Ali Farhan Thesis

Republic of Iraq
Ministry of Higher Education and
Scientific Research
AL-Mustansiriya University
College of Engineering
Study and Evaluation of the Operation
Characteristics for the Condensation Load
Distribution in Hybrid Systems on the
Condenser Side
A thesis
Submitted to the Mechanical Engineering Department
College of Engineering Al-Mustansiriya University in
Partial Fulfillment of the Requirements for the Degree of
Master of Science in Mechanical Engineering
By
Ali Farhan Muwayez Altemimi
(B.Sc.2007)
Supervised by
Dr. Ali Hussain Tarrad
(Assist. Prof.)
Rajab 1432 June 2011

To
My Family:
Father
Mother
And Brothers
Ali F. Muwayez

I
Acknowledgement
FIRST OF ALL, I THANK ALLAH FOR ALL HIS GIVING
The author would like to thank Dr. Ali Hussain Tarrad for his
supervision, invaluable advise, thankful guidance, and continuous
encouragement given during this study.
Thanks to Mr.Ali Jaber, an engineer who works in the Combustion
Laboratory in mechanical engineering department.
Thanks also to Mr.Ali Abdul-Kareem Mohammed, Mechanical
Engineer in Central Organization for Standardization and Quality Control
(COSQC), Ministry of Planning, for his assistance.
Thanks extended to the stuff of the Fluid, Turbomachineary and
Thermodynamic Laboratories in Mechanical Engineering Department,
University of Al-Mustansiriya for their help in using the various
facilities.
Finally, I would like to thank my beloved parents, for their great
patient and support during the period of this study and before this study,
Thank you for everything.
Ali F. Muwayez
June 2011

II
ABSTRACT
In this study, the performance of a vertical air cooled condenser is
investigated under both moderate and hot ambient conditions. Air pre-
cooling method was used to cool the entering air to the condenser.
However, a new combination (Hybrid) is described by using air and
water in the same time for separated condensers, (ACC and WCC). The
water cooled condenser is oriented vertically and is connected in parallel
with the ACC. The experimental work is accomplished in an ambient
temperature ranging between (20.7 °C) to (42°C) with air velocity range
of (3) to (6) m/s for ACC, inlet temperature range of (15) to (23) °C and
water volumetric flow rate of (0.2) to (1) m3
/h for WCC. Hybrid mode is
used in hot ambient period with air dry bulb temperature ranging between
(31°C) to (42°C), air velocity of (3) m/s, water inlet temperature at
(30°C) and volumetric flow rate range of (0.2) to (0.4) m3
/h.
The ACC showed a reduction in steam capacity and condensation
load in hot ambient. Therefore a reduction of air temperature from (42 to
20.7°C) at design air velocity of (3 m/s), would lead to an increase in the
steam mass flow rate and thermal load by (52 %) and (55%) respectively.
The WCC showed an increase in steam capacity of (54%) to (48%)
with increasing water volumetric flow rate from (0.2) to (1) m3
/h inlet
temperature variation from (15°C) to (23°C) respectively.
The overall steam capacity of the hybrid system showed an increase
of (14.2%) and the thermal load of (22%) when it is compared with that
of the ACC unit at dry-bulb temperature of (42°C). Pre-cooling of the air
will contribute by the steam loading of ACC of about (0.66) kg/h per each
degree reduction of air dry-bulb temperature.

III
The variation of mean overall heat transfer coefficient for ACC shows
a value of (89 to 60) 2
.
W
m C
with air velocity range of (6 to 3) m/s, steam
saturation temperature of (100°C), air inlet temperature of (20.7°C) to
(42°C),steam mass flow rate of (35.7) to (17.7) kg/h. WCC mean overall
heat transfer coefficient was (835 2
.
W
m C
) and (196 2
.
W
m C
) respectively,
with water volumetric flow rate of (0.2) m3
/h, water inlet temperature of
(15) to (23°C), saturated steam temperature of (100°C) and steam mass
flow rate of (24.24) to (10.2) kg/h.
A computer model is built for both ACC and WCC, ACCRP and
WCCRP programs respectively. The rating models for each condenser
showed a good agreement with the experimental measured values with
deviation of (+12) to (-5%) for ACC load and (-5%) for air exit
temperature, (+ 13) to (-10 %) for WCC load and (+ 5%) to (- 5%) for
water exit temperature.

IV
LIST OF CONTENTS
TITLE PAGE
Acknowledgment……………………………………………………… I
Abstract ………………………………………………………………. II
List of Contents……………………………………………………...... IV
Nomenclature ...……………………………………………………………….... VIII
CHAPTER ONE: INTRODUCTION…………………………………....... 1
1.1 General............................................................................................................. 1
1.1.1 Water Cooled Condenser ……………………………………….. 3
1.1.2 Surface Condenser with Once-Through Cooling Water System... 4
1.1.3 Surface Condenser with Evaporative Cooling Towers ………… 4
1.2 Air Cooled Condenser ……………………………………………. 5
1.2.1 Advantages and Disadvantages of Dry Cooling with an Air
Cooled Condenser ……………………………………………………………...
6
1.3 Alternative Condensing Systems……………………......................... 8
1.4 Aims of The Work ………………………………………………... 9
CHAPTER TWO: LITERATURE SURVEY ……………………..... 16
2.1 General …………………………………………………………… 16
2.2 Literature Categories ……………………………………………………... 16
2.2.1 Air Cooled Condenser Performance Evaluation and Modeling… 16
2.2.2 Water Cooled Condenser Performance Evaluation and
Modeling ……………………………………………………………...
21
2.2.3 Air-Water Cooling System Performance Evaluation.................... 23
2.3 Scope of The Work ………………………….................................. 24
CHAPTER THREE: EXPERIMENTAL APPARATUS …………... 25
3.1 Test Rig …………………………………………………………... 25
3.2 Steam Generator (SG) ……………………………………………. 28
3.3 Shell and Tube Condenser ……………………………………….. 29
3.4 Air Cooled Condenser (ACC) …………………………………… 31
3.5 The Air Cooled Condenser Fan ………………………………….. 35

V
TITLE PAGE
3.6 Duct System ……………………………………………………….. 35
3.7 Piping Lines ………………………………………………………... 36
3.8 System Insulation ………………………………………………….. 37
3.9 Water Loops ……………………………………………………….. 38
3.10 Instrumentation ………………………………………………….... 39
3.10.1 Temperature Measurements …………………………………...... 39
3.10.2 Pressure Measurements ………………………………………..... 40
3.10.3 Water Measurements …………………………………………..... 41
3.10.4 Front Face Air Velocity Measurements ……………………….... 42
3.10.5 Condensate Measurements.................... ……………………….... 42
3.11 The Electrical Board and Temperature Controllers ……………...... 43
3.12 Calibration of Measuring Instrumentation ……………………….... 43
3.12.1 Calibration of The Thermometers ……………………………...... 43
3.12.2 Calibration of The Pressure Gauges …………………………....... 44
3.12.3 Calibration of The Rotameter ………………………………….... 44
3.13 Experimental Procedure …………………………………………... 44
3.13.1 Primary Check up ……………………………………………...... 44
3.13.2 Test Procedure …………………………………………………... 45
CHAPTER FOUR: THEORETICAL RATING MODEL
DEVELOPMENT....................................................................................
50
4.1 Introduction......................................................................................... 50
4.2 Data Reduction.................................................................................... 50
4.2.1 Load Capacity................................................................................... 50
4.2.2 Overall Heat Transfer Coefficient.................................................... 51
4.2.3 Comparison of the Results.............................................................. 51
4.3 Evaluation of Thermo-Physical Properties ………………………… 51
4.4 Condensation Heat Transfer Coefficient …………………………… 52
4.4.1 Effect of Condensation Film Sub-cooling………………………… 53

VI
4.4.2 Effect of Condensate Waves And Turbulence …………………… 54
4.4.3 Effect of Vapor Velocity …………………………………………. 55
4.5 Forced Convection Inside Tube ……………………………………. 58
4.6 Air Side Heat Transfer Coefficient ………………………………… 60
4.7 Shell Side Heat Transfer Coefficient for Single Phase …………….. 61
4.8 NTU Effectiveness Method ………………………………………… 62
4.9 Air Cooled Steam Condenser Modeling ……………………………. 64
4.9.1 Air Cooled Steam Condenser Geometry.......................................... 65
4.9.1.1 Tube Geometry.............................................................................. 65
4.9.1.2 Air Side Geometry ……………………………………………… 66
4.9.2 Computational Procedure ………………………………………… 68
4.10 Water Cooled Steam Condenser Modeling ……………………… 72
4.10.1 Computational Procedure ……………………………………… 73
4.11 Model Verification............................................................................ 77
4.11.1 Air Cooled Condenser Verification................................................ 77
4.11.2 Water Cooled Condenser Verification........................................... 77
4.11.3 Overall Heat Transfer Coefficient Distribution............................. 77
CHAPTER FIVE: RESULTS AND DISCUSSIONS............................. 90
5.1 Introduction......................................................................................... 90
5.2 Experimental Results........................................................................... 90
5.2.1 Air Cooled Condenser Mode............................................................ 90
5.2.1.1 Air Flow Rate................................................................................ 90
5.2.1.2 Dry-Bulb Temperature.................................................................. 92
5.2.1.3 Condenser Steam Loading............................................................ 93
5.2.2 Evaporative Cooling Mode............................................................... 94
5.2.2.1 Air Flow Rate................................................................................ 94
5.2.2.2 Dry-Bulb Temperature.................................................................. 94
5.2.3 Water Cooled Condenser Mode....................................................... 95

VII
5.2.3.1 Cooling Water Flow Rate.............................................................. 95
5.2.3.2 Inlet Water Temperature............................................................... 96
5.2.4 Hybrid (Separated Condenser Mode)............................................... 96
5.2.4.1 Contribution of ACC and WCC in Hybrid System Overall Load. 98
5.2.5 Overall Heat Transfer Coefficient.................................................... 98
CHAPTER SIX: CONCLUSIONS AND RECOMMENDATIONS.... 112
6.1 Conclusions…………………………………………………………. 112
6.2 Recommendations…………………………………………………... 113
REFERENCES…………………………………………………………………... 114
APPENDICES……………………………………………………………………. A-1
Appendix (A) …………………………………………………………… A-1
Appendix (B) ………………………………………………………….... B-1
Appendix (C) …………………………………………………………… C-1
Appendix (D) …………………………………………………………… D-1
Appendix (E) ………………………………………………………….... E-1

