Design and Construction of a Continuous

Solar Absorption Refrigeration Unit

A thesis submitted to the University of Khartoum in fulfillment

of the requirement for the degree of Ph.D.
in Mechanical Engineering

By
Manahil Zaki A/Wahab Ali
B.Sc. (1993) University of Khartoum
M.Sc. (1999)University of Khartoum

Supervisor: Dr. M. A/Bagi Siraj

Co-Supervisor: Dr. Kamal Nasr ElDin Abdalla
Faculty of Engineering/Mechanical
Engineering Department
 Acknowledgements

I wish to express my gratitude to all the people that

encouraged and supported me during this project.

In the first place, I wish to thank my supervisor, Dr. M.

A/Bagi Siraj and my co-supervisor, Dr. Kamal Nasar
Aldin for their support. I am grateful to the Department
of Mechanical Engineering , Mechanical Workshop at U.K
and Energy Research Institute for their provision of
facilities with which the work was accomplished.
 Abstract
The main objective of this thesis is to design and construct a continuous
solar absorption refrigeration machine. A solar heated ammonia-water
refrigeration machine with 0.111kw cooling capacity was designed and
built. It consists of five heat exchangers (generator, condenser,
evaporator, absorber and solution heat exchanger),two pumps, two
throttle valves, two cooling water tanks and solar collectors with storage
tanks(two evacuated and two flat plate collectors). The refrigeration unit
was set up on roof of the north building of Faculty of Engineering,
Khartoum University. A series of experiments were carried out to
evaluate the performance of the heat exchangers and the solar collector
fields. Then the absorption machine was tested to validate the design
procedure and to evaluate the impact of the system components on the
overall performance of the refrigerating system. It was found that
reduction of heat losses, increase of streams flow rates, increase of
delivery temperature of the evacuated collector field and further
optimization of the system components are necessary for better system
performance.
 ‫اىخالصح‬
‫ماُ اىهذف اىشئٍسً ىهزا اىثحث هى ذطىٌش اىَعشفح واىخثشج اىعَيٍح ىرصٌٍَ وذْفٍز‬
‫وحذج ذثشٌذ شَسً اٍرصاصً ٍسرَش ‪ .‬ىقذ ذٌ ذصٌٍَ وذْفٍز وحذج ذثشٌذ ذعَو تاىطاقح‬
‫اىشَسٍح وذسرخذً ٍخيىط األٍىٍّا واىَاء مسائو ذثشٌذ ‪,‬وذثيغ قذسج اىىحذج ‪0.111‬‬
‫(اىَىىذ ‪،‬اىَنثف‪ ،‬اىَثخش‪،‬‬ ‫مٍيى واط ‪ .‬ذحرىي اىىحذج عيى خَس ٍثادالخ حشاسٌح‬
‫اىَثاده اىحشاسي ىيسائو واىَاص )‪ٍ ،‬ضخرٍِ‪،‬صَاًٍ خْق‪،‬خضاُي ٍاء ىيرثشٌذ‬
‫جاٍعح اىخشطىً ‪ .‬ىقذ‬ ‫وٍجَعاخ شَسٍح‪ .‬ذٌ ذصٍْع وذشمٍة اىىحذج تنيٍح اىهْذسح‪-‬‬
‫أجشٌد سيسح ٍِ اىرجاسب اىعَيٍح عيى اىَثادالخ اىحشاسٌح واىَجَعاخ اىشَسٍح‬
‫وذَد ٍقاسّح اىْرائج اىعَيٍح باىرصٌٍَ اىْظشي‪ .‬وقذ ذثٍِ أُ هْاك حاجح ىشفع دسجح‬
‫خفض فقذاُ اىطاقح‬ ‫حشاسج اىَاء تىاسطح ٍجَعاخ شَسٍح راخ مفاءج أعيً و‬
‫اىحشاسٌح وذقٌٍٍ قذسج اىجهاص ذحث ظشوف ذشغٍو ٍخريفح ورىل ىرحذٌذ عىاٍو‬
‫اىرشغٍو اىَثاىً ٍِ اجو أداء أفضو‪.‬‬
Table of Contents

1 INTRODUCTION 1
1.1 Background 1

1.2 Objective and Approach 3

1.3 Structure of This Thesis 3

2 LITERATURE REVIEW 4

2.1 The Basic Principles of Refrigeration 4

2.2 Solar Cooling Technologies 6

2.2.1 Solar Collection Technologies: 7
2.2.1.1 Solar (Photovoltaic) Cell: 7
2.2.1.2 Solar Thermal 11
2.2.2 Cooling Technologies 16
2.2.2.1 Photovoltaic Operated Refrigeration Cycle (Solar Electric) 17
2.2.2.2 Solar Mechanical Refrigeration 17
2.2.2.3 Solar Absorption Refrigeration 18
2.2.2.4 Comparison of Solar Cooling Technologies 19

2.3 The Absorption Cooling Process 20

2.3.1 The Components of the absorption cooling Process 21
2.3.1.1 The Condenser 21
2.3.1.2 The Evaporator 22
2.3.1.3 The Generator 22
2.3.1.4 The Absorber 23
2.3.1.5 The Solution Heat Exchanger 23
2.3.2 Physical Principals of the Absorption Process 23
2.3.2.1 Vapor Pressure Curve of Material Pairs 23
2.3.2.2 Ideal Performance Figures 25
2.3.2.3 Real Performance Figures 26
2.3.2.4 Mass and Energy Balances[14] 26

2.4 The Solar Energy Collection 29

2.4.1 Thermal Analysis of Flat Plate Collector[1] 29
2.4.2 Performance Prediction of Solar System[1] 31

3 SYSTEM CONFIGURATION AND DESIGN APPROACH 33

3.1 Configuration of the Refrigerator 33

3.1.1 Description of the System and Physical Layout of Components 33
3.1.2 Design and Operation Strategy 35

3.2 Design Approach 40

3.2.1 Pre-Design Condition 40
3.2.1.1 Cooling Capacity 40
3.2.1.2 The Temperature Limits of The processes 40
3.2.1.3 The System Operating Pressures and Solution Concentrations 41
3.2.2 Energy Balances 43
3.2.2.1 Enthalpies at State Points 43
3.2.2.2 Heating and Cooling Loads 45

3.3 Sizing of Heat Exchangers 46

3.3.1 Generator Design 50
3.3.2 Condenser Design 57

i
 3.3.3 Evaporator Design 67
3.3.5 Absorber Design 76

3.4 Sizing and Performance Prediction of Solar Collectors 81

3.4.1 Solar Collectors Sizing 81
3.4.2 Performance Predication of Solar Collectors 85
3.4.2.1 Long Term Performance of Evacuated Tube Collector 87

4 CONSTRUCTION AND PERFORMANCE TESTING 93

4.1 Construction and Assembly 93

4.2 Materials and Methods 95

4.2.1 Description of Experimental Set Up 95
4.2.1.1 Overall Description of Experimental Unit 95
4.2.1.2 Measuring Instruments 98
4.2.1.3 Detailed Description of the Test Components 98
4.2.2 Test Procedure and Result 104
4.2.2.1 Initial Observed Operation of the System Components 104
4.2.2.2 Test Procedure of Experimental Unit and Result 105

4.3 Analysis of Result 114

4.3.1 Shell and Tube Heat Exchanger 114
4.3.1.1 Energy Gain and Loss between Cold and Hot Streams 114
4.3.1.2 Temperature Ranges 115
4.3.1.3 Flow Rates 115
4.3.1.4 Heat Transfer Coefficient 116
4.3.2 Double Tube Heat Exchanger 116
4.3.2.1 Energy Gain and Loss between Cold and Hot Streams 117
4.3.2.2 Temperature Ranges 118
4.3.2.3 Flow Rates 118
4.3.2.4 Heat Transfer Coefficient 118
4.3.3 Flat Plate Collector Field 118
4.3.3.1 Efficiency Test of Flat Plate Collector 119
4.3.3.2 Variation of Storage Tank Temperature with Ambient Conditions 119
4.3.4 Evacuated Collector Field 120
4.3.4.1 Effect of Constant Load Withdrawal On Storage Tank Temperature Variation. 120
4.3.4.2 Variation of Storage Tank Temperature under Different Weather Conditions 120

5 CONCLUSION AND RECOMMENDATION 122

5.1 General Conclusion and Recommendation 122

5.2 Further Work 122

References 123
Appendix 124

ii
 Chapter 1
1. Introduction
1.1 Background
As the world concerns more and more on global climate changes and depleting energy
resources, solar cooling technology receives increasing interests from the public as an
environment-friendly and sustainable alternative. The coincidence of solar intensity
and cooling demand has long been inspiring people to invent a machine that cools
when the sun shines[16]. It motivated the first solar cooling machine of the history
dating back to as early as 1878 when the French mathematician Augustin Mouchot
demonstrated his solar engine with the absorption cooling machine of Edmond Carré
to produce ice at the World Exhibition in Paris[22]. But as energy prices went down
with diversifying energy sources and developing transportation technologies, the idea
of using solar energy became less attractive. It was not until the 1970s that solar
cooling received great interests again from the public, when the world suffered from
the oil crisis that had been initiated by Arab members of OPEC for political
motivations. The world realized that they could no longer depend on cheap oil prices
and began to look for alternatives. Industries tried to reduce energy consumptions by
improving energy efficiency on one hand and diversifying energy sources on the
other. There were many projects for development or demonstration of solar cooling
technologies and solar cooling continued to be an important issue in the 1980s [16]
Solar cooling and refrigeration is technically possible for a wide variety of
applications. There are two quite different major applications. The first is the cooling
of buildings of all types, ranging from individual houses to large public buildings such
as hospitals. The second is refrigeration for food preservation with a second
application in the storage of vaccines for medical purposes- a vital need and an
attractive option in many of the developing countries particularly in remote regions
with no access to electricity supply.
Of the several refrigeration system alternatives which might be considered for solar
cooling applications, the absorption system stands out as one most promising at the
present state. Absorption cooling is the first and oldest form of air conditioning and
refrigeration. The principle of absorption refrigeration was first demonstrated by
Faraday as early as 1825. An absorption refrigerator or air conditioner does not use an
electric compressor to mechanically pressurize the refrigerant. Instead, heat is
supplied to the system and the cooling effect is produced directly with no significant

1
amount of rotating machinery. The absorption device uses a heat source, such as
natural gas or a larger solar collector, to evaporate the already pressurized refrigerant
from an absorbent/refrigerant mixture.
Unlike electric or gas powered air conditioners, solar-powered refrigerators and air
conditioning are not widely acceptable. The popularity of these conventional air
conditioners and refrigerators has been attributed to low cost, compactness, and the
universal availability of electricity and gas. Now that oil, coal and natural gas have
risen to the present price levels, solar thermal driven or assisted absorption cooling
machines are gaining increasing importance due to the continually growing operating
cost of conventional units. Solar-powered machines combine the advantage of low
cost operation, widespread energy availability, and unlimited energy supply.
However, no cost–competitive system for widespread application exists. As already
stated, two main reasons that solar powered refrigeration and air conditioning are not
in widespread use today are high initial cost and large equipment size. Much of the
high cost and bulky equipment is due to the fact that most of the refrigerators and air
conditioners built to date have been custom made for case of testing rather than
compactness. With additional research, a more compact and inexpensive refrigerators
and air conditioner will be produced. A major obstacle in achieving reliable cost-
effective solar thermal absorption coolers is the problem of adapting individual
components of the system to operate at the normal working temperatures for solar
collectors, between 82 to 121ºC. Specially designed units for use with a wide range
of temperatures and in a variety of climates need to be developed.
There are currently much stronger economic and political drives to promote solar
cooling technology in the market. However, making a competitive solar cooling
machine for the market still remains a challenge to the academic and industrial
communities. The common goals of present day research activities on solar assisted
cooling are to find, for each different application of cooling, an optimum combination
of collector and cooling system that matches the special cooling demands and also the
constraints of the available solar radiation in the best way, with only marginal need
for fossil fuels.

2
1.2 Objective and Approach
In this thesis, a prototype of a continuous absorption refrigeration machine
(0.111kW) with water-cooled condenser and absorber has been developed. Its input
energy is hot water between 83 to 100ºC. The working fluid is ammonia solution with
ammonia as refrigerant and water as absorber. The driving heat for the absorption
refrigerator comes from the solar heat production system, which uses water as a heat
transfer fluid and includes the evacuated collectors and the hot water storage tanks.
The main target of the design has been to keep the driving heat as low as possible so
that to guarantee effective operation of the system using evacuated collectors with
maximum temperature below 100 ˚C. This unit has been conceived as an
experimental laboratory test device with removable components to facilitate
modifications with respect to the initial design.

1.3 Structure of This Thesis

This thesis is intended to provide detailed information on the design, construction, and
performance potential of 0.111kw solar-driven ammonia/water continuous absorption
machine.
Chapter two of this thesis, introduces the basic concepts and principles of
refrigeration, reviews and compares the solar cooling technologies that can provide
refrigeration and air conditioning. In addition to reviewing the potential technologies,
it presents the principles and operating characteristics of the most viable technology.
The third chapter deals with the design and construction of the cooling machine.
Chapter four concentrates on the testing and performance of the unit. Finally, chapter
5 gives overall conclusions and recommendations regarding the research activities
reported in this thesis. The recommendations given are focusing on what has been
missing or was insufficient in the course of the research hoping for a follow-up
research in the future.

3
 Chapter 2

2. Literature Review
The first aim of this chapter is to introduce the basic concepts and principles of
refrigeration and to give an overview of the state-of-the-art of the different
technologies that are available to deliver refrigeration from solar energy. The second
aim is to compare the potential of these different technologies in delivering
competitive sustainable solutions. Finally, the basics of absorption refrigeration and
solar collectors are illustrated.

2.1 The Basic Principles of Refrigeration

A refrigeration machine consumes energy to transfer heat from a source at a low
temperature, to a sink at a higher temperature. The heat extracted from the low
temperature source is the useful cooling. The basic principles of a refrigerating
machine working between two temperatures are illustrated by figure 2.1. Heat Q 0
collected at a lower temperature level T0 is lifted and ejected at a higher temperature
level T1. To accomplish this lift W power is needed, during the process this
t
power is degraded to heat and ejected together with collected heat at temperature level
T1 (see Eqn. 2.1.1)
Q1  Q 0  W t ----------------------------------------------------------------------(2.1.1)

T1
Q1

Wt

Q 0
T0

Fig. 2.1 Refrigeration Machine Working Between Two Temperatures[17]

4
To achieve any cooling in a refrigeration machine, a refrigerant (working substance):
a gas, vapor or a vapor/liquid combination has to be circulated through a cycle in a
refrigeration system. The refrigeration cycle involves movement of refrigerant
through a system, to ultimately provide either comfort cooling, beverage cooling, to
preserve food over a period of time, and to control humidity in a "refrigerated space".
The refrigeration cycle repeats as many times as necessary in a system until the
desired temperature is achieved.
Refrigeration is firmly rooted in two basic principles known as the first law and
second law of thermodynamics. The first law states that energy neither created nor
destroyed. If energy disappears in one form, it must reappear in another. The second
law states that no system can receive energy at low temperature and reject it at higher
temperature without receiving work from the surroundings.
A key figure to characterize the energy performance of a refrigeration machine is the
Coefficient of Performance, COP. In refrigeration systems the Coefficient of
Performance C.O.P is a term used to compare the performance of different units. The
C.O.P is defined as the ratio of the refrigerant effect, or heat removed in the
evaporator, to the energy input to a system.

Heat Energy Removed from Evaporator

COP= Energy Supplied From External Source

The Ideal Refrigerating Cycle

The ideal refrigeration process is given by the so-called Carnot refrigeration cycle.
The Carnot cycle is a cycle composed of the totally reversible processes of isentropic
compression and expansion and isothermal heat addition and rejection, so it is the
most efficient refrigeration cycle. The thermal efficiency of a Carnot cycle depends
only on the temperatures in kelvins of the two reservoirs in which heat transfer takes
place. Figure 2.2 is a T-s diagram of a Carnot cycle refrigerating system.

5
 3 2

4 1

Fig. 2.2 Ideal Reversed Carnot Cycle

This Carnot cycle is composed of four reversible processes:

1. An isothermal process 4-1 in which heat is extracted at constant temperature Te
per lb (kg) of working substance
2. An isentropic compression process 1-2
3. An isothermal process 2-3 in which heat is rejected at constant temperature Tc per
lb (kg) of working substance
4. An isentropic expansion process 3-4
The Carnot refrigeration cycle shows the highest coefficient of performance COP
between the two temperature limits. The Coefficient of Performance of Carnot
refrigeration cycle is given by:
TL
COP  --------------(2.2.2)
TH  TL

where TL is the lowest cycle temperature and TH the highest.

