

Pipes Refrigerator

The experiment to determine performance of heat pipes

and thermoelectric refrigerator was conducted in variation

of the input power 40 Watts (8.2V, 4.9A), 72 Watt (11.4V,

6.29A), 120 Watt (14V, 8.73A) with 1000mL water cooling

loads and operated at ambient temperatures range 27-30˚C.

4.2.1 Performance of Refrigerator Operated in Input

Power 40W

Figure 8 shows the cabin temperature conditions and water

temperature of the cooling load 1000mL at ambient

temperature conditions. It can be seen that the refrigerator

can be operated on low power. Within 35 minutes the

temperature decreased significantly from the cabin

temperature ambient. Then, the decrease of temperature

fluctuation is small, ranging from 0.01 to 0.03˚C. After 2

hours of operation, the cabin temperature increase in

fluctuates around 2˚C due to higher ambient temperatures.

Ambient temperature increases due to heat from the cooling

system is wasted to the environment and cause the

temperature around the test increased slightly. In the end,

the cabin temperature reached at T = 16.06˚C. Meanwhile,

the temperature of the water cooling loads tend to decrease

linearly up to 150 minutes of operating time. Then the

temperature tends to fluctuate with the small end

temperature reached 15.3˚C. The end of the cabin

temperature slightly greater differences compared to the

final temperature of the cooling load due to forced

convection in the have ability to maintain than the air

temperature. It means that water have a good stability than

air by fluctuation of temperature.

Figure 8. Performance of Thermoelectric and Heat Pipes

Refrigerator in various input power 40W


Power 72W

Figure 9 shows the performance conditions refrigerator

operated at input power 72Watt with 1000mL water cooling

load remains at a stable ambient temperature conditions

ranging from 27-28˚C. It can be seen that. cabin

temperature has decreased dramatically in the first hours of

operation and then remained stable until the temperature

reaches the end of 12.67˚C. In addition, the cooling load

temperature chart also had a significant reduction in cabin

temperature approaching 160 minutes of operation and

produce a final temperature that is slightly above the

temperature of the cabin 13.17˚C. This suggests that the

performance refregerator operated at 72W power has a

higher temperature difference than the power of 40W at

nearly the same ambient temperature conditions.

Figure 9. Performance of Thermoelectric and Heat

Pipes Refrigerator in various input power

72W

1.1 1.2 1.3


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Power 120W

Figure 10. Performance of Thermoelectric and Heat

Pipes Refrigerator in various input power

120W

Figure 10 shows the chart performance of heat pipes and

thermoelectric refrigerator with input power of 120 W at

ambient temperature of 27-28˚C. It explains that the cabin

temperature has decreased significantly during the 40

minutes at the start of the operation. Then, tend to be stable

until it reaches the end temperature is 10.63˚C. Decreasing

of cabin temperature is proportional in decrease of water

cooling load temperature. It decreases linearly until reaches

the end of the temperature is 10.96˚C. Decreasing of

temperature in both conditions is the largest compared to

operation on the input power 40W and 72W at the same

ambient temperature conditions. This illustrates that the

value of the temperature difference between the

environment and the cabin refrigerator depends on the

performance of the cooling system works through a given

input power. The input power will affect the inseide of

cabin temperature that happen heat transfer in forced

convection between coldsink and the air of the cabin.

4.3 Performance of Cabin Temperature in Variation of

Input Power

Figure10. Graph testing performances of cabin temperature

the power variation

Figure 10 above shows the temperature conditions in the

test cabin refrigerator with the power variation. Testing was

done by giving the power 40, 72 and 120Watt on the

refrigerator, with environmental temperature 28°C and

loaded by water with a volume of 1000ml. By looking at

these graphs can be concluded that the higher power that is

given then the lower the cabin temperature. As said in

previous discussions, this happens in the cabin heat transfer

by forced convection.

4.4. Result of COP Calculation

COP result calculation is given below show that the

performance of cabin temperature due to variation of

cooling load and ambient temperature.

Table 1. COP Value of thermoelectric and heat pipes

Refrigerator

abased on equation 4 bbased on equation 2 with the Qc values was based on equation 3

Table 1 shows performance comparison of thermoelectric

refrigerator in input power variation of running condition.

The COP values are decrease with increasing the input

power. The decreasing of actual COP values is caused by

the value of Qc that is tend to increase due to increasing of

cooling load, QL while the input power applied stay in small

difference of cooling capacity. It means the highest COP

values (optimum COP and actual COP) reached when the

refrigerator operated at input power 40W in constant

1000mLwater cooling load. The lowest of actual COP when

operated at input power 120W and cabin temperature is

10.63oC.

ACKNOWLEDGMENT

We would appreciate Assistant of Professor Ridho

Irwansyah, S.T., M.T. dan Wayan Nata, S.T., M.T. at the

Applied Heat Transfer Laboratory Faculty of Engineering

Universitas Indonesia for helpful advices and kindly

providing us with thermoelectric and its measurement

technique. We are also thankful to Dikti by the project in

Students Creativity Programme for funding this research

and member of Universitas Indonesia‘s Robotics Team for

technical support.