VIII
NOMENCLATURE
A Area m 2
B Central Baffle Spacing m
Bc Baffle Cut %
C Clearance Between
Adjacent Two Tubes
m
Cr Heat Capacity Ratio -
cp Specific Heat at Constant
Pressure
J/kg.°C
D Diameter or Depth m
Fp Fin Pitch m
g Gravitational Acceleration m/s2
G Mass Velocity of Fluid kg/m2
s
h Specific Enthalpy kJ/kg
hc Steam Condensing Heat
Transfer Coefficient
W/m2
°C
ho Out side Heat Transfer
Coefficient
W/m2
°C
H Height m
k Thermal Conductivity W/m °C
K Constant in equation(4-8) -
L Tube Length m
lf Fin Length m
m
Fluid Mass Flow Rate kg/s
n Parameter Defined by
Equation (5-46), Correlation
index
-
N Number -
Nu Nusselt Number -
p Operating Absolute
Pressure of Steam
bar
P Perimeter of Tube m
Pr Prandtle Number
Pr
pc
k
Q
Condensation Load kW
Re Reynolds Number
Re hud
Sm Hollow Fin Cross Sectional
Area
m2
t Thickness m
T Fluid Temperature °C or K
T Temperature Difference °C
u Fluid Velocity m/s
Uo Overall Heat Transfer
Coefficient
W/m2
°C

IX
W Air Cooled Condenser
Width
m
w Specific Humidity of Air kg. steam/kg.dry air
xeq Steam Local Quality -
XT Transverse Tube Pitch to
the Flow Direction
m
XL Longitudinal Tube Pith in
Flow Direction
m
z Heat Exchanger Cooled
Length
m
ABBREVIATIONS
ACC Air Cooled Condenser
ACCRP Air Cooled Condenser Rating Program
CCS Combined Cooling System
CW Cooling Water
EPA Environmental Protection Agency
EPRI Electric Power Research Institute
GEA Aktiengesellschaft (Germany Group)
GIEST Gryphon International Engineering Services Incorporated
HRSG Heat Recovery Steam Generator
ITD Initial Temperature Difference
LMTD Log Mean Temperature Difference
NTU Number of Heat Transfer Unit
OCI OCI Instruments Incorporated (USA Company)
OHTC Overall Heat Transfer Coefficient
WCCRP Water Cooled Condenser Rating Program
WCC Water Cooled Condenser
WETSAC Wet Surface Air Condenser
GREEK SYMBOLS
f Fin Efficiency defined by eq.(5-73)
o Total Surface efficiency defined by eq. (5-75)
Effectiveness of the Condenser
Density (kg/m3
)
Viscosity (kg/m.s)
Percentage of Error in Thermo-physical Properties
Condensate Tube Loading (kg/s.m)
Parameter defined by eq. (5-79)
Parameter defined by eq. (5-76)
Ψ Parameter defined by eq. (4-2)

X
SUBSCRIPTS
a Air
ass. Assumed Value
att. Attached
b bundle or base
c Condensate , cold
cBK Boyko-Kruzhilin condensation coefficient
cal. Calculated Value
cond Condenser, Condensation
db Dry-Bulb
e Exit or Equivalent
eff. Effective
eq. Equivalent
est. Estimated Value
f Fouling , Fin or fluid
g gas
h Hydraulic diameter, hot fluid
i Inside or Inlet
inc increment
l Liquid
lo Liquid Only
m mean
max maximum
min minimum
o Outside
r ratio,row
row Row Value
s Steam Value or Surface
sc Sub-cooling
sh Shell
t Tube
tb Tube bundle
tr tube per row
Tube Tube Value
Vapor
w Wall Value
wb Wet-Bulb

Chapter One Introduction
1
Chapter One
Introduction
1.1 General:
A power plant may be defined as a machine or assembly of
equipment that generates and delivers a flow of mechanical or electrical
energy .The main equipment for the generation of electric power is the
generator .When coupling it to prime movers, the electricity is generated.
The type of prime move determines the type of power plants. The major
power plants are:
1. Steam Power Plant.
2. Diesel Power Plant.
3. Gas Power Plant.
4. Nuclear Power Plant.
These power plants are called THERMAL POWER PLANT, because
it converts heat into electric energy. A steam power plant includes a
Boiler, Turbine, Generator, and feed water pump as shown in Figure (1-
1). Low pressure water condensate in the steam condenser is pumped to
high pressure before entering the boiler where superheated steam is
produced. The superheated steam is then sent to the steam turbine where
the steam expands to low pressure providing energy to drive a generator.
This low pressure steam has to be condensed in a condenser in order to
complete the steam cycle.

2
The steam condenser is one of the major components of the steam
cycle in steam power plant. It condenses the exhaust steam from the
turbine and thus recover the high-quality feed water for return to the
boiler and increases the cycle efficiency by allowing the cycle to operate
at largest possible temperature and pressure difference between the boiler
and condenser, Matthew (2002) [1].
The function of the condenser is to condensate the exhaust steam
from the steam turbine by rejecting the heat of condensation to the
circulating cooling water passing through the condenser. If the circulating
cooling water temperature is low enough, it will then create a low back
pressure (which is usually vacuumed below atmospheric) for the turbine.
This pressure is equal to the saturation pressure that corresponds to the
condensing steam temperature, and which in turn, is a function of the
cooling water temperature. The turbine work per unit pressure drop at the
low pressure is much greater than the high pressure end of turbine, El-
wakil (1985) [2].
A condenser by lowering the back pressure will lead to increase the
work of the turbine and then increases the plant efficiency, and reduces
the steam flow for a given plant output, El-wakil (1985) [2].
One of the important features of the thermal power plants is the
condensation of the vapor exhausted from the low pressure turbine
stage.The condensate being recycled back through the system by three
major condensation systems that are employed:-
1. Water – Cooled surface condensers and wet condensing systems.
2. Air – Cooled Condensers.
3. Alternative Condensing systems.

3
1.1.1 Water Cooled Surface Condenser:
In surface condensers, there is no direct contact between the
steam and cooling water. The condensate can be re-used in the boiler in
such condenser even impure water can be employed for cooling purpose.
Although more capital cost and space are needed in surface condensers,
but it is justified by the saving in running cost and the increase in
efficiency of plant achieved by using this condenser. Depending upon the
position of condensate extraction pump, flow of condensate and
arrangement of tubes, the surface condensers may be subdivided into:-
1. Down Flow Type:- The down flow type condenser depends on the
steam flow direction The exhaust steam enters the shell from the
top flows down and over the tubes and gets condensed and is then
finally removed by an extraction pump. The water flowing through
the tubes in one direction from inlet to exit end. Figure (1-2) shows
a longitudinal section of a two pass down-flow condenser.
2. Central Flow Condenser: - In the central flow condenser type, the
steam passages are all around the periphery of the shell as shown in
fig. (1-3) .Air is pumped away from the center of the condenser
and the condensate moves radially towards the tube nest.
3. Evaporation Condenser: - This type of condensing is essentially
used when a limited quantity of water is available. Its quantity is
needed to condense the steam which can be reduced by causing the
circulating water to evaporate under a small partial pressure. The
exhaust steam enters at the top of the condenser through gilled
pipes. The water pump sprays water on the pipes and descending
water condenses the steam. The water which is not evaporated,
falls into the open tank (cooling pond) under the condenser, from
which it can be drawn by circulating water pump and used over
again. The evaporative condenser is placed in open air and finds its

4
application in small size plants. Figure (1-4) illustrate the layout of
this condenser, Raja (2006) [3].
1.1.2 Surface Condenser With Once –Through Cooling
Water System:
This is the most efficient cycle heat rejection system design.
The type of circulating water system uses cold water from a nearby
river, lake or ocean and then circulates it through the condenser to
remove condensation heat and returns the hot water to the source.
Figure (1-5) shows a typical flow diagram for a once-through
cooling system. Special design for intake and outfall must be
recommended preventing trash and debris from reaching the main
circulating water pumps. Flow straightening elements designed to
provide streamlined flow to the pumps are also contained in the intake
structure with compatibility of handling changes in the water level due
to seasonal or tidal variations. After flowing through the condenser,
the hot water is returned to the environment through an outfall
structure which is designed to discharge the water at an optimal
velocity. The plant performance in terms of power output and
efficiency is regarded as best result for the once-through option,
Rattan (2003), Beychok (2010), [4,5].
1.1.3 Surface Condenser with Evaporative Cooling
Towers:
Evaporative cooling towers provide cooling based on the
design of wet bulb temperature. It have the advantage that cooling
water chemistry is better and is able to be controlled, and thus
reducing the loss of condenser performance due to fouling. The
concrete hyperbolic cooling towers structure that are part of so many
plants may not be allowed in a specific areas because the plume of

5
cooling tower also introduce problems .There are two major types of
cooling towers, natural draft and mechanical draft. In the first type, it
depends on air flow upon the natural driving pressure caused by the
difference in density between the cool outside air and the hot, humid
air inside. In the second type the air is moved by one or more
mechanically driven fans. Figure (1-6) shows a power plant diagram
with cooling tower for condenser cooling water, Beychok (2010) [5].
1.2 Air Cooled Condenser:
Air cooled condenser in power plant termed as a dry cooling
system transfers heat to the ambient by convection and radiation.
Instead of evaporation as the wet tower does, Robert and Green (1997)
[6]. Atmospheric air has been used for many years to cool and
condense fluid in sites of water scarcity. During the 1960s, the use of
air cooled condenser grew up rapidly in the United States and
elsewhere. This is because of the heat transfer process which removes
latent heat from the condensing steam to the sensible heat of ambient
air. The ambient dry bulb temperature controls the condensing
temperature and pressure achievable with dry cooling, Rattan (2003)
[4].In an air cooled condenser (ACC), the rejection heat from the
steam is dissipated directly to the atmosphere without using an
intermediate medium such as cooling water, as shown in figure (1-7).
In this arrangement, the steam exhausting the turbine is piped to the
condenser by a large diameter duct .The steam condenses in the air
cooled tube bundles. The finned tube bundles are mounted in a
horizontal, vertical or A- frame configuration and mounted on a steel
structure support. The steam enters the tube bundles at the top of the
A-frame or vertical air cooled condenser and condenses in the tube.
The heat is removed by air blown over finned-tube bundle by one or
more fan. The fans can be either forced or induced draft, as shown in

6
figure (1-8). This depends on the weather where the air is pushed or
pulled through the finned tube bundle. The condensate drains into a
collection tank and is then sent back to the steam generator feed water
system, EPA (2000) [7].
The location of the ACC must consider the large space
requirements and the possible recirculation of heated air. In the
horizontal position of ACC, the height of the finned tube bundle above
ground must be one-half of the tube length to produce an inlet velocity
equal to the face velocity [7].
1.2.1 Advantages and Disadvantages of Dry Cooling with
an Air- Cooled Condenser:
In the steam power plant industry, the reduced availability of
water as the cooling medium for the condensation of exhaust steam
often makes the selection of an air-cooled condenser as alternative to
the traditional steam surface condenser. Advantages of air cooling
include the following:
1. Dry cooling system minimizes the water usage requirements and
there will be no issues which are associated with blowdown
disposal and plume formation.
2. Air is more abundant quantity compared to water with no
preparation costs and this has brought an increase interest in the
use of the air as a cooling medium in place of water.
3. (ACC) designed for power plant technology evolved into a
configuration with ability of condensing a large volume of low
pressure vapor as well as the removal of non-condensibles.
4. (ACC) minimizes the environmental impacts from withdrawing
water and quickly dropping it back to the cooling water source

7
during the summer months, so causing damage to the surrounding
ecosystem.
5. Water is a corrosive fluid which requires a chemical treatment to
control the scaling leading to a reduction in the heat transfer rate if
the deposits are allowed to accumulate. This is in contrast to air
which is mostly noncorrosive.
6. Air- cooled condensers make it possible to build a power plant in
location without adequate cooling water resources.
7. Maintenance may be reduced due to elimination of water fouling
characteristics which could require frequent cleaning of water
cooled heat exchanger, Cheremisinoff (2000) [8].
Disadvantages of the Air- cooling are:
1. Energy consumption per (kg) condensate is higher for dry cooling
than for wet cooling (WCC) and the atmospheric emissions
associated with that energy consumption is also higher.[7]
2. Air-cooled condensers require larger surface area to obtain a given
heat removal rate, and the area increasing with the design ambient
air temperature. Water cooled condensers require much less heat-
transfer surface.
3. Air-cooled condensers are more expensive that it's water-cooled
equivalent.
4. Air cooler is noisy because of the large number of fans working to
move air to finned-tube bundle.
5. Air-cooled condensers affected by ambient air temperature and site
cyclonic winds. The energy penalty also increases with the ambient
air temperature.
6. The external surfaces of finned tubes on air cooled condensers are
affected by fouling from dust.