In an actual refrigeration cycle reversibility does not exist and therefore there will be
losses, which means that this cycle is ideal and cannot be achieved in a real machine,
but it gives a yardstick for comparison of real refrigeration machines and processes.
Any real refrigeration machine would have a COP less than COP Carnot.

2.2 Solar Cooling Technologies

From a sustainability perspective, directly using solar as a primary energy source is
attractive because of its universal availability, low environmental impact, and low or
no ongoing fuel cost. But there are many problems associated with its use. The main
problem is that it is a dilute source of energy. Even in the hottest regions on earth, the
solar radiation flux available rarely exceeds 1kW/m2, which is a low value for

6
technological utilization. Consequently, large collection areas are required in many
applications and this results in excessive costs.
A second problem associated with the use of solar energy is that its availability varies
widely with time. The variation in availability occurs daily because of the day-night
cycle and also seasonally because of the earth's orbit around the sun. In addition,
variations occur at a specific location because of local weather conditions.
Consequently, the energy collected when the sun is shining must be stored for use
during periods when it is available. The need of storage also adds significantly to the
cost of any system. Thus, the real challenge in utilizing solar energy as an energy
alternative is of an economic nature. One has to strive for the development of cheaper
methods of collection and storage so that the large initial investments required at
present in most applications are reduced.
In principle, there are many different ways to convert solar energy into cooling or air-
conditioning processes; an overview is given in Fig. 2.3 which describes the cooling
technologies capable of utilizing solar radiation as an energy source.
A main distinction can be made between thermally and electrically operated systems.
Among the thermally driven processes, thermo-mechanical processes and processes
based on heat transformation can be distinguished. The latter are all based on
reversible thermo-chemical reactions with relatively low binding energies.
There are two main concepts that can be combined with each other for cooling with
solar energy:
(A) Solar collection technology
(B) Technologies for cold production

2.2.1 Solar Collection Technologies:

Utilization of solar energy requires solar collectors. There are two general types:
1- Solar cells which can be used to produce electricity.
2- Solar thermal collector which can be utilized to generate heat.

2.2.1.1 Solar (Photovoltaic) Cell:

A solar cell or photovoltaic cell is a device that converts solar radiation energy
directly into electrical energy.
The solar cell consists of a disc or surface with two thin layer of differently doped
semiconductor material, often silicon, forming a junction in between (see Fig.2.4). Metal
stripes runs along the front of the surface and along the back is a metal plate. When solar

7
radiation hits the top of the upper layer, the disc is polarized. The upper layer becomes
negatively charged and the lower layer becomes positively charged [17]. If the metal stripes
and plate are connected in a closed circuit, an electrical current will flow through the circuit.
Thus electrical power is accessible. The voltage obtained from a single disc is rather low, in
the range of 0.5 V. To obtain higher voltage, several discs are connected in series. To increase
the current rows of serially connected solar cells can be connected in parallel. Thus solar cell
panels, also called modules, are constructed. The cells are encapsulated in a transparent
material (often plastic and low-iron glass) to protect them from the environment (but not to
heat insulate them). Several solar cell panels can be combined into a solar cell array. This is
illustrated in (Fig. 2.5). Commonly the output voltage from solar cell panels seems to be in
the range of 12-24 V. In this study solar cells will be referred to only as a source of
electric current, hence readers are referred to text books dealing with fundamentals of
photovoltaic technology.

8
 Solar Cooling
Technologies

Solar Radiation Cold Production

Collectors System

Photovoltaic Cells Thermal Collectors Electric Systems Thermal Systems

Flat Plate  Vapor compression Thermo- mechanical

 Evaporative cooler Systems

 Rankine/vapor
Rankine/vapor compression
compression
Evacuated  Stirling
Stirling Cycle/Vuilleumier
Cycle/Vuilleumier
 Stem
Stem JetJet
Cycle
Cycle

Parabolic Trough Heat Transmission

Systems

Parabolic Dish Closed Cycle Systems

Central Receivers Solid Sorbent

 Water/silica –gel
 Ammonia/salt

Liquid Sorbent

 Water /Libr
 Ammonia/Water

Open Cycle Systems

Solid Sorbent

 Desiccant wheel
 Fixed Bed Process

Fig. 2.3 Solar Cooling Technologies [18](slightly adapted) Liquid Sorbent

 Plate Absorbers
 Packed Bed Absorbers
10
 Fig. 2.5 Solar cell, module and
Fig. 2.4 Principles of a silicon array [17]
photovoltaic cell [17]
2.2.1.2 Solar Thermal
(i) Basic Principles
The basic principle of solar thermal collection is that when solar radiation strikes a surface,
part of it is absorbed, thereby increasing the temperature of the surface. The central
component in each solar collector is the absorber. Here, the absorbed solar radiation is
transformed into heat; part of this heat is transferred to the heat transfer fluid and another part
is lost to the environment.
The useful energy output ,Qu, per unit time of collector of area ,Ac ,is the difference between
the absorbed solar radiation,S, and the thermal loss ,UL,and is given by:


Qu  Ac S  U L (Tpm  Ta 
Where Tpm is plate mean temperature and Ta is ambient temperature
(ii) Classification of Solar Thermal Collectors
Based on the techniques employed in heat collection and losses reduction solar thermal
collectors may be classified according (a) their collecting characteristics, (b) operating
temperature ranges (c) the way in which they are mounted, and (d) the type of transfer fluid
they employ. Table 2.1 lists the relevant solar collector technologies.
)a) Collecting characteristics:
1) A non –concentrating collector is one in which the absorbing surface for
solar radiation is essentially flat with no means for concentrating the
incoming solar radiation. The technologies considered relevant are:
i. Flat plate collectors: consist of an absorbing surface with
passages for heat transfer fluid enclosed in an insulated casing
with a transparent cover.

11
 ii.Evacuated tubular collectors: consist of an absorbing surface
mounted in a vacuum to eliminate convection heat loss.
2) A concentrating or focusing collector is one which usually contains reflectors
or employs other optical means to concentrate the energy falling on the aperture
to a heat exchanger of surface area smaller than the aperture.
The technologies considered relevant are:
i. Dish type concentrating collectors.
ii. Linear concentrating collectors(parabolic trough)
(b) Operating Temperature Range[17]: A division can be made between high
temperature collectors with a temperature range above 150 °C,
medium temperature collectors with a temperature range of 30-150 °C,
and low temperature collectors with a temperature range below 30°C.
(c) Mounting. A collector can be mounted to remain stationary, be adjustable so as tilt
angle (measured from the horizontal and equivalent to latitude) to follow the change
in solar declination or be designed to track the sun. Tracking is done by
employing either an equatorial mounting or an azimuth mounting, for the purpose
of increasing the absorption of the daily solar radiation. The operation of the
tracking mechanism can be either manual or automatic.
(d) Types of fluid. A collector will usually use either a liquid or a gas as the transfer
fluid. The most common liquids are water or a water- ethylene glycol solution.
The most common gas is air.
Table 2.1 Solar Collectors Types [17]
Type of Collector Flat Plate Evacuated Parabolic trough Parabolic dish
Collectors Collectors collectors collectors
Concentration 1 1 15-45 100-1000
Ration
Typical Working 30-80 50-200 80-300 100-500
TemperatureºC

Diagrams

(iii) Types of Solar Thermal Collectors

a) Non –concentrating Collectors
1) Flat Plate Collectors
12
The main components of the flat plate collector are (see Fig. 2.6) :
 The absorber plate which absorbs solar radiation into heat and then transfer it
fluids passing through flow passages attached to the absorber plate. (Note: for
highly efficient flat plate collector, the absorber is coated with a selective
material to reduce radiative heat losses)

 The Cover plate which is made of glass or various plastic transparent

materials. It reduces the convective and radiative heat losses to the outside air
Insulation at the back and sides of the collector reduces heat losses.
 Enclosure- a box to hold collector components together and protect them from
the weather.
(Note: low temperature flat plate collectors do not have covers, enclosure, or insulation).

Fig. 2.6 Flat Plate Collector [17]

2) Vacuum Tube Collector

In vacuum collectors the absorbers are separated thermally from their surroundings by
a transparent cover. In addition to the reduction in radiation exchange, convective heat
transport is reduced by a very good vacuum of approximately 10-3 to 10-2 Pa. The
result is heat transfer coefficient of values of around 1W/m2K. A range of different
vacuum tube geometries is available on the market (see Fig. 2.7). The absorber is
either a standard finned tube design within an evacuated glass tube or coated directly
on the inner surface of a double glass tube, with the heat transferred to liquid
circulating in separate tubes on the inside of the double glass. This technology results
in better collector performance at higher temperatures (see Fig. 2.8)

13
 Vacuum Tube

Fig.2.7 Different Vacuum Tube Geometries [1,23]

Fig. 2.8 Performance Comparing of Solar Collectors

b) Concentrating Collectors
Solar concentrator is a device which concentrates the solar energy incident over a
large surface onto a smaller surface. The concentration is achieved by use of suitable

14
reflecting or refracting elements which results in an increased flux density on the
absorber surface as compared to that existing on the concentrator aperture. In order to
get a maximum concentration, an arrangement for tracking the sun's virtual motion is
required. An accurate focusing device is also required. Thus, a solar concentrator
consists of:
 The receiver: is that element of the system where the radiation is absorbed and
converted to some other energy form; it includes the absorber, its associated
covers, and insulation.
 The concentrator: or optical system, is the part of the collector that directs
radiation onto the receiver.
 A tracking arrangement.
1) Classification of Concentrating Collectors
Solar concentrators may be classified as (i) tracking type and(ii) non-tracking type. Tracking
may be continuous or intermittent and may be one-axis or two axes. As the sun may be
moving either the focusing part or the receiver or both; concentrators may be classified
accordingly. Further, the system may have distributed receiver or central receiver.
The concentrators may also be classified on the basis of optical components. They may(i)
reflecting or refracting type,(ii) imaging or non-imaging type, and (iii) line focusing or point
focusing type.
2) Types of Concentrating Collectors(Note: only relevant types will be
considered)
.i Parabolic Trough Concentrator(one-axis tracking)
A parabolic trough concentrator is a conventional optical imaging device used as a solar
concentrator. It consists of a parabolic reflector and a metal tube receiver at its focal plane.
The receiver of a parabolic trough is linear. Usually, a tube is placed along the focal line to
form an external surface receiver as shown in (Fig 2.9). The size of the tube, and therefore the
concentration ratio, is determined by the size of the reflected sun image and the
manufacturing tolerances of the trough. The surface of the receiver is typically plated with
selective coating that has a high absorbance for solar radiation but a low remittance for
thermal radiation loss. A glass cover tube is usually placed around the receiver tube to reduce
the convective heat loss from the receiver, thereby further reducing the heat loss coefficient.

15
 Fig. 2.9 Parabolic Dish Solar Collector [17]

.ii Parabolic dish concentrator (Two-axis tracking)

A parabolic dish reflector, shown schematically in Fig. 2.10, is a point-focus collector
that tracks the sun in two axes, concentrating solar energy onto a receiver located at
the focal point of the dish. The dish structure must track fully the sun to reflect the
beam into the thermal receiver. The receiver absorbs the radiant solar energy,
converting it into thermal energy in a circulating fluid. The thermal energy can then
either be converted into electricity using an engine-generator coupled directly to the
receiver, or it can be transported through pipes to a central power-conversion system.

Fig. 2.10 Parabolic Dish Collector

2.2.2 Cooling Technologies

Solar energy can be used as a primary energy input to different kinds of cooling
systems. For example, electricity generated by PV can be used to drive a vapour
compression system. Solar thermal collectors can be used to run a sorption cooling
system or a thermo-mechanical cooling system.

16
In this section, three approaches will be reviewed and their operating characteristics will be
highlighted.
The cooling technologies considered relative are:
1) Solar Electric
2) Solar Mechanical
3) Solar Absorption

2.2.2.1 Photovoltaic Operated Refrigeration Cycle (Solar Electric)

A solar electric cooling system consists mainly of photovoltaic panels and an
electrical cooling device. In concept, the operation of a PV-powered solar
refrigeration cycle is simple. Solar photovoltaic panels produce dc electrical power
that can be used to operate a dc motor, which is coupled to the compressor of a vapor
compression refrigeration system (see Fig. 2.11).

Batteries

Fig. 2.11 Photovoltaic Operated Refrigeration Cycle[21]

Firstly, the systems should be equipped with some means to cope with the varying
electricity production rate with time, e.g. electric battery, mixed use of solar- and grid
electricity or a variable-capacity compressor and so on.
Secondly, the price of a solar electric panel should be further decreased to compete
with other solar cooling technologies.

2.2.2.2 Solar Mechanical Refrigeration

Solar mechanical refrigeration uses a conventional vapor compression system driven
by mechanical power that is produced with a solar-driven heat power cycle. The heat
power cycle usually considered for this application is a Rankine cycle in which a fluid
is vaporized at an elevated pressure by heat exchange with a fluid heated by solar
collectors. A storage tank can be included to provide some high temperature thermal
storage.

17
The vapor flows through a turbine or piston expander to produce mechanical power,
as shown in figure 2.12. The fluid exiting the expander is condensed and pumped
back to the boiler pressure where it is again vaporized. The efficiency of the Rankine
cycle increases with increasing temperature of the vaporized fluid entering the
expander. High temperatures can be obtained from concentrating solar collectors that
track the sun’s position in one or two dimensions. Tracking systems add cost, weight
and complexity to the system.
Solar mechanical systems are competitive only at higher temperatures for which
tracking solar collectors are required. Because of its economy-of-scale, this option
would only be applicable for large refrigeration systems (e.g., 1,000 tons or 3,517
kWT)

Fig. 2.12 Solar Mechanical Refrigeration [21]

2.2.2.3 Solar Absorption Refrigeration

An absorption cooling system is considered as a “heat driven” system. It has a
unique capability of transforming thermal energy directly into cooling power. The
working fluid for the system is a solution containing a refrigerant and an absorbent
which have strong chemical affinity for each other. To examine the operation of such
a system, consider the generator in figure 2.13. Heat is transferred to the solution of
refrigerant and absorbent contained in the generator and as result, refrigerant is
vaporized from mixture, leaving a solution having a low refrigerant concentration
behind. The liberated vapor is free to flow to the condenser where heat is removed
from it to bring about its liquefaction. The generator and the condenser comprise the
high pressure section of the system. Liquid refrigerant which is accumulated in the
condenser is available for expansion from this high pressure portion of the system into
the low pressure evaporator, wherein vaporization of refrigerant takes place and the
cooling effect is achieved. Once the refrigerant has been vaporized, the evaporator
and has removed heat from cold storage or space being cooled it is discharged to the

18
absorber, which is weak in refrigerant concentration. Since this combination reaction is
exothermic, heat must be removed from the absorber to maintain its temperature at a
sufficiently low value to assure the desired high chemical affinity between the refrigerant and
the solution. The resulting absorber solution which is rich in refrigerant is collected in the
bottom of the absorber and pumped back into the generator to maintain the generator solution
level and concentration. It is the pump shown in Figure 2.13 that maintains the desired
pressure difference in the system.
The requirement to circulate refrigerant –weak solution continuously from the high
temperature generator to the low temperature absorber and refrigerant-strong solution in the
opposite direction necessitates the installation of the generator-absorber solution heat
exchanger shown. The solution heat exchanger is a simple counter –flow heat exchanger
which minimizes the heat loss associated with the fluid transfer between these two
components. Without this exchanger, the heat load on the collector and the heat rejection
load associated with the absorber would both be increased with consequent reduction in the
coefficient of performance of the system. The absorption refrigeration system includes five
heat exchangers and a pump together with the necessary piping and controls (see Fig.2.13).