V. CONCLUSION

1. Decreasing of temperature ambient affects the

decreasing of cabin temperature

2. Increasing of input power applied to system causes the

increasing temperature difference of system so that it


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affects the cabin and cooling load temperature of

thermoelectric and heatpipes refrigerator.

3. The measurements of cabin temperature with power 120

Watt (14V, 8.73A) produces temperature of

compartment is 10.63˚C in steady temperature indicates

effective performance work-based thermoelectric and

heat pipe cooling applications.

4. The COP values is decrease caused by the value of Qc

that is tend to increase due to increasing of cooling load,

QL. The highest COP values (optimum COP and actual

COP) reached when the refrigerator operated at input

power 40W in constant 1000mLwater cooling load.

5. As the result, we can see that thermoelectric and heat

pipes refrigerator has potency to replace the

conventional refrigerators. Non-CFC refrigeration

products could be replacement of the box of

conventional cooling or heating due its ability to keep

the temperature steady so that the materials can be

stored in optimal conditions.

REFERENCES [1] S.B. Riffat, Gouquan Qiu, Comperative investigation of

thermoelectric air-conditioners versus vapour compression

and absorption air-conditioners,Applied Thermal

Engineering 24 (2004) 1979-1993.

[2] Kong Hoon Lee, Ook Joong Kim, Analysis on the cooling

performance of the thermoelectric micro-cooler,

International Journal of Heat and Mass Transfer 50 (2007)

1982-1992.

[3] Nandy Putra, Desing, Manufacturing And Testing Of A

Portable Vaccine Carrier Box Employing Thermoelectric

Module And Heat Pipe, Journal of Medical Engineering &

Technology, 33 (2009) 232-237.

[4] Nandy Putra et al, The characterization of cascade

thermoelectric cooler in cryosurgery device, Cryogenics 50

(2010) 729-764

[5] S.B. Riffat, Xiaoli Ma, Thermoelectrics : a review of present

and potential applications, Applied Thermal Engineering, 23

(2003) 913-935

[6] Hsiang-Sheng Huang et. al, Thermoelectric water-cooling

device applied to electronic equipment, International

Communications in Heat and Mass Transfer 37(2010) 140-

146.

[7] Rieyu Chein, G. Huang, Thermoelectric cooler application in

electronic cooling, Applied Thermal Engineering 24 (2004)

2207-2217.

[8] Miguel A. Sanz-Bobi et al, Thermoelectricity applied to the

cryoconcentration of orange juice, 15th International

Conference on Thermoelectric (1996) 259-263

[9] A. Hamilton, J.Hut, An electronic cryopore for cryosurgery

using heat pipes and thermoelectric coolers : a preliminary

report, Journal of Medical Engineering & Technology 12

(1993) 104-109.

[10] Hiroki Takeda et al, Development and estimation of novel

cryoprobe utilizing the peltier effect for precise and safe

surgery, Cryobiology 59 (2009) 272-284.

[11] M.R. Holman, S.J Rowland, Design and development of new

surgical instrument utilizing the peltier thermoelectric effect,

Journal of Medical Engineering & Technology 21 (1997)

106-110.

[12] S.B Riffat et al, A novel thermoelectric refrigeration system

employing heat pipe and phase change material : an

experimental investigation, Renewable Energy 23 (2001)

313-323

[13] Faghri A. Heat pipe science and technology, Taylor &

Francis, 1995.

[14] David Reay, P.A. Kew, Heat pipes theory, design and

applications, Elsevier, 2006.

[15] Y.H Yau, M. Ahmadzadehtalatapeh, A review on the

application of horizontal heat pipe heat exchangers in air

conditioning system in the tropics, Applied Thermal

Engineering 30 (2010) 77-84.

[16] Te-En et al, Dynamic test method for determining the thermal

performances of heat pipes, International Journal of Heat

and Mass Transfer 53 (2010) 4567-4578.

[17] Leonard L. Vasilev, Heat pipe in modern heat exchangers,

Applied Thermal Engineering 25 (2005) 1-19.