8
1.3 Alternative Condensing Systems:
In this new technology, more adequate configuration is used in
power plants to compensate the most popular already discussed
problems. This is accomplished by using ACC or water-cooled
condenser (WCC) in condensing system. The combined configuration
formed by parallel connected of ACC and WCC as a facility to obtain
a good performance and compatibility for work in hot ambient and
water scarcity conditions. There are many types of combination exist
in the power plant field depending on system configuration and water
consumption. Combined (Parallel) condensing systems, have been
developed to save water while avoiding the high cost of dry cooling
systems and to ensure a relatively low steam turbine back pressure at
high ambient conditions. An excessive rise in steam turbine
backpressure during periods of peak ambient temperatures and
demand will result in a loss of efficiency of the steam turbine
generator set. The dry section of the system may be designed to reject
the total heat load at a low ambient temperature while maintaining the
turbine backpressure with specified limits, at high ambient
temperatures using the wet part of the system. One way of the sizing
the wet part of a combined condensing system (CCS) is to limit the
quantity of make-up water according to the local water availability. In
CCS the water consumption is reduced compared to 100% wet system,
the performance is improved compared to a 100% dry system and the
capital cost decreases as the proportion of wet in the CCS system is
increased, Backer and William (2003) [9],Figure (1-9) shows a
configuration of dry/wet cooling system (CCS).
Another technologies are used for condensing system in power
plants like a wet surface air cooled condensers (WETSAC) and
Cooling ponds, A cooling pond is a human-made pond which is used

9
for purposes of cooling. Cooling ponds are utilized most commonly
with power plants which use water for cooling, although there are
alternatives such as cooling towers which can be considered for
cooling needs at such facilities. In a cooling pond, hot water is pumped
through the pond to cool down. The heat is transferred through the
water, bringing the temperature down so that the water can be used
again. This is part of a process known as (wet cooling). One big
drawback to wet cooling is water loss through evaporation, which can
become a problem if water supplies are limited. In this case, water is
not a great choice of cooling material because the plant must import
more water on a regular basis to keep the cooling systems operational,
EPA (1970) [10]. About (1-2) acres (1 acre = 4046.86 m2
) of water
surface area required per (MW) of condensing, GIESI (2010) [11].
1.4 Aims of the Work:
The prediction of performance of heat exchanger is becoming
very important from the economical and thermal points of view.
Therefore, the present work will give details about rating of steam
turbine condensers. The main objects of the present work are as
follows:
1. To build an experimental rig to provide a proper operation
conditions for the condensers. Also, to prepare procedure for the
performance prediction of air-cooled and a vertical single tube pass
water-cooled condensers.
2. To investigate the performance enhancement of ACC condenser
when applying a prior cooling of air at inlet of ACC especially
during the summer season (with high ambient temperature in Iraq).
3. To build a computer program for each condenser used in the
present work and then carrying out proposed procedure using a
step by step calculation.

10
4. To provide an assessment for the advantage of using the
Combined Cooling System in Iraq depending on the ambient
temperature and availability of water, and determine if it represents
the best technique available for minimizing the disadvantages of
both ACC and WCC.
Figure (1-1): Schematic diagram of Simple Steam Power Plant.

11
Figure (1-2): Longitudinal section of a two pass down-flow condenser
[3].
Figure (1-3): Central-flow condenser. [3]

12
Figure (1-4): Side -view of evaporative condenser. [3]
Figure (1-5): Typical Flow Diagram for a Once-Through Cooling
System [4].

13
Figure (1-6): Surface condenser with Cooling Tower.
Figure (1-7): Typical ACC Flow Diagram. [4]

14
Figure (1-8): Typical components of ACC, Forced and Induced
Draft,[5].

15
Figure (1-9):Combined Condensing System (Dry/Wet cooling system).[8]

Chapter Two Literature Survey
16
Chapter Two
Literature Survey
:General.12
It has been recognized recently that the air cooled condensers are
environmentally preferable to the traditional water cooled condensers for
rejecting heat in thermal power plants. Since air cooled condenser is
considered a water saving system, it has a broad prospect in many regions
in the world. Therefore, the performance of air cooled condensers has
been studied by several investigators, and various methods were
developed to alleviate the air cooled condenser performance decline with
increasing ambient air temperature. A new approach is proposed in this
study that has the ability to improve the air cooled condenser in hot
ambient condition. This can be performed by combining the (ACC) with
an individual water cooled condenser operates with a specific amount of
water cooling flow rates according to water availability.
2.2 Literatures Categories:
The previous works which are related to the present study was
divided into three main categories according to fluid flow conditions and
types of heat exchangers used. This will be explained in the following
sections.
2.2.1 Air Cooled Condenser Performance Evaluation and
Modeling:
Larinoff and Moles (1978) [12], identified the problems that may be
associated with the operation of the air cooled steam condensers in power
plants. They investigated the effect of the ambient air temperature on the
turbine backpressure and recommended the range from (16) to (43) o
C as
limitation for the turbine back pressure.

17
Kutscher and Costenaro (2002) [13 ], developed a model to assess
the cost and performance of four methods for using supplemental
evaporative cooling to boost the summer performance of air cooled
condenser in geothermal power plants. The four modes is pre-cooling
with spray nozzles, pre-cooling with munters media, a hybrid
combination of nozzles and munters media and direct deluge cooling of
the air cooled condenser tubes. A (1 MW) binary- cycle plant at Empire,
Nevada,(US). Proposed to employ the air cooling and augment summer
heat rejection with some form of evaporative cooling figure (2-1). A
system in which water directly contacts the condensate tubes (deluge
cooling) has the highest performance and is economically the most
attractive. However, consideration of scaling and corrosion must be
addressed.
Figure (2-1): Monthly Electricity Production for evaporating cooling
Enhancement Methods.

18
Jabardo and Mamani (2003) [14], investigated the performance
evaluation of parallel flow micro channel finned tube condenser for
automotive application. A simulation model has been developed based on
dividing the condenser into three zones related to the thermodynamic
state of the refrigerant in the condenser. The condenser simulation model
assumes that the heat transfer surface is divided into three regions:
superheating, condensing and sub-cooling, each region are treated as an
independent heat exchanger. The discrepancy between the experimental
data and the simulation results has shown a good agreement.
Wilber and Zammit (2005) [15], outlines the problems found in the
field of operation and procurement of air cooled steam condensers. They
recommended an appropriate design condition for air cooled condenser
by using an ACC for installation at a 500MW, gas fired combined cycle
plant located in an arid, desert region. For a nominal (500MW,2 x 1 )
combined cycle plant, the steam side capacity is approximately about
(170 MW) with a corresponding steam flow of about (1.1 x 106
lb/hr ~
0.5 x 106
kg/hr). The performance requirements for ITD will typically in
the range (14) to (33.3) o
C , steam turbine exhaust pressure (back
pressure) in the range (0.084) to (0.24) bar, and the ambient temperature
will typically in the range (-17.8) to (43.3) o
C .
Gadhamshetty V., et.al. (2006) [16], proposed a new approach to
alleviate the performance decline in air cooled condenser with increasing
the air dry-bulb temperature. A chilled water thermal energy storage
system is used to pre-cool the inflow air to the ACC whenever the
ambient air temperature increases above (20 o
C). A 500 MW combined
cycle power plant with steam turbine net output power of 170.9 MW was
chosen to evaluate the application of this approach for power plant to be
located in southern New Mexico. Figure (2-4) shows the proposed

19
procedure used the 171 MW plant saves (2.5%) of the power (4.2MW)
without using any water or incurring any water treatment cost.
Figure (2-4): The Proposed System for ACC performance enhancement
with Increasing Air Dry-Bulb Temperature.
Jung and Assanis (2006) [17], presented a predictive numerical
model for the cross flow type single phase flow finned tube heat
exchanger with louvered fins. Using finite difference method with
staggered grid system based on the thermal resistance concept. The model
can predict the effect of detailed design variable and geometry change on
the heat rejection performance, for the (0.25 m2
) frontal core area and hot
side (water) volume flow rate of (4.8 m3
/h) the heat rejection increase by
(50%) with increasing of air velocity by (100%).
Abood (2007) [18], presented an experimental and theoretical study
of the best tilted angle of the A-frame air cooled condenser and the effect
of fan speed variation and backpressure on the performance of the air

20
cooled condenser with shading and without shading for condensing unit.
The study shows that the best tilted angle is (30 o
) and the performance
enhancement is achieved by using a spray cooling technique.
Tarrad et.al., (2008) [19], investigated the performance of the cross
flow air cooled heat exchanger with no phase change. Experimental and
theoretical studies were conducted on two different types and sizes of
finned tube air cooled heat exchanger. The values of the overall heat
transfer coefficient are ranged from (34) to (41) W/m2
K for (1250) CFM
and from (40) to (53) W/m2
K for (2500) CFM.
Tarrad et.al., (2009)[20], investigated the performance of a cross
flow air cooled exchanger with no phase change . The experimental work
was carried out on a finned tube air cooled heat exchanger, the OHTC for
inlet water (hot side) temperature of (50°C) and volume flow rate of (2
m3
/h) varied between (30-47), load was increased by approximating (3)
times with air velocity raised from (1.2) to (4.6) m/s. A quasi two-
dimensional model was built to perform the overall dimension of the air
cooled heat exchanger. The maximum discrepancy between the
experimental data and those calculated by model for overall performance
was about (3 %) for the given range of the simulated conditions.
Tarrad (2010) [21], developed a numerical model for performance
prediction of dry cooling of the air cooled condensers applied in power
plants technology. A computer code was built that depends on the idea of
using the row by row technique for estimating the heat transfer
coefficient , air temperature and air physical properties distribution in the
air flow direction from row to row. The model results showed an
improvement in the condensation load up to (23%) when air pre-cooling
mode applied to inflow air to the ACC to lower the dry-bulb temperature
from (45) to (28) o
C at constant wet-bulb temperature and fixed heat

