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HEAT EXCHANGER DESIGN

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Heat Transfer Equipment Types

Type Service

Double pipe exchanger Heating and cooling

Shell and tube exchanger All applications

Plate heat exchanger Heating and cooling

Plate-fin exchanger

Spiral heat exchanger

Air cooled Cooler and condensers

Direct contact Cooling and quenching

Agitated vessel Heating and cooling

Fired heaters Heating

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Double Pipe Heat Exchanger

Consists of two concentric pipes with one fluid flowing through the inner pipe while the other fluid flowing through the annular space

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Shell and Tube Heat Exchanger

Consists of tube bundles enclosed in a cylindrical shell with one fluid flowing through the tubes and the other flowing outside of the tubes

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LOGOHeat Transfer Equipment in

Industries

Exchanger: heat exchanged between two process streams

Heaters and coolers: where one stream is plant service

Vaporiser: if a process stream is vaporised

Reboiler: a vaporiser associated with distillation column

Evaporator: if concentrating a solution

Fired exchanger: if heated by combustion gases

Unfired exchanger: not using combustion gases

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LOGOHeat Transfer Equipment in

Industries1 9 8 5

MODES of HEAT TRANSFER

1. Conduction Transfer of heat from one part of a body to another part

of the same body or between two bodies in physical contact, without significant displacement of the particles of the two bodies

2. Convection Transfer of heat from one point to another within a fluid

or between a fluid and a solid or another fluid, by the movement or mixing of the fluids involved

3. Radiation Transfer of heat by the absorption of radiant energy

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BASIC THEORY

General equation for heat transfer across a surface for DPHE is:

Q =heat transferred per unit time, W U=the overall heat transfer coefficient, W/m2oC A= heat-transfer area, m2

Tm= the mean temperature difference,oC

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lmTUAQ

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BASIC THEORY

General equation for heat transfer across a surface for STHE is:

Q =heat transferred per unit time, W U=the overall heat transfer coefficient, W/m2oC A= heat-transfer area, m2

Tm= the mean temperature difference,oC

Y = geometric correction factor

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lmTUAYQ

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Tube-Side Passes

One tube pass

Two tube pass

Three tube passes

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Geometric Correction Factor1 9 8 5

Also refer to Figure 11-4, Perry 7th Edition

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Geometric Correction Factor

For design to be

practical, Y ≥ 0.85

212

212

21

2

)1(12

)1(12ln1

11

ln1

ZZX

ZZXZ

ZXX

ZY

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LOGO Logarithmic Mean Temperature Difference

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ΔT2ΔT1

1

2

12

lnTTTT

Tlm

If ΔT1 < ΔT2 and (ΔT2/ΔT1) ≤ 2, then ΔTlm is the arithmetic mean temp difference

LOGOOverall Heat Transfer

Coefficient

Rearranging the General Equation in terms of driving force and total resistance:

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lmTUAQ UA

TQ lm

1

Driving Force

Total Resistance

LOGOOverall Heat Transfer

Coefficient

The overall coefficient is reciprocal of the overall resistance to heat transfer, which is the sum of several individual resistances. Individual resistance is the reciprocal of individual HTC.

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totRUA

1

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Total Resistance

the sum of several individual resistancesIndividual resistance is the reciprocal of

individual HTC.

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conduction and convection from sresistance individual of sum1

totRUA

inside

ConvectionConductionConvection

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Total Resistance

Conduction Heat Transfer is governed by Fourier’s Law!

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dx

dTkA

dt

dQ

k = thermal conductivity of the Solid (BTU/hr-ft2-(OF/ft))

A = Area perpendicular to the direction of heat transfer

x = distance of heat flow

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Total Resistance

At Steady State:

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dx

dTkAq

dx

dTkA

dt

dQ

invariant time

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Total Resistance

If k is constant:

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kAxxTT

q

xx

TTkAq

)()(

)(

)(

12

21

12

21

Define R = Δx/kA

Thus, q= - ΔT/R

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If k is not constant:

2

1

2

1

T

T

x

xkdT

A

dxq

If k varies slightly with Temp:

AkxxTT

q

m

)()(

12

21

**km is evaluated at the mean temperature

Total Resistance1 9 8 5

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If k is not constant:

2

1

2

1

T

T

x

xkdt

A

dxq

If A varies slightly with Thickness:

mmAkxx

TTq

)()(

12

21


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Convection Heat Transfer

q = hcA (T1 – T2) Where:

hc- convection heat transfer coefficient, Btu/hrft2°F -similar to k/∆x A – Heat transfer Area T1 – temperature at surface 1

