Chapter 5 - Heat Exchanger

Production of n-propanol Chapter V

5-1

CHAPTER V

HEAT EXCHANGER DESIGN

5.0 INTRODUCTION

The process of heat exchanger between two fluids that are at different temperature

and separated by a solid wall occurs in many engineering applications. The device

used to implement this exchange is termed a heat exchanger, and a specific

applications may be found in space heating and air-conditioning, power production,

waste heat recovery and chemical processing.

A heat exchanger is a device used to passively transfer heat from one

material to another. These materials may be liquid or gaseous, depending on the

situation in which the heat exchanger is being used. There are many models and

types of heat exchangers, but they essentially work based on the laws of

thermodynamics. One of those laws states that when an object is heated, the heat

energy contained within that object will diffuse outward to the surrounding

environment, until the heat energy in the object and in the environment have reach

equilibrium.

(Source: Fundamentals of Heat and Mass Transfer , 6th Edition)

5.1 TYPE OF HEAT EXCHANGER

Heat exchangers may be classified according to the following main criteria:

1. Recuperators and regenerators

2. Transfer process: direct contact and indirect contact

3. Geometry of construction: tubes, plates and extended surface

4. Heat transfer mechanisms: single phase and two phase

5. Flow arrangements: parallel, counter and cross flow

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5-2

The simplest heat exchanger is one for which the hot and cold fluids move in

the same or opposite directions in a concentric tube heat exchanger (or double pipe)

construction. In the parallel-flow arrangement, the hot and cold fluids enter at the

same end, flow in the same direction, and leave at the same end. In the counter flow

arrangement, the fluids enter at opposite ends, flow in opposite directions and leave

at opposite ends. Alternatively, the fluids may move in cross flow (perpendicular

each other), by the tubular heat exchanger. The principal types of heat exchanger

used in the chemical process and allied industries are as below:

1. Double-pipe exchanger

2. Shell and tube exchangers

3. Plate and frame exchangers

4. Double-pipe exchanger

5. Shell and tube exchangers

6. Plate and frame exchangers

7. Plate-fin exchangers

8. Spiral Heat exchangers

9. Air cooled: cooler and condensers

10. Direct contact: cooling and quenching

The common configuration use is the shell-and-tube heat exchanger.

Specific forms differ according to the number of shell-and-tube passes, and the

simplest form. Baffles are usually installed to increase the convection coefficient of

the shell-side fluid by inducing turbulence and cross-flow velocity component. The

shell and tube exchanger is by far the most commonly used type of heat-transfer

equipment used in the chemical and allied industries.

(Source: Chemical Engineering Design 5thEdition)

5.2 TYPE OF SHELL AND TUBE EXCHANGER

The principal types of shell and tube exchanger are:

1. Fixed tube plate

2. U-Tube

3. Internal floating head without clamp ring (pull through design)

4. Internal floating head with clamp ring (split flange design)

5. External floating head, packed gland

6. Kettle reboiler with U-tube bundle


5-3

The characteristics of shell and tube exchanger types are listed in Table 5.1.

For this design, the shell and tube type of internal floating head (split-ring floating

head) has been chosen according to the advantages compared to the others.

Table 5.1: Comparison of types of shell and tube exchanger

Advantages Disadvantages

Fixed Tube plate

Simplest

Cheapest

Tube bundle cannot be

removed for cleaning.

No provision for differential

expansion of shell and tubes.

Limited to temperature

differences up to 80°C.

Limited to low shell pressure

up to 8 bar.

U-Tube (U-Bundle)

Requires only one tube

Limited in use to relatively

clean fluids as the tubes and

bundle are difficult to clean.

Too difficult to replace a tube.

Internal Floating Head

(Split-ring floating head)

Suitable for high

temperature differentials

The tubes can be rodded

end to end and the bundle

easily to remove and

repairs.

Easier to clean and can be

used for fouling liquids.

Separate the shell and

tube side fluid at the

floating head end.

Increase efficiency.

Internal Floating Head

(pull through design)

Same as Internal Floating Head (split-ring floating head)

Clearance between the

outermost tubes in the bundle

and the shell must be made

greater than the fixed and U-


5-4

tube designs.

External Floating Head

Floating head joint is located

outside the shell, and the

shell sealed with a sliding

gland joint employing stuffing

box and makes a danger of

leaks through the gland.

Limited to about 20 bar.

The shell side is not suitable

for flammable or toxic

materials.

Kettle Reboiler

Same as U-tube Same as U-tube

(Source: Coulson & Richardson’s Chemical Engineering Volume 6, 1999)

5.3 CHEMICAL ENGINEERING DESIGN

Figure 5.1: Heat exchanger E101

Step to Design Heat Exchanger

1. Define the duty: heat-transfer rate, fluid flow-rates, temperatures.

2. Collect together the fluid physical properties required: density, viscosity,

thermal conductivity.

3. Decide on the type of exchanger to be used.

4. Select a trial value for the overall coefficient, U.

5. Calculate the mean temperature difference.

6. Calculate the area required from equation.

7. Decide the exchanger layout.

8. Calculate the individual coefficients.

Stream 15 Stream 17

Heat Exchanger (E101)


5-5

E-1

9. Calculate the overall coefficient and compare with the trial value. If the

calculated value differs significantly from the estimated value, substitute the

calculated for the estimated value and return to step 6.

10. Calculate the exchanger pressure drop; if unsatisfactory return to steps 7 or

4 or 3, in that order of preference.