Heat Out Heat in

Pressure Relief Valve

Condensor Evaporator

Pressure Relief Valve

Solution
Generator Heat Exchanger Absorber

Pump
Heat In Heat Out

Fig. 2.13 Absorption Refrigeration Cycle

2.2.2.4 Comparison of Solar Cooling Technologies

The current commercial status of different solar cooling technologies may be quickly
viewed in a comparison of the initial costs of various cooling systems. From the
review of three previously mentioned solar cooling technologies (see Table 2.2), it

19
was concluded that solar electric and thermo mechanical technologies are currently
not competitive with solar thermal absorption technology in terms of initial cost.
Absorption cooling is found to be the most cost-effective for solar cooling
applications and it is a promising technology and can play a vital role to reduce GHG
emission. It was also concluded that the direction of future R&D would better be
focused on low temperature-driven cooling system. It is because initial cost can be
lowered significantly, if cheaper solar collectors are used.
Table 2.2 Comparison of Solar Refrigeration Systems[2]

Comparison of Solar Refrigeration Systems

Cold Production System Electric System Thermo mechanical Sorption Systems
 Solar Energy Collection Photovoltaic Panels Concentrating Collectors Non-Concentrating
System Collectors

 Refrigeration Cycle Vapor Compression Vapor Compression/ Absorption

Rankine
 Minimum Driving inapplicable Above 200 ºC 80-150˚C
Temperature
 Price per W or kW €1,667/Kw(cost of solar €7.71/Watt(not €500-800/KW
panels alone) including the price of the
heat engine)
 C.O.P 0.25-0.3(due to low solar 0.3-0.4 0.6-0.7
cell efficiency)

2.3 The Absorption Cooling Process

The simplest absorption cycle, single effect cycle, have five main components,
namely generator, absorber, condenser, evaporator and a solution heat exchanger. In
order to clarify the principle, figure 2.14 will be discussed briefly. A binary mixture
(refrigerant/absorbent) circulates between generator and absorber. At point (1)
mixture with high concentration of refrigerant (strong) enters the generator. When
heat QG supplied on the high temperature level TG, refrigerant is driven out of this
mixture and it becomes weak (point 2 ). The escaping refrigerant vapour (point 5 )
flows to the condenser (point 6) where it condensates at temperature Tc (point 7),
provided that the vapour pressure at the condenser is lower than at the generator.
Condensation heat Qc must be removed. This liquid refrigerant is throttled to low
pressure (point 8 ) and fed to evaporator. If heat QE is supplied, the liquid evaporates
at temperature TE (point 9). The vapour flows to the absorber, provided that the
vapour pressure at the absorber is lower than at the evaporator. In the absorber, weak
mixture that has left the generator and has been throttled to low pressure, enters (point 3).
This weak mixture absorbs the vapour and absorption heat QA must be rejected. The

20
strong mixture is pumped back to the generator (point 1), which completes the cycle. The
work done by the pump (Wp) is small compared to the heat flows. In the considered
system, QE will be provided by the storage room.
Through the representation of the components in the Pressure-Temperature-concentration
diagram (P-T-x diagram with the concentration of the solution , x , as parameter) the
individual process steps can be reconstructed (see Fig 2.15). On the high pressure side,
with pressure Ph, are the condenser and generator, on the low pressure side, with pressure
level Pl, are the evaporator and absorber. In the evaporator and condenser the refrigerant
concentration is 100 %, which corresponds to a concentration of solution of x =1.0 . The
lowest refrigerant concentration in the solution is produced in the generator (
concentration lines on the right of the P-T-x diagram)

High Pressure Ph

Condenser Generator

TC
TG
Solution Heat
Exchanger

Solution
Throttle TA Throttle
Pump

Evaporator Absorber
TE

Low Pressure PL

Fig. 2.14 Absorption Cycle (state Points)

2.3.1 The Components of the Absorption Cooling Process

2.3.1.1 The Condenser
The condenser is the device that transfers heat from the refrigeration system to a
medium which can absorb and move it to a final disposal point. Heat release takes
place either by air or a liquid circuit. It is in the condenser that superheated, high
pressure refrigerant vapor is cooled to its boiling (condensing) point by rejecting

21
sensible heat. The additional rejection of latent heat causes the vapor to condense
into the liquid state.

Ph

Pl

TE TC TA TG

Fig. 2.15 Absorption Cycle on the P-T-X Diagram[2]

2.3.1.2 The Evaporator

The evaporator is that part of the refrigeration system in which the refrigerant boils
and, in doing so, absorbs heat. Heat uptake takes place either by air or a liquid circuit.
Before the condensed refrigerant enters the evaporator, the pressure must be reduced
to the low evaporator pressure. This is usually done by a throttle valve. The purpose
of the evaporator is to receive low-pressure, low temperature fluid from the throttle
valve and to bring it in close thermal contact with the load and leaves the evaporator
as a dry gas.

2.3.1.3 The Generator

The rich(strong) solution (rich in refrigerant) contained in the generator is heated by
conventional heat sources like gas , other fossil sources, or by solar thermal energy
using collectors. The rise in temperature of the rich solution raises the vapor pressure
such that the vapor pressure in the condenser equals the saturation pressure in the

22
generator. The refrigerant, expelled in the vapor form goes to the condenser , and the
weak solution returns through a throttle valve to the absorber.

2.3.1.4 The Absorber

The refrigerant-poor solution flows back into the absorber from the generator. The
refrigerant vapor produced in the evaporator is absorbed there as function of the
absorber temperature and solvent concentration. The evaporator and absorber are at
the same refrigerant pressure level. The refrigerant poor solution in the absorber must
constantly take up the refrigerant produced in the evaporator, since otherwise the
evaporator pressure would rise. Through refrigerant absorption the concentration of
refrigerant vapor in the solution rises. The concentration modification between the
rich and the poor solution  r and  p is called the degassing width.

2.3.1.5 The Solution Heat Exchanger

The requirement to circulate refrigerant weak solution continuously from the high
temperature generator to the low temperature absorber and refrigerant-strong solution
in the opposite direction necessitates the installation of the solution heat exchanger
shown in Fig 2.14. The installation of this counter flow heat exchanger minimizes the
heat loss associated with the fluid transfer between these two components. Without it,
the heat load on the collector and the heat rejection load associated with the absorber
would both be increased with consequent reduction in the coefficient of performance
of the system. This exchange of heat may bring the rich solution close to the boiling
point or even higher.

2.3.2 Physical Principals of the Absorption Process

2.3.2.1 Vapor Pressure Curve of Material Pairs
In thermal equilibrium, a saturation vapor pressure arises over a pure liquid,
depending solely on the temperature. As a function of the evaporation enthalpy of the
pure working material, an exponential rise of the saturation vapor pressure, P, with the
1
negative reciprocal value of the temperature  results [2] . For pure ammonia, this
T
results in an approximation solution with logarithmic function( see Eqn. 2.3.2.1.1):

b
log10 p  a  ------------------------------------(2.3.2.1.1)
T

23
Indicated by the coefficients a= 10.018 and b= 1204.3 for pressures up to 25 10 5 Pa
(T in Kelvin). According to the equation 2.3.2.1.1 (Chaperon's equation) if a curve is
plotted between logarithm of vapor pressure and reciprocal of the absolute
temperature, then the lines of constant concentration will be practically straight lines
(see Fig.2.16). According to the pressure-temperature-concentration diagram (P-T-X
diagram), a fixed relation exists between temperature, pressures and concentrations of
the absorbent-refrigerant pair in the absorption cooling process.
The low pressure level P of the cooler is determined by the saturation vapor pressure
at the desired evaporator temperature Te. The temperature level of the condenser Tc
determines the high pressure level Ph in the absorption cooler.
The maximum concentration of solution at which absorption can still occur is
determined by the absorber temperature Ta. The generator temperature Tg on the high
pressure side determines the minimum solution concentration and so determines the
degassing width, i.e. the concentration difference , between the strong solution in the
absorber and the weak solution in the generator. As shown in the diagram, the
concentration of solution varies in 10% steps from 0.1 to 1 (pure ammonia
corresponds to x =1, left curve).
As illustrated above, the performance of an absorption cooler like the one shown
schematically in Figure 2.14 is determined by the temperatures of the various
components. The condenser and absorber temperatures are determined primarily by
the available heat rejection temperature. The evaporator temperature must be
sufficient to produce the desired cooling effect in the space being cooled. With these
three temperature prescribed, the generator temperature and hence heat source
temperature necessary to effect proper system operation are fixed by thermodynamic
consideration. The maximum generator temperature results when refrigerant vapor
has already been expelled from the solution, i.e. at the end of the expulsion process.
As shown in figure 2.15 &2.16, the process runs via the evaporator(status point1),
condenser(2), the absorption of the refrigerant in the solution (3), the entry of the rich
solution into the generator (4), the weak solution at the end of the expulsion
process(5) and the weak solution cooled by the solution heat exchanger before re-
entry into the absorber(6). The refrigerant vapor in the generator is then in equilibrium
with the refrigerant-weak solution leaving the generator.

24
Fig. 2.16 Vapor Pressure Curve of Ammonia-Water Solution in the log -1/T Diagram

2.3.2.2 Ideal Performance Figures

In an ideal absorption process the cyclic process of the refrigerant is regarded as loss-
free and thermodynamically reversible. Based on the principle of conservation of
energy, the heat taken up in the evaporator and in the generator must equal the
delivered heat in the condenser and absorber (see Eqn. 2.3.2.2.1).
QE  QG  QA  Qc --------------- (2.3.2.2.1)
Since an ideal process runs reversibly, the entropy must remain constant according
to the second law of thermodynamics. The reduced entropy in the condenser
corresponds to the entropy increase in the evaporator, and the entropy decrease in the
absorber corresponds to the entropy increase in the generator. If instead of the energy
balance a power balance is set up and the power is related to the circulating
refrigerant mass flow, the result is
Q e / m v Q c / m v
 ------------------- (2.3.2.2.2)
Te Tc

Q a / m v Q g / m v
 ---------------------- (2.3.2.2.3)
Ta Tg

25
 The coefficient of performance (COP) of an absorption cooler is defined by the
relation of the power taken up in the evaporator to the supplied power in the
generator, and can by reformulation the above equations be represented as a
temperature relation.
Q E TG  Ta TE
COP    ----------------------- (2.3.2.2.4)

QG TG TC  TE
The COP of an absorption cooler is thus the product of a right-circulating Carnot
thermal engine between the temperatures of the generator and the absorber and a left-
circulating Carnot cooling machine between the temperatures of the evaporator and
the condenser.
Good COPs result if the condenser and absorber temperature can be kept low, since
on the one hand the temperature lifting capacity of the thermal machine rises between
the absorber and generator, and on the other hand the temperature difference between
the evaporator and condenser for an efficient cooler cyclic process remains small.

2.3.2.3 Real Performance Figures

In an ideal absorption cooling process, COP over 1.0 can readily occur. In a cooler,
however, irreversible processes occur during absorption of the refrigerant in the
solution. The real COPs depend furthermore on whether the freed amounts of heat in
the absorber, condenser can be recovered and supplied to the process again.

2.3.2.4 Mass and Energy Balances[14]

(i) Generator
At the generator, 3 mass flows occur and heat is supplied externally (see Fig.2.17) .
Rich(strong) *solution is supplied to a generator at the flow rate m r , concentration xr
 p,
and Temperature T7. Poor(weak)* solution leaves the generator at flow rate m

concentration xp and T8. Superheated vapor leaves for condenser with flow rate m v

Temperature T1., pressure PH. Heating medium enters the generator at m w and
Temperature T11.

QG  m
 8h8  m
 1h1  m
 7 h7 -----------------------------(2.3.2.4.1)

p m
xpm  v  xr m
r -----------------------------(2.3.2.4.2)

* rich and poor solution are strong and weak solutions respectively.

26
 m v , hv

m r , xr , h8
 r , x p , h7
m

Fig. 2.17 Mass Flows at the Generator

(ii) Absorber
 p ,concentration of x p ,
Poor solution is supplied to absorber at a rate of m

temperature of T10. Rich solution leaves the absorber at flow rate m r , concentration

x r and temperature T5. Saturated vapor from the evaporator enters at flow rate m v .
Coolant enters the absorber at temperature T13 (see Fig. 2.18).
m r , xr , h4 m p , x p , h10

m v h10

Fig. 2.18 Mass Flows at the Absorber

QA  m
 10h10  m
 4 h4  m
 5h5 ---------------------------(2.3.2.4.3)

(iii) Condenser
 v and at a temperature T2
The vapor condenses in a condenser with at the rate of m
Coolant at temperature T15 (see Fig. 2.19).

m 1  m 2 --------------------------------------(2.3.2.4.4)
Qc  m
 2 h2  m
 1h1 ----------------------------(2.3.2.4.5)

27
 m v h1

m v h2
Fig. 2.19 Mass Flows at the Condenser

(iv) Evaporator
The magnitude of the heat transfer and fluid flow rates within the system are
determined by the power level of the evaporator. Refrigerant evaporates under
pressure PL, flow rate m v . Chilled water enters the evaporator at T17 (see Fig. 2.20)

m v h3

m v h4
Fig. 2.20 Mass Flows at the Evaporator
3  m
m  4 ----------------------------------(2.3.2.4.6)

QE  m
 4 h4  m
 3h3 -------------------------(2.3.2.4.7)

(v) Solution Heat Exchanger

Rich solution at the flow rate of m r is heated from T6 toT7 and poor solution at m
 p is
cooled down from T8 to T9 (see Fig. 2.21).

m r , xr , h6  p , x p , h9
m

Fig. 2.21 Mass Flows at the Solution Heat Exchanger

28
 m r (T7  T6 )  m p (T9  T8 ) ------------------------------(2.3.2.4.8)

 5 x5  m
m  1 x1  m
 2 x2  m
 6 x6 ------------------------(2.3.2.4.9)

2.4 The Solar Energy Collection

The driving heat for the absorption refrigeration system comes from the solar heat
production sub-system, which includes the collectors and storage tanks.

2.4.1 Thermal Analysis of Flat Plate Collector[1]

The important parts of a typical liquid heating flat-plate solar collector are shown in
Fig.2.22. In steady state, the performance of a solar collector is described by an
energy balance that indicates the distribution of incident solar radiation into useful
energy gain, thermal losses, and optical losses. The solar radiation absorbed by a
collector per unit area of absorber, S, is equal to the difference between the incident
radiation It and the optical losses (average-absorptance-transmitance ratio ( ) av ) as
defined by:.
S  ( ) av I t --------------------------------------------------(2.4.1.1)
The thermal energy lost from the collector to the surrounding by conduction,
convection, and infrared radiation can be represented as the product of heat transfer
coefficient UL times the difference between the mean absorber plate temperature Tpm
and the ambient temperature.
The useful energy output of a collector area, Ac, is the difference between the
absorbed solar radiation and the thermal loss:

Qu  Ac [S  U L (Tpm  Ta )] ---------------------------------(2.4.1.2)

Where UL=Us (side- losses) + Ub (bottom-losses) + Ut( top Losses)

The temperature on the absorber sheet metal Tpm is, however, a complicated function
of the distance from the heat-removing fluid tubes and the flow length, so an average
value can only be determined very laboriously from a measured temperature
distribution. What is measurable, however, is the fluid inlet temperature into the
collector or the mean fluid temperature, which at not too low flow rates is given by
arithmetical average value between entry and exit temperatures. Above all the
representation of the available energy as a function of the fluid inlet temperature is
very useful for system simulations, since the fluid inlet temperature is given by

29
storage return temperature. To be able to determine analytically the available energy
at a given fluid inlet temperature or mean fluid temperature, the temperature
distribution on the absorber sheet metal must first be calculated as the solution of a
thermal conduction problem(See Fig. 2.22). Subsequently the local fluid temperature
is calculated by heat transfer to the fluid.

Transparent
cover Absorber
Plate
Flow
Passages

Enclosure

Insulation

Fig. 2.22 Flat plate Collector[1]

At a given mass flow, the entire rise in temperature and available energy can then be
calculated by an integration over the flow length, and represented as a function of the
fluid inlet temperature as shown by the following equation:

Qu  Ac Fr [S  U L (Ti  Ta ) -------------------------------------(2.4.1.3)

Where Fr is the collector heat removal factor and is defined as the ratio of the actual
useful energy gain to the useful energy gain if the entire collector were at the fluid
inlet temperature Ti, and can be expressed as

 C f /( AcU l )][1  exp(  AcU L F  /( m

Fr  [m  C f ))] ------------(2.4.1.4)

Where F  is the collector efficiency factor and is defined as the ratio of actual heat
collection rate to the useful heat collection rate when the collector absorbing plate Tp
is at the local fluid temperature Tf can be written as:

1
F  --------------(2.4.1.5)

 1 1 1 

U LW    
U L [ D  (W  D) F ] CB  Di h f
 


30
Where F is the fin efficiency factor and is given by equation:

tanh m(W  D)l 2

F --------------------------- (2.4.1.6)
m(W  D) / 2

Where m  U L / k

Where k is the plate thermal conducdivity,  is the plate thicknes, Di is the inside tube diameter and hfi
is the heat transfer coefficient between the fluid and the tube wall. The bond conductance C b can
estimated from knowledge of the bond thermal conductivity kb, the average bond thickness  , and the
bond width b. On a per unit length basis
kb b
, Cb 

2.4.2 Performance Prediction of Solar System[1]
The solar system can be considered to consist of solar collectors with a fully mixed
sensible heat storage units supplying a load at a fixed flow rate at a constant
temperature. Thus, the energy provided from the collectors to the storage and then to
the process load depends on the storage tank temperature. The energy balance
equation for the whole system during sunshine hours can written as:
dTs
( ms C s )  l Cl (Ts  Tl ,r )  ( ) Fr I (t ) Ac ----(2.4.2.1)
 Ac FrU (Tm  Ta )  (UA) s (Ts  Ta )  m
dt
Rate of Rate of heat Rate of heat Energy to Rate of solar
heat lost from lost from load radiation
stored collector Storage
absorber by
collector
Where ms is the mass of stored water, Cs is the specific heat capacity Ts is the storage temperature
Ac is the collector area Fr heat removal factor U collector loss coefficient Ta ambient temperature
 l rate of water flow to the load C l is the specific heat
(UA)s is the loss coefficient-area product m
capacity Tl , r return load temperature ( ) absorptivity-transmitivity product , I solar radiation
intensity.
By integrating the above equation an expression for the change in storage tank
temperature for a certain time period can be obtained.
t
Ts  Tsi  Fr Ac [( ) I (t )  U (Tm  Ta )]  (UA) s (Ts  Ta )  m
 l Cl (Ts  Tl ,r )
2(ms C s )

-----------------------(2.4.2.2)
Thus, the elevation in storage tank temperature during a given time increment can be
calculated if the following parameters were determined:
 The collector parameters
 The storage size and loss coefficient
 The energy to load
 The meteorological data
31
Note that since the storage is generally only discharged down to the return load
temperature, the minimum tank temperature should be the load return
temperature Tl , r

32
 Chapter 3
3. System Configuration and Design Approach
The purpose of this analysis is to present a step by step detailed design procedure of a
continuous solar absorption refrigeration system using aqua-ammonia as a working
fluid and operates under fixed set of conditions. This analysis includes refrigerant and
solution flow rates, heating and cooling requirements, solution concentrations,
pressures and temperatures, heat exchangers dimensions and collectors areas.