21
exchanger surface area, the OHTC range of (82) to (80.4) 2
W
m C
, and air
face velocity of (3.6) m/s.
2.2.2 Water Cooled Condenser Performance Evaluation and
Modeling:
Yusur (1997) [22], investigated a step by step method for
thermal and hydraulic design of a horizontal shell and tube condenser. He
presented a design procedure based on Silver Bell Ghaly method for
multi-component condensation. In this method the baffle spacing is
subdivided into a number of steps selected by the designer, and the mean
output conditions of each step are taken as the input to the next step.
Tarrad and Kemal (2001) [23], investigated a step by step
method for thermal and hydraulic design of steam power plant condensers
for one and two passes. A computer program proposed to calculate the
condensation rate distribution and temperature of vapor, condensate,
outer tube wall, pressure distribution in shell side, outlet cooling water
and mean cooling water for each bay in the condenser. A comparison was
made between the field data (Al-Daura Thermal Power Station) and the
simulation model and the comparison show a good agreement.
Tarrad and Majeed (2007) [24], investigated a step by step
technique to predict the performance of the power plant condenser. A
computer code was built by using FORTRAN 90 language to evaluate
the vapor temperature, pressure distribution, condensation temperature,
water inlet and outlet temperature and some other parameters. A
comparison was made between field data and that predicted shows a good
accuracy.
Yusuf A.K. and O.Guraras (2004) [25], provided a computer-
based design model for a shell and tube heat exchanger with no phase
change fluid flow both on shell and tube side. The program covers

22
segmental baffled U-tube, and fixed tube sheet heat exchangers, one-pass
and two-pass for tube-side flow. The program determines the overall
dimensions of the shell, the tube bundle, and optimum heat transfer
surface area required.
Mohammed (2005) [ 26 ], investigated experimentally and
theoretically the thermal-hydraulic design of shell and tube heat
exchanger using the step by step technique for a single tube pass. They
provided a computer program to simulate the performance of heat
exchanger for a smooth tube bundle and a low finned tube bundle.
Tarrad (2007) [ 27 ], provided a computer program design model for the
thermal-hydraulic design of a single shell-single pass of enhance tube
bundle heat exchanger using the step by step technique. The model basic
design depends on the thermal and economical point of view by selection
of the low finned tube characteristics to avoiding the space restriction of
the equipment layout in the failed application and reduction in the cost of
manufacturing machining.
Thirumarimurugan et.al. (2008)[28] investigated the performance
of a single shell-single pass shell and tube with water is the hot fluid at
(100°C), whereas water and Acetic acid-water miscible solution serves as
cold fluid with constant inlet temperature of (29°C). The experimental
work conducted at a variable water discharge from (0.12-0.72) m3
/h, and
a volume fraction of Acetic acid is varied from (10-50) %. The OHTC
increased from (126.167) to (150.15) W/m2
.°C according to the heat
exchanger used. A mathematical model was built using MATLAB
program to calculate the outlets temperatures of both the shell and tube
side fluids and the overall heat transfer coefficient.

23
2.2.3 Air-Water Cooling System Performance Evaluation:
EPRI (2004) [29], identified the numerous design arrangements
that may exist for hybrid systems. Hybrid systems employ a combination
of both wet and dry cooling technologies. The hybrid system is water
conservation and plume abatement design. Water conservation system is
intended to reduce (but not completely eliminate) the use of water for
plant heat rejection. A limited amount of water is used during the hottest
periods of the year to mitigate the large losses in steam cycle capacity and
plant efficiency associated with all-dry operation. These systems can limit
annual water use to (2 to 5%) of that required for all-wet systems (with
water cooled condenser) and still achieve substantial efficiency and
capacity advantages during the peak load periods of hot weather, as
compared to an all-dry system.
Heyns (2008) [30], Studied the performance characteristics of a
power plant incorporating a steam turbine and a direct air cooled dry/wet
condenser operating at different ambient temperatures. The proposed
cooling system uses existing A-frame air cooled condenser with a hybrid
(dry/wet) dephlegmator to achieve enhancement in cooling performance
when ambient temperatures are high. The study showed that for the same
turbine power output the water consumed by an air cooled condenser with
a hybrid dephlegmator is at least (20 %) less than an air cooled condenser
with spray cooling of the inlet air.
NREL (2011) [31], identified and evaluate methods by which the
net power output of an air-cooled geothermal power plant can be
enhanced during hot ambient conditions using minimal amounts of water.
Geothermal power plants that use air-cooled heat rejection systems
experience a decrease in power production during hot periods of the day,
therefore hybrid cooling options, which use both air and water, have been

24
studied to assess how they might mitigate the net power decrease in hot
periods.
2.3 Scope of the Work:
There are many studies conducted for the enhancement air
cooled condenser performance because of the water scarcity and
environmental protection. However the studies did not focused on the
using a water cooled condenser as a method to handle the air cooled
condenser performance deterioration. Water cooled condenser must be
restricted to the small amount of cooling water for the summer condition
working in a combined mode with air cooled condenser to improve the
performance requirement in power plant.
All of the above investigations of air cooled condenser and water
cooled condenser are conducted to the usual installation of these
condensers in power plant technology or industrial fields.
In the present work, according to the experimental conditions
conducts and the study purpose, two condensers are combined in parallel
and in a vertical orientation with instrumentation. Each condenser can
work individually with a specified steam loading according to the steam
generator feeding. The combined mode for the water cooled condenser
under a low cooling water flow rate is considered as a new proposed
method to investigate the air cooled condenser capability in a hot ambient
condition. Also evaporative cooling for air inflow is used to decrease the
high air dry-bulb temperature.
For the model development, it is suggested to use the segment by
segment technique to analyze the heat exchangers. A tracking approach
was applied in the modeling simulation to evaluate the heat transfer
characteristics.

Chapter Three Experimental Apparatus
25
Chapter Three
Experimental Apparatus
3.1 Test Rig:
An experimental facility was constructed to allow two types of
condensing system worked as a test arrangement. Each one represents a
separate unit having all the specifications and instruments which allows
condensation data to be collected over a range of operating conditions.
Air cooled condenser (finned tube heat exchanger), and shell and tube
heat exchanger were used as a single or in a hybrid arrangement.
The objectives of the experimental facility are:-
1. To obtain different sets of experimental heat load and condensation
rate over a variety range of service stream conditions, temperature
and flow rates, for single or combined form of ACC and WCC.
The experimental apparatus was modified after a first set of data to
ensure an accurate operation condition compatibility to weather
criteria in Iraq in hottest period. This was accomplished by adding
an electrical heating coil works as air heater for a set exit
temperatures.
2. To study the ability of this type of combination between the ACC
and WCC to compensate the deterioration of air cooled condenser
performance in hot ambient or the water scarcity when the power
plant condensing system depending on (100%) water as a cooling
medium.
Figure (3.1) presents photographic views of the experimental test
system. It consists of two heat exchangers for condensation purposes
with process and service fluids in an open loop cycle. Figure (3.2)
shows a schematic diagram of the experimental system.

26
Figure (3-1): Photographic views of the experimental test facility.
Duct Heating System
Steam Generator Feeding System
Electrical Board
Water cooled condenser
"Shell and Tube"
Air cooled condenser
Glass level Indicator
Graded Cylindrical Container
Ducting system with instrumentation
Steam Generator
Thermostat Controller
Cooling Water Supplying System
Centrifugal Fan

27
Figure (3-2): Schematic diagram of experimental set-up*
.
* All fluid power equipments are validated with references, [32, 33].

28
3.2 Steam Generator (SG):
The steam generator is constructed so that it can have the ability
to provide wet steam at atmospheric pressure. It is made of galvanized
steel sheet of (3 mm) wall thickness to ensure operating under high
pressure safely. However, the pressure did not exceed (2 bar) in the
present study. The steam generator core capacity is about (120
litter).Water is pumped to the SG with a specific quantity, which must
be suitably cover (at height of 50 to 60 cm) the heating element inside
SG. This can be checked by the sight glass level (high temperature
proof) installed on the SG, The steam production is about (9.57 kg/h)
at low pressure (1 atm).A safety valve is fixed at the top of SG and
pressure gauge is installed for pressure measurement and monitoring.
Figure (3-3) illustrates the construction of the steam generator.
Figure (3-3): Steam generator used in the present experimental set-up.

29
3.3 ٍShell and Tube Condenser:
Shell and tube heat exchanger which is a Flovex-milan type
(BEM/S 301.2-A-1) are used as condenser and placed vertically with
condensation on the shell side. The heat exchanger of (325mm) tube
effective length contains (28) tube. Each tube is made of stainless steel
with (9.45 mm) outer diameter and (8.68 mm) inner diameter, as
shown in figure (3-4).
The tubes are distributed as a rotated square of (45o
) tube layout.
The clearance between two adjacent tubes is (3 mm), and the tubes
pitch is (12.94 mm).The shell diameter is (81.64 mm). The baffle
space is (70 mm) and baffle cut is (25%).The shell and tube condenser
constructed for counter-flow configuration, in which the process fluid
(steam) in shell side flows in opposite direction to the cold water
which flows in tube side. The shell side inlet and outlet nozzles are of
(42.5 mm), the tube side inlet and outlet connections are of (50.8 mm),
and the total heat transfer area is (0.270) meter square.
Instruments for temperature and pressure are connected to the tubes
and shell sections. The gauges are fitted in prepared pockets on the
required locations.

30
(a) dimensions in cm.
(b)
Figure (3-4): (a) Test section of Shell and Tube Condenser (b) Tube
layout.

31
3.4 Air Cooled Condenser (ACC):
Air cooled condenser is a finned tube heat exchanger, typically
used for the process which consists of a finned-tube bundle with
rectangular box headers on both ends of tubes. The "ACC" used is a
vertical type (single pass, two rows) with flatted tubes occupied each
row with equal numbers. as shown in figure (3-5) and (3-6), In ACC,
heat is transferred from the process fluid (steam) to the cooling
medium (air) through the fin tube bundle. It depends on the
temperature difference (driving force) between air and steam so that
the dry bulb temperature of air is a key control of the ACC
performance. Therefore, the dry cooling system with ACC is less
efficient in hot ambient. The centrifugal fan installed in front of the
ACC to move air towards finned tube bundle, flatted tube geometry
used to enhance heat transfer inside tube with extra heat transfer area
and reduce pressure drop outside tube. The tubes material are brass
which has excellent physical properties [34], compared with other
materials. The fins are constructed to the tubes so each fins row
contains (70 fins) attached to tube in equal pitch about (4.18 mm).The
fins material is a copper, the physical characteristic and dimensions
are listed in table (3.1). The whole assembly is mounted on legs with a
rubber and fasteners to well-set during operation.