Heat Load Of Heat Exchanger

Figure 5.2: Heat Exchanger ( E-101 )

Duty for this heat exchanger (E-101) is obtained from the equation 1.1

)( 21 TTmCQ p (5.1)

)70150(246.33600

/20890

s

hkg

KW87.1506

The duty of tube-side is equal to the duty of shell-side. From this value of duty

calculated, the flow rate of the cooling water can be determined. This is done by

using equation 5.2.

Cooling water flow rate, ṁw ))(( 12 ttCp

Q

water (5.2)

))3060(2.4(

87.1506

KW

skg /0.12

The inlet and outlet temperature of the tube have been assumed. The inlet

temperature of water is 30oC and the outlet temperature of water is 60oC due to

commonly used in industry. Table 5.3 shows the physical properties of the

component in the tube and shell side. All the physical properties are taken based on

the mean temperature.

Stream 17

( 70oC)

Stream 15

( 150oC)


5-6

Table 5.3: Physical Properties in Shell and Tube Side

Physical Properties Shell Side Tube Side

Temperature, T (°C)

(a) Inlet 150 30

(b) Outlet 70 60

(c) Mean 110 45

Specific Heat, Cp ( kJ/kg°C) 3.246 4.200

Thermal Conductivity, k ( W/m°C) 0.1150 0.6376

Density, (kg/m3) 13.704 995.818

Viscosity, (mNs/m2) 0.0001621 0.5986

Duty, Q (kW) 1506.87 1506.87

Flow rate (kg/s) 5.80 12.0

(Source:Fundamentals of Heat And Mass Transfer Sixth Edition )

To design or to predict the performance of heat exchanger, it is essential to relate

the total heat transfer rate to quantities such as the inlet and outlet fluid temperature,

the overall heat transfer coefficient and the total surface area for heat transfer.

The logarithmic mean temperature different (LMTD) method has been

choose. It is because the fluid inlet temperature is known and the outlet temperature

is specified.

(source: Fundamentals of Heat and Mass Transfer , 6th Edition)

Counter-flow arrangement is selected as the temperature difference is

greater compared to cross flow. For the LMTD involved, the following assumptions

are made:

1. The overall coefficient of heat transfer is constant

2. The rate of flow of each fluid is constant

3. There is no phase change during cooling process

4. There is an equal amount of cooling surface in each pass



5-7

5.3.1 Determination of Heat Transfer Area

To find the heat transfer area of heat exchanger, the true temperature different mT

must known first. By relating the total heat transfer rate q to the temperature

different between the hot and cold fluid, expression 5.3 can be produce.

ch TTT (5.3)

Since T varies with the position in the heat exchanger, it is necessary to

work with equation 1.4, where mT is an true temperature difference. From

equation 5.4, heat transfer area of heat exchanger can be obtained.

mTUAq (5.4)


5.3.1.2 Determination of Log Mean Temperature Difference

The form of mT may be determined by applying an energy balance to the

differential elements of length dx and surface area dA in the hot and cold fluids. The

energy balances and the subsequent analysis are subject to the following

assumption:

1. The heat exchanger is insulated from its surrounding, in which case only

heat exchange between the hot and cold fluid.

2. Axial conduction along the tubes is negligible.

3. Potential and kinetic energy changes are negligible.

4. The fluid specific heat is constant.

5. The overall heat transfer coefficient is constant.

The specific heat may change as a result of temperature variations, and the

overall heat transfer coefficient may change because of variations in fluid properties

and flow conditions. However, in many applications such variations are not

significant, and it is reasonable to work with average ofcCp ,

hCp and U for heat

exchanger. Applying all this assumption, equation 5.5 can be obtained.

)(

)(ln

)()(

12

21

1221

tT

tT

tTtTTlm

(5.5)


5-8

where;

1T = inlet shell-side fluid temperature, °C

2T = outlet shell-side fluid temperature, °C

1t = inlet tube-side temperature, °C

2t = outlet tube-side temperature, °C


)3070(

)60150(ln

)3070()60150(

lmT

CT o

lm 66.61

5.3.1.3 True Temperature Difference of Heat Exchanger

The usual practice in design of heat exchanger is to estimate the true temperature

different from the log mean temperature difference by applying the correction factor,

tF to allow for the departure from true counter current flow. True temperature

difference is shows in equation 5.6. The correction factor is the function of the shell

and tube fluid temperature, and the number of tube and shell passes. It is normally

correlated as a function of two dimensionless temperature ratio which are equation

5.7 and equation 5.8.

lmtm TFT (5.6)

)(

)(

12

21

tt

TTR

(5.7)

)3060(

)70150(

R

67.2

)(

)(

11

12

tT

ttS

(5.8)

)60150(

)3060(

S

33.0


5-9

Based on the value of R and S calculated, the correction factor tF can be

found. In order to find the correction factor, two shell passes and four tube passes

on the heat exchanger have been choose. This is due to, the value of tF cannot be

obtained when one shell pass and two tube passes is used. In addition, an

economic exchanger design cannot normally be achieved if the correction factor tF

falls below about 0.75. in these circumstances, an alternative type of exchanger

should be considered that gives a closer approach to true counter current flow. The

use of two side shell pass and four tube passes will give a closer approach to true

counter current flow.

