Basics of Rating and Thermal Design of Heat Exchangers · King Abdulaziz University Faculty of Engineering . Mechanical Engineering Department. Basics of Rating and Thermal Design

King Abdulaziz University

Faculty of Engineering

Mechanical Engineering Department

Basics of Rating and Thermal

Design of Heat Exchangers

General notes prepared for the course

MEP 460: Design of Heat Exchangers

by

Abdul-Rahim A. Khaled, Professor

2014-2015

�سم ا� ا�ر�ن ا�رحيم

Basics of Rating and Thermal Design of HXs. Prof. A.-R.A. Khaled

(i)

Table of Contents

No. Topic Page

1 Revision 1

2 General Correlations for Internal Flow 8

3 Classifications of Heat Exchangers 16

4 Basic Equations for Heat Exchangers 27

5 Fouling of Heat Exchangers 45

6 Double Pipe Heat Exchangers (DPHXs) 54

7 Shell-and-Tube Heat Exchangers (STHEs) 65

8 Compact Heat Exchangers (Compact HXs) 86

9 Boiling 100

10 Condensation 109

11 The Gasketed-Plate Heat Exchangers 117

12 Condensers and Evaporators 133

13 References 146

�سم ا� ا�ر�ن ا�رحيم


(1)

Internal Flow

General Information

Mass flow rate, m :

Cm Aum ,

CA : Flow area; mu : Mean flow velocity normal to cross-section

Mean bulk Temperature, Tm:

p

ACp

mcm

TdAucT C

Newton’s law of cooling for the Local Heat Flux, q

ms TThq

Reynolds number, ReD:

w

hmD P

mDuRe

4

Hydraulic diameter, Dh:

w

Ch P

AD

4 , wP : Wetted perimeter


(2)

For circular pipe,

ihiwiC dD;dPP;dA 2

4

Onset of turbulence occurs at a critical Reynolds number of

2300cr,DRe

Fully turbulent conditions occurs when

00010,ReD

Fully developed flow

A) Hydrodynamically fully developed flow

Hydrodynamic Entry Length

Laminar: Dh,fd Re.Dx 050

Turbulent: 6010 Dx h,fd

Requirement for fully developed hydrodynamically flow condition:

0

h,fd

w

h,fd x

d

x

u


(3)

B) Thermally fully developed flow

Thermal Entry Length

Laminar: PrRe.Dx Dt,fd 050

Turbulent: 6010 Dx t,fd

For laminar flow, how do hydrodynamic and thermal entry lengths compare for a gas?

And oil? A liquid metal?

Can a flow be developing hydrodynamically and be thermally fully-developed?

Requirement for thermally fully developed flow condition:

0

t,fdms

s

xTxT

x,rTxT

x

xfk

h

TT

kq

TT

rT

TT

TT

r ms

s

ms

rr

rrms

s i

i

xfh


(4)

Calculation of mean bulk temperature

mpmmmpconv dTcmTdTTcmdq

Integrating from the tube inlet to outlet Lxx 0 : i,mo,mpconv TTcmq

Uniform Heat Flux ttanconsqs

PdxqdTcmdq smpconv

p

sm

cm

Pq

dx

dT


(5)

Integrating from the tube inlet to outlet Lxx 0

mss

p

si,mo,m TThq;

cm

PLqTT

Uniform Surface Temperature ttanconsTs

dxTThPdTcmdq msmpconv

dxcm

hP

TT

dT

pms

m

Integrating from the tube inlet to outlet Lxx 0

L

pi,ms

o,ms hdxL

h;cm

PLhexp

TT

TT

0

1


(6)

Uniform External Fluid Temperature ttanconsT

totpi,m

o,m

Rcmexp

TT

TT

1

oo

W

ii

totAh

RAh

R11

LdA ii ; LdA oo ;

Lk

ddlnR

tube

ioW

2

i,m

o,m

i,mo,m

tot

i,mo,mpconv

TT

TTln

TTTT

RTTcmq

1

Overall heat transfer coefficients U,U,U oi

ooii

totAUAUUA

R111

itube

ioi

oo

i

iii

o

tube

ioo

oo hk

ddlnd

hd

d

U;

hd

d

k

ddlnd

hU

1

2

111

2

11


(7)

Problems

P1. Exhaust gases from a wire processing oven are discharged into a tall stack,

and the gas and stack surface temperatures at the outlet of the stack must be

estimated. Knowledge of the outlet gas temperature Tm,o is useful for

predicting the dispersion of effluents in the thermal plume, while knowledge

of the outlet stack surface temperature Ts,o indicates whether condensation

of the gas products will occur. The thin-walled, cylindrical stack is 0.5 m in

diameter and 6.0 m high. The exhaust gas flow rate is 0.5 kg/s, and the inlet

temperature is 600 C. Estimate the outlet gas and stack surface

temperatures if the convection heat transfer coefficients are

KmW.hi

2210 and KmW.ho

2913 . Take CT 4 and

kgKJc air,p 1104 .

P2: A hot fluid passes through a thin-walled tube of 10-mm diameter and 1-m

length, and a coolant at free stream temperature of 25 C is in cross flow

over the tube. When the flow rate is 18 kg/h and the inlet temperature is 85

C, the outlet temperature is 78 C. Assuming fully developed flow,

determine the outlet temperature if the flow rate is increased by a factor of

2. That is, the flow rate is 36 kg/h, with all other conditions the same.


(8)

General Correlations for Internal Flow

Friction factor for hydrodynamically fully developed flow

The pressure drop may be determined from knowledge of the friction

factor f, where,

22

mu

Ddxdpf

Fanning friction factor fa:

4ffa

Laminar flow in a circular tube:

D

a

D Ref;

Ref

1664

Turbulent flow in a smooth circular tube:

621053000641790

DD Re,.Reln.f

621053000283581

DDa Re,.Reln.f


(9)

Turbulent flow in a roughened circular tube (e: surface roughness):

equationColebrook,

fRe

.

.

Delog.

f D

512

7302

110

aDa fRe

.

.

Delog.

f

2551

7304

110

equationHaaland,.

De

Re

.log.

f

.

D

111

1073

9681

1

111

1073

9663

1.

Da.

De

Re

.log.

f


(10)

Pressure drop for fully developed flow from x=0 to x=L

210 ppLxpxpp

2

2

mu

D

Lfp

24

2

ma

u

D

Lfp

Pumping power requirement P

p

pmP

p : Pump total efficiency

Nusselt number for fully developed flow condition

Laminar flow in a circular tube

- Uniform surface heat flux sq : 364.k

hDNuD

- Uniform surface temperature sT : 663.k

hDNuD

Turbulent flow in a circular tube

200050105103

127121

10002

187121

10008

63

3221

3221

Pr.,Re

Prf.

PrRefNu

ncorrelatioGnielinski,Prf.

PrRefNu

D

a

DaD

DD


(11)

Laminar flow in a noncircular tube

w

Ch P

AD

4


(12)

Laminar flow in annulus region with insulated outer surface 0oq

ioo,h

o,ho

omi,soi DDD;k

DhNu;TThq

- Uniform surface heat flux iq

20893615237821

499382208172

oioi

oio

DD.DD.