3.1 Configuration of the Refrigerator

3.1.1 Description of the System and Physical Layout of Components
The designed cooling capacity of the refrigerator is 0.111 kW(nominal capacity is
0.1kW+11% efficiency losses) evaporator temperatures of +5°C with indirect heating
through commercial vacuum tube collectors. Its input energy is hot water below
100˚C. It has five heat exchangers (generator, condenser, evaporator, absorber, and an
solution heat exchanger ). The generator is heated by hot water provided by a solar
vacuum collector. The condenser and the absorber are water cooled. The evaporator
provides the cold storage. The solution heat exchanger makes possible the exchange
of heat between the low and high temperature ammonia solution. Finally two flat plate
solar collectors are inserted between the solution heat exchanger and absorber, which
works as a pre-heater to raise the cold rich solution temperature to an intermediate
level. The basic components of the system are shown schematically in Figure
3.1(adapted from reference) and the types of components are presented in table 3.1.

 Cold Water
 Refrigerant
 Solution
 Hot Water

Fig.3.1 Absorption Refrigeration Cycle[20]

33
 Table 3.1 Types of Components of the Refrigeration System

No Name Type Schematic Diagram

Generator Shell and Tube

1

2 Condenser Coil
(in Stagnant Water)

4 Evaporator Coil
(in Stagnant Water)

5 Absorber Header- Riser

(in Stagnant
Water)

6 Solar Warm up Cycle Header - Riser

(Flat Plate Collector)

8 Solution Heat Exchanger Double Tube

9 Solar Heating Cycle Dewar Flask

(Evacuated Collector)

34
3.1.2 Design and Operation Strategy
The physical layout of the prototype and the proposed operation scheme are illustrated
in figure 3.2 and table 3.2 respectively. The Methodology employed in designing this
refrigeration unit is presented in Figure 3.3. Main decisions during the design
process of the refrigerating system comprise the following measures:
 The design of this system requires six effective solar hours to generate the
refrigerant needed by the refrigerator to work twenty four hours daily. A
liquid ammonia storage tank is installed in the system to store the refrigerant
sufficient for twenty four hour operation.
 To achieve a continuous and thereby effective refrigerant generation process
even under variable solar irradiation, a minimum hot water storage tank
volume is required. The solar collectors system are equipped with storage
tanks.
 The driving heat for the sorption refrigerator comes from the solar s collectors
system. The cycle is started after the minimum driving temperature is
reached. The solar collectors system will be in the preheating phase during the
early morning hours.
 The primary-circuit pump of the solar system, i.e., the pump which moves the
fluid from the evacuated solar collector to the generator and the secondary
pump which moves the rich ammonia solution to the generator are switched
on when minimum driving temperature is reached.
 The solution heat exchanger makes possible the exchange of heat between the
low and high temperature ammonia solution, and it raises the strong ammonia
solution temperature to the designed generation temperature.
 The solar warm up cycle, which consists of two flat plat collectors with
different storage capacities, work alternatively as a pre-heater, i.e. raising the
rich ammonia solution temperature during the early morning hours to an
intermediate level while solution heat exchanger raises it to the final
temperature level that is actually required by the generator. The solar
collector, with the minimum capacity and the higher storage temperature,
supplies the system with the warm rich ammonia solution during the first hour
of operation.

35
 The physical layout of the components allows the flow of the working fluid by
gravity, and not by pressure gradient from the condenser and the absorber.
 The arrangement of the components in the prototype is determined by the
common pressure level of the generator and condenser on the one hand, and
the evaporator and the absorber on the other hand. The generator and
condenser are located in the upper part of the prototype, while the evaporator
and absorber are located in the lower part. This arrangement allows flow of
working fluid by gravity from the upper to the lower components of the
system .

36
1 Evacuated Collector

2 Flat Plate collector

3 Hot Water Tank

4 Cold Water Tank

5 Liquid Ammonia Reservoir

6 Generator

7 Condenser

8 Solution Heat Exchanger

9 Absorber

10 Evaporator

Hot Water Circulation

Refrigerant Circulation

Strong Ammonia Solution Circulation

Weak Ammonia Solution Circulation

Fig. 3.2 Physical Layout of the Solar Absorption Refrigeration Prototype

37
 Table 3.2 Operation Scheme of the Refrigeration System

No Components Quantity of Circulating Fluids during 1- 7 8 9 10 11 12 1 2 3 4 5 6 7-12

6am am am am am am pm pm pm pm pm pm pm
(refer to Fig. 3.2) operation *

Solar Warm up Cycle 15 kg of strong ammonia Solution during first hour

of operation
- - - - - - + - - - - - - -
(Flat Plate Collector 2A)
The Solar Collector

Solar Warm up Cycle 65 kg of strong ammonia solution during next 5

hours of operation
- - - - - - - + + + + + - -
(Flat Plate collector 2B)
Evacuated Collector - - - - - - + + + + + + + -
(Component 1A) 35.7 kg /hr of hot water during 3 hours of
Evacuated Collector Operation
- - - - - - + + + + + + + -
System

(Component1B)
Hot Water Pump - - - - - - + + + + + + + -
Generator 71 kg of strong ammonia Solution during 6 hours - - - - - - + + + + + + + -
The cold production system

of operation
Condenser 9.16 kg of ammonia vapor during 6 hours of - - - - - - + + + + + + + -
operation
Evaporator 9.16 kg of ammonia liquid during 24 hours of + + + + + + + + + + + + + +
operation
Absorber 61kg of weak ammonia Solution during next 5 + + + + + + + + + + + + + +
hours of operation+9.16 kg of ammonia vapor
during 24 hours of operation
Double Tube Heat - - - - - - + + + + + + + -
Exchanger
Ammonia Solution ِ - - - - - - + + + + + + + -
Pump

*Note: refer to section 3.3 for quantities of circulating fluids

38
Stage 1: Pre Design Conditions Stage 2: Energy Balances Stage 3: Sizing of Heat
Exchanging Components

Stage 4: Sizing of Solar Collectors

Fig. 3.3 Design Methodology

39
3.2 Design Approach
Design activities undertaken during the development of the refrigerator system as
illustrated by Figure 3.4 are subdivided into four main stages as follows;

Stage 1 : Pre design Conditions Stage 2 : Energy Balances Stage 3 : Sizing of Heat
Exchanging Surfaces

 Refrigeration Cycle  Temperature  Specification of

 Working Pair Differences& Types of Heat
 Cooling Capacity Volume flows Exchangers
 Evaporator  Determination  Determination of
Temperature of heating and Thermo physical
 Cooling Water Cooling Loads Properties of
Temperature fluids.
 Heating Water  Flow Conditions
Temperature  Flow
 Cycle High &Low Correlations
 Solar Collector
Pressure
Area
 Strong & Weak
Solution Concentration
Stage 4: Solar Collector Area

Fig. 3.4 Design Stages

3.2.1 Pre-Design Condition

3.2.1.1 Cooling Capacity
The designed cooling capacity of the system is chosen to be 0.111Kw.

3.2.1.2 The Temperature Limits of The processes

According to the second law of thermodynamics, it is necessary to assume operating
temperatures for the generator, evaporator and condenser before establishing
properties at the state points of the system cycle. The operating temperatures are
chosen as follows:-
1- The temperature of the condenser and absorber may be selected if the temperature
of the available cooling water is known. According to the temperature expected in
Sudan, the cooling water will have a temperature Tcw of 25 to 35˚C.
2- The designed evaporator temperature is 5 ˚ C.
3- The generator temperature is selected on the basis of the available hot water
temperature. If flat plate collectors are to be used, the maximum hot water that can be
attained is 100˚ C.
This gives the process limits:

40
Minimum and Maximum Generator Temperature Tg are 73 and 83 ºC respectively.
Minimum and Maximum Absorber Temperature Ta are 38 and 46˚ C respectively.
Condenser Temperature Tc is 38 ˚ C
Evaporator Temperature Te is 5 ˚ C

3.2.1.3 The System Operating Pressures and Solution Concentrations

The influence of the limiting temperature on the system pressures and concentration
can be shown clearly on the equilibrium diagram for aqua ammonia (P-T-X), see Fig.
3.5 According to the diagram,
Tc determines the high pressure Pc
Tg and Pc determine the poor solution concentration Xp, so
Xp is determined by Tc and Tg
Te determines the low pressure Pe
Ta and Pe determine Xp, so
The rich solution concentration is Xr is determined by Te and Ta

Fig. 3.5 P-T-X Diagram

41
The major pre-design parameters with their corresponding values and ranges are listed
in table 3.3. The values or ranges of these parameters are determined by conventional
design practice.
Table 3.3 Pre-design Conditions
Parameter Standard Range Remark
value
1 Cooling Capacity 0.111KW required cooling power of the
cycle
2 Evaporator Temperature 5˚C required design temperature
3 Condenser Temperature 38˚C Assuming cooling water is
available for a heat sink at
35˚Cand allowing temperature
differential of 3 degrees,
4 Absorption Temperature 38-46˚C Assuming cooling water is
available for a heat sink at
35˚Cand allowing temperature
differential of 3 degrees,
5 Generator Temperature 73-83˚C Determined by the collector
temperature
6 Cycle High Pressure 15.5bar P-T-X diagram*
7 Cycle Low Pressure 5.5bar P-T-X diagram
8 Ammonia Solution 46%-54% P-T-X diagram
Concentration
9 Mass of pure Refrigerant 9kg Determined by power level of
Circulated the cycle
10 Mass of Rich Solution 71kg The amount of solution need to
Circulated be circulated to generate the
required refrigerant
11 Mass of Poor Solution 62kg The amount of solution need to
Circulated be circulated to generate the
required refrigerant.

* Note P-T-X diagram is an abbreviation of Pressure- Temperature-Concentration Diagram

42
3.2.2 Energy Balances
3.2.2.1 Enthalpies at State Points
The quantities of heat transferred to and from the solution during the different
processes can be found from the enthalpies of solution and vapor(see Table 3.4)In
this analysis of the absorption refrigerator, it is necessary to make several
assumptions, so referring to Figure 3.6
1- The fluids at points 2, 4, 5 , and 8 can be assumed to be under equilibrium
conditions.
2- The enthalpies at points 6 and 5 are assumed to be equal. Also, the
temperature rise through pump is neglected.
3- It is assumed that there are no pressure drops through the connecting lines
and heat exchangers.

High Pressure Ph

Condenser Generator

TC
TG
Solution Heat
Exchanger

Solution
Throttle Pump Throttle
TA

Evaporator Absorber
TE

Low Pressure Pl
Fig. 3.6 State Points at Absorption Cycle Schematic Diagram

Note that the numbers in the above diagram refer to the following state points:
1- ammonia vapour out of the generator
2- ammonia liquid out of the condenser
3- ammonia vapour/liquid going to the evaporator
4- ammonia vapour out of the evaporator
5- strong ammonia solution out of the absorber (rich in ammonia weak in water )
7- strong ammonia solution into the generator after passing through the solution heat
exchanger.
8- Weak ammonia solution out of the generator (weak in ammonia rich in water)
10-ammonia solution into the absorber after passing solution heat exchanger (weak in
ammonia rich in water)

43
 Table 3.4 Enthalpies at State Points

State Compone Flow Temperature Pressure Ammonia Enthalpy Equilibrium

Point nts Direction )˚C) (bar) Solution (kJ/Kg) Status
Conc.
1533.99 Superheated
1 Generator 73-83 14.5 (See eqn. Ammonia

3.2.2.1.2) Vapor

346.736 Saturated
2 Condenser 38 14.5 (from Ammonia

tables) Liquid
Throttle
3 Valve

1384.11 Saturated
4 Evaporator 5 5.5  (from Ammonia
tables) Vapor

23 Strong
5 Absorber 38-46 5.5 0.54 See eqn. Ammonia
(3.2.2.1.1) Solution

Solution 67.48 Strong

6 Heat 59 14.5  See eqn. Ammonia
Exchanger (3.2.2.1.1) Solution
221.69 Strong
7 Generator 73 14.5 0.54 See eqn Ammonia
(3.2.2.1.3). Solution
221.933 Weak
8 Generator 83 14.5 0.46 See eqn Ammonia
( 3.2.2.1.1) Solution
Solution 304 Weak
9 Heat 66 14.5  See eqn. Ammonia
Exchanger (3.2.2.1.4) Solution

ِ 142.44 Weak
10 Absorber 66 14.5 0.46 See eqn. Ammonia
(3.2.2.1.1) Solution

The enthalpy of an aqua-ammonia solution [15]at an ammonia concentration x and

temperature T (˚ F) is:

h( x,T )    33.5  334 x  414 x 2  (1.03  0.401x  0.435x 2)T  (0,00021  0.00385x  0.00334x 2 )T 2 Btu / lb
------------------------------------(3.2.2.1.1)
hsv(T ,T21 )  611  0.641T1  0.398T2 Btu / lb ---------------------------------------(3.2.2.1.2)
h7  c p (T7  T6 )  h6 ---------------------------------------------------------------(3.2.2.1.3)
h9  c p (T9  T8 )  h8 ----------------------------------------------------------(3.2.2.1.4)

44
 3.2.2.2 Heating and Cooling Loads
The heating and cooling loads are obtained from the following equations (see section
2.3.2.4):
Generator
Heating load of the generator:
QG  m
 8h8  m
 1h1  m
 7 h7 ---------------------------------(3.2.2.2.1)

Condenser
Cooling load of the condenser:

Qc  m
 2 h2  m
 1h1 ---------------------------------------------------------(3.2.2.2.2)
Evaporator
Cooling effect at the evaporator:

QE  m
 4 h4  m
 3h3 ----------------------------------------------------------(3.2.2.2.3)
Absorber
Cooling load at the absorber:
QA  m
 10h10  m
 4 h4  m
 5h5 ------------------------------------------------(3.2.2.2.4)
Solution Heat Exchanger
Heat exchanged at the solution heat exchanger:
m r (T7  T6 )  m p (T9  T8 ) ---------------------------------------------------------(3.2.2.2.5)

After the quantities of heat exchanged at the system components were obtained, the
general design conditions were specified (see Table 3.5).
Table 3.5 General Design Conditions
General Design Condition of the Refrigeration System
Design Temperature of Single Components
Generator Temperature 73-83˚C
Condenser Temperature 38˚C
Evaporator Temperature 5˚C
Absorber Temperature 38-46˚C
Low Pressure 5.5bar
High Pressure 14.5 bar
Degassing Width in the Generator 8%
NH3 Rich Solution Concentration 54%
NH3 Poor Solution Concentration 46%
Heat Supplied
Generator Qg( required heat power) 0.5574 kW
Evaporator Qe( designed cooling power) 0.111 kW
Heat Dissipated
Condenser Qc(required cooling) 0.533 kW
Absorber Qa (required cooling) 0.294 kW

45
3.3 Sizing of Heat Exchangers
Approach to Heat Exchanger Design:
In a heat exchanger, heat energy is transferred from one body or stream to another.
Temperature difference between the source of heat transfer and receiver of heat is the
driving force in heat transfer. The heat passing from one body to another travels
through a medium which in general offers resistance to the heat flow. Both these
factors, the temperature difference and the resistance to heat flow, affect the rate of
heat transfer. In the design of heat exchange equipment, heat transfer equations can be
applied to calculate this transfer of energy so as to carry it out efficiently and under
controlled conditions. The equation for the heat exchanger heat transfer :
--------------------------------------(3.3.1)
Where
Q= heat transfer rate between the fluids
U= Overall heat transfer coefficient
A = heat transfer area
∆T= log mean temperature difference
After accomplishing the two stages of design procedure (see Fig. 3.4), there are
important decisions that need to be taken in the third stage i.e. heat exchanger sizing:
1- A starting point in sizing a heat exchanger is specifying the purpose of the
exchanger, so the five exchangers included in the refrigeration cycle are
designed to fulfill the need for generator, condenser, evaporator, solution
heat exchanger and absorber.
2- What kind of construction is to be used, for example, double tube or shell and
tube. Table 3.6 shows the type of exchanger construction chosen for each
component of the system.
3- Then the other decision that need to be taken is fluid paths through the heat
exchanger. Difference is made between a parallel flow, counter flow and
cross flow. In designing this refrigeration unit, a counter flow is chosen
where ever applicable.
4- The state of the media in the heat exchanger. For example, liquid-to-liquid,
gas- to-gas or liquid-to-gas heat exchanger (or vice versa). The states of the
media in this design are either liquid-to-liquid or gas-to-liquid (see table 3.1).
5- Mechanism of heat transfer: The basic mechanism of heat transfer are
conduction, convection, boiling, condensation and radiation. Of these,
radiation is usually significant only at temperatures higher than those
ordinarily encountered in tubular process heat transfer equipment, therefore,

46
 radiation will not be considered in this work . All of others play a vital role in
equipment design. The mechanism of heat transfer for each heat exchanger
will be given separately in the following sections, the emphasis will be upon a
qualitative description of the process and the basic equation. Table 3.6 gives a
summary of the correlation used to describe the mechanism of heat transfer.