2012

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Preliminary Study on Length of Candle Filter Surface on the Flow Pattern

in Freeboard of Fluidised Bed Gasifier

A. Farhan Faudzi1*

, Kahar Osman1, Nor Fadzilah Othman

2,

Mohd Hariffin Bosrooh2

1 Computational Fluid Mechanics Laboratory, Faculty of Mechanical Engineering,

Universiti Teknologi Malaysia,

81310 UTM Skudai, Johor, Malaysia, [email protected]

2 TNB Research Sdn. Bhd.,

No. 1, Lorong Ayer Itam,

Kawasan Institusi Penyelidikan,

43000, Kajang Selangor, Malaysia,

ABSTRACT

This paper presents numerical study on flow pattern in

freeboard of fluidized bed gasifier during syngas filtration

due to different length of candle filter surface which are

0.25m, 0.30m and 0.35m. This numerical study was done

by considering two flows which are air flow and coal flow

by using ANSYS Fluent 14. The air inlet velocity flows

from previous study are 0.11m/s, 0.16m/s and 0.21m/s

whereas coal inlet velocities choose to be 10% of air flow

velocity. From the results, observed that flow from mixture

of air and coal for 0.25m of candle filter surface has

uniform distribute as flow move throughout freeboard. On

the other hand, for 0.30m and 0.35m of candle filter

surface, flow tend to become higher in velocity nearer wall

of freeboard. By increasing the flow inlet velocity, the flow

pattern in freeboard remain same as before but increase in it

magnitude. In comparison for all length of candle filter

surface in this study, 0.25m considered best as flow

produce in freeboard is more uniformly.

Keywords : Example: Freeboard fluidized bed, candle

filter, flow pattern

1. INTRODUCTION

The candle filter design much important in order to enhance

the ability to filter any impurities or contaminations that

exist in freeboard bed fluidized. The higher filtration rate

mean that higher in provide uniform flow distribution in

freeboard of fluidized bed gasifier. This is because the non-

uniformities may lead to insufficient filter cleaning [1].The

previous study shown that velocity profile not achieve

better at inlet vent and distribution of inlet velocity no fully

understood. Apart from that, flow velocity at inlet vent not

uniform distributed lead to lower usage rate of filter media.

However, a more uniform flow inlet can be obtained by

adjustment of length or angle of candle filter surface [2].

From previous study, results indicate that distribution

uniformity of flow could be improved by reducing inlet

velocity, diverging angle and particle diameter [3]. On the

other hand, changes of velocity magnitude during inlet

process also important to see as it is adversely affect

operation of plant and filter durability [4]. This research

was conducted to seek the flow uniform distribution in

freeboard of fluidized bed gasifier due to the different

length of candle filter surface.

2. CFD MODELLING DESCRIPTIONS

In this paper, the flow pattern inside the freeboard of

fluidized bed gasifier study using ANSYS Fluent 14.0 by

dimension as 2D model with viscous model of k-epsilon

model. In the multiphase model, it stated to be mixture with

2 eularian phases which are air and coal. The geometry of

freeboard of fluidized bed gasifier was simplified as in

figure 1 with varies in length of candle filter surface. The

flow of air and coal into freeboard was at bottom part which

was inlet vent. The vertical length of freeboard is 0.3m with

width of 0.05m whereas candle filter surface stated at

opposite of inlet vent with vertical length of 0.03m from

inlet vent [5].

Figure 1: 2D model for freeboard of fluidized bed

gasifier

The air inlet velocity varied at 0.11m/s, 0.16m/s and

0.21m/s meanwhile coal inlet velocity is 10% of air inlet

velocity which were 0.011m/s, 0.016m/s and 0.021m/s.

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Pressure stated as 2000pa at surface of inlet vent. The

simulation study on candle filter neglect it effect in filtering

particle as study much concern about velocity flow pattern

in freeboard of fluidized bed gasifier. Any chemical

reaction and temperature neglected.


All the results from simulation by ANSYS Fluent 14.0

presented were converged by running 40 to 80 iterations.

Figure 2 show the mixture flow pattern of air and coal in

fluidized bed gasifier with different in length of candle

filter surface of 0.25m, 0.30m and 0.35m with different

flow inlet velocity.

Figure 2 (a): Flow pattern for 0.11m/s air inlet

Figure 2 (b): Flow pattern for 0.16m/s air inlet

Figure 2 (c): Flow pattern for 0.21m/s air inlet

Figure 2: Flow velocity coloured by contour vector

From Figure 2, left show 0.30m, upper part show 0.35m

and lower part show 0.25m of candle filter surface. Candle

surface length of 0.25m show flow velocity of air and coal

not reach higher value as shown by contour vector.

Moreover, as filter media introduce in this model, the

filtering of particle become easier as flow exist nearer filter

surface. The flow pattern for 0.30m of length of candle

filter surface stated almost same with 0.25m but the flow

form has higher magnitude in velocity. However, at length

of 0.30m the velocity starts to become higher at nearer wall

of freeboard. This show that flow exist caused fluid or solid

particle exist will flow away from candle filter and filtering

process become more inefficient. For 0.35m of candle filter

surface, flow pattern more toward at nearer freeboard wall,

moreover the contour vector stated at high value.

Furthermore, flow exist less at candle filter surface and the

vortex can occur at candle filter surface due to different in

velocity profile exist. The vortex exist prove of circulating

movement which is non uniform flow which lead to low

rate usage of filter surface [6].

