project

“THERMAL DESIGN OF EVAPORATIVE CONDENSER” i
Project Report (181909) entitled
“Thermal Design of Evaporative Condenser”
Submitted by
Patel Simit J. (Enrollment No. 100110119009)
Joshi Mohit H. (Enrollment No. 100110119011)
Patel Raghav S. (Enrollment No. 100110119012)
Chauhan Ronit A. (Enrollment No. 100110119049)
Academic Year 2013-14
(Second Semester)
In partial fulfillment of the requirements for
Bachelor of Mechanical Engineering
Faculty Guide Industry Guide
Prof. Sankalp Kulkarni Dr. H. C. Patel
Department of Mechanical Engineering
G H Patel College of Engineering & Technology
Gujarat Technological University
Ahmedabad, Gujarat.

“THERMAL DESIGN OF EVAPORATIVE CONDENSER” ii
G H Patel College of Engineering & Technology
Charutar Vidya Mandal Institution
Vallabh Vidyanagar – 388 120
CERTIFICATE
Date: DD/MM/YYYY
This is to certify that the Project Report entitled,
“Thermal Design of Evaporative Condenser”, submitted by Patel
Simit J. (Enrollment No. 100110119009), Joshi Mohit H.
(Enrollment No. 100110119011), Patel Raghav S. (Enrollment No.
100110119012), Chauhan Ronit A. (Enrollment No. 100110119049) is
original work and literature used from other sources have been
acknowledged in the report. The Project Work is part of curriculum
of the degree of Mechanical Engineering at Gujarat Technological
University (GTU), Ahmedabad pursued during the second semester
of academic year 2013-14.
Place: V. V. Nagar Patel Simit J. (Enrollment No. 100110119009)
Date: DD/MM/YYYY Joshi Mohit H. (Enrollment No. 100110119011)
----------------------------------------------------------------------------------------------------------
This is to certify that the above mentioned “seminar” is studied and
presented with our coordination.
Prof. Sankalp Kulkarni Dr. Darshak Desai
Project Guide Head of the Dept.

“THERMAL DESIGN OF EVAPORATIVE CONDENSER” iii
ABSTRACT
Energy is one of the major inputs for economic development of the industry. Energy sector
assumes critical importance in a view of increasing energy need and requiring huge
investment to meet them. Reduction in energy consumption helps industry to reduce its
energy bill, improve system performance and enhance its competitive position. Energy
consumption also has a significant impact on human environment, so it is require finding
out ways of decreasing energy consumption.
This report contains the Calculation of Thermal Design parameters of Evaporative
condenser of refrigeration unit at Amul dairy, overview of refrigeration system
arrangement at project site, industry problem, and Calculation of Thermal Design of
Evaporative condenser. According to calculation, total condensing load is 1625.13 kw or
463TR. Tube outer diameter of condenser is selected from standard manufacturer’s
catalog which is 26.67mm. Total heat transfer surface area in DE superheating section is
8.15m2
and in condensing area it is 330.28 m2
.
It is our belief that it will be helpful to industry.

“THERMAL DESIGN OF EVAPORATIVE CONDENSER” iv
ACKNOWLEDGEMENTS
We are highly indebted to our college, G.H. Patel College of Engineering And Technology, for
providing us with this excellent learning opportunity which greatly helped us in honing our skills.
We would also like to thank AMUL ltd for their guidance and constant supervision as well for
their support in pursuing the project.
We are thankful and fortunate enough to get constant encouragement, support and guidance
from industrial guide Dr. Harikrishna C. Patel (Chief engineer of Amul dairy)
We owe our profound gratitude to our project head Dr. Hemant Thakkar and project guide
Prof. Sankalp Kulkarni who guided us all along, till the completion of our project by giving us
the necessary information for developing a good system.
Signature
Patel Simit J. (Enrollment No. 100110119009)
Joshi Mohit H. (Enrollment No. 100110119011)

“THERMAL DESIGN OF EVAPORATIVE CONDENSER” v
LIST OF FIGURES
Figure No Figure Description Page No
Figure 1 Heat flow diagrram of refrigerator and heat pump 1
Figure 2 Combined Cross Flow 5
Figure 3 Parallel Flow 6
Figure 4 Counter flow 6
Figure 5 Axial Fan 7
Figure 6 Centrifugal fan 7
Figure 7 Layout of -2 °C system of the plant 11
Figure 10 Chilled water cycle 12
Figure 11 Tubes arrangement 23
Figure 12 Outer look of evaporative condenser 24
Figure 13 Graph of correction factor 27
Figure 14 Arrangement of tubes 29

“THERMAL DESIGN OF EVAPORATIVE CONDENSER” vi
Symbols and Abbreviations
QL magnitude of heat removed from the refrigerated space at
temperature TL
QH magnitude of heat rejected to the warm space at
temperature TH
W net Net work input to the refrigerator
COP Co-efficient of performance
HP Heat pump
R Refrigerator
h Enthalpy
U Overall heat transfer co-efficient
hi convection heat transfer co-efficient tube side
ho convection heat transfer co-efficient shell side
Do Outside Diameter of tube
Di Inside Diameter of tube
Ai Inner side tube surface area
Ao outer side tube surface area
K Thermal conductivity of the material
∆Tm logarithmic mean temperature difference
Q Amount of heat transfer
SL Longitudinal pitch of tube sheet
SD Diagonal pitch of tube sheet
ST Transverse pitch of tube sheet

