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International Journal of Mechanical Engineering and Technology (IJMET), ISSN 0976 –

6340(Print), ISSN 0976 – 6359(Online) Volume 4, Issue 2, March - April (2013) © IAEME

209

THEORITICAL INVESTIGATIONS OF INJECTION PRESSURE IN A

FOUR STROKE DI DIESEL ENGINE WITH ALCOHOL AS FUEL

S.Sunil Kumar Reddy1

and Dr. V. Pandurangadu2

1 Associate Professor, Mechanical Department, N.B.K.R.I.S.T, Vidyanagar, Nellore, A.P

2 Professor, Mechanical Department, Jawaharlal Nehru Technological University, Anantapur.

A.P

ABSTRACT

An intensive search for alternate fuels is going on due to the stringent emission

legislation all over the world for diesel engines which produce more environmental pollution.

The major pollutants from these engines are oxides of nitrogen (NOx), smoke and particulate

matter. The difficulty in meeting the increasingly stringent limitations on emissions has

stimulated interest in alcohol -fueled diesel engines because it is a renewable bio-based

resource and it is oxygenated, thereby providing the potential to reduce particulate emissions

in compression–ignition engines and ethanol diffusion flames produce virtually no soot. With

the high latent heat of vaporization, alcohol absorbs heat from the combustion chamber and

makes it cools. This reduces the efficiency of the engine. So, more amount of fuel cannot be

injected in to combustion chamber. But due to the low viscosity of alcohol more fuel will be

injected in to the combustion chamber with the available fuel injection pump, which normally

operates at 180 bar pressure. This makes the starting of the engine difficult.

In order to compensate this, the fuel injection pressure is to be reduced. So an attempt is

made to find an injection pressure for the suitability of using alcohol in diesel engines. In the

present theoretical investigation, the performance parameters for normal diesel engines are

obtained by using a computer program. Then the performance of the diesel engine is compared

with alcohol at different injection pressures. The injection pressure of alcohol fuel is selected

for the further experimental work in such a way that the injection pressure at which the

performance of alcohol and diesel fuel are in close agreement. So to study the effect of

injector opening pressures, five injector opening pressures (180, 175, 170, 165 and 160 bar)

are considered. From the theoretical results, it is observed that the injector opening pressure

of 165 bar results in higher brake thermal efficiency and is in close agreement with diesel

fuel.

INTERNATIONAL JOURNAL OF MECHANICAL ENGINEERING

AND TECHNOLOGY (IJMET)

ISSN 0976 – 6340 (Print)

ISSN 0976 – 6359 (Online)

Volume 4, Issue 2, March - April (2013), pp. 209-216 © IAEME: www.iaeme.com/ijmet.asp Journal Impact Factor (2013): 5.7731 (Calculated by GISI) www.jifactor.com

IJMET

© I A E M E



210

KeyWords: alcohols, fuel pump, injection pressures and emissions

INTRODUCTION

During recent years INDIA imported 75% of crude oil from other countries to meet

energy requirements. This intensified the research for discovering the new type of engine

design and alternative fuels for better control over pollution, and further leads to the stringent

emission norms. Alcohols are being considered to be supplementary fuels to the petroleum

fuels in India because these are derived from indigenous sources and are renewable. Due to

the high self-ignition temperature and latent heat of alcohols, it requires abnormally high

compression ratios to use them in conventional diesel engines [2].

J P Subrahmanyam [6] developed a computer simulation model for the single cylinder

DI diesel engine with diesel and alcohol as fuel. This model illustrates the simulation of

overall cycle consisting of compression, combustion, expansion and exhaust processes and

predicts various combustion and performance parameters. Further, this model is validated

with available experimental results. Nadir Yilmaz et al [3] identified some of the practical

problems encountered during the usage of alcohol in the diesel engines due to its

characteristics (high latent heat of vaporization, high auto ignition temperature) in which the

reaction transport mechanism is absent. He developed a model which measures the chemical

reactions in the combustion, which further models cylinder pressure and attendant extent of

reaction. Saeed et al [5] conducted experiments with alcohol in single cylinder diesel engine

to find effect of alcohol to diesel fuel on the ignition delay period and concluded that with

increasing the alcohol content ignition delays are prolonged and this can be reduced by air

preheating and/or supercharging.