Meanwhile, for all length of candle filter surface as

increasing air inlet velocity, the magnitude of flow inside

freeboard bed gasifier increase. Figure 3 show the change in

magnitude of flow inside freeboard at candle filter surface

as air inlet velocity in varies. From the figure 3, green line

represent length of 0.25m, red line for 0.30m and blue line

for 0.35m of candle filter surface. The velocity in freeboard

show steady flow as it passes through candle filter then start

to produce higher flow velocity at nearer wall of freeboard

as length of candle filter surface increase.

Figure 3 (a): Magnitude of flow velocity at

0.11m/s air inlet


2012

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Figure 3 (b): Magnitude of flow velocity at 0.16m/s

air inlet

Figure 3 (c): Magnitude of flow velocity at 0.21m/s

air inlet

Figure 3: Magnitude of flow velocity at candle filter

surface with different length of candle filter surface.

4. CONCLUSION

The numerical result shown that the shorter length of candle

filter surface provide more uniform flow hence lead to high

rate of filter surface. Moreover, flow pattern exist at nearer

candle filter surface enhance filter process.

ACKNOWLEDGMENT The authors would like to thank Universiti Teknologi

Malaysia and TNB Research Sdn. Bhd. for supporting this

research activity.

REFERENCES

[1] T.G. Chuah, C.J. Withers and J.P.K. Seville,

―Prediction and measurement of the pressure and

velocity distributions in cylindrical and tapered rigid

ceramic filters‖, Separation and Purification

Technology 40, 2004, 47-60

[2] Chia-Jen Hsu and Shu-San Hsiau, ―Experiment study

of the gas flow behavior in the inlet of a granular bed

filter‖, Advanced Powder Technology 22, 2011, 741-

752

[3] Gong Jinke, Tian Chan and Wu Gang, ―Numerical

simulation on distribution characteristics of particle

distribution uniformly in a radial style diesel

particulate filter‖, Advances in Computer Science and

Engineering, 2012, 795-805

[4] S. Ito, T. Tanaka and S. Kawamura, ―Changes in

pressure loss and face velocity of ceramic candle

filters caused by reverse cleaning in hot coal gas

filtration‖, Powder Technology 100, 1998, 32-40

[5] Andrea Di Carlo and Pier Ugo Foscolo, ―Hot syngas

filtration in the fluidized bed gasifier: Development of

a CFD model‖, Powder Technology 222, 2012, 117-

130

[6] Chia-Jen Hsu, Shu-San Hsiau, Yi-Shun Chen and Jiri

Smid, ―Investigation of the gas inlet velocity

distribution in a fixed granular bed filter‖ Advanced

Powder Technology 21, 2010, 614-622


2012

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Preliminary Study on the Effect of Type Distributor Plate on Airflow

Pattern in Bubbling Fluidised Bed

Nofrizalidris Darlis1*

, Kahar Osman1, Ab Malik A. Hamat

1 , Nor Fadzilah Othman

2, Mohd

Hariffin Bosrooh2

1 Computational Fluid Mechanics Laboratory, Faculty of Mechanical Engineering,

Universiti Teknologi Malaysia,

81310 UTM Skudai, Johor, Malaysia,

[email protected]

2 TNB Research Sdn. Bhd.,

No. 1, Lorong Ayer Itam,

Kawasan Institusi Penyelidikan,

43000, Kajang Selangor, Malaysia

ABSTRACT

This paper presents numerical study on airflow pattern in

bubbling fluidized bed due to the different shapes of

distributor plate which are flat, convex and concave. The

numerical study was done at a grace of a Commercial

Computational Fluid Dynamics ANSYS Fluent 14. The air

inlet velocity based on previous studies which are 0.28m/s,

0.33m/s and 0.37m/s. From the results, observed that

airflow for flat plate distributor has uniform upward

movement. Meanwhile, concave and convex has airflow

movement upward and downward resulting circulating flow

which is good for mixing process. By increasing the air

inlet velocity, the airflow inside the bubbling fluidized bed

increase. Compared all the distributor plate in this study,

convex shape considered the best distributor plate in mixing

process.

Keywords Example: Gasification,distributor

plate,airflow

1. INTRODUCTION

The gas distribution plate is the key element in fluidization

technology. There are many types of distributor plate

differentiated by hole type, number of hole and plate shape.

Main requirement for a distributor plate are to promote

uniform distribution of fuel particle to make sure a good

chemical conversion and uniform temperature throughout

the bed [1]. In the same time, mixing pattern will affect the

performance of gasification [2]. This type of plate will

influenced the mixing pattern between air and fuel particles.

In the other hands, distributor plates also govern to the

particles circulation behavior in the gasification chamber

[3]. Although many research regarding distributor plate

have been carried out, the effect of different shape type of

the plate still in minimum study. The airflow and its

distributor primary factor influenced fluidized bed

processing [4]. The higher heating value reached its peak

value at a fluidization velocity of 0.28 m/s but remained

fairly constant at the fluidization velocities of 0.33 and 0.37

m/s [5]. Therefore this research was conducted to seek the

air distribution in bubbling fluidized bed due to the

different shapes of distributor plate which are flat, convex

and concave. The geometry of the bubbling fluidized bed

was referring to laboratory scale fluidized bed gasifier at

TNB Research Sdn. Bhd.