“THERMAL DESIGN OF EVAPORATIVE CONDENSER” vii
AT Area of transvers passage
AD Area of diagonal passage
V Velocity
Re Reynolds number
Nu Nusselt number
Pr Prandtle number
L Length
µ Viscosity
þ Density
Cp Specific heat at constant pressure
NTU Number of transfer units
€ Effectiveness
γ mass flow rate of water

“THERMAL DESIGN OF EVAPORATIVE CONDENSER” viii
CONTENTS
Abstract viii
Acknowledgement viii
List of Figures viii
Symbols and Abbreviations viii
Contents VI
Chapter : 1 INTRODUCTION
1.1
1.2
1.3
1.4
Project statements
Project goals
Refrigeration system
The Actual Vapour-Compression Refrigeration Cycle
1
1
1
4
Chapter : 2 Evaporative Condenser
2.1 Introduction 4
2.2
2.3
Evaporative Condenser
Typical Applications
5
9
Chapter : 3 Experimental setup
3.1 System Introduction 11
3.2 Observed data 15

“THERMAL DESIGN OF EVAPORATIVE CONDENSER” ix
Chapter : 4 Calculation of design parameters for evaporative
condenser
4.1
4.2
4.3
4.4
4.5
4.6
4.7
Design of tubes in condenser
Average Velocity of Ammonia
Heat transfer co-efficient for inside of tubes
Heat transfer co-efficient for outside the tubes
Overall Heat Transfer co-efficient
Calculation of Tube Areas of De-superheating and
Phase Change process
Calculation of air flow rate and water evaporation rate
18
20
21
23
25
26
29
Chapter : 5 RESULTS AND CONCLUSION
5.1 Results 31
5.2 Suggestions 31
5.3 Conclusion 32
FORM 1 Application For Grant Of Patent
FORM 2 Provisional Specification
FORM 3 Statement And Undertaking Under Section 8
References
33
37
39

INTRODUCTION
“THERMAL DESIGN OF EVAPORATIVE CONDENSER” 1
Chapter 1: Introduction
1.1 Project Statement
“Thermal Design Of Evaporative Condenser”
1.2 Project Goals
 To Understand the Concept of Evaporative Condensers.
 To Calculate Design parameters for Evaporative Condensers according to operational
limits of Plant.
 To suggest which Industry provides optimum design parameters for same.
1.3 Refrigeration system
A major application area of thermodynamics is refrigeration, which is the transfer of heat
from a lower temperature region to a higher temperature one. Devices that produce
refrigeration are called refrigerators, and the cycles on which they operate are called
refrigeration cycles. The most frequently used refrigeration cycle is the vapor-compression
refrigeration cycle in which the refrigerant is vaporized and condensed alternately and is
compressed in the vapour phase. Another well-known refrigeration cycle is the gas
refrigeration cycle in which the refrigerant remains in the gaseous phase throughout. Other
refrigeration cycles discussed in this chapter are cascade refrigeration, where more than one
refrigeration cycle is used; absorption refrigeration, where the refrigerant is dissolved in a
liquid before it is compressed; and, as a topic of Special Interest, thermoelectric refrigeration,
where refrigeration is produced by the passage of electric current through two dissimilar
materials.

INTRODUCTION
We all know from experience that heat flows in the direction of decreasing temperature, that
is, from high-temperature regions to low-temperature ones. This heat-transfer process occurs
in nature without requiring any devices. The reverse process, however, cannot occur by itself.
The transfer of heat from a low-temperature region to a high-temperature one requires special
devices called refrigerators. Refrigerators are cyclic devices, and the working fluids used in
the refrigeration cycles are called refrigerants. A refrigerator is shown schematically in Fig. Here
QL is the magnitude of the heat removed from the refrigerated space at temperature TL,QH is
the magnitude of the heat rejected to the warm space at temperature TH , and Wnet,in is the
net work input to the refrigerator.
Another device that transfers heat from a low-temperature medium to a high-temperature one
is the heat pump. Refrigerators and heat pumps are essentially the same devices; they differ in
their objectives only. The objective of a refrigerator is to maintain the refrigerated space at a
low temperature by removing heat from it. Discharging this heat to a higher-temperature
medium is merely a necessary part of the operation, not the purpose. The objective of a heat
pump, however, is to maintain a heated space at a high temperature. This is accomplished by
absorbing heat from a low-temperature source, such as well water or cold outside air in
winter, and supplying this heat to a warmer medium such as a house.
The performance of refrigerators and heat pumps is expressed in terms of the coefficient of
performance (COP), defined as
A comparison of both Eqs reveals that for fixed values of QL and QH. This relation implies
that COPHP = 1 since COPR is a positive quantity. That is, a heat pump functions, at worst, as
a resistance heater, supplying as much energy to the house as it consumes. In reality.

INTRODUCTION
Fig. 1 heat flow diagrram of refrigerator and heat pump
However, part of QH is lost to the outside air through piping and other devices, and COPHP
may drop below unity when the outside air temperature is too low. When this happens, the
system normally switches to the fuel (natural gas, propane, oil, etc.) or resistance-heating
mode. The cooling capacity of a refrigeration system—that is, the rate of heat removal from
the refrigerated space—is often expressed in terms of tons of refrigeration. The capacity of a
refrigeration system that can freeze 1 ton (2000 lbm) of liquid water at 0°C (32°F) into ice at
0°C in 24 h is said to be 1 ton. One ton of refrigeration is equivalent to 211 kJ/min or 200
Btu/min. The cooling load of a typical 200-m2 residence is in the 3-ton (10-kW) range.