32
(a) (b) [21]
(c)
(d)
Figure (3-5): (a) Photographic views of ACC (b) Schematic of ACC
(c) ACC tube layout (d) Part of tubes in two rows.
Air flow

33
Figure (3-6): Test Section of the ACC (dimensions in cm).

34
Table (3-1): ACC Geometry and Physical Characteristic*
.
DimensionParameter
590 mmWidth (W)
30 mmDepth (D)
320 mmHeight (H)
110No. of Tubes
55No. of Tubes/Row
2No. of Rows
140No. of fins/ Tube (on both sides)
11.25 mmTransverse Distance (XT) to flow
15 mmLongitudinal Distance (XL) to flow
0.1888 m2
Frontal Area (Aƒron )
Core
4.18 mmPitch (ƒΡ)
8.05 mmLength (lƒ)
0.18 mmThickness(tƒ)
36.12mm2
Area of single Fin (Aƒ)
CopperMaterial
388Thermal conductivity (W/m.C).[35 ]
Fin
2.35 mmHeight (Ht )
12 mmDepth (Dt )
0.24 mmThickness (tt)
BrassMaterial
SmoothInner tube surface
119Thermal Conductivity(W/m.o
C).[35 ]
Tube
3.935 m2
Total Surface Area (Atotal )
3.029 m2
Total Fin Surface Area
0.906 m2
Total Bare Tube Area
Area
* All dimensions are measured by electronic Caliper and compared with standard data.

35
:The Air Cooled Condenser Fan5.3
The ACC fan is of the centrifugal type with forward
configuration. This is a part of evaporative air cooler construction
manufactured by (M. P. Co.), model (BA28). The delivered air
volumetric flow rate has two different values either (1200) or (2400)
CFM. This is coincident with two speeds of electrical motor.
:Duct System6.3
The main object of this system is to direct air from the centrifugal
fan upstream towards ACC at a specific quantity of flow required.
Flexible type of (1-1.5 m) length with (30 cm) diameter is used in
experimental set-up to join the exit side of the fan to the inlet square duct
(590 x 300 mm2
) fixed on ACC frontal. Connections were sealed to
prevent air leakage and minimize this loss.Flexible duct is very
convenient for attaching supply air outlets to the rigid ductwork.
However, the pressure loss through flex is higher than that for most other
types of ducts. An attempt was done to keep the installed length short,
less than (4.572 m) or so to minimize turns and high pressure loss, Kinks
in flex must be avoided [36]. Figure (3-7) shows the duct system
assembly. As a facility to control the air inlet temperature rise, the duct
heating part of (165 x 310 mm2
) is added to the ductwork system. It has
(10 kW) coiled heating element work with (220-240 volt), situated inside
the inlet air construction with suspended screws to adjust heater to centre.
The sides in the front of heater were covered with special design wings to
prevent air by-passing and to direct air toward the heating section to heat
up the air to a desired value coincides with hot ambient temperature
measured in Iraq.

36
Figure (3-7): Duct System Assembly and Heating Section Specification.
:sLinePiping7.3
One inch diameter commercial galvanized steel pipe with
different lengths was used to connect the water feeding system. Screwed
fittings are used to prepare a suitable jacket for temperature and pressure
gauges on specific locations. To facilitate the assembly of screwed
fittings small lengths of pipe called pipe nipples are used between fittings
along the desired pipe line length, figure (3-8-b), [37].
Union was used to join the main parts of the water cooling loop to easily
remove for maintenance or alternate working conditions, figure (3-8-a).
A (3/4" - 1") diameter commercial galvanized steel pipe is used to
connect the process fluid (steam) system side. Instruments are mounted
with tee connection, all pipe sections are insulated with (25 mm) glass
Flexible connections

37
wool commercial insulation. Short length flexible join similar to that used
in automobile is positioned to transfer steam to the ACC.
(a) (b)
Figure (3-8): (a) Positioning of Unions. (b) Screwed Fittings Drawing
Symbols. [37]
:System Insulation8.3
The water cooled condenser, condensate tank, together with the hot
side piping system are all closely insulated with glass wool insulation.
The glass wool is the most popular and widely used insulation material. it
is made from fiberglass arranged into texture similar to wool, with very
low thermal conductivity (k=0.035 W/m.C).The WCC and piping system
are insulated with a sheet of glass wool having a thickness of (25 mm)
[38].

38
:Water loops9.3
There are two main water loops used in the present work
arrangement. The first loop is in which one flow through the WCC side
while the other was used to feed the steam generator.
On the WCC side, the water was used as a cooling medium
removing the latent heat from the condensing steam. Water was pumped
by a centrifugal pump driven by an electrical motor (370 watt) having
(2900 RPM) at (220-240 volt). The trade name is 'CRAFFT ALFA', with
a maximum capacity of (30 l/min).and (30 m) head. The pump takes the
water from main reservoir with (200 liter) capacity. Another reservoir
equipped with heater (3000-6000 watt) was used with ability of heating to
operate as an emergency tank when water cut-off occurs and to give
warm water depending on the experiments conducted, figure (3-2).
On the steam generator feeding side, the water was pumped to the
intake portion of steam generator with a limited quantity. This will insure
the safety work of heating element and will minimize the time that
required producing steam.
A hand controlled valve is used to adjust the water flow rate through the
test section (WCC or ACC). The flow rate of the water through the test
sections was measured with variable area flow meter. The outlet water
from the WCC exit section is drained to sewage or is returned to
emergency tank.
Over a range of experimental sets, water in both steam generator and
water supply tanks must be blown down to prevent scale forming and
corrosion reactions.

39
:Instrumentation10.3
:Temperature Measurements.110.3
Five stem glass thermometers were installed in the water and steam
side to measure the temperature of steam (process fluid) and water
(service fluid) at different locations. The technique is used by installing
the thermometers with a prepared jacket to immerse the thermometer
through the flow by using screwed fittings mentioned earlier, fig.(3-9).
This method confirms the accurate measured values for temperature
measurements that it provides a direct contact between the bulb and fluid.
Temperature gauges have a range of (0 o
C – 120 o
C) at a division
of (1o
C), maximum and minimum errors are corresponding to (2%) and
(-1.2%) respectively. Digital thermometers (type-K) are used to measure
the temperature on each inlet and exit sides of ACC condenser. Table (3-
2) shows the specification of digital thermometer. The wet bulb
temperature of air entering the condenser was measured by using the wet
cotton wick at the bulb of thermometer sensor.
Table (3-2): Temperature Sensor Characteristics.
-20o
C to 1000o
C
Resolution 1o
C
-20o
C to 0o
C ± (5% of rdg +4 digits)
0o
C to 400o
C ± (1% of rdg +3 digits)Accuracy
400o
C to 1000o
C ± (2% of rdg +3 digits)
Note: Accuracy is specified for a period of year after manufacture calibration at
(18 o
C to 28 o
C) with relative humidity to 75%.

40
Figure (3-9): Temperature gauge mounted on pipe.
:Pressure Measurements.210.3
Five stem glass pressure gauges were used to measure the
pressure at each exit and inlet ports of service fluid (water side) and also
at the inlet side of process fluid (steam) of ACC, and on WCC. No
pressure gauges are needed for the exit side of steam loops because the
system is opened to atmosphere. One gauge was installed on the steam
generator to control the build up the inside pressure with presence of
safety valve at steam generator exit point. Two of the pressure gauges on
the water side have ranges of (0-2.5 bar).The division (0.05 bar),
maximum and minimum errors are (0) and (-0.3) respectively, which
were connected to the tube side of water cooled condenser. One have a
range of (0 – 60 kpa), the division (1 kpa) with error not exceed (3 %)
which connected to the shell side to measure the condensation pressure.
The last one was installed on the inlet to the air cooled condenser and has
a range of (0 – 2.5 bar) the division (0.05 bar) and maximum error is
about (2 %).

41
:Water Measurements.310.3
A rotameter KDG 2000 was used to measure the flow rate of the
circulating water in (l/h) has a range of (200 – 1800 l/h), as shown in
figure (3-10). It is a ' G.CUSSONS ' type variable area flow meter.
Calibration of the rotameter has an error of about (1 %) l/min.
(a) (b)
Figure (3-10): (a) Rotameter with Control Valve (b) A Schematic
Diagram of the Rotameter.

42
:Front Face Air Velocity Measurements.410.3
Front face air velocity was measured by using digital anemometer
provided with liquid crystal display (LCD /10 mm /4digits).manufactured
by ' Electronics Company Italy'. Using (6F22-9V battery), as shown in
figure (3-11).
Figure (3-11): Microprocessor Digital Anemometer.
Table (3-3): Electrical specifications of digital anemometer.
AccuracyResolutionRange
±(2% + 1rdg)0.10.4 – 30 (m/s)
2% + 3rdg)(±0.11.4 – 108.0 (km/h)
2% + 2rdg)(±180 – 5910 ( ft/min )
2% + 2rdg)(±0.10.8 – 58.3 ( knots )
Note: Operating temperature (0 o
C to 50 o
C) and humidity less than 80%.
:Condensate Measurements.510.3
The quantity of steam condensate was collected in graded cylindrical
glass container with capacity of (2L) prepared at exit ports of each type of
condensing system (ACC and WCC).
This method used to determine the mass flow rate of steam per unit
time (kg/s) that condensed in each condenser for each experimental run

43
with a limited time period. where the typical cylindrical container used in
this work. This type of graded laboratory cylinder is used to calibrate the
rotameter for cooling water side.
:The Electrical Board and Temperature Controllers11.3
The electrical board contains main circuit breaker and other
secondary switches, which supply the electrical power to the whole
system components of the experimental rig. The control panel consist of
an electrical contactor (1 x 28 Amp), which is connected to the steam
generator heater and the scale thermostat (0 o
C to 150 o
C). This contactor
is controlled by thermostat electrical signal to switch on and off the
electricity to the heater.
There are two switches, each one. controls one of the two single
stage pumps. Each of these switches is connected with a small light to
show and ensure that all equipments are in or out of operation.
Moreover, there are two add-on circuits one for the air heater and
the other for water heater to produce a warm water according to the
experimental condition.
3.12 Calibration of Measuring Instruments:
All the measuring devices have been calibrated before the
installation in the experimental test rig. The calibration process of each
individual device will be described as below:
3.12.1 Calibration of the Thermometers:
The five thermometers are calibrated in the Central Organization
for Standardization and Quality Control (C.O.S.Q.C), Iraq-Baghdad.
Calibration curves of the gauges are shown in Appendix (A).