2. CFD MODELLING DESCRIPTIONS

In this paper, the airflow pattern inside the gasification

chamber will study by using Commercial Computational

Fluid Dynamics ANSYS Fluent 14. Standard K-epsilon

model with enhanced wall treatment was used as the model

of this study.

Figure 1: 2D geometry modeling for flat, concave and

convex distributor plate

The geometry of bubbling bed fluidized was simplified as

figure 1 and the air inlet was state at the holes plate. The

inner diameter 250mm and its bed height is 300mm [6].

The air velocity inlet is varied at 0.28m/s, 0.33m/s and

0.37m/s.

Figure 2: Types of distributor plate [1]

Figure 2 show the types of distributor and the angle of

convex and concave plat was 20 degree. The simulation

study was neglected any chemical reaction and temperature.

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All simulation results presented throughout this paper were

converged by running 800 to 1500 iterations. Figure 3 show

the air flow pattern in bubbling fluidized bed by using flat,

concave and convex distributor plate with different air inlet

velocity.

Figure 3 (a): Airflow pattern for 0.28m/s air inlet

Figure 3 (b): Airflow pattern for 0.33m/s air inlet

Figure 3 (c): Airflow pattern for 0.37m/s air inlet

Figure 3: Airflow velocity coloured by contour vector

Flat plate distributor show that the airflow uniformly

moving upward without any circulating. From the airflow

pattern, it is reasonable to say that this type of distributor

plate will give low mixing and turnover. A. E. Ghaly and K.

N. MacDonald (2012) reported that localized mixing caused

upward movement of the bubbles by flat distributor plate

was clearly evident but no bed material turnover was

observed. Besides, by using concave plate observed that

upward movement at the center and downward at near the

wall. The airflow velocity near the wall region is less than

the center. Meanwhile, convex plate type show the upward

movement beside the wall and downward at the center. The

velocity of the airflow near the wall is higher than the

center region. Both concave and convex distributors give

upward and downward movement of the air inside the

fluidized bed resulting circulation movement which is good

for the mixing process. The different speed of airflow can

be used to give more time for carbon conversion of the fuel

and also minimize the bed material from leave the reactor.

This is concurrent with past research that to improve

mixing properties of the binary mixture, which has great

tendency for segregation due to density differences, an

angled distributor plate should be used [1].

For all type of distributor, by increasing the air inlet

velocity, the airflow inside the bubbling fluidized bed

increase. Figure 4 show clearly the graph of velocity

changed for each type of distributor plate.

Figure 4 (a): Distribution of airflow velocity at

0.28m/s air inlet

Figure 4 (b): Distribution of airflow velocity at 0.33m/s air

inlet


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267

Figure 4 (a): Distribution of airflow velocity at 0.37m/s air

inlet

Figure 4: Airflow distribution at 0.05m of bed high with

different type of distributor plate

The airflow velocity change for convex distributor was

highest compared with concave and flat plate. The rapidly

drop of airflow at the center region resulting circulating

flow. Increasing circulating flow will increase the particle

turnover, thus increasing mixing efficiency.

4. CONCLUSION

In this paper, the effect of type distributor plate shape has

been studies using ANSYS Fluent 14. These numerical

results give a good agreement with the past study by A. E.

Ghaly and K. N. MacDonald (2012). Convex distributor

plate shows very good airflow movement inside the

bubbling fluidized bed reactor compared with others. It is

reasonable to say that by using convex distributor plate will

give high efficient mixing process during gasification.

ACKNOWLEDGMENT The authors would like to thank Universiti Teknologi

Malaysia and TNB Research Sdn. Bhd. for supporting this

research activity.

REFERENCES

[1] E. Ghaly and K.N. MacDonald, ―Mixing patterns and

residence time determination in a bubling fluidized

bed system‖ American Journal of Engineering and

Applied Sciences, 2012, 5 (2), 170-183

[2] Chih-Jung Chen, Chen-I Hung, and Wei-Hsin,

―Numerical investigation on performance of coal

gasification under various injection pattern in an

entrained flow gasifier,‖ Applied Energy, in press.

[3] Zbigniew Garncarek, Longin Przybylski, John S.M.

Botterill and Christopher J. Broadbent, ―A quantitative

assessment of the effect of distributor type on particle

circulation,‖ Powder Technology 91, 1997, 206-216

[4] Frederic Depypere, Jan G. Pieters and Koen

Dewttinck,―CFD analysis of air distribution in

fluidized bed equipment‖, Powder Technology 145,

2004, 176-189.

[5] Sadaka S.S., A. E. Ghaly and M. A.