Hence, based on Figure E.1(Appendix E), with the value of R is 2.67 and

the value of S is 0.33, the correction factor is 0.88. By using equation 5.6, true

temperature difference was calculated.

lmtm TFT

66.6188.0

Co26.54

5.3.1.4 Overall Coefficient

The most essential part of any heat exchanger analysis is determination of the

overall heat transfer coefficient,U . This overall heat transfer coefficient is defined in

terms of the total thermal resistance to heat transfer between two fluids. For this

case, gases is taking as a hot fluid and water as a cold fluid. Based on table E2

(Appendix E),the overall coefficient for this heat exchanger is 160 W/m2 oC

By inserting the value of heat load q , true temperature difference mT , and

overall heat transfer coefficient, U into equation 5.4, the heat transfer area can be

calculated.

mTUAq

mTU

qA

)26.54()/160(

1506870

2 CCmW

W

oo

257.173 m


5-10

5.3.2 Tube Exchanger

The most basic and the most common type of heat exchanger construction is the

tube and shell. This type of heat exchanger consists of a set of tubes in a container

called a shell. The fluid flowing inside the tubes is called tube side fluid and the fluid

flowing on the outside of the tubes is the shell side fluid. There are several factor

that must be take into account before allocate the suitable fluid in the shell and tube

side. In this case there is no phase change in both fluid. All the factor is shown in

table 5.4.

(source:www.engineersedge.com/heat_exchanger/tube_shell)

Table 5.4 : General consideration of fluid allocation in shell and tube

Factor Description

Corrosion The more corrosive fluid should be allocated to the tube

side. This will reduce the cost of expensive alloy or clad

components. It is advantageous in cooling to connect the

stream to the tubes of the cooler rather than the shell. In

this way, since the stream may be corrosive, the action can

be confined to the tube side alone, whereas if the steam is

introduced into the shell, both may be damaged.

Fouling The fluid that has the greatest tendency to foul the heat-

transfer surface should be placed in the tubes. This will

give better control over the design fluid velocity, and the

higher allowable velocity in the tubes will reduce fouling.

Also the tubes will be easier to clean.

Fluid temperature If the temperatures are high enough to require the use of

special alloys placing the higher temperature fluid in the

tubes will reduce the overall cost. At moderate

temperatures, placing the hotter fluid in the tubes will

reduce the shell surface temperatures, and hence the need

for lagging to reduce heat loss, or for safety reasons.

Operating

pressures

The higher pressure stream should be allocated to the

tube-side. High-pressure tubes will be cheaper than a high-

pressure shell.



5-11

Factor Description

Pressure drop For the same pressure drop, higher heat-transfer

coefficients will be obtained on the tube-side than the shell-

side, and fluid with the lowest allowable pressure drop

should be allocated to the tube-side.

Viscosity The higher heat-transfer coefficient will be obtained by

allocating the more viscous material to the shell-side,

providing the flow is turbulent.

Stream flow rate Allocating the fluids with the lowest flow-rate to the shell-

side will normally give the most economical design.


Based on the factors in Table 5.4, it can be conclude, the fluid that want to

be cool down was allocate in shell side which is gasses, and the cooling fluid which

is water will be allocate in the tube side.

5.3.2.1 Number Of Tube

Before the number of tube in the heat exchanger is determine, the arrangement of

tube inside the heat exchanger must be known first. The tubes in an exchanger are

usually arranged in an triangular pattern as shown in Figure 5.3 . The triangular

pattern gives higher heat transfer rates. This type of tube arrangement was

commonly used in industry.

The recommended tube pitch (distance between tube centres) is 1.25 times

the tubes outside diameter, and this will normally be used unless process

requirements dictate otherwise.


Figure 5.3.: Triangular tubes pattern

Flow

Pt


5-12

The preferred lengths of tubes for heat exchanger that commonly used in

industries are 16ft (4.88mm). For diameter size of tube, 16-25 mm is preferred for

most duties, as they will give more compact and therefore cheaper exchanger.

Hence, take the value of outside and inside diameter in this range. 20 mm for

outside diameter and 16 mm for inside diameter have been choose for this purpose.


Area of one tube ( neglecting thickness of tube sheet ) :

Lda o (5.9)

88.402.0 a

2307.0 m

Hence, the total number of tubes are :

a

AN t (5.10)

2

2

307.0

57.173

m

m

565

Taking the number of passes of tube is 4 as mention earlier, therefore the number of

tubes per pass, pN are :

4

tp

NN (5.11)

4

565

141

5.3.2.2 Tube Side Velocity

Tube cross sectional area, 4

2

ics

dA

(5.12)

4

)016.0( 2

410011.2


5-13

Area per pass, pcsp NAA (5.13)

14110011.2 4

2028.0 m

Water mass velocity, p

tt

A

mV (5.14)

20284.0

/12

m

skg

2/57.428 smkg

Tube side velocity,

tt

Vu (5.15)

3

2

/818.995

/57.428

mkg

smkg

sm /430.0

5.3.2.3 Tube Side Heat Transfer Coefficient

Reynolds number, t

ittt

du

Re (5.16)

where ;

tRe = Reynold number of fluid in tube side

t = fluid (water) density of tube side, 995.818 kg/m3

tu = fluid velocity of tube side, m/s

t = fluid dynamic viscosity of tube side, Ns/m2

id = inside diameter of tube side, m

Therefore;

23

3

/105986.0

016.0/430.0/818.995Re

mNs

msmmkgt

11445

Prandtl number, t

tp

k

C Pr (5.17)


5-14

where;

Pr = Prandtl number

pC = fluid heat capacity, J/kgoC

tk = fluid thermal conductivity of tube side, W/moC

therefore;

CmW

mNsCkgJo

o

/6376.0

/105986.0/102.4Pr

233

94.3

And ratio of, 305016.0

88.4

m

m

d

L

i

(5.18)

Then, find the heat transfer factor,hj . It is often convenient to correlate heat transfer

data in terms of a heat transfer. The hj value can be obtained from Figure E.3

(Appendix E) based on the Reynolds number and the ratio ofidL / . Hence, the heat

transfer factor is 0.0039.

