WATER POLLUTION CONTROL RESEARCH SERIES • 16T30 DHS 07/69
A SURVEY OF ALTERNATE METHODS
FOR COOLING CONDENSER DISCHARGE
WATER
LARGE-SCALE HEAT
REJECTION EQUIPMENT
ENVIRONMENTAL PROTECTION AGENCY • WATER QUALITY OFFICE
-------
WATER POLLUTION CONTROL RESEARCH SERIES
The Water Pollution Control Research Series describes
the results and progress in the control and abatement
of pollution in our Nation’s waters. They provide a
central source of information on the research, develop-
ment, and demonstration activities in the Water Quality
Office, Environmental Protection Agency, through inhouse
research and grants and contracts with Federal, State,
and local agencies, research institutions, and industrial
organizations.
Inquiries pertaining to Water Pollution Control Research
Reports should be directed to the Head, Project Reports
System, Office of Research and Develop ient, Water Quality
Office, Environr cntal ?retection Agency, Roo i llO ,
Washington, . C. 2O2 2
-------
A SURVEY OF ALTERNATE METHODS
FOR COOLING CONDENSER DISCHARGE WATER
LARGE-SCALE HEAT REJECTION EQUIPMENT
by
DYNATECH R/D COMPANY
A Division of Dynatech Corporation
Cambridge, Massachusetts 02139
for
WATER QUALITY OFFICE
ENVIRONMENTAL PROTECTION AGENCY
Project No. 16130 DHS
July, 1969
For sale by the Superintendent of Documents, U.S. Government Printing Office, Washington, B.C. 20402 - Price $1.25
Stock Number 5501-0126
-------
EPA Review Notice
This report has been reviewed by the Water
Quality Office, EPA, and approved for publication.
Approval does not signify that the contents
necessarily reflect the views and policies of
the Environmental Protection Agency, nor does
mention of trade names or commercial products
constitute endorsement or recommendation for
use.
-------
TABLE OF CONTENTS
Section Page
INTRODUCTION 1
1.1 ScopeofTaski 1
1.2 Format of Task 1 Report I
1. 2. 1 Section 2 - Thermodynamic Aspects of the
Steam Power Cycle 1
1.2.2 Section 3-Heat Rejection Equipment -
Technical Considerations 2
1.2.3 Section 4 - Heat Rejection Equipment -
Operating Considerations 2
1.2.4 Section 5 - Methods of Comparison and
Selection 2
1.2.5 Section 6 - Potential Extensions of the
State-of-the-Art 2
2 THERMODYNAMIC ASPECTS OF THE STEAM POWER CYCLE 3
2. 1 Basic Steam Cycle 3
2. 2 Thermodynamic Efficiency ( th)
2. 3 Methods of Decreasing Heat Rejection (Increasing
2. 3. 1 Lowering the Condenser Pressure 7
2.3.2 Increasing the Boiler Pressure 9
2.3. 3 Increasing the Amount of Superheat 10
2. 3. 4 Reheat Cycle 10
2. 3. 5 Regenerative Cycle 11
2. 3. 6 Supercritical and Binary Cycles 13
3 HEAT REJECTION EQUIPMENT - TECHNICAL CON-
SIDERATIONS 15
3. 1 Introduction 15
3. 2 Once-Through Cooling 16
3. 3 Cooling Ponds 22
3.3. 1 Completely Mixed Ponds 24
3.3.2 Flow-Through Ponds 25
3.3. 3 Internally Circulating Ponds 26
3.3.4 Sizing of Cooling Ponds 26
3.4 Spray Ponds 28
3. 5 Wet Cooling Towers 30
3. 5. 1 Introduction 30
3.5.2 Atmospheric Wet Cooling Towers 31
3. 5. 3. 1 Natural Draft - Counterfiow 35
3.5. 3.2 Natural Draft - Crossflow 35
-------
Section Page
3.5.4 Mechanical Draft (MD) Wet Cooling Towers 36
3. 5.4. 1 Mechanical-Induced Draft-
Counterflow 37
3. 5.4. 2 Mechanical-Induced Draft-
Crossflow 38
3. 5.4. 3 Mechanical Forced Draft 40
3. 5. 5 Cooling Tower Theory 41
3. 5. 5. 1 Derivation of Basic Equations 41
3. 5. 5. 2 Counterfiow Tower 49
3.5.5.3 Crossflow Tower 51
4. 5. 5. 4 Discussion and Summary .51
3. 5. 6 Size Estimation 51
3. 5. 7 Cost Information 54
3.5. 7.1 ND Towers - Capital Cost 56
3. 5. 7. 2 MD Towers - Capital Cost 58
3. 5. 8 Maintenance Considerations 62
3. 5. 9 Secondary Pollution Considerations 64
3. 6 Dry Cooling Towers 67
3. 6. 1 Introduction 67
3.6.2 Description 67
3. 6. 3 Theory of Operation 69
3.6.4 Size Estimation 71
3. 6. 5 Cost Information 72
3.6. 6 Maintenance Requirements 76
3. 6. 7 Secondary Pollutant Considerations 76
3. 7 Evaporative Condensers 76
3. 7. 1 Introduction 76
3. 7. 2 General Description 77
3. 7. 3 Theory of Operation 79
3.7.4 Size Estimation 82
3.7.5 Cost Information 82
3. 7. 6 Maintenance Considerations 82
3. 7. 7 Secondary Pollution Considerations 84
3. 8 Air Cooled Condensers 84
3.8.1 Introduction 84
3. 8. 2 General Description 84
3. 8. 3 Theory of Operation 86
3. 8. 4 Size Estimation 87
3. 8. 5 Cost Information 88
3. 8. 6 Maintenance Requirements 89
3. 8. 7 Seconda Pollution Considerations 89
IV
-------
Section Page
3.9 Advanced Concepts 90
3.9.1 Introduction 90
3.9.2 The Froth Contact Heat Exchanger 90
3. 9. 3 Ellipsoidal Cooling Tower 91
3. 9.4 Rotating Arm Cooling Tower 92
3. 9. 5 Rotating Cooling Tower 93
3. 9.6 Powered Hyperbolic Towers 94
3. 9. 7 New Materials 96
4 HEAT REJECTION EQUIPMENT - OPERATIONAL
CONSIDERATIONS 103
4.1 Introduction 103
4. 2 Once-Through Cooling and Cooling Ponds 103
4.3 Spray Ponds 104
4.4 Wet Cooling Towers 105
4.4. 1 Off-Design Performance 105
4.4.2 Unusual Characteristics 107
4. 5 Dry Cooling Towers 108
4. 5. 1 Off—Design Operation 108
4.5.2 Operating Procedure 108
4.5.3 Unusual Characteristics 109
4.6 Evaporative Condensers 110
4.6.1 Off-Design Performance 110
4.6.2 Operational Procedures 112
4. 6. 3 Unusual Characteristics 112
4. 7 Air Cooled Condensers 113
4.7.1 Operating Procedures 113
4. 7. 2 Off-Design Operation 113
4. 7. 3 Unusual Characteristics 113
5 METHODS OF COMPARISON AND SELECTION 116
5.1 Introduction 116
5.2 Economics 116
5. 3 Size and Capacity 124
5. 4 Geographic Limitations 124
5.5 Availability 125
6 SUMMARY OF TASK I 127
6. 1 Present Status 12
6. 2 Relationship with Other Tasks 127
-------
Section 1
INTRODUCTION
This report presents the results of Task I, Phase I of an overall program
to determine the technologicai and economic consequences of alternative methods of
cooling condenser discharge water from thermal power plants.
1. 1 Scope of Task I
The first task of this study, which is reported on in this document, is con-
cerned with various schemes for heat rejection on a large scale. A large part of the
problem in this area is attributable sin ply to the sheer magnitude of the quantities of
heat which must be rejected. In torder_of_magnitude! terms, a 500 megawatt plant
with an efficiency of —25% must reject approximately five billion Btui’hr (5. 0 x i0 )
of heat to its surroundings. In terms used in conventional cooling equipment, this
represents 400, 000 tons of cooling (1 ton = 12, 000 Btu/hr). (For purposes of com-
parison, the capacity of the air-conditioning system in Boston’s Prudential Center, a
fifty-two story office building, is approximately 6, 000 tons.) Clearly, the difficulties
of extrapolating performance data on existing equipment are severe. In this task, in-
formation is presented to identify present state-of-the-art methods of large scale heat
rejection and to determine their applicability to this problem.
1. 2 Format of Task I Report
This report is divided into six sections plus a bibliography. This first
section serves both to define the scope of this task and where it fits into the overall
program goals and also to provide a “road-map through the report itself. The follow-
ing paragraphs give a brief synopsis of each of the remaining sections.
1. 2. 1 Section 2 - Thermodynamic Aspects of the Steam Power Cycle
In Section 2, a brief review of the thermodynamics of power generation is
presented. This is not intended to be comprehensive nor to deal with the practical
1
-------
aspects of real systems. It serves simply to provide some understanding of how the
heat rejection requirement affects the total power cycle, and to set some fundamental
thermodynamic limits on the nature of the solution. It will further define much of
the nomenclature for the remainder of the discussion. A detailed treatment of the
practical aspects of power generation will be done in Task II.
1.2.2 Section 3 - Heat Rejection Equipment - Technical Considerations
In Section 3, information is presented on all heat rejection schemes which
may be considered candidates for use with large power plants. These include once-
throughcoolin coolingponds, spray ponds, wet cooling towers, dry cooling towers,
evaporative condensers, air-cooled condensers, and some advanced concepts. The
material presented includes sizing procedures, capital and operating cost data.
maintenance, and secondary pollution considerations.
1. 2. 3 Section 4 - Heat Rejection Equipment - Operating Considerations
Section 4 is organized in the same way as Section 3 and concerns the same
types of equipment. It deals, however, with the operating considerations such as theory
of operation at design, off-design operation, running procedures, and any unusual char-
acteristics affecting the interface of the cooling system with the power plant.
1. 2.4 Section 5 - Methods of Comparison and Selection
In Section 5, the material describing the alternative cooling methods is
integrated and criteria are developed for the selection of a particular unit in terms
of geographic location, capacity requirements, and economic factors.
I. 2. 5 Section 6 - Potential Extensions of the State-of-the-Art
In Section 6. some preliminary thoughts are presented to serve as a basis
r the work of Task II, Phase II which will deal with advanced concepts in large-scale
heat rejection equipment.
2
-------
Section 2
THERMODYNAMIC ASPECTS OF THE STEAM POWER CYCLE
2.1 Basic Steam Cycle
A block diagram of the basic Rankine steam power cycle is presented in
Figure 2.1. All existing steam power plants are variations of this basic cycle. A
particular plant may have a number of turbines with steam being reheated between
turbines or it may use a multiple stage turbine with a portion of the steam being with-
drawn between stages to preheat the water before it re-enters the boiler; there are
many variations which may be employed to increase the cycle efficiency above that
of the basic steam cycle. However, the performance of the basic Rankine cycle is
representative of all steam power systems.
Heat is transferred to the boiling fluid at a constant pressure either from
a burning combustible fuel or from a thermonuclear reaction, in order to provide steam
at a high pressure and temperature at location ®. The steam is then expanded through
a turbine to generate the output power from the cycle. The turbine is generally connected
to an electrical generator to provide electrical power. At location J the steam has a
lower temperature and a much lower pressure. In the condenser, heat is rejected from
the cycle to an external heat sink at a constant pressure, causing the steam to condense.
Thus, all the steam has been converted to water at location ® . Between locations J
and ® a pump raises the pressure of the water to that of the boiler and moves the water
into the boiler to complete the cycle.
An excellent way to show the thermodynamic processes occurring in the basic
Rankine cycle is the temperature-entropy diagram of the working fluid. (See Figure 2.2
The solid lines in Figure 2.2 indicate the basic Rankine cycle. The pumping
(3-4) and the expansion in the turbine (1—2) are shown as reversible adiabatic processes
for simplicity. It is conventional to install a superheater between the boiler and the tur-
bine. This moves the point (1) into the superheat region point 1’ so that the end of the
expansion is now 2’, a point of high quality. Some water droplets are present but in
tolerable amounts. The basic power plant Rankine cycle is shown by the points 3-4-1’-
2’—3 in Figure 2.2.
3
-------
Electricity Out
Turbine
Boiler
Heat In
Heat Rejected
Work In
Figure 2.1. Schematic of Basic Rankine Power Cycle
-------
T
— basic cycle
superheat cycle
S
Figure 2.2. Temperature-Entropy Diagrams of Rankine Cycles
2.2 Thermodynamic Efficiency (‘ h )
The efficiency of a steam power cycle is defined as the net work output
divided by the heat input to the system:
w . -w
turbine pump
th Q -
IN
Conservation of energy requires that the equation below be satisfied
Q + W — Q . - W . - P (Plant Losses) = 0 (2-2)
IN pump rej turbine Loss
Equations (2-1) and (2-2) may be combined and written as follows;
rej = ( - Wturbine - ( th - 1) Wpump - FLoss (2-3)
Liquid
p = const.
Vapor
5
-------
The second term in Equation (2-3) is very small, since the pump power represents a
negligible portion of the turbine power. Therefore, the balance between turbine power
and heat rejected may be approximated by the following relation:
rej = ( _ ) - (2-4)
For a plant which must generate a given amount of power, it is clear that any decrease
in thermodynamic efficiency is accompanied by an increase in rejected heat.
The thermodynamic efficiency can also be written in terms of the thermo-
dynamic properties.
IN - heat added in the steam generator
1N = th(h 1 ,-h 4 )
W . - work extracted from the turbine
turbine
Wt
- work required to pump the liquid
W = th(h -h)
p 4 3
rej - heat required to condense the exit steam from the turbine
rej = th (h - h 3 )
where m = flow rate of water through the system
h 1 = enthalpy of the water at various stations.
Thus, the expression for efficiency can be rewritten as follows:
6
-------
( h 1 , - h 2 ,) — (h 4 - h 3
th (h 1 ,-h 4 ) -
or ( h 1 , - h 4 ) - (h 2 , - h 3 )
1 7 th = (h 1 , — h 4 ) (2-5a)
2. 3 Methods of Decreasing Heat Rejection (Increasing 17th )
Any decrease in the amount of heat rejected resulting from internal power
plant characteristics is obtained from altering the thermodynamic power cycle. The
main area of concern to power plant designs in the past has been increasing the thermal
efficiency, since this change decreases the cost per kwh. The present state of the art in
power plant design indicates that sizable gains in efficiency are a thing of the past.
In this section, methods currently used in power plants to increase the
thermal efficiency (thus to decrease as shown in Equation (2-4) are discussed
and equations for computing 17 th are presented.
2.3.1 Lowering the Condenser Pressure
The thermal efficiency of a plant is largely controlled by the pressure at
which condensation occurs in the condenser. In a two-phase mixture, only one thermo-
dynamic property is independent. Usually the temperature or the pressure is specified
and the other properties are determined by the saturation line of the fluid. As a result,
when the temperature of the condenser is decreased, the corresponding pressure will
also decrease. The highest cycle efficiency is obtained by operating with the condenser
pressure as low as possible. Calculations of thermal efficiency and heat rejection rate
have been performed for a hypothetical steam power plant which generates power at the
rate of 120Mw. The boiler provides steam at a pressure of 1500 lbf/in 2 and a tempera-
tureof 1000°F and the turbine efficiency is 80%. Figure 2.3 illustrates the values of thermo-
dynamic efficiency and heat rejection rate for this plant using a range of values of condenser
pressure and temperature while maintaining a constant power output.
7
-------
--- -
-
- ‘zoo
- 1,0
- /80
- 170
- I 0
- gSO
- l4O
- 130
- /20
hO
35
34
JO’-
z8
z6
- - -.— - I
0 2 3 4 .5 6 7 8 9 ia I I Z ‘3 s4 iS
Condenser Pressure, psia
_____ I - . . ... - . . . - L _
i 6 153 !70 ‘83 l 2OZ ZC ’ z 16
Condenser ‘l’emperature, ° F
Figure 2. 3. Thermodynamic Efficiency and heat Rejection Rate As
A Function of Condenser Pressure
C)
Q )
C.)
‘. 1-4
c x i
0
I
CD
CD
CD
C)
CD
0
‘1
CD
w
-------
The effect of condenser temperature variations upon the condenser heat
rejection rate is clear from Figure 2.3. The condenser temperature is determined
by the rate at which heat can be removed from the condenser to a heat sink or cooling
source. The method presently employed in many power plants is to draw water from
a nearby river and use this to cool the condenser coils, thereby transferring heat from
the condenser to the diverted river water. Regardless of the method chosen to remove
beat from the condenser, the ability to reject this heat is an essential part of the steam
power cycle and is a vital factor in determining the overall efficiency of the power plants
as well as the amount of heat rejected and the temperature at which it is transferred to
the heat sink.
2. 3. 2 Increasing the Boiler Pressure
In Figure 2.4 a Rankine cycle is shown with two upper pressure levels p 1
Figure 2.4. Effect of Increased Boiler Pressure
on Power Cycle
S
and p 2
7-
I
4 ,
4
p 1 > p 2
cond
9
-------
The efficiency of a cycle operating at p 2 is:
( h 1 - 112) - (h 4 - h 3 )
tb = (h 1 - h 4 ) (2-7)
at p 1
( h 1 , — h 2 ,) — (h 4 , — h 3 )
th = (h 1 , - h 4 ,) (2-8)
The numerator (net work) of both equations remains approximately con-
stant, while the denominator (the heat added at the steam generator) is smaller at
the higher pressure p 1 . Thus at constant power the amount of heat rejected at the
condenser would decrease as the pressure increased.
2. 3. 3 Increasing the Amount of Superheat
An increase in the superheat in the boiler, increases the work obtainable
from the turbine while it also increases the heat added in the boiler. However the
percentage increase in net work is greater than the percentage increase in additional
heat required since the heat is being added at a higher average temperature. Thus the
overall effect is an increase in the efficiency of the cycle. At constant power level,
this results in less heat rejected.
2. 3.4 Reheat Cycle
The reheat cycle has been developed to take advantage of the increased
efficiency with higher pressures, and yet avoid excessive moisture in the low pres-
sure stages of the turbine, without going to excessive turbine inlet temperatures.
In this cycle steam is extracted from the turbine after being partially
expanded and is reheated in the boiler before it is expanded further in the turbine.
This cycle is shown on a T-S diagram with solid lines in Figure 2. 5
10
-------
7 -
Figure 2. 5. Temperature-Entropy Diagram of’Reheat Cycle
The thermal efficiency of this process is given by:
( h 5 —h 6 )+(h 1 --h 2 )-(h 4 -h 3 )
1 th (h 5 - h 4 ) + (h 1 - h 6 ) (2-9)
This modification is not a primary improvement on the cycle. If superheat
were added up to point 1’, as shown in Figure 2. 5, and no reheating was done, the ther-
mal efficiency for the simple cycle (dashed lines) would be greater than that indicated
by Equation (2—9). The reason the reheat cycle is used is to overcome equipment
limitations which would not allow operation at temperatures existing at point 1’.
2. 3. 5 Regenerative Cycle
This is a modification to the Rankine cycle which extracts steam from some
intermediate point in the turbine and heats the water entering the boiler thus decreasing
the amount of heat added at the boiler, The net work of the cycle is also decreased since
i
4
3
a
5
11
-------
all the steam is not expanded in the turbine. However, since the average temperature
at which heat is added to the system from outside is increased, the net effect is an
increase in thermal efficiency of the cycle and a subsequent reduction of the heat re-
jected at the condenser. There are two types of feedwater heaters: open and closed.
The open feedwater heater has the steam from the turbine and the liquid from the
condenser in actual contact with each other. The closed feedwater heater keeps
the two fluids separate while the transfer of heat takes place.
Open feedwater heaters have the advantage of being less expensive coupled
with the disadvantage of requiring a pump to handle the feedwater between each heater.
Figure 2.6 shows a schematic of a cycle with one open feedwater heater
and Figure 2. 7 shows the accompanying T-S diagram for the cycle.
Figure 2. 6. Schematic of Regenerative Cycle
EJ
12
-------
T
S
Figure 2. 7. Diagram of Regenerative Cycle
The expression for the thermal efficiency for this system is:
th
where mT = total mass flow rate
= mass flow extracted from the turbine
2.3.6 Supereritical and Binary Cycles
Two other cycle modifications have been used in modern power plants;
supercriticai steam generation and binary cycles (mercury-steam). They have not
been applied to large nuclear installations for the following reasons. The pressures
and temperatures encountered in a supercritical cycle (10000 F and 3200 psia) exceed
those permitted in a water cooled reactor. Typically, conditions in nuclear plants run
about 600° F and 1000 psia for boiling water reactors and 600°F and 2000 psia for pres-
surized water reactors.
4.
6
2:
- h 3 )
13
-------
The binary-vapor cycle has been used as a “topping unit’ 1 in small plants
in an attempt to overcome some undesirable thermodynamic characteristics of water:
(1) the high specific heat of the liquid phase; (2) low latent heat of evaporation at
high pressures, and (3) high boiler pressures. However, the differences in latent heat
are such that the mass flow rate of mercury must be 7 - 10 times that of the steam.
The result is that the properties which determine the plant size and cost outweigh the
thermodynamic attractiveness.
REFERENCES - Section 2
The following references contain general treatments of the thermodynamic
aspects of steam power cycles. Task 2 will examine the design and operation of power
plants in detail.
2. 1 Sorenson, Harry A., Principles of Thermodynamics , Holt Rinehart
and Winston, New York, 1961.
2.2 Keenan, Joseph H., Thermodynamics , John Wiley & Sons, Inc.,
New York, 1941.
2. 3 Mooney, David A., Mechanical Engineering Thermodynamics ,
Prentice-Hall, Inc. , Englewood Cliffs, New Jersey, 1953.
