SCHOOL OF CHEMICAL ENGINEERING

Design of a shell and tube hydrocarbon preheater

Group : 3

MODULE CODE: ENCH3EC

Name: Waseem Amra Student Number: 210512847

Date: 30/09/2013

Academic Supervisors: Mr. C.A Baah

Dr S.L Kiambi

STATEMENT OF AUTHORSHIP

DESIGN OF A SHELL AND TUBE HYDROCARBON PREHEATER

by

Waseem Amra

I hereby declare that this design and the associated report is my own work (except where formally

acknowledged in the section headed (“Acknowledgements”)

Student Number: ………………

Signature: …………………..

Date: ……………….

Abstract This report serves as a guideline to the design of a shell and tube heat exchanger for the client

ABC chemicals, who wish to embark on the expansion of their current facilities. The objective

entails the preheating of a hydrocarbon binary mixture of hexane and n-heptane, via the heating

fluid dowtherm-A, which is sent to a distillation column. The design follows the standards of the

Tubular Exchanger Manufactures Association (TEMA) and the exchanger is of TEMA type AFT.

Using the specified information given by ABC Chemicals the heat exchanger is designed to

operate under worse conditions, which is taken as winter and corresponds to the hydrocarbon

binary mixture to be heated from 19 to 77.15 °C. The heat duty of the heat exchanger was

calculated to be 1456.87 kW. Estimation of the heat transfer coefficient on the tube and shell side

was done using the Sieder-Tate and Bell Delaware method, respectively. The overall heat transfer

coefficient was determined by an iterative procedure which converged to a value of 220.43

W/m2.K after six iterations. The pressure drops on the tube and shell side were evaluated using

Frank’s adaption and the Bell Delaware method, respectively. It was found that the pressure drop

on the tube side was within specification however the pressure drop on the shell side was low.

Mechanical design of the heat exchanger was performed in order to evaluate the effects of internal

pressure on the vessel hence corresponding minimum thickness and flanges were able to be sized

for the vessel. A technical drawing was also done and is provided in the Appendix. Cost analysis

was performed using two methods: mass-basis of the exchanger and Purohit’s method. Purohit’s

method was selected as it takes into account several factors which the mass-basis does not include.

i

Table of contents

Abstract .......................................................................................................................................................... i

1. Introduction .............................................................................................................................................. 1

2. Theoretical Background ............................................................................................................................ 2

2.1. Background: ....................................................................................................................................... 2

2.2. Physical Properties: ............................................................................................................................ 2

2.3. Thermal Design: ................................................................................................................................. 3

2.3.1. Mass Balance: ............................................................................................................................. 3

2.3.2. Energy Balance: ........................................................................................................................... 3

2.3.3. Velocity, tube and shell specifics: ............................................................................................... 4

2.3.4. Estimation of heat transfer coefficients: .................................................................................... 5

2.3.5. Pressure Drop.............................................................................................................................. 6

2.4. Mechanical Design: ............................................................................................................................ 7

2.5. Costing: .............................................................................................................................................. 8

3. Results ....................................................................................................................................................... 9

4. Discussion ................................................................................................................................................ 11

5. Conclusion ............................................................................................................................................... 15

Appendix A – Nomenclature ....................................................................................................................... A1

Appendix B - References ............................................................................................................................. B1

Appendix C – Process Instrumentation Diagram ........................................................................................ C1

Appendix D – Source Data ..........................................................................................................................D1

Appendix E - Sample Calculations ............................................................................................................... E1

Appendix F – Technical Drawing ................................................................................................................. F1

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1. Introduction A major process in industry is the heating of fluids. Generally this operation is done in a shell

and tube heat exchanger. Optimisation of such a unit is of high importance as this process is a

major contributor to operational costs. Hence, when designing a shell and tube heat exchanger

consideration is taken to minimise fabrication and operational cost. (Geankoplis, 1993)

The client, ABC chemicals, wishes to introduce a new chemical into the market and have

decided to expand on their current facilities. This expansion involves the installation of a new

distillation tower for the separation of a binary mixture of hexane and n-heptane. Further the

client wishes to install a heat exchanger before the line of the distillation tower, which will

preheat the mixture before entry into the distillation tower. This will lead to better separation as

the mixture components will be closer to their boiling points. (Seader, Henley, & Roper, 2010).

The process description is as follows: A hydrocarbon binary mixture of hexane and n-heptane is

sent to a fractionating column which operates isobarically at 101.3 kPa absolute pressure. 38500

kg/hr of the binary mixture is pumped from a storage tank to a shell and tube preheater and then

to the fractionating column. This stream contains 46.4 and 53.6 mass percentage. The

temperature of the fluid in the storage tank depends on the time of year and averages from 19°C

in winter to 32°C in summer. The preheater increases the fluid temperature to 77.15°C. The

binary mixture is heated by dowtherm-A from a hot reservoir. The temperature of the dowtherm-

A in the hot reservoir is maintained at 98.7°C. The dowtherm-A leaves the preheater at 54.45°C

and is then pumped to a cold reservoir. The distillate is condensed and cooled in a condenser.

Table (6) shows the stream data for the heat exchanger and column.

The objective task is to design the preheater before the column. The design entails the use of a

rear end head type setup for pull-through with floating head. This report serves to show the steps

involved in the designing of the preheater.

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2. Theoretical Background

2.1. Background:

The shell and tube heat exchanger is the most common type of heat transfer equipment used in industry. The

design consists of a bundle of tubes enclosed in a cylindrical shell, with the ends of the tubes fitted in tube

sheets that separate the shell and tube side fluids. The advantage of this construction is as follows:

1. A large surface area in a small volume.

2. It is easily cleaned.

3. There is a wide range of materials used for fabrication, hence caters for most processes.

4. Good mechanical layout.

Functioning of this unit is as follows: Process fluids flow through the shell and tube side. Heat transfer

between these fluids is through conduction through the tube wall and convection through the films at the outer

and inner tube walls.