DD..Nu

- Uniform surface temperature i,sT

2859854133811

517355590571

oioi

oio

DD.DD.

DD..Nu

Turbulent flow in a noncircular tube and in an annulus region

w

Ch

P

AD

4 ,

h

Ce

P

AD

4 ,

hmD

DuRe

h ,

k

hDNu e

De

wP : Wetted perimeter; hP : Heat transfer perimeter

200050105103

127121

10002

187121

10008

63

3221

3221

Pr.,Re

Prf.

PrRefNu

ncorrelatioGnielinski,Prf.

PrRefNu

h

h

h

D

a

Da

eD

D

eD


(13)

Nusselt number for developing flows (Entry region effects)

Laminar flow in a circular tube

- Thermal Entry region

3204001

06680663

PrReLD..

PrReLD..Nu

D

DD

- Combined Entry region

14031

861

.

s

bDD

L

DPrRe.Nu


(14)

Problems


(15)

Classifications of Heat Exchangers


(16)

Classifications according to transfer processes


(17)

Classifications according to construction features


(18)


(19)


(20)


(21)


(22)


(23)

Classifications according to surface compactness


(24)

Classifications according to number of fluids

Heat exchangers with as many as 12 fluids streams have been used in

some chemical processes applications.

The design theory of three- and multi fluid heat exchangers is

algebraically very complex.

The present notes covers only design theory of two-fluid heat exchangers.

Classifications according to heat transfer mechanism

Classifications according to flow arrangements


(25)


(26)


(27)

Basic Equations for Heat Exchangers

Heat Exchangers: Devices used to exchange thermal energy between at least two fluids.

They encompass a wide range of flow configurations.

Classifications of Heat exchangers based on flow arrangement


(28)


(29)

The overall heat transfer coefficient

Bare circular clean tubes of Nt number of tubes

ii

W

ooiioo AhR

AhAUAUUA

11111

LNdA;LNdA tiitoo

LNk

ddlnR

ttube

ioW

2

ii

o

tube

ioo

oo hd

d

k

ddlnd

hU

1

2

11

itube

ioi

oo

i

i hk

ddlnd

hd

d

U

1

2

11


(30)

Bare circular fouled tubes of Nt number of tubes

iii

fi

W

o

fo

ooiioo AhA

RR

A

R

AhAUAUUA

11111

cetanresisunitfoulingoutside:R fo

cetanresisunitfoulinginside:R fi

ii

ofi

i

o

tube

ioofo

oo hd

dR

d

d

k

ddlndR

hU

1

2

11

i

fi

tube

ioifo

o

i

oo

i

i hR

k

ddlndR

d

d

hd

d

U

1

2

11

Finned circular fouled tubes of Nt number of tubes (rectangular fins)

iiiii

fi

W

oo

fo

oooiioo AhA

RR

A

R

AhAUAUUA

11111

i

fi

fii

o

fo

fooA

A;

A

A 1111


(31)

LNNHdA;LNNHdA tfifiiitfofooo 22

LNNHA;LNNHA tfiififitfoofofo 22

LNNdA;LNNdA tifiiuitofoouo

fiuiifouoo AAA;AAA

efficiencysurfaceinner,outer:, io

surfaceinner,outeronfinsof.no:N,N fifo

lengthfininner,outer:H,H fifo

thicknessfininner,outer:, fifo

tyconductivithermalfininner,outer:k,k i,fino,fin

efficiencythermalfininner,outer:, fifo

ii,fin

iifi

ii,fin

iifi

fi

oo,fin

oofo

oo,fin

oofo

fo

k

hH

k

hHtanh

;

k

hH

k

hHtanh

2

2

22

22

22


(32)

iii

o

i

fi

i

o

ttube

ioo

o

fo

ooo hA

AR

A

A

LNk

ddlnAR

hU

1

2

11

iii

fi

ttube

ioi

o

fo

o

i

ooo

i

i h

R

LNk

ddlnAR

A

A

hA

A

U

1

2

11

Log Mean Temperature Difference (LMTD) method of analysis

Assumptions: a) Fully developed conditions, b) Constant cross-sectional areas, c) Constant

properties


(33)

1) Counter Flow Heat Exchangers

21

12

21122112

ch

ch

chchoohhphhccpcc

TT

TTln

TTTTAUTTcmTTcmQ

phhpcc

oo

ch

ch

cmcmAUexp

TT

TT

11

21

12

2) Parallel Flow Heat Exchangers

11

22

11222112

ch

ch

chchoohhphhccpcc

TT

TTln

TTTTAUTTcmTTcmQ

phhpcc

oo

ch

ch

cmcmAUexp

TT

TT

11

11

22

3) Other Types Heat Exchangers

lmoohhphhccpcc TAUTTcmTTcmQ 2112

21

12

2112

ch

ch

chchcf,lmlm

TT

TTln

TTTTFTFT

1Ffactorcorrection:F

tarrangemenflow,

TT

TTP,

TT

TT

cm

cmRfF

ch

cc

cc

hh

h,ph

c,pc

11

12

12

21

efficiencyeTemperatur:P

ratiocapacityHeat:R


(34)


(35)


(36)


(37)

The-NTU method of heat exchanger analysis

Assumptions: a) Fully developed conditions, b) Constant cross-sectional areas, c)

Constant properties

Definitions and relationships

maxmax

minr

min

oo

Q

Q;

C

CC;

C

AUNTU

UnitsTransferofNumber:NTU

hcmin C,CofMinimumC

hcmax C,CofMaximumC

phhhpccc cmC;cmC

flowfluidhot,coldofcapacitythermal:C,C hc

esseffectivenexchangerheat:

11 chminmax TTCQ


(38)

2112 hhhccc TTCTTCQ

11

21

11

12

11 chmin

hhh

chmin

ccc

chmin TTC

TTC

TTC

TTC

TTC

Q

arragementflow,C,NTUf r

arragementflow,C,gNTU r

1

1

1

1

1

NTUnNTU

nNTUNTU

;n

Q

NTU

NTU

Cr

6

5

4

3

2

1

1

1


(39)

Heat exchanger with condensing fluid

2

1

2112

ch

ch

chchoofghccpcc

TT

TTln

TTTTAUhmTTcmQ

001 rh C;C;.F

Heat exchanger with evaporating fluid

ch

ch

chchoohhphhfgc

TT

TTln

TTTTAUTTcmhmQ

1

2

1221

oo

r

AU

NTU

NTU

E

F

C

8

7

6

5

4

3

2

1

1

1

1

1

1

1

1

NTUnNTU

nNTUNTU

;n


(40)

001 rc C;C;.F

Basic heat exchanger equation under variable Uo-coefficient

21

12

2112

ch

ch

chchom,o

TT

TTln

TTTTFAUQ

oA

o

o

m,o dAUA

U1

Linear variation of Uo-coefficient with A

2

21 ,o,o

m,o

UUU

Small variation of Uo-coefficient with A

2oom,o AAUU


(41)

Heat exchanger rating calculations

Known quantities: Inlet temperatures, mass flow rates, outer surface area

1) Calculate the overall heat transfer coefficient Uo.