44
3.12.2 Calibration of the Pressure Gauges:
The calibration of the pressure gauges is carried out by
comparing their reading with respect to a standard pressure gauge
reading. The standard pressure gauge is calibrated in (C.O.S.Q.C)
manufactured by (OCI Instruments /USA), with grade A. The calibration
curves of the gauges are shown in Appendix (A).
3.12.3: Calibration of the Rotameter:
The flow meter is calibrated using a known volume graded
container and a stop watch. The reading of the flow meter is recorded and
compared to the water accumulated in the graded container for specific
time interval.
Various flows have been rated to obtain the rotameter readings via
the volume to time reading. The calibration curve for the rotameter is
shown in Appendix (A).
3.13 Experimental Procedure:
3.13.1 Primary Check up:
Before starting the experimental tests, it was necessary to do
several inspections to ensure that all testing components are well installed
and ready for operation process. The following checks have been
accomplished before starting the tests:
(1) Checking the water level in the SG, which must be suitably above
the heating element in the SG core, this can be checked by the
sight glass level installed on the SG.
(2) Checking the cold water level in the tank, and opening the source
of water that feeds this tank.
(3) Opening all valves installed before pressure gauges of the system.
These valves are installed in order to protect the pressure gauges

45
and to avoid errors in their readings during and after finishing
tests.
(4) Valve installed on the SG is kept closed during the initial starts of
tests, and until steam mass flow rate reaches the required quantity
according to test conducts. Then opening valves and steam
(primary fluid) start flow to the test section. Pumps and air cooler
fan start to ensure circulation of these secondary fluids in the
system.
(5) Checking all switches and wiring that control the electrical
equipments in experimental rig and ensure the safety side for
operation.
3.13.2 Test Procedure:
After completing inspection steps, the first step of the test process
is to turn on the main switch of the electrical control panel that supplies
power to the whole system. Then do the following procedures:-
1. Set the thermostat of the SG to desire temperature to put the
electrical power on to the heater inside SG. Heater indicator (Red)
on the electrical board will be flashed. The process of steam
production takes about (1.5-2) hour to produce the required steam
depending on the water volume occupied in the SG core, and the
water inlet temperature.
2. with regard to the experimental work which is divided into four
categories as below then:-
a. Use water cooled condenser (WCC) alone.
b. Use the air cooled condenser (ACC) alone.
c. Use both condensers with steam mass flow rate through both
of them.

46
d. Use evaporative cooling (cooler) of air prior to the inlet of
air cooled condenser.
Each mode of condensation described above tested with a specific
operation conditions described in Table (3-4), to evaluate the
performance and effect of using each mode with comparison.
3. When using mode (a), switch on the cold water pump prior to the
steam flowing for homogeneous water flow and temperature
distribution, steam mass flow rate is determine by collecting a
specific amount of condensate through a limited time period, and
the volumetric water flow rate is set according to the required flow
rate, ranging between (200 l/h) to (1200 l/h).
4. It is recommended to circulate the cooling water throughout the
system prior to steam flow to avoid temperature mal-distribution of
water throughout the condenser.
5. Open the main valve at the top of the boiler where steam is allowed
to pass to the system and record the time of flowing process.
6. Collecting the following, after reaching the steady state condition:-
a. The ambient conditions (Air dry-bulb and wet-bulb
temperatures).
b. The water flow rate and temperatures of water at the inlet
and exit sides of the condenser.
c. The steam inlet side pressure and temperature.
d. The amount of condensate collected along a specific time
(l/s).
e. The condensate exit temperature.
7. When using mode (b), switch on the fan of evaporative cooler to
delivered air to the air cooled condenser with air velocity of (3) or
(6) m/s, at the same time of steam flowing. If the heating up of

47
entering air is required then turn on the air heater to warm air to
desired temperature according to temperature controller installed,
as in figure (3-7).
8. Collecting the following data, after reaching the steady state
condition:
a. The ambient conditions (Air dry-bulb and wet-bulb
temperatures).
b. The Air face velocity.
c. The temperatures of air at inlet and exit sides of the
condenser, In addition the wet-bulb temperature was
measured on the inlet and exit sides of condenser. A mesh of
thermocouples with a small ball tips, allocate at different
locations at condenser exit to determine the mean air exit
temperature and to reduce the error occurs in the reading
when only one thermocouple used, (5 thermocouples with 1
mm ball tips),fig.(3-12).
d. The steam entering pressure and temperature.
e. The amount of condensate collected along a limited time
(l/s).
f. The condensate exit temperature.
9. When using mode (c), applying all the steps for mode (a) and (b),
only for the temperature of air inlet with the range of (31°C) to
(42°C) were corresponding to hot ambient conditions for air side,
and (30°C) inlet water temperature.
10. For the mode (d), the evaporative cooler is used with water
circulation to cool down air before to the condenser. Therefore,
improvement can be attained for the performance of air cooled
condenser. Data will collected as shown in mode (b).

48
Table (3-4): The Experimental Testing Map.
Cooling Mode
Air
Moderate
Temperatures: (20.7
°C) to (25.2°C).
Boiler Pressure
Variation (1.2 bar) to
(1.8 bar).
Air Flow Rate
Variation (1200)
CFM, (2400) CFM.
Hot Ambient
Temperatures (31°C)
to (42°C).
Boiler Pressure
(1.8 bar).
Air Flow Rate
Variation (1200 to
2400) CFM.
Evaporative Cooling
Inlet Air
Temperatures (27°C) to
(37.5°C)
Air Flow Rate
Variation (1200 to
2400) CFM.
Boiler Pressure
(1.8 bar).
Water
Inlet Water
Temperatures:
(15°C),(19°C),(23°C).
Water Flow Rate
Variation: (200 l/h) to
(1000 l/h).
Boiler Pressure fixed
to (1.8 bar).
Hybrid (Separated
Condensers)
Air Side:
Temperatures
Variations (31°C) to
(42°C).
Air Flow Rate is fixed
to design value of
(1200) CFM.
Water Side:
Inlet Water
Temperature is fixed
to (30°C) design value
in summer condition,
[39 and 40].
Water Flow Rate is
fixed to (20 to 40 %)
of the available water
flow rate of (1000 l/h).
* Boiler Pressure is Fixed to
(1.8 bar).

49
Figure (3-12): Thermocouples Distribution in Exit Air Duct (dimensions
in cm).

Chapter Four Theoretical Rating Model Development
50
Chapter Four
Theoretical Rating Model Development
:Introduction.14
This chapter deals with the mathematical model for rating both
of heat exchangers used in the present work. A step by step approach will
be used to simulate the performance prediction for each type of
condensing system used under a variable operation conditions. This
technique is the base on the idea of dividing the condenser to sub-
condenser segments and then studying each segment as an individual
condenser modeled one by one along the steam flow direction.
4.2 Data Reduction:
The experimental data which listed in appendix (B) are used to
obtain the following parameters:-
4.2.1 Load Capacity:
The condenser heat rejection is evaluated as follows:-
stcond fgQ x m h (4-1)
Where (x ) is the steam quality. All dry steam at saturation condition has
quality equal ( 1x ).therefore equation (4-1) can be written as:
stcond fgQ m h (4-2)
When the sub-cooling occurs according to minimum steam mass flow
rate or the minimum value of entering temperature or maximum flow
rates, the single-phase heat duty can be calculated as below:
( )sc st l sat eQ m cp T T (4-3)

51
The total condensation load for both ACC and WCC is calculated by:
total region cond scQ Q Q Q (4-4)
The data uncertainty for the (Q ) calculation is illustrated in appendix
(C).
4.2.2 Overall Heat Transfer Coefficient:
The overall heat transfer coefficient for each region and overall fixed
area can be calculated from the following equation:
LMTD
Q
U
A T
(4-5)
Where (A) represent the surface area according to the region considered
and the total area is:
cond scA A A (4-6)
The data uncertainty for the (U) calculation is illustrated in appendix (C).
4.2.3 Comparison of Results:
It is necessary to compare the performance of each type of
enhancement mode used in present experimental work at different
conditions as a percentage by using the following expression:
2 1
1
( %) 100 (4-7)
where ( ) represents any performance parameter ( , exitQ T ).
4.3 Evaluation of Thermo-physical Properties:
All the thermo-physical properties data required in the modeling
procedure were formed as curve fitting equations. The Lab Fit curve
fitting software version 7.2.37 (2007) [41] was used to generate the
equations with acceptable small error difference for a specified limits of
temperature and pressure as shown in appendix (D). The properties
equations for dry and mist air were used from the open literature, Kroger
(2004) [42], Rogers and Mayhew (2005) [43].

52
4.4 Condensation Heat Transfer Coefficient:
Condensation inside and outside vertical tubes depends on the
vapor flow direction and its magnitude. During the downward steam
flow, and if the vapor velocity is very low, then the condensate flow is
controlled by gravity. Then it will be within the control limits of Nusselt
model (1916) [44], conservative unless the tube inside diameter is very
small and tube wall curvature effects become important, [45].The Nusselt
model for a laminar condensation of a pure single component saturated
vapor is expressed as:
1/33
( )
0.925
( )
l l v l
c
l
gk
h
z
(4.8)
Where vertical tube loading on cooled length (z), condensate rate per unit
tube perimeter is estimated as follows:-
.
( ) / iz m D , when condensing inside tube (4.9)
.
( ) / oz m D , when condensing outside tube (4-10)
For a tube bundle:
Inside condensation:
.
( ) / t b iz m N D (4.11-a)
Outside Condensation:
.
( ) / t b oz m N D (4.11-b)
Equation (4.8) is applicable up to Reynolds number of (30), Moran and
Shapiro (2006) [46], the Reynolds number for the condensate film is
given by:
Re 4 ( ) /c lz (4.12)
Substituting (4.12) into (4.8) and rearranging, the mean heat transfer
coefficient up to point z is, [47]:

53
1/3
2
1/3
1.47Re
( )
c l
l l l
h
k g
(4.13)
In the present work, the step by step increments technique is
achieved by dividing the tubes into small length segments with a
specified condensate load according to the vapor fraction distribution
along each segment length. This approach produces a thermal
performance calculation coincide with each segment condition.
Therefore, the above relations are used to a specified length (z).
Many investigators have made significant improvements to the
original Nusselt work to include the different effects which are not
included into the original theory. However, they indeed exist during an
actual condensation process such as the effects of condensate film
subcooling, vapor velocity and the wave effect. The most common
theoretical treatment is to consider these coefficients as separate
correction factors applied to the Nusselt equation.
4.4.1 Effect of Condensate Film Subcooling:
The latent heat of vaporization is the heat released as a unit mass
of vapor condenses at a specified temperature. It represents normally the
heat transfer per unit mass of condensate forming during condensation.
However, the condensate in an actual condensation process is cooled
further to some average temperature between the saturation temperature
of steam and surface temperature of the tube, releasing more heat in the
process. Therefore, the actual heat transfer will be larger. Bromley (1952)
[48], extended the Nusselt theory to include subcooling of the condensate
in the heat balance.