Following TEMA standards a heat exchanger is divided into three parts. These are: the front end, shell, and

rear end. Figure (1) shows the divided sections:

Figure 1

Choice of each section depends on the process conditions, which is outlined in the discussion.

2.2. Physical Properties: The evaluation of physical properties for the binary mixture was estimated using the following mixing rules:

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Property Equation Equation no.

Heat Capacity (J/kg.K) ∑

(2.1)

Density (kg/m3) ∑

(2.2)

Conductivity (W/m.K)

( ) ∑

( )

(2.3)

Viscositiy (Pas.s) ( ) (2.4)

( ) ( ( )

)

(2.5)

Table 1 – Mixing Rules

Use of the mixing rules is valid as the system operates under ideal conditions and the chemical constituents of

the binary components are similar.

2.3. Thermal Design: The procedure adopted in this report is outlined in figure (8).

2.3.1. Mass Balance:

The mass balance of the exchanger, following the conservation of mass, reduces to a simple in=out as there is

no accumulation or generation terms.

2.3.2. Energy Balance:

The primary objective for this design is to determine the required heat transfer surface area for a specified

heat duty using the available temperature differences. (Sinnott, 2005)

(2.6)

(2.7)

Equation (2.6) gives the heat transfer duty of the exchanger and equation (2.7) gives the heat duty of either the

shell/tube side. The overall heat transfer coefficient ( ) is the sum of several individual resistances and is

defined as follows:

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(

)

(2.8)

The driving force for the transfer is the temperature difference between the fluids. For counter-current flow

heat exchangers the driving force is taken as the logarithmic mean temperature difference and is defined as

follows:

( ) ( )

( )

( )

(2.9)

However the nature of counter-current flow assumes several other factors to be constant - overall heat transfer

coefficient, and that the temperature difference of the streams are not excessively large. Hence, a correction

factor is applied to allow for the departure from true counter-current flow and the temperature difference is

found using the following equations: (2.10)

The correction factor is found via parameters R and S, which is used to find from correction charts,

shown in figure (2).

(2.11)

(2.12)

Once the heat duty is known it is possible to compute the surface area off heat transfer, assuming an initial

value for the overall heat transfer coefficient. Using equation (2.6) rearranged:

(2.13)

2.3.3. Velocity, tube and shell specifics:

Velocity of each side has to be determined in order to proceed with the estimation of heat transfer coefficients.

Using the determined heat transfer area the number of tubes and number of tubes per pass can be determined

using the following equations:

(2.14)

(2.15)

(2.16)

Dimensions of the tubes are chosen from the Standard of the Tubular Manufacturers Association (TEMA)

handbook.

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The tube velocity is then determined using the following equation:

(2.17)

Tube side velocity has an optimum range between 1-2 m/s recommended by Sinnott(2005). With the amount

of tubes known, the computation of the tube bundle, pitch and shell diameter can be found using the following

equations: (

)

(2.18)

(2.19)

(2.20)

Where is the bundle diameter clearance found from figure (3). Bundle diameter clearance is taken into

account as spaces must be left in the pattern of tubes on the tube sheet to accommodate the pass partition

plates. The bundle diameter clearance depends on what rear end head type configuration is used. The

parameters and depend on what tube arrangement has been chosen.

The shell side velocity is determined by first assuming a baffle spacing ( ) and baffle cut( ). Baffle spacing

is taken as a ratio between 0.2 – 1 times the diameter of the shell and typically a baffle cut of 25% is used.

Shell side velocity is determined using the following equations:

( )

(2.21)

(2.22)

(2.23)

Optimum range for shell side pressure drop: 0.3 – 1 m/s (Sinnott, 2005)

2.3.4. Estimation of heat transfer coefficients:

Since heat is transferred via convection the heat transfer coefficients for each side have to be determined,

which in turn allows for the overall heat transfer coefficient to be estimated.Correlations are used to determine

the heat transfer coefficients with the help of the following dimensionless numbers:

Prandtl Number:

(2.24)

Reynold’s Number:

(2.25)

Tube side: Estimation of the tube side heat transfer coefficient is found using the Sieder-Tate correlation:

(

)

(2.26)

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Shell side:

Estimation of the shell-side heat transfer coefficient is found using the Bell Delaware Method:

(2.27)

In Bell Delaware method the heat-transfer coefficient is estimated from correlations for flow over ideal tube-

banks, and the effects of leakage, bypassing and flow in the window zone are allowed for by applying

correction factors. (Sinnott, 2005)

The ideal cross flow coefficient ( ) is determined by the following equation:

(

)

(2.28)

The correction factors , , and depend on various structural parameters which is shown in Table (3).

The tube row correction factor ( ) and window correction factor ( ) is found from figure (4) and figure (5),

respectively. The bypass correction factor ( ) and leakage correction ( ) are determined using the following

equations:

*

( (

)

)+ (2.29)

*( )

+ (2.30)

2.3.5. Pressure Drop The tube side and shell side pressure drop were determined using Frank’s Adaption and the Bell Delaware

method, respectively. Recommended ranges for pressure drops of liquids according to Sinnott (2005) are as

follows:

Viscosity (mN s/m2) Pressure (kPa)

< 1 mN s/m2 ≈ 35kPa

1 to 10 mN s/m2 50-70 kPa

Table 2 – Pressure Ranges

Frank’s Adaption is as follows: * (

) (

)

+

(2.31)

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For the shell-side the pressure drops in the cross-flow and window zones are determined separately, and

summed to give the total shell-side pressure drop.

(2.32)

(

)

(2.33)

( )

(2.34)

*( )

+ (2.35)

( ) (2.36)

The bypass ( ) and leakage ( ) correction factors are determined in the same manner as for the heat

transfer coefficients.