2) Calculate thermal capacities Cc and Ch.

3) Calculate Cmin, Cmax and Cr.

4) Calculate NTU.

5) Calculate the heat exchanger effectiveness using -NTU relationships.

6) Determine the heat transfer rate Q.

Heat exchanger design calculations using LMTD method

Known quantities: Inlet temperatures, mass flow rates, one outlet temperature

1) Calculate the heat transfer rate Q.

2) Calculate the unknown outlet temperature using the conservation of energy

principle.

3) Calculate cf,lmT .

4) Calculate the correction factor F.


6) Determine Ao.

Heat exchanger design calculations using -NTU method

Known quantities: Inlet temperatures, mass flow rates, one outlet temperature

1) Calculate the heat transfer rate Q.

2) Calculate the unknown outlet temperature using the conservation of energy principle.


4) Calculate thermal capacities Cc and Ch.

5) Calculate Cmin, Cmax and Cr.

6) Calculate the heat exchanger effectiveness .

7) Calculate the NTU using NTU- relationships.

8) Determine Ao.


(42)

Heat Exchanger Design Methodology


(43)

Problems


(44)


(45)

Fouling of Heat Exchangers

Fouling: It is accumulation of undesirable substances on a surface.

Examples on fouling

Deposit of cholesterol on the inner surface of the artery wall. This deposit results in

narrowing the blood flow cross-sectional area. Thus more pumping power is required

by the heart to circulate the blood.

Deposit of ashes on the tube surface which forms of a less thermally conducting layer

on the surface. Thus, heat transfer rate is reduced.

Disadvantages of fouling on heat exchanger operation

Reducing heat transfer rate due to the increased thermal resistance of the deposit

layer.

Increasing pumping power requirement due to narrowing of the flow passages.

Definitions

conditionfouledonbasedtcoefficientransferheatOverall:Uof

conditioncleanonbasedtcoefficientransferheatOverall:Uoc

ft

ocof

RUU

11


(46)

factorfoulingTotal:R ft

tubecircularbare,d

dRRR

i

ofifoft

tubefinned,A

ARRR

i

o

i

fi

o

fo

ft

Design of heat exchangers based clean condition

lmococ TAUQ

Design of heat exchangers based fouled condition

lmofof TAUQ

Relationship between clean and fouled conditions based designs

lmococlmofof TAUTAU

ococofof AUAU

011 .RUU

U

A

Aftoc

of

oc

oc

of

Heat exchanger operation under fouled condition

A. Effect of fouling on pressure drop and pumping power

22

22f,m

if

ff

c,m

ic

cc

u

d

Lfp;

u

d

Lfp

fc ff

22

44iff,micc,mfc dudumm


(47)

2

if

ic

c,m

f,m

d

d

u

u

01

5

.d

d

p

p

P

P

if

ic

c

f

c

f

B. Effect of fouling on flow rate

L

pdm;

L

pdm

f

iffc

icc

44

128128

fc pp

4

ic

if

c

f

d

d

m

m

C. Effect of fouling on heat transfer

f,lmooff TAUQ

c,lmoocc TAUQ

c,lm

f,lm

ftocc,lm

f,lm

oc

of

c

f

T

T

RUT

T

U

U

Q

Q

11

Cost of fouling on industrial sector

Increased capital cost (large pumps are required, large heat transfer area is required,

stand by heat exchangers are required).

Increased maintenance cost (to clean the deposits).

Loss of production cost (shutting down the plant for cleaning the heat exchangers).

Energy losses (Large electrical energy is required to operate the large pumps).

053251

251

5..

P

P

.d

d

c

f

if

ic

4096080

251

4..

m

m

.d

d

c

f

if

ic


(48)

Categories of fouling

A. Particulate fouling: e.g. dust, ash.

B. Crystallization fouling: This arises due to presence of salts in the fluid.

C. Corrosion fouling: corrosive fluids may react with tube material producing corrosion

products.

D. Biofouling: e.g. bacteria, algae and molds.

E. Chemical reaction fouling: The tube surface can act as catalyst expediting

chemical reaction on the surface. e.g. polymerization.

Fundamental process of fouling

1. Initiation of the surface: e.g. removing of any protective layers.

2. Transport of undesirable substances: Mechanisms of this process are: a) mass

diffusion, b) Inertial impaction, c) Sedimentation, d) Thermophoresis (motion of

particles due to temperature gradients), d) Electrophoresis (motion of particles due to

electrical potential gradients).

3. Attachment: depends on adhering forces between the deposits and the surface.

4. Removal: removal mechanisms are: a) Dissolution (removal by ions), b) Erosion

(removal by small fouling masses), c) Spalling (removal by large fouling masses).

5. Aging: growing effect.

(Fouling is an extremely complex phenomenon)

Cleanliness Factor (CF) of the heat exchanger:

ftococ

of

RUU

UCF

1

1

designsTypical,.CF 850

Percent Over Surface (OS) of the heat exchanger:

%RU%A

AOS ftoc

oc

of1001100

designsTypical%,.OS 617


(49)

TEMA values of fouling factors to be used in heat exchangers design

TEMA: Tubular Exchangers Manufacturers Association

Cleaning of heat exchanger can be started once the fouling factors reach TEMA

values shown in the next tables.


(50)


(51)

Prediction of fouling

Techniques to control fouling


(52)

Problems


(53)


(54)

Double Pipe Heat Exchangers (DPHXs)

DPHX consists of one set of pipes of smaller diameter (tubes) placed concentrically inside a

pipe of larger diameter (pipe or shell) with appropriate fittings to direct the flow

from one section to the next.

Annulus is the volume between the outer surface of the tubes and the inner surface of the

pipe.

DPHXs are used majorly for sensible cooling/heating processes.

DPHXs exist in industry in form of hairpin heat exchangers as shown in the figures below.

DPHXs can be arranged in series and parallel arrangements or combined arrangements.


(55)

Advantages of DPHXs: a) easy to be cleaned, b) easy to be maintained, c) fluids can be

gases or liquids, d) tubes may have fins on their inner surface,

outer surface or on both inner and outer surfaces (fins are

usually placed on the surface that has minimum convection heat

transfer coefficients).


(56)

Disadvantages of DPHXs: a) bulky, b) have small heat transfer area per unit volume, c)

expensive per heat transfer area.