EVAPORATIVE CONDENSER
Chapter 2: Evaporative Condenser
2.1 Introduction
Condenser is basically an example of Heat Exchangers. Heat exchangers are devices that
facilitate the exchange of heat between two fluids that are at different temperatures while
keeping them from mixing with each other. Heat exchangers are commonly used in practice
in a wide range of applications, from heating and air-conditioning systems in a household to
chemical processing and power production in large plants.
Heat exchangers differ from mixing chambers in that they do not allow the two fluids
involved to mix. In a car radiator, for example, heat is transferred from the hot water flowing
through the radiator tubes to the air flowing through the closely spaced thin plates outside
attached to the tubes.
Heat transfer in a condenser usually involves convection in each fluid and conduction
through the wall separating the two fluids. In the analysis of heat exchangers, it is convenient
to work with an overall heat transfer coefficient U that accounts for the contribution of all
these effects on heat transfer. The rate of heat transfer between the two fluids at a location in
a heat exchanger depends on the magnitude of the temperature difference at that location,
which varies along the heat exchanger. In the analysis of heat exchangers, it is usually
convenient to work with the logarithmic mean temperature difference LMTD, which is an
equivalent mean temperature difference between the two fluids for the entire heat exchanger.

2.2 Evaporative Condenser
Principle of Operation :
The vapor to be condensed is circulated through condensing coils, which is continually
wetted on the outside by water sprayed over tubes. Air is pulled over the coil, causing a small
portion of the water to evaporate. The evaporation removes latent heat from the vapor in the
coil, causing it to condense.
Types of Evaporative condensers :
1) According to directions of Air and Water Flow :
 Combined Flow : Combined flow is the use of both a condensing coil and fill
surface
for heat transfer in an evaporative condenser. The addition
of fill surface to the traditional evaporative condenser design
reduces evaporation in the coil section, reducing the potential
for scaling and fouling. Combined flow evaporative
condensers utilize parallel flow of air and spray water over
the coil, and cross flow air/water flow through the fill
surface.
Fig. 2 Combined Cross Flow
 Parallel Flow : In parallel flow, air and water flow over the coil in the
Same direction. In the fill section of BAC’s combined flow
evaporative condensers, air and water interact in a crossflow

configuration: water flows vertically down the fill as air
flows horizontally across it.
Fig. 3 Parallel Flow
 Counter Flow : In counter flow evaporative condenser design, the flow of the
Air in the opposite direction of the spray water. In
counterflow evaporative condensers, air travels vertically up
through the unit while the spray water travels vertically down
over the coil.
Fig. 4 Counter Flow
2) According to the type of Fan used :
The flow of air through most factory assembled evaporative condensers is provided
by one or more mechanically driven fans. The fans may be axial or centrifugal, each
type having its own distinct advantages.
 Axial Fan : Axial fan units require approximately half the fan motor
horsepower comparably sized centrifugal fan units, offering
significant life-cycle cost savings.

Fig. 5 Axial Fan
 Centrifugal Fan : Centrifugal fan units are capable of overcoming
reasonable amounts of external static pressure (≤ 0.5”),
making them suitable for both indoor and outdoor
installations. Centrifugal fans are also inherently quieter
than axial fans, although the difference is minimal and can
often be overcome through the application of optional low
sound fans and/or sound attenuation on axial fan units.
Fig. 6 Centrifugal Fan
3) According To Draft :
 Forced Draft : Rotating air handling components are located on the air
inlet face at the base of forced draft units, facilitating easy
access for routine maintenance and service. Additionally,
location of these components in the dry entering air stream

extends component life by isolating them from the corrosive
saturated discharge air.
 Induced Draft : The rotating air handling components of induced draft
equipment are mounted in the top deck of the unit,
minimizing the impact of fan noise on near-by neighbors and
providing maximum protection from fan icing with units
operating in sub-freezing conditions. The use of corrosion
resistant materials ensures long life and minimizes
maintenance requirements for the air handling components.
An Evaporative condenser is a device which facilitates the condensation, or return to a liquid
state, of a hot gas or vapour in a tube system using the evaporation of water flowing over the
tubes. In simple terms, an evaporative condenser uses a system of tubes exposed to a constant
flow of water to cool and condense a hot gas. The cooling and subsequent condensation of
the gas is caused by a process of heat transfer that takes place when the water flowing over
the gas-filled tubes evaporates. This process is used extensively in the air conditioning
industry as a means of condensing refrigerant gas.
All refrigeration and air conditioning processes use a gas to facilitate the transfer of heat
between the air-conditioned area and the outside atmosphere. This process relies on the use of
a compressor to increase the pressure in the evaporator section, or air conditioned area, of the
system which in turn allows the absorption of heat from that area. This absorbed heat and any
heat generated by the compression process needs to be rapidly removed from the refrigerant
in the outside, or condenser, part of the system so that the cycle of heat transfer can be
repeated. This is typically done by allowing the hot refrigerant to circulate through a series of
tubes exposed to a fan induced airflow.
Although effective, this method of cooling or condensation of the refrigerant is not
particularly efficient, particularly in large commercial systems. This lack of efficiency
becomes pronounced when the ambient, or outside, temperature rises above 100 degrees
Fahrenheit (37 degrees Celsius). In these conditions, a conventional air-cooled system may