( )

(2.37) (

) (

(2.43)

( )

(2.38) ( ) (2.44)

(2.39)

( )

(2.45)

(2.40)

( )

(2.46)

(2.41) ( ) (2.47)

(2.42) √ (2.48)

Table 3 – Parameters for Bell’s Method

2.4. Mechanical Design: For a vessel under internal pressure the minimum thickness of the cylindrical shell, taking corrosion into

account, is given by the following equation:

(2.49)

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The minimum thickness for a torispherical head is determined using the following equation:

( ) (2.50)

The minimum thickness for a flat end is determined using the following equation:

√

(2.51)

Tube plates separate the shell and tube side fluids and are the main support of the shell and tube heat

exchanger. They are exposed to the pressures on both sides and are designed to withstand maximum

differential pressure. The ligament efficiency of tube plates is given by:

(2.52)

Minimum thickness of the plates to resist bending is:

√

(2.53)

Minimum thickness of the plates to resist shear stress is:

(2.54)

The Harker equation is used to evaluate the nozzle diameter and is given by:

(2.55)

Standard flanges are used and are sized according to nozzle diameters. Dimensions for the flanges are

provided in any standard neck flange table.

2.5. Costing: The cost of the heat exchanger was evaluated using two methods. These are Purohit’s method and a mass

basis method. Purohit’s method includes the effects of the shell and tube diameter, tube construction, wall

gage, pitch, layout angle and length. The mass basis method is used by finding the volume of the equipment

and multiplying it by density.

∑ (2.56)

The summation term takes into account the volumes of the shell, tubes, domed heads, flanges and tube

plates. Analysis for the volume of equipment is shown in Table (11) in Appendix E. The price of carbon

steel is taken as /kg.

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3. Results

Heat Exchanger data sheet

Equipment No. (Tag) H-101

Descript. (Func) Hydrocarbon PREHEATER

Sheet No. 1

Operating data 1

2

SIZE 7.32m long TYPE SHELL AND TUBE 3

SHELLS PER UNIT 2 No. OF UNITS 1 4

SURFACE PER UNIT 270 m2 5

Performance of one Unit 6

7

SHELL SIDE TUBE SIDE 8

FLUID CIRCULATING HYDROCARBON MIXTURE DOWTHERM-A 9

TOTAL FLUID ENTERING (kg/h) 38500 68364 10

IN OUT IN OUT

VAPOUR (kg/h) none None 11

LIQUID (kg/h) 38500 68364 12

SPECIFIC GRAVITY LIQUID (pref=

998kg/m3) 0.6503 1.016 13

VISCOSITY LIQUID (Pa.s) 0.00027 0.00132 14

SPECIFIC HEAT (J/kg.K) 2343 1734 15

THERMAL CONDUCTIVITY

(W/m.K) 0.1219 0.1216 16

TEMPERATURE (°C) 19 77.15 98.7 54.45 17

OPERATING PRESSURE (kPa) 101.3 101.3 18

VELOCITY (m/s) 0.48 1.04 19

NO. OF PASSES 2 4 20

PRESSURE DROP (kPa)

ALLO

W 35

CALC

. 37

ALLO

W 70 CALC. 65 21

FOULING COEFFICIENT (W/m2.K) 5000 2000 22

HEAT EXCHANGED (kW) 1457 MTD CORRECTED (°C) 24.6 23

24

Construction of one Shell 25

26

DESIGN PRESSURE (kPa) 111 27

DESIGN TEMPERATURE (°C) 100 28

TUBES No. 614 OD (mm) 19.05 THICKNESS(mm) 3.4036

LENGT

H 7.32 PITCH (mm) 23.81 29

SHELL ID (mm) 827 THICKNESS 5 30

TUBE SHEET

STATIONARY Y

FLOATING

HEAD Y 31

BAFFLES TYPE SEGMENTED

SPACING(mm) &

CUT 413mm, 25% 32

LONG BAFFLE TYPE

LONGITUDINA

L SEAL ALL-METAL T4 SEAL 33

MATERIAL OF

CONSTRUCTION SHELL SIDE

CARBON-

STEEL TUBE SIDE CARBON-STEEL 34

CORROSION ALLOWANCE SHELL SIDE 2 TUBE SIDE 2 35

WEIGHT OF ONE UNIT EMPTY (tonnes) 6.72 36

Heads 37

TORISPHERICAL FLAT-END 38

PRESSURE 111 kPa PRESSURE 111 kPa 39

CROWN RADIUS 827 mm

THICKNES

S 11.64 mm 40

KNUCKLE RADIUS 82.7 mm 41

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THICKNESS 5 mm

42

43

Tube Plate 44

45

DIFFERENTIAL PRESSURE 111 kPa 46

THICKNESS 16.3 mm 47

DIAMETER 827 mm 48

Flanges and Nozzles 49

50

SHELL TUBE 51

PRESSURE 111 111 kPa 52

NOMINAL SIZE 250 350 mm 53

OUTER DIAMETER 139.7 355.6 mm 54

BRANCH THICKNESS 8 12 mm 56

d1 139.7 355.6 mm 57

D 375 490 mm 58

b 22 22 mm 59

h1 60 62 mm 60

d4 365 415 mm 61

f 3 4 62

BOLTING M16 M20 63

d2 22 22 mm 64

k 335 445 mm 65

d3 290 385 mm 66

h2 15 15 mm 67

r 12 12 mm 68

69

Table 4 - TEMA SHEET

COSTING

Exchanger cost (R) 668249.6

Insulation cost ( R) 53459.97

Installation cost ( R) 173744.9

Total cost ( R) 895454.5

Table 5 - Costing

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4. Discussion The heat exchanger specified by the client was a shell and tube type heat exchanger. There are several

other types of heat exchangers that could be chosen, such as the double-pipe and cross-flow heat

exchanger. However, the shell and tube type heat exchanger is the most common type of heat transfer

equipment used in industry (Geankoplis, 1993). The shell and tube type heat exchanger is preferred over

others due to several advantageous factors. The setup of a shell and tube type heat exchanger gives a large

surface area in a small volume. Since the vessel is cylindrical in nature, and often has domed ends, it

enables good pressure operation. Due to its high frequency in industry there is a wide range of fabrication

techniques, which are well established, and there is a wide range of materials which can be used for

construction. Further, and one of the most important factors, a shell and tube type heat exchanger is easily

cleaned since it can be dismantled easily, hence allows it to cater for fluids which are corrosive in nature

(Sinnott, 2005).