DPHX cross sections; See next figures for various cross-sections

Pressure drop in single hairpin heat exchanger

o,Co

oo,m

i,Ci

ii,m

A

mu;

A

mu

2

24

2

i,m

ioutletinletS,bend

i

i,ai

uKKK

d

Lfp

2

24

2

o,m

ooutletinletS,bend

o,h

o,ao

uKKK

D

Lfp

025001 .K;.K;.K S,bendoutletinlet


(57)

Pressure drop in NHP hairpin heat exchangers arranged in series

o,Co

oo,m

i,Ci

ii,m

A

mu;

A

mu

22

24

2

i,m

iioutletiinletHP

L,ibend

S,ibend

i

i,ai

uKKN

KK

d

Lfp

22

24

2

o,m

oooutletoinletHP

L,obend

S,obend

o,h

o,ao

uKKN

KK

D

Lfp

51025001 .K;.K;.K;.K L,o,ibendS,o,ibendo,ioutleto,iinlet

Thermal/hydraulic analysis of tubes with NHP=1

i,h

i,C

i,e

i,w

i,C

i,hP

AD;

P

AD

44

i

i,ei

D

i

i,hi,mi

Dk

DhNu;

DuRe

i,ei,h

Bare circular tubes of number Nt

tii,C NdA 2

4

tii,hi,w NdPP

LNdA tii 2

Finned circular tubes of number Nt (rectangular fins)

ti,fii,ftii,C NNHNdA

2

4

ti,fi,ftii,hi,w NNHNdPP 2

11

HPii NAA


(58)

ii,fi,fti,fii,fiti,u HLNNA;NdLNA 22211

i,fi,fiti HNdLNA 221

Convection heat transfer coefficient inside tubes (hi)

Developing Laminar flow in a circular tube


i

ii,mi

d

duRe

i

320401

06680663

idi

idi

i

iid

PrReLd.

PrReLd..

k

dhNu

i

i

i


14031

861

.

i,s

i,biid

i

iid

L

dPrRe.

k

dhNu i

i

Turbulent fully developed flow

200050105103

127121

10002

63

3221

iD

ii,a

iDi,a

D

Pr.,Re

Prf.

PrRefNu

i,h

i,h

i,e

tubesmooth;Re,.Reln.fi,hi,h DDi,a

621053000283581

tuberoughened;.

De

Re.

log.f

.

Di,a i,h

111

10 7396

631

Thermal/hydraulic analysis of annulus with NHP=1

o,h

o,C

o,e

o,w

o,C

o,hP

AD;

P

AD

44


(59)

o

o,eo

D

o

o,ho,mo

Dk

DhNu;

DuRe

o,eo,h

Bare circular tubes of number Nt

toio,C NdDA 22

4

too,htoio,w NdP;NdDP

LNdA too 21

Finned circular tubes of number Nt (rectangular fins)

to,foo,ftoio,C NNHNdDA

22

4

to,fo,ftoo,hto,fo,ftoio,w NNHNdP;NNHNdDP 22

oo,fo,fto,foo,foto,u HLNNA;NdLNA 22211

o,fo,foto HNdLNA 221

Convection heat transfer coefficient outside tubes (ho)

Developing Laminar flow


32

0401

06680663

oDo,h

oDo,h

o

o,eoD

PrReLD.

PrReLD..

k

DhNu

o,h

o,h

o,e


14031

861

.

o,s

o,bo,hoD

o

o,eoD

L

DPrRe.

k

DhNu o,h

o,e

11

HPoo NAA


(60)


200050105103

127121

10002

63

3221

oD

oo,a

oDo,a

D

Pr.,Re

Prf.

PrRefNu

o,h

o,h

o,e

tubesmooth;Re,.Reln.fo,ho,h DDo,a

621053000283581

111

1073

9663

1.

Do,a.

De

Re

.log.

fo,h

Thermal analysis of NHP hairpin series heat exchangers

21

12

21122112

ch

ch

chchoohhphhccpcc

TT

TTln

TTTTAUTTcmTTcmQ

111

HPooHPoo NAA;NAA

Bare circular tubes

ii

ofi

i

o

tube

ioofo

oof hd

dR

d

d

k

ddlndR

hU

1

2

11

Circular tubes with fins on their outer surfaces

ii

o

fi

i

o

ttube

ioo

o

fo

ooof hA

AR

A

A

LNk

ddlnAR

hU

1

22

11

1

1

1

11

Circular tubes with fins on their inner surfaces

iii

o

i

fi

i

o

ttube

ioo

fo

oof hA

AR

A

A

LNk

ddlnAR

hU

1

22

11

1

1

1

11


(61)

Finned circular tubes with fins on inner and outer surfaces

iii

o

i

fi

i

o

ttube

ioo

o

fo

ooof hA

AR

A

A

LNk

ddlnAR

hU

1

22

11

1

1

1

11

1

1

1

1 1111i

fi

fii

o

fo

fooA

A;

A

A

ii,fin

iifi

ii,fin

iifi

fi

oo,fin

oofo

oo,fin

oofo

fo

k

hH

k

hHtanh

;

k

hH

k

hHtanh

2

2

22

22

22

Design and operational features

DPHXs have four key design components:

1. Shell nozzles

2. Tube nozzles

3. Shell-to-tube closure

4. Return-bend housing


(62)

Problems


(63)


(64)


(65)

Shell-and-Tube Heat Exchangers (STHEs)

STHE basic components:

Tube bundle Shell Front end head Rear end head Baffles Tube sheets


(66)

STHEs have larger heat transfer area per unit volume than double pipe heat exchangers.

STHEs can be used for large pressure applications.

STHEs are easier to be cleaned than many types of heat exchangers such as compact heat

exchangers.

TEMA Shell and end heads types


(67)

ShellEShellF pp 8

ShellEShellG pp

ShellEShellJ pp 8

1

E-Shell: least expensive shell

E-Shell and 1P tubes: counter flow heat exchanger

F-Shell and 2P tubes: counter flow heat exchanger

J-Shell: can be used as shell side horizontal condenser

G-Shell, J-Shell and X-Shell: cross flow heat exchanger

Tubes covers 60% of the shell diameter of K-Shell

Tube bundle

They should accommodate thermal expansion.

They should furnish ease of cleaning.

They should provide least expensive construction.


(68)


(69)

Tubes and tube passes

Tube material: low carbon steel, low alloy steel, stainless steel, copper, admiralty,

cupronickel, inconel, alumnium (in form of alloys) or titanum.

Tube standards


(70)


(71)

Tube layouts

PT: tube pitch; C: clearence; oT dPC

mmC 7 : cleaning requirment for square pitch


(72)

Traingular pitch results in largest tube density inside the shell

51251 .d

P.

o

T : to avoid weak structurally tube sheets while enhancing heat

transfer

layouttube,P,N,d,DfN TPoimaxt

maxtN : Maximum tubes to be fit inside the shell – (Tube count)


(73)


(74)


(75)


(76)

Allocations of streams

The more seriously fouling fluid flows through the tubes.

The high pressure fluid flows through the tubes.

The corrosive fluid must flow through the tubes.

Fluids producing lower heat transfer coefficients flows on the shell side.

Baffle types and geometry

Baffles function

Supporting the tubes to prevent vibrations.

Diverting shell side fluid flow across the bundle to obtain higher convection heat

transfer coefficients.

Baffle types

Transverse baffles (Plate baffles, Rod baffles)

Longitudinal baffles.