54
Rohsenow (1956) [49], showed that empirically adding a sensible
heat term to the latent heat of vaporization gave reasonable results. The
effective latent heat of vaporization is estimated as:
*
( )
1 0.68
p l sat w
g g
g
c T T
h h
h
(4.14)
Drew (1954) [50] proposed that other physical properties to evaluate at an
effective film temperature in which he established the following equation:
0.25( )eff w sat wT T T T (4.15)
Recently, Cengel (2008) [51] recommends that the properties of
liquid should be evaluated at the film temperature, which is
approximately equal to the average temperature of the liquid and it is
estimated as:
( ) / 2f sat wT T T (4.16)
4.4.2 Effect of Condensate Waves and Turbulence:
At Reynolds numbers greater than about (30). It was observed
that a wave form at the liquid-vapor interface although the flow in liquid
film remains laminar. The flow in this case is said to be wavy-laminar.
The waves at liquid-vapor interface tend to increase heat transfer. The
increase of heat transfer is due to the wave effect is on average about
(20%). But it can exceed 50% in some cases. Kutateladze (1963)[52]
recommended the following correlation which could be applied to
equation (4.13):
0.11
0.8Re
wavy
c
h
h
(4.17)
As the local condensate film thickness (the film Reynolds number)
increases, the film will become unstable and then the turbulence will
develop in the condensate film. The heat transfer mechanism then

55
undergoes a significant change. This is because the heat is transferred
across the condensate film by turbulent mixing as well as by molecular
conduction. The main controlled parameter concerning the flow regimes
is the condensate velocity, which in expressed by its Reynolds number as
follows:-
4
Re l
l
i lD
(4.18)
And:
4
Re l
l
o lD
(4.19)
Where l is defined as:
( ) )l eqz x (4.20)
For Rel < 30, use equation (4.13).
For 30 < Rel < 1600, Kutateladze recommended the correlation for
laminar wavy zone:
1/3
2
1.22
Re
( ) 1.08Re 5.2
c l l
l l l l
h
k g
(4.21)
For Rel > 1600, Labuntsov correlation for turbulent zone [53]:
1/3
2
0.5 3/4
Re
( ) 8750 58Pr (Re 253)
c l
l l l l l
h
k g
(4.22)
4.4.3 Effect of Vapor Velocity:
When the vapor velocity is high, the vapor will pull the liquid
along the interface. This is because the vapor velocity at the interface
must drop to the value of the liquid velocity. If the vapor flows downward
this addition force will increase the average velocity of the liquid and thus
the film thickness is decreased. This in turn, will decrease the thermal
resistance of the liquid film and thus increasing heat transfer. For

56
condensation inside tubes in air cooled condenser model the tube inside
diameter is very small and the influence of tube wall curvature is
important in condensation phenomenon. Tubes with small inside diameter
of (1~5 mm) are controlled by shear stress and surface tension, Xiaoze
and Buxuan (2002) [54]. Therefore, the correlation by Boyko-Kruzhilin
(1967) [55] was used to estimate the mean condensing coefficient for a
steam between inlet quality ix and outlet quality ox inside tubes as below:
( ) ( )
2
m i m o
c BK loh h (4.23)
Where:
1 l
i
m i
x (4.24)
1 l
o
m o
x (4.25)
loh Sensible heat transfer coefficient assuming that the total fluid is
flowing with condensate properties (condensate filled the tube and was
flowing alone). This can be evaluated using any suitable correlation for
forced convection in tubes. Boyko and kruzhilin used the correlation:
0.8 0.43
0.021( / )Re Prlo l i lo lh k D (4.26)
In a condenser, the inlet stream assumes to be saturated vapor and the
vapor will be totally condensed. For these conditions equation (4.23)
becomes, Sinnott (2005) [56]:-
1 ( )
2
l
c BK loh h (4.27)
In the present work it is suggested to divide the tube length into
equal increments. The quality change across each increment of length
(∆z) is calculated to be (∆x),(xi –xe ). The heat transfer of steam in two

57
phase region depends on the quality change as a demonstrative factor to
describe the amount of heat released at constant temperature. After the
determination of the quality limits at each increment, it was utilized to
calculate the local condensing heat transfer coefficient at the midquality
magnitude of each ( x ). Assuming that the local value is constant over
particular quality range ( x ). Then:
1/ 2
m m meq i o
(4.28)
1/ 2 1 1l l
i o
m eq
x x (4.29)
1 l
eq
m eq
x (4.30)
where eqx represents the midquality which is equal:-
( )
2
i o
eq
x x
x (4.31)
So that equation (4.23) for modeling becomes:
1 l
c BK lo eqh h x (4.32)
At high condensing loads, with vapor shear dominating, tube
orientation has no effect, and equation (4.23) may also be utilized for
horizontal tubes, Perry (1999) [57].

58
4.5 Forced Convection Inside Tube:
Turbulent flow:
Numerous relations have been proposed for predicting fully
developed turbulent flow in uniform cross-section tubes. The Dittus and
Boelter correlation is suitable for moderate temperature variation.
Incropera and Dewitt (1996) [58].for the fluid flow the Dittus-Boelter is
usually given in the form:
0.8
0.023Re Prn f
i
k
h
D
(4.33)
Where Reynolds number and Prandtl number are estimated as follows:-
Re t i iu D GD
(4.34)
Pr
p
f
c
k
(4.35)
And 0.4n for heating ( Ts > Tm) and 0.3n for cooling ( Ts < Tm).
This mathematical relation has been confirmed experimentally for the
following ranges of conditions:
0.7 Pr 160
Re 10,000
/ 10
D
iL D
For the small condensing load, vertical tube condenser may be
maintained sub-cooling in the bottom end of the tube. For this condition,
if the temperature difference at the inlet and exit is greater than 10o
C,
then the moderate temperature variation assumption is invalid. However,
for flows characterized by large property variations, Sieder and Tate
correlation was used to calculate the heat transfer coefficient. Incropera
and Dewitt (2006) [59] recommends:
0.14
4/5 1/3
0.027Re Pr f
i s
k
h
D
(4.36)

59
0.7 Pr 16,700
Re 10,000
/ 10
D
iL D
Laminar flow:
Below a Reynolds number of about (2000), the flow in pipes will be
laminar. Sieder and Tate (1930) [60] recommends a simple correlation to
estimate the film heat transfer coefficient as:
0.33 0.140.33
1.86(RePr) /i sh D L (4.37)
where (L) is the length of the tube or conduits.
Water Correlation:
Special correlation for water could be used for more accurate
estimation for inside coefficient. Eagle and Ferguson (1930)[61],had
given data adapted by Sinnott (1999) [62] to develop specifically
correlation for water as follows:-
0.8
0.2
4200(1.35 0.02 ) t
i m
i
u
h t
D
(4.38)
Equation (4-38) is applicable for all range of Reynolds number.
However, if the estimated inside heat transfer coefficient by this equation
is large compared with another used correlation, section (4.5), the
smallest value will be taken.
For noncircular tubes all the above equations may be applied by
using an effective diameter as a characteristics length. It is termed as the
hydraulic diameter and is defined as:
4 c
h
A
D
P
(4.39)
where Ac and P are the flow cross sectional area and wetted perimeter.

60
4.6 Air Side Heat Transfer Coefficient:
Heat transfer performance of the tube bank is determined by flow
pattern, which is strongly dependent on the arrangement of the tubes. The
longitudinal tube spacing and transverse tube spacing could influence the
thermal characteristics performance of heat exchanger. This have been
studies by Grame [63] and Jones [64].On the other hand fin spacing, fin
thickness and fin height also affect the performance of finned tube heat
exchangers as reported by Briggs (1963) [65]. In the present work, the
used heat transfer coefficient for the triangular fins could be predicted
approximately using correlation of internal flow in ducts for laminar or
turbulent flow, which was described in section (4.5). Reynolds number
for air flow inside a triangle fin with using a hydraulic diameter approach
is calculated as below:
max
Re hu D
(4.40)
The velocity umax in eqn. (4.40) is calculated at the minimum cross flow
area Sm in tube bundle, that is:
.
max
a
m a
m
u
S
(4.41)
In the present work minimum cross flow area was calculated according to
the air side geometry as:
1
( )
2
m p T tS F X H (4.42)
. .
, /a inc a incm m H (4.43)

61
4.7 Shell Side Heat Transfer Coefficient for Single Phase:
Sometimes when the amount of vapor in the shell side of the vertical
condenser is low, the entire vapor quantity is condensed before the tube
length was reached. In this case, heat transfer occurs between condensed
vapor and cooling water inside tube, which lead to sub-cooling of the
condensate. Therefore, this section will deal with the method of how to
calculate shell side heat transfer coefficient for single phase as described
by Mohammed (2005)[26].
The procedure for calculating the shell side heat transfer coefficient
for a single shell pass exchanger is given below:
1. Calculate the bundle cross flow area, given by:
s
s t o
t
D
A P d B
P
(4.44)
2. Calculate the shell side equivalent diameter. For Square pitch:
2
2
4
4
o
t
e
o
d
P
D
d
(4.45)
3. Calculate the shell side mass flow rate per unit area by:
.
s
s
s
m
G
A
(4.46)
4. Calculate the Reynolds number for the shell-side, given by:
.
Re s e
s
s l
m D
A
(4.47)
5. Calculate the heat transfer coefficient:
0.55 1/3
0.36Re Pr s
o s s
e
k
h
D
(4.48)

62
4.8 NTU Effectiveness Relations:
The NTU method offers more advantages for analyzing heat
exchangers. It shows that iterative procedure is not required when inlet
and outlet temperatures are unknown. For any heat exchanger, the total
heat rejected from the hot fluid to the cold fluid is dependent on the heat
exchanger effectiveness and also on the heat capacity of each fluid. This
can be calculated as follows:-
m i n , ,h i c iQ c T T (4.49)
The heat capacity, c, the extensive equivalent of the specific heat,
determines the amount of heat a substance absorbs or rejects per unit
temperature change, where:-
c mcp (4.50)
The effectiveness is the ratio of the actual amount of heat transferred to
the maximum possible amount of heat transferred and defined as:
maxQ
Q
(4.51)
In the present work, the heat duty of each increment is calculated by the
effectiveness-NTU method. The effectiveness relation for single-phase
fluid cross flow is given below, Holman (2002) [66]:-
exp( ) 1
1 exp r
r
NC n
C n
(4.52-a)
And for Counter flow:
1 exp (1 )
1 exp (1 )
r
r r
N C
C N C
(4.52-b)
where:
0.22
n N (4.53)
min
max
r
c
C
c
(4.54)