The client specified that a rear end head type configuration for pull-through with floating head design be

selected. In this type of configuration the tube sheets are made small enough so that it and its gasketed

bonnet may be pulled completely through the shell. This allows for ease of inspection and cleaning on the

shell-side (Wolverine, 2007). A disadvantage of the pull-through design is that the clearance between the

outermost tubes in the bundle and the shell must be made greater to accommodate the floating head

flange, allowing fluid to bypass the tubes. For the front end stationary head type a channel and removable

cover was selected. This type of configuration allows for ease of cleaning for the tube-side (Hyprotech,

2002).

The heat exchanger is designed to operate under worst conditions, which is in winter where the fluid has

to be heated from 19 to 77.15 °C. This corresponds to the maximum possible heat duty that could occur as

this is the greatest temperature difference achieved.

The Dowtherm-A heating fluid was chosen to be place in the tube-side and the hydrocarbon binary mixture

was placed in the shell-side. The factors involved in deciding the choice of fluid allocation are: Corrosivity,

fouling, temperatures, pressures, viscosity and flowrate. Both fluids are considered non-corrosive, hence

corrosivity was not considered as a determining factor in fluid allocation. The fouling factor of the Dowtherm-

A at 0.0002 m2.

K/W is lower than that of the binary mixture, which is 0.0005 m2.

K/W. Generally, the more

fouled fluid is put in the tube-side, however the pull-through with floating head allows for ease of cleaning in

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the shell-side therefore this factor was catered for. Further the Dowtherm-A temperature and mass flowrate is

higher than that of the binary mixture, which is favoured for in selection of the tube-side. Allocating fluids

with the lowest flowrate to the shell side generally gives the most economical design (Sinnott, 2005). Both

streams operate at the same pressure hence this was not a determining factor. Hence, the factors contributing

for the dowtherm-A to be in the tube-side outweighed the factors of the binary mixture in the tube-side.

The physical properties of the fluids were determined using mixing rules, equations (2.1 – 2.5). The pure

component properties for the binary mixture, which consists of hexane and n-heptane, are acquired from

Perry’s Chemical Engineers’ handbook. All properties are taken at the mean temperature of the inlet and outlet

stream. Mixing rules are acceptable for the binary mixture as the chemical nature of the components’ are

similar, as well as the system behaves ideally.

The procedure adopted for the thermal design is shown in figure (8). The counter-current flow arrangement

was chosen for the heat exchanger. Although there are other types of flow arrangements, such as co-current

and cross flow, the counter-current arrangement proves to be the most effective. The selected tube arrangement

is for a square pattern. This provides further ease of cleaning of the tubes. All dimensions within the design

were selected keeping in accordance with the TEMA standards. The tube dimensions selected were for a 19.05

mm (

) outer diameter of BWG gage 10. Tube length of 7.32 m (24ft.) was chosen. Hence, the heat

exchanger fitted in the allocated space of 10 by 5 m, and was chosen to be in the horizontal position.

The given specifications by the client made it possible to determine the heat duty of the exchanger. The

unspecified mass flowrate for dowtherm-A was then determined and found to be 18.99 kg/s which is higher

than the shell-side flowrate. Evaluation of the temperature correction factor was done using figure (2). The

parameters R and S did not fall within the correction chart for a one-shell and single-tube pass heat exchanger.

Hence, the correction factor was found on a two-shell pass with multiple of four passes, which is figure (2).

Four tube passes was selected for the design of the heat exchanger. The overall heat transfer coefficient was

first assumed as 180 W/m2.K, which allowed for the heat transfer area to be evaluated. The range for the

overall heat transfer coefficient of dowtherm and heavy oils is 50 to 300 W/m2.K (Sinnott, 2005). The design

was then iterated, following the procedure in figure (8), until convergence yielded a zero percentage error

between the assumed and calculated overall heat transfer coefficient. The iteration is shown in Table (5) and

the overall heat transfer converged at the value of 220.43 W/m2.K. The use of equation (2.8) takes into account

fouling coefficients and the material of construction.

The velocity of the tube-side was found to be 1.036 m/s which just fell within the range of 1 – 2 m/s

(Sinott, 2005). An increase in the tube-side velocity in principle is possible if a smaller tube is chosen

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however the amount of tubes with the bundle will increase, which in turn will increase fabrication costs.

Furthermore, the pressure drop will become excessively large leading to the minimum thickness which is

required for the vessel under internal pressure to increase, increasing fabrication costs. The heat transfer

coefficient on the tube-side was determined using the Sieder-Tate correlation. The Sieder-Tate correlation

provides an accurate representation for the convective heat transfer coefficient. dowtherm-A is considered

a non viscous liquid, hence for non-viscous liquids the viscosity ratio of equation (2.26) is taken as unity

(Sinnott, 2005). The velocity on the shell-side was found to be 0.48 m/s which falls in the range of 0.3 to

1.0 m/s (Sinnott, 2005). Baffle spacing of 413 mm was selected using the factor discussed in Section

(2.3.3). This yielded a total of 17 baffles. The baffle spacing is acceptable as it is above the minimum

required baffle spacing of 165.4 mm and the selected spacing falls within the optimum range of 0.3 – 0.5

times the diameter of the shell, which corresponds to 248.1 to 413 mm. The shell-side heat transfer

coefficient was determined using the Bell Delaware method. The Bell Delaware method is estimated from

correlations which take into account flow over ideal tube-banks, the effects of leakage, bypassing and

flow in the window zone and is allowed for by applying correction factors (Sinnott, 2005).