14
-------
Section 3
HEAT REJECTION EQUIPMENT—TECHMCAL CONSIDERATIONS
3. 1 Introduction
It has been established in Section 2 that heat rejection at the low temperature
of the steam power cycle is necessary to its operation. To transfer this heat from the
condenser a driving potential either in the form of temperature or enthalpy (for evapora-
tive systems) must be present. If the immediate sink is of finite size, it will increase
in temperature until heat can no longer be transferred unless an equal amount of heat is
also removed from the sink. The rejected heat is thus transferred from one sink to
another. From this point of view, deep space is the ultimate heat sink. For example,
in run-of-river cooling, the rejected heat follows a path from the condenser through the
river water to the atmosphere and is then radiated to deep space. The temperature of
the river and/or the atmosphere may increase in the process.
Present heat rejection systems transfer the heat from the condensing steam
to either water which is eventually cooled in the atmosphere or directly to the air.
The heat may be transferred by two modes; latent and sensible. Latent heat
transfer, or more commonly evaporation, occurs at constant temperature and involves
the amount of heat (energy) used in transforming the water from liquid to the vapor state.
The amount of energy is called the latent heat of vaporization and is taken from the re-
maining water, lowering its temperature. Although evaporation is usually called heat
transfer, it is actually a removal of energy by mass transfer. The difference between
the partial pressure of water in the air and the vapor pressure of the water is the driving
potential.
The sensible heat transfer occurs due to a difference in water temperature
and air dry-bulb temperature. As sensible heat is transferred, the air heats up and the
water cools down. Often, both modes operate simultaneously with approximately 75%
of the total heat transferred by evaporation and the remaining 25% by sensible heat
transfer in cooling towers, for example.
The following heat rejection systems have been considered as candidates
for use with large power generation systems:
15
-------
Once-through cooling
Cooling ponds
• Spray ponds
• Cooling towers (wet and dry)
• Evaporative condensers
• Air cooled condensers
A discussion of several advanced systems not yet in use has also been included.
3.2 Once-Through Cooling
Once-through cooling is the most common method for removal of heat from
power plant condensers in the United States. In this system cool water is diverted from
a river or estuary and pumped through the condenser tubes where it picks up the latent
heat of the steam condensing on the outer surface of the tubes. The warmed water is
then discharged to the river at a point where recirculation of the warm water will not
occur. No attempt at cooling the warm water prior to discharge is made in contrast
to the other systems considered in this report.
Once-through cooling is the least expensive method, provided a suitable site
can be found. Capital and installation costs are usually low although the sea water in-
stallations require a slightly more expensive condenser and longer piping. Operating
and maintenance costs are also low. However, various forms of aquatic life are forced
through the condenser tubes often resulting in high mortality rates and the discharged
water is 10 to 20 degrees warmer than the inlet. This sudden change in temperature
upsets the ecology of the water lowering the oxygen content, increasing the metabolic
rate of the fish and enhancing growth of undesirable bacteria. Many miles are required
before the water temperature returns to normal due to the very slow cooling process.
The temperature increase of the cooling water through the condenser may
be calculated from:
heat rejected to water, (Btu/hr)
water flow rate through condenser, (ibm /hr
16
-------
The specific heat of water is 1. 0 Bti /lbm_° F and has been omitted in the
equation. This warm water is then discharged to a region of cool water. If the warm
and cool waters mix thoroughly, the temperature will decrease quickly to a mixed
temperature given by
= Tp÷T
mixed cool
p is the percentage of total water flow pumped through the condenser.
On the other hand, thermal stratification may occur, that is, the warm
water may simply form a layer on top of the denser cool water and very little mixing
between the two layers will occur. Criteria for the establishment and stability of ther-
mally stratified flows have not been fully determined, but the placement of the discharge,
the flow patterns of the stream or estuary which govern the advection, and the local fluid
velocities and channel characteristics which govern the turbulent mixing of the layers
seem to be the more important parameters. The shape and extent of the “thermal plume’
often observed at the discharge of plants located on estuaries can be greatly altered by
the tidal stage. Movements of the plume upstream or spreading out across the stream
followed later by a return to a plume that hugs the bank near the discharge have been
observed. Almost complete mixing due to wave action is observed when the warm water
is discharged to a bay.
Once the warm water has been discharged, the heat must be dissipated so
that the temperatures will return to “normal”. The net rate at which heat is transferred
across the water surface is determinated by the algebraic sum of seven quantities:
1. Short-wave solar radiation directly from the sun.
2. Long-wave atmospheric radiation primarily from water
vapor in the upper atmosphere.
3. Reflected solar radiation.
4. Reflected atmospheric radiation.
5. Long—wave back radiation from the water to
the atmosphere.
17
-------
6. Evaporation of water into the air above the water; and
7. Conduction from the water to the air and surrounding
cooler regions of water and land.
A complete description of these phenomena is given in Reference 3. 1.
Neglecting the conduction to the surrounding water and land, we can write
dH
dt = _K(Tw TE)
where dH/dt net rate of heat exchange per unit area.
K thermal exchange coefficient.
T\v surface temperature of the water.
TE equilibrium temperature.
The equilibrium temperature, TE, is the water temperature at which there is no net
exchange of energy between the water and its surroundings. Determination of TE
involves an iteration where one guesses a value of TE, calculates several functions
which are dependent on TE, and then calculates a new value of TE. The process is
complete when the initial and calculated values of TE lie within a certain range. A
thorough discussion of the method and tables of the required functions is given in
Reference 3. 1
The energy exchange coefficient K is a function of wind speed, evaporation
rate, and equilibrium temperature. The following equation from Reference 3. 1 may be
used
K =15.7 +(O.26 + 1 3)bW
where W = wind speed
b empirical evaporation coefficient
/3 de/dT (Figure 3.1, Ref. 3.1)
e saturation vapor pressure
18
-------
Solutions of the heat transfer equation yielding 4 enipe ature distributions
are obtainable only after simplifying assumptions are mad . Edinger and Geyer (Ref.
3. 1) present an expression for the exponential decay in ten erature downstream of a
point where the discharge and stream are completely mixed. This analysis is rather
simple in that it assumes the heated layer is completely mixed at a given downstream
position; that is, there are no horizontal or vertical temperature gradients present.
As heat is carried away from the stream, the longitudinal temperature decays according
to
[ Qp Tp + - Q ) TR 1 (/PCpUdm)X
T=T + —T e
where Q = plant cooling water discharge rate
= total flow rate of stream
T = plant discharge temperature
TR river temperature prior to mixing
u = average stream velocity
dm = average depth of heated water
p = density of water
C = specific heat of water
More complicated models than the one presented above have been formulated.
However, solution is generally performed numerically using a computer. Typically, a
large number of cases must be run to study the effect of time of day, meteorological con-
ditions, and condition of the body of water at different times of the year. The results are
not, in general, applicable to a large number of cases.
Capital costs for once-through cooling generally include the condenser, cir-
culating water pumps, screen house, and associated piping. The pumps and piping sizes
19
-------
are determined by the volume of water that must be passed through the condenser.
The amount of heat removed by the cooling water is equal to
H thC .A T
where th = mass flow rate of cooling water, (lbm/hr)
= specific heat of water, (1. 0 Btu/lbm 1°F)
= temperature rise of cooling water through
the condenser, (°F).
A condenser may be designed to raise a small amount of water a relatively large AT
or a large amount of water a small T. In the past the availability of water at the plant
site in conjunction with economic studies has determined which route would be followed.
Pumping through a large amount of water and raising it just a few degrees allows a smaller
condenser to be used but the pumping costs are higher.
In addition to the temperature rise and amount of heat to be transferred,
the condenser size is determined by the inlet water temperature, the temperature of
the condensing steam and the overall heat transfer coefficient. Cost of the condenser
will be raised by requiring special construction features and/or special materials
(marine use).
The cost of single pass shell and tube condensers as a function of required
tube area is given in Figure 3. 1 (Ref. 3.4). The area may be computed from
H =UAi T
0
where H = heat removed from condensing steam, (Btu/hr)
U = overall heat transfer coefficient ,(approximately
0
500 Btu/hr-ft —°F)
= log mean temperature difference between steam and
cooling water, (° F).
The following values are typical for power plants:
H = 3840 Btu/kwh for a fossil fuel plant with 40%
thermal efficiency
H = 6410 Btu/kwh for a nuclear plant with 33% thermal
efficiency. 20
-------
required tube area - ft 2
Figure 3.1.
Condenser Capital Cost as a Function
10
0
4. )
C l )
0
0
C d
4.)
0
a)
U)
0
0
1
1 O
1 3
io 4 io 6
of Tube Area
-------
A study by Shade and Smith (Ref. 3.2) run—of-river cooling on an 1800 Mw
fossil fueled plant with 550 F inlet water temperature, 170 F temperature rise through
the condenser and 1.2 inches Hg condenser pressure yielded a cost of $5. 00/kw for
the installed condenser, pumps, and piping. The condenser for a hypothetical 1000 Mw
nuclear plant with an efficiency of 33%, 90°F condensing temperature, 65° F inlet water
and 15°F temperature rise through the condenser would cost $1. 321kw using Figure
3. 1. Adding this to the $0. 84/kw cost for the pumps and doubling the total to account
for installation and then adding the $1. 007kw for the intake crib (Steur, Ref. 3. 55) yields
$5. 32/kw. This cost would be slightly higher for a smaller unit, other factors being
equal, since specific machinery costs decrease as unit size goes up.
The pump cost was based on a figure of $1. OO/gpm. A nuclear plant typically
uses 0. 84 gpm/kw while a fossil plant uses about 0.42 gpm/kw. Kovats (Ref. 3. 3) outlines
some of the aspects of water pump selection and specification. Consideration of small,
multiple pump designs versus large. single units is made.
Condensers for marine installations will cost about 25% more than conventional
units judging from data in Reference 3.4. If we use this figure and identical installation
costs and assume $0. 50/kw for additional piping, the cooling system cost rises to $6. 15/kw
for our hypothetical 1000 Mw nuclear plant. Shade and Smith give a figure of $6. 00 /kw for
a bay or lake cooling scheme where the additional cost is due to longer piping. Eicher
(Ref. 3.56) lists the additional cost of a once-through marine installation at $1. 00/kw.
In summary, once through cooling of large power plants ranges from $5. 001kw
to $6. 507kw. Run-of-river cooling will generally provide the most economic? 1 operation
while installations requiring the use of sea water and long runs of piping will be the most
expensive. It is clear that although simple run-of-river cooling is quite economical, con-
siderations of available sites and stream temperature increases will virtually eliminate
this method as a heat rejection scheme.
3. 3 Cooling Ponds
A cooling pond is a body of water into which the warm condenser discharge
is pumped so that it may be cooled and eventually recirculated through the condenser.
22
-------
PUMP MA c(UP
-yAr 3 (AM
DISC ,44 £
- -
CONDtNfl £7
4 . - ) __________
S1 -. p
‘ i T 7 ___ __
Figure 3.2. CoolIng Pond
-------
The heat is dissipated in the same fashion as described in Section 3.2. That is, there
is a net positive heat exchange across the water surface effected by conduction, evapora-
tion, and radiation. Cooling capacity must be sufficient to insure satisfactory condenser
intake temperatures. Large amounts of surface area are required since the cooling
process is so slow. The land cost then becomes the major economic factor in cooling
pond cost.
According to Edinger and Geyer (Ref. 3. 1) cooling ponds can be classified,
on the basis of circulation pattern and temperature distribution, as
1. completely mixed ponds,
2. flow-through ponds, or
3. internally circulating ponds.
The pond circulation patterns are determined by the geometry, placement of intake and
discharge, wind patterns, and cooling water flow rate.
3.3.1 Completely Mixed Ponds
A completely mixed pond is one in which the pond temperature is almost
uniform except in a small region near the plant discharge. Typically this temperature
is just a few degrees (1 - 2°F) above the equilibrium temperature. This situation occurs
when the product of the exchange coefficient (K) and the pond area (A) is much greater
than the input heat rate (H ) determined by
where p = water density, (Ibm/ft 3 )
= specific heat of water, (Btu/lbm-°F)
= plant discharge rate, (ft 3 /day)
Since the temperature (T) is uniform in the pond, the heat balance can be
written as
P C Qp (Tp - T) = KA (T - TE)
24
-------
The solution is
TD+rT
T= E
1 +r
where r = KA/PCQ
The pond temperature may also be written as
T TE - 1
Tp TE l÷r
T approaches TE as r becomes very large.
The value of the exchange coefficient is usually an average taken over some
period of time (day, week, etc.) so that the actual pond temperature will vary with time.
A shallow pond will exhibit greater variations in temperature when compared to a deeper
pond of equal volume.
3. 3.2 Flow-Through Ponds
In a flow-through pond the temperatures decrease away from the plant dis-
charge. Sometimes the pond discharge is returned to a stream or bay to satisfy local
temperature regulations. More commonly, the plant intake is located at the low tem-
perature end of the pond.
Since the temperature decreases away from the plant discharge, the heat
transfer rate also decreases. Writing a heat balance on a differential element of pond
area (dA) yields:
dT
= K(T_TE)
The solution is
T-T
E -r
Tp TE = e
25
-------
Again, r KA
-
The pond temperature at a point is thus a function of the area between it and the dis-
charge. This analysis is strictly valid for a long narrow pond with no lateral or vertical
temperature gradients since a T widet pond would have vertical and lateral tempera-
ture variations.
If we compare the expressions for the temperature excesses of thoroughly
mixed and flow through ponds of equal area, we see that
T -T r
m E e
Tf TE - 1+r
where Tm temperature of mixed pond
Tf = temperature of flow-through pond
The above expression shows that the temperature at the outlet of a flow-through pond is
always less than that of a fully—mixed pond of equal area.
3. 3. 3 Internally Circulating Ponds
An internally circulating pond exhibits a flow pattern in which the warm
plant discharge flows on the surface of the pond while the cooled water returns to the
plant intake by flowing under the warm layer. Two analyses of such ponds are de-
scribed in Reference 3.1 with mixing between the two layers included in one and neg-
lected in the other.
3. 3.4 Sizing of Cooling Ponds
The cooling ponds must have sufficient surface area that it cools the water
to a temperature that will insure satisfactory operation of the power plant. Assuming that
we know the flow patterns of the pond and can obtain reliable estimates of the exchange
coefficient and equilibrium temperature we can select A such that the pond outlet will
provide cool water. Actually it would require a pond of infinite size to cool the warm
water to the equilibrium temperature. Usually a 2 - 3°F approach is the lowest practicable
in a pond of reasonable size.
26
-------
A quick estimate of the size of a cooling pond may be made by using the
equation
H = hALT
H is the heat rejected to the water, iT is the temperature difference between the
air and average water temperature and a typical value of his 3.5 Btu/hr-ft 2 -°F
may be used. McKelvey and Brooke (Ref. 3. 6) present a chart with typical values
of solar heating and a nomograph to calculate pond size.
‘ Rule-of-Thumb sizing is also used with some engineers suggesting
1 acre/Mw + 20% for surrounding land and others suggesting 2 acres/Mw. The for-
mer is used for fossil fueled plants while the latter is used for nuclear plants.
Variations in calculated pond size can result depending upon which formula
is used. As an illustration, consider the hypothetical 1000 Mw nuclear plant consid-
ered in Section 3.2. Plant inlet water temperature is 65°F and discharge is 80°F.
For simplicity, K was assumed equal to 150 Btu/ft 2 -dav-°F, the equilibrium tempera-
ture was assumed equal to 63°F and the air temperature to 65°F. The results are
given below.
Method of Analysis Required Pond Size Acres
fully mixed pond 11, 700
flow through pond 3, 360
H=3.5A T 5,600
1. 2 acre/Mw 1. 200
2.0 acre/Mw 2,000
Of course, if the above conditions are changed, the pond will have a new design point
and the above comparison may not be valid.
The required surface area of a pond is independent of pond depth for all
practical purposes. However, a minimum depth of three feet appears advisable to
prevent excessive channeling of flow in ponds with irregular bottoms and to avoid
large day-to-day changes in outlet temperature. Some ponds have been designed
with deep sections near the discharge and gradual reduction in depth towards the intake.
27
-------
Conduction to the earth surrounding the pond is almost always neglected
but Nelson (Ref. 3. 7) has given an average heat transfer coefficient of 0. 5 Btu/ft 2 -hr-
°F. Wet soil such as river bottom mud may be as high as 2.5, however. In any event,
it provides a safety factor in the calculations.
Usually natural lakes or bays are not available for use as cooling ponds.
The warm discharge water will not be allowed to enter a natural lake at a very high
temperature and most natural lakes will not be large enough for the bigger nuclear
plants. If a cooling pond is decided upon, an artificial lake must be created by dam-
ming and/or excavating. A study must be made to insure that rainfall or runoff is
sufficient to provide makeup lost by evaporation or a pumping system must be installed
to insure an adequate supply. Maintenance problems are minimal other than checking
and up-keep of dan , Juices, and spiliways, etc.
Cooling pond costs are a very strong function of the amount of land re-
quired. Shade and Smith, using the 1. 0 acre/Mw + 20% for surrounding land (at $iooo/
acre), calculated the cooling pond and piping would cost an additional $5. 001kw. How-
ever, they used a more expensive dual pressure condenser in the calculations compared
to a simple condenser assumed in the run—of—river calculations. Kolfiat (Ref. 3. 20) and
others (Refs.3.55 and 3.56) have estimated the additional cost of a cooling pond as $2. 50
kw. This appears to be a more reasonable figure if we assume 2. 0 acre/Mw, $1000 per
acre for land and excavation, and $0. 5/kw for piping, dams, etc. The cooling pond may
prove to be more expensive in the future as larger amounts of more expensive land are
required for the big nuclear plants.
3.4 Spray Ponds
Spray ponds operate on the same principle as cooling ponds. However,
the evaporation resulting from exposure to the air is enhanced by spraying the warm
water into the air over the pond. Residence time in the air is thus increased as is
the relative velocity between the water droplets and the air. The spray nozzles are
usually located 5 to 10 feet above the surface of the pond. The performance of the
pond is a strong function of the design of these nozzles. Poor designs require high
pressure and power, do not mix the air and water efficiently (poor spray patterns),
and have excessive water losses. Water losses can, however, be cut down with
28
-------
louvered fences. The effect of seven different nozzles on the performance of a spray
pond is given in Reference 3. 10. Unfortunately, performance is limited by the relatively
short contact time between the air and the water spray. Also, impurities can easily enter
the open water area and be carried into the condenser, increasing maintenance require-
ments.
Spray ponds have handled as much as 120,000 gpm. The pumping costs
are less than for cooling towers due to the low head requirements. The heads usually
range from 4 to 30 feet. A spray pond needs only about 5% of the area of a cooling pond
due to the increased transfer coefficient.
There is not a large amount of data on design of spray ponds available
in the open literature. Badger and McCabe (Ref. 3. 11) found that spray ponds will
cool 15- 20 gallons per hour per square foot when cooling water from 110° F or 120° F
down to 70° F. Using the mean values yields a transfer coefficient of 146 Btu/hr-ft 2 -
°F Figure 3.3 from Perry (Ref. 3.12) indicates the performance of a spray pond
for steam condensing service. Robinson and Roll (Ref. 3.13), using a spray pond
4 feet long and 14 feet long with parallel air and water flow, give the following formula
for h in spray ponds: h = 0.0078 L’
where L = water load in lb water sprayed/ft 2 -hr
Typical values of h range between 20 and 160.
0
bt
S
4
temperature) water leaving -
Figure 3. 3. Performance of a Spray Pond
* 10
o0
0
•70
29
-------
Spray ponds require very little maintenance other than occasional cleaning
of the pipes and nozzles and routine maintenance on the pump. The danger of freezing
and/or poor heat transfer due to climatic conditions may be present. It is most im-
portant to start off on the right foot with a proper choice of nozzles and piping layout
to obtain efficient spraying. Low heat transfer coefficients, dead zones where the
circulation is almost zero, and high power consumption are the undesirable alternatives.
Sizing and capital costs may be estimated as follows. AFsuming land at
0. 1 acre/Mw and $1000/acre yields $0. 1/kw. The spray system (piping, nozzles,
pumps and installation) costs about $2. 50/kw. Thus, the spray pond would add
about $2. 50/kw to the base. This price includes the costs of a relatively simple
single pass condenser, circulating water pump, screen house, and piping. Other
equipment such as dual pressure condensers could raise this figure by several
dollars per kw.
3. 5 Wet Cooling Towers
3. 5. 1 Introduction
A wet cooling tower is a device which cools water by bringing the water
into contact with air. The water and air flows are directed in such a way as to provide
maximum heat transfer. The heat is transferred to the air primarily by evaporation
(about 75 c) while the remainder is accomplished by sensible heating.
The rate of heat transfer depends upon several factors other than the tem-
perature and enthalpy differences. They are:
a. The interfacial (contact) area between the water and
air. Packing is often used to break the water up into
smaller droplets which evaporate more easily (splash
type packing) and/or provide more surface area upon
which to develop liquid films (film type packing).
b. The relative velocity between the air and the water.
Higher velocities allow more heat to be carried away
in a given amount of time.
30
-------
c. The residence time of the air in the heat transfer
section of the tower. Sufficient packing height should
be provided so that the air leaving the tower will be
saturated.
A cooling tower is usually comprised of the following parts (Ref. 3. 14):
1. Inlet-water distributor
2. Packing
3. Air moving equipment
4. Inlet-air louvres
5. Drift (carryover) eliminators
6. Cooled water storage basin.
In some towers air movement is due to wind or natural draft.
Tn the past, cooling towers were used primarily as water conservation
devices and were required by law in some cases. Presently, as the number of large
bodies of water which would serve as heat sinks diminishes and the problems associated
with thermal pollution are being realized, cooling towers are being studied as a prin-
cipal method of satisfactory heat rejection.
3. 5. 2 Atmospheric Wet Cooling Towers
Atmospheric towers utilize the naturally occurring winds as the air flow.
Water is sprayed down into the air stream and the heat is carried away. A sketch of
a typical tower is shown in Figure 3. 4. Since the air flow is dependent upon natural
winds, these towers are usually built with a high, broad side exposed to the prevailing
wind. They may be thought of as enclosed spray ponds; the “tower” and louvres keep
drift losses down and provide somewhat greater surface area for water films.