For the complete combustion in the diesel engine very short time is available. So the

liquid fuel should be injected in droplets of smallest size to obtain largest surface-volume

ratio. But the rate of burning depends primarily upon the rate at which the products of

combustion can be removed from the combustion chamber and replaced by fresh oxygen. So

in diesel engines for the efficient combustion, fuel injection pump plays an important role [7].

An attempt is made for theoretical investigations with different injection pressures for

suitability of using alcohol in diesel engines.

The diesel engine performance is obtained with a computer programming. In such

analysis, if all the variables are taken into account, the computer capability and time required

will be beyond those available for this work. Hence the aim of this theoretical analysis is

restricted only to identify the important variables affecting the performance of the insulated

engine and to know the trends.

The general assumptions that are made in developing this model for the diesel engines

are as follows [4]:

(a) The charge inside the cylinder at any instant consists of a non-reacting mixture

of air and residual gases.

(b) The fuel is assumed to mix homogeneously with air.

(c) The pressure and temperature are spatially uniform.

The performance equations used for the development of computer program for the

diesel engine at various stages is explained below briefly



211

)(P

)(P

θ

θθ d+

COMPRESSION PERIOD

During the compression period the charge consists of air and residual gas.

The general equation for the work can be written as follows

WCOMP = NA CVR (T2-T 1)

Assumed wall temperature is the temperature at the end of the compression stroke.

With this value of P (θ +dθ ) and knowing the value of index and pressure P (θ ) and the

temperature T (θ +dθ ) can be computed.

T (θ +dθ ) = [ ]( r/1r − )

* T (θ )

COMBUSTION PERIOD

The combustion model for the simulation is based on the following assumptions and

simplifications.

i) The ideal gas law is applicable.

ii) The cylinder content consists of a homogeneous mixture of air and combustion

products at all times.

Work done is calculated using the equation.

WCOMB = Σ ((P + ∆P)/2) ∆V

EXPANSION PERIOD

During this process the computations carried out are all similar to the computations

carried out earlier during the compression period.

HEAT TRANSFER MODEL

The heat transfer (hc) in the engine can be calculated with the HOHENBERG RELATION

hc = 0.13 * V-0.06

* [P(W+1.4)]0.8

T-0.4

( KW / m2 K)

HEAT RELEASE MODEL

The rate of heat release (Hr) is calculated using an empirical relation

Hr = WC * C1 (m+1) exp (-WC (θ /θ C) m

IGNITION DELAY MODEL

Ignition delay is understood as being the period of time elapsing between the start of

injection nozzle needle lift and the rise in cylinder pressure which is indicated by a marked

deviation of the cylinder pressure from the compression pressure.

ID (CA) = [0.044 exp (45 / T) / P 1.1

] * RPM



212

FRICTION MODEL

Because of various parameters like gas pressure, wall tension of rings, blow by loss,

pumping losses, throttling loss etc;, the frictional mean effective pressure varies in the

combustion chamber and is given by the GOSCH equation.

FMEP = 0.6 (CR-4) + RPM / 200 +1.1x10 6

V2

COMPUTER PROGRAM

The computer program for this model is written in C language. The computer

Graphics has been developed to plot various output parameters on the monitor theoretically

[1]. The various engine geometry parameters such as bore, stroke, connecting rod length and

combustion chamber geometry are given as inputs to the program. The engine variables such

as compression ratio, intake temperature, intake pressure, injection advance, calorific value

and combustion duration are also given as inputs.

For the practical results experiments are conducted on 4-stroke 3.68 KW Kirloskar

water cooled DI Diesel engine at various loads with an injection pressure of 180 bar. Air

suction rate and exhaust air flow rates were measured with the help of an air box method.