loose up to 25% of its operational efficiency. The evaporative condenser is a far more
efficient condensation mechanism in larger systems loosing only a fraction of its effective
capacity in all ambient conditions.
The evaporative condenser system typically consists of a series of pipes or tubes that carry
the hot refrigerant gas. These tubes are simultaneously exposed to a spray of water and fan
facilitated airflow. A portion of the water flowing over the gas filled tubes evaporates due to
a combination of being heated by the tubes and the flow of air. This evaporation is the
mechanism that allows the rapid cooling of the refrigerant gas which is then pumped back
into the building to resume the air conditioning process. The remaining water is then
collected and re-circulated over the condenser coils.
Although a common feature in large commercial air-conditioning systems, the evaporative
condenser is seeing an increase in use in smaller domestic air conditioners. These systems use
approximately 5-8 gallons (19-30 litters) of water which is circulated over the condenser for
eight hours before being purged from the system completely and replaced with fresh water.
The excellent efficiency characteristics of an evaporative condenser system translates into
considerable financial savings and this type of system is sure to see increased use in
residential air conditioning applications.
2.3 Typical Applications
1) Reduces Fouling Tendency
Advanced coil technology, applied on CXV Evaporative Condensers, is used to
reduce the tendency to accumulate fouling and scale on the coil’s exterior surface.
Four facets of the unique product design contribute to the reduced tendency for
fouling:
2) The Air and Water Flow in a Parallel Path
Better water coverage over the coil is maintained because the air and spray water flow
in a smooth, parallel, downward path over the coil. With this parallel flow, the spray

water is not stripped from the underside of the tubes by the upward air flow, as on
other, conventional designs. This eliminates scale-producing dry spots on the coil.
3) Increased Water Flow Over the Coil
The spray water flow rate over the coil plan area is more than twice that of
conventional units. This heavy coverage provides continuous flooding of the primary
heat transfer surface for decreased fouling potential. Improved spray water coverage
is provided at no increase in pumping horsepower due to the unique heat transfer
system of the design.
4) Evaporative Cooling Occurs Primarily in the Fill
CXV models incorporate combined flow technology, using both primary and
secondary heat transfer surfaces. The primary heat transfer surface is the serpentine
coil, which is the most important and expensive component in the unit. In BAC’s
combined flow design, more than 80% of the latent heat transfer occurs in the
secondary surface, PVC cooling tower fill , effectively moving the evaporation
process away from coil. The coil is protected from detrimental fouling and scale since
it relies primarily on sensible conduction/convection heat transfer and, therefore, is
less susceptible to scale formation than are other designs that rely primarily on latent
(evaporative) heat transfer.
5) Colder Spray Water
Spray water at a colder temperature has a lower propensity to form scale because
scale-forming compounds remain in solution, rather than deposit as solids on the coil
exterior surface. Spray water flowing over the coil is commonly 6°F to 8°F colder
than on other designs due to the addition of the secondary heat transfer surface.
Colder spray water alone typically reduces the scaling potential* by 25% compared to
other designs. This is over and above the fouling reductions achieved by the first three
factors described above

EXPERIMENTAL SETUP
Chapter 3: Experimental setup
3.1 System Introduction
There are 3 types of Refrigeration system in the Plant of AMUL , ANAND.
1) -2 °C system
2) -10 °C system
3) -30 °C system
 -2 °C system
Fig.7 Layout of -2 °C system of the plant
As shown in the figure 1 economiser and accumulator are used as throttling devices. From
the receiver tank liquid ammonia first come to economiser under the high pressure of 13 bar
(absolute). Here in the economiser small amount of liquid ammonia is taken and is expanded
through throttle valve and liquid ammonia releases its heat to the low pressure ammonia

EXPERIMENTAL SETUP
present at the shell side. Here in the economiser temperature of the liquid ammonia reduce to
10 C. further this low temperature liquid ammonia fed to the accumulator. In the accumulator
pressure of ammonia reduced and simultaneously temperature also reduced to -2 C. after the
accumulator all the ammonia is still in the liquid phase and feed to the different evaporators
in the plant using centrifugal pumps. At the evaporator liquid ammonia gain the heat from
the water and chilled it and converted into vapour form. This low temperature low pressure,
vapour ammonia is fed to the screw compressors. At the exit of the compressor high pressure
high temperature vapour ammonia is achieved. This high pressure high temperature vapour
ammonia is fed to the vertical shell and tube type water cooled condensers. After the
condensation saturated liquid ammonia is collected in the receiver tank, and whole the cycle
is continuously repeated.
 -10 °C system
As described above for the -2 C whole the cycle is repeated for the -10 C also. but for this
system different accumulator and economisers are used. Except economiser and accumulator
all the instruments are common for both the -2 C and -10 C system. -2 C system is used
mainly for the space cooling and to maintain the temperature of the milk and butter storage

EXPERIMENTAL SETUP
section while -10 °C is mainly used in ice silos to chill the water which is used in
pasteurization process.
 -30 °C system
In this system some liquid ammonia is taken from the -10 C system’s accumulator and
fed to the -30 C system’s accumulator in which the temperature of the liquid ammonia is
reduced to the -30 C. this much low temperature is used in milk pouch packing unit. Than
whole the cycle is repeated as describe above. Arrangement of the evaporater units are
described in the following diagrams.