31
-------
I -, - - t 1
liii. * 4 _____
w -.
Figure 3.4. Atmospheric Spray-Filled Tower
According to McKelvey and Brooke (Ref. 3. 6) they should only be used where
the cooling load is less than 1. 8 x i0 6 Btu/hr and the water loading is from 5 to 13 lbs/
mm-ft 2 . These towers are simple and relatively trouble free; packing and filtration
equipment maintenance are the most important items. However, they do have several
disadvantages:
1. Low or change in direction of winds produce a
temperature rise in the exit water.
2. The tower must be installed in an open area where
prevailing winds are not obstructed.
3. The cooling range is low.
4. Pumping losses are higher than for atmospheric
packed towers which usually do not atomize the
water as thoroughly.
The atmospheric packed tower is another variation of the atmospheric wet
tower. These towers are similar to the atmospheric spray tower but are filled with
wooden slats in the zone between the spray nozzles and water. As the water drops over
the packing, some of it breaks up into fine droplets which fall through the air stream
and some forms a film on the packing. This increased contact area increases heat
transfer and pumping costs are lower since high pressures are not necessary.
Neither of these types of towers is built today for use in power plants due
to the large amount of space (large open surrounding area) required, the relatively high
material cost, and the relatively narrow width which necessitates a large support struc-
ture to withstand high winds.
32
-------
3. 5. 3 Natural Draft (ND) Wet Cooling Towers
The movement of air in these towers is a natural draft due to a difference
in density between the inlet and exit air streams. That is, the air outside the tower
has a slightly greater (0. 005 ibm/ft 3 is a typical number) density than the air inside the
tower above the packing. The difference in hydrostatic pressure between two columns
of air of equal height, one inside the tc wer and the other outside, drives the air through
the packing and up the tower. The driving potential is directly proportional to the density
difference and the height of the columns.
The concept is illustrated in the sketch below.
B 1 U —
ambient air density
air density in tower
VE = exit velocity
A
The operation is governed by the following equation
- - — h = pressure drop through packing
0 + frictional losses in tower
driving force .
+ kinetic energy of exit air stream.
The pressure drop through the packing is due to area contractions and ex-
pansions and friction over the packing and spray. It is usually expressed as a re-
sistance in velocity heads (based on inlet air velocity) per foot of packing as a function
of water loading (lbm/hr-ft5. The resistance increases as the water loading increases
and as the inlet velocity level decreases. Obviously, some compromises must be made
to obtain sufficient packing to break up the spray or form films for adequate heat transfer
without a large pressure drop. The frictional losses in the remainder of the tower are not
great since typical towers have about an L/D ratio slightly greater than 1.
33
-------
The kinetic energy of the exit air stream (1/2 PTVE 2 ) is usually specified
to insure that the vapor plume will rise away from the tower under all conditions.
Typical exit velocities are 10 to 12 fps. In the absence of streamline curvature, the
static pressure across the exit plane is equal to the atmospheric pressure at the same
height outside the tower.
Therefore, in order to overcome the pressure drops in the tower and pro—
vide specified velocities at the exit plane and through the packing (the pressure drop
is much greater than the exit kinetic energy loss) in the face of such small density
differences, a large height is needed. The latest designs used with large power stations
are the huge hyperbolic units with heights of 375 feet being common, both in the United
States and England. A sketch of a typical unit is shown in Figure 3. 5. The hyperbolic
design is due to the fact that as the air enters the base it is moving in the radial direction.
As it moves up through the packing, it still possesses some inward radial momentum
which causes the air stream to form a vena contracta at some point above the packing.
The new hyperbolic design thus closely matches the path of the air flow through the
tower whi1e resulting in decreased material costs. In addition, the concrete shell is
actually stiffer against wind forces.
L, 1 e1 1 t7- SL,41, 1 v,47 P 4T 4 E
L /S7 /g /7 7N
wA7 .e c’i-
Figure 3. 5. Natural Draft-Counterfiow Cooling Tower
A/ acir
W Z’E7E
SHEL 1-
,qiR IN
R41ArIM $
w q7 C_ I A !
34
-------
The latest ND towers designed in this country have ranges which are greater
than the approaches by as much as 100 F. Commonly quoted figures (Ref. 3. 20) for op-
timum range and approach for such units are 26° F and 18° F, respectively. The per-
formance is a function of the inlet wet-bulb temperature and relative humidity. The
density difference decreases as the inlet wet bulb temperature increases at constant
relative humidity while the driving force increases as the relative humidity increases
at constant wet bulb temperature. The volume of air that passes through the tower is
therefore dependent upon atmospheric conditions. Design point for these towers is
nearly always based on a “worst” set of atmospheric conditions. Somewhat oversized
towers result but satisfactory operation under adverse conditions is assured.
ND towers may be of the counterfiow or crossflow variety. The two types
are described in the following sections.
3.5. 3. 1 Natural Draft - Counterfiow
The counterfiow tower is depicted in Figure 3.5. The packing is situated
so that the air stream flows vertically upwards through the packing while the water is
sprayed downward. Either splash or film packing may be employed. The splash packing
is usually constructed of redwood timber while the latest film packings (which are more
prevalent in the new units) are asbestos sheets separated by plastic spacers. The sheets
are generally 3/16” thick on 1-1/2” centers (Ref. 3.24).
The tower is equipped with a drift eliminator plate above the water distribu-
tion system to trap water droplets suspended in the air stream. The pressure drop
across it is negligible. A typical water distribution system consists of one or two main
conduits feeding several thousand nozzles to insure proper distribution of hot water.
Pumping head requirements generally run 40 to 50 feet.
3. 5. 3. 2 Natural Draft - Crossflow
The crossflow unit is similar to the counterfiow in outside appearance (see
Figure 3.6) but the packing is set up around the periphery of the base. About half of the
new hyperbolic units are couriterfiow. They generally use timber splash type packing.
When compared to a counterfiow unit of given height and diameter, the crossflow unit
will typically have a smaller pressure drop and lower cooling capacity due to the lower
volume transfer coefficient of the packing.
35
-------
C L W A7
Jr
Figure 3.6. Natural Draft-Crossflow Cooling Tower
3. 5. 4 Mechanical Draft (MD) Wet Cooling Towers
The air flow in a MD tower is produced by a fan which may be placed at
either the inlet (forced draft) on the exit (induced draft) of the tower. The air flow
rate can thus be controlled to satisfy off-design cooling requirements. The use of
the fan also means that the great height of the ND unit is not necessary. Large MD
units are 60-75 feet high. Unfortunately, the lower height also means that the problems
of ground fogging and/or warm air recirculation may present themselves.
The fan is usually a low—head rise (less than 0.5 inches of wateri, large
volume flow rate unit. Propeller-type fans are by far the most common although some
forced draft units do use squirrel—cage fans. Fans on forced draft units are often driven
directly or with multiple V-belts while induced draft units use right angle remote drives
to remove the motor and gearing from the warm humid exit air stream. The fan does, of
course, produce a certain amount of noise, but this has not yet been shown to be a problem.
:: eirr
Ale /N
36
-------
Drift eliminators are placed at the discharge to prevent most of the suspended
water droplets from leaving the tower. Air velocity levels in these towers are in the range
5 to 10 fps. Pumping heads (for the tower alone) for large power station units are 50 - 60
feet.
The initial cost of an MD tower is les than that of comparable capacity ND
unit. However, this is offset by higher operating and maintenance costs over the life of
the tower.
Mechanical draft units may be classified as:
1. induced draft-counterfiow,
induced draft crossflow, or
forced draft.
They are discussed in the following sections.
3. 5.4. 1 Mechanical-Induced Draft-Counterfiow
A sketch of the configuration and operation of a typical unit is given in Fig-
i7,C/FT EL/,4i’/,t, q 7 ’ 1 e
/‘L.M 7} P 4 f
A/,ç’ /4/
Figure 3. 7. Mechanical-Induced Draft Counterfiow-
Cooling Tower
2.
3.
ure 3.7.
FfW
4/ IA! ‘ 1 /
/
37
-------
This arrangement has several advantages:
1. The warmest water contacts the most humid air,
while the coldest water contacts the driest air, thus
utilizing the evaporative process most effectively.
2. Warm air recirculation is minimized by pushing the
discharge straight up.
3. Very large fans up to 60 feet in diameter) can be used.
This avoids going to multiple fans.
4. Close approach and long cooling range is possible (at
increased cost).
The disadvantages are:
1. The typically small open area at the base causes high
velocity inlet air and therefore increasing the fan horse-
power.
2. Baffles must be provided to achieve uniform air velocities
through the packing.
3. Requires greater pumping head than a counterflow unit
due to taller packing.
The newest counterfiow towers use the more efficient film packing similar
to that of the ND counterfiow towers. The pressure drop through the tower is slightly
greater than a comparable crossflow unit since the air must travel upward against
the falling water, but the heat transfer per unit volume is greater so that less air is
needed. The total horsepower requirements are therefore quite similar.
3. 5.4. 2 Mechanical-Induced Draft-Crossflow
An induced draft crossflow is illustrated in Figure 3. 8.
3S
-------
Figure 3. 8. Mechanical-Induced Draft-C ros silow
Cooling Tower
When compared to a counterfiow unit, the crossflow towers have the following advantages:
1. Lower pumping head since the packing is lower.
2. Lower pressure drop through the packing.
3. Higher water loadings are possible for a
given height.
4. The overall height of the tower is less.
They do have the foliowing disadvantages, however:
1. A substantial crossflow correction factor must be applied
to the driving force, particularly where long ranges and.
close approaches are desired. In such cases for some
pumping heads, a crossflow tower may need more ground
area and more fan horsepower than a counterfiow tower
(Ref. 3.57).
2. The packing is not so efficient and more air flow is required
for an equivalent capacity tower.
3. Winter operation is sometimes a problem since ice may form
on the inlet louvres and surrounding structure.
A 1 - Ot
39
-------
4. The packing is in contact with the warmest water
causing more rapid deterioration. Unfortunately,
inspection is more difficult to perform.
Splash or film type packings are available with these units. More durable
materials (asbestos, clay) are available which lengthen the life of packings. The
greater initial cost is offset by reduced maintenance costs.
3.5.4.3 Mechanical Forced Draft
See Figure 3. 9.
The forced draft towers have the fan mounted at the air inlet of the tower.
aA e— C
,- :
Figure 3. 9. Mechanical-Forced Draft Cooling Tower
The advantages of this arrangement are:
1.
2.
Vibration and fan noise are reduced since the fan
is mounted low near the base.
Problems of blade erosion and condensation in gear-
boxes are greatly reduced.
3. These units are slightly more efficient than the induced
draft type since some of the pressure of air velocity
is converted into static pressure in the tower and recovered
as work.
L4 1 ,- I, .,
c k
40
-------
On the other side of the ledger:
1. Practical fan size is limited (up to 12 feet in diameter)
so that mulliple fans must be used for high air flow rates.
2. Baffles are recessary to insure uniform air distribution.
3. Recirculation of the hot, humid discharge may be a
problem since the air may tend to flow back to the
low pressure intake,
4. In cold weather, ice may form on the fan blades, clogging
air inlets and sometimes causing breakage.
Forced draft units are not, strictly speaking, either counter- or crossflow
but are a combination of the two. In some portions of the tower (usually near the en-
trance) the air flow is generally perpendicular to the water flow while at points near
the exit the flow pattern becomes 1 nore similar to a counterflow unit. Splash type
packings are usually employed.
3. 5. 5 Cooling Tower Theory
Cooling tower performance is almost universally based on the theory pre-
sented by Merkel (Ref. 3. 15) in 1925. He showed how the overall heat transfer com-
bined sensible and latent) from the water to the air could be calculated using the enthalpy
potential as the driving force. The derivation of the basic equations is given first and
then modifications for counterfiow and crossflow are given.
3. 5. 5. 1 Derivation of Basic Equations
The following derivation is largely taken from Baker and Shryrock (Ref.
3. 16) and illustrates the assumptions and methods used in cooling tower analysis . A
nomenclature is provided at the end of this section.
Consider an increment of a cooling tower having one square foot of plan
area, and a cooling volume, V, with a square foot of exposed water surface per cubic
foot of volume. There are L. lbs. of water and G, lbs of dry air per hour flowing through
the tower at steady state. A schematic is shown as Figure 3. 10.
41
-------
L. /L/,,,
T-c/T
‘ fAren* /
j 14.e, dV
Figure 3.10. Differential Element of Cooling Tower
Within the volume each particle of water is assumed to be surrounded by
an interface to which heat is transferred from the water.
, ,
I
W< w W’
\
&_ • g .°
. £ ‘‘-
r-i rni
¼ S€ & )
j! tc c *
7’er •7
s 5)
dni =Ka. ctV(w -w)
/7 Lrn’ dc L 5 k v(w -w)
Figure 3. 11. Drop of Water Undergoing Simultaneous Heat
and Mass Transfer
t
/
1
42
-------
The rate of heat transfer from the bulk water to the interface is:
dq = CL dt = KL a dV (t - T’) (3-1)
A portion of this heat is transferred from the interface to the air as sensible heat.
Thus,
dq = KG a dV (T’ - T) (3-2)
The resistance to mass transfer from the water to the interface is neglected,
and the mass transfer of vapor from the film to the air is
dm K’ a dV (W” - W) (3-3)
This may be converted to a heat rate by multiplying by hfg Therefore, the
latent heat transfer is given by
bfg K’ a dV (W” - W) (3-4)
At steady state, the mass rate of water evaporated equals the rate of
humidity increase of the air. Therefore
dm = GdW (3-5)
The heat lost by the water is equal to the heat gained by the air. The heat
balance is
G (h 2 — h ) = [ L (t 1 — 32) — (L — LE) (t 2 — 32)] C
= [ L (t 1 — t 2 ) + LE (t 2 — 32)] C (3-6)
and since LE = G 2 - W 1 )
G (h 2 - h 1 ) = [ L (t 1 — t 2 ) + G (W 2 - W 1 ) (t 2 — 32)] C
In practice, the second term on the right-hand side is small and is neg-
lected so that we may write
43
-------
Gdh = C Ldt (3-8)
pw
The enthalpy of moist air is defined as:
h = C (T - T 0 ) +W [ hfg + C (T - T 0 )1 (39)
and
dh Cp + W C ) dT ÷ E h g + C (T - T 0 )] dW (3-10)
The specific heat of the moist air is
C + WC (3-11)
p pv
dh = C dT + {h 1 g + C (T - T 0 )1 dW (3-12)
and combining Equations (3-8) and (3-12) we can write
C Ldt G C dT + [ h + C (T - T )} GdW (3-13)
fg
The first term on the right represents sensible and the second latent heat.
Since
dq GC dT,
S P
dm= GdW
and K
G
CK’ 1
p
for air-water vapor mixtures we can write
Ldt= K’ adV {c (T’ - T) + [ hfg + Cp (T - T 0 )] T”_W)} (3-14
44
-------
The enthalpy of the air stream may be written as
h C T - Cpa T 0 + W (hfg - C , T ) (3-15)
the enthalpy of the interface is
h” = C T’ - C T + W” (h - C T ) (3-16)
fg p °
Solving Equations (3-15) and (3-16) for T arid T’, and substituting the results in Equation
(3-14)
C Ldt = K ’ adV (h” - h) ÷ C T (W” - W) (3-17)
w pv
Merkel showed the second term on the right to be small so it is neglected; thus
C Ldt = K’adV (h” - h) = Gdh (3-18)
w
This eqUation relates the air stream to the interfacial film, the conditions
of which are indeterminate for all practical purposes. This is overcome by assuming
V = T. The coefficients KG and K’ are then replaced by overall coefficients Kg and
K. Assuming the Lewis relation still applies,
K
g
KC
Actually it is more nearly 0. 9, but it has been common to assume it does
apply. Using these assumptions yields
C Ldt KadV (h’ - h) = Gdh (3-19)
p
w
45
-------
and
= f2 h’— h (3-20)
or h
Kay j2 h’-h (3-21)
Equations (3—20) and (3-21) are convertible into one another and are two forms of the
basic equation. They conform to the NTU (Number of Transfer Units) heat exchanger
design concept in the sense that the value of NTU specifies the size of the equipment
necessary to achieve the maximum possible effectiveness. It can be shown that heat
exchangers with values of NTU greater than those necessary to produce the maximum
effectiveness are too large (excessive amount of heat transfer area) and therefore more
expensive than necessary. The NTU corresponding to a set of hypothethical conditions
is called the required coefficient. The same calculations applied to a set of test con-
ditions is called the available coefficient of the tower.
The designer must provide sufficient packing depth for good heat transfer
effectiveness without excessive pressure drop. The latter point is particularly im-
portant in ND towers since the height of the tower increases in direct proportion to
the packing pressure drop.
The group KaV/L is often called the “volume transfer coefficient.” It is
independent of temperature for all practical purposes but is a function of both water
and air flow rates. Lowe and Cristie (Ref. 3. 58) tested some 60 packings suitable for
use in cooling towers and found that the experimental results for all but one could be
represented by an equation of the form
Ka L
L
The two factors X and n thus characterize the transfer characteristics of the packing.
The packing depth (V) is computed to satisfy Equation (3-20).
46
-------
Nomenclature for Section 3. 5. 5. 1
a area of water interface, (sq ft/cu ft)
C unit heat capacity (humid heat) of moist air, (Btu/° F-lb dry air)
CPa specific heat of dry air at constant pressure, (Btu/lb-° F)
specific heat of water vapor at constant pressure, (Btu/lb-° F)
C specific heat of water at constant pressure, (Btu/lb—°F)
G air flow rate, (lb dry air/hr-sq ft)
h enthalpy of moist air, (Btu/lb dry air)
h 1 enthalpy of moist air, entering tower
h 2 enthalpy of moist air, leaving tower
h’ enthalpv of moist air, at bulk water temperature
h” enthalpy of moist air, at interface temperature
hfg latent heat of evaporation, assumed constant in system
K overall unit conductance, mass transfer between saturated air
at mass water temperature and main air stream, (lb/hr-sq it-
Ib/ib)
K’ unit conductance, mass transfer, interface to main air stream,
( [ b/hr-sq ft-lb/lb)
K overall unit conductance, sensible heat transfer between main
g water body and main air stream,( Btu/hr-sq ft-° F)
KG overall unit conductance, sensible heat transfer between
interface and main air stream, (Btu/hr-sq ft-° F)
KL unit conductance, heat transfer, bulk water to interface,
(Btu/hr—sq ft- F)
L mass water rate, (lb/hr-sq it)
LE mass evaporation loss, (lb/hr-sq ft)
47
-------
Nomenclature for Section 3. 5. 5. 1 Continued
m mass transfer rate, interface to air stream,( lb/hr-sq ft)
rate of latent heat transfer, interface to air stream,
(Btu/hr-sq ft)
rate of sensible heat transfer, interface to air stream,
(Btu/hr-sq ft)
rate of heat transfer, bulk water to interface, (Btu/hr-sq ft)
t bulk water temperature, (° F)
t 1 bulk water temperature at inlet (hot water) , (° F)
t 2 bulk water temperature at outlet (cold water), (° F)
T dry-bulb temperature of air stream, (°F)
T datum temperature for water vapor enthalpy, ( F)
T ’ dry-bulb temperature of air at interface, (°F)
TWB wet-bulb temperature, air stream, (°F)
V active tower volume, (Cu ft/sq ft plan area> (packing height)
W absolute humidity (humidity ratio) of main mass air,
(lb vapor/lb dry air)
absolute humidity of main air mass entering tower
absolute humidity of main air mass leaving tower
absolute humidity at interface
absolute humidity, saturated at water temperature
48
-------
3. 5. 5.2 Counterfiow Tower
The figure below is a schematic of a simplified counterfiow tower. The
enthalpy of the air increases as it rises in the tower while the enthalpy of the saturated
interfacial film on the water drop decreases as the drop falls.
Ri-, ‘ 4
1
©
‘ .5 /
G
0
Figure 3. 12. Schematic of Coimterflow Cooling Tower
A graphic representation of the enthalpies and temperatures for a counter-
flow unit is given in Figure 3. 13. The water entering at the top of the tower (point A)
is surrounded by an interfacial film at the bulk water temperature t 1 . The enthalpy
of the film at point A is h . As the water is cooled, the film follows the saturation
curve to point B, where the temperature is t 2 and the enthalpy h 2 ’.
The air enters with an enthalpy h 1 at point C and leaves with an enthalpy
of h 2 at point D. The potential at the base of tower is h 2 ’ - h 1 , represented by the
vertical distance BC. Heat removed from the water is added to the air and from
Equation (3-18) dh = L/Gdt. The air operating line is therefore straight.
#
d— .c-i
I®
49
-------
The heat transfer driving potential at any point in the tower is represented
by the vertical distance between AB and CD. Starting at the bottom, since air and
water conditions are known there, a numerical integration is performed. The tower
is divided into a finite number of volumes with the divisions having equal temperature
drops. The equation
= C L t = KadV (h’ - h)
p m
applies to the increment where (h’ - h)m is the mean enthalpy difference in the increment
of volume and delta refers to quantities for the increment. Summation of the increments
yields
KaV LC t (h’ -h)
Illustrations of the use of this method are given in References (3. 16) and (3. l7 .
60
-4-
4C
h. r_ — --
Sc 6C ‘O 80
Figure 3. 13. Enthalpy-Temperature Diagram of Air
and Water in Cooling Tower
— —
&
‘a
F
80—
90 tOe’
,‘e Ye -‘ e#-, ‘-c O
t
-
“ 120
50
-------
3. 5. 5. 3 Crossflow Tower
A crossflow tower has horizontal and vertical variations in potential. The
cross sections are divided into unit volumes z x wide and y high. The mean driving
force in crossflow cooling is calculated by averaging the reciprocals of the inlet and
exit potential differences. This straight line relationship is a necessary assumption
because calculation of the mean driving force is not easily done since each unit—volume
is as complex as the tower as a whole. Excellent discussions and examples are given
in References 3.15 and 3. 57.
3.5.5.4 Discussion and Summary
The analysis presented here included several assumptions:
1. The bulk water and interface were assumed to be at
the same temperature.