63
min
UA
N NTU
c
(4.55)
This is the effectiveness relationship for a cross flow, single-pass heat
exchanger with both fluid unmixed. For tow-phase fluid flow, the
effectiveness relation is:
1 exp( )NTU (4.56)
The overall heat transfer coefficient (U), takes into consideration the total
thermal resistance to heat transfer between two fluids. Even though the
convective heat transfer coefficients may be different on the two sides of
the heat exchanger, where the (UA) product is the same on either side.
This is because all of heat taken from hot side must be transferred to the
cold side. Overall heat transfer coefficient defined as, Sinnott (1999)
[62]:-
, ,
ln( / )1 1 1 1 1
2
o o i o o
o o o fo o o wall i fo i i i i i
D D D A A
U h h k A h A h
(4.57)
By neglecting effect of fouling on both sides of the heat exchanger, and
inside surface efficiency, the overall heat transfer coefficient is reduced
to:-
ln( / )1 1 1
2
o o i o
o o o wall i i
D D D A
U h k A h
(4.58)

64
4.9 Air Cooled Steam Condenser Modeling:
The air cooled steam condenser is shown in figure (4-1).The
steam flows inside the flat tube. The air flow is perpendicular to the tubes
across the fins so that both process and service fluids pass in a cross flow
pattern on both sides of the exchanger wall. Row by row technique was
used and each row in the single pass condenser is divided into segments.
Each segment is treated as a small condenser with a specified geometry.
Each tube in small condenser is associated with steam parameters,
specific air mass flow rate and inlet air temperature. The flow chart of the
modeling programs are presented in appendix (E).The assumptions of the
ACC modeling are:
1. The air mass flow rate is assumed to be distributed uniformly over
the whole face of the air cooled steam condenser. To insure this
assumption, the inlet duct configuration and mal-distribution was
checked.
2. Homogenous temperature distribution of air all over the frontal
face area of the heat exchanger and hence for each segment.
3. The mean air exit temperature of each row is considered to be the
inlet to the next row.
4. The steam temperature variations between the rows were assumed
to be negligible.
5. The inlet air velocity for each row was assumed to be uniform and
is represented by a measured value.
6. The steam mass flow rate from the main header is divided equally
for each row.
7. Heat transferred away by convection and radiation is neglected.
8. Material properties of the finned tube bundle are constant with
temperature variations.

65
4.9.1 Air Cooled Steam Condenser Geometry:
4.9.1.1 Tube Geometry:
The cross section of the ACC tube is as shown in figure (4-2.a).The
sides of the tube are assumed to be semicircular with a diameter equal to
the tube height. Therefore, the tube cross-sectional area is calculated as
follows:-
2
, , , ,( )
4
t t i t i t i t iA H W H H (4.59)
The tube wetted perimeter is the total tube perimeter in contact with
the steam. It is given by:
, , ,2( )t t i t i t iP H W H (4.60)
The hydraulic diameter is given by:
,
4 t
h t
t
A
D
P
(4.61)
The tube walls wetted by direct contact with steam are transfer heat
directly from the steam to the outside air. This constitutes the direct heat
transfer area:
, , ,2( )s t i t i t t i tA W H L H L (4.62)
Each fin attached to the tube with approximately (1 mm) thickness
along the depth. The contact area is calculated as:
att att tu b eA t D (4.63)
The total heat transfer area is given by:
,o t s attA A A (4.64)

66
5.9.1.2 Air Side Geometry:
The cross section of the tube outer surface is shown in figure (4-2.b).
The face area of one tube and fin set in one increment, figure (4-2.c) is
defined as:
, ,( )fa inc tube o h incA H F H (4.65)
Therefore, the area blocked by the fins is given by:
inc
ba f f f f
p
H
A l t l t
F
(4.66)
The area available for air flow is the total area less than the area blocked
by the fins together with the area occupied by the tube for steam flow as
follows:-
, ,( )a fa inc ba tube o incA A A H H (4.67)
The perimeter of the tube which is directly in contact with air is given
by:-
2 ( )inc f
a f f inc inc
p p
H t
P l t H H
F F
(4.68)
The hydraulic diameter is:
,
4 a
h a
a
A
D
P
(4.69)
The tube wall which is in contact with steam on the inside surface and
with air on the outside surface directly transfers heat from the steam to
the outside air. This constitutes the heat transfer area:
, , , ,2( ) 1 f
a s t o t o t o inc
p
t
A D H H H
F
(4.70)
The fins also confirm the heat transfer from the steam to air. Therefore fin
surface area is defined as follows:-

67
, 2 inc
f s f f
p
H
A l D
F
(4.71)
The total heat transfer area become:
, ,o a s f f sA A A (4.72)
Where ηf is the fin efficiency, when ( 0,
dT
x l
dx
) efficiency given by,
Kreith (1999) [67], figure (4-3):
tanh
f
ml
ml
(4.73)
a a
a a
h P
m
k A
(4.74)
Therefore, the total surface efficiency of the fin, ηo is therefore expressed
as below:
1 1f
o f
o
A
A
(4.75)

68
4.9.2 Computational Procedure:
In the present work model, there are two separate sections. These are
generally the condensing section and the sub-cooled section. The
subcooler section is treated by using single-phase flow correlations. In the
condensing section and since the local heat transfer coefficient varies
along the tube axis, a segmentation technique is suggested to be used in
the present work, as in figure (4-1). For the present work, the following
procedures in describing the calculation technique of the model are given
as follows:-
1. Input condenser operating conditions:
i. Steam side (Pcond ,Tcond ,
.
stm ).
ii. Air side (Ti ,
.
airm , airu ).
2. Input Surface Characteristics:
i. Tube geometry: tube height (Ht), tube depth (Dt), tube
thickness (tt), transverse tube pitch (XT), longitudinal tube
pitch (XL), as shown in table (3-1).
ii. Fin details: fin length (lf), fin depth (Df), fin pitch (fp), fin
thickness (tf), as shown in table (3-1).
iii. Core dimension: height (H), width (W), depth (D).
3. Calculate ( , ,f t oA A A ) using equations in section (4.9.1).
4. Estimate the inlet enthalpy of steam if the assumed quality chosen
is not to be at saturated state, by using properties equations in
Appendix (D) . A usual form for mixed region is shown below:
= 1 f gx x (4.76)
5. Set the main loop for the rows number 1...... ... rJ To N .
6. Specify the steam mass flow rate for each row (
.
.
. /stst row trm m N ).

69
7. Determine the number of horizontal increments dividing each row
of the single pass tubes.
8. Set the secondary loop for the horizontal increments number in
each row ( 1...... ..... incI To N ).
9. for each increment:
i. Assume exit air temperature ( aeT ).
ii. Assume exit steam side enthalpy ( ,inc eh ).
iii. Assume tube wall temperature ( wallT ).
10. Estimate the initial set of both air and steam sides properties,
Appendix (D).
11. From the assumed value of exit steam side enthalpy (step 9)
Calculate the exit steam quality from the relation:
e l
e
g l
h h
x
h h
(4.77)
12. Calculate the midquality value, Equation (4.31).
13. Estimate on the steam side liquid only heat transfer coefficient
according to the flow condition from equations ( 4.33, 4.36 and
4.37).It is conservative to use equation (4.26) to compute the value
of liquid only heat transfer coefficient as Boyko and Kruzhilin
recommended. And condensation heat transfer coefficient eq.(4.32).
14. Estimate the air mass flow rate for each increment, equation
(4.43).
15. Determine the minimum flow cross section area, equation (4.42)
to calculate the maximum velocity equations (4.41), and Reynolds
number, equation (4.40).
16. Estimate the air side heat transfer coefficient section from
eq.(4.33) when the flow is turbulent or from eq.(4.37) for laminar
flow condition.

70
17. Calculate the fin efficiency, over all efficiency, and design
(fouled) overall heat transfer coefficient ( dU ), or clean overall heat
transfer coefficient ( cU ). Using equations, (4.73, 4.75, 4.57 and
4.58).
18. Calculate the exact value of the wall temperature, used the
relation, Perry [57]
:
o
wall c
U
T T T T
h
(4.78)
Wall temperature can be calculated for inside and outside
surfaces according to each heat transfer coefficient (h) calculated for
inside or outside streams. It has been assumed that no variation
occurs across two sides due to the high thermal conductivity of the
tube wall material.
19. Check the error difference between the assumed and the
calculated value by using:
0.1%calculated assumed
calculated
ABS (4.79)
Where ( ) represent any variable need to check. If the check is not
satisfying then let ( assumed calculated ) and update the solution from
step 13.
20. Estimate minimum heat capacity, and determine the number of
transfer unit (NTU) and the effectiveness (ε) by equations (4.50,
4.55, and 4.56).
21. Calculate the heat duty using equation (4.49).
22. Calculate the exact value of the exit specific enthalpy as follows:
, .e cal i
st
Q
h h
m
(4.80)

71
Compare the calculated value of enthalpy with that assumed.
Then check the error difference between the values as in step (19),
update for new value of exit enthalpy if the check fails to be
satisfied.
23. For the sub-cooling zone, correct the value of the heat duty by
calculating a new value of (ε) and update ( minC ), using equations
(4.50, 4.55, 4.52-a, and 4.49).
24. If satT T , L coreH , then after calculation the (Q ) determine
the exit condensate temperature as:
.exit in
p
Q
T T
mc
(Single-phase) (4.81)
25. Calculate the exact value of the air exit temperature as shown
below:
.ae ai
a pa
Q
T T
m c
(4.82)
Compare the calculated value of air exit temperature with that
assumed, at step (9), and check the error difference between tow
values as in step (19).Update for new value of exit air temperature if
the error percent was failed to be satisfied.
26. When the end criteria achieved (total length, increments counter)
stop the program and record the final simulated data (hc , ha , Uo , Rel
xeq , Tw,Tae and Q), see the program flow chart in Appendix (E).

72
4.10 Water Cooled Steam Condenser Modeling:
Counter-Current single pass shell and tube heat exchanger was used
in the present work. The shell and tube condenser is oriented vertically
and on the shell side where condensation occurs. In this type of unit, the
condensation coefficient depends more on the vapor velocity than on the
film thickness. It is a conservative assumption that is to calculate the
liquid film resistance as if the baffles is not present, Cao (2010) [68]. In a
vertical tube bundle, the presence of one or more tube does not alter the
heat transfer coefficient, Kern (1950) [69]. The flow charts of the
modeling programs were presented in appendix (E).The assumptions of
the WCC modeling which are taken into account are as follows:-
1. Saturated steam at condenser shell side inlet according to the steam
generator design criteria.
2. Each tube in prototype tube bundle carrying an equal heat load.
3. The water mass flow rate is assumed to be distributed uniformly
between the tubes in tube bundle.
4. Each increment has a temperature rise according to the
condensation condition outside the tube. Therefore, more rise of
temperature occurs near the water exit end.
5. The steam mass flow rate from inlet side is distributed uniformly
through the vertical tubes bundle.
6. Material properties of the tubes bundle and shell are constant with
temperature variation of both streams.

Ali Farhan Thesis

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