The design of a process heat exchanger usually proceeds through the following steps:
1- Process conditions (stream compositions, flow rates, temperatures, pressure),
must be specified.
2- Required physical properties over the temperatures and pressure ranges of
interest must be obtained .
3- The type of heat exchanger to be employed is chosen.
4- A preliminary estimate of the size of the exchanger is made.
5- A first design is chosen, complete in all details necessary to carry out the
design calculations.
6- The design chosen in step 5 is evaluated, or rated, as to its ability to meet the
process specification with respect to heat transfer.
7- On the basis of result of step 6, a new configuration is chosen if necessary and
step 6 is repeated. If the first design was not adequate to meet the required heat
load, it is usually necessary to increase the size of the exchanger.
8- The final design should meet process requirement (within reasonable
expectations of errors.
The design procedure of heat exchangers are presented in details in the following
sections according to the following order.
1. Generator
2. Condenser
3. Evaporator
4. Double Tube Exchanger
5. Absorber

47
 Table 3.6 Flow Streams and Correlations

Component Geometry Refrigerant Heating Cooling Correlation Schematic Diagram

Name Phase(at state Agent Agent (see symbols definition in following sections)
points referred
to in 3.6)
Internal Flow: Two Phase Forced Convective Boiling

Hot Water (External

Strong Ammonia
[5,6]:
Shell & Tube


Generator

0.5
 1 
Solution
1 k  d Re Pr 
0.33
ho

Flow )
hi  1.86 0   3.5 
L  L  hi  X tt 
External Flow: Single Phase Forced
0.33
Convection
k  d Re Pr 
ho  1.86 0 
L  L 
Internal Flow:
Superheated Ammonia Gas

Condensation of Superheated Film Condensation of

Coil in a Stagnant Pool

Vapor 0.33 Superheated Vapor[10] 1

k  d Re Pr   gl ( l   v )kl3hfg  4
hi  1.86 0 
Condenser

 h  0.555  
Water L  L  i
 l (Tsat  Ts ) Di 
2
External Flow: Natural Convection[5]
  g 2C p
n

n
 T 
ho  c  k  
 d   k 
 i   
Internal Flow: Heat Flux in Nucleate Boiling Regime[8]
Liquid Ammonia

1 3
 g (  l   v )  2  C p ,l Te 
Coil in a Stagnant

q  l h fg   C n 
Evaporator

     S , f h fg Prl 
Water

3 
External Flow: Natural Convection
n
 T 
n
 g 2C 
k  
Pool

ho  C   
 l   k 

48
Table 3.6 Flow Streams and Correlations (continued)

Component Geometry Refrigerant Heating Cooling Correlation Schematic Diagram

Name Phase Agent Agent

Internal Flow: Transient Heat Transfer by conduction in a

Header-Footer in a Stagnant

Weak Ammonia Solution

cylinder[8]
Qo C p l L(Ti NH3  TbH2 0 )
2


Absorber

Water
 L L
4
External Flow: Natural convection

  g 2C 
n
 T 
n

 C    
Pool

ho k  
 l   k 
Internal Flow:
Single Phase Forced Convection for
Laminer flow.  
0.3 3
Strong Ammonia

 
Weak ammonia

 k 
Solution Heat

 Di Re Pr 
Double Tube

hi   1.86
Exchanger

 D  
Solution

 i  L
solution

 
 
  
5
External Flow: Single Phase Forced Convection for
Laminer flow.
 k 
0.33

   De Re Pr 
ho   1.86 
 De  L 
Internal Flow [1]:
Solar Warm up Cycle

\
Strong Ammonia

Solar Radiation

 t
 Ts  Ts  [Qu  (UA)(Ts  Ta )]
Solution
Header-Footer

6 mC p

49
3.3.1 Generator Design
Description: It consists of a shell on the outside and tubes placed inside the shell, the
tubes are attached on the front and rear ends to tube sheets and by baffles which are
placed to redirect the shell fluid past the tubes to enhance heat transfer (see Fig.3.7)
Mechanism of heat transfer:
1- Inner Tubes: The flow is a two phase flow in which the liquid and vapor are
multi-components. The thermodynamic relationships are complex, the
temperature, for example, being variable over a range values at a given
pressure, but with a changing ratio of total liquid to total vapor and with
changing composition of each phase. Prediction of the amount and
composition of each phase is relatively well understood and easily done in few
cases, as for mixtures of hydrocarbons, other cases require laboratory
thermodynamic data.
For design consideration, the effective heat transfer coefficient can be
considered to be made up of convective and nucleate boiling. The convective
boiling coefficient is estimated using an equation for single phased forced
convection heat transfer modified by a factor to allow for the effects of two
phase flow (see flow correlation on table 3.6)
2- Shell: The flow in the shell is a forced convection flow. When a fluid is forced
past a solid body and heat is transferred between the fluid and the body, this is
called forced convection heat transfer (see flow correlation on table 3.6).

Calculation Procedure:
The sequence of calculations in shell and tube type heat exchanger are as follows:
a- Area of flows[8]:
(i) Through tubes,
Where, N is the number of tubes, di inner tube diameter.
d s Ct B
(ii) Through shell, A2 
PT
Where, B is the baffle length, PT is distance between tube centres, Ct is clearance
between tube bundle and shell, ds is shell diameter
b- Equivalent diameter for shell:
For triangular pitch[8],

Where, do is outer diameter

50
c- Velocity of flow[8]:
(i) For flow through tubes,
m w
(ii) For flow through shell, V2 
 w A2
d- Reynolds number[8]:
d iV1  NH 3
(i) For flow through tubes, Re 
 NH 3
DV 
(ii) For flow through shell, Re  e 2 w
w
e- Individual heat transfer coefficient:
(i)Shell heat transfer[5,6,8],
0.33
k  d Re Pr 
ho  1.86 0 
L  L 

0.5
ho  1 
 3.5 
hL  X tt 
(ii)Tube side
0.33
k  d Re Pr 
hi  1.86 0 
L  L 
f- Overall heat transfer coefficient[8]:
r
r0 ln o
1 r ri 1
 0   f
U ri hi k ho
g- Log-mean temperature difference[8]:
(Twi  TNH3i )  (Two  TNH3O )
m 
(Twi  TNH3i )
ln
(Two  TNH3O )
Where, Twi &Two are inlet and temperature of water, and TNH3i & TNH30 are inlet and outlet temperature
ammonia solution

Fig. 3.7 Shell and Tube Heat Exchanger

51
The following tables shows the details of the calculation procedures and the resulting dimension specifications of the generator.
Table 3.7 Generator Design Shell Side

Heating Load Overall Heat Specific heat Inlet water Outlet water Inlet aqueous Oulet aqueous Logarithmic Mass Heat
Quantity QG Transfer capacity of water temperature temperature Ammonia Ammonia Mean Flow of transfer
coefficient Cpw Twi Two Temperature Temperature Temperature heating Area
U TNH3i TNH3o m water A

w
m
2
Numerical 0.3328 KW 100W/m K 4.226 KJ/kgK ( at 88˚C 80˚C 72 ˚C 83˚C 6.382 0.00994 0.5215m2
Value (preliminary mean temp108˚C ) kg/s
estimation)
Remarks Design From tables( see Design Design Design Design See equation See See
Appendix) 3.3.1.2 equation equation
3.3.1.1 3.3.1.3

Heat gained by ammonia solution = Heat lost by water

QG  m
 w (Two  Twi ) -----------------------------------------------{3.3.1.1}
Twi

Two
(Twi  TNH3i )  (Two  TNH3O ) TNH3o
m  --------------------------------------------{3.3.1.2}
(Twi  TNH3i )
ln
(Two  TNH3O ) TNH3i
Counter flow

QG  UA m ------------------------------------------------{3.3.1.3}

52
Table 3.8 Generator Design Shell Side (continued)

Heat Outer Inner Tube Surface No of Tubes Shell Bundle Pitch Diametrical Shell
Quantity transfer diameter diameter of length Area of N Equivalent diameter Clearance diameter
Area of tube tube l one tube diameter db between shell ds
A (do) ID At De and tubes Ct
[5]
Numerical 0.5215m2 20 mm 14.8mm 1.6 m 0.10053 5 14.22 mm 81 mm Triangular pitch 9 mm 90 mm
Value mm2

Remarks Design Design See See equation 3.3.1.8 See Design See
3.3.1.5
equation equation equation
3.3.1.4 3.3.1.6 3.3.1.7

At  di  l ---------------------------------------------------- {3.3.1.4}
N  A / At ----------------------------------------------------------------------- {3.3.1.5}
N  a(d b / d o ) b ------------------------------------------------{3.3.1.6}
ds = db + Ct ------------------------------------------------ --{3.3.1.7}

{3.3.1.8}

Where, PT = 1.25 do , a & b are constants that take the values of 0.319 & 2.142 respectively for triangular pitch

53
Table 3.9 Generator Design Shell Side (continued)

Shell velocity of Mean bulk Density Viscosity of Thermal Prandtl Reynold Shell side
flow area water temperature of water water conductivity number Number heat
Quantity transfer
of water of water of water Of water
w
coefficient
A2 V2 Tb k Pr Re
ho

Numerical 0.0018m2 0.01262 357K 968.6

343  10 6 673x10-3 184
Value m/s kg/m3 2 W/m.K 2.068 506 W/m2K
N. s/m

Remarks See See equation See equation From From From From See equation See
equation 3.3.1.10 3.3.1.11 tables(see tables(see tables(see tables(see equation
3.3.1.9 appendix appendix appendix appendix 3.3.1.12 3.3.1.13
3.3.1.14
3.3.1.15

d s Ct B
A2  -------------------------------------------------------------{3.3.1.9}
PT
m w
V2  ----------------------------------------------------------------------------------- {3.3.1.10}
 w A2

Twi  Two
Tb  -------------------------------------------------------------------------------------------------------------------- {3.3.1.11}
2

DeV2  w
Re  -------------------------------------------------------------------------------{3.3.1.12}
w
0.33
k  d Re Pr 
hi  1.86 0  ------------------------------------------------------------------{3.3.1.13}
L  L 

hl De
Nu  ---------------------------------------------------------------{3.3.1.14}
k
0.5
ho  1 
 3.5  --------------------------------------------------------------------- {3.3.1.15}
hL  X tt 

Where, B is the baffle length

54
Table 3.10 Generator Design Tube Side Ammonia Solution
Ammonia Prandtl Bulk Temp. of Ammonia Ammonia Ammonia Solution Reynold Shell side Heat Tube side Heat Fouling Thermal Overall Heat
Solution Number Ammonia solution Solution Solution viscosity Number Transfer Transfer Factor Conduct-ivity Transfer
Quantity Velocity Of Pr Of density conductivity Of NH3 coefficient( coefficient( f of Coefficient
NH3 NH3 Tb
 NH Liquid Phase) Gaseous Phase) Steel
V1
 NH 3
kNH3 3 Re hLi hLi Kw

Numerical
Value 0.00892 1.423 78 ˚C 860 0.526 w/mk 163x10-6 697 184 831 0.00018 15 W/mK 130
m/s (220˚F) kg/m3 N. s/m2 W/m2K W/m2K W/m2K

Remarks See equation From From From From tables(see See equation See equation From See table in See equation
3.3.1.16 tables(see Design tables(see tables(see appendix 3.3.1.18 3.3.1.19 tables(see appendix 3.3.1.21
appendix appendix appendix 3.3.1.20 appendix

-------------------------------------------------{3.3.1.16}

-----------------------------------------------{3.3.1.17}
d iV1  NH 3
Re  ---------------------------------------------------------------{3.3.1.18}
 NH 3
hi d i
Nu  ------------------------------------------------------{3.3.1.19}
k NH 3
0.33
k  d Re Pr 
hi  1.86 0  --------------------------------------{3.3.1.20}
L  L 

ro
r0 ln
1 r ri 1
 0    f --------------------------------- {3.3.1.21}
U ri hi k ho

55
Table 3.11 Generator Specifications
Name: Generator Dimensional Specifications:
Type Shell and Tube Heat Exchanger Number of Tubes 5
Flow Type Counter Flow Inner Diameter 14.8mm
Operation Mode : Outer Diameter 20.8
Component vaporizing 54% Length of Tubes 1.6m
Inlet Temperature 72˚C Pitch Triangular
Outlet Temperature 83˚C Bundle Diameter 81 mm
Operating Pressure 14.5 bar Shell Diameter 90mm
Heating Load 0.3328kW Baffle Spacing 5cm
Heat Transfer Coefficient 130 W/m2K Shell Clearance 9mm
Thermodynamic Mode: Material:
Vapor Phase Superheated Vapor Inner Tube Stainless Steel
Liquid Phase Ammonia – Water Mixture Shell Stainless Steel
Heating Agent:
Agent Name Water
Rate 35kg/h
Inlet Temperature 88˚C
0utlet Temperature 8o˚C

Fig. 3.8 Generator Dimensions

56
 3.3.2 Condenser Design
Description: The condenser is a coil in stagnant water basin type.
Mechanism of heat transfer:
1. Condenser Coil: The superheated vapor enters the coil and the vapor is cooled to
saturation temperature and then condensed as it losses heat to the water in the tank. The
flow is subdivided into two regions: the superheated vapor region and the saturated
vapor region.
a- The super heated vapor region: In the superheated region the flow is a forced
convection flow in which heat is transferred from the vapor to the surrounding liquid.
Flow Correlation : Experimentally, it has been shown that forced convection heat
transfer can be described in terms of these factors grouped in dimensionless numbers :
Nusselt number (Nu) = (hcD/k)

Prandtl number (Pr) = (cpμ/k)

Grashof number (Gr) = (D3ρ2g β ΔT/μ2)
Nu= f(Re and Pr)
(see flow correlation on table 3.6)
b- The saturated vapor region : The liquid and vapor are the same pure
component. The pressure temperature relationship in this case is the vapor pressure
curve for the components. The mode or mechanism of condensation is filmwise
condensation inside horizontal tubes. The actual heat transfer mechanism that operates
in the film wise condensation is closely related to the two phase flow mechanism
described in the previous section. Conditions within the tube are complicated and
depend strongly on the velocity of the vapor flowing through the tube. If this velocity is
small, condensation occurs in the manner depicted by figure 3.9 for horizontal tube.
That is, the condensate flow is from the upper portion of the tube to the bottom, from
whence it flows in a longitudinal direction with the vapor(see flow correlation on table
3.6).