Temperatures at the inlet and exhaust valves are monitored using Nickel-Nickel Chromium

thermocouple thermocouples. Time taken to consume 20 cc of fuel was noted using a digital

stop watch.

Figure1. Experimental set up of Engine Test Rig

Engine RPM is measured using an electro-magnetic pick up in conjunction with a digital

indicator of AQUTAH make. The experimental set up used is as shown the following

Figure.1.



213

RESULTS

Pressure The computed values of the engine parameters for the normal engine are evaluated. The

computed values of pressure during compression and expansion at various crank angles for

0

20

40

60

80

100

120

-120 -60 0 60 120

Pre

ssu

re (

bar)

Crank Angle (degree)

Practi…Theo…

Exhibit 1 Variation of Pressure with Crank angle

theoretical and practical normal engine are shown in Exhibit 1 and are observed that the peak

pressure is higher for the theoretical engine than for the practical normal engine and increases

substantially with increase of crank angle. The cycle peak pressure for a normal practical

engine is 72.76 bar and for theoretical engine it is 79.34 bar.

Temperature

Exhibit 2 shows the cycle peak temperature for theoretical and practical normal

engines. The cycle peak temperature for a normal practical engine is 1156 K and for

theoretical engine it is 1375 K. The rise is about 219K at the peak value. At the end of the

expansion the cycle temperature for the practical engine is 719 K and for theoretical engine it

is 790 K. The rise is about 69 K for the theoretical engine.

0

400

800

1200

1600

-120 -60 0 60 120

Tem

pera

ture

(K

)

Crank Angle (degree)

Practic

al

Exhibit 2 Variation of Temperature with Crank angle



214

This concludes that both the theoretical and experimental values are in close

agreement. So this program can be used to verify at which injection pressure the diesel engine

has better performance characteristics when Alcohol is used as fuel. In the computer program

instead of diesel fuel properties, alcohol properties are introduced with different injection

pressures to know at which injection pressure the alcohol performance will be in close

agreement with diesel fuel. Injection pressures are varied from 180 bar to 160 bar to test the

performance of

the engine. The optimum alcohol pressure obtained is used in diesel engines for the further

experimental investigations.

Brake Thermal Efficiency

The variation of Brake Thermal Efficiency with power output using alcohol as fuel is

shown in exhibit.3 with different injection pressures and the same is compared with diesel

fuel performance. It is evident from the graph that diesel has the highest Brake Thermal

Efficiency. Brake Thermal Efficiency depends on combustion process which is very complex

phenomenon that depends on chamber design, viscosity of the fuel, latent heat of

vaporization and the fuel injection pressure. It is observed that at 165 bar injection pressure,

the brake thermal efficiency is slightly better than the other and is in close agreement with

0

5

10

15

20

25

30

35

0 1 2 3 4

Bra

ke T

herm

al E

ffic

ien

cy

(%

)

Power (KW)

Pure diesel

Alcohol-180

Alcohol-175

Alcohol-170

Alcohol-165

Alcohol-160

Exhibit 3 Comparison of Brake Thermal Efficiency with Power

Output with Different Injection Pressures

diesel fuel. The remaining values of Brake Thermal Efficiency of alcohol are in between 180

bar and 165 bar pressure.

Indicated Thermal Efficiency

It is evident from the graph that diesel has the highest Indicated Thermal efficiency.

From the graph shown in exhibit 4, it is observed that at 165 bar injection pressure, the

indicated thermal efficiency is maximum compared to other injection pressures. This is due to

the amount of alcohol entered in to the combustion chamber is reduced with the reduction of



215

fuel injection pressure. This further reduces the amount of heat absorbed by the alcohol for

the evaporation from the combustion chamber. At 160 bar fuel injection pressure, the amount

of alcohol injected is less and this reduces the power output. So at 165 bar fuel injection

pressure, optimum amount of fuel is injected in such way that the indicated thermal

0

5

10

15

20

25

30

35

40

45

0 1 2 3 4

Ind

ica

ted

Th

erm

al E

ffic

ien

cy (%

)

Power (KW)

Pure diesel

Alcohol-180

Alcohol-175

Alcohol-170

Alcohol-165

Alcohol-160

Exhibit 4 Comparison of Indicated Thermal Efficiency with Power

Output with Different Injection Pressures

efficiency is higher. The remaining values of Indicated Thermal Efficiency of alcohol are in

between 180 bar and 165 bar pressure.