Sabbah,―Development of an air-stream fluidized bed

gasifier‖, Misr Journal of Agricultural Engineering,

1998, Vol 15(1), 47-52.

[6] Nor Fadzilah Othman, Mohd Hariffin Bosrooh and

Kamsani Abdul Majid,―Partial gasification of different

types of coals in a fluidized bed gasifier‖, Jurnal

Mekanikal, 2007, 40-49


2012

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Natural Convection in A Differentially Heated Cavity

Using Splitting Method

Ubaidullah S.a, Kahar Osman

a


Universiti Teknologi Malaysia, Skudai, Johor, Malaysia


ABSTRACT

The solution of thermally-driven flow of Navier-

Stokes equation, in Boussinesq approximation, is

presented. The results are obtained using splitting

method and in good agreement with available

benchmark numerical solutions. The convection

terms are explicitly integrated using 3-step Range-

Kutta scheme. Buoyancy-driven fluid flows data in a

differentially heated cavity for low Rayleigh

numbers (R=103, R=10

4 and R=10

5) are presented.

Keywords : Natural convection, thermally driven,

splitting method

1. INTRODUCTION

Thermally-driven fluid/gas flow can be found in

many engineering applications such as room

ventilation system, heat exchanger, solar energy

collector, cooling of electronic components, factory

stack emission etc. The fluid velocity is caused by

buoyancy force generated from the change of fluid

density as a result of significantly hot fluid

temperature. This type of flows are normally

investigated to improve performance of engineering

systems, increase efficiency of ventilation systems

that may reduce the cooling energy required, predict

contaminants between buildings and other purposes.

Researches on thermally-driven flows have been

done in many areas of engineering. In solar collector

study, Fan et. al. [1] studied the standby heat loss

effects to thermal performance of a heat system by

analyzing the flow around the hot storage tank. In

construction and design of buildings, Chun. al. [2]

investigated the buoyancy-driven ventilation in a

reduced-scale building, Rakesh [3] studied the flow

in solar chimneys, Guohui [4] studied the impact of

computational domain in buoyancy –driven flow

simulation. In air pollution research, Baik and Kim

[5] investigated the urban street canyon flows with

street-heating in between two buildings. Various

solar radiation conditions in urban street canyon

have been studied by Xiaomine et. al. [6]. Rezwan

et. al. [7] investigated the building aspect ratio and

wind speed effect on temperature distribution in

urban street canyon. A lot of problems related to

thermally-driven flows can be found in the

literature. In brief, natural convection flows are still

a major interest for researchers in heat and fluid

flow studies.

There are numerous approaches to solve Navier-

Stokes equation that governs the thermally-driven

flow problems. One of the most successful

approaches nowadays is the projection methods.

Projection methods lead to easy-to-implement and

efficient algorithms by decoupling the diffusion and

convection terms of the Navier-Stokes equation.

Fractional step method and pressure correction

methods are two kinds of projection methods.

Fractional step method is based on a full splitting of

the diffusion and incompressibility constraint

(pressure) in different sub steps (Karniadakis [8]).

Full splitting suffers from erroneous solutions as a

result of improper boundary condition of pressure.

However, the pressure boundary conditions have

been discussed extensively in the literature [8], [12],

[13], [14]. Meanwhile, pressure correction methods

are based on predictor-corrector procedure between

velocity and pressure fields. Initial approximation of

pressure allows the momentum equation to be

solved without satisfying the divergence-free

constraint and requires additional pressure

correction procedure. In this method, a Poisson

equation for a new defined quantity is solved instead

of pressure Poisson equation. A homogeneous (zero)

Neumann condition need to be used to ensure

divergence-free condition of the velocity which is

not valid for pressure itself [14]. However, the final

velocity fields satisfy the divergence-free condition

for semi-discrete formulation.

A splitting algorithm based on pressure correction

method is chosen in this study for solving the

thermally-driven flow problems in a differentially

heated cavity, the standard problem used for

benchmarking new computer programs.

2. NUMERICAL MODEL

The so called Boussinesq equations to model non-

isothermal flow problems are used

(1)

(2)

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(3)

Where

R = (gβ∆Tl3)/(k/ν) (4)

Pr = ν/k (5)

are commonly used Rayleigh and Prandtl numbers

and g = (0, 1)T . Here, u, p and T are fluid velocity,

pressure and temperature respectively, g is

acceleration of gravity, β is thermal expansion

coefficient, l is characteristic length, ∆T is

characteristic temperature difference, k is thermal

diffusivity and v is kinematic viscosity of the fluid.

The equations are non-dimensionalised according to

Paolucci and Chenowith [10]. The equations are

solved using algorithm by Minev [9].

Instead of using spectral element method as in the

pressure correction method by Minev [9], present

study applied finite difference method. The

equations are discretized in time with a second order

backward difference scheme and a second order

central difference scheme for space. The algorithm

leads to two Helmholtz equations for velocity

components, and a Poisson equation for pressure.