EXPERIMENTAL SETUP
Corrugated plate type evaporators are used as shown in the diagram to chill the water.
This chilled water is further used in different applications. Shell and tube type evaporators
are also used as ice silos. In ice silos water is at the shell side and ammonia is passed
through the pipes.
Fig.10 Chilled water cycle

EXPERIMENTAL SETUP
3.2 Observed data
To Design the Evaporative condenser some basic data of refrigerant properties are required
which are as follows:
 Properties Of Ammonia :
1) For 100 ° c
 Density = 7.9 kg/m3
 Enthalpy = 1650 KJ/kg
 CP = 2570 J/kg K
 Viscosity = 12.8 µ Pa. s
 K = 0.0348 W/m K
 Pr = 0.948
 Entropy = 5.470 KJ/kg K
2) For 34 ° c
 Density = 589 kg/m3
 S = 1.26 KJ/kg K
 CP = 4.87 KJ/kg K
 Viscosity = 121 µ Pa. s
 K = 0.46 W/m K
3) For 32 ° c
 Density = 592 kg/m3
 CP = 4.85 KJ/kg K
 Viscosity = 123 µPa.s
 K = 0.466 W/mK
 Pr = 1.28
 Entropy = 1.23 KJ/kg K

EXPERIMENTAL SETUP
 Total Refrigeration Load = 463 × 2 = 926 TR
 1620 KW = 463 TR
 Wet bulb temperature of air = 28 ° c
 Condensing inlet temperature = 100 ° c
 Condensing outlet temperature = 34 ° c
 Inlet pressure of Ammonia = 13.5 bar
 Outlet pressure of Ammonia = 13.5 bar (Assume no pressure drop )
 Mass Flowrate of Ammonia = 1.24 kg/s
 Evaporative Temperature = -2 ° c
 Inlet Air Properties:
 Air dry bulb temperature=35 ° c
 Air wet bulb temperature=28 ° c
 Relative Humidity = 60
 Specific humidity = 0.0211
 Specific Volume = 0.9016 m3
/kg
 Enthalpy = 89.33 KJ/kg
( All Data are for Anand sea level elevation 39 m )
 Inlet water temperature = 30 ° c
 Outlet water temperature = 33.5 ° c

EXPERIMENTAL SETUP
 Water inlet enthalpy ( hiw ) = 125.7 KJ/kg
 Water outlet enthalpy ( how ) = 140.0 KJ/kg

Chapter 4: Calculation of design parameters for evaporative
condenser
4.1 Design of tubes in condenser
 Total condensing load if one compressor is running
= 463 TR
= 463 × 3.51
= 1625.13 KW
 Change in enthalpy for water
= 140 - 125.7
= 14.3 KJ/kg
 Flow rate of water Q = ṁ CPW ΔT
Q = ṁ Δh
1625.13 = ṁ ( 14.3 )
. .
. ṁ = 113.64 kg/s
Tube in pressure view

 According to thin cylindrical formula wall ,
Thickness of Pipe ,
P I ( D × L ) = Ʈ × t × 2 × L
.
.
. t = ( Assume Diameter of Pipe is 26 mm. )
.
.
. t = For SS (304) Tensile stress is 205 Mpa.
.
.
. t = 0.08878 mm
A little consideration will show that the thickness of wall as obtained by the above relation is
too small. Therefore , for design of pipes a certain constant is added for SS(304) , It is 1.5 .
.
.
. t = + C
= 0.08878 + 1.5
= 1.58878 mm
So, Final dimensions of Tube,
 Outer Diameter ( Do ) = 26 mm
 Inner Diameter ( Di ) = 22.8224 mm
 Thickness (t) = 1.58878 mm
Material of pipe is SS(304), Thermal conductivity is K = 16.2 W/m2
K . But in market
standard dimensions of pipes are available as below.
 Outer Diameter ( Do ) = 26.67 mm
 Inner Diameter ( Di ) = 23.368 mm
 Thickness (t) = 1.651 mm

Now, Enthalpy values of Ammonia
 Ammonia inlet enthalpy at 100 ° c hiw (superheated state) = 1650 KJ/kg
 Ammonia outlet enthalpy at 34 ° c how (saturated) = 342 KJ/kg
 Ammonia enthalpy at 32 ° c how (saturated) = 332 KJ/kg
 Change in enthalpy for Ammonia
= 1650 - 332
= 1318 KJ/kg
 Flow rate Q = ṁ CPW ΔT
Q = ṁ Δh
1625.13 = ṁ ( 1318 )
. .
. ṁ = 1.2330 kg/s
ṁ = 1.24 kg/s

Cross section area of pipe ( Inside ) = × Di
2
= × (23.368)2
= 428.65 mm2
4.2 Average /velocity of Ammonia
 Average velocity of Ammonia vapour at inlet
1) For Ammonia Vapour 100°c & 13.5 bar
Volumetric Flow rate Q=
= = = 0.15696 m3
/s