(77)


(78)

Baffle window (single/others segmental):


(79)

B : Baffle spacing WA : area of baffle window

Single segmental baffle %%D

HcutBaffle%

b

W 3010025

Other segmental baffle %D

AAcutBaffle

b

WW 10042

21

For strong structurally tube

sheets, less vibration, and

augmented heat transfer

6040 .D

B.

i

Basic design procedure of a heat exchanger


(80)


(81)

Shell side pressure drop

140

2

2

1.

w,ooo,eo

ibooo

D

DNGfp

Square pitch:

o

oTo,e

d

dPD

44 22

Triangular pitch:

2

8434 22

o

oTo,e

d

dPD

T

oio,C

P

dBDA 1

o,C

oo

A

mG

1B

LNb ; bN : no. of baffles

oo Reln..expf 1905760

o

oo,e

o

GDRe

6101400 oRe

Tube side pressure drop (bare circular tubes)

P

tii,C

i,Ci

ii,m

N

NdA;

A

mu 2

4

; PN : no. of tube passes

P

i,m

i

i

i,ai Nu

d

Lfp

244

2

Convection heat transfer coefficient inside tubes (hi)

Developing Laminar flow in a circular tube


i

ii,mi

d

duRe

i

320401

06680663

idi

idi

i

iid

PrReLd.

PrReLd..

k

dhNu

i

i

i


(82)


14031

861

.

w,i

b,iiiid

i

ii

idL

dPrRe.

k

dhNu


200050105103

127121

10002

63

3221

iD

ii,a

iDi,a

D

Pr.,Re

Prf.

PrRefNu

i,h

i,h

i,e

tubesmooth;Re,.Reln.fi,hDi,hDi,a

621053000283581

tuberoughened,.

De

Re.

log.f

.

Di,a i,h

111

10 7396

631

Shell side heat transfer coefficient using Kern method

63

140

31

550

101102

360

o

oo,e

o

.

w,o

oo

.

o

oo,e

o

o,eo

GDRe

PrGD

.k

Dh

Square pitch:

o

oTo,e

d

dPD

44 22

Triangular pitch:

2

8434 22

o

oTo,e

d

dPD

T

oio,C

P

dBDA 1

o,C

oo

A

mG

%cutBaffle 25


(83)

Thermal analysis of shell-and-tube heat exchangers

21

12

21122112

ch

ch

chchoohhphhccpcc

TT

TTln

TTTTFAUTTcmTTcmQ

Bare circular tubes

ii

ofi

i

o

tube

ioofo

oo hd

dR

d

d

k

ddlndR

hU

1

2

11

LNdA too

Finned circular tubes with fins on inner and outer surfaces

(rectangular fins)

iimLi

mLo

i

fi

mLi

mLo

ttube

iomLo

o

fo

ooo hA

AR

A

A

Nk

ddlnAR

hU

1

2

11

1

1

1

11

mLi

mLfi

fii

mLo

mLfo

fooA

A;

A

A

1

1

1

11111

ii,fin

iifi

ii,fin

iifi

fi

oo,fin

oofo

oo,fin

oofo

fo

k

hH

k

hHtanh

;

k

hH

k

hHtanh

2

2

22

22

22

ii,fi,fti,fii,fiti,u HLNNA;NdLNA 2

i,fi,fiti HNdLNA 2

oo,fo,fto,foo,foto,u HLNNA;NdLNA 2

o,fo,foto HNdLNA 2


(84)

Problems


(85)


(86)

Compact Heat Exchangers (Compact HXs)

Compact HXs is used for gas flow applications.

Compact HX types are: a) Plate-fin type, and b) Tube-fin type.

Compact HX has heat transfer surface area per unit volume larger than 700 m2/m

3.

32700 mmVAt

Microheat exchanger has heat transfer surface area per unit volume larger than

10000 m2/m

3.

3200010 mm,VAt

Compact HXs are used for gas to gas or liquid to gas heat exchangers.


(87)

Compact HXs applications: a) air conditioning condensers and evaporators, b)

automotive radiators, c) intercoolers of compressors.

Heat transfer enhancement techniques

Active techniques (requiring external power to enhance heat transfer), e.g. surface

vibration, acoustic techniques, using electrohydrodynamic or hydromagetic effects.

Passive techniques (do not require external power to enhance heat transfer), e.g.

using fins, twist tapes to gererate turbulence, surface roughness.

Heat transfer in compact heat exchangers

o,po

oo

cG

hSt:StnumbertontanS

o

o,ho

o

o

o,po

o

DGRe;

k

cPr

omin,

oomax,oo

A

mUG


(88)

areaflowexternalimummin:A omin,

areatransferheatexternal:AA ot

areafrontalflowexternal:A o,fr

HWA o,fr

t

minho

A

ALD 4

3232

o

o,po

oooo,H Pr

cG

hPrStj

Important relationships

HWLV;A

A;

V

A

o,fr

omin,t

4 o,h

omin,

t D;LA

A

o,,o

o,fr

o

omin,

oo

U

A

m

A

mG 1

o

o,,o

o,fro

o

o

o,ho

o,h

U

A

mDGRe 144

Overall heat transfer coefficient

cmthicknessfin:o

1cmpitchfin:bo

t

o,f

oA

A


(89)

ofoo 11

thicknesstube:tube

perimeterinnertube:Pi

perimeteroutertube:Po

Circular tube with fins on outer surface and no-fins on inner surface

ootube

io

o

ooo

io

oo

i

o

o hk

ddlnbd

h

b

d

d

U

1

21

11

1

11

Rectangular tube with fins only on outer surface

ootube

tube

o

oo

i

o

io

oo

i

o

o hk

b

P

P

h

b

P

P

U

1

1

11

1

11


(90)

Thermal analysis of compact heat exchangers

21

12

21122112

ch

ch

chchtohhphhccpcc

TT

TTln

TTTTFAUTTcmTTcmQ

Gas-side pressure drop for compact heat exchangers

21

11

2

11

,o,oo

Tube-fin compact heat exchangers

11

2 2

121

1

2

,o

,o

o

,o

omin,

to

,o

oo

A

Af

Gp

Plate-fin compact heat exchangers

2

121

2

12

1

2

11212 ,o

,o

e

o

,o

omin,

to

o

oc

,o

oo k

A

Afk

Gp

tcoefficienlossncontractioinlet:kc

tcoefficienlossansionexpoulet:ke

Gas-side convection coefficient, ho, using different charts

The h-coefficient shown in the subsequent figures is the ho-coefficient.

The f-factor shown in the subsequent figures is the fo-factor.


(91)

rowsof.no:NL

rowpertubesof.noimummax:NT

inch.PL 8660

inch.PT 001

LLL PPNL 1

TTT PPNH 1


(92)

rowsof.no:NL


inch.Df 8610

inch.PL 9000

inch.PT 9750

fLL DPNL 1

fTT DPNH 1


(93)

rowsof.no:NL


inch.D f 1211

inch.PL 351

inch.P;inch.P B,TA,T 84812321

fLL DPNL 1

fTT DPNH 1


(94)


(95)


(96)


(97)


(98)

Problems


(99)


(100)

Boiling

Boiling is associated with transformation of liquid to vapor at a

solid/liquid interface due to convection heat transfer from the solid.

