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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
Agitated vessel
Fired heaters
Cooling and quenching
Heating and cooling
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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Heat 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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Heat Transfer Equipment in
Industries
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  UATlm




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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BASIC THEORY
General equation for heat transfer across a
surface for STHE is:
Q  UAYTlm





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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Tube-Side Passes
One tube pass
Two tube pass
Three tube passes
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Geometric Correction Factor
Also refer to Figure 11-4, Perry 7th Edition
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Geometric Correction Factor
Y
Z
2

1
1
2
 1 X 
ln

 1  ZX 
1
2

2  X Z  1  ( Z  1) 2 


Z  1 ln
1
2

2  X Z  1  ( Z  1) 2 


For design
to be
practical,
Y ≥ 0.85
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Logarithmic Mean
Temperature Difference
ΔT1
ΔT2
T2  T1
Tlm 
T2
ln
T1
If ΔT1 < ΔT2 and (ΔT2/ΔT1) ≤ 2, then ΔTlm
is the arithmetic mean temp difference
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Overall Heat Transfer
Coefficient
Rearranging the General Equation in terms of
driving force and total resistance:
Driving Force
Q  UATlm
Total Resistance
Tlm
Q
1
UA
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Overall 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.
1
 Rtot
UA
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Total Resistance
the sum of several individual resistances
Individual resistance is the reciprocal of
individual HTC.
Convection
Conduction
Convection
inside
1
 Rtot  sum of individual resistance s from convection and conduction
UA
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Total Resistance
Conduction Heat Transfer is governed by Fourier’s Law!
dQ
dT
  kA
dt
dx
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:
dQ
dT
 t imeinvariant  kA
dt
dx
dT
q   kA
dx
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Total Resistance
If k is constant:
(T1  T2 )
q  kA
( x2  x1 )
(T1  T2 )
q
( x2  x1 )
kA
Define R = Δx/kA
Thus, q= - ΔT/R
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Total Resistance
If k is not constant:
q
x2
x1
If k varies
slightly with
Temp:
T2
dx
   kdT
T1
A
(T1  T2 )
q
( x2  x1 )
km A
**km is evaluated at the mean temperature
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Total Resistance
If k is not constant:
q
x2
x1
If A varies
slightly with
Thickness:
T2
dx
   kdt
T1
A
(T1  T2 )
q
( x2  x1 )
km Am
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Total Resistance
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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Total Resistance
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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Total Resistance
Convection
Conduction
Convection
inside
1
1
1
x
1
1
 Rtot 




UA
hi Ai hi ,d Ai km Am ho Ao ho,d Ao
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Total Resistance
in
si
d
e
1
1
1
1



UA U o Ao U i Ai U m Am
Ao
Ao
Ao x 1
1
1



 
U o hi Ai hi ,d Ai km Am ho ho,d
Ai x
Ai
Ai
1
1
1
 



U i hi hi ,d km Am ho Ao ho,d Ao
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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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Invidual Heat Transfer
Coefficient
HT w/o Phase Change: DPHE
For Long Tubes (L/D) > 50, Tube-side
1   
hi d i
0.8
Nu 
 0.023N RE N Pr 3  
k
 w 
Applicabilty:
1. Non-metallic fluid
2. 0.5 < NPr < 100
3. NRE > 10,000
0.14
  d 0.7 
1    
 L 


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Invidual Heat Transfer
Coefficient
HT w/o Phase Change: DPHE
For Long Tubes (L/D) > 50, Annular Space
Nu 
ho d eq
k
 0.023N RE N Pr
0.8
1
Applicabilty:
1. Non-metallic fluid
2. 0.5 < NPr < 100
3. NRE > 10,000
3
  
 
 w 
0.14
  d 0.7 
1    
 L 


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Invidual Heat Transfer
Coefficient
HT w/o Phase Change: DPHE
For Short Tube (L/D < 50)
his
D
 1  
hi
L
0.7
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Invidual Heat Transfer
Coefficient
HT w/o Phase Change: DPHE
Laminar Flow, Forced Convection
N NU  2 N GZ
N GZ 
1
3
m cp
kL
 