Nusselt number, 33.0PrRe tht jNu (5.19)

33.0)94.3(114450039.0

18.70

Tube side heat transfer coefficient:

i

tti

d

kNuh

(5.20)

m

CmW o

016.0

/6376.018.70

CmWo2/67.2796

5.3.3 Bundle and Shell Diameter

The bundle diameter depends not only the number of tubes but also on the number

of tube passes, as space must be left in the pattern of tubes on the tube sheet to

accommodate the pass partition plates.


5-15

An estimation of the bundle diameter bD can be obtained from equation 5.21

which is an empirical equation based on standard tube layouts. The constant for use

in the equation, for triangular and square pattern are given in table E4 (Appendix E).


By choosing the triangular pattern of tube and 4 tube passes, the value of 1K is

0.175 and the value of 1n is 2.285. Hence ,

1

1

1

nt

obK

NdD

(5.21)

285.2

1

175.0

565020.0

m

m687.0

The clearance required between the outermost tubes in the bundle and the shell

inside diameter will depend on the type of exchanger and the manufacturing

tolerance. Split ring floating head exchangers have been choose for efficiency and

ease of cleaning. Based on figure E5 (Appendix E) the shell clearance is 64mm.

Hence, shell inside diameter,

bs DD shell clearance (5.22)

064.0687.0

751.0

5.3.3.1 Shell-side Heat Transfer Coefficient

The complex flow pattern on the shell side, and the greater number of variables

involved, make it difficult to predict the shell-side heat transfer corfficient and

pressure drop with complete assurance.

Kern’s method was choose to determine these heat transfer in shell side.

This method is base on experimental work on commercial exchanger with standard

tolerances and will give a reasonably satisfactory prediction of the heat transfer

coefficient for standard designs.



5-16

In order to calculate the heat transfer on the shell side, the number of baffle

spacing must be estimate first. Baffle spacing are used in the shell to direct the fluid

stream across the tubes, to increase the fluid velocity and so to improve the rate of

transfer. The most commonly used type of baffle is the single segmental baffle

spacing. Take the baffle spacing equal to 5 because this spacing should give good

heat transfer.


Therefore, take the baffle spacing equal to 5, hence :

5

sb

DI (5.23)

5

751.0 m

150.0

For triangular pitch, tube pitch ot dp 25.1 (5.24)

020.025.1

025.0

Hence, cross flow area ;

t

bsots

p

IDdpA

)( (5.25)

025.0

150.0751.0)020.0025.0( sA

202253.0 m

The shell side equivalent diameter ( hydraulic diameter ).

)917.0(10.1 22

ot

o

e dpd

d (5.26)

))020.0(917.0025.0(020.0

10.1 22

20142.0 m


5-17

Volumetric flow rate on shell side ;

m

s (5.27)

3/704.13

1

3600

120890

mkgs

h

h

kg

sm /423.0 3

Therefore, shell side velocity ;

s

ss

Au

(5.28)

2

3

02253.0

/423.0

m

sm

sm /77.18

Reynolds number, s

esss

du

Re (5.29)

23

3

/101621.0

0142.0/77.18/704.13

mNs

msmmkg

22532

Prandtl number, s

ss

k

Cp Pr (5.30)

CmW

mNsCkgJo

o

/1150.0

)/101621.0)(/10246.3(Pr

233

6.4Pr

In order to find the heat transfer coefficient of shell, the baffle cut must select first.

Baffle cut is used to specify the dimension of a segmental baffle, expressed as a

percentage of the baffle disc diameter. Baffle cut from 15% to 45% are used.

Generally , a baffle cut of 20% to 25% will be the optimum, giving good heat transfer

rates, without excessive pressure drop. From this selection of baffle cut, the heat

transfer factor,nj can be determine.


Based on Figure E.6 (Appendix E), at baffle cut percent equal to 25% and Reynold

number equal to 22532, heat transfer factornj equal to 0.004.


5-18

Hence, shell-side heat transfer coefficient,

3/1

Re rsn

f

es Pjk

dh (5.31)

where;

sh = heat transfer coefficient of shell-side, W/m2°C

ed = inner diameter of tube-side, m

fk = thermal conductivity of shell tube, W/m°C

nj = heat transfer factor of shell-side

sRe = Reynolds number of shell side

Pr = Prandtl number of shell-side

Therefore,

e

renf

sd

PRjkh

33.0

0142.0

)6.4(22532004.01150.0 33.0

CmWo2/76.1207

5.3.4 Overall Heat Transfer

Taking material of construction is carbon steel, wk = 55 W/m°C

Overall heat transfer coefficient,

0003.01

2

ln

0002.011

sw

i

o

o

i

o

to hk

d

dd

d

d

hU (5.32)

0003.076.1207

1

)55(2

016.0

020.0ln020.0

016.0

020.00002.0

67.2796

1

WCmo

/1077612.1 23

CmWUo

o

2/03.563

This is above the initial estimation of 160 W/m2 0C. The number of tube could

possibly be reduced, but first check the pressure drop.


5-19

5.3.5 Pressure Drop

In many applications, the pressure drop available to drive the fluids through the

exchanger will be set by the process conditions. When the designer is free to select

the pressure drop, an economic analysis can be made to determine the exchanger

design that gives the lowest operating costs. The value that suggested in designing

of this heat exchanger are shown in table 5.5.