The tube-side pressure drop was evaluated using Frank’s Adaption, which is the most realistic pressure

drop estimation (Sinnott, 2005). The pressure drop on the tube-side was found to be 64.6 kPa. This falls

within the range given in Table (2). The pressure drop on the shell-side was estimated using the Bell

Delaware method. The pressure drop was calculated to be 2.43 kPa which does not fall within the ranges

of Table (2). However, the Bell Delaware method is a highly inaccurate representation of pressure drop

for a divided shell. For such designs, a divided-flow model based on Tinker’s work should be used.

Devore’s method can also be considered, providing the exchanger layout conforms with those covered in

his work (Sinnott, 2005). As a cross-check estimation of the shell-side pressure drop using Kern’s method

was evaluated and this yielded a pressure drop of 37.53 kPa, which lies in Table (2). Hence Kern’s

method is a better approximation for a divided shell. Calculation of the overall heat transfer coefficient

takes into account fouling factors on each side.

The shell diameter was determined to be 827 mm. Hence the L/D ratio fell within the optimum range.

Carbon steel was the selected material for the construction of both the shell and tube side, since both the

fluids are considered non-corrosive. The calculated total number of tubes is 614, with 153 tubes per pass.

The TEMA type for this design is A-F-T. Where ‘A’ is the channel and removable cover, ‘F’ represents

two-shell pass and allows for true counter-current flow and ‘T’ represents pull-through with floating head.

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Recommendation of a rear end head type of a U-tube is advised as there is an ease in bundle removal and

is the simplest design.

The mechanical design followed the standards of PD5500 and ASME8 heat exchanger design. For the

shell under internal pressure of 111 kPa, it required a minimum calculated thickness of 2.27 mm, inclusive

of corrosion allowance. However, according to the standards of PD5500, the minimum thickness of a

shell, less than 1 m in diameter, under internal pressure and which must be able to withstand its own

weight and any incidental loads is 5 mm, inclusive of corrosion allowance. Hence 5 mm was chosen. For

TEMA type AFT the channel and removable cover and pull-through with floating head requires a flat end

and a domed head. A torispherical domed head was selected as the heat exchanger is a low pressure

system. The torispherical head was designed with crown radius equal to the shell diameter and 10% of the

crown radius was used as the knuckle diameter to ensure a ratio of knuckle of crown to radii greater than

0.06 to prevent buckling (Sinnott, 2005). The minimum thickness required for the torispherical head is

2.417 mm, inclusive of corrosion. For ease of fabrication purposes, the thickness of the head is taken as

the same thickness of the shell wall. The minimum thickness of the flat end was evaluated to be 13.64

mm, inclusive of corrosion allowance. The tube thicknesses under internal pressure was found to be less

than the value used in the thermal design, hence the thermal design thickness was used. One could reduce

costs by reducing the thickness of the tubes, however this would lower the tube velocity. The nozzle

diameter was evaluated using the inlet-velocity relationship to area and Harker’s equation. The nozzle

diameters using Harker’s equation was small for the corresponding inlet flowrate hence the inlet-velocity

relationship to area was used to determine the diameter of the nozzles. Because there was no phase change

and change in velocity, an assumption was made that the outlet nozzle sizes are the same as the inlet.

Appropriate flanges were sized for the nozzles. The tube plate minimum thickness was evaluated by

considering two factors i.e. bending and shear stress. However, the value calculated for both of these

factors was less than the standard sized thickness of 16.3 mm. Hence the standard size of 16.3 mm,

inclusive of corrosion allowance, was used.

Costing was determined using two methods. These were the mass-basis method, in which the volume of

each component was determined and multiplied by the density, and Purohit’s method. The estimated costs

take into account insulation, installation and labour of the vessel (Purohit, 1983). The mass-basis method

yielded a cost price of R 1121186 and Purohit’s method yielded a cost of R 895454.5. Purohit’s method

was selected as it takes into account several factors which the mass-basis method does not account for.

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5. Conclusion A shell and tube heat exchanger of TEMA type AFT was selected where ‘A’ represents a

channel and removable cover, ‘F’ two-shell pass with longitudinal baffle and ‘T’

represents pull-through with floating head design.

With the choice of winter conditions selected as worst operating conditions the heat duty

was determined to be 1456.87 kW.

Dowtherm-A was selected to be placed in the tubes.

Evaluation of the thermal design estimated the use of 614 tubes with 4 tube passes.

The overall heat transfer coefficient was found to be 220.43 W/m2.K

Mechanical design of the exchanger yielded appropriate minimum thicknesses and sizing

of nozzles and flanges.

Purohit’s method was selected for cost analysis and the total cost of the heat exchanger is

R 895454.5.

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Appendix A – Nomenclature

A Heat transfer area (m2)

Ab Clearance area between the bundle and the shell(m2)

AL Total leakage area(m2)

Apass Area of a pass (m2)

As Maximum area for cross flow in the shell (m2)

Asb Shell-to-baffle clearance area (m2)

Atb Tube to baffle clearance area (m2)

AW Area of the window-zone (m2)

Bb Bundle Cut (%)

BC Baffle Cut (%)

Cph Design Factor

cS Baffle-to-shell clearance

ct Diametrical tube=to=baffle=clearance (mm)

Cs Stress concentration factor

di Tube internal diameter (m)

do Tube outer diameter (m)

DS Shell diameter

Db Bundle Diameter (m)

e Minimum thickness (mm)

f Design stress

Ft Temperature correction factor

h Heat transfer coefficient (W/m2.K)

Hb Height from baffle cord to top of tube bundle (m)

Hc Baffle cut height (m)

jf Heat transfer factor

jt Friction factor

k Thermal conductivity (W/m.K)

lb Baffle spacing (m, mm)

L Tube length (m)

m Mass flow rate (kg/s)