 w



0.14
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SHELL AND TUBE
HEAT EXCHANGER
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Invidual Heat Transfer
Coefficient
HT w/o Phase Change: STHE, ho
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Invidual Heat Transfer
Coefficient
HT w/o Phase Change: STHE, hi
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Heat Transfer WITH
Phase Change
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Invidual Heat Transfer
Coefficient
HT w/ Phase Change: STHE
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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Invidual Heat Transfer
Coefficient
HT w/ Phase Change: STHE
Film-type Condensation on Vertical Surface, Laminar
 kl 3 l  l   v H v g 
h  0.943

l LTv  T1 


1
4
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Invidual Heat Transfer
Coefficient
HT w/ Phase Change: STHE
Film-type Condensation on Vertical Surface, Turbulent
 kl 3 l  l   v H v g 
h  1.13

l LTv  Tl 


1
4
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Invidual Heat Transfer
Coefficient
HT w/ Phase Change: STHE
Film-type Condensation on Horizontal Surface
 kl l l   v H v g 
h  0.725

l DTv  Tl  

3
1
4
 If the amount of condensate is unknown
For Nre > 40, h is multiplied by 1.2
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Invidual Heat Transfer
Coefficient
HT w/ Phase Change: STHE
Film-type Condensation on Horizontal Surface
 kl l l  v gL 
h  0.95

lW


3
1
3
 If the amount of condensate is known
For Nre > 40, h is multiplied by 1.2
LOGO
Invidual Heat Transfer
Coefficient
HT w/ Phase Change: STHE
Film-type Condensation on Horizontal Surface, Banks of Tubes
 kl l l   v H v g 
h  0.725

 l NDTv  Tl  
3
For Nre > 40, h is multiplied by 1.2
1
4
LOGO
Invidual Heat Transfer
Coefficient
HT w/ Phase Change: STHE
Film-type Condensation on Horizontal Surface, Banks of Tubes
h N  hN
N
1
4
1
4
N1  N 2  ...  N n
 3
3
3
4
4
N1  N 2  ...  N n 4
LOGO
Invidual Heat Transfer
Coefficient
HT w/ Phase Change: STHE
Film-type Condensation on Horizontal Surface, Banks of Tubes
 kl l l   v H v g 
h  0.725

 l NDTv  T1  
3
1
4
 w/o splashing
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Invidual Heat Transfer
Coefficient
HT w/ Phase Change: STHE
Film-type Condensation on Horizontal Surface, Banks of Tubes
 kl l l   v H v g 
h  0.725
2

3
 l N DTv  T1  
3
1
4
 w/ splashing
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Invidual Heat Transfer
Coefficient
Film Temperature
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
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Invidual Heat Transfer
Coefficient
Film Boiling on Submerged Horizontal Cylinder or Sphere
 qCpl 
h  0.225

 A 
0.69
 Pkl 
 


0.31
 l 
  1
 v 
0.33
LOGO
Invidual Heat Transfer
Coefficient
Film Boiling on Submerged Horizontal Cylinder or Sphere


3

gkv  v l   v    0.4C pv Ts  Tsat  
q
 C

A
v DTs  Tsat 


LOGO
Invidual Heat Transfer
Coefficient
Film Boiling on Submerged Horizontal Cylinder or Sphere
Nusselt-type Equation by Rohsenow:
 DG   C p 
hD


 Cr 


 k
k


 
2
3
Cr varies from 0.006 to 0.015




0.7
LOGO
Invidual Heat Transfer
Coefficient
Film Boiling on Submerged Horizontal Cylinder or Sphere
Nusselt-type Equation by Forster and Zuber:
 DG 
hD