2. The resistance to mass transfer from the water to
the interface was negligible.
3. The quantity of water evaporated during the cooling
process was neglected.
The first two assumptions indicate that the calculated energy transfer will
be higher than can be verified by experiments. The effects become more pronounced
at greater temperature differentials between the air and the water and at higher L /G
ratios since evaporation effects become more important at the higher L /G ratios.
This is the reason for the use of ‘hot water correction factors, ‘ discussed in Ref-
erences 3.18 and 3.19.
3.5.6 Size Estimation
The size of a wet cooling tower is dependent upon four factors: (1) heat load,
(2) range, (3) approach, and (4) wet bulb temperature. The heat load (in Btu ‘hr) can be
calculated from the following equation
51
-------
Iii 1
H 3413 x — 1.01 x (plant size — kw)
L th J
where piant efficiency (0. 95 for nuclear: 0. 85 for fossil fuels)
th = thermodynamic efficiency of the Rankine cycle.
The range is usually specified by the increase in temperature through the condenser
or local temperature regulations. The approach is generally governed by economics
—-the smaller the approach. the greater the capital cost of the unit. Usually a 5 ap-
proach is the coldest water outlet temperature that a cooling tower manufacturer will
guarantcc. Even this figure is seen only in very well designed mechanical draft towers.
Natural draft units run with 15 - 20°F approaches. In sizing a tower, approach. fol-
lowed closely by gpm is the most important factor. The wet-bulb temperature usually
used for design purposes is one somewhere near the maximum for year-round variations.
It can be higher than the value obtained from meteorological data if other towers or pro-
cess equipment are in the immediate vicinity.
An important parameter is the circulation rate (L) which is governed b
the heat load and range according to the following equation
Q Btu/hr)
L(gpm )
range x 300
The circulation rate is therefore inversely proportional to the range. Addition of water
is required to replace water loss by evaporation, carryover, blowdown, and leakage.
This makeup amounts to about 3. 5 of the circulation rate with about 1. 5 due to evapora-
tion and the majority of the remainder attributable to blowdown. Properly designed drift
plates practically eliminate carryover.
Natural draft towers are necessarily larger than mechanical draft types. A
rough idea of the relative plan areas may be made by comparing packing water loads.
Natural draft units run with 2 to 4 gpm/ft 2 while mechanical draft towers average about
6 gpm/ft 2 . It is possible to achieve up to 10 gpm/ft 2 with a doubleflow MD tower. A
natural draft tower will therefore require 2 to 3 times more ground area than an equiva—
lent capacity MD unit.
52
-------
There is an even greater contrast in the heights of the two towers. Large
counterfiow MD towers require about 20 feet below the packing for inlet area and 5 to
10 feet above the packing for water distribution system, fans and stacks. The packing
for a large power station tower will be 20 to 30 feet high. Using the maximum values
gives 60 feet high.
As shown in Section 3. 5. 3, the great height of the ND tower is necessary to
overcome the losses thru the packing, in the tower and at the discharge. Assuming
an average density difference of 0.005 Ibm/ft 2 and a total pressure loss of 0. 4 inches
of water (including the kinetic energy loss at the exit) yields 416 feet for the required
height. The density difference cou1d represent an ambient condition of 87° F with
relative humidity of 50 and a discharge at 100°F,100% relative humidity. This dis-
charge temperature may sound high. but the large hyperbolic tower at Ft. Martin,
West Virginia cools water from 114°F to 90° F, for example. The figure of 0.4 inches
of water includes a packing pressure drop of about 3. 7 velocity heads (based on 3 fps)
per foot of packing for a 40 foot deep packing based on data obtained from the same
tower (Ref. 3.59).
Tower plan area estimates were made for a 1000 Mw nuclear plant using
0. 8 gpm/kw. The following required base areas were calculated:
Type Loading Required Area Area/2
ND 3 gpm/ft 2 2.67 x 10 ft 2 1.33 x 10
MD 6 gpm/ft 2 1.33 x 10 ft 2 0.67 x 10
The required area of the ND tower could be provided by two units 412 feet in diameter.
The MD required area could be provided by a combination of 10 square cells. The
dimensions would then be 115 feet wide and 1150 feet long.
A size comparison has been given by Kolfiat (Ref. 3. 20). A hyperbolic
natural draft tower for a 500 Mw station would be about 370 feet high and have a 400
foot base while a mechanical draft unit for the same station would be about 600 feet
long, 70 feet wide and 60 feet high. This yields the following base areas:
53
-------
ND 1.26 x10 5 ft 2
MD 0.42 x 1O 5 ft 2
These figures compare favorably with the example given above (compare with Area/2
Column).
In addition to the area required for the tower itself, land is required for
reservoirs, blowdown ponds, and pumping and storage areas. Large towers cannot
be packed too closely together since high winds produced by a Venturi-effect can
cause excessively high wind loads.
The size of a cooling tower can also vary with location and design capability.
As the meteorological conditions vary from place to place, tower performance (es-
pecially wet natural draft units) will be altered. The size can also be reduced if a
risk is taken that the cooling tower will not be able to accommodate the design heat
load 100% of the time. A reduction in capacity is then incurred under adverse con-
ditions. Reference 3.21 describes a method of approaching this problem and calculating
the possible reduction in capacity.
3. 5. 7 Cost Information
Cooling tower costs can be grouped under one of three headings:
1. Initial Costs,
2. Annual Fixed Costs, and
3. Annual Operating Costs.
The initial costs are a strong function of the plant site. Land costs, installation and
accessory costs, and design ambient conditions may all be affected by location. In
some locations it may be necessary to construct reservoirs and dams to insure suf-
ficient makeup water supply while in other locales rivers or lakes may be available.
Since a cooling tower installation with, for example, run-of-river makeup will not
require the more expensive intake and discharge tunnels, screen house, screens
and rakes that a run-of-river once-through cooling system requires, the cost data
54
-------
will be presented for the complete cooling system of condenser, tower, piping, etc.
From this point of view the cooling system may be considered as a piece of equipment
which will be interfaced with the plant at the turbine exhaust.
There are two types of cost information available in the open literature.
One gives ball-park estimates usually expressed as $/kw while the other gives the
cost in dollars expressed as an equation or graph. Both of these types of numbers
are derived from correlations of past data, but the latter type (typically from process
mechanical draft units) is expressed as a function of the heat load, range, approach,
and wet-bulb temperature so that the effect of changing conditions may be studied.
The following table gives a breakdown of cooling tower costs.
Table 1
COOLING TOWER COSTS
1. Initial Costs
a. Capital Cost
b. Installation Cost
c. Accessory Capital Costs
2. Annual Fixed Costs
a. Amortization and Depreciation
b. Interest Costs
c. Taxes
d. Insurance
e. Rents
3. Annual Operating Costs
a. Power (Pump and Tower Fan)
b. Power (Other)
c. Maintenance
d. Water Treatment
e. Makeup \Vater
f. Sewer Charges
55
-------
The approximate initial costs for the ND and MD towers are discussed in
the following sections. In Task I, Phase II. we shall carefully correlate the most
recent data on cooling towers and determine cost-size relationships for use in the
system studies.
The bulk of annual fixed costs is also a strong function of the particular
plant. Taxes, insurance, rents can vary widely from area to area. Interest charges
are usually the same across the country but may vary from year to year. Capital
costs for a power plant are generally amortized over 30 years.
Operating and maintenance costs for pumps vary as to location and type of
tower. According to Kovats (Ref. 3.3), operating costs for different brands may be
taken as identical and does not depend on whether large single or smaller multiple
setups are used. Downtime is greater with large single, however, and four pumps
in steps of about 32, 67, 82 and 100% of total flow gives good range of flow adjust-
ment and flexibility.
Kovats also says that pumps used in cooling tower service require 1. 0 to
1.2 percent of generator load capacity and Berg and Larson (Ref. 3.26) suggest that
maintenance costs can be taken as equal to power costs for a first estimate. Accurate
operating and maintenance data is not available from manufacturers due to lack of
feedback from users. Over the 30 year life of a pump, approximately another 2/3 of
the initial cost will be spent on replacement parts.
Costs of water treatment are given by Lane and Larson Ref. 3.27), Jones
(Ref. 3.28), and Berg and Larson (Ref. 3.26). Lane and Larson present costs for
lime softening. zeolite softening. Jones draws attention to the much lower costs at-
tainable when the unit is run at higher concentrations. Discussion of the limits of
amounts of certain chemicals, particularly calcium and sulfates, is also given by
Lane and Larson (Ref. 3.26).
3. 5. 7. 1 ND Towers - Capital Cost
Some recent economic studies have indicated that the wet ND hyperbolic
design is the most economical tower when such a unit is required. ‘Unfortunately,
56
-------
only a few of these have been built in the United States and fifteen or so are in the
construction stage so a large amount of data is not available. The British have built
some 60 to 70 of these towers but labor and material costs are quite different and
design wet-bulb temperatures run 15 to 20 degrees lower than in the United States.
Due to the higher wet—bulb temperatures, U. S. plants with hyperbolic towers usually
run at higher exhaust pressures (2. 5 - 3. 5 inches Hg) and resultant lower efficiency.
This is necessary to achieve suitable ranges and practical approaches.
It must be noted that the studies referred to above are overall studies which
included fuel, operating. and usually transmission costs. A typical situation occurs
where the use of a cooling tower allows the plant to he located on less expensive land
or closer to a mine mouth or a large user of electricity. Water around mine mouths
is generally acidic and requires treatment before use in the condenser. A cooling
tower that recirculates the water allows treatment to he performed economically.
That is. man of the studies show that cooling towers are attractive due to reduced
fuel, land, or power transmission costs.
The following table is a compilation of cooling system costs using ND wet
cooling towers.
Table 2
Ref. Plant Size Fuel $/kw Comments
3.20 1000 Mw N 15.00*
Hypothetical plants
3.23 1000Mw N 11.50*
3.24 540 Mw f 8. 65* Ft. Martin plant
3. 2 1800 Mw f 7. 50 Run—of—river makeup
3.60 1800 Mw f 7. 90* Keystone plant
3.61 1600Mw 1 9.40 Reservoir makeup
S4/kw was added to tower cost for condenser, piping, pumps, etc. to compute
system cost.
57
-------
A study on cooling towers for a 400 Mw fossil-fueled plant cooling 150, 000
gpm (Ref. ) yielded the following costs as a function of approach.
Tower Diameter Approach, °F Tower Cost, $/kw System Cost, $/kw
210 21.6 3.0 7.0
230 18.8 3.43 7.43
250 16.5 4.12 8.12
270 13.9 4.70 8.70
If we assume a 38 thermal efficiency, the range is 22. 6°F. The above table clearly
shows the increasing cost of closer approaches. The study does not account for any
increases in cost duetoplant location.
The estimated cooling system cost using ND wet cooling towers is sum-
marized below.
Fuel Approach, °F System Cost, $/kw Makeup
1 22 7. 00 river
f 18 7.50 river
1 18 10.50 reservoir
N 18 11.50 river
N 18 15.00 reservoir
The cost of pumping runs about 0.4 . of generator capacity for a nuclear
plant. The tower has an 18°F approach and a 25°F range. The cost under other
conditions may be predicted from graphs given in Section 3.5. 7.2.
3. 5. 7.2 MD Towers - Capital Cost
Prior to 1962, I\ID draft towers dominated the power plant cooling tower
field. Recently, however, the large power plants have been equipped with ND hyper-
bolic towers. In spite of their higher initial cost, ND towers have lower operating
58
-------
and maintenance costs. Pumping costs will be similar for ioth tvers. Fans for
MD units consume about 0.8% of generator capacity (Ref. . 20). This may be even
higher if a very close approach (less than 10° F) is desired. ND towers do not require
any fan power. The use of a mechanical draft tower results in a penalty of 2 to 3% of
plant capability compared to 1 to 1. 5% for a natural draft unit.
Kolfiat (Ref. 3. 20) has estimated that an induced draft wet MD tower for a
nuclear plant would add $6/kw to the plant cost. Cooling towers for the 540 Mw nuclear
plant at Vernon, Vermont would cost an additional 6 million dollars or $11. 1 /kw (Ref.
3. 22). The Vernon towers have a very close approach which accounts for the high cost.
The effect of approach on initial cost may be seen in the following data taken from Ravet
(Ref. 3.62):
400 Mw fossil fueled plant 150, 000 gpm TWB = 76 53% R. H.
Approach, °F Tower Cost, $ S/kw
21.4 4.8x10 5 1.20
16.6 6.0 1.50
13.6 7.2 1.80
12.0 8.0 2.00
9.6 9.6 2.40
Lockart et al (Ref. 3.63) correlated the data obtained by Knowlton Ref.3. 64)
for 52 mechanical draft cooling towers. The data are presented in the form of a graph
on Figure 3.14. It was necessaryto ad just the costs since the data was obtained in 1955
and some preliminary calculations indicated that costs have approximately doubled since
then. Extrapolation of the curves was also performed to extend the useful range.
59
-------
i
-4-.—.
—
•1--
:: :
:i
t
:
: ‘.:
::
-
:
-.
-_ -. .
60 65 70 75 80
WB
0
C)
a.
0
0
0
5
4
3
2
1
10 15 20 25 30
Range -
7
I 1 I I 10
I.iiti:
.
,
,
.
- ..
. .
T
-
.:::
— —
-k+
-
— -•
:.
: :
—
::.
•—
—
—
+
--
. , .‘t.::..
— 7 —j -. -—
r’V
: -.
— -----: - - -
:.:.
- —--.— ... —4— —
-
:
::::
- -‘ -
- -I
E
1. Total Fan and Pumping Horsepower
2. Tower Fan Horsepower
3. Total Installed Cost of Tower and Accessories
4. Cooling Tower Cost
I i I I I I I I I I I I I I
— - 1H ! .,.T
0 L fH I jj fL i f I
0 105 106 io
(gpm) K) C\\ B)
Figure 3.14. Costs and Power Consumption of MD
Wet Cooling Towers
60
1.4
1.2
1.0
.8
0
0
10,00(
1,00k
2’
— : /
7’
0
--
—
0
6
10
ci
0
0
-------
To illustrate use of the curves, we have compared data from the graphs
with Ravet’s data.
TWB= 76
Range = 23°F
CWB 0.90
gpm 1.5x10 5
21.4
16. 6
13. 6
12.0
9.6
Tower Cost, $/kW
3.60
4.75
5.87
6.30
7.50
5.37
7.13
8.88
9.55
11.25
9.37
11.13
12.88
13. 55
15. 25
The cost figures given by Kolfiat and that of the Vernon, Vermont plant lie within the
range given in the third column. It is felt, then, that Figure 3. 14 will deliver reasonable
estimates of wet mechanical draft cooling tower initial costs. The total system cost at
18°F approach is found to be $10. 50/kw. This compares to $11. 501kw for an ND unit
with run-of-river makeup or $15. 001kw for ND tower with reservoir makeup (figures
are for nuclear plants).
Kolfiat has also estimated that the fans for a wet mechanical draft unit will
consume about 0. 8% of the generator capacity. Figure 3. 14 gives 0. 46% of generator
Approach, °F K
21.4 1.8
gpmxKx CWB
Tower Cost,$
$/kw
1.80
Tower Cost, (ReL 3.62)
4.8x10 5
2.43x10 5
7.2x10 5
16.6
2.35
3.17 x 10
9.5 x 1O
2.37
6.0 x 1O 5
13.6
2.90
3.92 x 10
11,8 x 1O 5
2.94
7.2 x
12.0
3.13
4.23 x
12.5 x iø
3.12
8.0 x 10
9.6
3.70
4.99 x 10
15.0 x 10
3.74
9.6 x 10
The costs obtained from the graph appear to be about 50% high.
if the 400 Mw plant was a nuclear installation, the gpm would be approxim-
ately doubled. The graphs yielded the following figures:
Approach, °F
Total Cost, Installed
Tower & Accessories, $/kw
Total System
Costs, $/kw
61
-------
capacity for the fan alone and 0. 85% for the fan and pump horsepower requirements.
This figure is for a nuclear plant with a design wet-bulb temperature equal to 72° F,
25° F range, 18° F approach and 0. 85 gpm/kw. A fossil-fueled plant operating under
the same conditions would use about 0.23% for fans alone and 0.42% for the total of
fans and pumps.
The power costs are a strong function of approach. If the approach was
10°F instead of 180 F for the nuclear plant, the consumptions would be 0. 78% and 1.41%
for the fans and the fans and pumps, respectively.
The pumping cost for a natural draft tower may be estimated from Figure
3. 14 by subtracting the fan horsepower requirement from the total. This is valid since
the pumping heads are quite similar for both types of towers.
3. 5.8 Maintenance Considerations
It has been pointed out by a number of authors that proper maintenance is
essential to the satisfactory performance of a cooling tower. The task is difficult.
Components of the tower may be flooded, splashed, sprayed, exposed to hot humid
air and yearly variations in ambient conditions. The capacity for heat transfer is
dependent upon water flow, air flow and the particular tower characteristics. Poor
maintenance can produce adverse effects in all three areas.
Water flow is often altered, particularly the distribution system. Flooding
due to corrosion or poor spray patterns due to corrosion or clogging are often seen.
Effects on the air flow are usually exhibited by increased static pressure. Scale, oil,
mud, and algae accumulated on the fill or mist eliminators result in decreased air flow
and poor performance. Flow pulsations within the fan sometimes occur causing un-
stable performance. Air leaks may decrease the capacity of the tower by allowing air
to flow through the tower without contacting the falling water. Within the water,
deterioration of the packing and support structures can cause unfavorable distributions
of air and water, decreasing the heat transfer coefficients. A long list of examples of
poor maintenance is given in Reference 3. 29.
62
-------
The key to success in this area is periodic maintenance and inspection.
Present day corrosion inhibitors and fungicides have extended the time required be—
tween replacement of items such as packing, fasthers, and support structure. A
periodic maintenance schedule allows work to be accomplished under the most favor-
able conditions.
Proper selection of materials is very important as are the particular
coatings and treatments selected. A compromise between long service life and
cost must be made. A set of “normal” operating conditions has been defined to de-
termine service life for materials and costs (Ref. 3. 30).
1. Circulating water pH between 6 and 8; chloride content
below 750 ppm as NaCl; sodium bicarbonate content
below 200 ppm as NaHCO 3 ; maximum temperature of
140F; no significant contamination with unusual chemicals
or foreign substances; and adequate water treatment to
minimize corrosion and scaling.
2. Chlorin if used, is added intermittently with a free
residual of one ppm or less maintained for short
periods.
3. Atmosphere around tower is not worse than moderate
ind s trial. Rainfall and fog are only slightly acid and do
not contain significant chlorides or hydrogen sulfide.
With regard to chlorine treatment, the following should
be noted: “The surface deterioration of redwood in cooling
tower service may be virtually eliminated by avoiding the
use of chlorine and using a non-oxidizing algaecide in the
treatment of the circulating water.”
Reference 3. 31 lists the standards for lumber, plastics, electrical service, metal
hardware, and safety usually considered in specifications of cooling towers. The
costs of maintenance have been discussed in Section 3. 5. 7.
63
-------
3. 5.9 Secondary Pollution Considerations
The use of a cooling tower may be accompanied by undesirable side effects.
Chemicals harmful to various forms of life may be discharged to nearby streams, rivers,
or lakes. The evaporation of water into the air stream results in an increase in the con-
centration of dissolved solids in the remaining water within the tower. Scale formation
or corrosion in various components of the cooling system will eventually occur depending
upon the pH of the water. Scale formation generally lowers the heat transfer ability of
the system while corrosion leads to material failures. Briefly, the saturation pH (pH )
is the minimum pH that a sample of water can have at a certain temperature and still
precipitate scale. If the pH of the water is above pH 5 , scale will be produced and the
water is essentially non-corrosive. However, if the pH is below pH 5 , scale (if any) will
be removed and the water may be corrosive. The lower the actual pH relative to pH 5 ,
the greater the tendency of the water to be corrosive. To maintain the pH of the re-
circulated water between 6 and 8, sulfuric acid or caustic soda is commonly added to
the makeup water to prevent scaling or excessive corrosion. Scaling, corrosion, and
wood deterioration all increase with increasing temperature.
To maintain the proper pH and keep the concentrations of dissolved solids
below certain levels, a portion of the recirculated water is bled-off and discharged
as waste. It is replaced with raw or treated water. This process is known as blow-
down and may be performed continuously or intermittently.
A typical water treatment program is made up of the following five steps:
1. Analysis of the available raw water.
2. Determination of maximum allowable concentration of
dissolved solids.
3. Determination of the optimum concentration for the system.
This involves a tradeoff between the additional cost of water
treatment and cost of deterioration of equipment due to scale
and corrosion.
64
-------
4. Determination of the optimum pH of the water system.
5. Selection of the proper scale and corrosion inhibitors.
It is possible to select an inhibition that will reduce
corrosion in one part of the system and also prevent
scale deposition in the hotter portions of the system.
Malley and Gasper (Ref. 3. 65) describe sixteen systems for bleeding cooling
tower water basins. They are divided into three categories:
1. Continuous bleed-off,
2. Proportional or semi-proportional bleed-off, and
3. Automatic units.
The first two systems utilize manual checking and adjustment of the concentration level
while the third monitors concentration via a conductivity measurement and regulates
bleed-off through a feedback control system.
The continuous systems generally have a pipe, trough, or weir somewhere
in the system. A portion of the water in the basin is allowed to flow out of the basin
and is discharged as waste. Some systems collect a portion of the water that has just
fallen through the packing and discharge it to the waste stream. Bleed-off may also
be taken from the circulating pump. The proportional systems permit a bleed-off
which is proportional to cooling load. They usually have valves whose activity is tied
to circulating water pump or fan operation.
According to Berg, et al (Ref. 3. 26) the blowdown in gallons per hour re-
quired to satisfy a given concentration ratio may be estimated as a fraction of the
evaporative loss from the following equation
0.06 x R x L
Blowdown = - 1
where R = rang (°F)
L = circulation rate, (gpm)
C = ratio of circulation water to raw-water minerals concentration.