NCV Number of constrictions in the cross flow section

Ns Number of sealing strips

Nt Number of tubes

Ntp Number of tubes per pass

NW Number of tubes in the window zone

Nwv Number of restrictions for cross flow in the window

zone

Nu Nusselt’s Number (Dimensionless)

θb Angle subtended by baffle chord (rads)

pt Pitch (mm)

Pr Prandtl number (Dimensionless)

Pi Operating pressure (kPa)

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∆P’ Maximum pressure difference (kPa)

Q Heat Duty (kW)

Ra Bundle cross sectional area to the total bundle cross

sectional area

Ra’ Bundle cross sectional area to the total bundle cross

sectional area

RC Crown radius (m)

RK Knuckle radius (m)

RW Ratio of the number of tubes in the window zone to

the total number of tubes

SA Surface area (m2)

t1 Cold inlet temperature (⁰C)

t2 Cold outlet temperature (⁰C)

T1 Hot inlet temperature (⁰C)

T2 Hot outlet temperature (⁰C)

∆TLM Logarithmic mean temperature difference (⁰C)

tp Minimum thickness to resist bending/shear

stress(mm)

us Velocity in the shell-side (m/s)

ut Velocity in the tube-side (m/s)

uw Velocity in the window zone (m/s)

uz Geometric mean velocity (m/s)

Uo Overall heat transfer coefficient (W/m2K)

V Volume (m3)

Vtube Volumetric flowrate in tubes (m3/s)

VBI Viscosity blending index

Ligament efficiency

μ Viscosity (Pa.s)

μwall Viscoisty at the wall (Pa.s)

Density (kg/ m3)

Table 6

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Appendix B - References

Anonymous. (2006). Overview of Pressure Vessel Design to AS 1210. Chicago: Proplin.

Barbee, J., Davis , M., Davis, S., & Gaddis, D. (2007). Standards of the Tubular Exchanger

Manufactures Association. New York: Tubular Exchanger Manufactures Association.

Geankoplis. (1993). Transport Processes and Unit Operations. New Jersey: Prentice-Hall

International.

Harding, J. (2009). Economics of New Reactors and Alternatives. Washington DC: Harding Consulting

.

Hyprotech. (2002). Shell-and-Tube Heat Exchangers. London: HTFS.

Lestina, T., & Serth, R. W. (2010). Process Heat Transfer: Principles, Applications and Rules of Thumb.

London: Academic Press.

Perry, R. H., & Green, D. W. (2008). Perry's Chemical Engineers' Handbook. New York: McGraw-Hill.

Purohit, G. P. (1983). Estimating costs of shell-and-tube heat exchangers. New York: Fluor Corp.

Seader, J. D., Henley, E. J., & Roper, D. K. (2010). Speration Process Principles. Utah: John Wiley &

Sons, Inc.

Sinnott, R. K. (2005). Chemical Engineering Design. Oxford: Elsevier Butterworth-Heinemann.

Wolverine . (2007). Construction of Shell and Tube Heat Exchangers. Wolverine Tube Heat Transfer

Data Book, 36.

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Appendix C – Process Instrumentation

Diagram

Included in the Group Hazop

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Appendix D – Source Data

Figure 2 – Temperature correction chart for a two shell pass with

multiple of 4 tube passes (Sinnott, 2005)

Figure 3 – Bundle diameter clearance (Sinnott, 2005)

Figure 4 – Row correction factor chart (Sinnott, 2005)

Figure 5 – Window correction factor (Sinnott, 2005)

Figure 6 – Baffle geometrical factors (Sinnott, 2005)

Figure 7 – Coefficient for Fl’ pressure drop (Sinnott, 2005)

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Figure 8 – Thermal Design Algorithm (Sinnott, 2005)

Figure 9 – Tube-gage cost multiplier (Purohit, 1983)

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Appendix E - Sample Calculations Evaluation of physical properties:

Component Mass Fraction

(wi)

Mass (kg/hr) Molecular

weight(g/kmol)

Moles (kmol/hr) Mole

Fraction (xi)

Hexane 0.464 17864 0.08618 207.29 0.502

n-Heptane 0.536 20636 0.1002 205.95 0.498

Table 6 – Cold stream data

DOWTHERM-A

Heat Capacity (J/kg.K) 1733.73

Density (kg/m3) 1014.29

Conductivity (mW/m.K) 129.6

Viscosity (mPas.s) 1.32

Table 7 – Properties of dowtherm-A

Properties acquired from () and (Perry’s)

Hexane n-Heptane

Heat Capacity (J/kg.K) 2354.16 2332.71

Density (kg/m3) 633.08 659.71

Conductivity (mW/m.K) 118.5 125.4

Viscosity (Pas.s) 2.37x10-4

3.06x10-4

Table 4 – Pure properties of hexane and n-heptane

∑

( ) ( )

∑

( ) ( )

( ) ∑

( )

( )

( )

For hexane:

Similarly for n-heptane:

( ( )) ( ( ))

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Similarly for n-heptane:

∑ ( ) ( )

( ( ( )

)) ( (

( )

))

( )

MIXTURE PROPERTIES

Heat Capacity (J/kg.K) 2342.66

Density (kg/m3) 649.02

Conductivity (W/m.K) 0.1219

Viscositiy (Pas.s) 2.71x10-4

Table 5 – Properties of the binary mixture

Thermal Design:

The following design sequence follows the algorithm from shown in figure (8) with the exception

of step 2, as it was done in the evaluation of properties.

Mass Balance:

Energy Balance:

According to step 1 of figure (8) the heat duty and unspecified mass flowrate are found using

equations (2.6) & (2.7):

( )( )

With the heat duty found the unspecified mass flowrate of the shell side can be found using

equation (2.7) rearranged:

( )

Following steps 3 & 4 of figure (8) the required area of heat transfer is required. First equations (2.9

– 2.12) are to evaluated:

( ) ( )

( )( )

( ) ( )

( )( )

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The parameters and need to be found to estimate :

cannot be read off the plot for a one shell pass with multiple of 2 tube passes hence was

read off the plot for two shell passes with multiple of four tube passes, shown in figure (2).