 0.0015
k
  
0.62
 Cp 


 k 
1
3
LOGO
HE DESIGN SPECS
LOGO
TOTAL HEAT
TRANSFER AREA
Q
A
UTlm
A  NT DL
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
LOGO
TOTAL HEAT
TRANSFER AREA
With an appropriate pitch to diameter ratio and
optimum pipe diameter chosen and the total
HT area,
Shell Diameter NC DO  NC  1C
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HE DESIGN SPECIFICATION
LAYOUT AND PITCH ARRANGEMENT
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HE DESIGN SPECIFICATION
LAYOUT AND PITCH ARRANGEMENT
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HE DESIGN SPECIFICATION
LAYOUT AND PITCH ARRANGEMENT
• 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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HE DESIGN SPECIFICATION
BAFFLES
 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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HE DESIGN SPECIFICATION
BAFFLES
Segmental Baffles
Baffle Cut:
Baffle Spacing:
25 to 45% of disk diameter
20 to 100% of Shell Diameter
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HE DESIGN SPECIFICATION
BAFFLES
Disk and Doughnut Baffles
•Reduces pressure drop by 50-60%
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HE DESIGN SPECIFICATION
BAFFLES
LOGO
HE DESIGN SPECIFICATION
BAFFLES
Minimum unsupported tube span (in.) acc. to Perry = 74d0.75
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HE DESIGN SPECIFICATION
BAFFLES THICKNESS: BENDING
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HE DESIGN SPECIFICATION
BAFFLES THICKNESS: SHEARING
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HE DESIGN SPECIFICATION
BAFFLES THICKNESS
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Pressure Drop
Tube-Side Pressure Drop (Coulson and Richardson, 2005)
Basic Equation for
isothermal system
Tube friction losses
only
jf = dimensionless friction factor
L’ = effective tube length
Di = inside tube diameter
ρ = density of fluid at bulk/film temperature
ut = velocity of fluid
LOGO
Pressure Drop
Tube-Side Pressure Drop (Coulson and Richardson, 2005)
For non-isothermal
systems
Tube friction losses
only
LOGO
Pressure Drop
Tube-Side Pressure Drop (Coulson and Richardson, 2005)
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 pass
2. Frank (1978) considers this to be too high, and recommends
2.5 velocity heads
3. Butterworth (1978) suggests 1.8
4. Lord et al. (1970) take the loss per pass as equivalent to a
length of tube equal to:
a. 300 tube diameters for straight tubes
b. 200 for U-tubes
5. Evans (1980) appears to add only 67 tube diameters per
pass.
LOGO
Pressure Drop
Tube-Side Pressure Drop (Coulson and Richardson, 2005)
W/ pressure losses due to contraction, expansion and flow reversal
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
LOGO
Pressure Drop
Tube-Side Pressure Drop (Coulson and Richardson, 2005)
W/ pressure losses due to contraction, expansion and flow reversal
For two tube passes, there will be:
1. two contractions (0.5)
2. two expansions (1.0)
3. one flow reversal (1.5)
LOGO
Pressure Drop
Tube-Side Pressure Drop (Coulson and Richardson, 2005)
W/ pressure losses due to contraction, expansion and flow reversal
LOGO
Pressure Drop
Shell-Side Pressure Drop (Coulson and Richardson, 2005)
LOGO
Pressure Drop
Shell-Side Pressure Drop (Coulson and Richardson, 2005)
Shell Equivalent Diameter (Hydraulic Diameter)
Square-Pitched Tube
Arrangement, de in
meter
Triangular-Pitched Tube
Arrangement, de in meter
LOGO
Pressure Drop
Shell-Side Pressure Drop (Coulson and Richardson, 2005)
Shell-Side Friction Factor???
LOGO
LOGO
Pressure Drop
Shell-Side Pressure Drop (Coulson and Richardson, 2005)
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)
LOGO
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)
•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% or
so of the upstream pressure
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Heat Exchanger
Temperature Limits
RULES OF THUMBS
•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
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Heat Exchanger
Temperature Limits
RULES OF THUMBS
•For the cooling water, on an open circulation
systems, the temperature of the cooled water is 813ºF 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
LOGO
Heat Exchanger
Temperature Limits
RULES OF THUMBS
•the greatest temperature difference in an
exchanger should be at least 36 degF, and;
•the minimum temperature difference should be at
least 10 degF