Table: 5.5 : Allowable pressure drop

Phase Allowable Pressure Drop

Liquid 35kN/m2

Gas 0.4-0.8kN/m2


5.3.5.1 Tube Side Pressure Drop

Based on Figure E.7 (Appendix E) at Reynolds number , 11445 the tube friction

factor fj equal to 0.0048.

Therefore, the pressure drop on tube side ;

25.28

2

t

m

wi

fpt

u

d

LjNP

(5.33)

Neglect the viscosity correction term,

m

w

equation 5.33 becomes;

25.28

2

t

i

fpt

u

d

LjNP

(5.34)

where;

pN = Number of tube passes, 4

fj = friction factor

L = tube length, 4.88m

id = inside diameter of tube, 0.016m

tu = fluid velocity in tube-side, 0.430m/s

= fluid density in tube-side, 995.818kg/m3


5-20

Therefore,

∆Pt =

2

/43.0/818.9955.2

016.0

88.40048.084

23 smmkg

m

mPt

= 5233.62 N/m2

The pressure drop is in range of specification.

5.3.5.2 Shell-side Pressure Drop

Based on figure E8 (Appendix E) at Reynolds number 22532 , the tube friction

factor fj equal to 0.044.

Therefore, shell-side pressure drop;

1402

28

.

w

s

Be

sfs

u

l

L

d

DjP

(5.35)

Neglect the viscosity correction term,

m

w

equation 5.35 becomes;

28

2

s

Be

sfs

u

l

L

d

DjP

(5.36)

where;

L = tube length, 4.88m

bI = baffle spacing, 0.150m

ed = equivalent diameter,0.0142 m

su = fluid velocity in shell-side, 18.77m/s

Therefore,

2

/77.18/7035.13

150.0

88.4

0142.0

751.0044.08

23 smmkg

m

m

m

mPs

= 1462.02 kN/m2

This value of pressure drop is exceeding specification and need to be modified. In

order to doing this, the shell side velocity must be reduced by increasing the baffle

spacing.


5-21

Hence, the new baffle spacing is ;

bI = old baffle spacing/0.3 (5.37)

3.0

150mm

mm500

By using the new value of baffle spacing, the new properties in shell side are;

20751.0 mAs

smus /63.5

6758Re s , 007.0nj , 053.0fj

CmWho

s

2/92.633

barP 47.0

This pressure drop is in a range of specification.

New overall heat transfer coefficient,

0003.01

2

ln

0002.011

sw

i

o

o

i

o

to hk

d

dd

d

d

hU

0003.092.633

1

)55(2

016.0

020.0ln020.0

016.0

020.00002.0

67.2796

1

WCmo

/105256.2 23

CmWUo

o

2/95.395


5-22

Table 5.6: Summary of Chemical Design of Heat Exchanger

Parameters SI unit

Process condition:

Heat load, Q

Heat transfer coefficient assume, Uass

Heat transfer coefficient calculate, Ucalc

Heat transfer area

∆Tlm

∆Tm

1.506 x 102 kW

160 W/m2.°C

395.95 W/m².°C

173.57 m2

61.66°C

54.26°C

Shell side

Inlet temperature, T1

Outlet temperature, T2

Flow rate,

m s

Shell side velocity, us

Diameter of shell, Ds

Bundle diameter, Db

Equivalent diameter, de

Shell passes

Heat Transfer Coefficient, hs

Pressure drop, ∆Ps

150°C

70°C

20890 kg/h

5.63 m/s

0.751 m

0.687 m

0.0142 m

2

633.92 W/m2.°C

47000 N/m2

Tube side: Water

Inlet temperature, t1

Outlet temperature, t2

Flow rate,

m t

Tube velocity, ut

Tube length

Outer diameter, do

Inner diameter, di

Birmingham Wire Gage (BWG)

Tube pitch, pt

Number of tube, Nt

Tube per pass

Tube passes, Np

Heat transfer coefficient, hi

Pressure drop, ∆Pt

30°C

60°C

12.0 kg/s

0.430 m/s

4.88 m

0.020 m

0.016 m

16

0.0238 m

565

150

4

2796.67 W/m2.°C

5233.62 N/m2


5-23

5.4 MECHANICAL DESIGN OF HEAT EXCHANGER

The mechanical design is a function of the equipment, the operating pressure and

temperature the equipment dimension, the opening and connection and the material

of construction. The main details of mechanical design include the followings:

Operating and design temperature and pressure

Material of construction

Corrosion allowance

Shell side

1. Shell thickness

2. Head and closures

3. Nozzles

4. Flanges

Tube side

1. Tube thickness

2. Nozzles

3. Flanges

Insulator thickness

Weight load heat exchanger

1. Vessel weight

2. Tubes weight

3. Weight of mixture to fill the shell vessel

4. Weight of water to fill the tubes

5. Weight of insulator

Support design



5-24

Previous calculations from chemical engineering design will be used in the

calculation of mechanical design.

Figure 5.4: Part of heat exchanger


Table 5.7: Part of heat exchanger

Number Description

1 Shell

2 Channel head

3 Channel cover

4 Nozzle

5 Tube


5.4.1 Design Pressure

The heat exchanger must be design to withstand the maximum pressure to which it

is likely to be subjected in operation. The design pressure is normally taken at 10%

above normal working operation. The purpose is to avoid spurious operation during

minor process upsets.