Agitation of the liquid by vapor bubbles is one mechanism that provides

for large convection coefficients and hence large heat fluxes at low-to-

moderate surface-to-fluid temperature differences.

Special form of Newton’s law of cooling:

esatss ThTThq

liquidofetemperatursaturation:Tsat

etemperaturexcess:TTT satse

Special cases

Pool boiling

Liquid motion is due to natural convection and bubble-induced mixing.

Forced convection boiling

Fluid motion is induced by external means, as well as by bubble-induced

mixing.

Saturated boiling

Liquid temperature is slightly higher than saturation temperature.

Subcooled boiling

Liquid temperature is less than saturation temperature.


(101)

The boiling curve

Reveals range of conditions associated with saturated pool boiling on es Tq

plot.

Free convection boiling atm@water,CTe 15

- Little vapor formation

- Liquid motion is due principally to single-phase natural convection.

Onset of Nucleate Boiling - ONB atm@water,CTe 15

Nucleate boiling atm@water,CTC e 1305

Isolated vapor bubbles atm@water,CTC e 1105

- Liquid motion is strongly influenced by nucleation of bubbles at the surface.

- h and sq increase sharply with increasing eT .

- Heat transfer is principally due to contact of liquid with the surface (single-phase

convection) and not to vaporization.


(102)

Jets and columns atm@water,CTC e 13010

- Increasing number of nucleation sites causes bubble interactions and coalescence

into jets and slugs.

- Liquid/surface contact is impaired.

- sq continues to increase with eT but h begins to decrease.


(103)

Critical Heat Flux - CHF, atm@water,CTq emax 130

- Maximum attainable heat flux in nucleate boiling.

- 21 mMWqmax for water at atmospheric pressure.

Potential burnout for power - Controlled heating - An increase in beyond maxq causes the surface to be blanketed by vapor, and the

surface temperature can spontaneously achieve a value that potentially exceeds its

melting point for water at atmospheric pressure CTe

1000 .

- If the surface survives the temperature shock, conditions are characterized by film

boiling.

Film boiling atm@water,CTe 1120

- Heat transfer is by conduction and radiation across the vapor blanket.

- A reduction in sq follows the cooling curve continuously to the Leidenfrost point

corresponding to the minimum heat flux minq for film boiling.

- A reduction in sq below minq causes an abrupt reduction in surface temperature

to the nucleate boiling regime.

Transition boiling for temperature - Controlled heating - Characterized by a continuous decay of sq (from maxq to minq ) with increasing

eT .

- Surface conditions oscillate between nucleate and film boiling, but portion of

surface experiencing film boiling increases with eT .

- Also termed unstable or partial film boiling.


(104)

Pool boiling correlations

Nucleate boiling

- Rohsenow Correlation

fgsvn

lfgf,s

eL,p

f

vLfgLs hAm

PrhC

Tcghq

321

L Dynamic visocity of liquid phase

fgh Enyhalpy of vaporization

g Gravitional acceleration

L Density of liquid phase

v Density of vapor phase

f Surface tension

L,pc Specific heat of liquid phase

LPr Prandtl number of liquid phase

n,C f,s Surface/fluid combination

- Critical heat flux of Zuber

fgsmaxv

v

vLf

vfgmaxs hAmg

Chq

41

2


(105)

1310.C Horizontal cylinder or sphere

1490.C Large horizontal plate

Film boiling

fgvsatsss hmTTAhq

313434 hhhh radconv

radconvradconv hhifhhh 4

3

413

satsvv

fgvL

v

conv

TTk

DhgC

k

DhNu

620.C Horizontal cylinder DLAs

670.C Sphere 2DAs

satsv,pfgfg TTc.hh 80

sats

satsrad

TT

TTh

44

ttanconsBoltzmannStefan:KmW. 42810675

Correlation for boiling inside tubes

fg

sAC

hm

Dxq

m

uXdAX C

X : Mean vapor mass fraction at a given section

f,s

.

.

fgC

s..

.

v

l

sp

GXhAm

qFrfXX.

h

h 80

70

640160

10

1 11058166830


(106)

f,s

.

.

fgC

s..

.

v

l

sp

GXhAm

q.FrfXX.

h

h 80

70

080720

450

2 1266711361

gDAmFr lC

2

01.Frf Vertical tubes

01.Frf Horizontal tubes with 040.Fr

30632 .Fr.Frf Horizontal tubes with 040.Fr

40

80

0230 .

i

.

l

l,ml

i

spPr

Du.

k

Dh

21 h,hMAXh

800 .X


(107)


(108)

Problems

P1 A long, 1-mm-diameter wire passes an electrical current dissipating 3150

W/m and reaches a surface temperature of 126 C when submerged in water

at 1 atm. What is the boiling heat transfer coefficient? Estimate the value of

the correlation coefficient Cs,f.

P2 A nickel-coated heater element with a thickness of 15 mm and a thermal

conductivity of 50 W/mK is exposed to saturated water at atmospheric

pressure. A thermocouple is attached to the back surface, which is well

insulated. Measurements at a particular operating condition yield an

electrical power dissipation in the heater element of 6.950x107 W/m

3 and a

temperature of To= 266.4 C.

(a) From the foregoing data, calculate the surface temperature, Ts, and the

heat flux at the exposed surface.

(b) Using the surface heat flux determined in part (a), estimate the surface

temperature by applying an appropriate boiling correlation.

P3 A 1-mm-diameter horizontal platinum wire of emissivity 0.25 is operated in

saturated water at 1-atm pressure. What is the surface heat flux if the surface

temperature is 800 K?

P4 Copper tubes 25 mm in diameter and 0.75 m long are used to boil saturated

water at 1 atm. (a) If the tubes are operated at 75% of the critical heat flux,

how many tubes are needed to provide a vapor production rate of 750 kg/h?

What is the corresponding tube surface temperature?

P5 A vertical steel tube carries water at a pressure of 10 bars. Saturated liquid

water is pumped into the =0.1-m diameter tube at its bottom end (x =0) with

a mean velocity of um=0.05 m/s. The tube is exposed to combusting

pulverized coal, providing a uniform heat flux of qs=100,000 W/m2.

(a) Determine the tube wall temperature and the quality of the flowing water

at x =15 m. Assume Gs, f=1.

(b) Determine the tube wall temperature at a location beyond x=15 m where

single-phase flow of the vapor exists at a mean temperature of Tsat.

Assume the vapor at this location is also at a pressure of 10 bars.


(109)

Condensation

Heat transfer to a surface occurs by condensation when the surface temperature

is less than the saturation temperature of an adjoining vapor.

Film condensation

Entire surface is covered by the condensate, which flows continuously

from the surface and provides a resistance to heat transfer between the

vapor and the surface.

Thermal resistance is reduced through use of short vertical surfaces and

horizontal cylinders.

Characteristic of clean, uncontaminated surfaces.

Dropwise condensation

Surface is covered by drops ranging from a few micrometers to

agglomerations visible to the ordinary eyes.

Thermal resistance is greatly reduced due to absence of a continuous film.