Fig.3.9 Film Condensation in Horizontal Tube (cross-section of condensate flow for low velocity vapor[10])

57
2. Water tank : Heat transfer in the water tank is by natural convection. Heat transfer by
natural convection occurs when a fluid is in contact with a surface hotter or colder than
itself. As the fluid is heated or cooled it changes its density. This difference in density
causes movement in the fluid that has been heated or cooled and causes the heat transfer
to continue
Flow Correlation: Experimentally, it has been shown that convection heat transfer can
be described in terms of these factors grouped in dimensionless numbers :
Nusselt number (Nu) = (hcD/k)
Prandtl number (Pr) = (cpμ/k)
Grashof number (Gr) = (D3ρ2g β ΔT/μ2)
and in some cases a length ratio (L/D).
If we assume that these ratios can be related by a simple power function we can then
write the most general equation for natural convection:
(Nu) = K(Pr)k(Gr)m(L/D)n
In natural convection equations, the values of the physical constants of the fluid are
taken at the mean temperature between the surface and the bulk fluid(see flow
correlation on table 3.6)..
Quantity of water in the tank[5]
In case of a storage tank with liquor of mass m and specific heat Cp, heated by vapor
condensing in a helical coil, it may be assumed that the overall transfer coefficient U
is a constant. If Ts is the temperature of the condensing vapor, T1 and T2 the initial and
final temperatures of the liquor at any time t , then the rate of transfer of heat is given
by:

58
 The following tables show the details of the calculation procedures and the resulting dimension specifications of the condenser.
Table 3.12 Condenser Design ( superheated coil /vapour side)

Bulk Temp. Bulk Surface Internal Thermal Viscosity Specific Reynold Prandtl Cooling Vapuor Temp. Surface vapouside
Number Difference Overallheat
of water Temp. of Temp. of diameter conductivity of Heat Number Load of Flow Rate Area of

Quantity transfer
vapour Vapour of Vapour Vapour Capacity Superheated
Tb Ts Ts
of tube
K Re Pr vapour Coil 
m Coil coefficient

do
 Cp
A
Qcsuperheat U

Numerical
Value 35˚ C 60.5 ˚C 47.75 ˚C 0.03175m 0.0380512 11.951 2.3012 2134.49 0.723 0.0626 0.0004240 60.5-35 0.5089 5
w/mk kJ/kg kW 74 Kg/s ˚C m2 w/m2k
N. s/m2
83  38 60.5  35 Evaluated at
Bulk Temp
Evaluated
at Surface
Evaluated at
Bulk Temp
hs  hsat 9.16kg
2 2 Temp. time (6hrs) 3600  6hrs

Remarks From From From From hsuperheat=

Design Design Tables(see Tables(see Tables(see Tables(s 1620.1 See See
Appendix) ee See
Appendix) Appendix) hsaturated= equation Equation
Appendi
1472.6 3.3.2.2 Equation 3.3.2.3
x) 3.3.2.1
KJ/Kg Preliminary
Value

A  d O l
-----------------------------------------------------{3.3.2.1}
 m  TbNH3  Tbw ------------------------------------------------------------ {3.3.2.2}

Q  UA m ---------------------------------------------------{3.3.2.3}

59
Table 3.13 Condenser Design continued ( superheated coil /vapour side)

Length of vapor side heat transfer vapor side heat transfer Coil Diameter
Quantity Tube coefficient coefficient
L hi dc
hicoil

Numerical
Value 5.1 m 4.187 w/m2k 5.73027 0.3 m

w/m2k

See See Design

Eqn3.3.2.4 Eqn 3.3.2.5

0.33
k  d Re Pr 
hi  1.86 0  ---------------------------------------------------------------------------------- {3.3.2.4}
L  L 

 d 
hi coil  hi ( straightpi pe)1  3.5  ------------------------------------------------------ {3.3.2.5}
 dc 

60
 Table 3.14Condenser Design continued (superheated coil/ water side)

Bulk Temp. Surface Superheated Thermal Temperature Specific Density of Viscosity Prandtle Cubic Tube Grashof External
of water Temperature conductivity of Difference Heat Water of Number Expansion Diameter Number transfer
Quantity Ammonia
w 
Water Capacity Ammonia coefficient
Bulk Temp. di For helical
T
Tbw Ts Kw of Water liquid Pr Gr
Tbv tube
(at 42 ˚ C) Cp
w ho

Numerical 36˚ C 48.25˚ C 60.5˚ C 0.634

Value 12.25 ˚ C 4.179KJ/ 991 631x10-6 4.16 400.4x10-6 1.8544186
w/m.k 0.03175 667
kg kg/m3 N.S/m2 29x1012
m w/m k2 w/m2k
Tbw T bv See Table
48.25-36
 c =
3.3.2.1 0.13&n=
2 0.33
From Tables(see From From From From See See
Remarks Design Appendix) Tables Tables(see Tables Tables(see Design equation equation
(see Appendix) (see Appendix) 3.3.2.6 3.3.2.7
Appendix Appendix
) )

Tbw  Ts
Note : Evaluated at 42˚ C
2
gTd i3  2
Gr  -----------------------------------------------------{3.3.2.6}

n
 T   g C p 
n 2

ho  c   k  
 
------------------------------- {3.3.2.7}
 d i   k 

* c  & n are constants c  0.13 & n =0.33 see Appendix

61
 Table 3.15 Condenser Design continued (superheated coil)

Heat Length of Internal External Condenser Internal External External Internal Thermal Tube Wall Overall Temp.
Transfer helical Diameter of Diameter Load Heat Heat Scale Scale Conducti Thickness Heat Differen
Quantity Area tube Qc -vity of
helical tube of helical Transfer Transfer factor factor Ri xw Transfer ce

(superheated Steel
di tube Coefficient Coefficient Ro Coefficient
l section) Wall
do hicoil ho U
kw
Numerical 0.5089 5.1 m 0.02655 m 0.03175m 5.73027 667 0.0002 0.0002 15 0.00026 m 5.62
Value m2 0.0626 KW w/m2k w/m2k W/mK W/ m2K 60.5-35
˚C

From From From Design

Remarks Design Design Design table(see table(see table(see Eqn.
appendix) appendix) appendix) 3.3.2.8

ro
r0 ln
1 r ri 1
 0    f ------------------------------------------------ {3.3.2.8}
U ri hi k ho

62
 Table 3.16 Condenser Design continued (saturated vapour coil /vapour side)

Bulk Bulk Surface Density of Density of Thermal Viscosity Tube Tube Internal Cooling Specific
Temp. of Temp. of Temp. of Conductivity of Internal Internal Heat Load of
Quantity water Vapour of Water
Capacity
w
water vapour Vapour Water Diameter Diameter Transfer Saturated
Tv Ts v KL
w Coefficient vapour Coil
Of Liquid
CpL
Tb Di Do hicoil
Qcsaturated

Numerical
Value 36˚ C 38 ˚C 37 ˚C 11.270 583 0.4514 0.0001244 0.03175 0.03213 12971 0.47093 4.845072
kg/m 3
kg/m 3 W/mK N. s/m2 m m w/m2k KW KJ/Kg K

36  38 Evaluated at
Temp
Evaluated at
Temp
Evaluated at
Temp
Evaluated
at Bulk
hg  h f Evaluated at
Temp
2 37˚C 37˚C 37˚C Temp time (6hrs) 38˚C

Remarks From From From From h g= From

Design Design Tables(see Tables(see Tables(see Tables(see Design Design See equation 1472.6KJ/kg Tables(see
Appendix) Appendix) Appendix) Appendix) 3.3.2.11 hf= Appendix)
362.1
KJ/Kg

1
 g l (  l   v )k l3 hfg  4
hi  0.555  -----------------------------------------------------------------------{3.3.2.9}

 l sat(T  T s ) Di 

hfg  h fg 
3
Tsat  Ts  --------------------------------------------------------------------------------------- {3.3.2.10}
8
 d 
hi (coil )  hi ( straightpi pe)1  3.5  -----------------------------------------------------------{3.3.2.11}
 dc 

63
 Table 3.17 Condenser Design continued (saturated vapour coil/ water side)

Prandtle Cubic Thermal Viscosity of Viscosity Reynold Specific Grashof External Temp. Surface Length of Overall
Conductivity Water of Number Capacity Number Heat Difference Area of Coil Condenser Heat
Quantity Number Expansion
w 
of Water Transfer
 Vapour Of Liquid Transfer A Coil
Pr Kw Coefficient
v Re Cpw Gr Coefficient
ho
L
U

Numerical
Value 4.62 361.9x10-6 0.628 695x10-6 11.905x10 1455 4.178 1.071x10-7 324.399 2 ˚C 0.84397 8.5 279
-6
W/mK N. s/m2 2 KJ/Kgk w/m2k m2 m w/m2k
N. s/m
Evaluated
at Surface
Evaluated at
Surface
Evaluated
at Surface
 (38 -36)˚ C 0.84397
C=0.53
Temp Temp Temp. 4m &
310k 310k
 v D
n=0.25

Remarks See See equation See See

D=0.03123 equation 3.3.2.13 Eqn. Equation equation
3.3.2.12 3.3.2.15 3.3.2.14 3.3.2.16

gTd i3  2
Gr  ---------------------------------------------------------------- {3.3.2.12}

n
 g 2C p 
n ------------------------------------------------------ {3.3.2.13}
 T 
ho  c  k 
 k 
 di   

A  d O l ----------------------------------------------------------------------------{3.3.2.14}

Q  UA m ----------------------------------------------------------{3.3.2.15}
r
r0 ln o
1 r ri 1
 0    f -------------------------------------------------{3.3.2.16}
U ri hi k ho

64
Table 3.18 Condenser Design continued (water tank/ superheated coil side)

Quantity Temperatu Initial Final Overall Heat Time For Specific Mass
re Temperatu Temperatu Heat Transfe Raising Heat of
of re re of Water Transfer r Area Tank Capacity Coolin
Condensing of Water T2 Coefficie A Temperatu of Water g
Vapor T1 nt re Cp Water
Tsv U T m
Numeric 83˚C 35˚C 37˚C 5.62 0.5089 6 hours 4200KJ/Kg 345.6
al Value m2 k kg

Remarks Design Design Design Table Table Design From See

3.15 3.12 Tables equatio
n 2.6.1

Table 3.19 Condenser Design continued (water tank/ saturated coil side)

Quantity Temperatu Initial Final Overall Heat Time For Specific Mass of
re Temperatu Temperatu Heat Transfe Raising Heat Cooling
of re re of Water Transfer r Area Tank Capacity Water
Condensing of Water T2 Coefficie A Temperatu of Water m
Vapor T1 nt re Cp
Tv U T
Numeric 38˚C 35˚C 37˚C 279 0.8439 6 hours 4200KJ/K 240.71k
al Value 7 gk g
m2
Remarks Design Design Design Table Table Design From See
3.17 3.17 Tables equation
3.3.2.17

Tv  T1 UA
ln  t ------------------------------- {3.3.2.17}
Tv  T2 mC p

Fig. 3.10 Condenser Coil and Water Tank

65
Table 20 Condenser Specifications
Name: Condenser Cooling Agent:
Type Coil in Stagnant Basin Agent Name Water
Operation Mode: Initial 35˚C
Temperature
Condensing Component 99% Ammonia Vapor Final Temperature 37˚C
Operating 38 ˚C Agent Quantity І 350 kg
Temperature
Operating Pressure 14.5 bar Agent Quantity И 250 kg
Cooling Load 0.53353kW Dimensions:
Heat Transfer Coefficient for Superheated 5.62 W/m2K Inner Diameter 31.23mm
Vapor
Heat Transfer Coefficient for Saturated 279W/m2K Outer Diameter 31.75mm
Vapor
Thermodynamic Mode: Tube Length for Superheated Vapor 5.1m
Vapor Phase І Superheated Ammonia Vapor Tube Length for Saturated 8.5m
Vapor
Vapor Phase И Saturated Ammonia Vapor Coil Diameter for Superheated 110 cm
Vapor
Liquid Phase Saturated Ammonia Liquid Coil Diameter for Saturated Vapor 110 cm
Coil
Material:
Tubes Stainless Steel
Tank Iron

66
3.3.3 Evaporator Design
Description: It is a coil in stagnant water basin type.
Mechanism of Heat Transfer:
1. Evaporator Tube: There are several mechanisms, or processes, through
which a liquid at the saturation temperature may be converted to a vapor by
the addition of heat. If the boiling or vaporization occurs on a hot surface in a
container, in which the liquid is confined, the process is called "pool boiling".
There are several quite different mechanisms by which pool boiling occurs
depending upon the temperature difference between the surface and the liquid,
and to a lesser extent upon the nature of the surface and liquid. The classic
curve of heat flux vs. temperature difference between surface and liquid
saturation temperature for saturated pool boiling is shown in Fig.3.11 There
are various correlations that have been proposed in the literature for this region
the correlation used in this design is for nucleate boiling((see flow correlation
on table 3.6)..

Fig.3.11 Curve of Heat Flux for Pool Boiling

2- Water Tank: Heat Transfer in the water tank is by natural convection (refer
to section 3.3.2 above)

67
 The following tables show the details of the calculation procedures and the resulting dimension specifications of the evaporator.
Table 3.21 Evaporator Design (saturated liquid side)

Quantity
Density Density Enthalpy Enthalpy Latent Heat Specific Surface Prandtl Heat Flux Prandtle Viscosity Temperature Evaporator
of of of of of Heat Tension Number Number of Difference Area
Ammonia Ammonia Ammonia Ammonia Evaporation capacity of [8] Pr Ammonia
Liquid Vapour liquid vapour of ammonia liquid A
Ammonia liquid q  T
Liquid
 l
l v hf
hg hfg
Prl
C pl
Numerical
633 3.9783 -319.27 216.72 1038 0.00018065 2.5774 0.027261462
Value 4.184 55.7x10-3N/m 1.672 628.7 1.58 ◦C m2
kg/m3 kg/m3 KJ/kgK KJ/kgK KJ/kg KJ/kgK 3
W/m2 N. s/m2
From From From From From From See From See Equation
Remarks Tables(see Tables(see Tables(see Tables(see hg-hf Tables(see From Tables(see Tables(see Equation Tables(see 3.3.3.2
Appendix) Appendix) Appendix) Appendix) Appendix) Appendix) Appendix) 3.3.3.1 Appendix)

1 3
 g (  l   v )  2  C p ,l Te 
q    l h fg    C h Pr n  ---------------------{3.3.3.1}*
    S , f fg l 
Qe
A -----------------------------------------------------------------{3.3.3.2}
q 

Tambient  Tevaporator
* C s , f  0.013 n  1.7 Te  Ts  Tevaporator Ts 
2

68
 Table 3.22 Evaporator Design continued (saturated liquid side)

Heat Flux Evaporator Heat Transfer Internal Heat Transfer

q  Area
A
Coefficient
hi

4035.2 W/m2 0.02726146 m2 2

1566 W/m k
Design
Q  Ahi t

Table 3.23 Evaporator Design continued (water side)

Water Density Cubic Conductivity Specific Prandtle Viscosity External Temperature

Quantity Reference of water of Water Heat Number of Water Heat Difference
Expansi
T
Temperat Kw Capacity Pr Transfer
on Of Water
ure
w 
of Water
w Coefficient

T
Cp w ho

Numerical
27 997 0.0002761 0.613 4.179 0.000855 Tb-Ts
Value 5.83 651
kg/m3 W/mK KJ/kgK N. s/m2 2
W/m k

Tsurface  Tbulk From From From Tables(see From From See Equation
Remarks 2 Tables(see Tables(see Appendix Tables(see Tables(see 3.3.3.3
Appendix) Appendix) Appendix Appendix
n
 T   g C  
n 2

ho  C   k
 l   k 
------------------------------------------{3.3.3.3}

C   0.53 n  0.25

Table 3.24 Evaporator Design continued

Internal External Internal External Stainless Fouling Overall Evaporator Length

Quantity Heat Heat Diameter Diameter Steel Tube factor Heat Heat of Coil
Transfer Transfer Di Do Thermal Transfer Transfer
Coefficient Coefficient Conductivity f Coefficient Area L
K U
A
hi ho

Numerical
Value 1566 0.02726146
2
651 0.0148 0.02 m 15 W/mk 0.0002 326 m2
0.433
2 2
W/m k W/m k m W/m k m
From
Remarks See Table See Table Design Design From Tables(see See Eqn. Design See
3.22 3.23 Tables(see Appendix) 3.3.3.4 Eqn.
Appendix)
3.3.3.5

ro
r0 ln
1 r 1 ri
fr
 0    o  f ---------------------- {3.3.3.4}
U ri hi k ho ri
A  DL --------------------------------------------------- {3.3.3.5}

69
Table 3.25 Evaporator Specifications
Name: Evaporator Cooling Agent:
Type Coil in Stagnant Basin Agent Name Water
Operation Mode : Initial 35˚C
Temperature
Evaporating Component Saturated Ammonia Liquid Final Temperature 37˚C
Operating 5 ˚C Agent 0.125m3
Temperature Quantity
Operating Pressure 5.5 bar Dimensions
Cooling Load 0.110kW Inner Diameter 14.8mm
Heat Transfer Coefficient 326 W/m2K Outer Diameter 20mm
Thermodynamic Mode: Tube Length for Superheated 0.433m
Vapor
Vapor Phase Saturated Ammonia Vapor Material
Liquid Phase Saturated Ammonia Liquid Tubes Stainless Steel

Fig. 3.12 Evaporator Coil and Cooling Tank

70
3.3.4 Double Pipe Heat Exchanger Design
Description:
A double-pipe heat exchanger consists of two concentric pipes or tubes. The outer tube is
called the annulus. In one of the pipes a warmer fluid flows and in the other a colder one. Due
to the temperature difference between the fluids, heat is transferred.
Mechanism of Heat Transfer
Inner Tubes & Outer Tubes :The flow in the annulus and inner pipes is a forced convection
flow (see section3.3.1 for details and see flow correlation on table 3.6).
Calculation Procedure:
The sequence of calculations in double pipe heat exchanger are as follows:
a- Area of flow[8]:
(i) Through the pipe, D1
Ai 
4
 ( D32  D22 )
(ii) Through the annulus Aa 
4
D 2  D22
b- Equivalent diameter of the annulus[8]: De  3
D2
c- Velocity of flow[8]:
M s
(i) For flow through pipe, u s 
 s Ai
M w
(ii) For flow through annulus, u w 
 w Aa
M s , M w e are mass flow rates of strong and weak ammonia
Where  s and ,  w densities of
strong and weak ammonia solutions respectively:
d- Reynolds number[8]:
Du 
(i) For through pipe, Re s  2 s s
s
D eu w  w
(ii) For flow through annulus, Re w  i
w
Where s , w are viscosities of strong and weak ammonia solution.
Individual heat transfer coefficient
0.33
 
 
(iii) For pipe,  k   D Re Pr 
hi   1.86 i  Two
 Di  
L
 TT
  so
  wi

 k 
0.33
 D Re Pr  Tsi
(iv) For annulus, ho   1.86 e  Counter flow
 De   L 
Where L is the length of the heat exchanger k is thermal conductivity.