5-25

By taking a safety factor of 10%;

For shell-side; Pi = Po x 1.1

= 20 bar

= 2 N/mm2 x 1.1

= 2.2 N/mm2

For tube-side; Pi = Po x 1.1

= 20 bar

= 2 N/mm2 x 1.1

= 2.2 N/mm2

Table 5.8: Design pressure of shell and tube

Parameter Shell side Tube side

Operating pressure, bar 20 20

Design pressure, N/mm2 2.2 2.2

5.4.2 Design Temperature

The strength of metals decreases with increasing temperature so the maximum

allowable design stress will depend on the material temperature. The design

temperature at which the design stress is evaluated should be taken as the

maximum working temperature of material, with due allowance for any uncertainty

involved in predicting vessel wall temperature, therefore taking a safety factor of

10% to cover uncertainties in temperature prediction.


By taking a safety factor of 10%;

For shell-side; Ti = To x 1.1

= 150°C x 1.1

= 165°C

For tube-side; Ti = To x 1.1

= 60°C x 1.1

= 66°C


5-26

Table 5.8: Design temperature of shell and tube

Parameter Shell side Tube side

Operating temperature, °C 150 60

Design temperature, °C 165 66

5.4.3 Material of Construction

Selection of a suitable material must take into account the suitability of the material

for fabrication as well as the compatibility of the material with the process

environment. The material which is fit to the chemical and mechanical requirements

and the same time the most economical should be selected. Carbon steel has been

choosing as a material of construction due to more cheaply than stainless steel and

high corrosion resistance. A few factors that should be considered while choosing

the material of construction are:

1. Corrosion resistance

2. Operating conditions

3. Economic feasibility

4. Suitability for fabrication

5. Process safety


5.4.4 Welded Joint Efficiency and Corrosion Allowance

The strength of welded joint will depend on type of joint and the quality of the

welding. The soundness of weld is then checked by visual inspection and by non-

destructive testing called radiography. The welded joint factor, J is taken as 1.0.

The corrosion allowance is additional thickness of metal added to allow for

material lost by corrosion and erosion or scaling. The allowance is based on

experience with the material of construction under similar service condition to those

for the purposed design. A minimum corrosion allowance used is 2 mm for carbon

steel material of construction.



5-27

5.4.5 Design Stress (Nominal Design Strength)

For the design purpose, it is necessary to decide a value for the maximum allowable

stress (nominal design stress) that can be accepted in the material of construction.

The allowable stress for the selected material of construction at the design

temperature shows in Table 5.10

Table 5.10: Design stress for material construction

Material used Design stress ,f (N/mm2)

Shell: Carbon steel 105 @ 200°C

Tube: Carbon steel 125 @ 100°C


5.4.6 Minimum Practical Wall Thickness

This is required to ensure that any vessel is sufficiently rigid to withstand its own

weight, and any incidental loads. From previous calculation in chemical engineering

design, the internal diameter of shell, Ds = 0.751 m. For a cylindrical shell, the

minimum thickness required to resist internal pressure can be determined as

follows:

Minimum wall thickness, i

ii

PJf

DPe

2 (5.38)

where;

iP = internal design pressure of shell, N/mm2

iD = Shell diameter, mm

J = Joint factor (J=1)

f = Design stress of shell, N/mm2

Therefore,

22

2

/2.2/10512

751/2.2

mmNmmN

mmmmNe

mm95.7

Adding corrosion allowance of 2mm;

mmmme 295.7

mm95.9


5-28

5.4.7 Minimum Thickness of Tube Wall

The minimum thickness required for the tube: i

ii

PJf

DPe

2 (5.39)

where;

iP = internal design pressure of tube, N/mm2

iD = internal tube diameter, mm


f = Design stress of tube side, N/mm2

Therefore;

22

2

/2.2/10512

16/2.2

mmNmmN

mmmmNe

mm17.0


mmmme 217.0

mm17.2

5.4.8 Head and Closure

There are several types of head and closure as describe in Table 5.11. For this

design, ellipsoidal heads was chosen since the operation pressure is less than 15

bar and this types of heads most economical. The standard ellipsoidal heads are

manufactured with a major and minor axis ratio of 2:1.

Table 5.11: Selection of head and closure

Types of head Advantages

Hemispherical heads Suitable for high pressure

Higher cost

The strongest shape

Ellipsoidal heads Most economical for operation above

15 bar

Torispherical Suitable for operation above 15 bar



5-29

5.4.8.1 Ellipsoidal Head

Minimum thickness required, i x P.-s x J x f

iDiPe=

202 (5.40)

where;

iP = internal design pressure of shell, N/mm2

iD = Shell diameter, mm


f = Design stress of shell, N/mm2

Therefore;

22

2

/2.22.0/10512

751/2.2

mmNmmN

mmmmNe

mm88.7


mmmme 288.7

mm88.9

5.4.8.2 Channel covers (Closures)

Minimum thickness required, 2

1

e

i

epD

PDCe (5.41)

where;

pC = design constant that depend on the edge constraint

eD = nominal plate diameter, mm

iP = internal pressure of shell, N/mm2

sf = design stress of shell side, N/mm2

Values for the design constant, Cp and the nominal plate diameter, De are given in

the design codes and standards for various arrangements of flat end closures (BS

5500, clause 3.5.5). From Coulson & Richardson’s Chemical Engineering Volume 6,

plates welded to the end of the shell with a fillet weld, angle of fillet 45oC and depth

equal to the plate thickness, take Cp as 0.4 and De=Di;


5-30

Therefore,

2

1

105

2.27514.0

e


mmmme 248.43

mm48.45

5.4.9 Weight Loads

5.4.9.1 The Shell Weight

For preliminary calculation, the approximate weight of a cylindrical vessel with dome

ends, and uniform wall thickness, can be estimated using the equation below:

Vessel weight, 310)8.0( tDLgDCW mmmvv (5.41)

where;

Wv = Total weight of the shell, N

Cv = Factor for the weight of nozzles for vessels with only a few internal

fittings (1.08)

L = Length of tube, m

g = gravitational acceleration, 9.81 m/s2

t = wall thickness, mm

m = density of vessel material, 7854 kg/m3

Dm = mean diameter of vessel, m


)10( 3 tDD im (5.42)

)1095.9751.0( 3

mm76.0

Therefore;

)00995.0))(76.0(8.088.4)(/81.9)(76.0)(/7854()08.1( 23 msmmmkgWv

N12730


5-31

5.4.9.2 Weight of Tubes

Weight of tubes, gddNW miott )(22

(5.43)

where;

Nt = number of tubes,

do = outside diameter of tube, m

di = inside diameter of tube, m

m = density of tube material, kg/m3

Therefore;

2322 /81.9/7854)))016.0()02.0((565 smmkgmmWt

N19693

5.4.9.3 Weight of fluid to fill the shell

Weight of fluid (gas), 4

2 x gs x L x ρsπ x D

Wg= (5.44)

where:

Ds = diameter of shell side, m

L = length, m

s = density of shell-side, kg/m3

g = gravitational acceleration, m/s2

Therefore;

4

/81.9/7035.1388.4751.0 232smmkgmm

Wg

N6.290


5-32

5.4.9.4 Weight of water to fill the tube

Weight of coolant,

4

22gLρddxN

W tiot

water

(5.45)

where:

t = density of water in tube, kg/m3

Therefore;

4

)/81.9)(/818.995)(88.4()016.0()02.0(565 2322 smmkgmmmWwater

N3.3046

5.4.9.5 Weight of Insulator

Material used as insulator is mineral wool. From Coulson & Richardson’s Chemical

Engineering Volume 6, the density of mineral wool insulation is 130 kg/m3.

Approximate volume of insulation, iDLeV (5.46)

where:

V = Approximate volume of insulation, m3

L = Length of tube, m

ei = Thickness of insulator, m

Volume of insulation, )0187.0)(88.4)(751.0( mmV

3215.0 m

Weight of Insulator, gVW ii (5.47)

where:

Wi = weight of insulation, kgm/s2

i = insulation density, kg/m3

Therefore;

233 /81.9/130215.0 smmkgmWi

N2.274


5-33

The total weight of heat exchanger;

iwatergtvT WWWWWW (5.48)

NNNNN 2.2743.30466.2901969312730

N1.36034

kN03.36

5.4.10 Baffles

Baffles are used in the shell to direct the fluid flow across tube and increase the fluid

velocity. When the fluid velocity increases, the rate of heat transfer is also improved.

The assembly of baffles and tubes inner diameter hold together by support rods and

spacers. The most commonly used type of baffle is the single-segmental baffle.

Baffle used to specify the dimensions of a segmental baffle. Generally, baffle cut of

20%-25% will be optimum. The value will give good heat transfer rate without

excessive drop. The function of baffles are to support the tubes for structural rigidity,

preventing tube vibration and sagging to divert the flow across the bundle to obtain a

higher heat transfer coefficient.


Baffle diameter = Ds – 3.2 mm (5.49)

= 751 mm – 3.2 mm

= 747.8 mm

Tolerance = Ds + 0.8 mm (5.50)

= 751 mm + 0.8 mm

= 751.8 mm

Baffle spacing, IB = Ds/5 (5.51)

= 751 mm /5

= 150 mm

Baffle modification; = 150mm/0.3

= 500mm


5-34

Number of baffle, Nb = (L/IB) -1 (5.52)

= (4880 mm/500 mm) -1

= 8.76 ≈ 9 baffles

Baffle Thickness = 4.76 mm (Adapted from British Standard (DIN 38025)

5.4.11 Nozzle (Branches)

Nozzles are used for entering and leaving the inlet and outlet stream of heat

exchanger. The nozzles are for channel side and the shell side of heat exchanger.

Standard steel pipe will be used for the inlet and outlet nozzles are obtained from

Perry’s Handbook (Table 10-18). It is important to avoid flow restrictions at the inlet

and outlet nozzles. It is also to prevent excessive pressure drop flow induced

vibration of the tubes. Material of construction for nozzle will be the same as the

heat exchanger body.

(source:Perry's Chemical Engineering Handbook)

5.4.11.1 Shell-side Nozzles

Table5.12: Properties for the shell-side

Properties Inlet Outlet

Temperature, °C 150 70

Density, , kg/m3 13.70 13.70

Flow rate, m, kg/s 5.80 5.80

Fluid velocity, u, m/s 5.63 5.63

Flow area, A, m2 [A = m/ x u] 2.38 2.38

Inside diameter, mm 751 751

By referring to the standard properties of steel pipe from Table 10-18

(Perry’s Handbook), the standard nominal pipe size was taken as 30 in. The

properties at this nominal size are show below:


5-35

Table 5.13: Properties of pipe of shell-side

Nominal

size, in

Outside diameter,

OD, in

Schedule

no.

Inside diameter,

ID, in

Flow area,

ft2

30 30 5S 29.5 4.746

5.4.11.2 Tube-side Nozzles

Table 5.14: Properties of tube-side

Properties Inlet Outlet

Temperature, °C 30 60

Density, , kg/m3 995.81 995.81

Flow rate, G, kg/s 12.00 12.00

Fluid velocity, u, m/s 0.43 0.43

Flow area, A, m2 [A = G/ x u] 0.0052 0.0052

Inside diameter, m [(4 x A/)1/2] 0.0813 0.0813

Inside diameter, mm 81.3 81.3

By referring to the standard properties of steel pipe from Table 10-18

(Perry’s Handbook), the standard nominal pipe size was taken as 4.0 in. The

properties at this nominal size are show below:

Table 5.15: Properties of pipe of tube-side

Nominal size,

in

Outside

diameter, OD, in

Schedule

no.