Surface coatings may be applied to inhibit wetting and stimulate dropwise

condensation.


(110)


(111)

General correlations for film condensation on vertical surface

fgLssatsL hmTTAhq

ssatl,pfgfg TTc.hh 680

312 gh

TTLkP

lfgl

ssatl

Wavy free laminar region 815.P

41

312

9430 P.k

ghNu

l

lLL


(112)

Wavy laminar region 2530815 P.

820

312

8906801 .

l

lLL .P.

Pk

ghNu

Turbulent region 012530 .Pr;P l

3421

312

895302401

l

l

lLL PrP.

Pk

ghNu

l Dynamic visocity of liquid phase

l kinematic visocity of liquid phase

fgh Enyhalpy of vaporization

l,pc Specific heat of liquid phase

g Gravitional acceleration

L Surface length along g-direction

lk Thermal conductivity of liquid phase

Lm Condensate mass flow rate at bottom section

A Surface area

Lh Average convection heat transfer coefficent

Correlations for film condensation on radial systems

fglssatsN,D hmTTAhq


Single smooth sphere and single horizontal tube (N=1)

413

ssatll

fgvll

l

DD

TTk

DhgC

k

DhNu

7290.C Horizontal tube DLAs

8260.C sphere 2DAs


(113)

Vertical tier of N horizontal tubes

n

DN,D Nhh

41n Analytical value, assuming continuous film between smooth tubes

61n Experimental value, due to dripping between smooth tubes

061 n Experimental value, finned or ribbed tubes


(114)

Correlation for film condensation in horizontal tube

(a) Small vapor mass flow rate

00035,Du

Rev

v,mv

i,v



(115)

413

550

ssatll

fgvll

l

DD

TTk

Dhg.

k

DhNu

(b) Large vapor mass flow rate

890

4080 22210230

.

tt

.

i

.

l,D

l

DD

X

.PrRe.

k

DhNu

m

mX;

D

XmRe v

l

l,D

14

1050901

.

v

l

.

l

v

.

ttX

XX:parameterMartinelli

Correlation for dropwise steam condensation on copper surfaces

ssatsdc TTAhq

CTC,T,h satsatdc

10022204410451

CT,KmW,h satdc

100510255 2


(116)

Problems

P1 Saturated steam at 1 atm condenses on the outer surface of a vertical, 100-

mm-diameter pipe 1 m long, having a uniform surface temperature of 94 C.

Estimate the total condensation rate and the heat transfer rate to the pipe.

P2 The condenser of a steam power plant consists of a square (in-line) array of

625 tubes, each of 25-mm diameter. Consider conditions for which saturated

steam at 0.105 bars condenses on the outer surface of each tube, while a

tube wall temperature of 17 C is maintained by the flow of cooling water

through the tubes. What is the rate of heat transfer to the water per unit

length of the tube array? What is the corresponding condensation rate?

P3 The condenser of a steam power plant consists of AISI 302 stainless steel

tubes (ks=15 W/m.K), each of outer and inner diameters do=30 mm and

di=26 mm, respectively. Saturated steam at 0.135 bar condenses on the

outer surface of a tube, while water at a mean temperature of Tm=290 K is in

fully developed flow through the tube. For a water flow rate iside the tube of

0.25 kg/s, what is the outer surface temperature Ts,o of the tube and the rates

of heat transfer and steam condensation per unit tube length? As a first

estimate, you may evaluate the properties of the liquid film at the saturation

temperature. If one wishes to increase the transfer rates, what is the limiting

factor that should be addressed?

P4 Refrigerant R-22 with a mass flow rate of 8.75x10-3

kg/s is condensed inside

a 7-mm-diameter tube. Annular flow is observed. The saturation

temperature of the pressurized refrigerant is Tsat=45 C, and the wall

temperature is Ts=40 C. Vapor properties are v=77 kg/m3 and v=15x10

-6

N.s/m2. Determine the heat transfer coefficient and the heat transfer and

condensation rates per unit length at a quality of X=0.5.


(117)

The Gasketed-Plate Heat Exchangers

Gasketed Plate Heat Exchangers (G-P HXs) were introduced mainly for the

food industry because of their ease of cleaning.

G-P HXs design becomes well identified in 1960s with the development of

more effective plate geometries, assemblies, and improved gasket materials.

G-P HXs can be used as an alternative to tube-and-shell type heat

exchangers for low-and medium-pressure liquid-to-liquid heat trasfer

applications.

Unlike to common design of heat exchangers, manufactrures of G-P HXs

have developed their own computerized design procedures applicable to the

exchangers they market.

Main Components

G-P HX components are the plate (the heat transfer solid surface) and the

frame.

Elements of the frame: fixed plate, compression plate, pressing equipment,

and connecting ports.

Elements of the heat transfer soild surface: series of plates, parts for fluid

entry and exit in the four corners.


(118)

The Plate Pack

Packing the plates make the holes at the corners to form continuous tunnels

or mainfolds.

The plate pack is tightnend by means of either a mechanical or hydraulic

tightening device.

The passages formed between the plates and corner ports are arranged so

that the two heat transfer media can flow through alternate channels, always

in counter-current flow.

The warmer medium will give some of its thermal energy through the thin

plate wall to the colder medium on the other side.


(119)

The medium are led into similar hole-tunels as in the inlets at the other end

of the plate package and are then discharged from the heat exchanger.

Several hundereds of plates can be stacked in a single frame which are held

together by the bolts that hold the stack in compression.

The two sides of the plate heat exchanger are normaly of identical

hydrodynamic characteristics.

The plate is a sheet of metal precison-pressed into a corrugated pattern.

The largest single plate is the order of 4.3 m heigh x 1.1 m wide.

The heat transfer area for a single plate lies in the range 0.01-3.6 m2.

The plate thickness ranges between 0.5 and 1.2 mm.

The plates are spaced with nominal gaps of 2.5-5.0 mm.

The hydraulic diameters for the flow channels ranges between 4-10 mm.

The fluid should be equally distributed over the full width of the plate. This

requires the minmum length/width ratio of the order of 1.8.

Leakage from the channels between the plates to the surrounding

atmosphere is prevented by the gasketing around the exterior of the plate.

The number and size of the plates are determined by the flow rate, physical

properties of the fluids, pressure drop, and the temperature requirements.


(120)

The Plate Types

Wide types of corrugated plates are available including chevron and

washboard types.

The most used type is the chevron type.

In washboard type, turbulence is promoted by continuously altering flow

direction and fluid's velocity.


(121)

In chevron type, adjacent plates are assembled such that the flow in the

channels provides swirling motion to the fluids.

The Chevron angle () varies between the extremes of about 65 and 25.

The Chevron angle () determines the pressure drop and heat transfer

characteristics of the plate.

The chevron angle () is reversed on adjacent plates so that when plates are

clamped together, the corrugations provide numerous contact points.

Because of the many supporting contact points, the plates can be made from

very thin material, usually 0.6 mm.


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Main Advantages

Flexibility of design through a variety of plate sizes and passes

arrangements.

Easily accessible heat transfer area, allowing changes in configuration to

suit changes in processes requirements through changes in the number of

plates.