All of them are solved implicitly in present study.

The boundary conditions are

Nusselt numbers (Nu) are calculated using simple

finite difference formula at the hot vertical wall

Nu = ∂T/∂x |x1=0 (6)

3. RESULTS

0 20 40 60 80 100 120 140 1600

20

40

60

80

100

120

140

160Streamline

5 10 15 20 25 30 355

10

15

20

25

30

35

(a) (b)

5 10 15 20 25 30 355

10

15

20

25

30

35

5 10 15 20 25 30 355

10

15

20

25

30

35

(c) (d)

Figure 1: (a) Streamlines (b) iso-U velocity

contour (c) iso-V velocity contour (d)

Isotherms for buoyancy driven flow at

Rayleigh number = 103.

0 20 40 60 80 100 120 140 1600

20

40

60

80

100

120

140

160Streamline

5 10 15 20 25 30 355

10

15

20

25

30

35

(a) (b)

5 10 15 20 25 30 355

10

15

20

25

30

35

5 10 15 20 25 30 355

10

15

20

25

30

35

(c) (d)


contour (c) iso-V velocity contour

(d) Isotherms for buoyancy driven

flow at Rayleigh number = 104.

0 20 40 60 80 100 120 140 1600

20

40

60

80

100

120

140

160Streamline Iso-U

10 20 30 40 50 60 70 80 90 100

10

20

30

40

50

60

70

80

90

100

(a) (b)

Iso-V

10 20 30 40 50 60 70 80 90 100

10

20

30

40

50

60

70

80

90

100

Isotherm

10 20 30 40 50 60 70 80 90 100

10

20

30

40

50

60

70

80

90

100

(c) (d)


contour (c) iso-V velocity contour

(d) Isotherms for buoyancy driven

flow at Rayleigh number = 105.

u1=0

u2=0

T=1

u1=0, u2=0,

dT/dn=0

u1=0

u2=0

T=0

u1=0, u2=0,

dT/dn=0

Ω

(0,0)

(0,1) (1,1)

(1,0)

x1

x2


2012

270

Table 1: Buoyancy-driven flow in an enclosed

cavity. Present results (P) compared with benchmark

numerical solution (B) by Minev [9] and the

derivation (D) for R=103, 10

4 and 10

5.

Variable Source R=103

R=104

R=105

U1,max B 3.630 12.627 34.73

P 3.675 14.893 37.12

D (%) +1.2 +12.2 +6.9

X1,max B 0.813 0.823 0.875

P 0.834 0.868 0.914

D (%) +2.6 +3.5 +4.4

U2,max B 3.693 19.617 68.59

P 4.232 19.520 58.72

D (%) +14.6 -8.0 -14.4

X2,max B 0.170 0.125 0.079

P 0.152 0.106 0.066

D (%) -10.6 -5.6 -16.5

Numax B 1.507 3.531 7.717

P 2.652 5.534 9.275

X2(Nu) B 0.08 0.143 0.080

P 0.13 0.140 0.067

Numin B 0.692 0.586 0.726

P 1.405 1.082 0.847

X2(Nu) B 1.0 1.0 1.0

P 0.98 1.0 1.0

0 2 4 6 8 100

1.0

0.75

0.5

0.25

Nusselt Number (Nu)

Y -

Coord

inate

R=104

R=105

R=103

Figure 4: Comparison of local Nusselt number

along the hot wall (X1=0).

0.25 0.5 0.75 1.00-60

-40

-20

0

20

40

60

X1 - Coordinate

U2-V

elo

city

R=105

R=103

R=104

Figure 5: Variation of vertical velocity at

X2=0.5.

-30 -20 -10 0 10 20 30

1.0

0

0.5

0.25

0.75

U1 - Velocity

X2 -

Coord

inate

R=104

R=105

R=103

Figure 6: Variation of horizontal velocity at X1=0.5.

0 1.00.50.25 0.750

0.2

0.4

0.6

0.8

1

X1 - Coordinate

Tem

pera

ture

R=105

R=104

R=103

Figure 7: Variation of temperature at mid-

height (X2=0.5).

4. DISCUSSION

The initial conditions are set as zero velocity, zero

pressure and zero temperature. The temperature on

the left cavity is then set to 1 and the algorithm is

then applied. After some initial transience, the

solutions reach steady-state values and plotted as in

the above figures. The number of grid used is 150 x

150 for all Rayleigh numbers. For transitional and

high Rayleigh number flows (R > 105), they are

unstable with present grid size. They require finer

mesh and are not presented here. The time step is

obtained experimentally and varies a lot with

Rayleigh number. Its value is 0.004 for R=103 and

become as small as 0.0004 for R=105. The algorithm

takes 1.51 s per time step on Intel Core i5 CPU (3.5

GB RAM) for second order central difference

approximation scheme.