Inside diameter,

ID, in

Flow area,

ft2

3.33 4.0 5S 3.834 0.08017

5.4.11.3 Flanged for Nozzle

Flanged joints are used for connecting pipes and instruments to vessel, for

manholes cover and for removal vessel head when ease of access is required.

Flanged may also be used on the vessel body, when it is necessary to divide the

vessel into sections for transport maintenance. Flanged joints are also used to

connect pipe to requirements such as pumps and valves. Flanges range size from a

few millimeters diameter for small pipes to several meters diameter for those used

as body or head flanges on vessels. Flanges dimension must be able to withstand


5-36

the hydrostatic ends loads and the bolt loads necessary to ensure tight joint in

service.

For the design of this heat exchanger, welding-neck flange are used. It is

because welding-neck flanges have a long trapped hub between the flange ring and

the welded joint. This gradual transition of the section reduces the discontinuity

stresses between the flange and branch. It is also can increase the strength of the

flange assembly.

Welding-neck flanges are suitable for extreme service conditions, where

flange are likely to be subjected to temperature, shear and vibration loads. They will

normally be specified for the connections and nozzles on process vessels and

process equipment. The dimensions of flanged for nozzle for nominal size 80 and

300 mm are show in Table 5.16.


Table 5.16: Dimensions of flanged for nozzle

Type

Nom. size

Pipe, o.d. d1

Flange Raised

face Bolting Drilling Neck

D b h1 d4 f No d2 k d3 Shell-side

300 323.9 440 22 44 365 4 M20 12 22 395 355

Tube-side

80 88.9 190 16 34 128 3 M16 4 18 150 110

(All units in mm)

d4

k

D

de

d3

d1


5-37

Figure 5.5: Typical standard flange design

5.4.12 Design of support saddles

The saddles must be designed to withstand the load imposed by the weight of the

vessel and contents. They are constructed of bricks or concrete, or are fabricated

from steel plate. The contact angel should not less than 120oC, and will not normally

be greater than 150oC. Wear plate often welded to the shell wall to reinforce the wall

over the area contact with the saddles.


Table 5.17: Dimensions for saddle support

Vessel

diameter,

(m)

Maximum

weight,

(kN)

Dimensions, (m) mm

V Y C E J G t1 t2 Bolt

dia.

Bolt

holes

0.8 50 0.58 0.15 0.70 0.29 0.225 0.095 8 5 20 25

Figure 5.6: Standard steel saddles


5-38

Table 5.18: Summary of mechanical design of shell and tube heat exchanger

PARAMETER SPECIFICATION

Design Pressure

+10% above normal working of operations.

Shell : 2.2 N/mm2

Tube : 2.2 N/mm2

Design Temperature

+ 10% to cover uncertainties in prediction.

Shell : 165°C

Tube : 66°C

Material of construction Shell : Carbon steel

Tube : Carbon steel

Design Stress

Take above or nearest the design temperature

Shell : 105 N/mm2@200°C

Tube : 125 N/mm2@100°C

Minimum Thickness Shell : 9.95 mm

Tube : 2.17 mm

Head and closure

Head: Ellipsoidal head type

Closure: Channels cover type

Thickness : 9.88 mm

Thickness : 44.48 mm

Weight Load

Weight of shell

Weight of tubes

Weight of gas to be filled the vessel

Weight of water to be filled in the tube

Weight of insulator

Total weight

12.73 kN

16.69 kN

0.29 kN

3.05 kN

0.27 kN

36.03 kN


5-39

REFERENCES

R K Sinnott. Third Edition, 1999. Coulson & Richardson’s Chemical Engineering

Volume 6. Elsevier Butterworth Heinemann. 634-869.

Frank P. Incropera & David P. Dewitt, Fifth Edition, 2002. Fundamentals of Heat and

Mass Transfer. John Wiley & Sons, U.K. 924

Robert H. Perry & Don W. Green. 1997. Perry’s Chemical Engineers Handbook.

Seventh Edition. Mc Graw Hill.

Carl L. Yaws. Chemical Properties Handbook (Physical Thermo, Environment,

Transport, Safety and Health for Organic and Inorganic Chemicals). Mc

Graw Hill.

Yunus A. Cengel & Michael A. Boles. Third Edition, 1998. Thermodynamics An

Engineering Approach. McGraw-Hill.

Dr. Brian Spulding & J.Tab Orela. 1990. Heat exchanger Theory and Design

Handbook. McGraw-Hill.


5-40

APPENDIX E


5-41

Figure E.1: Temperature correction factor : two shell passes and four tube

passes


5-42

Table E.2: Typical Overall Coefficient


5-43

Figure E.3: Tube Side Heat Transfer Factor


5-44

Table E.4: Constant for Tube Arrangement

Figure E.5: Shell-bundle Clearance


5-45

Figure E.6: Shell Side Heat Transfer Factor


5-46

Figure E.7: Tube Side Friction Factor


5-47

Figure E.8: Shell Side Friction Factor


5-48

Chapter 5 - Heat Exchanger

Documents

process of heat exchanger

simplest heat exchanger

tubular heat exchanger

concentric tube heat

heat energy

fundamentals of heat

types of heat exchangers

used type of heat