Efficient heat transfer; high heat transfer coefficients for both fluids because

of turbulence and a small hydraulic diameter.

Very compact (large heat transfer area/volume ratio yet 2500 m2 of surface

area is available in a single unit), and low in weight.

Only the plate edges are exposed to the atmosphere, no insulation is

required as heat losses are negligible.

Inter-mixing of the two fluids cannot occur under gasket failure.


(123)

Plate units exhibit low fouling characteristics due to high turbulence and

low residence time. The transition to turbulence occurs at low Reynolds

number of 10 to 400.

More than two fluids may be processed in a single unit.

Less expensive than tubular heat exchangers if the tubes are made from

stainless steel.

Performance limits

The gaskets impose restriction on operating temperatures (160C-250C).

The gaskets impose restriction on operating pressures (25-30 bar).

The gaskets impose restriction on nature of fluids that can be handled.

The upper size of the G-P HX is limited by the presses available to stamp

out the plates from the sheet metal.

G-P HXs with sizes larger than 1500 m2 are not normally available.

It is possible to have a maximum design pressure of up to 2.5 MPa;

normally, the design pressure is around 1.0 MPa.

Operating temperatures are limited by the availability of suitable gasket

materials.

G-P HXs are not suitable for air coolers or gas-to-gas applications.

Velocities lower than 0.1 m/s are not used in plate heat exchangers.

High viscous fluids are not preferred to be used in G-P HXs.

Specially designed G-P HXs are now available for duties involving

evaporation and condensation systems.

Passes and Flow Arrangements

The term "pass" refers to a group of channels in which the flow is in the

same direction.

Single pass arrangement can be of U- and Z-arrangement types.

The U-arrangement: all four ports in this arrangement will be on the fixed-

head plate. This permits disassembly of the heat exchanger for cleaning or

repair without disturbing any external piping. The flow distribution for this

arrangement is less uniform than the Z-arrangement.

The Multipass arrangement: consists of passes connected in series.

Arrangements abbreviation is written as:

pN of 1st fluid no. of channels for 1st fluid/ pN of 2nd fluid no. of channels for 2nd fluid


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Applications

G-P HXs are used in chemical, pharmaceutical, hygiene products,

biochemical processing, food, and dairy industries as they can meet health

and sanitation requirements.

G-P HXs are mainly used as liquid-to-liquid turbulent flow HXs.


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Fouling of G-P HXs

G-P HXs have less fouling than tubular heat exchangers because of the

following reasons:

High turbulence maintains solids in suspension.

Velocity profiles across a plate are uniform with almost absent zones of

low velocities.

The plate surfaces are generally smooth and can be further electro-

polished.

Deposit of corrosion products are absent because of low corrosion rates.

The high film coefficients maintain a moderately low metal wall

temperature. This helps preventing crystallization growth.

The plates can be easily cleaned.

Heat Transfer and Pressure Drop Calculations

Surface enlargement factor ,

lengthojectedPr

lengthDeveloped , 251151 .. .

pA

A

1

1 ;


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1A : Actual effective area provided by manufacturer;

pA1 : The projected plate area;

wpp LLA 1 ; pvp DLL ; phw DLL ;

vL : The vertical ports distance;

wL : The horizontal ports distance.

The mean channel spacing b :

tpb ;

p : The plate pitch; t : the plate thickness

The compressed plate pact length cL , t

c

N

Lp , tN : total number of

plates.

The hydraulic diameter of the channel hD :

b

Lb

bLD

w

wh

2

2

4

;

hcDG

Re ;

wcp

cbLN

mG

;

cpN : Number of channels per pass;

p

tcp

N

NN

2

1 , pN : number of passes

Convection HT Coefficients Correlation:

170

31

.

w

bnh

h PrReCk

hD

The frictional pressure drop cp :


(129)

1702

24

.

w

bc

h

pv

c

G

D

NLfp

,

m

p

Re

Kf

The port pressure drop pp :

241

2p

pp

GN.p ,

42p

pD

mG

.

The total pressure drop tp :

pct ppp .

The overall heat transfer coefficient:

whcc k

t

hhU

111;

fcfh

whcf

RRk

t

hhU

111

Required Heat Duty, eA is the total developed area for all thermally

effective plates 2tN :

2112 hhhpcccpr TTcmTTcmQ ;

cf,lmeff TFAUQ ;

2

1

21

T

Tln

TTT cf,lm

;

121 ch TTT ; 212 ch TTT

The safety factor of the design sC ,

r

f

sQ

QC


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(131)

Problems


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Condensers and Evaporators

Horizontal shell side condensers


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Horizontal tube side condensers

Vertical shell side condensers


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Vertical tube side condensers

1. Vertical in-tube down flow condenser


(137)

2. Reflux condenser


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Air cooled condensers

Direct contact condensers


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Failure of condenser operation

1. The tubes may be more fouled than expected – a problem not unique

to condensers.

2. The condensate may not be drained properly, causing tubes to be

flooded. This could mean that the condensate outlet is too small or too

high.

3. Venting of non-condensable gases may be inadequate.

4. The condenser was designed on the basis of end temperatures without

noticing that design duty would involve a temperature cross in the

middle of the range.

5. Flooding limits have been exceeded for condensers with backflow of

liquid against upward vapor flow.

6. Excessive fogging may be occurring. This can be problem when

condensing high molecular weight vapors in the presence of non-

condensable gases.

7. Severe maldistribution in parallel condensing paths is possible,

particularly with vacuum operation. This occurs because there can be

two flow rates which satisfy the imposed pressure drop.

Condensers for refrigeration and air conditioning

1. Water-cooled condensers

Horizontal shell-and-tube

Vertical shell-and-tube

Shell-and-coil

Double pipe

2. Air-cooled condensers

3. Evaporative condensers


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Evaporative condensers


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Evaporators for refrigeration and air conditioning

Water cooling evaporators (Chillers)


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Air cooling evaporators (Air coolers)


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Standards for evaporators and condensers

Air-conditioning and Refrigerating Institute (ARI) Standards

American Society for Heating, Refrigerating and Air-Conditioning

Engineers (ASHRAE) Standards


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Problems


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References

[1] F. P. Incorpera, D. P. DeWitt, T. L. Bergman, A. S. Lavine, “Fundamentals

of Heat and Mass Transfer-6th

Edition”, John Wiley, 2006, New York.

[2] F. P. Incorpera, D. P. DeWitt, T. L. Bergman, A. S. Lavine, “Fundamentals

of Heat and Mass Transfer-7th

Edition”, John Wiley, New, 2011, New York.

[3] R. K. Shah, D. P. Sekulic, Fundamental of Heat Exchanger Design, John

Wiley, 2003, New York.

[4] S. Kaka, H. Liu, Heat Exchangers: Selection, Rating, and Thermal Design,

CRC Press, 2002, Florida.

Documents

Basics of Rating and Thermal Design of Heat Exchangers · King Abdulaziz University Faculty of Engineering . Mechanical Engineering Department. Basics of Rating and Thermal Design