design of heat exchanger 2

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Design of Heat Exchanger (2)

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Dr Abrar Inayat

Chemical Engineering Department Universiti Teknologi PETRONAS

Tronoh

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Recap

• Basics of heat exchangers

• Types of heat exchangers

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Outcome

• Calculation used to design a heat exchanger for a process plant

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Shell and Tube Exchangers

• Tube-sheet layout (tube count)

• The bundle diameter will depend not only on thenumber of tubes but also on the number of tubepasses, as spaces must be left in the pattern of tubeson the tube sheet to accommodate the passpartition plates

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(3)

mm diameter, outside tube

mm diameter, bundle

tubesofnumber

where,

0

1

1

0

0

1

1

1

d

D

N

K

NdD

d

DKN

b

t

n

tb

n

bt

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3


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Table 4. Constants for use in eq. 3

where Pt = tube pitch, mm

Shell and tube exchangers

• Shell types (passes)

• The principal shell arrangements are shown inFigure 12.12a-e.

• The letters E, F, G, H, J are those used in theTEMA standards to designate the varioustypes

• The E shell is the most commonly usedarrangement

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Shell and tube exchangers

• Two shell passes (F shell) are occasionally usedwhere the shell and tube side temperaturedifferences will be unsuitable for a single pass

• The divided flow and split-flow arrangements(G and J shells) are used to reduce the shell-side pressure drop; where pressure drop,rather than heat transfer, is the controllingfactor in the design

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Shell and tube exchangers• Shell and tube designation:

• A common method of describing an exchangeris to designate the number of shell and tubepasses: m/n

• Where, m is the number of shell passes and nthe number of tube passes

• Example 1: 1/2 describes an exchanger with 1shell pass and 2 tube passes

• Example 2: 2/4 an exchanger with 2 shellpasses and 4 four tube passes

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Shell and tube exchangers• Baffles:

• Baffles are used in the shell– to direct the fluid stream across the tubes

– to increase the fluid velocity

– and so improve the rate of transfer

• The most commonly used type of baffle is thesingle segmental baffle shown in Figure12.13a, other types are shown in Figures12.13b, c and d.

• Only the design of exchangers using singlesegmental baffles will be considered

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Mean temperature difference• In order to determine the heat transfer area

required for a given duty, the meantemperature difference must be calculated.

• That is, the difference in the fluidtemperatures at the inlet and outlet of theexchanger

• logarithmic mean temperature difference Tlmis only applicable to sensible heat transfer intrue co-current or counter-current flow (lineartemperature enthalpy curves)

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Mean temperature difference• For counter-current flow, Fig 12.18a, the logarithmic

mean temperature is given by

• The equation is the same for co-current flow, but theterminal temperature differences will be (T1 - t1) and(T2 - t2).

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outlet re, temperatufluid cold t

inlet re, temperatufluid cold t

outlet re, temperatufluidhot T

inlet re, temperatufluidhot T

difference raturemean tempe log T

where,

ln

2

1

2

1

lm

12

21

1221

tT

tT

tTtTTlm (4)

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Mean temperature difference• The usual practice in the design of shell and

tube exchangers is to estimate the truetemperature difference from the logarithmicmean temperature by applying a correctionfactor to allow for the departure from truecounter-current flow

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factor correction re temperatu

difference re temperatu true

where,

t

m

lmtm

F

ΔT

TFT (5)

Mean temperature difference• The correction factor is a function of the shell and

tube fluid temperatures, and the number of tube andshell passes

• It is normally correlated as a function of twodimensionless temperature ratios

• R is equal to the shell-side temperature differencedivided by the tube-side temperature difference

• S is a measure of the temperature efficiency of theexchanger

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12

21

tt

TTR

11

12

tT

ttS

(6)

(7)

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Mean temperature difference• For a 1 shell : 2 tube pass exchanger, the correction

factor is given by

• The equation for a 1 shell : 2 tube pass exchanger can beused for any exchanger with an even number of tubepasses, and is plotted in Figure 12.19

• The correction factor for 2 shell passes and 4, ormultiples of 4, tube passes is shown in Figure 12.20, andfor divided and split, flow shells in Figures 12.21 and12.22

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112

112ln1

1

1ln1

2

2

2

RRS

RRSR

RS

SR

Ft(8)

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• Fluid allocation: shell or tubes

• Where no phase change occurs, the followingfactors will determine the allocation of thefluid streams to the shell or tubes

• Corrosion:

– The more corrosive fluid should be allocated tothe tube-side. This will reduce the cost ofexpensive alloy or clad components.