65
-------
Values of C run from 2 to 10. For example, a large ND cooling tower
might have the following values:
R 24°F
L = 250,00 0gpm
C=5
Blowdown 0.06 (?4) (250, 000 ) 90,000 gal/hr
This water requirement must, of course, be added to the amount of water loss via
evaporation and other operations.
Careless handling of blowdown and corrosion inhibitors can discharge
pollutants into the stream. This discharged water may contain high concentrations
of acids, caustic, or chlorine (used for control of algae anJ slime growth). Chlor-
inated phenols and quaternary salts are also used for algae and slime control. Chlorine
is not usually a problem since halogens are oxidizing agents and dissipate quickly.
The salts are toxic but can be reduced to low concentrations and biodegraded quite
easily. The chlorinated phenols, on the other hand, are not as easily degraded and
remain toxic for long periods of time.
In cases where very stubborn growths are observed heavy metal salts
such as copper or mercury compounds may be used. These materials remain toxic
as long as they are water soluble. This causes organisms to demand even more oxygen
from the already oxygen deficient water. Higher temperatures also tend to raise the
toxicity of many materials.
Abnormal situations can enter the secondary pollution picture. During
startup, lignin, chromated copper arsenate (the most popu1ar preservative), and
other preservatives may be leached out of wood very rapidly and enter the waste
stream in high concentrations. Use of asbestos and plastics which offer better re-
sistance to fungal attack alleviate this problem to some extent. Heavy-duty coatings
such as epoxy coal tar may have been used on steel and cast iron components. Poly-
chlorophenates have been used for years as an algaecide and fungicide in concentrations
66
-------
up to 300 ppm as a shock treatment (Ref. 3. 66). Dust and other materials from the
air stream can collect in the basin of the cooling tower. If not removed, they may
enter the stream. Acid cleaning of cooling equipment is also a potential problem
unless the acid is properly drained and diluted before disposal.
Cooling towers can be set up so that the rate of pollutants entering the
streams is within some accommodation limit. Operator and equipment failures can
destroy this balance by releasing large amounts of pollutants in a short amount of
time. Several methods of sampling the waste stream involving chemical or elec-
trical measurements are available. The sampling and analysis must be done rapidly
so that adequate corrections may be made. The control methods are dependent on the
comparative sizes of pollutant discharge and receiving stream. If the amount of pol-
lutants is small, dilution combined with controlled release rates can be used. If the
ratio is large, coagulation, neutralization, or softening must be employed. Under
severe conditions it may be necessary to shut down the tower for several hours while
the water is treated.
3. 6 Dry Cooling Towers
3.6. 1 Introduction
The dry cooling tower is a relatively recent development. Only a few power
plants have been built which utilize this cooling technique. Notable examples are a 120
Mw plant at Rugeley, England and a 15 Mw plant in West Pakistan. The dry cooling tower
technique appears to be a very promising alternative to once-through cooling.
3.6.2 Description
The dry cooling tower is usually employed as a component of the “Heller
System” which is illustrated in Figure 3. 15. The turbine exhaust is condensed in
a direct contact spray condenser. A portion of the condensate flows to the boiler,
while the remainder is pumped through the dry cooling tower in a closed loop, re-
turning to the condenser. The dry cooling tower itself consists of an air-cooled heat
exchanger mounted inside a cooling tower chimney. The airflow past the heat ex-
changer may be driven by fans (forced draft), or it may result from the natural draft
of the heated air inside the tower. The natural draft tower requires a larger tower than
the forced draft, but it requires no mechanical air moving equipment.
67
-------
Air Flow
1. Boiler
2. Turbine
3. Direct Contact Spray Condenser
4. Feedwater Pump
5. Cooling Loop Water Pump
6. Air Cooled Heat Exchanger
7. Cooling Tower Chimney
8. Water Turbine
Figure 3.15. Basic Diagram of the “Heller SystemY A Direct Contact Spray
Condenser With a Dry Cooling Tower
-------
The heat flow from the water to the air is dependent only upon the mech-
anisms of conduction and convection; no evaporation takes place within the tower. The
heat exchanger is generally a fin-tube, cross-flow type.
The direct contact spray condenser consists of a steel shell into which
the turbine steam is exhausted. Cool water from the cooling tower is introduced into
the condenser as a spray of water droplets from a series of evenly distributed nozzles.
Upon contact with the water spray, the steam condenses and is cooled to a temperature
nearly equal to the temperature of the spray water. The condensate and cooling water
collect at the bottom of the condenser. This water is then pumped to the boiler and
through the cooling tower as indicated in Figure 3.15. This type of condenser is used
because it has a much higher thermal efficiency and can be made much more compact
than conventional condensers. However, since the cooling water mixes with the con-
densate, the cooling water must be satisfactory for use in the boiler.
The pressure in the condenser is always well below atmospheric. In order
to prevent air leakage into the cooling water through joints in the piping and heat ex—
changer, the water is pressurized to a level higher than atmospheric by a pump. This
represents an additional auxiliary power cost for the systems. However, in large
systems a hydraulic turbine can be placed before the condenser to recover some of
the energy required for pressurization.
3. 6. 3 Theory of Operation
The heat transfer between the water and the air in the heat exchanger portion
of the system, when a cross flow heat exchanger is used, is governed by the following
equation. (See Figure 3.16.)
(T - T )-(T - T )
Q UAF 1 3 2 4 (3-22)
G
Li’ T
2 4
69
-------
Q
A
FG
T 1
T 2
T 3
T 4
= heat transfer rate, (Btu/hr)
= heat exchanger surface area,(ft 2 )
= cross-flow correction factor,( 1.0)
= inlet water temperature, (° F)
= outlet water temperature, (° F)
= inlet air temperature, (°F)
= outlet air temperature, (° F)
U = overall coefficient of heat transfer, (Btu/hr-ft 2 -° F)
Air Cat in 2 , T 4
Figure 3.16.
Water Out
rñ
T 2
where
1
‘1
Water In
T
1
q
1”
Air
T 1
in 2 , T 3
> T 2 , T 4 > T 3
Cross Flow Air Heat Exchanger
70
-------
The following relations are also obtained from heat balances upon the
water and air, respectively
‘i C 2 (T 1 - T 2 ) (3—23)
C 2 (T - T 3 ) (3-24)
where = mass flow rate of water
mass flow rate of air
C 1 specific heat of water at constant pressure
C 2 specific heat of air at constant pressure
Equations (3-22), (3—23), and (3-24) can be used in two basic ways: to solve
for the heat transferred by a specified heat exchanger (UAFG) given ‘ , ‘ ‘ T 1 and T 3 ,
or to determine a design which will transfer a given amount of heat (Q) for given values
of the same variables. When the unit is forced draft, the air flow rate is known
from the operating level of the fans. However, when a natural draft system is used,
m 2 depends upon T 4 , T 3 and the airflow resistance of the heat exchanger and tower
walls.
For a given plant, a value of rej is known for each value of fti 1 , and T 1
(Ta). From this, a heat exchanger and cooling tower can be designed to transfer the
proper amount of heat for each value of T 1 . To develop a system which will function
properly under adverse-conditions, the designis usually based upon high values of Q and
T 1 , and the highest expected value of T 3 for the geographic location of the plant.
The actual design calculations for designing a dry cooling tower, particularly
a natural draft unit, are more involved than those given here. These equations are in-
tended to provide an introduction to the concept. References 3. 33, 3. 34, 3. 35, and 3. 36
discuss in greater detail the engineering aspects of these heat exchangers.
3.6.4 Size Estimation
Equations of the type presented in Section 3. 6. 3 are used to estimate the
size of the dry cooling tower required to reject a specified amount of heat to air at a
71
-------
given ambient temperature. In order that the plant function properly under most at-
mospheric conditions, a design air temperature near the highest expected value for
that locality is used. The system can then be expected to perform effectively for all
lower temperatures, since the tower will be able to reject more heat than required
under those circumstances.
Because of the high cost of a dry cooling tower system compared to con-
ventional cooling towers, particularly close attention is paid to the plant costs.
It does not appear to be economical to replace the cooling equipment of an existing
plant with a dry cooling tower because either the tower would have to be extremely
large, or the plant would not operate at its design point. The best design requires
an overall economic optimization study, as discussed further in Section 3. 6. 5.
There are only a few power plants in operation which use dry cooling
towers, so it is difficult to predict what size cooling towers are necessary for given
power requirements. However, the dry cooled system at Rugeley, England should
serve to illustrate the expected component sizes for a medium size power plant.
This particular plant generates 120Mw using steam at 1,500 psig and 1,000°F re-
heated to 1,000°F between turbine stages. The hyperbolic cooling tower, built of
reinforced concrete, is 350 feet high with a base diameter of 325 feet and a throat
diameter of 205 feet. The heat exchanger has a total frontal area of 80, 000 square
feet and will reject 5. 87 x io8 Btu/hr. The turbine operates at a vacuum of 28. 7
inches Hg. The corresponding saturation temperature is about 80° F. A plant built
to operate almost anywhere in the United States would require slightly higher exhaust
pressure and temperature to avoid unreasonably large exchanger areas since our
average yearly temperatures run higher than those in England.
3. 6. 5 Cost Information
A dry cooling tower is generally larger and more expensive than a con-
ventional (wet cooling tower. As a result, the design of a particular power plant
should be preceded by an extensive economic optimization study. Although it is
technically possible to build a dry cooling tower to produce turbine exhaust pres-
sures as low as those obtained with water cooled condensers, this is usually not
72
-------
e onomically feasible. Instead, the system as a whole must be analyzed to determine
the design which will result in the lowest cost per kilowatt generated, taking into ac-
count the added expenditures associated with decreased vacuum levels.
One optimization approach is to treat heat exchanger area and air through-
flow rate as independent variables. Associated with these variables are condenser cost,
tower cost, piping and construction cost, turbine price for various operating vacuum
levels, pump and fan power costs, turbine power output at various vacuum levels, and
the capital value of the plant power output. The procedure is to seek values of the in-
dependent variables which will minimize the power generation cost ($/kwh). An ex-
ample of this technique has been presented by Oplatka (Ref. 3.37). For the particular
cost criteria he selected, the variation of cost per kilowatt-hour as a function of con-
denser size is illustrated in Table 3. Note that the optimum condensing pressure is
greater than that generally used for water cooled condensing plants.
Table 3
ECONOMIC OPTIMIZATION OF DRY COOLING TOWER SYSTEM
Condenser
Surface Area
too Small
Optimum
-
Condenser
Surface Area
too Large
Relative Price of Condenser and
Accessories
0.8
1.0
1.21
Pressure in Turbine Exhaust, inHg
3. 56
2. 64
2. 07
Difference in Power at Turbine
Shaft Compared with Optimum,
% of Rated Output
-2. 35%
+1.84%
Fan Power as % of Turbine Rating
1.6%
2. 0%
2.4%
Difference in Cost of Generation
Compared with Optimum
Cents/kwh
±0.0018
+0.0012
An optimization study of this type should be performed for all proposed
power plant designs. However, such an analysis cannot be carried out for an arbitrary
plant. The entire calculation is dependent upon the relative price of materials, labor
73
-------
and capitalized cost of power output, plus local temperature conditions for the specific
location of interest. At the present time an insufficient number of plants utilizing dry
cooling towers have been built to permit the determination of average costs. Oplalka
has presented some relative cost data comparing river cooled, wet tower cooled and
dry tower cooled plants (Ref. 3. 38). These data are contained in Table 4.
An economic study is also necessary to best determine the location and
cooling technique for a new power plant. When cooling water is available in large
quantities at low cost, the dry cooling tower is not competitive with once-through
cooling and conventional wet cooling towers. However, the dry cooling tower concept
relaxes the requirement for an ample water supply. Land sites which are less ex-
pensive and closer to sources of fuel can now be considered. An entirely new approach
is needed to determine the best site and appropriate cooling technique to be used.
Oplatka (Ref. 3. 38) has developed a method for comparing three types of
cooling for a particular plant: direct condenser cooling using river water, a water
cooled condenser with a conventional cooling tower, and a “Heller System’ direct
contact spray condenser with a dry cooling tower (Oplatka refers to this as an air
cooled condenser; we reserve this classification for conventional condensers reject-
ing heat directly to the air). Each system is required to generate the same amounts
of power, 750 Mw. A conceptual plant is developed for each of the three cooling tech-
niques, each design being optimized to produce the lowest cost per kilowatt-hour.
Costs are then assigned to each plant including capital outlay and auxiliary power
requiremement.s. but neglecting the cost of providing cooling water. From these
data, one can determine ranges of water cost for which each system is desirable.
The cost of water should include facilities for pumping, handling and treatment e.g.,
cc. truction and maintenance of a dam and reservoir). The proper type of plant de-
s ’--i can thus be determined for a specific combination of location, water cost and
re juired power output. By extending this type of study to a number of potential sites
which may possess different land costs, cooling water costs and fuel availability, an
economically optimum choice can be made. Other factors such as thermal or secondary
pollution effects can be incorporated into the analysis by assigning additional costs
related to the magnitude of the undesirable characteristic.
74
-------
I II III
Cooling By Wet Cooling Dry Cooling
River Water Tower Tower
1. Pressure at Exhaust Bar. inches Hg 1.50 1.90 3.50
2. Net Output of Turboset After Subtracting
Power to Drive Feed Pump, Mw 750 750 750
3. Power for Cooling Water Pump, Mw 1 3
4. Power for Pumps and Fans in Dry
Cooling Tower, Mw - - 12
5. Efficiency 0.409 0.405 0.394
6. Extra Capital Outlay Compared to I, io6 $ 2.48 3.07
7. Capitalized Value of Power for Pumps and
Fans Compared tol, i0 6 $ 0.16 0.83
8. Capitalized Value of Efficiency Difference
Compared to I, 106$ 0.60 2.18
9. Total Extra Capital Outlay Compared to I, i0 6 $ 3.23 6. 09
10. Total Extra Capital Outlay Compared to TI, TO 6 $ 2.84
Table 4
COMPARISON OF THREE COOLING TECHNIQUES
-------
3.6. 6 Maintenance Requirements
Operational experience at Rugeley has shown that very little maintenance
is required by a dry cooling tower system. Originally, it was feared that the circulating
water would absorb aluminum and iron from the heat exchanger system, leading to boiler
scale and corrosion of the heat exchanger and piping interior surfaces. After operation,
however, internal corrosion of the heat exchanger was undetectable and the aluminum
content of the cooling water was very low, of the order 0.02 parts per million. There-
fore, water quality control for such a system is no more difficult than that required for
a conventional water cooled condenser plant.
Exterior corrosion on the heat exchanger surface was found to be a problem
of the Rugeley plant. This was attributed to atmospheric pollution, high humidity and a
high concentration of salt in the air. Protective coatings for the heat exchanger surfaces
are presently being developed which promise to eliminate the problem. External cor-
rosion has not been a problem with dry cooling towers in less humid climates, such as
Hungary, which implies that corrosion protection may only be necessary under par-
ticularly severe climatic conditions.
3. 6. 7 Secondary Pollutant Considerations
There is no evidence that the use of a dry cooling tower system leads to any
significant secondary pollution. Since the cooling water flows in a closed loop through
the plant and heat exchanger, the only external effect is to increase the temperature of
the air passing across the heat exchanger surface. Fog and misting, which can be a
problem with conventional cooling towers, cannot occur because there is no water
evaporated into the air stream. Boiler blowdown can introduce contamination if dumped
into a river, but this is associated with the boiler and has no relation to the cooling tech-
nique being used.
3. 7 Evaporative Condensers
3. 7. 1 Introduction
The evaporative condenser, like the cooling tower, uses the latent heat of
vaporization of water to achieve cooling. Existing units of this type have capacities
76
-------
ranging from 2 to 375 tons; no units of the capacity required by a power plant are
known to exist at this time. Three hundred-seventy five tons would be sufficient
to cool a 750 kw nuclear plant. They are generally used with industrial water
coolers and air-conditioning systems, particularly in areas with water shortage
problems.
3. 7. 2 General Description
Figure 3. 17 is a diagram of an evaporative condenser. The sub-
stance to be cooled enters the coils in a vapor state, is condensed, and
leaves the coils as liquid. Atmospheric air is pulled or blown by fans so that it
flows over the outside surface of the coils. A continuous water spray is directed
over the coils by a separate water circulation system. Heat is neither added nor
taken from the circulating water, so it reaches an equilibrium temperature which
depends upon the operating conditions. This temperature is always above the dew
point of air, which causes the water wetting the coils to evaporate into the air stream
absorbing the latent heat of vaporization from the coils (approximately 990 Btu/pound
of evaporated water).
A big disadvantage with this arrangement is the fact that the ratio of
vapor specific volume to liquid specific volume is several orders of magnitude
greater for steam than for any of the common refrigerants. For this reason,
steam is condensed on the outside of tubes in a condenser since very large diam-
eter piping would be required for condensation on the inside. It is possible, how-
ever, that a power plant using a fluid other than steam may find an evaporative
condenser attractive.
77
-------
r
Liquid Out
1
Air In
G 1 , h 1 , w 1
Makeup
cY
./
Q)
ci
Q.T
Figure 3.17. Schematic Diagram of an Evaporative Condenser
Fan
Air Out
G 1 , h 2 ,w 9 ..ui*
Vapor In
Eliminators
)
/
)
Water Tank
Pump
78
-------
3. 7. 3 Theory of Operation
The heat flux transferred from the condensing refrigerant to the air is
given by the following equation:
-f A/O.24 G
H=G(h_h 1 )(1_e g ) (3-25)
where A = outside area of condenser tubes, (ft 2 )
f = coefficient of heat transfer between air and
water. (Btu/hr-ft - DF)
G = mass flow rate of air, (lbm/hr)
h = enthalpy of air at a weI bulb temperature equal
to the spray water temperature, (Btu/lbm)
= enthalpy of entering air,( Btu/lbm)
H = heat flux, (Btu/hr).
In order to use Equation (3-25), the enthalpy h must be determined,
which requires a Imowledge of the spray water temperature. The overall heat trans-
fer coefficient from the refrigerant to the air (U) is given by the relation below:
1 1 B
+ -i—-- (3-26)
w R
where B = ratio of outside to inside tube surface areas
= heat transfer coefficient between refrigerant and
tube wall,( Btu/hr—ft 2 —” F)
f = heat transfer coefficient between water and tube
w
wall, (Btu/hr—ft’ —° F).
79
-------
By balancing the heat flux across the tube wall, and neglecting the thermal resistance
of the tube material, the equation below is obtained.
I -f A/O.24G
T —T
h-h IJA
w I
where TR = condensing temperature of the refrigerant, (°F)
T = spray water temperature, (° F)•
By creating a modified psychrometric chart, as illustrated in Figure 3.18, Equation
(3-27) can be solved graphically. The angle U is defined by
I —f A/O.24G
)
tanU = UA
Point A in the figures is defined by the condensing temperature of the refrigerant (TR)
and the entering air wet bulb enthalpy (h 1 ). From point A, a line is drawn upward
at an angle of 6 from the vertical. The intersection of this line with the saturation
line locates point C. The coordinates of point C are T and h . Thus, a knowledge
of 1 g’ A, G, U, TR. and T 1 is sufficient to determine the spray water temperature,
and the air enthalpy at a wet bulb temperature equal to the spray water temperature.
With h known, Equation (3-2 5) is used to calculate the heat flux from the refrigerant
to the air.
80
-------
o
/
I I
85 -7;;9 .c
Condensing Temperature - °F
Figure 3.18. Modified Psychrometric Chart for Determining the Air
Enthalpy at a Wet Bulb Temperature Equal to the
Spray Water Temperature
8O
/
75
81
-------
3.7.4 Size Estimation
For a power plant, the heat rejected is known as a function of condensing
temperature. Hence, the equations above are used to determine the condenser design
parameters (e. g., A and G) so that the condenser will supply the proper condensing
temperature and reject the required amount of heat. A discussion of heat transfer co-
efficients can be found in Reference 3.44. Judging from presently available equipment,
the evaporative condenser is a very compact unit since it combines the function of a
condenser and cooling tower.
3. 7. 5 Cost Information
Since evaporative condensers are not built in sizes large enough for even
small power plants, no reliable cost information is available. However, Stoeker
Ref. 3.41) states that: “Cooling towers usually dominate installations larger than
several hundred tons of refrigeration. In sizes below several hundred tons, the two
compete vigorously.”
With comparison to an air cooled condenser, Wile (Ref. 3.44) says: “The
air-cooled condenser will have higher operating costs due to the higher operating pres-
sure and generally larger fan motor as compared to the fan motor and pump of the evapo-
rative condenser.”
A very rough approximation to costs for very large units may be made by ex-
trapolating manufacturers’ costs of present day units. Such costs may be approximated
as a straight line on a log-log plot of cost vs. unit size. The results are shown in Fig-
ure 3.19.
3. 7. 6 Maintenance Considerations
One of the principal problems in operating the evaporative condenser is the
buildup of excessive salt concentrations in the circulated water. Martin (Ref. 3.42)
recommends adding an additional quantity of makeup water on the order of one gallon
per hour per ton into the spray system while subtracting a like amount from the sump.
82
-------
I I I
±T1i
:
I.
A 0
-
::: :
,
i :
I ’ z :w
:
F
i: 1
i:
7 B 9j
2
4
B
plant size - Mw
Figure 3.19. Capital Cost of Evaporative Condensers for
Nuclear (77 = 33 ) Plants
83
6 7
7 ,
2
—4
1 ci
4:
2
Cl)
0
(I )
0
0
1c
6
5
-------
A second problem is scale formation. It has a deleterious effect upon heat
transfer (“fouling factor”) and excessive amounts can affect proper distribution of the
spray water. Observations in the field (Ref. 3. 43 indicate, however, that evaporative
condensers are less sensitive to scale formation than shell and tube condensers unless
very heavy deposits are allowed to accumulate. The use of large water quantities is a
simple method of reducing scale buildup and corrosion.