T2 – temperature at surface 2


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Convection Heat Transfer: Rearranging

q = (T1 – T2)/(1/hcA) Where:

hc- convection heat transfer coefficient, Btu/hrft2°F -similar to k/∆x A – Heat transfer Area T1 – temperature at surface 1

T2 – temperature at surface 2


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odooommidiiitot AhAhAk

x

AhAhR

UA ,,

11111

inside

ConvectionConductionConvection

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doomm

o

idi

o

ii

o

o hhAk

xA

Ah

A

Ah

A

U ,,

111

inside

mmiioo AUAUAUUA

1111

odo

i

oo

i

mm

i

diii Ah

A

Ah

A

Ak

xA

hhU ,,

111

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Typical Fouling Factor (Foust, 1980)

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Heat Transfer Without Phase Change

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DOUBLE PIPE HEAT EXCHANGER

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LOGOInvidual Heat Transfer

Coefficient

HT w/o Phase Change: DPHE

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Applicabilty:1. Non-metallic fluid2. 0.5 < NPr < 1003. NRE > 10,000

For Long Tubes (L/D) > 50, Tube-side

7.014.0

Pr8.0 1023.0 3

1

L

dNN

k

dhNu

wRE

ii


Coefficient


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Applicabilty:1. Non-metallic fluid2. 0.5 < NPr < 1003. NRE > 10,000

For Long Tubes (L/D) > 50, Annular Space

7.014.0

Pr8.0 1023.0 3

1

L

dNN

k

dhNu

wRE

eqo


Coefficient


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For Short Tube (L/D < 50)

7.0

1

L

D

h

h

i

is


Coefficient


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Laminar Flow, Forced Convection

14.0

31

2

wGZNU NN

kL

mcN p

GZ

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SHELL AND TUBE HEAT EXCHANGER


CoefficientHT w/o Phase Change: STHE, ho

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CoefficientHT w/o Phase Change: STHE, hi

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Heat Transfer WITH Phase Change

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CoefficientHT w/ Phase Change: STHE

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Film-type Condensation on Vertical Surface

Assumptions:

1. Pure vapor is at its saturation temperature.

2. The condensate film flows in laminar regime and heat is transferred through the film by condensation.

3. The temperature gradient through the film is linear.

4. Temperature of the condensing surface is constant.

5. The physical properties of the condensate are constant and evaluated at a mean film temperature.

6. Negligible vapor shear exists at the interface



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Film-type Condensation on Vertical Surface, Laminar

41

1

3943.0

TTL

gHkh

vl

vvlll



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Film-type Condensation on Vertical Surface, Turbulent

41

313.1

lvl

vvlll

TTL

gHkh



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Film-type Condensation on Horizontal Surface

41

3

725.0

lvl

vvlll

TTD

gHkh

For Nre > 40, h is multiplied by 1.2

If the amount of condensate is unknown



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Film-type Condensation on Horizontal Surface

31

3

95.0

W

gLkh

l

vlll


If the amount of condensate is known



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Film-type Condensation on Horizontal Surface, Banks of Tubes

41

3

725.0

lvl

vvlll

TTND

gHkh




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41

NhhN

43

43

43

41

...

...

21

21

n

n

NNN

NNNN



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41

1

3

725.0

TTND

gHkh

vl

vvlll

w/o splashing



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41

32

1

3

725.0

TTDN

gHkh

vl

vvlll

w/ splashing


CoefficientFilm Temperature

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Condensate Properties are evaluated at the Film Temperature

Tf = ½(Tsv + Tw) by Kern, D.Q., Process HT

Tf = Tsv - 0.75ΔT by McAdams, W.H., Heat Transmission,

3rd. Ed.

ΔT = Tsv - Tw


CoefficientFilm Boiling on Submerged Horizontal Cylinder or Sphere

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33.031.069.0

1225.0

v

llp Pk

A

qCh l



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satsv

satspvlvv

TTD

TTCgkC

A

qv

4.03



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7.032

k

CDGC

k

hD pr

Nusselt-type Equation by Rohsenow:

Cr varies from 0.006 to 0.015


Coefficient

Film Boiling on Submerged Horizontal Cylinder or Sphere

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3162.0

0015.0

k

CDG

k

hD p

Nusselt-type Equation by Forster and Zuber:

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HE DESIGN SPECS1 9 8 5

LOGOTOTAL HEAT

TRANSFER AREA1 9 8 5

lmTU

QA

DLNA T

A compromise between NT and L is chosen based on (L/Dshell) between 5 to 10

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HE DESIGN SPECIFICATION

No. of Tubes in Conventional Tubesheet Layout

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LOGOTOTAL HEAT

TRANSFER AREA1 9 8 5

CNDNDiameterShell COC 1

With an appropriate pitch to diameter ratio and optimum pipe diameter chosen and the total HT area,

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LAYOUT AND PITCH ARRANGEMENT

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LAYOUT AND PITCH ARRANGEMENT

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• Optimum Pitch to Diameter Ratio: 1.25 to 1.50

• Suggested clearance: 6.4 mm

Tube layout normally follows symmetrical arrangement having the largest number of tubes at the center

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BAFFLES

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Used to support tubes against sagging and vibrations

Direct the flow of fluid and control velocities

Types: Segmental Disk and Doughnut Type

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BAFFLES

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Segmental Baffles

Baffle Cut: 25 to 45% of disk diameterBaffle Spacing: 20 to 100% of Shell Diameter

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BAFFLES

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Disk and Doughnut Baffles

•Reduces pressure drop by 50-60%

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BAFFLES

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BAFFLES

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Minimum unsupported tube span (in.) acc. to Perry = 74d0.75

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BAFFLES THICKNESS: BENDING

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BAFFLES THICKNESS: SHEARING

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BAFFLES THICKNESS

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Pressure Drop

Tube-Side Pressure Drop (Coulson and Richardson, 2005)

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Basic Equation for isothermal system

Tube friction losses only

jf = dimensionless friction factorL’ = effective tube lengthDi = inside tube diameterρ = density of fluid at bulk/film temperature

ut = velocity of fluid

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Pressure Drop


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For non-isothermal systems

Tube friction losses only

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Pressure Drop


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W/ pressure losses due to contraction, expansion and flow reversal

Suggestions for the Estimation of these Losses:

1. Kern (1950) suggests adding 4 velocity heads per pass2. Frank (1978) considers this to be too high, and recommends

2.5 velocity heads3. Butterworth (1978) suggests 1.84. Lord et al. (1970) take the loss per pass as equivalent to a

length of tube equal to:a. 300 tube diameters for straight tubesb. 200 for U-tubes

5. Evans (1980) appears to add only 67 tube diameters per pass.

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Pressure Drop


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The loss in terms of velocity heads can be estimated by:

1. counting the number of flow contractions, expansions and reversals, and;

2. using the factors for pipe fittings to estimate the number of velocity heads lost

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Pressure Drop


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For two tube passes, there will be:

1. two contractions (0.5)2. two expansions (1.0)3. one flow reversal (1.5)

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Pressure Drop


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Pressure Drop

Shell-Side Pressure Drop (Coulson and Richardson, 2005)

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Pressure Drop


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Shell Equivalent Diameter (Hydraulic Diameter)

Square-Pitched Tube Arrangement, de in meter

Triangular-Pitched Tube Arrangement, de in meter

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Pressure Drop


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Shell-Side Friction Factor???

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Pressure Drop


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Shell-Side NOZZLE Pressure Drop

1 ½ velocity heads for the inlet

½ for the outlet

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Pressure Drop

RULES OF THUMBS (Silla, 2003)

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Pressure Drop

RULES OF THUMBS (Coulson and Richardson, 2005)

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Pressure Drop

RULES OF THUMBS (Couper, Penny, Fair & Wallas, 2010)

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•vacuum condensers be limited to 0.5–1.0 psi (25–50 Torr)

•In liquid service, pressure drops of 5–10 psi are employed as a minimum, and up to 15% orso of the upstream pressure

LOGOHeat Exchanger

Temperature Limits

RULES OF THUMBS

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•At high temperature, water exerts corrosive action on steel and scaling is increased

•To minimize scale formation, water temperature should not be more than 120ºF

•To protect against fouling and corrosion, water temperature (outlet) should not be more than158F

LOGOHeat Exchanger

Temperature Limits

RULES OF THUMBS

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•For the cooling water, on an open circulation systems, the temperature of the cooled water is 8-13ºF above the wet bulb temperature

• When using cooling water to cool or condense a process stream, assume a water inlet temperature of 90oF (from a cooling tower) and a maximum water outlet temperature of 120oF

LOGOHeat Exchanger

Temperature Limits

RULES OF THUMBS

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•the greatest temperature difference in an exchanger should be at least 36 degF, and;

•the minimum temperature difference should be at least 10 degF

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