.Therefore:

( )

The required heat transfer area can now be determined using equation (2.6) rearranged. An

initial value for has to be assumed. The range for Dowtherm and heavy oils is 50-300

W/m2K. An iterative procedure is used to determine the required heat transfer area until the

percentage difference between the ( ) and ( ) is in the range 0-30%.

(Sinnott, 2005)

An initial value of the overall heat transfer coefficient ( ) was assumed to be 180

W/m2K and yielded a calculated value of ( ) of 203.302 W/m

2K. This produced a

percentage error of 12.95%. To reduce this percentage error the calculated overall heat transfer

coefficient of the first iteration was then used as the initial overall heat transfer coefficient in the

next iteration until the percentage error was close to 0%. Table (5) shows the iterative

procedure.

Iteration No. (W/m

2K)

(W/m2K) Percentage Error(%)

1 180 203.30 12.94

2 203.30 213.54 5.04

3 213.54 217.72 1.96

4 217.72 219.37 0.76

5 219.37 220.02 0.294

6 220.02 220.43 0 Table 6 - Iterations

Iteration no.6 is the last iteration as the value converged at W/m2K.The following

calculations are based on iteration no.6. The required heat transfer area can now be determined

using equation (2.6) rearranged:

( )

Velocity, tube and shell specifics:

Step 6, 7 and 8: To determine the velocity of the tube side, standard tube dimensions have to be

chosen. Table (6) shows the tube dimensions chosen for this design.

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Outer diameter, do

(m)

B.W.G. Gage Wall Thickness (m) Inner diameter, di (m)

do/di Tube Length

(m)

0.019050 10 0.003404 0.012243 1.556017 7.32

Table 7 – Tube Dimensions

The area of a single tube is found using equation (2.15):

( )( )

The total number of tubes is found using equation (2.14):

The amount of tubes per pass is found next. Number of tube passes ( ) for this design is 4. The

number of tubes per pass was found using equation (2.16):

The area of a pass ( ) can now be evaluated:

( )

The volumetric flowrate in a single tube is found via the following equation:

Using equation (2.17) the tube side velocity is evaluated:

The tube-side velocity is acceptable as it falls within the optimum range given in Section (2.3.3).

The bundle diameter was then determined. A tube arrangement for a 90 degree was selected

hence the parameters and are 0.158 and 2.263, respectively. Using equation (2.18):

(

)

(

)

The type of rear end configuration for this design is a Pull-through with floating head (TEMA

type-T) and corresponds to a bundle diameter clearance (BDC) of 0.092 m, as shown in figure

(Sinnott, 2005). The diameter of the shell is then evaluated using equation (2.19):

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The pitch is found using equation (2.20): ( )

The baffle spacing ( ) and baffle cut ( ) were selected as follows:

; ( )

The velocity of the shell side is calculated as follows:

( )

( )( )( )

The shell-side velocity is acceptable as it falls within the optimum range given in Section

(2.3.3.).

Estimation of heat transfer coefficients:

Steps 9, 10 and 11 are used to determine the tube and shell heat transfer coefficients.

Tube side:

Evaluation of the tube-side heat transfer coefficient is calculated using the Sieder-Tate method

as follows:

( )( )( )

( )( )

(

)

( ) ( ) ( )

The viscosity term is taken as unity as the fluid is non-viscous as well as the temperature

difference between the wall and the fluids are not excessively large. The tube-side heat transfer

coefficient is determined by equation (2.26) rearranged:

( )

Shell-side:As discussed in Section (2.3.4.) the Bell Delaware method is used to determine the

shell-side heat transfer coefficient.

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The ideal cross flow coefficient ( ) is determined by the following equation:

(

)

( )( )

( )

( )( )

( )

The heat transfer factor is read off figure (12.31) of (Sinnott, 2005). Hence, the ideal cross flow

coefficient is:

(

)

( )( )( ) ( ) ( )

( )

To evaluate the correction factors , and the structural parameters have to be determined. Using

equations (2.37 – 2.48) the parameters are determine

( )

( )

( )

( ( ))

( )

(

)

(

( )

) ( )

( )

( ) ( )

( )

( )

( )

( )

( )

( )

( ) ( )

√ √ ( )

Table 8 – Parameters for Bell’s method

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The parameters , and are read off figure (6). The velocity in the window-zone is determined

using the following equation:

( )

The baffle-to-shell-clearance( ) is evaluated using the following equation:

mm

For a Re > 2000, figure (4) is used to determine the tube row correction factor . The number of

constrictions encountered in the cross-flow section ( ) was found to be 17. For this the tube row

correction factor was determined to be:

The window correction factor ( ) is found from figure (5). For a corresponding of 0.3 the window

correction factor was determined to be: The bypass correction factor ( ) is determined using

equation (2.29): *

( (

)

)+ *

( (

( )

)

)+

One sealing strip was used for 5 vertical rows

The leakage correction factor ( ) is determined using equation (2.30):

*( )

+ *

( ( ))

+

Where is obtained from figure (7).

The shell-side heat transfer coefficient can now be determined using equation (2.27):

( )( )( )( )( )

Finally, the overall heat transfer coefficient can be determined using equation (2.8):

(

)

( )

( )

Pressure Drop:

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Step 12 of figure (8).

Tube Side: The tube-side pressure drop was determined using Frank’s Adaption, described by

the following equation:

[ (

) (

)

]

[ ( ) (

) ( ) ]

( )( )

The friction factor ( ) is obtained from figure (12.36) of (Sinnott, 2005). This pressure drop is

applicable as it falls in the range given in Table (2).

Shell-side: The shell-side pressure drop was determined using Bell Delaware method. The

pressure drops in the cross-flow and window zones are determined separately, and summed to

give the total shell-side pressure drop.