71
 1 1 1D2
h- Overall transfer coefficient[8]:   f
U ho hi D1
Where f is the fouling factor
(Two  Tsi )  (Twi  Tso )
i- Log mean temperature difference[8]: Tm 
T T 
ln  wo si 
 Twi  Tso 
Where are inlet and outlet temperatures of strong and weak solutions
j- Area of heat transfer and length and length of pipe: QI  UATm
Where A area of heat exchanger.

The following tables show the details of the calculation procedures and the resulting dimension specifications of the double tube
heat exchanger.
Table 3.26 Solution Heat Exchanger Design

Bulk Bulk Density of Density Viscosity Viscosity Thermal Specific Specific Inlet 0utlet Inlet Outlet
Quantity Strong Weak Temp. of Temp. of strong of weak Of strong Of strong Conductivit heat heat Temperat- Temperat Temperat- Temperat
Ammoni Ammonia
a mass mass of strong of weak ammonia Ammonia ammonia ammonia y of capacity of capacity of ure of ure of ure of weak -ure of
Flow Flow Rate ammonia ammonia solution solution solution solution Aqueous strong weak Strong Strong Ammonia weak
s w s w
Rate solution solution ammonia ammonia ammonia Ammonia Ammonia Solution Ammonia
m s m w Tbs Tbw  Cp Cp Solution Solution Solution
Tsi Tso Twi Two

Numeri
cal
0.0066 0.0057 764 788 0.000188 0.000214 0.5 WK/m 4614 4552 73˚C 83˚ C 66˚ C
Value 66˚ C 75˚ C 59˚ C
kg/s kg/s kg/m 3
kg/m 3 N/m2s Ns/m2 J/kgK J/kgK

Remarks From From From From From From From Design Design Design Design
Design Design Design Design Tables(see Tables(see Tables(see Tables(see Tables(see Tables(see Tables(see
Appendix) Appendix) Appendix) Appendix) Appendix) Appendix) Appendix)

72
 Table 3.27 Double Pipe Heat Exchanger Design (continued)

Log Mean Outer Internal Internal Cross Cr2oss Equivalent Velocity Velocity Reynold Reynold Prandtl Prandtl
Temperture diameter Diameter 0f Diameter of Sectional Sectional Diameter of through through Number of Number Number of Number
Quantity Difference of Inner Outer tube Inner tube Area of Area of Annulus strong of weak of strong of of
Inner Annulus
Tm tube D2 D3 D1 inner
Tube
Annulus
Tube
De
tube tube
Ammonia Res Ammonia
Rew
Ammonia
Prs
strong
Ammonia
Ai Ao us uw Prw

Numerical 8.4 ˚ C 0.0254m 0.03256m 0.01986m 0.00031 0.00033 0.01633 m 0.0279 0.035099 2333 1507 1.74 4.047
Value m2 m2 m/s m/s
See equation D See See equation See See equation See equation See See equation See
Ai 
Remarks 3.3.4.1 Design Design Design 4 equation 3.3.4.3 equation 3.3.4.5 3.3.4.6 equation 3.3.4.8 equation
3.3.4.2 3.3.4.4 3.3.4.7 3.3.4.9

(Two  Tsi )  (Twi  Tso )  ( D32  D22 )

Tm  --------------------------- {3.3.4.1} Aa  ---------------------------------------------------------------{3.3.4.2}
 TwoTsi  4
ln  
 Twi  Tso 

D32  D22 M s
De  -----------------------------------------------{3.3.4.3} us  -----------------------------------------------------------------------------{3.3.4.4 }
D2  s Ai

M w
uw  ----------------------------------------------------{3.3.4.5} D2 u s  s -----------------------------------------------------------------------------{3.3.4.6}
 w Aa Re s 
s

C p s
D i eu w  w --------------------------------------------------------{3.3.4.7} Prs  ----------------------------------------------------------------------{3.3.4.8}
Re w 
w k

C p w
Prw  ------------------------------------------------------{3.3.4.9}
k

73
Table 3.28 Double Pipe Heat Exchanger Design continued

Overall Heat Heat Transfer External Overall Heat Overall Heat Total Tube
Heat Mass Flow Temperature Heat Transfer Coefficient of Scale factor Transfer Transfer Area Length
Quantity Exchanger Rate through Difference Transfer Coefficient of weak solution Coefficient A
Rf L
Load Inner Tube
T Coefficient Strong side U
Q 
m U Solution Side
hi ho

Numerical 87.8 85.94 47.6

Value 426W 0.0066 17˚ C 68 W/m2 K 0.0002 W/m2 K 0.7487 m2 12m
Kg/s W/m2 K W/m K 2

Preliminary See equation See equation From Tables(see see equation See equation See equation
Tso  Tsi
Remarks Value 3.3.4.10 3.3.4.11 Appendix) 3.3.4.12 3.3.4.13 3.3.4.14
Qm
 cpt Design

0.33
 
  ------------------------------------------------------- {3.3.4.10}
 k   Di Re Pr 
hi   1.86 
 Di  
L

 
 
 k 
0.33
 D Re Pr 
ho   1.86 e  ------------------------------------------- {3.3.4.11}
 De   L 
1 1 1D2
   R f ----------------------------------------------------- {3.3.4.12}
U ho hi D1
QI  UATm ------------------------------------------------------------- --- {3.3.4.13}

A  D2 L ------------------------------------------------------------------------{3.3.4.14}

Fig. 3.13 Double Tube Heat Exchanger

74
Table 3.29 Double Pipe Heat Exchanger Specifications
Name: Heat Exchanger Thermodynamic Mode:
Type Double Tube /Counter Flow Liquid Phase Ammonia- Water Mixture
Heating Specification: Dimensions:
Component Heated 54% Ammonia Mixture Inner Diameter of Inner 19.86 mm
Tube
Flow Rate 23.76kg/hr Outer Diameter of Outer Inner 25.4mm
Tube
Exit Temperature 73˚C Inner Diameter of Outer 32.56 mm
Tube
Inlet Temperature 59˚C Tube Length 12m
Heat Transfer Coefficient 47.6 W/m2K Material:
Heating Load o.426KW Tubes Stainless Steel
Heating Agent
Agent Name 46% Weak Ammonia Solution
Initial Temperature 83˚C
Final 66˚C
Temperature
Agent 20.52 kg/hr
Rate

75
3.3.5 Absorber Design
Description: The absorber is a header-riser type cooled by stagnant water in a tank. The
weak ammonia solution is distributed in the row and columns of the absorber. Ammonia
vapor is distributed through nozzles which are located in the rows of the absorber.
Heat Transfer Mechanism.
1. Flow Inside Tubes: Unsteady state heat flow exists in tubes. The solution in tubes is
stagnant and its temperature changes with time as it is cooled by the water in the tank.
The problem is treated as a transient heat conduction in a cylinder..
The form of the solution is:
ƒ(  ) = F{(  )(Bi)}
where ƒ and F indicate functions of the terms following,  is dimensionless
temperature,  is called the Fourier number (this includes the factor k/cρ the thermal
conductance divided by the volumetric heat capacity, which is called the thermal
diffusivity) and Bi is the Biot number.
The exact analytical solution is very complicated. However, an approximate solution
can be obtained by using graphical tools. The graphs allow you to find the centerline
temperature at any given time, and the temperature at any location based on the
centerline temperature((see flow correlation on table 3.6)..
2. Water Tank: Heat Transfer in the water tank is by natural convection (refer to
section 3.3.2 above)
The following tables show the details of the calculation procedures and the resulting
dimension specifications of the absorber.

Table 3.30 Absorber Design (Aqueous Ammonia Solution)

Bulk Density Prandtl Viscosity Thermal Specific Volume of Diameter Total Surface
Quanti Temp. of of weak Number Of weak Conductivit heat Absorber of Length Area
-ty of weak Ammoni Pr ammonia y of capacity of Va Absorber of A
ammoni- -a solution Aqueous weak Tubes Tube
-a solution
w ammonia

ammonia D L
solution
Tbw
w Cp

Numer
-ical 0.000315 0.5 WK/m 4879 0.0965 0.1282m 7.5 m
52˚ C 822.2 3.0206
Value 2.47 Ns/m2 J/kgK
kg/m3 m2

Remar 110% Mass D 2 L

66  38 From From From From From V
-ks
Tables(see Tables(see  4
Design Design
Tables(se Tables(see Tables(see
2 Appendix)
e Appendix) Appendix)
Appendix)
Appendi
x)

76
 Table 3.31 Absorber Design (Cooling Water Side)

Bulk Temp. Specific Density Prandtl Viscosity Thermal Specific Volume of Diameter of Grashof Nusslet External
of Water Expansion of Water Number Of Water Conductivit heat Absorber Absorber Number Nu Heat
Quantity
Tb H 2 0  H o 2
Pr
H o2
y of Water

capacity of
weak
Va Tubes
D
Gr Transfer
Coefficient
ammonia
Cp ho

Numerical
Value 35˚ C 320.6  106 K 1 995
5.2
0.000769 0.62WK/m 4879 0.0965 0.1282m 7.7  10 4 16.52 79.9152
W/m2 K
kg/m 3 Ns/m2 J/kgK

Remarks 110% Mass D 2 L See Equation. See See Equation

V
37  35 From From From From From From
Tables(see  4
3.3.5.1 Equation 3.3.5.3
Tables(see Tables(see Tables(see Tables(see Tables(see 3.3.5.2
2 Appendix) Appendix)
Appendix) Appendix) Appendix) Appendix)

gTl 3  2
Gr  -------------------------------------------- {3.3.5.1}
2
0.25
 Pr 
Nu  0.683Gr 0.25
Pr 0.25
  ------------------- {3.3.5.2}
 0.861  Pr 
hl
Nu  -------------------------------------------------------- {3.3.5.3}
k

77
 Table 3.32 Absorber Design continued

Bulk Density Viscosity Thermal Specific heat Characteristic Thermal Biot Time Fourier External Dimensionless Total Total Heat
Quantity Temperature of Water Of Water Conductivity capacity of Length of the Diffusivity Nunber Heat Heat Transfer Length Transferred
 Number

 NH  NH
of Water Tube

Of Ammonia weak Transfer Q of
Solution 3 3 ammonia lc
Bi
Coefficie Tube
Q
TiNH3 Cp nt Qo L

ho
Numerical
Value 0.5WK/m 10.243 6 Hours 0.65186 79.9152 0.9 7.5 m 6355872.18
0.124  10 6
52˚ C 822.23 0.000769 4879 0.064m
Ns/m2 J/kgK W/m2 K J
kg/m3 0.294KW

Remarks 66  38
From From From From Tables(see Design See From Graphs Design See
2 D 
k
t
Tables(see Tables(see Tables(see Appendix) section [8] Equation
cp hl
Bi  
Appendix)
Appendix) Appendix) 2 5.2 3.3.5.4
K l2 Q
6  3600

Qo C p l L(Ti NH3  TbH2 0 )

2

 ---------------------------------------- {3.3.5.4}
L L

Total heat need to be removed = mTC p =61(66-38)4879= 8.3MJ

78
Table 3.33 Absorber Design continued

Quantity Heat Heat Bulk Bulk Surface Temperatur Heat

Transfe Transfe Temperatur Temperatur Temperature e Transfer
r r Area e of Weak e of Water Ts Difference Coefficien
Q A Ammonia T t
TbNH3 U
2
Numerica 0.294KW 3.0206m 52˚ C 36˚ C 44˚ C 8˚C 12.17 W/m2
l Value
K

Remark
s
See
Section
TbNH3  Tbw Ts  Tbw See Equation
3.3.5.5
5.3
2

Q  AUT ------------------------------------------------------ {3.3.5.5}

Table 3.34 Absorber Design continued (Water Tank side )

Quantity Temperatu Initial Final Overall Heat Time For Specific Heat Mass
re Temperatu Temperatu Heat Transf Raising Capacity of of
of Weak re re of Water Transfer er Area Tank Water Coolin
Ammonia of Water T2 Coefficie A Temperatu Cp g
TbNH3 T1 nt re Water
U T m
Numeric 52˚C 35˚C 37˚C 12.17 3.0206m 6 hours 4200KJ/kgK 1511kg
2
al Value W/m2 K gk

Remarks Design Design Design See table See Design See

3.33 table equatio
3.33 n 2.6.1
See
equatio
n 3.3.5.6

TbNH3  T1 UA
ln  t ----------------------------------------------- {3.3.5.6}
TbNH3  T2 mC p

79
Table 3.35 Absorber Specifications
Name: Absorber Thermodynamic Mode
Type Riser Header Liquid Phase Ammonia- Water Mixture
Cooling Specification Dimensions
Component Cooled 46% Ammonia Mixture Inner Diameter 0.1282 mm
Flow Rate 20.33kg/hr Total Length of Absorber 7.5m
Bulk Temperature 52˚C Material
Heat Transfer Coefficient 12.17 K W/m2K Tubes Galvanized Steel
Cooling Load o.2943kW Tank Polyethylene
Cooling Agent
Agent Name Water
Bulk Temperature 35˚C
Agent Quantity 1500 kg

Gas Inlet Pipes

Fig. 3.14 Absorber Components (Header-Riser pipes, Ammonia vapor Inlet Pipes and Water Tank )

80
3.4 Sizing and Performance Prediction of Solar Collectors
3.4.1 Solar Collectors Sizing
3.4.1.1 Flat Plate Collector Design
Description: The flat plate collector field consists of two collectors of header-riser type with
storage tanks of different capacities.
Basic Flat-Plate Energy Balance
There are three parameters which need to be specified in designing a solar collector: optical
efficiency (transmittance-absorbtance factor), collector heat loss coefficient and heat removal
factor. The useful energy output,Qu, of a collector area, Ac, is the difference between the
absorbed solar radiation, S, and the thermal loss , UL and is given by the following equation:

Qu  Ac Fr [S  U L (Ti  Ta )
[1]

Where S   av (optical efficiency)  It(incident solar radiation)

UL=Us (side- losses) + Ub (bottom-losses) + Ut( top Losses)

Ti & Ta are inlet fluid and ambient temperatures respectively.

The following tables show the details of the calculation procedures and the resulting
dimension specifications of the collector.