3. 7.7 Secondary Pollution Considerations
There was no evidence of secondary pollution problems particular to the
evaporative condenser reported in the literature. Ground fogging might be a problem
in a very large unit since the height dimensions and air flow rates would be comparable
to a mechanical draft cooling tower. The possibility of polluting a stream with the con—
centrated salt bleed water should also be considered.
3. 8 Air Cooled Condensers
3.8.1 Introduction
The air cooled condenser rejects heat directly to air flowing over its ex-
terior surface. Although this name is used by some authors when referring to dry
cooling towers, we shall treat these as two separate classes of cooling devices.
3. 8.2 General Description
A diagram of a basic air cooled condenser is presented in Figure 3.20.
Steam flows through a long coil or matrix of tubes which have finned exterior surfaces.
Heat transfer from the hot steam through the tube walls and fins to the cool outside air
causes the steam to condense. In refrigeration service, the vapor usually enters the
condenser in a superheated state, requiring that the condenser desuperheat the vapor
as well as condense it. The steam entering the condenser of a steam power plant is
usually a high quality binary mixture.
84
-------
Air Out
I’
Finned Tube Surface
Vapor In
_______- Liquid Out
/
Air In
Air In
I _______
Figure 3.20. Schematic of Air Cooled Condenser
The heat transfer characteristics of an air cooled condenser are similar
to the heat exchanger portion of a dry cooling tower; heat is transferred from the sub-
stance within the tubes to the exterior air stream. The basic difference is that vapor
enters the condenser, is condensed, and possibly subcooled, whereas the heat
exchanger portion of a dry cooling tower just cools a liquid within the tubes. Due
to condensation, the air cooled condenser must transfer directly to the air the
85
-------
large amount of latent heat of vaporization (approximately 1,000 Btu/lbm, depending
upon the condensing pressure) as well as the heat corresponding to the amount of
liquid cooling.
Wet and dry cooling towers may be either forced or natural draft types.
Because of the poor heat transfer characteristics of air, air cooled condensers nearly
always use forced draft. The increased heat transfer rate to a fast moving air stream
driven by fans permits size reductions which more than offset the cost of the air moving
equipment.
It is difficult to obtain air cooled condenser information which is applicable
to systems as large as thermal power generating plants. Most existing air cooled con-
densers are used with small refrigerating and air conditioning systems where compact
size is a much more important design parameter than it would be for a power plant. A
few large air cooled condensers are being used with power plants in Europe. One par-
ticularlv successful unit is part of a 48 Mw power plant at Volkswagenwerke, Wolfsburg,
Germany. A 3 Mw plant using an air cooled condenser is also in operation at Wyodak,
\Vvoming. Their results have been so successful that a 20 Mw air cooled system is
being considered for that location.
3. 8.3 Theory of Operation
An air cooled condenser may contain a series of three heat exchangers.
The first acts as a desuperheater, lowering the steam temperature to the condensation
temperature. This step would be rare in a steam power plant, however. The second
is a condenser, transferring the latent heat of vaporization from a constant temperature
liquid-vapor mixture between an initial state of pure vapor to a final state of pure liquid,
and the final section acts as a subcooler, lowering the water temperature below the con-
densing temperature. The degree of desuperheating required depends upon the entering
steam temperature, while the amount of subcooling depends upon the condenser size
and the air temperature.
Each portion of the condenser transfers heat according to the heat exchanger
equation as given below
86
-------
( T - T )-(T 2 - T 4 )
Q = UAFG < (3-28)
in [ i ::]
where T 1 water temperature entering heat exchanger section
T 2 water temperature leaving heat exchanger section
T 3 air temperature entering heat exchanger section
T 4 air temperature leaving heat exchanger section
U heat transfer coefficient
A surface area of heat exchanger section
FG design section, experimentally determined for a
particular heat exchanger configuration.
In the first section, T 1 is the entering steam temperature and T 2 is the
condensing temperature corresponding to the operating pressure of the condenser.
In the second section, T 1 = T 2 condensing temperature, but Q is known from the
heat of vaporization corresponding to the condenser pressure.
= hfg (3-29)
where th 1 is the mass flow rate of steam and hfg is the enthalpy difference between
saturated steam and saturated vapor. In the third section, T 1 is the condensing tem-
perature. The theory of air-cooled heat exchangers is discussed in more detail in
Section 3. 6. 3. Various design considerations and air cooled condensers can be ob-
tained from numerous tests on heat transfer and fluid mechanics.
3.8.4 Size Estimatation
There are no standard techniques for sizing air cooled condensers since
so few units of power plant size have been built. Data from small units cannot be ex-
trapolated since they are designed for minimum size and usually operate at refrigerant
condensing temperatures of approximately 105° F. Condensers for power plants operate
at lower temperatures (70 - 80° F) for efficient system performance, and an economic
design would be based on criteria other than minimum size.
87
-------
The size of the air cooled condenser at Wyodak, Wyoming should serve
as an example for a small power plant. The power output is rated at 3 Mw. The
condenser is built in four sections, each section composed of a single layer of one
inch diameter tubes thirty feet long. The tubes are spaced such that each layer is
eight feet wide. The fins on each tube are 2-1/4 inches in diameter, giving a total
exterior surface area of 101, 048 square feet. Six fans, driven by 40 horsepower
motors, develops an air flow of 3,180,000 lbm/hr. The cooling area of the condenser
is approximately 35. 0 square feet per kilowatt.
The plans for the 20 Mw unit are based upon the same ratio of cooling area
per kilowatt. The turbine is expected to operate at an exhaust pressure of 6. 0 inches
of Hg. This value, higher than usually used for power plants, is more economical when
the cost and effectiveness of the air cooled condenser are considered.
3. 8. 5 Cost Information
Except for the cooling system, the cost of components (turbine, boiler,
etc.) will be about the same for plants using air cooled condensers or cooling towers.
However, the capital cost of an air cooled condenser is approximately twice as much
as a water cooled condenser and conventional cooling tower. This larger initial in-
vestment can be justified for applications where the air cooled plant operating costs
are considerably lower. The operating cost of a conventional cooling tower (including
pumping and water treatment) is approximately 0.35 mills/kw as opposed to a cost of
0.1 miU/kw for the additional fuel need when operating an air cooled condenser. This
figure does not include the fan power, however, which is 1-1/2% of generator capacity.
In addition, when the available water is of particularly poor quality, the water treat-
ment cost for a cooling tower may be as much as 75% higher, as was the case atwyodak.
Converse (Ref. 3.22) reports that an air—cooled condenser was considered
for the Veri on, Vt. plant. Data from GEA Luftkulilergesellschaft yielded an initial cost
figure of $15. 6/kw (approximately 2x the cost for the cooling system with the ND tower).
He also estimated electrical production would cost an additional 0.44 mills/kwh when
the air-cooled condenser was used.
Other factors which enhance the desirability of air-cooled condensers are
related to their low water requirement. The Wyodak plant, for example, is constructed
88
-------
at the mouth of a mine which produces sub-bituminous and lignite coal. Although this
coal is very cheap to produce, the heating value per pound is so low that it is not eco-
nomical to transport or store it. By producing power at the mine mouth, great savings
in fuel costs can be realized.
As water costs and treatment requirements increase, the air cooled
condenser system gains an edge over the river or cooling tower cooled plants. The
water at mine mouth locations is generally very hard and acidic. Even when large
amounts are available, treatment costs for a cooling tower become prohibitive. Thus,
to utilize the low fuel cost of a mine mouth plant, air cooled condensers and dry cool-
ing towers are particularly attractive.
3. 8. 6 Maintenance Requirements
Maintenance procedures for an air cooled condenser are primarily limited
to the air moving equipment. Periodic oiling of fan motors, lubrication of bearings
and adjustment of driver mechanisms is necessary. Occasional cleaning of the ex-
terior finned surfaces to remove dirt and dust is suggested for efficient operation.
Since the water flows in a closed loop through the boiler, water treatment is needed
only to maintain a water quality which will not corrode the boiler or cause scaling.
This degree of purification is sufficient to prevent corrosion and scaling of the in-
terior surfaces of the condenser coils.
A wind deflector may be desirable when prevailing winds are likely to
interfere with the discharge of air passing through the condenser system. Corrosion
of the exterior surface will not be a problem if non-ferrous materials are used. When
steel is required, a protective coating, such as those being developed for dry cooling
tower heat exchangers, can be applied.
3. 8. 7 Secondary Pollution Considerations
The air cooled condenser introduces no secondary air or water pollution.
The circulating air is heated as it passes over the condenser surface, but the humidity
ratio is not changed. Noise generated by the fans and the air flow may be a nuisance
in some locations although the sound level is about the same as that generated by a
conventional mechanical draft cooling tower.
89
-------
3.9 Advanced Concepts
3. 9. 1 Introduction
There are several large—scale heat rejection schemes which may be clas-
sified as advanced concepts. Discussions have appeared in the open literature and,
in some cases, models have been built and tested. None, however, has been utilized
in a full scale plant so cost and operating information are not available.
Also included in this section is a discussion of some new materials for
use in cooling towers. Strictly speaking, this is not an advanced concept but it is
possible that new materials may reduce capital cost, construction cost, and/or
maintenance costs. Such changes may alter the economic balance and favor an al-
ternate heat rejection scheme.
3. 9. 2 The Froth Contact Heat Exchanger
A schematic of a froth-contact heat exchanger is shown below.
Air Out
Process
Fluid
Sieve Plate
In this device, first described by Poll and Smith (Ref. 3.51), the process stream flows
inside tubes which are submerged in a pool of water. Air is blown up into the water
through the sieve plate creating a bubbly froth which covers the tubes. The process
involves both latent and sensible heat transfer although the major portion is accom-
plished by latent transfer. Higher overall heat transfer coefficients than in conventional
evaporative condensers have been observed because of the highly agitated liquid in close
contact with the tubes and the high interfacial area between the air stream and the liquid
in the froth.
90
-------
Tater requirements for a froth contact exchanger are comparable to other
evaporative cooling devices. Slightly less may be necessary since preliminary tests
indicate that a somewhat greater portion of the heat may be transferred as sensible
heat. This is not yet definitely established, however. It is not necessary to circulate
the froth water, but makeup to replace that lost by evaporation and to keep the level
of dissolved solids and pH under control is still required. Much less air is required
than for an equal capacity air-cooled condenser. According to Smith and Poll (Ref.
3. 51), “The amount of air required by a froth—contact heat exchanger for a given load
is, at a maximum, one-quarter of that required by the equivalent air-cooled unit and
is normally much less.
Several people in this country and in Britain are interested in the froth
contact exchanger, primarily for use in refrigerant condenser service. However,
due to the problems encountered when attempting to condense steam on the inside of
tubes, a more likely function in a power plant would be replacement of the dry cooling
tower in a “Heller system” Section 3.6). These dry cooling towers are quite large
and it is conceivable that a froth contact device could reduce the size appreciably
with only a small increase in the amount of water required.
3. 9. 3 Ellipsoidal Cooling Tower
The following discussion is taken from Reference 3. 52. To make cooling
towers less conspicuous, the tendency has been to make them broader rather than
taller. However, a point has been reached where performance may be adversely
affected by wind. Under certain circumstances the wind flowing over the leeward
rim of the tower may cause a down draft. This may happen in either wet or dry
towers although it is more likely to occur in a dry tower because of the large
diameter. Secondly, observations at Rugely have shown that winds adversely
affect the heat transfer ability of the tower by causing poor airflow distributions.
To overcome this problem, engineers at Britain’s Central Elecricitv
Research Laboratories are working on a new semi-ellipsoidal tower as shown on
the following page.
This novel idea evolved from the knowledge that flow over an elliptic air-
foil separates at a point about 60 to 70 percent of the chord from the leading edge.
Model studies have shown this to be true for the ellipsoidal shape tower since the
91
-------
separation point is downstream of the exit. Downdrafts and warm air recirculation
are thus eliminated and the models indicate that performance is likely to improve as
a result of the small but significant suction produced by the flow over the shell.
The design is applicable to wet or dry towers but present thinking indicates
that the radiators of a dry tower could be arranged on a shallow inverted cone with its
apex at the center of the tower base. Air would flow in and under the radiators as shown
in the sketch above. It is estimated that the cooling capacity of a 300 foot high 1500 foot
diameter dry tower would be sufficient for the needs of a 2000 Mw station. CERL en-
gineers are testing a 1/40th scale, 36-ft—dia. wooden model using electrically heated
radiators to test aerodynamic and thermodynamic efficiency of the design (Ref. 3.68).
CERL emphasizes that this project is still in an early stage of development and many
problems must be solved before a full scale installation is attempted.
3.9.4 Rotating Arm Cooling Tower
An unusual cooling tower design is described in Reference 3. 53. A sketch
of the unit is shown below. The force due to reaction of the curtain of water directed
at the packing propels the distributor arm at 25 to 30 rpm. The plastic fortified wood-
cellulose packing has numerous tiny vertical openings throughout its entire depth of only
18 inches. More space is saved by a special baffle attached to the trailing edge of each
distributor arm only four inches above the fill which acts as a drift eliminator.
o icc 1 W FEET
L I
92
-------
Because of these innovations the tower takes only about one quarter the
space of more conventional cooling tower designs. It combines efficient cooling in
the range of 120 to 500tons (Gx 106 Btu/hr) with minimum weight and space.
[ F:n
_________________________________ Rotating Arm
_____________ - _____________
I ________________ _________________
Hot Water ____________
H Cooled Water
Sump —j________________________
The results obtained so far are encouraging; the improved heat and mass
transfer result in smaller towers. Unfortunately, the cooling capacity of the units
described is several orders of magnitude below that necessary for a moderate size
power plant.
3.9.5 Rotating Cooling Tower
The following description is taken from Reference 3.53.
The rotary cooling tower has a revolving circular packing composed of
plastic extrusions. Tip speed of the packing is about 1200 fpm. The packing has
thousands of passageways traveling from a central core to the outside, following
the involute of a circle. Water is sprayed onto the packing core from a stationary
distributor running down the center of the core. Centrifugal force forces the water
through the passageways as a thin film. As it leaves the outside of the revolving
packing, the water strikes a honeycombed louvre which deflects it to the sump.
An induced-draft fan, mounted at the center of the core, draws air
from the outside through the louvres and then through the passageways to the core
93
-------
where it is discharged to the atmosphere. The high velocity (30 fps) turbulent air
stream flows countercurrent to the thin water film. The high relative velocity prom-
ises increased heat transfer rates and resultant decreases in size and weight.
3. 9.6 Powered Hyperbolic Towers
The powered hyperbolic tower has been suggested as a solution to the
problems of unfavorable climate and need for better control associated with the ND
hyperbolic tower (Refs. 3.69 and 3. 70). The design (see Figure 3.21) is basically
a natural tower with supplementary air moving capacity supplied by cells of axial
flow fans mounted around the base. According to Reference 3. 70 the capacity of an
ND hyperbolic tower was doubled when fans were installed.
The benefits claimed for the design include:
1. More uniform outlet water temperature throughout the year.
2. Reduced ground area.
3. Improved control, enabling operation of associated equipment
at higher efficiency.
The operating curve at constant heat load, range, and relative humidity is shown below.
THE EFFECT OF WOT ON CWT FOR -
PURE ND AND POWERED HYPERGOLC
*ERS AT CONSTANT HEAT LOAD.
COOLING RANGES AND RELATIVE I& Oiry
‘5
I.
• • - -- 4 15
/ Pf NT 0 PANS AT FULL S ED
50 10 10
WET BULR TEMPERATURE. F
94
-------
Shell
Figure 3.21. Sketch of Powered Hyperbolic Tower
Eli min tors
Fan Housing
Profile
1 ._
Diffuser Pond
95
-------
3.9.7 New Materials
The utilization of new materials is aimed at improving the operational and
maintenance side of the economic picture. Longer, trouble-free part life is obtained
at slightly higher original cost. In the following paragraphs examples of the use of
various modern materials are presented. The list is not intended to be exhaustive
but merely to illustrate present day trends in material selection.
Big improvements have been made in fill materials. The high humidity,
high temperatures, and concentrations of acids and salts contribute to rapid deteriora-
tion of the fill. Chemical treatment improves the service life of the standard redwood
fill but presents the possibility of stream contamination. Hard-burned clay acid-
resistant) and high—impact plastics are being considered as alternatives for splash
type packing.
In film tpe packing, where the water must wet the surface of the fill.
several improvements are on the horizon. Plastic-fortified cellulose sheet which
looks like corrugated cardboard is being tested. It is easily treated to improve
fire resistance and to check growth of algae. Asbestos—cement sheets are already
being extensively used. They have a rough absorbent surface that wets well. A
stack of these sheets 5 feet high is equivalent to a 20 to 30 foot depth of splash packing.
Minimum expected life is 15 to 20 years. Metal and plastics are also being formed
into wavy sheets. A rather thick layer of water is formed in each of the hollows. Any
solids precipitated during evaporation remain in the film and are washed down into
the sump. Further water treatment may be necessary with this type.
Considerable effort has been made to cut down the noise generated b the
fan. Some manufacturers are leaning toward squirrel-cage fans since they are usually
quieter. However, they operate most efficiently in a specific speed range not corn-
monlv used in cooling towers, so power requirements may be high. Large multibladed
fans using glass-reinforced plastic blades are also being designed. The favorable
damping qualities of the plastic tends to eliminate high frequency vibrations that
generate noise.
96
-------
The exterior of almost all new cooling towers is composed of ceramic or
concrete. Even smaller units are constructed in this manner. The durability of these
materials combined with the scarcity of satisfactory lumber has been a deciding factor.
Fiberglass is gaining favor in several areas. Drift plates are now made
of fiberglass; it allows greater flexibility in design. Fan stacks are also made of
fiberglass. Fiberglass requires virtually no maintenance, no painting or treatment
is necessary. A small all plastics cooling tower is described in Reference 3.54.
97
-------
REFERENCES - SECTION 3
3. 1 Edinger, J. E. and Geyer, J. C., “Heat Exchanger in the Environment,”
Edison Electric Institute Publication No. 65-902, 1965.
3.2 Shade, W. R. and SmitI H. F., Discussion of “Economic Considerations in
Thermal Discharge to Streams,” National Symposium on Thermal Pollution
Part II, August 14 — 16, 1968.
3. 3 Kovats, A., “The Economics of Condenser Circulating Water Supply in Power
Stations,” ASME Paper 67-WA/PWR-2.
3.4 Bauman, H. C., “Fundamentals of Cost Engineering in the Chemical Industry,
Reinhold Publishing Corp., New York, 1964.
3. 5 Harbeck, G. E., “The Use of Reservoirs and Lakes for the Dissipation of
Heat,” U. S. Geological Survey: Circular 282, Washington, D. C.’ 1953.
3. 6 McKelvey, K. K., and M. Brooke, “The Industrial Cooling Tower,” Elsevier
Publishing Company, Amsterdam, 1959.
3.7 Nelson, \V.L., Oil and Gas Journal , 47, No. 18, 1948, p. 73.
3. 8 Utlev, E. E., “Lake Size Based on Cooling Needs of Four Unit Plant,” Elec.
World , Vol. 166, No. 6, August 8, 1966.
3. 9 Derrick, A. E., “Cooling Pond Proves to Be the Economic Choice at Four
Corners,” Power Eng’g. , Vol. 67, November 1963.
3.10 Hori, S., Patchett, U. A. and Boelter. L. M. K., “Design of Spray Cooling
Ponds,” ASVHE Journal Section, Heating, Piping, and Air Conditioning ,
October 1942.
3. 11 Badger, W. L. and McCabe, W. L., “Elements of Chemical Engineering,”
2nd Ed. McGraw-Hill Book Co., New York, 1936.
3. 12 Perry, J. H., “Chemical Engineers Handbook,” 3rd Ed., McGraw-Hill
Book Co. , New York, 1950.
3.13 Robinson, P. and Roll, C. S., Thesis, MIT. 1922.
3. 14 DeMonbrun, J. R., “Factors to Consider in Selecting a Cooling Tower,”
Chem. Eng’g. , Sept. 9, 1968.
3. 15 Merkel, F ..“Verdunstimgskuehlung, ” VDI Forschungsarbeiten No. 275,
Berlin, 1925.
3. 16 Baker, D. R., and Shryock, H. A., “A Comprehensive Approach to the
Analysis of Cooling Tower Performance,” Jouriial Heat Transfer , August
1961.
98
-------
3. 17 Stoecker, W. F., “Refrigeration and Air Conditioning,” McGraw-Hill Book
Co., Inc., New York, 1958.
3.18 Baker, D.R., and Mart, L. T., “Cooling Tower Characteristics as Deter-
mined by the Unit-Volume Coefficient, “ Refrig. Eng. , Vol. 60, No. 9, P. 965.,
September 1952.
3. 19 Engalitcheff, J., Couriterfiow Cooling Tower Selection and Performance Chart,
Refrig. Eng. , Vol. 64, No. 12, p. 35, December 1956.
3.20. Koiflat, T. D., “How to Beat the Heat in Cooling Water,” Electrical World ,
October 14, 1968.
3. 21 Hollmeier, R. J., “Risk Criteria Method for Selection of Cooling Towers to
Control Steam Temperature,” Proceedings of the Am. Power Conf., Vol.
XXIX, 1967, pp. 441-447.
3.22 Converse, A. 0., “Thermal Energy Disposal Methods for the Proposed Nuclear
Power Plant at Vernon, Dartmouth College, Hanover, New Hampshire, 1967.
3. 23 Pacific Northwest Laboratories, Nuclear Power Plant Sorting in the Pacific
Northwest for the Bonneville Power Administration, Contract No. 14-03-67868
(Battelle Memorial Institute, Richiand, \Vas hington). 1967.
3.24 “Keystone Cooling Towers Will Be Record Breaking,” Power Engineering ,
September 1964.
3.25 Paige, P. M., “Costlier Cooling Towers Require a New Approach to Water—
Systems Design,” Chemical Engineering , July 3, 1967.
3.26 Berg, B-and Larson, T. E., “Estimation of Costs of Cooling Tower Installation
and Operation,” Proceedings of the American Power Conference, Vol. XXV,
1963, pp. 678—686.