 Q = V × A
Vi =
Vi = 366.18 m/sec
2) For Ammonia Liquid Vapour 34°c & 13.5 bar
Volumetric Flow rate Q=
=
= = 2.105 × 10 -3
m3
/s
 Q = V × A
Vi =
Vi = 4.911 m/sec
4.3 Heat Transfer co-efficient for inside the tubes
 Calculation for Reynolds number of ammonia flowing inside the tube
1) For Ammonia vapour ( 100° c - 34° c )
(Re)v =
= = 4.56 × 106
2) For Ammonia at saturation
(Re)s =

=
= 0.5586 × 106
 Average Re for ammonia flow inside the pipe, (Re)A = 2.559 × 106
 Average Prandtle No. (Pr)a = 1.114.
 For smooth tubes , the friction factor in turbulent flow can be determined from the
explicit first PETUKHOV EQUATION.
 For Smooth Tubes f = [ 0.790 × ln Re – 1.64 ] -2
= [ 0.790 × ln (2.559 × 106
) – 1.64 ] -2
= 9.96 × 10- 3
 Nu = ( 0.125 × f × Re × Pr ) 1/3
= [ 0.125 × (9.96 × 10- 3
) × (2.559 × 106
)×1.114 ]1/3
= 3302.69
 Nu =
hi =
hi = 30636.87 W/m2
K
Here, K is the average thermal conductivity of Ammonia.

4.4 Heat Transfer co-efficient for outside the tubes
Let’s assume the tube arrangement,
Fig. 11 Tubes arrangement
From this arrangement of Tubes , It is Staggered Arrangement.
 ST = 57.4 mm
 SL = 50 mm
 Inclination of the tube is 1.5° .
 Velocity of water on the outer surface of tube due to Gravity.
Experimentally V = 0.58/1.25 = 0.464 m/s
 Water will create a thin film at the outer surface of the tubes.
 Reynolds number for flow of water
(Re) =
=
= 15445.33895
= 0.15445 × 105

 Let’s assume that length of condenser tube is 6 mm , tube bundle width is 3 m and
pitch is 57.4 mm. These assumptions are made on the basis of available floor area on
the plant site.
Fig. 12 outer look of evaporative condenser
 So, one layer of tubes contain tubes =
= 52.26 = 52 tubes.
 Area of one tube = 2ΠrL
= 2 × 3.14 × 0.0267 × 6
= 1.006 m2
 Let’s assume number of tubes is 300.
 Now the film heat transfer co-efficient from outside of the tubes in a counter flow
horizontal tube Evaporative Cooler are obtained from Mizushincl equation at which is
also recommended by Dreyer.

 Ho = 2102.9 [ ]1/3
{ with 0.195 < < 5.56 }
 Where [ ] =
=
= 0.58456
Which is within given range.
 Ho = 2102.9 [ ]1/3
= 2102.9 × 058456
= 1758.3087 W/m2
K
4.5 Overall Heat Transfer co-efficient
 U × A =
=
=
= 117.94
 U =

=
= 1408.36 W/m2
K
4.6 Calculation of Tube Areas of De-superheating and Phase
Change process
1) Area for De-superheating Section
 In De-superheating section NH3 with 100° c temperature will cool down to 34°c
temperature. So, enthalpy drop in desuperheating section
 Δ h = 1650 – 14
= 180 KJ/kg
 LMTD =
( )
= 15.676
 LMTDf = F* LMTD
= 0.98*15.676
= 15.3624
Calculation of correction factor (F):

R = 0.04772
P = 0.94
From the standard graph value of correction factor is obtained which is 0.98.
Fig. 13 graph of correction factor
 Q = U × A × LMTD
180 ×103
= 1408.36 × A × 15.676
A = 8.153 m2
2) Area for Phase change Section
 Now, For Phase section AMTD is used .
 So for phase change process Q = U × A × AMTD

 AMTD =
= 2.425
 Q = U × A × AMTD
1128 × 103
= 1408.36 × A × 2.425
A = 330.28 m2
 Total Tube Area = Area of De-superheating + Area of Phase change
= 8.15 + 330.28
= 338.43 m2
So from this result our assumption of 300 tubes is nearly correct.
For desuperheating 10 tubes beneath the blower is arranged.
For condensing area 330.28/52 = 6.35 layers of tubes required so beat arrangement is of
7 layers of tubes having each layer 52 tubes.

Fig. 14 arrangement of tubes
4.7 Calculation of air flow rate and water evaporation rate
Air is entering in the condenser at some average temperature with some relative humidity that
is varying continuously with environment change so it is required to take average properties
of air in calculation.
So ,
Air inlet DBT temperature = 35 °C and WBT = 28°C
Relative humidity = 60 %
Enthalpy (h1) = 89.64 KJ/kg
Density = 0.906535 m3
/kg
Air outlet DBT temperature = 37 °C and WBT = 33.67°C
Relative humidity = 80 %
Enthalpy (h2) = 120.98 KJ/kg
Density = 0.9284 m3
/kg

Applying energy equation for heat transfer between air-water mixture and ammonia (Q˚≈
1630 KW)
m˚ a (h2 – h1) = 1630
m˚ a =
m˚ a = 52.04 kg/s (Average density of air = 0.91745 kg/m3
)
m˚ a = 52.04 * 0.91745
m˚ a = 47.744 m3
/s
 For Water evaporation rate
Water evaporation enthalpy = 2260 KJ/kg