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General Design Considerations

• Fouling: The fluid that has the greatesttendency to foul the heat-transfer surfacesshould be placed in the tubes.

• This will give better control over the designfluid velocity, and the higher allowablevelocity in the tubes will reduce fouling. Also,the tubes will be easier to clean

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• Operating pressures:

– The higher pressure stream should be allocated tothe tube-side

– High-pressure tubes will be cheaper than a high-pressure shell

• Pressure drop:

– For the same pressure drop, higher heat-transfercoefficients will be obtained on the tube-side thanthe shell-side

– Fluid with the lowest allowable pressure drop should be allocated to the tube-side

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• Viscosity:

– Generally, a higher heat-transfer coefficient will beobtained by allocating the more viscous materialto the shell-side, providing the flow is turbulent

• If turbulent flow cannot be achieved in theshell it is better to place the fluid in the tubes,as the tube-side heat-transfer coefficient canbe predicted with more certainty

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• Stream flow-rates: Allocating the fluids withthe lowest flow-rate to the shell-side willnormally give the most economical design

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Shell and tube fluid velocities• High velocities will give high heat-transfer

coefficients but also a high-pressure drop

• The velocity must be high enough to preventany suspended solids settling, but not so highas to cause erosion

• High velocities will reduce fouling

• Liquids: Tube-side, process fluids: 1 to 2 m/s,maximum 4 m/s if required to reduce fouling;water: 1.5 to 2.5 m/s.

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Shell and tube fluid velocities

• Shell-side: 0.3 to 1 m/s.

• Vapors: the velocity used will depend on theoperating pressure and fluid density

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Fluid physical properties• The fluid physical properties required for heat-exchanger

design are:

• density, viscosity, thermal conductivity and temperature-enthalpy correlations (specific and latent heats)

• The thermal conductivities of commonly used tubematerials are given in Table 12.6

• In the correlations used to predict heat-transfercoefficients, the physical properties are usually evaluatedat the mean stream temperature

• This is satisfactory when the temperature change issmall, but can cause a significant error when the changein temperature is large

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Tube side heat transfer coefficient and pressure drop: Single phase

• Heat transfer, turbulent flow:

• Heat-transfer data for turbulent flow insideconduits of uniform cross-section are usuallycorrelated by an equation of the form

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wallat the viscosityfluid

re, temperatufluidbulk at the viscosityfluid

area,unit per flow mass velocity,mass

ty,conductivi thermalfluid

m/s velocity,fluid

for tubesperimeter wetted

flowfor area sectional cross 4

m diameter,mean hydraulicor equivalent

t coefficien inside

number Prandtl Pr

number Reynolds Re

number Nusselt where,

PrRe

p

w

t

f

t

ie

e

i

f

p

etet

f

ei

c

w

ba

C

μ

μ

G

k

u

dd

d

h

k

μC

μ

dG

μ

dρu

k

dhNu

CNu

W/moC

kg/m2s

Ns/m2

J/kgoC

Tube side heat transfer coefficient

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liquids sfor viscou 0.027

liquids viscous-nonfor 0.023

gasesfor 0.021 C

where,

PrRe

designexchanger for eq generalA (1936), Tate andSieder

0.14 c

heatfor 0.4 b and coolingfor 0.3 b 0.8, a

14.0

33.08.0

w

CNu

(10)


Tube side heat transfer factor • Heat-transfer factor, jh:

• It is often convenient to correlate heat-transfer data in termsof a heat transfer j factor,

• The heat-transfer factor is defined by:

• more convenient form is given by

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14.0

67.0Pr

w

h Stj

14.0

33.0PrRe

w

h

f

ii jk

dh

(14)

(15)

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• Coefficients for water

• The equation below has been adapted from data given by Eagle and Ferguson (1930):

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mm diameter, inside tube

m/s ocity, water vel

perature, water tem

for water,t coefficien inside

where,

02.035.142002.0

8.0

i

t

i

i

ti

d

u

t

h

d

uth

W/m2oC

oC

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Tube-side pressure drop

• There are two major sources of pressure losson the tube-side of a shell and tubeexchanger:

– the friction loss in the tubes

– the losses due to the sudden contraction andexpansion and flow reversals that the fluidexperiences in flow through the tube arrangement

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Tube-side pressure drop• The tube friction loss can be calculated using the familiar

equations for pressure-drop loss in pipes

• The basic equation for isothermal flow in pipes (constanttemperature) is

• Where, jf is the dimensionless friction factor and L’ is theeffective pipe length.