As in figure 1, figure 2 and figure 3, the streamlines,

iso-velocity contour for both directions x and y and

the isotherms plots are in very good agreement with

the benchmark numerical solution of a differentially

heated cavity flow. Although consistent contour


2012

271

plots are obtained, errors can be found in the

position of maximum velocities and its values. The

spatial discretization of the algorithm is expected to

be the main contributing factor for errors as the

benchmark solution used high order spectral element

discretization technique. Besides, Karniadakis [8]

recommended the use of high order pressure

boundary condition especially in splitting algorithm

for low Reynolds number flow which is not being

implemented in the algorithm. In addition, present

study employed full-discrete formulation, discrete

time and space, instead of continuous projection,

discrete time with continuous space, as in Minev

[9].

As for Nusselt number, the values are very sensitive

to the point where it is calculated especially for

R=105. As in figure 4, the Nusselt number values

along the X2 axis has the same trend as Wan et. al.

[11]. For U1, U2 and temperature variation along X1

and X2 axes as in figure 5, figure 6 and figure 7,

they are also in good agreement with Wan et. al.

[11]. However, their values are shifted slightly

higher from the benchmark solution due to the

prescribed reasons mentioned above.

5. CONCLUSION

Present study has successfully investigated the

natural convection in a differentially heated square

cavity using splitting method. Iso-contours of

streamlines, horizontal velocity, vertical velocity

and temperature are the same as available

benchmark numerical solutions. Present algorithm

has slightly over-predicted the Nusselt number

values although they have relatively similar trends

along vertical hot wall as benchmark solutions.

REFERENCES

[1] J. Fan and S. Furbo, ―Buoyancy Driven Flow In

A Hot Water Tank Due To Standby Heat

Loss‖, Solar Energy, vol. 50, 2012, pp 1266-

1274.

[2] P. L. Chun, H. T. Lin, J. H. Chou, ―Evaluation

Of Buoyancy-Driven Ventilation In Atrium

Buildings Using Computational Fluid

Dynamics And Reduced-Scale Air Model‖,

Building and Environment, vol. 44, 2009, pp

1970-1979.

[3] K. Rakesh, L. Chengwang, ―Flow Reversal

Effects On Buoyancy Induced Air Flow In A

Solar Chimney‖, Solar Energy, vol. 86, 2012,

pp2783-2794.

[4] G. Guohui, ―Simulation Of Buoyancy-Driven

Natural Ventilation of Buildings – Impact Of

Computational Domain‖, Energy and

Buildings, vol 42, 2010, pp 1290-1300.

[5] J. J. Kim, J. J. Baik, "Urban street-canyon flows

with bottom heating," Atmospheric

Environment, vol 35, no. 20, 2001,

pp.3395-3404.

[6] X. Xiamoine, H. Zheng, W. Jiasong, Z. Xie,

"Impact of solar radiation and street layout

on pollutant dispersion in street canyon,"

Building and Environment, vol. 40, no. 2,

2005, pp. 201- 212.

[7] A. M. Rezwan, Y.C.L. Dennis, C. H. Liu,

"Effects of building aspect ratio and wind

speed on air temperatures in urban-

like street canyons," Building and

Environment, vol. 45, no. 1, 2010, pp.

176-188.

[8] G. E. Karniadakis, M. Israeli, S. A. Orszag,

―High Order Methods for The Incompressible

Navier-Stokes Equation,‖ International

Journal of Computational Physics, vol. 90,

1991, pp 414-443.

[9] P. D. Minev, F. N. Van De Vosse, L. J. P.

Timmermans and A. A. Van Steenhoven, ‖A

Second Order Splitting Algorithm For

Thermally-Driven Flow Problems‖,

International Journal of Numerical Method

and Fluid Flows, vol 6. No. 2, 1995, pp 51-60.

[10] S. Paolucci and D. R. Chenowith, ―Transition

To Chaos in a Differentially Heated Vertical

Cavity,‖International Journal of Fluid

Mechanic, vol. 201, 1989, pp 379-410.

[11] D. C. Wan, B. S. V. Patnaik, and G. W. Wei,

―A New Benchmark Quality Solution for the

Buoyancy-Driven Cavity by Discrete Singular

Convolution‖, Numerical Heat Transfer, vol.

40, 2001, pp 199-228.

[12] S. A. Orszag, M. Israeli and M. O. Deville,

―Boundary Conditions for Incompressible

Flows‖, Journals of Science Computing, vol 1,

1986, pp 75-111.

[13] P. M. Gresho and R. L. Sani.‖On Pressure

Boundary Conditions for the Incompressible

Navier-Stokes Equation‖, International

Journals of Numerical Methods in Fluids, vol

7, 1987, pp 1111-1145.

[14] L. J. P. Timmermans, P. D. Minev and F. N.

Van De Vosse, ―An Approximate Projection

Scheme for Incompressible Flow Using

Spectral Elements‖, International Journals of

Numerical Methods in Flu

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