Specific enthalpy for air =1.026 KJ/kg
ΔT = 2 °C
According to energy equation water evaporation rate can be find out
1630 = 2260 * m˚ w + 52.04 *1.026 *2
m˚ w = 0.6739 kg/s
Chapter 5: Results and Conclusion
5.1 Results
 Refrigeration and cooling system of the plant and its working is thoroughly
understood.
 Thermal analysis of the condensers is done.
 Convection heat transfer co-efficient(ho ) outer side, Calculation of heat transfer co-
efficient inside tube side & overall heat transfer co-efficient(U) has been calculated.
ho = 1758.3087 W/ m2
K
hi = 30636.87 W/ m2
K

U = 1408.36 W/ m2
K
 Total heat transfer surface area to condense ammonia from its superheated state of
100˚C and 13.5 bar pressure to 34˚C and 13.5 bar pressure liquid state
A = 338.43 m2
 Total number of tubes = 10 tubes for desuperheating area + 364 tubes for condensing
area
= 374 tubes
5.2 Suggestions
 Top of the condenser must be arranged in such a manner so that hot humid air
coming out of the condenser directly mix with the atmospheric air and don’t be
recirculated in the condenser.
 Considerable fouling of the tubes can reduce the heat transfer rate of the condenser,
chemical treatment and cleaning of the tubes must be done 4 times a year at regular
interval.
 Water to be used in condenser has pH between 7.3 to 8.1.
5.3 Conclusion
Main advantage of evaporative condenser is energy conservation , two
evaporative condenser , replace system of 5 condenser, 5 cooling tower
and 10 water circulation pumps which result in huge power saving.

FORM 1
THE PATENTS ACT 1970 (39 of 1970)
& THE PATENTS RULES, 2003
APPLICATION FOR GRANT OF
PATENT [See section 7, 54 & 135 and
rule 20(1)]
(FOR OFFICE USE ONLY).
Application No: 010593
Filing Date: 05/05/2014
Amount of Fee Paid:1000INR
CBR No:44 Signature:
1 APPLICANT(S)
NAME NATIONALITY ADDRESS
Mohit Joshi
Simit Patel
Raghav Patel
Ronit Chauhan
Indian
Indian
Indian
Indian
G H Patel College of Engineering
& Technology, Near Bakrol Gate,
Vallabh Vidyanagar,Anand
Gujarat, India.
Email:mohithjoshi19@gmail.com,
Gujarat, India.
Email:patelsimit649@gmail.com,
Gujarat, India.
Email:raghavspatel@gmail.com,
Gujarat, India.
Email:ronit7050@gmail.com,
2 INVENTOR (S)
NAME NATIONALITY ADDRESS
Mohit Joshi
Simit Patel
Indian
Indian
Gujarat, India.
Email:mohithjoshi19@gmail.com,

Raghav Patel
Ronit Chauhan
Indian
Indian
Gujarat, India.
Email:patelsimit649@gmail.com,
Gujarat, India.
Email:raghavspatel@gmail.com,
Gujarat, India.
Email:ronit7050@gmail.com,
3 TITLE OF INVENTION: THERMAL DESIGN OF EVAPORATIVE CONDENSERS.
4 Address for Correspondence of
Applicant
Mohit Joshi
G H Patel College of Engineering &
Technology, Near Bakrol Gate,
Vallabh Vidyanagar,Anand,Gujarat,
India.
Telephone No:02690-266931
Mobile No.: +91-9998331131
E-mail: mohithjoshi19@gmail.com
5 PRIORITY PARTICULARS OF THE APPLICATION (S) FILED IN CONVENTION
COUNTRY
-
Country Application
No.
Filing Date Name of
the
Applicant
Title of the Invention
India 010593 05/05/2014 Mohit
Joshi
THERMAL DESIGN OF
EVAPORATIVE
CONDENSERS
6 PARTICULARS FOR FILING PATENT COOPERATION TREATY (PCT)
NATIONAL PHASE APPLICATION
International application number International filing date as allotted by the
receiving office
382565 09/06/2014

7 PARTICULARS FOR FILING DIVISIONAL APPLICATION
Original (First) application number Date of filing of Original (first) application
010593 01/02/2014
8 PARTICULARS FOR FILING PATENT
OF ADDITION
Main application/patent Number
010593
Date of filing of main application
17/03/2014
9 DECLARATION :
(i) Declaration by the inventor(s)
We the above named inventors are the true and first inventors for this invention and
declare that the applicant herein is one of the true inventors and assigned as an applicant
by us.
(a) Date: 05/05/2014
(b) Signature (s):
(c) Name (s) : Mohit Joshi
(a) Date: 05/05/2014
(b) Signature (s):
(c) Name (s) : Simit Patel
(a) Date: 05/05/2014
(b) Signature (s):
(c) Name (s) : Raghav Patel
(a) Date: 05/05/2014
(b) Signature (s):
(c) Name (s) : Ronit Chauhan
(ii) Declaration by the applicant(s) in the convention country
I/We, the applicant(s) in the convention country declare that the applicant(s) herein is/are
my/our assignee or legal representative.
(a) Date: 05/05/2014
(b) Signature (s):
(c) Name of the signatory (s) : Mohit Joshi
(iii) Declaration by the applicant:
I, the applicant hereby declare(s) that:—
* I am in possession of the above-mentioned invention.
* The provisional specification relating to the invention is filed with this application.