• The flow in a heat exchanger will clearly not beisothermal, and this is allowed for by including anempirical correction factor to account for the change inphysical properties with temperature

• Normally only the change in viscosity is considered12/2/2014 46

28

2

t

i

f

u

d

LjP

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20

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Tube-side pressure drop

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2100 Re flow,ent for turbul 0.14

2100 Re flow,laminar for 25.0

28

2

m

u

d

LjP

m

w

t

i

f

• The pressure losses due to contraction at the tubeinlets, expansion at the exits, and flow reversal in theheaders, can be a significant part of the total tube-sidepressure drop

• The loss in terms of velocity heads can be estimated bycounting the number of flow contractions, expansionsand reversals, and using the factors for pipe fittings toestimate the number of velocity heads lost

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Tube-side pressure drop• For two tube passes

– there will be two contractions, two expansionsand one flow reversal

• The head loss for each of these effects is: contraction0.5, expansion 1.0, 180o bend 1.5

• Hence, for two passes the maximum loss will be

• 2 × 0.5 + 2 × 1.0 + 1.5 = 4.5 velocity heads

= 2.25 per pass

• Frank’s recommended value of 2.5 velocity heads perpass is the most realistic value to use

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Tube-side pressure drop• Pressure drop equation becomes

• Another source of pressure drop will be the flow expansionand contraction at the exchanger inlet and outlet nozzles.

• This can be estimated by adding one velocity head for theinlet and 0.5 for the outlet, based on the nozzle velocities

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tubeone oflength

m/s velocity,side tube

passes side tubeofnumber

(Pa) drop pressure side tubeΔ

where,

25.28

2

L

u

N

P

u

d

LjNP

t

p

t

t

m

wi

fpt

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Shell side heat transfer and pressure drop: single phase

• Flow pattern:

• The flow pattern in the shell of a segmentally baffledheat exchanger is complex

• This makes the prediction of the shell-side heat-transfercoefficient and pressure drop very much more difficultthan for the tube-side

• Though the baffles are installed to direct the flow acrossthe tubes, the actual flow of the main stream of fluid willbe a mixture of cross flow between the baffles, coupledwith axial (parallel) flow in the baffle windows; as shownin Figure 12.25.

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Flow pattern• Not all the fluid flow follows the path shown

in Figure 12.25

• Some will leak through gaps formed by theclearances that have to be allowed forfabrication and assembly of the exchanger

• These leakage and bypass streams are shownin Figure 12.26, which is based on the flowmodel proposed by Tinker (1951, 1958).

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Flow pattern• In Figure 12.26, Tinker’s nomenclature is used to

identify the various streams, as follows

• Stream A is the tube-to-baffle leakage stream. Thefluid flowing through the clearance between thetube outside diameter and the tube hole in the baffle

• Stream B is the actual cross-flow stream

• Stream C is the bundle-to-shell bypass stream Thefluid flowing in the clearance area between the outertubes in the bundle (bundle diameter) and the shell

• Stream E is the baffle-to-shell leakage stream. Thefluid flowing through the clearance between theedge of a baffle and the shell wall

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Kern’s method• The shell area is calculated using the flow area between the

tubes taken in the axial direction (parallel to the tubes) andthe wetted perimeter of the tubes; see Figure 12.28

• Shell-side jh and jf factors for use in this method are given inFigures 12.29 and 12.30, for various baffle cuts and tubearrangements. These figures are based on data given by Kern(1950) and by Ludwig (2001)

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Kern’s method• The procedure for calculating the shell-side heat-transfer

coefficient and pressure drop for a single shell passexchanger is given below:

• Calculate the area for cross-flow As for the hypotheticalrow of tubes at the shell equator, given by:

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centres bebetween tu distance total

theand besbetween tu clearance the

of ratio theis termthe

m spacing, baffle

m diameter, inside shell

diameter outside tube

pitch tube

where,

0

0

0

tt

B

s

t

t

Bsts

pdp

l

D

d

p

p

lDdpA

Kern’s method• Calculate the shell-side mass velocity Gs and

the linear velocity us

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3

s

kg/m density, fluid side-shell

kg/s side-shell on the rate flow fluid W

where,

s

s

s

ss

Gu

A

WG

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Kern’s method• Calculate the shell-side equivalent diameter

(hydraulic diameter), Figure 12.28.

• For a square pitch arrangement

• For an equilateral triangular pitch arrangement:

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2

0

2

00

2

0

2

785.027.14

4

dpdd

dp

d t

t

e

m diameter, equivalent where,

917.010.1

2

42

187.0

44

2

0

2

00

2

0

d

dpdd

dp

p

d

e

t

tt

e

Kern’s method• Calculate the shell-side Reynolds number, given by:

• From Reynolds number, read the value of jh fromFigure 12.29 for the selected baffle cut and calculatethe shell-side heat transfer coefficient hs from:

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eses dudG

Re

14.0

31PrRe

w

h

f

es jk

dhNu

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Kern’s method

• From shell-side Reynolds number, read thefriction factor from Figure 12.30 and calculatethe shell-side pressure drop from:

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0.142

82

s ss f

e B w

B

B

D uLP j

d l

where,

L tube length

l baffle spacing

the term L l is the number of times the

flow crosses the tube bundl

Example 1

• A horizontal shell-and-tube heat exchangerwith two tube passes and one shell pass isbeing used to heat 9 kg/s of 100% ethanolfrom 15 to 65oC at atmospheric pressure.

• The ethanol passes through the inside of thetubes, and saturated steam at 110oCcondenses on the shell side of the tubes. Theexchanger contains a total of 50 tubes perpass.

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Example 1

• Data to design a shell and tube heatexchanger are given in Table 1.

• Estimate the overall heat transfer coefficientof the exchanger.

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Property at 40oC Ethanol

Heat capacity, cp 2.594 kJ/kgK

Density, 785 kg/m3

Viscosity, µ 0.0009 Pa.s

Thermal conductivity, k 1.63×10-4 kJ/s.m.K

Fouling coefficient, hi,d 5000 W/m2.K

Exchanger configuration

Tube outside diameter, do 0.019 m

Tube inside diameter, di 0.015 m

Flow area per tube, Ac 0.000177 m2

of ethanol at 89oC 0.0004 Pa.s

Steam film coefficient, ho 1.0×104 W/m2.K

for steel 0.045 kJ/s.m.K

w

wk

Table 1: Properties of ethanol and exchanger configuration

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Formulas

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c

iA

mG

iiGd

N Re

k

cpPr

14.0

31

Pr

8.0

Re023.0

wi

i NNd

kh

oo

i

w

i

o

i

diii hd

d

k

d

dd

hhU

1

2

ln111

,

Solution

• Mass velocity in each tube

• Reynolds number

• Prandtl number

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2k/s.m95.101650000177.0

9

c

Ti

A

mG

15.169490009.0

95.1016015.0Re

iiGd

32.141063.1

0009.0594.2Pr

4

k

Cp

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Solution

• Ethanol film coefficient

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KW/m58.1643

0004.0

0009.032.1415.16949

015.0

10001063.1023.0

023.0

2

14.0

318.04

14.0

31

Pr

8.0

Re

i

i

wi

i

h

h

NNd

kh

Solution

• Overall heat transfer coefficient:

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KW/m1079U

K/Wm102677.91

1

2

ln111

2

i

24

,

i

oo

i

w

i

oi

diii

U

hd

d

k

d

dd

hhU

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Example 2: Kern’s Method

• Design an exchanger to sub-cool condensatefrom a methanol condenser from 95oC to 40oCusing Kern’s method. Flow-rate of methanol100,000 kg/h. Brackish water will be used as thecoolant, with a temperature rise from 25oC to40oC.

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Using Kern’s method

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Thank You

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design of heat exchanger 2

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