CONCLUSIONS

• The pure Alcohol at 165 bar pressure has higher Brake Thermal Efficiency and

indicated thermal efficiency than all other. This is due to the entering of optimum

amount of alcohol in to the combustion chamber and at the remaining pressures more

amount of alcohol is entered and made the combustion chamber cool.

• It is concluded that Alcohol can be used in diesel engines with 165 bar pressure at

which the performance of alcohol is in close agreement with diesel fuel. At this

pressure the low viscosity of the alcohol is compensated. So the same fuel injection

pump can be used for the experiments by reducing the injection pressure to 165 bar.

REFERENCES

1. Y.Miyairi,” Computer Simulation of an LHR DI diesel engine”, SAE Paper No.880187.

2. Dr.V.Ganesan., “Internal Combustion Engines”

3. Nadir Yilmaz, A. Burl Donaldson “Modeling of Chemical Processes in a Diesel Engine

with Alcohol Fuels”, Journal of Energy Resources Technology, December 2007,

Volume 129, Issue 4, pp 355-359.

4. Dr.V. Ganeshan, “ C.I. Engine Simulation”

5. Saee, M.N, Henein, N.A “Combustion phenomenon of alcohols in C.I. Engines”,

Journal of Engineering for Gas Turbines and Power, Vol/Issue 111:3, 1999.



216

6. J P Subrahmanyam, Rafiqul Islam “ computer simulation studies of an alcohol-fueled,

Low hear Rejection, Direct- Injection diesel engine”, SAE 972976

7. Roy Kamo, Nagesh S “ Injection characteristics that improve performance of ceramic

coated diesel engines”, SAE 1999-01-0972

8. Ram Chandra, T.K. Bhattachary, “Performance Characteristics of a Stationary Constant

Speed Compression Ignition Engine on Alcohol-diesel Micro emulsions”, Agricultural

Engineering International: the CIGR E Journal, Vol VIII, June 2006.

9. Shri. N.V. Hargude and Dr. S.M. Sawant, “Experimental Investigation of Four Stroke

S.I. Engine Using Fuel Energizer for Improved Performance and Reduced Emissions”,

International Journal of Mechanical Engineering & Technology (IJMET), Volume 3,

Issue 1, 2012, pp. 244 – 257, ISSN Print : 0976 – 6340, ISSN Online: 0976 - 6359.

10. A. P. Patil and H.M.Dange, “Experimental Investigations of Performance Evaluation of

Single Cylinder, Four Stroke, Diesel Engine, using Diesel, Blended with Maize Oil”,

International Journal of Mechanical Engineering & Technology (IJMET), Volume 3,

Issue 2, 2012, pp. 653 - 664, ISSN Print : 0976 – 6340, ISSN Online: 0976 - 6359.

11. N.V. Hargude, “An Experimental Investigation for Performance Analysis of Four

Stroke S.I. Engine using Oxyrich Air”, International Journal of Mechanical Engineering

& Technology (IJMET), Volume 3, Issue 2, 2012, pp. 532 - 542, ISSN Print : 0976 -

6340, ISSN Online: 0976 - 6359.

NOMENCLATURE

V = Cylinder volume, m

3

P = Cylinder pressure, atm.

W = Mean piston speed, ms-1

T = Cylinder temperature, K

WC = Wibe’s constant (6.908)

m = constant (0-2)

θ = crank angle under consideration (CAD)

θ C = combustion duration (CAD)

C1 = FKG* HV/COMDUR

CA = Crank angle

T = Temperature in the combustion chamber

P = Pressure in the combustion chamber

RPM = Speed of the engine

FMEP = frictional Mean Effective pressure

CR = Compression ratio

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