* There is no lawful ground of objection to the grant of the Patent to me.
* I am the true inventor cum applicant of this Patent application.
* The application or each of the applications, particulars of which are given in Para 5
was the first application in convention country/countries in respect of my invention. -----
NA-----
* I claim the priority from the above mentioned application(s) filed in convention
country / countries and state that no application for protection in respect of the invention
had been made in a convention country before that date by me/us or by any person from
which I/We derive the title. -------NA-------
* My/our application in India is based on international application under Patent
Cooperation Treaty (PCT) as mentioned in Para - 6. -------NA-------
* The application is divided out of my/our application particulars of which are given in
Para - 7 and pray that this application may be treated as deemed to have been filed on
under section 16 of the Act. -------NA-------
* The said invention is an improvement in or modification of the invention particulars of
which are given in Para - 8. -------NA-------

FORM 2
THE PATENT ACT 1970
(39 OF 1970 )
&
THE PATENTS RULES, 2003
PROVISIONAL SPECIFICATION
( See section 10 and rule 13 )
“Thermal Design of Evaporative Condensers”
Mohit Joshi,
G H Patel college of Engineering and Technology,
Vallabh Vidyanagar,
Anand, Gujarat, India.
The following specification describes the nature of invention:
Thermal Design of Evaporative Condensers
 Field of the Invention
This invention relates to Calculation of thermal parameter analysis of design of
Evaporative Condensers.
 Object of the Invention.
The objective of this Invention is to study the refrigeration system situated at
AMUL, Anand, Gujarat. The need of changing shell & tube type condensers eith
Evaporative condensers for energy saving purpose. It contains the Calculation of
Thermal Design parameters of Evaporative condenser of refrigeration unit at
Amul dairy. The main objective is to design the thermal parameters of evaporative

condensers for the load parameters related to load and load variation according to
different seasons and different parameters.
Dated 5th
day of May 2014
Inventor and Applicant
(Mohit Joshi)

FORM 3
THE PATENTS ACT, 1970 (39 of 1970)
&
THE PATENTS RULES, 2003
STATEMENT AND UNDERTAKING UNDER SECTION 8
(See section 8, rule 12)
1. Name of the applicant(s) : I , Mohit Joshi having Indian nationality and residing at G
H Patel College Of Engineering And Technology,Anand, Gujarat, India.
2. Email: mohithjoshi19@gmail.com , Tel (M): +919998331131.
3. Name, address and nationality of the joint applicants : Patel Simit, Patel Raghav,
Chauhan Ronit, All having Indian nationality and residing at G H Patel College Of
Engineering And Technology,Anand, Gujarat, India.
4. Email: patelsimit649@gmail.com ; 9726090666 , patelraghav@gmail.com ;
9427484496 , ronit7050@gmail.com ; 8866224435.
We hereby declare
(ii) that we have not made any this application for the same/substantially the same
invention outside India.
3. Name and address of the assignee:…….NA……..
(iii) that the rights in the application (s) has/have been assigned to……NA…….that
We undertake that upto the date of grant of the patent by the
Controller, We would keep him informed in writing the details regarding
corresponding applications for patents filed outside India within three months from
the date of filing of such application.
Dated this 5th
day of May 2014
4. To be signed by the applicant or Signature
___________________________
his authorized registered patent agent. (Mr. Mohit Joshi)
5. Name of the natural person who has signed. (Inventor and Applicant)
To,
The Controller of Patents,
The Patent Office,
At Mumbai.
Note : Strike out whichever is not applicable.

References
 Research papers
 International Journal of Refrigeration 29 (2006) 645–658 Bilal A. Qureshi, Syed
M. Zubair*
Mechanical Engineering Department, KFUPM Box #1474, King Fahd University
of Petroleum and Minerals, Dhahran 31261, Saudi Arabia
Received 10 November 2004; received in revised form 17 August 2005; accepted
20 September 2005
Available online 15 December 2005
 W. Goodman, The evaporative condenser, Heat Piping Air
Cond 10 (1938) 165–328.
 REFRIGERANT FORCED -CONVECTION
CONDENSATION INSIDE HORIZONTAL
TUBES Soonhoon Bae
John S. Maulbetsch
Warren M. Rohsenow Department of Mechanical
Engineering
Engineering Projects Laboratory
Massachusetts Institute of Technology
 A Study of Evaporation Heat Transfer
Coefficient Correlations at Low Heat
and Mass Fluxes for Pure Refrigerants
and Refrigerant Mixtures
M. K. Smith, J. P. Wattelet, and T. A. Newell
ACRCTR-32
 Websites
 https://www.irc.wisc.edu/properties/
 http://www.sugartech.co.za/psychro/
 http://www.che.com/nl/YToyOntpOjA7czo0OiI4OTQ5IjtpOjE7czo4NjoicHJvY2
Vzc2luZ19hbmRfaGFuZGxpbmcvdGhlcm1hbF9hbmRfZW5lcmd5X21nbXQvaG
VhdF9leGNoYW5nZXJzX2NvbmRlbnNlcnNfYW5kX2Nvb2xlcnMiO30=/
 http://www.engineeringtoolbox.com/arithmetic-logarithmic-mean-temperature-
d_436.html
 http://www.mcsgl.com/methods-of-dehumidifying,179.html
 Books
 Heat_Transfer_2nd_Ed._by_Cengel

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