3.27 Lane, R. W., Larson, T. E., “The Role of Water Treatment in the Economic
Operation of Cooling Towers,” Proceedings of the American Power Conf.,
Vol. XXV.
3.28 Jones, D. B., “Can You Afford Your Cooling Water Treatment Program,”
Materials Protection , March, 1965.
3. 29 Willa, J. L., “Effect of Maintenance on Cooling Tower Performance,” Materials
Protection , October, 1964.
3. 30 “Cooling Towers Get Bigger and Better As Water Demands Grow,” Power Eng’g. ,
Vol. 69, No. 12, December 1965.
3.31 Bailiff, N. L., “Material Standards and Testing for Water Cooling Towers,”
Heating, Piping, and Air Conditioning , March 1966.
99
-------
3. 32 Forbes, M. C.., “Cooling Towers Not Unmixed Blessing in Pollution Control,”
The Oil and Gas Journal , April 24, 1967.
3. 33 Cheshire, L. J., Daltry, J. N., “A Closed Circuit Cooling System for Steam
Generating Plant,” The South African Mechanical Engineer , February 1960.
3. 34 Christopher, P. J.,, “The Dry Cooling Tower System at the Rugeley Power
Station of the Central Electricity Generating Board,” English Elec . , Vol. 20,
N.1, January-February, 1965.
3. 35 “Dry Cooling Tower at Rugeley,” The Engineer . Vol. 214, No. 5579, December
28, 1962.
3. 36 Heller, L., Forgo, L., “Recent Operational Experiences Concerning the ‘Heller
System’ of Air Condensation for Power Plants. Latest Results of Developments,”
Sixth World Power Conference, Melbourne, October 1962.
3. 37 Oplatka, G., “Air-Cooled Condensing Plants,” Brown Boveri Review, Vol. 49,
No. 7/8, August 1962.
3. 38 Oplatka, G., “Economics of Air-Cooled Condensers in a 750 Mw Installation,”
Brown Boveri Review, Vol. 53, No. 3, March 1966.
3. 39 Goodman. W .,“The Evaporative Condenser,” Heating, Piping and Air Con-
ditioning , Vol. 10, No. 3, March 1938.
3.40 Thomsen, E.G., “Heat Transfer in an Evaporative Condenser,” Refrig. Eng’g. ,
May 1946.
3. 41 Rish, R. F., “The Design of a Natural Draught Cooling Tower,” mt. Heat
Transfer Conf., 1961, Vol. V-A pp. 951—958.
3.42 Martin, R. J.. “Evaporative Condensers - - Their Uses in Water Shortage Areas,”
Refrig. Eng’g. , Vol. 55, No. 5, p. 452, 1948.
3.43 Wile, D. D., “Evaporative Condenser Performance Factors,” Refrig. Eng’g. ,
Vol. 58, No. 1, p. 55. January 1950.
3.44 Wile, D. D., “Evaporative Condensers,” Heating, Piping, and Air Conditioning ,
August 1958.
3.45 Edwards, R. C., “Air Condensers with Gravity Flow,” Heating, Piping and Air
Conditioning , October 1958.
3.46 Nussbaum, 0. J., “Air-Cooled Condensers,” Heating, Piping, and Air Con-
ditioning , September 1958.
3.47 Nussbaum, 0. J., “Air Cooled Condensers For Large Tonnages,” Industrial
Refrigeration, March 1955.
100
-------
3.48 Parce, J.Y., “Cooling Water Scarce? Use Air!” Electric Light and Power,
November 1967.
3. 49 Viessman, W., “How, When, Arid Where to Use Air Cooled Condensers,”
Heating, Piping, and Air Conditioning . September 1956.
3.50 Wolf, J. L., “Economic Evaluation of Condensing Methods,” Heating, Piping,
And Air Conditioning , August 1958.
3.51 Poll, A., and Smith, W., “Froth-Contact Heat Exchangers,” Chemical Eng’g. ,
October 26, 1964.
3. 52 “A New Squat Shape for Cooling Towers,” Engineering , Vol. 199 No. 5160,
March 12, 1965, p. 328.
3.53 Elonka, S., “Cooling Towers,” Power Special Report, March 1963.
3.54 “All Plastics Cooling Towers:’ Oil, Engine, and Gas Turbine , Vol. 30, January
1963, p. 294.
3. 55 Steur, W. R., “Artificial Cooling Can Broaden Choice of Sites for Today’s New
Steam Stations,” Electrical World , Feb. 1962, pp. 42—43.
3.56 G. J. Eicher, “Cooling Lakes can be a pleasant solution,” Electrical World ,
April 14, 1969, pp. 90—92.
3.57 Molyneux, F., “Counter- and Cross-flow Cooling Towers,” Chemical and
Process Engineering , May, 1967, pp. 56-60.
3.58 Lowe, H. J., and D. G. Christie, “Heat Transfer and Pressure Drop Data
on Cooling Tower Packings, and Model Studies of the Resistance of Natural
Draught Towers to Airflow,” mt. Heat Transfer Conf., 1961, Vol. V-A.
3. 59 “World’s Largest Hyperbolic Tower Cools a Power Plant,” Engineering News-
Record , May 4, 1967, pp. 58-59.
3. 60 “Electric Utilities Include Built-in Safeguards for Environmental Protection,”
Environmental Science and Technology , June, 1969, pp. 523-525.
3. 61 Smith, A. F., and R. E. Bovier, “Hyperbolic Cooling Towers with Reservoir
Storage of Makeup to Serve the Proposed Keystone Generating Station,” Proc.
of the Am. Power Conf., Vol. XXV, 1963, pp. 406-415.
3. 62 Ravet, Louis, C., “Natural-Draft Cooling Towers - The Shape of Things to Come,”
Power Engineering , Feb., 1963, pp. 50—53.
3. 63 Lockart, F. J., et al, “Cooling Towers for the Power Industry,” Proceedings
of American Power Conference, 17: 320, 1955.
101
-------
3. 64 Knowlton, A. E., “Looking Over the Statistics of Cooling Towers,” Electrical
World, 142 , No. 10, Sept. 6, 1954, pp. 91-98.
3. 65 Malley, R. J. and Gasper, K. F., “Bleed-off is Essential in Recirculating
Cooling Water Systems,” Materials Protection , October, 1964, pp. 38-44.
3. 66 Willa, J., “Cooling Towers: Step Children of Industry,” Heating, Piping & Air-
Conditioning , January, 1966, Pp. 153-158.
. 67 Smith, W. W., “Corrosion Resistant Materials for Cooling Tower Hardware,”
Cooling Tower Institute Bulletin TPR-126 , July, 1962.
3.68 “Squat Tower Cuts Cooling Costs,” Engineering News Record , March 16, 1967,
p. 72.
3.69 Campbell, C., “A New Look at Cooling Towers for the Power Generation
Industry,” presented at Cooling Tower Institute Meeting, January 19-22, 1969,
Houston, Texas.
3.70 Davidson, W. C., “Tower’s Cooling Doubled by Fan-assisted Draft,” Electrical
World , March 25, 1968, pp. 19—21.
3. 71 Buss, J. R., “How to Control Fog From Cooling Towers,” Power , Jan., 1968,
pp. 72-73.
3. 72 Calbert, J. D., “Cooling Reservoir Study,” Power Engineering , Dec., 1967,
pp. 52-53.
3.73 Hansen, E. P. and J. J. Parker, “Status of Big Cooling Towers,” Power
Engineering , May, 1967, pp. 38-41.
3. 74 Kelly, A. G., and N. R. Lawless, “Economic Sizing of Cooling Towers,”
Combustion , August, 1962, pp. 41-47.
102
-------
Section 4
HEAT REJECTION EQUIPMENT - OPERATIONAL CONSIDERATIONS
4.1 Introduction
Off-design operation and other operational considerations affect both the
capability and economics of heat rejection systems. Ambient conditions (tempera-
ture, relative humidity, amount of sunlight, wind, and rainfall) can alter the op-
erating point of the system. Conditions within the system itself may change; water
flow rate (or velocities), air flow rate or heat transfer coefficient may be a function
of the operating point. Unusual problems which may affect the system performance
are also discussed in this chapter. Clogging of nozzles, buildup of slime and algae
on packing, and ice formation on inlets and fan blades all lower system effectiveness
and result in higher operating costs.
4. 2 Once-Through Cooling and Cooling Ponds
Changes in water flow rate, wind, and ambient conditions can alter the
performance of these systems. A reduction in the flow rate raises the temperature
of a riveç for example, even though the upstream temperature remains constant
since the water temperature rise at the discharge is given by
H
temperature rise =
mC
p
Thus, for a given H, the temperature rise increases as ñi decreases. Changes in
river flow rate also effect the exponential temperature decay. A stream with a
given overall exchange coefficient and equilibrium temperature will have both a
greater average depth and mean velocity when passing a greater flow rate. This
increases the value of a in the equation
T(x) (TO_TE)e +TE
resulting in a decay to equilibrium temperature within a shorter distance.
103
-------
The energy exchange coefficient (K) and equilibrium temperature (TE)
are functions of the meteorological conditions. The value of K increases as relative
humidity, wind, and equilibrium temperature increase. The effect of these factors
on the determination of the equilibrium temperature is more complicated and is de-
pendent on several variables which are nonlinear functions of equilibrium temperature
in addition to the energy exchange coefficient, air temperature, and net radiative input.
The equations for TE may be combinedto yield
0.26T +e ± C43)
.051TE 2 +KTE = 1 -1801±(K-15.7) [ 0.26 a ]
The overall exchange coefficient is
K = 15.7 + (0.26 ÷ 3)bW.
The functions p and C (3) can be represented by the following equations:
= 1.965 x 10 (TE)’
—6 3.49
C 3) = 5.9 x 10 (TE)
It can be seen that while qualitative statements as to the effect of various parameters
can be made, quantitative statements can be made only after numerical investigation.
In summary, it may be generally said that increased wind increases the
effectiveness of rivers and ponds for cooling while increased sunlight, air temperature
ind relative humidity decrease their effectiveness.
4.3 Spray Ponds
Much of the discussion in the preceding section applies to spray ponds as
well. In addition, the water flow rate, as determined by the pumping capacity, is a
104
-------
variable which may be affected by head (water depth), clogging of nozzles and pipe-
lines, arid deterioration of components of the pumping systems. Data is not available
on, say, flow rate versus time in use. Obviously, a reduction in the flow rate will
lower the effectiveness of the spray pond. A restricticm in the nozzle or line will
also lower the effectiveness by lowering the height of the water spray which decreases
the droplet exposure time. This problem can be redi.iced through proper maintenance.
In extreme circumstances, icing may occur around the nozzles and on con-
trol valves. Depending on the circulation patterns within the pond, algae and slime
can accumulate on various parts and cause fouling.
Increased winds may increase carryover substantially. Since the water
droplets are pushed higher into the air, particles of water may fall into regions where
they will not flow back into the pond. This problem is not likely to affect the capacity
of a large pond, however.
4.4 Wet Cooling Towers
4. 4. 1 Off-Design Performance
In a given natural draft tower off—design operation may stem from changes
in water loading or changes in ambient conditions. The available data seems to in-
dicate that tower effectiveness is lower at reduced water loadings particularly in
natural draft towers due to decreased air-water relative velocity. Mechanical draft
towers have an advantage in this respect since they are able to maintain satisfactory
air flow with reduced water loadings. If the heat exchange capacity falls off, the
condenser back pressure will increase resulting in reduced plant capacity.
The following items may be considered ambient condition variables: wind,
humidity, and wet and dry bulb temperatures Wind can have a considerable effect on
range and heat rejected. Strong winds seem to adversely affect NDtower performance
by creating poor airflow distribution. The effect is more pronounced on the windward
side of the tower. Crawshaw (Ref. 4. 1) found that the effect of wind was stronger 2t
reduced water loadings. There is a small increa5e in draft but this is outweighed b
the poor heat transfer due to the nonuniformities. If the pond is outside the tower and
exposed to the winds, water loss will be increased, but the water will be cooler.
*Relative humidity is, of course, fixed once the wet and dry bulb temperatures are known.
105
-------
Operation at a given wet-bulb temperature is more efficient when the
relative humidity is higher due to the greater density difference. This is especially
true in an ND tower. If ambient wet-bulb temperature, heat load, and water circula-
tion rate are held constant for a given tower, the higher relative humidity results in
cooler discharge water. The range will be the same, but the condenser pressure will
be lower. Lower wet-bulb temperatures have the same effect. The driving force for
heat transfer is greater dn both ND and MD towers) and the density difference is
greater so that the ND towers operate more efficiently due to the higher velocities
through the packing. With regard to this, it may occur that peak loads will occur in
the summer due to high use of air conditioning, for example. This is an unfavorable
situation for an ND tower since the wet-bulb temperatures are also highest then. The
combined effects of relative humidity and wet-bulb temperature are shown in the figure
below (taken from Ref. 4. 7 __________________________
0 0
S
a
. !
a
S
•0 O .o
5 (1 UL T( P (PATU E.
On very tall towers, appreciable vertical temperature gradients may be
r esent. The variations inside the shell are of secondary importance. However, changes
in temperature over the packing can produce considerable changes in frictional re-
sistance and L/G ratios. Outside the tower, temperature inversions and vertical
lapses reduce the range of the unit. Studies to date indicate barometric pressure
has no measurable effect.
THE (EFECTS OF *(1’ BU1 Tf D€ ATU5E
AY V( WUM O T Y
( ‘ OE’.S TV T E t .C ( OR V G FQSC(
- t: ”
106
-------
The frictional resistance of the tower is a function of the L/G ratio. Lowe
and Christie (Ref. 4. 8) found that the pressure drop across many of the packings doubled
as the water loading went from 0 to 3000 lb/hr—ft 2 at low velocities (approximately 3 fps).
The effect was less pronounced at higher velocities (7-Sfps). The spray above and below
the packing also causes a pressure loss. To illustrate the effect of L/G ratio on the
losses as the following data on a 238 high ND tower from Reference 4. 9 is presented.
The loss values are given in number of velocity heads:
L/G 1.0 2.0 3.0
packing 32.7 34.9 36.5
spray 3.8 9.6 14.8
eliminator 5.2 5.2 5.2
outlet 10.5 10.5 10.5
inlet 13.5 13.5 13.5
columns, etc. 2.3 2.3 2.3
Total 68.0 76.0 82.8
The packing resistance is due to skin friction over the water film and a form drag on the
ripples observed on the sheets. The spray loss is assumed to be due to water droplets
falling through the air stream above the packing and between the sheets. It is a very
strong function of L/G. The last four losses are constant since the air flow was held
constant.
4. 4. 2 Unusual Characteristics
Mechanical draft units seem to exhibit more problems than the natural draft
units. Ground fogging has a greater tendency to occur with the mechanical units since they
have much lower heights. The lower height also results in greater warm air recirculation
which leads to problems of fan blade erosion and moisture condensation in gear boxes as
well as reduced efficiency.
During cold weather operation, ice may form on inlets of both natural and
mechanical draft units. The problem is more serious with the MD units since ice can
actually break fan blades. Natural draft towers have been known to exhibit a mode of
operation in which ice forms, is melted by the hot water, and then forms again. Such
operation is not desirable and design precautions such as piping the hot water to regions
around inlet should be taken.
Large natural draft towers do not have a history of producing fog or drizzle.
Plumes are usually dissipated a few hundred feet from the tower and the great height
keeps them well above ground level so recirculation is not a problem.
107
-------
4. 5 Dry Cooling Towers
4. 5. 1 Off-Design Operation
A dry cooling tower is designed to supply the required degree of cooling
at the highest expected ambient air temperature. At lower temperatures the cooling
rate is increased unless control techniques are applied. To maintain a fixed value
of turbine exhaust pressure the airflow through the heat exchanger can be decreased
by adjusting shutters in the tower, or by removing sections of the heat exchanger
from servicewhenlarge cooling reductions are required. Controlling the airflow
is generally sufficient unless freezeup becomes imminent, at which point a reduction
in the heat exchanger area is necessary. A cooling rate larger than design is not
serious; in fact, it leads to a higher power output by decreasing the turbine exhaust
pressure. However, for overall economic reasons, it may be best to maintain the
vacuum level at the design value by controlling the volume of air passing through
the tower.
When the air temperature rises above the design value, the turbine exhaust
pressure and temperature increase until the temperature of the water in the tower
heat exchanger is high enough that the tower can transfer to the air the amount of
heat rejected in the condenser. The increased turbine exhaust pressure reduces
the system efficiency. However, since the system is designed for the highest ex-
pected air temperature, the probability of occurrence of such an off-design operating
condition is very small.
4.5.2 Operating Procedure
Freezing of the circulating water inside the heat exchanger during cold
weather is a potential hazard with the dry cooling tower system. This is avoided by
designing the heat exchanger in the form of separate units with control valves to put
each individual unit out of service as desired. As the ambient air temperature de-
creases, the number of cooler segments actually in use is decreased by progressively
removing more of them from service and draining them. At the Rugeley plant, an
alarm is given when the water temperature drops to 45° F. The operator then reduces
the effective surface area by actuating control valves which remove a heat exchanger
section from use. 108
-------
It has been shown that a particular plant should be designed to operate at
an optimum turbine exhaust vacuum level. As the ambient air temperature changes,
it is necessary to alter the cooling tower performance level to maintain the proper
vacuum. Under extreme conditions, this is done by putting segments of the heat ex-
changer in or out of service as is done to prevent freezeup. However, for minor
adjustments control is maintained by adjusting the air flow rate through the tower.
Changing the fan speed will accomplish this in a mechanical draft unit. With natural draft
towers a series of shutters within the tower can be positioned to vent the air flow to
the outside of the tower before it has traveled through all the active heat transfer
elements. Since the system is generally designed to operate properly at the highest
expected ambient temperature, reducing the air flow through the heat exchanger
provides a versatile method of controlling the turbine exhaust pressure.
4. 5. 3 Unusual characteristics
The most significant characteristic of the dry cooling tower system is its
small water requirement. The paper Dry Cooling Towers” (Ref. 4.4) presents the
following estimated water requirements, which includes boiler make-up.
Once through water cooling 600 gpm/M v
Conventional Cooling tower 30 gpm/Mw
Dry cooling tower 2 gpm/Mw
Clearly, the water requirement for a dry cooling tower is an order of magnitude less
than that for a wet cooling tower, and several orders of magnitude less than that needed
for a system using once—through cooling. Thus, a dry cooling tower is particularly
attractive in locations where water is scarce or expensive. Less expensive land areas
where water would have to be piped in for a conventional plant are particularly attrac-
tive locations. In arid regions, dry cooling tower or air-cooled condenser plants may be
the only practical systems for power generation.
The winter operation of a dry cooling tower is very simple, requiring only
that the proper number of cooling surfaces be disconnected to prevent freezeup. Since
there is no moisture added to the circulating air, ice formation on fan blades or inside
109
-------
the tower is not a problem, as it is with a wet cooling tower which requires direct
contact between the air and the cooling water.
Since the dry cooling tower requires a very small amount of make-up
water, a minimum amount of water treatment is required. This is an added ad-
vantage over the conventional cooling tower, which requires thirty times as much
make-up due to evaporation.
The “Heller System” does require controls not needed with a conventional
cooling tower. Regulators must be installed which will close the spray valves if the
circulating water pumps fail; otherwise, the condenser would flood. In addition, a
throttle valve in the spray system is needed to prevent water hammer when the spray
is shut off.
4. 6 Evaporative Condensers
4.6.1 Off-Design Performance
The effects of changes in the wet-bulb temperature of the entering air and
condensing temperature are illustrated in Figure 4. 1. Clearly, the unit is able to
handle higher loads as the wet-bulb temperature decreases. The capacity of the unit
may also be increased by increasing the fan speed. The larger volume of air will raise
the capacity but the extent depends on the refrigerant loading (tons per 100 sq. ft. of
external area). Goodman (Ref. 4.1) indicates that, for a given U and fg the loading
has a maximum value which is reached asymptotically as shown in Figure 4. 2. Ob-
viously, an increase in the air flow rate at operating point A will produce a larger
gain in capacity than a corresponding increase at point B.
When a refrigerant is used, the vapor entering the condenser is some-
times superheated. This does not have a noticeable effect on the operation, how-
ever. The only adverse effect seems to be a slight increase in scale formations.
Present experimental evidence indicates that normal increases in water flow
rate tend to increase the condenser rating. At high water flows, turbulence in the water-
film and impinging spray increases the water film coefficient and the overall coefficient.
The inside film coefficient remains reasonably constant.
110
-------
130
120
:110
100
90
80
70
60
50
Figure 4.1.
I. . :
cs
0
0
Temperature
Figure 4. 2. Evaporative Condenser \Vater Loading
as a Function of Air Flow
115
N
N
Nc ’
io s
N
95
90
60 65 70 75 80 85 90
wet bulb temperature of entering air - ° F
Performance Curves of an Evaporative Condenser
(Condensing Temperatures Are For Refrigerant- 12)
A
CFM PEJ TON
111
-------
4.6.2 Qperational Procedures
For a very large plant, it is possible to combine several units to obtain
the desired capacity. Water and vapor flow distribution problems tend to crop up,
though, so care must be taken in controls design. At the present state it would re-
quire a very large number of these units to satisfy the cooling requirements of a large
power plant (about two units per Mw).
During cold-weather, it might seem advisable to drain the water from an
evaporative condenser and operate the coil dry. Figure 4. 3 shows, however, that
even at extremely low temperatures the dry coil capacity is below that possible with
wet-coil operation.
‘4 ,
Jc —
‘O 0 ‘0 ZO 30 40 Sc’
y .3 ,j 5 TEMP( 4TCIRE
O ENT5R/N A/
Figure 4. 3. Capacity of an Evaporative Condenser
Operating with a Dry Coil
4. 6. 3 Unusual Characteristics
Undesirable characteristics associated with evaporative condenser are
generally associated with the fact that a refrigerant is being condensed in the tubes.
For example, very warm days result in high gas discharge temperatures which tend
to deteriorate the oil and lead to reduced compressor life. This, of course, would
not be a problem in a conventional power plant where the condensate is returned to
the feedwater pump.