The cross-flow pressure drop is determined using equations (2.32-2.36):

(

)

( )( )( )( )

( )

The bypass correction factor for pressure drop ( ) is determined using equation (2.29) again,

with the only difference being:

The leakage correction factor for pressure drop ( ) is determined using equation (2.30) again,

with the only difference being is obtained from figure (7) and corresponds to 0.65.

The bypass and leakage correction factors for pressure drop are given below.

Hence the cross-flow pressure drop is:

( )( )( )

The pressure drop in the window zone and end zone is found using equation (2.34) and equation

(2.35), respectively.

( )

( )( ( ))

( )( )

*( )

+ ( ) *

( )

+ ( )

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The number of baffles (Nb) needs to be determined, followed by the total-shell side pressure

drop using equation (2.36).

[(

) ] [(

) ]

( ) ( ) ( )( ) ( )

This pressure drop is not within the range given in Table (2).

Mechanical Design:

The operating conditions of the heat exchanger is given in Table (8) as well as the material of

construction parameters of the shell.

Operating Pressure, P (N/mm2) 0.111

Shell inner diameter, Di (m) 0.827

Design Temperature (°C) 98.7

Type Carbon-

Steel

Tensile strength (N/mm2) 460

Design stress at design temperature (N/mm2), f 170.26

Table 9 – Mechanical design parameters

The minimum thickness of the cylindrical shell is found using equation (2.49):

( )( )

( ) ( )

Taking corrosion considerations into account 2 mm is added to the minimum thickness

However for a vessel approximately 1m in shell diameter, according to TEMA standards, the

minimum thickness, to ensure that any vessel is stable and can support its own weight or any incidental loads, is

5mm. Hence, the thickness of the cylindrical shell is taken as 5mm.

Domed ends: Torispherical head

Torispherical head

Crown Radius, Rc (equal to shell diameter) 0.827 m

Knuckle Radius, Rk (10% of crown radius) 0.0827 m

Length of shell, Lshell 8.174 m

Joint factor (no joints in head) 1

Stress concentration factor, Cs 1.54

Table 10 – Domed head factors

The stress concentration factor Cs is determined using the following equation:

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( √

)

The minimum thickness for a torispherical head is found using equation (2.50):

( )

( )( )( )

( )( ) ( )( )

Adding 2mm for corrosion allowance: Take as the same

thickness as the wall

Flat end: The minimum thickness for a flat end is determined using the following equation:

; √

( )√

Adding 2mm for corrosion allowance:

Tube plates:

Max differential pressure in heat exchanger: Design

factor Cph for a welded plate = 0.45; Dp=Di Maximum

allowable shear stress = 0.5(170.26) = 85.13 N/mm2

Minimum thickness of the plates to resist bending is : √

Minimum thickness of the plates to resist shear stress is:

Generally the greater value between equations (2.53) & (2.54) is selected, however, both the

values are not greater than the standard minimum plate thickness, Hence, the standard minimum

plate thickness is used. (2mm was added for corrosion allowance)

Nozzle Diameters: Nozzle diameters were sized according to the inlet-velocity relationship to area and

also by using the Harker’s equation, equation (2.55).

Method Shell nozzle diameter(mm) Tube nozzle diameter(mm)

Inlet-velocity 208.66 303.35

Harker 130.75 67.40

Table 11 – Nozzle diameters

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Flanges: Flanges were sized according to the nozzle diameters. Dimensions for the flanges are

provided in any standard neck flange table. Shown in the results

Costing:

To determine the cost, the volume of each component has to be found:

Volume of the tube bundle (m3)

(

) 0.7518

Volume of the cylindrical shell (m3)

(

) 0.0477

Volume of 2 tube plates (m3)

(

) 0.0118

Volume of flat end (m3)

0.00633

Volume of 4 flanges ( 2 for the shell + 2 for the tube) (m3) 0.00706

Torispherical head

Surface area of crown (m2) 0.8327

Surface area of knuckle (m2)

(

)

0.0416

Volume of a torispherical head (m3) ( ) 0.00437

Table 12 – Volumes of equipment

Where the height of the torispherical head is determined using the following equation:

√(

)(

)

The density of carbon steel:

∑ ( )( )

Total cost has to include insulation, installation and labour costs and is shown below:

( )

( )

( )

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Purohit’s Method:

Tema Type parameters

Channel and Removable cover (A) f = 1.03

Two-shell pass with longitudinal

baffle (F)

Baseline = 1.15

Pull-through with floating head (T)

r = 1.05

(Based on a square tube

arrangement)

1

Table 13 – TEMA type parameters

Purohit’s method involves the use of cost factors based on the TEMA type, shown in Table(12),

and is used to determine a base cost of the unit. Purohit’s method adopts the use of American

units. The following equations are used to determine the base price:

( ) ( ) ( ) ( )

(

)

(

)( )( )( )

Since a TEMA shell type E a correction of the shell type was found using Table (5) of (Purohit,

1983). Correction for the shell type ( ) was found to be:

With the tube length being < 20 ft. and no expansion joint used the correction factor for tube length and

expansion joint were not used. More than 2 tube passes were used hence a correction factor tube pass

( ) was used.

Since pressures were less than 150 psig and carbon steel was fully used correction factors for design

pressure and material of construction were not used. Material of construction was neglected since Purohit’s

equations are based on carbon-steel. Since BWG gage of 10 was used a correction factor for tube gage

determined using the following equation : ( ) (

( ) )

( ) ( )

The total correction factors were totalled:

The total cost of the exchanger was found using the following equations:

( ( ) ) ( ( ) )( )

Where A is the area of heat transfer and N is the number of shells. Since Purohit’s method is based on

costing in 1982 today’s price was determined using cost indices.

( )

( )

( )

Purohit’s method is inclusive of labour

costs. The cost index for 1982 was taken from (Sinnott, 2005) and for 2013 was taken from (Harding, 2009).

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Appendix F – Technical Drawing