112
-------
4. 7 Air Cooled Condensers
4.7.1 Operating Procedures
Except for periodic maintenance, the air cooled condenser requires no
operational control procedures. Although freezeup may seem to be a possibility,
no problems of this nature haveoccurred with systems presently in operation.
Thermostatic controls which close air shutters or louvres as the ambient tempera-
ture decreases below freezing maintain the condensate temperature above the freezing
point. The air cooled condenser plant at Wolfsburg, Germany has performed satisfac-
torily with ambient air temperatures ranging from - 7°F to 90°F.
4.7.2 Off—Design Operation
Since both the dry cooling tower and the air cooled condenser transfer heat
to the air by conduction and convection heat transfer, they operate in basically the
same manner under off-design conditions. The air flow can be reduced when the am-
bient temperature is low, or sections of the condenser can be taken out of operation
under particularly cold conditions, usually when the air temperature is below freezing.
Because most air cooled condensers are of the forced draft type, the air flow can be
controlled by either changing the fan power or adjusting shutters and vents.
When the air temperature rises above the design value, the turbine exhaust
temperature and pressure rise to new equilibrium values just as occurs with a dry
cooling tower system. However, if the fans can be operated at a speed greater than
design without damage, the cooling rate of the condenser can be increased during
short periods of very high air temperature. Hence. it may be possible with some
systems to maintain the proper system operating level by driving the forced draft
fans at higher speeds.
4.7.3 Unusual Characteristics
The most significant characteristic of the air cooled condenser is its low
water requirement; only boiler make-up is needed. As a result, it has many of the
desirable characteristics of the dry cooling tower system. The economic aspects of
a system which needs very little water is discussed in the section on dry cooling
towers in Section 3.6.5.
-------
Air leakage into the condenser can be a problem, since the condensing
pressure is always well below atmospheric. Leaks of this type are much more
difficult to detect than are leaks in the “Heller System where water leaking out-
side the coils should be readily visible.
The biggest drawback to large scale air cooled condenser is the large
piping required. Steam has a very small density (0. 0053 ibm/ft 3 at 79°F and a
pressure of 1.0 inches of Hg) compared to water (62.4 lbm/ft 3 ). To pass a given
mass flow rate at similar velocities, the steam pipe would have to have an area
approximately 10, 000 ames as large as the water pipe. Even though the steam will
flow at a greater velocitly for the same pumping power, the steam pipe will still
have to be several orders of magnitude larger than that required for a water pipe.
Attemptlng to use smaller steam pipes would require very high head pressures to
move the steam, which would necessitate a large amount of pumping power. The
“Heller System,” dry cooling tower concept was developed from efforts to avoid
this problem.
114
-------
REFERENCES FOR SECTION 4
4.1 Crawshaw, C. J., “Effect of Wind Velocity on the Performance of a
Natural Draught Cooling Tower,” The En neer , May 25, 1962.
4. 2 Crawshaw, C. J., ‘Investigation into Variations of Performance of a
Natural-Draught Cooling Tower, Proc. Instr. Mech. Engrs., Vol.
178 pt. 1 No. 3s.
4.3 McKelvey, K.K. and M. Brooke. “The Industrial Cooling Tower,”
Elsevier Publishing Company, Amsterdam, 1959.
4.4 Reti, G.R., “Dry Cooling Towers,” Proc. American Power Con-
ference, 1963.
4.5 Reti, G.R., “Dry Cooling Tower Shows Promise for Arid Areas,”
Power Engineering, April 1963.
4.6 Ritchings, F.A. and A.W. Lotz, “Economics of Closed Versus Open
Cooling Water Cycles,” Proc. American Power Conference, 1963.
4. 7 Campbell, J. C., “A New Look at Cooling Towers for the Power
Generation Industry,” presented at CTI meeting, January 19-22, 1969,
Houston, Texas.
4. 8 Lowe, H. J. and D. G. Christie, “Heat Transfer and Pressure Drop
Data on Cooling Tower Packings, and Model Studies of the Resistance
of Natural Draught Towers to Airflow,” mt. Heat Transfer, Conf.
1961, Vol. V-A.
4. 9 Rish, B. F., “The Design of a Natural Draught Cooling Tower,” mt.
Heat Transfer Conf.., 1961, Vol. V-A, pp. 951-958.
115
-------
Section 5
METHODS OF COMPARISON AND SELECTION
5. 1 Introduction
When comparing the heat rejection schemes discussed in this report, several
factors must be considered. Economics, unit size, local climate, and secondary pol-
lution are among the more important considerations. Some categories, esthetics, for
example, are less amenable to quantitative comparison than others making selection
somewhat more difficult. Fortunately, the factors presently considered most important
are easily defined and compared.
It is inevitable that certain assumptions be made in order to compare costs.
Whenever this was done, care has been taken to list the assumptions clearly and, if
possible, to estimate probable variations and their effect.
The systems discussed in Task I present various avenues of heat rejection
to the atmosphere. None of the schemes makes an attempt at waste heat utilization
which we feel is the ultimate solution so long as Rankine cycle steam plants are em-
ploved for power generation.
5. 2 Economics
Data has been gathered on capital and operating costs for the various
systems and has been presented in various sections of Chapters 3 and 4. This section
is a summary of the data obtained. In some cases data on installations of the large
size required for power plants were not available so estimates were obtained by ex-
trapolation of the present data.
Two alterations to capital cost were made in this section. The cooling
pond cost was set at $6. 50 /kw for fossil plants and $7. 501kw for the nuclear plants
to reflect the greater amount of land necessary for the nuclear plant. The spray
pond cost was raised to 58. 10/kw for the nuclear plant since larger pumps are re-
quired.
116
-------
A summary of the data collected in Task I is presented in Tables 5, 6, 7,
and 8. The capital cost includes the price of the condenser, its associated pumps and
piping, the heat rejection unit, and its accessories. The equipment cost was not given
as the increase over once-through cooling because some of the schemes (particularly,
dry cooling towers, evaporative condensers, and air-cooled condensers) use different
condensing arrangements. The cooling system may he then thought of as a complete
package which would be interfaced to the turbine exhaust and accomplish the condensa-
tion of the steam.
The operating costs listed in the tables are primarily power costs for pumps
and fans and costs for makeup water and treatment water. Power requirements were
established and converted to cost information assuming that electricity was available
at 4 mill/kwh. Figures for water loss, usually as a percentage of water circulated,
were used to determine power requirements for makeup water pumps.
The total cost column lists the equipment cost in dollars /kw-yr. The
average annual fixed charge rate was assumed to be equal to 9%. The figure for a
particular plant may be different since taxes and installation costs vary widely from
place to place. The capital costs were amortized over a thirty year period to obtain the
yearly figure.
Calculations were performed for both nuclear and fossil-fueled plants since
the rates of heat rejection are quite different. The nuclear plant was assumed to have a
thermal efficiency of 33% and an in-plant loss of 5 . The fossil-fueled plant was as-
sumed to have a thermal efficiency of 40% and in-plant losses of 15%. The heat re-
jected to the cooling water is then;
nuclear plant; 6410 Btu/kwh
fossil-fueled plant; 3840 Btu/kwh
Calculations were performed for each plant allowing two temperature rises (10° F and
20° F) through the condenser. The following schemes have costs which are functions
of the temperature rise of the cooling water through the condenser;
• once-through cooling
• cooling ponds
117
-------
:
Cooling System
Equipment
Clj)1t Ll
Cost
$/kw
Operating
Cost
$/kw-yr
Maintenance
Cost
$/kw-yr
Total Cost
$/kw-yr
Additional
Cost to
Consumer
%
Once Through Run-of--River
5.30
0.59
-
0.59
1.66
0
Once Through Estuary
6.30
0.59
0.59
1.75
0.05
Cooling Pond
6.50
0.74
0.74
2.06
0.23
Spray Pond
7.60
1.18
0.59
2.46
0.46
ND Wet Tower River Makeup
8.50
1.18
1.18
1.18
1.44
ND Wet Tower ReservoirMakeup
11.50
1.18
1.18
4.44
1.59
MD Wet Tower
7.20
1.54
1.54
4.43
1.58
ND Dry Cooling Tower
20.0
1.19
1.19
5.58
2.24
MD Dry Cooling Tower
13.0
1.54
1.54
5.65
2.28
Evaporative Condenser
io.o
1.05
1.05
3.70
1.17
Air Cooled Condenser
is. 0
0.60
0.30
3.65
1.14
Table 5. Fossil Fueled Plant - T 10°F
-------
Cooling System
Equipment
Capital
Cost
$/kw
Operating
Cost
$/kw-yr
Maintenance
Cost
$/kw-yr
Total Cost
$/kw—yr
Additional
Cost to
Consamer
%
Once Through Rtm-of-River
5.00
0.30
0.30
1.05
0
Once Through Estuary
6.00
0.30
0.30
1.14
0.05
Cooling Pond
6.50
0.38
0.38
1.34
0.16
Spray Pond
7.60
0.60
0.30
1.58
0.30
ND Wet Tower River Makeup
7.50
0.60
0.60
2.92
1.17
ND Wet Tower ReservoirMakeup
10.50
0.60
0.60
3.20
1.23
MD Wet Tower
7.20
0. 94
0. 94
3.23
1.25
ND Dry Cooling Tower
20.0
0.60
0.60
4.40
1.91
MD Dry Cooling Tower
13.0
0. 94
0. 94
4. 45
1. 94
Evaporative Condenser
10.0
1.05
1.05
3.70
1.52
Air Cooled Condenser
15.0
0.60
0.30
3.65
1.49
Table 6. Fossil Fueled Plant - = 20°F
-------
Cooling System
Equipment
Capital
Cost
$/kw
Operating
Cost
$/kw-yr
Maintenance
Cost
$/kw-yr
Total Cost
$/kw-yr
Additional
Cost to
Consumer
%
Once Through Run-of-River
5.88
0.99
0.99
2.51
0
Once Through Estuary
6.88
0.99
0.99
2.60
0.05
Cooling Pond
7.50
1.24
1.24
3.16
0.37
Spray Pond
8.10
1.98
0.99
3.70
0.67
ND Wet Tower River Makeup
12.50
1.98
1.98
6.16
2.08
NI) Wet Tower ReservoirMakeup
15.00
1.98
1.98
6.38
2.21
MD Wet Tower
9.40
2.34
2.34
6.23
2.12
ND Dry Cooling Tower
22.0
2.00
2.00
7.38
2.78
MD 1)ry Cooling Tower
15.0
2.34
2.34
7.43
2.82
Evaporative Condenser
13.0
1.40
1.40
4.67
1.34
Air Cooled Condenser
17.0
1.00
0.50
4.43
1.10
Table 7. Nuclear Fueled Plant - T 10°F
-------
Cooling System
Equipment
C’ipital
Cost
$/kw
Operating
Cost
$/kw-yr
Maintenance
Cost
$/kw-yr
Total Cost
$/kw-yr
Additional
Cost to
Consumer
%
Once Through Run-of-River
5.24
0.50
0.50
1.47
0
Once Through Estuary
6.24
0.50
0.50
1.56
0.05
Cooling Pond
7.50
0.62
0.62
1.92
0.26
Spray Pond
8.10
1.00
0.50
2.23
0.43
ND Wet Tower River Makeup
11.50
1.00
1.00
4.08
1.49
ND Wet Tower Reservoir Makeup
14.00
1.00
1.00
4.31
1.62
MD Wet Tower
9.40
1.33
1.33
4.20
1.56
ND Dry Cooling Tower
22.0
1.00
1.00
5.38
2.24
MD Dry Cooling Tower
15.0
1.33
1.33
5.41
2.26
Evaporative Condenser
13.0
1.40
1.40
4.67
1.83
Air Cooled Condenser
17.0
1.00
0.50
4.43
1.70
Table 8. Nuclear Fueled Plant - L T = 20°F
-------
• spray ponds
• wet and dry cooling towers
Operating costs are, of course, lower when the temperature rise is higher because
less circulating water is required. Maintenance costs are also lower since the pumps,
etc., are smaller and replacement parts are generally less expensive and easier to
replace.
Evaporative condensers and air—cooled condensers remove essentially
the latent heat of the condensing steam so the size is specified for a plant operating
at a given condensing pressure. The unit size for a nuclear plant will be greater,
of course, due to the larger amount of heat rejected. Performance does vary with
changing ambient conditions, but these calculations are more extensive and will be
performed in Phase 11.
The total cost of some of the systems also includes the cost of running
at higher condenser temperatures. The percentage increase in power production
was calculated from the equation given by Converse.
increase = 0.2 (Td - T)
where Td = design condenser temperature
T = actual condenser temperature
Td was assumed to be 10O F 1. 93 inches Hg) for once-through cooling. The following
condenser temperatures were assumed for the other systems:
System T - F p - “Hg Power Increase-
Cooling Pond 100 1.93 0
Spray Pond
ND \Vet Cooling Tower 115 2.99 3. 0
MD Wet Cooling Tower 110 2.60 2. 0
Dry Cooling Tower 120 3.45 4.0
Evaporative Condenser 110 2.60 2.0
Air Cooled Condenser 120 3.45 4.0
122
-------
A figure of 0. 42 gpm/kw for circulating water rate fOf fossil fueled
plants has been used in several places in the report. This corresponds to a plant
with a thermal efficiency of 40% and a temperature rise through the condenser of
18°F. In this section T T s of 10°F and 20°F were assumed. The circulating water
rates are found to be:
Fuel gpm/kw
10 f 0.76
20 f 0.38
10 N 1.28
20 N 0.64
Since reliable information on maintenance costs was not available, the
maintenance cost was set equal to the power cost following the recommendation of
Lane and Larson except for spray pond and air cooled condensers. It was felt that
this method yielded figures which were excessive for these units.
The total cost column contains some interesting results. The costs of
cooling are much higher for the nuclear plants. Secondly, the power penalty attached
to some of the systems is evidently quite substantial. For example, in Table 6 the
capital costs for the spray pond and the ND wet tower with river makeup are ap-
proximately equal. The operating and maintenace costs are also about the same.
However, the 3% power penalty attached to the tower results in 0. 6% additional
cost to the consumer. It is possible that a power penalty should be attached to the
spray and cooling ponds as well.
There is also quite a difference between the towers which require reser-
voirs to be built and those which have natural water available for makeup and purge
duty. The additional cost is about $0. 25/ kw-vr. The dry towers do not require such
reservoirs but initial cost is very high. The air-cooled condenser is more expensive
than the wet towers but may prove to be the most economical choice when water is
scarce. The evaporative condenser is really an unknown quantity and, of course,
makeup and purge water are still necessary.
123
-------
The last column in Tables 5, 6, and 7 lists the additional cost of the cooling
system to the consumer. Generally the cost to the consumer decreases as the range is
increased except for the evaporative and air—cooled condensers which increase relative
to the other systems as the range is increased.
5. 3 Size and Capacity
There is considerable variation in the amount of land required by the various
cooling schemes. However, a direct comparison of the cooling schemes might he some-
what misleading if one is not familiar with power plant design. For example. one would
expect a considerable saving in space required when an evaporative condenser was
compared to a mechanical draft wet cooling tower. It is common practice to place
the condenser directly under the turbine—generator set. Installation costs would
prohably be higher for the MD wet tower scheme but the evaporative condenser would
probably be larger since the condensation occurs on the inside of the tubes so the two
units might have equal land requirements.
Generally, we can say that the size is inversely proportional to the heat
transfer ability of the unit. Improvements such as spraying the cooling water or in-
creasing the velocity of the airstream will result in decreased unit size. Units which
use only air for cooling are much less efficient than evaporative units.
It is possible that the last three schemes will not be practical on very
large stations due to the large amount of space required. Air-cooled condensers
are a particu1ar problem since they have both the problem of poor outside heat
transfer tube to air even with fans and the problem of trying to condense steam
on the inside of the tubes. Perhaps a binary cycle using a fluid on the low-end that has
a small ratio of specific volumes is a possible solution.
5.4 Geographic Limitations
Several geographic considerations have been mentioned in the preceding
sections. Extremes in temperature, relative humidity and other climatic conditions
such as excessive rainfall can affect the cooling systems. The location of a plant
can have a considerable effect on the economics. Plants located at mine-mouth, for
example, can usually use more costly cooling systems but can still be competitive
124
-------
due to lower fuel costs. Systems which use very little water are necessary there
since the water requires a large amount of treatment.
The availability of water has obvious economic implications. Runnoff may
be necessary to maintain the level of a cooling pond or to provide a supply for cooling
tower blowdown or makeup. The quality of the water is also important since water
treatment is expensive both in terms of dollars per pound of chemicals and increased
maintenance costs.
The topography of the region surrounding the plant may influence the
operation of the cooling system. A particular arrangement of mountains could
cause winds and air circulation patterns that force vapor plumes from cooling
towers to remain around the Plant. Poor draft patterns could also result. High
surrounding mountains could raise the cost of fuel transportation.
The most important climatic factors are temperature wet-and dry-
bulb), amount of rainfall, and wind. One must remember that evaporative cooling
devices can ‘approach” the wet bulb temperature of the air while the dry devices
can cool only to the air dry bulb temperature. This means that a dry cooling device
would be unable to maintain proper condensing pressure during a very hot summer
day. On the other hand there are no ice formation problems in winter months.
The effect of rainfall on the operation of the various systems has not
been studied. At this point we feel that this should not present a problem in any
part of the United States. The amount of wind has a negligible effect on mechanical
draft units. In natural draft units the wind does increase the draft somewhat but tends
to produce nonuniform air distributions which lower the heat transfer ability,
5 5 Availability
The following cooling systems described in Sections 3 and 4 have been
used in large power stations (500 - 1000 Mw) in the United States:
• once-through cooling
• cooling ponds
• ND and MD wet cooling towers
125
-------
Smaller stations have utilized MD wet cooling towers, spray ponds, and air-cooled
condensers. No one in the United States has used the dry cooling tower or evaporative
condenser concept in a power plant. In all of the most recent studies of situations
where once-through cooling was not practical, the ND wet cooling tower was found to
be the most economical. However, they still require large amounts of makeup and
purge water (about i. x gal/hr for a 1000 Mw plant) so that if it becomes necessary
to site plants in areas where water is not plentiful or requires much treatment, more
intensive studies will be made of the dry units.
126
-------
Section 6
SUMMARY OF TASK I
6. 1 Present Status
At this stage of the investigation a complete documentation of all available
techniques and equipment for rejecting large quantities of heat either to waterways or
to the atmosphere is in hand. Information is available whereby estimates may be made
of equipment size, capital cost, and operating cost, both in terms of maintenance dollars
and parasitic power drain. These estimates, however, are restricted primarily to so-
called conventional operating conditions, and are based on standard design procedures.
This information is useful primarily to provide baseline results relevant to
the simplest kind of solution to the thermal pollution problem. That is, these results
may be used to estimate the incremental operational and cost penalties involved in going
beyond run-of-river cooling.
6. 2 Relationship With Other Tasks
The output from Task II, Phase I, will be similar to that of this task, but
will deal directly with the power cycles. The results of the two tasks can then be com-
bined to map out a “state -of-the-art” operating region with associated cost factors for
power production and related thermal controls under ‘ ordinary” conditions; that is, for
the plant operating at conventional temperatures and pressures, utilizing conventional
cooling requirement of standard design.
In Task IV, Phase I. these results will be used as a starting point to deter-
mine what the desirable characteristics of a total—community power generation system
would be. These results will also serve as the base data to Task I, Phase II, in which
a model of the power generation process will be constructed, parametized, and op-
timization criteria developed.
The remaining tasks will deal with substantial advances in departures from
the state-of-the-art when operating conditions, now considered far off-design, will be
investigated. The present results will serve not only as a starting point for design mod-
ifications but also as a basis for comparision with existing cost and operational complexity.
127
-------
Acce .ion Number
W
2_ [ SubJect Fzeld & Group
050
SELECTED WATER RESOURCES ABSTRACTS
INPUT TRANSACTION FORM
5_jor anization
Dynatech P / O Company, 17 Tudor Street, Cambridge, Mass. 02139
L TitJe
o Survey of Alternate aetnous for Coo1 ng Condenser Discharge Water--
Large-Scale Heat Rejection Equimment
10 Authoi
John H. Carey
John T. Canley
John S. !‘iaulbetsch
16 (Project Dea ØzatIon
FWQA Contract 1 -12-’477
Pro ect Uo. 16130 OHS 07/69
Note
22 Citation
Environmental Protection Agency, Water Pollution Control Research Series
1 l30 OHS 07/69, pm. 127, July 1969, 31 fig., 8 tab., 86 ref.
2 J DescHplors (Starred First)
Thermal mower plants , cooling towers , heat, condensers
25 Identifiers (Starred First)
Cooling ponds
- Abstroct
In this remort, various tyoes of heat transfer ecuipment which would be
applicable to the task of dissipating condenser discharge heat from large
power olants are identified. Their performance is analyzed, performance
mrediction methods are presented, and the ca ita1 and operatinq costs are given.
The report begins with a brief review of the basic thermodynamics of the
Rankine power cycle in order to provide some basis for understanding the
effect of the heat rejection process on the total cycle. Information including
sizing procedures, capital and omerating cost, duty maintenance requirements,
and possible secondary pollution considerations are presented for candidate
systems such as once-through cooling (to serve as a base), cooling ponds,
smrav ponds, wet cooling towers, dry cooling towers, evaporation condensers,
and air-cooled condensers.
Operational considerations are discussed including, the theory of o:eration
at the design point, off-desiqn performance, running mrocedures, and any
unusual characteristics affecting the interface between the cooling system and
the power plant.
Finally, the material describing each of the individual units is inte-
grated and procedures and criteria are developed for the selection of an
optimum unit given a particular geographic location, capacity requirements,
and economic constraints.
Ahstract :r - 1n ,t (u( n
- OvrRtech RID 17 Tir - - ‘ ° ‘
C
SCN O DO: I O : ’ CLS SCENT IC 4DORMAT ON
N F T ’1ENT O Ti-IE INTE, OR
OT ON. 0. C.
fl 71 T C
GPO: 1 7O 407 - fl
------- |