Table 3.36 Flat Plate Collector Design (Overall Heat Loss Coefficient Evaluation)

Quantity
Internal Plate to External External Top Loss Back Edge Heat Overall
Convection Cover Convection Radiation Coefficient Surface Coefficient Heat
Coefficient Radiation Coefficient Coefficient Ut Coefficient UE Transfer
Coefficient h3 h4 UB Coefficient
h1 UL
h2

Numerical
Value 4 W/m2K 7 W/m2K 7.3 5.5 5.915W/m2K 0.45 1.75 8.04
W/m2K W/m2K W/m2K W/m2K W/m2K

Remarks
Design Design Design Design See Eqn. Design Design See Eqn.
3.4.11 3.4.1.2

1
UT  ------------------------------{3.4.1.1}
1 1

h1  h2 h3  h4
U L  U T  U B  U E ----------------------------{3.4.1.2}

81
 Table 3.37 Flat Plate Collector Design continued(Collector Parameters)

Quantity Overall Tube Tube Inside Plate Plate Thermal Collector Heat Collector Mass Specific Collector Heat
Spacing Diameter Thickness Conductivity Efficiency Transfer Efficiency Flow Heat Area Removal
Loss
Coefficient Di  K F Coefficient Factor Rate Capacity Ac Factor
F
Cp
UL
W Inside Tubes 
m Fr
hf
Numerical
Value 8 W/m2K 150 mm 31.75 mm 1.5 mm 237 W/mK 0.95418578 57.64W/m2 0.782 0.0066 3540 2 m2 0.782
8 K kg/s J/kgK
Remarks
See Previous Design Design Design From Tables See Eqn. See Eqn. See Eqn. Design From Design See Eqn.
Section 3.4.1.3 3.4.1.4 3.4.1.4 Tables 3.4.1.5

tanh[m(W  D) / 2] ----------------------------------------------------------{3.4.1.3}*
F
m(W  D) / 2
1
F  -------------------------{3.4.1.4}*

 1 1 1  
U LW    

U L [ D  (W  D ) F ] C B  Dh f 
m Cp   A U F   -------------------------------------------{3.4.1.5}
Fr  1  exp  c L 
AcU L F    m  C p 


h f Di UL kbb
Nu   3.66 , m2  CB 
k k 

82
Table 3.38 Collector Specifications
Name: Solar Collector Dimensional Specifications:
Type Flat Plate Length of Riser Tubes 1m
Flow Type Thermosyphon Riser Tube Diameter 0.03175m
Structure Riser –Header Riser Tube Thickness 2.5 mm
Operating Pressure 5.5 bar Number of Riser Tubes 14
Collector Tilt Angle 15 ˚ Tube center to center distance 15 cm
Heat Transfer Fluid Strong Ammonia Solution 54% Length of Header Tubes 210 cm
Specific Heat Capacity of Ammonia Solution 3540J/kgK Header Tube Inner Diameter 0.03175m
Heat Removal Factor 0.605 Header Tube Thickness 2.5mm
Transmittance –Absorptance Factor 0.55 Absorber Plate Thickness 2.7mm
Overall Heat Transfer Coefficient 8.04 W/m2K Collector Case Dimensions LxWxH 210x125x32cm
Storage Tank Heat Loss Coefficient 7.838 W/m2K Insulating Material Thickness(collector/storage) 5cm
Absorber Plate Thermal Conductivity 237 W/mK
Materials:
Riser Tubes Stainless Steel
Header Tubes Stainless Steel
Transparent Cover Window Glass
Case Iron + Galvanized Steel Sheets
Storage Tank Stainless Steel
Insulation Glass Wool

83
 183 cm

1
1 inch
4
1m

15cm
2m

66 cm

22cm

1
1 inch
4 1m

15cm

2m

Fig. 3.15 Flat Plate Collectors Field

84
3.4.1.2 Evacuated Collector Specification
Description: The evacuated collector field consists of two all-glass collector type connected
in series. The storage capacity of each collector is around 95 liter and hot water is delivered
from the collector field to the generator by a hot water circulation pump. The specification of
collector is presented in Table 3.39.

Table 3.39 Evacuated Collector Specification

Type All Glass Evacuated Tube Collector Length of Evacuated Tubes 1.5 m
Flow Type Thermosyphon Inner Tube Diameter 37mm
Structure All glass coaxial double-layer tubes Outer Tube Diameter 47mm
Operating Pressure 1MPa Tube Thickness 1.6 mm
Collector Tilt Angle 45˚ Number of tubes 16/15
Heat Transfer Fluid Water Volume of Tank 100 L
Absorption 0.94~0.96 Insulating Layer Thickness 5cm
Emittance 0.04~0.06 Materials
Vacuum P≤5×10-3Pa Evacuated Glass Tubes borosilicate
Transmittance of Outer Tube 0.91 Solar Selective Surface AlN/AIN-SS/Cu
Collector Heat Loss Coefficient ≤0.6W/ m2K Insulation Foaming polyurethane
Storage Tank Heat Loss Coefficient 10W/m2·K Storage Tank Stainless Steel

Storage
Tank

Vacuum
Tubes
Fig. 3.16 Flat Plate Collectors Field

3.4.2 Performance Predication of Solar Collectors

The solar energy collection system consists of the evacuated collectors which provides hot
water to heat the generator, and flat plate collectors which preheat the rich ammonia before
been pumped to the generator. Both collector fields have storage units attached to them. To

85
 Id
rd rd  Ib
z Hd
cos 
Rb 
cos  z
Rb
 Hd

s Id
 KT
Hd
cos s   tan  tan  Kt 
H I  I d  Ib
  H

I
Pg I
rt 
rt H

 1  cos    1  cos  
I T  I b Rb  I d    IIgT  
 2   2 

Fig. 3.17 Estimation of Hourly Radiation using the Isotropic Sky Correlation

86
check the system delivery capacity at the specified design temperature under the local
meteorological conditions, the variables that must be considered are daily thermal energy
delivery (load), mean inlet and delivery temperature, storage size & heat loss , irradiation on
the collectors ,ambient temperature. The variation of storage temperature with time were
predicted using an expression for the change in storage tank temperature for the specified
time period (hourly) in terms of the above mentioned quantities (see equation below).
t
Ts  Tsi  Fr Ac [( ) I (t )  U (Tm  Ta )]  (UA) s (Ts  Ta )  m
 l Cl (Ts  Tl ,r ) [1]
2(ms C s )
Estimation of Incident Radiation
Hourly values of solar radiation in the above mentioned performance prediction equation
were estimated from daily values using the Isotropic Sky correlation. Several terms related to
solar geometry and basic concepts referring to solar radiation are used, but not explained(to
obtain more details refer to specific textbooks[1]. The procedure of hourly radiation
estimation is illustrated in Fig. 3.17.

3.4.2.1 Long Term Performance of Evacuated Tube Collector

The evacuated collector field supplies the generator with hot water at a temperature of 88
˚C and a fixed flow rate of 35kg/hr and returns water back to the tank at 80˚ C. The
performance of the collector field was predicted under the above mentioned conditions
throughout the year for the period from one to four O'clock afternoon using recorded
meteorological data(monthly average). The result of performance prediction is presented in
the following charts(see Fig. 3.18) and the calculation procedure is presented in tables 3.40
&3.41 According to the result obtained it was decided to use two evacuated collectors to
supply the load with the hot water needed at the generator.

3.4.2.2 Performance Predication of Flat Plate Collector

The function of the flat plate collector field is to preheat the strong ammonia solution before
delivery to the generator. The performance was predicted under storage capacity of 65, 45
and 15 kg for December only because, according to Sudan meteorological statistics,
December is the month with the minimum radiation level. The performance prediction
procedure is presented in Tables 3.42 and 3.43 According to the result obtained it was
decided to use two collectors with 65 and 15 kg capacity to supply the system with preheated
strong ammonia solution necessary to operate the refrigeration unit.

87
 Temperature of Storage Tank At one O'Clock Temperature of Storage Tank at Two O'Clock

89.6 90.5

89.4
90

Temperature of Storage Tank In Degree Celsuis

Temperature of Storage Tank in degree celsuis
89.2

89.5 Storage Tank

89 Temperature

88.8 89

88.6
88.5

88.4
88
88.2

87.5
88

87.8 87
1 2 3 4 5 6 7 8 9 10 11 12 1 2 3 4 5 6 7 8 9 10 11 12
Months of the Year Month of the Year

Temperature of Storage Tank at 3 O'Clock Temperature of Storage Tank at 4 O'Clock

89 89

88.5

88

Temperature of Storage Tank in Degree Celsuis

88
Temperture of Tank in Degree Celsuis

87.5 87

87

86
86.5

86
85

85.5

84
85

84.5
83
1 2 3 4 5 6 7 8 9 10 11 12
1 2 3 4 5 6 7 8 9 10 11 12
Month of the year
Month of the Year

Fig. 3.18 Predicted Storage Tank Temperature

88
Table 3.40 Radiation on Tilted Surface for the Period for Evacuated Collector from 12 to 4 O`Clock
Quantity Hourly Sunset Latitu Ratio of Ratio of Hourly Daily Total Dialy Day Hourly Ground Zenith
Angle of Hourly Hourly Radiation
Hourly
Clearance Beam
Slope Ratio of Angle

Radiation Hour Total Diffuse Diffuse Beam Angle
de Declinatio Total Diffuse Radiation Angle Radiati Reflecta of
 z
on Tilted Angle n Radiation Radiation Radiation Inciden
Radiation Radiation on
Surface To Daily to Daily H Hd KT nce on Tilted
Surface to
ce
s  Total
Radiation
Diffuse
Radiation I Id Ib Pg that on 
IT Horizontal
rt Surface
rd Rb
Numerical
Value See Eqn See qn. From From See Eqn. From Tables From From See Eqn
15.5˚ See See Eqn 0.2 65˚ See Eqn From From
3.4.2.1 3.4.2.2 Graphs Graphs 3.4.2.4 3.4.2.5 Graphs Tables 3.4.2.6 3.4.2.3 Graphs Graphs
Table

Units MJ / m 2 MJ / m2 / Day MJ / m 2 MJ / m 2 MJ / m 2

12 to 1 1 to 2 2 to 3 3 to 4 12 to 1 1 to 2 2 to 3 3 to 4
O'Clock O`Clock O'Clock O'Clock O'Clock O`Clock O'Clock O'Clock
Hourly Radiation Estimation I 1201 I 0 1 0 2 I 0203 I T 1201 I T 1201 I T 1201 I T 1201
I 0304
Month
 s KT
H
Jan -21.27 83.80 0.70 20.30 3.2886 3.842 2.194 1.1774 3.8665 3.3608 2.7355 1.1978
Feb -13.29 86.24 0.72 23.10 3.6498 3.234 2.541 1.571 3.8287 3.3178 2.7348 1.8542
March -2.28 89.37 0.69 24.50 3.969 3.43 2.646 1.421 4.0238 3.4913 2.4995 1.3594
April 9.41 92.63 0.69 25.90 3.95 3.45 2.7 1.65 3.3066 2.7864 2.1693 1.2954
May 18.79 95.414 0.65 24.70 4.0508 3.555 2.717 1.5808 3.0098 2.5949 1.8915 1.0426
June 23.31 96.86 0.62 23.60 4.1536 3.5872 2.596 1.3216 2.8883 2.4895 1.6724 1.2115
July 21.52 96.28 0.60 23.00 3.0554 2.5933 1.2121 0.7648 3.0515 2.5933 1.2121 0.7648
August 13.78 93.9 061 22.80 4.1602 3.527 2.487 1.7572 3.2676 2.7347 1.9090 0.8724
September 2.22 90.62 0.64 22.90 3.6084 3.034 2.3977 1.2830 3.6084 3.0304 2.3977 1.2830
October -9.60 87.31 0.67 21.90 3.5478 3.066 2.3652 1.2702 3.6113 3.1168 1.9964 1.0516
November -19.15 84.47 0.71 21.00 3.318 2.898 2.268 1.386 3.8102 3.495 2.7515 1.7686
December -23.24 83.160 0.72 20.10 3.4617 3.5062 2.863 1.8904 3.4617 3.5062 2.8634 1.8905
 1  cos  
I T  I b Rb  I d 
 1  cos   -----{3.4.2.1}[1]
  I g   cos s   tan  tan  ------- --------------------------{3.4.2.2}
 2   2 
cos  --------------------------------------------------{3.4.2.3} rt 
I -----------------------------{3.4.2.4}
Rb 
cos  z H

rd 
Id ----------------------------{3.4.2.5} I  I d  I b ---------------------------{3.4.2.6}
Hd

89
 Table 3.41 Storage Tank Temperature Variation with Radiation for Evacuated Collector.

Number of Collectors Total Collectors Area Surface Area of Mass of Water in Heat Removal Absorptance- Heat Loss
Storage Tank the Storage Tank Factor Transmittance
Ac M Fr Factor
Coefficient
Quantity
As  Of Storage
Tank
UL
Numerical
Value 2 1.62 m2 4.032 m2 190 kg 0.9 0.5 O.78

From Tables
Design Collector Area = 0.81 Tank Surface Each Storage Tank Design Design (apply a Safety factor
Remarks Area = 2.016 = 95 kg of 2)

Flow Rate of Hot Water = 35.8kg/hr Flow Rate of Hot Water = 26.8kg/hr
Load Temperature Return = 80˚C Load Temperature Return = 80˚C
Storage Tank Storage Storage Tank Storage Tank Storage Tank Storage Tank Storage Tank Storage Tank Storage Tank
Temperature Tank Temperature Temperature Temperature Temperature Temperature Temperature Temperature
at 12 0'Clock Temperatur at 2 at 3 at 12 0'Clock at 1 0'Clock at 2 at 3 at 3
Month e
0'Clock 0'Clock 0'Clock 0'Clock 0'Clock
Ts 12 at 1 0'Clock Ts 02 Ts 03 Ts 12 Ts 01 Ts 02 Ts 03 Ts 04
Ts 01
Jan 88 ˚C 88.7405 88.8757 88.375 88 88.9932 89.3173 89.0859 87.2733
Feb 88 88.7618 88.9169 88.4679 88 89.0144 89.413 89.1859 88.0922
March 88 89.0604 89.4552 88.7766 88 89.313 89.96 89.5137 87.966
April 88 88.5027 88.4314 87.76 88 88.7553 88.9196 88.4529 87.1548
May 88 88.2812 87.1668 87.37 88 88.5338 88.6477 88.0481 86.6614
June 88 88.1106 87.8273 86.77 88 88.3632 88.3029 87.434 86.2102
July 88 88.1537 87.846 86.194 88 88.4063 88.3229 86.856 85.1232
August 88 88.3736 88.1702 87.1709 88 88.6262 88.5983 87.7971 86.0464
September 88 88.7874 88.8883 88.3443 88 89.4094 89.6983 89.3282 87.572
October 88 88.7828 88.9668 87.9938 88 89.0354 89.4636 88.7085 88.0526
November 88 88.798 89.1755 88.7504 88 89.0514 89.6734 89.4727 88.2982
December 88 88.3547 88.7343 88.4129 88 88.6073 89.2176 89.1086 88.0202

t
Ts  Tsi  Fr Ac [( ) I (t )  U (Tm  Ta )]  (UA) s (Ts  Ta )  m
 l Cl (Ts  Tl ,r )
2(ms Cs )

90
Table 3.42 Estimation of Hourly Radiation on a Sloped Surface from Daily Data for Flat Plate Collector (Constant Values)

Declination Latitude Sunset Hour The Monthly Average Daily Diffuse Tilt Angle Ground Diffuse
Quantity
Angle
 Angle Average Radiation Component of  Reflectance
 s Clearance H(June) Daily Radiation Factor
Angle HD Pg
KT
Numerical
Value -23 15.5 83.24 0.72 20.10 3.892869711 15.5 0.2
MJ/day/m2

Remarks From Tables From Tables

cos s   tan tan  From Tables From Tables From Graphs Design Constant

Table 3.43 Estimation of Hourly Radiation on a Sloped Surface from Daily Data for Flat Plate Collector continued(Time Dependent Variables)

Angle of Zenith Geometric Ratio of Ratio of hourly Hourly Radiation Hourly Diffuse Hourly Hourly Radiation On Tilted
Quantity Incidence Angle Factor hourly Total Diffuse I Radiation Beam Ground
 z Rb Radiation to Radiation to Daily Id Radiation IT
Daily Total Total Ib
rt rd
From Graphs From cos  From Graphs From Graphs
Rb  I Id
Remarks Graphs cos  z
rt  rd  Ib  I  Id MJ W
H Hd
7---8 0.35 0.25 1.4 0.03 0.04 0.603 0.1556 0.4474 0.7813 217W
8---9 0.55 0.45 1.22 0.07 0.078 1.407 0.30342 1.10358 1.649 458W
Time of the Day

9--10 0.725 0.6 1.21 0.106 0.11 2.1306 0.4279 1.7027 2.488 619W
10-11 0.85 0.725 1.172 0.132 0.13 2.6532 0.5057 2.1475 3.023 840W
11-12 0.925 0.775 1.194 0.15 0.14 3.015 0.5446 2.4704 3.495 971W
12--1 0.925 0.775 1.194 0.15 0.14 3.015 0.5446 2.4704 3.495 971W
1---2 0.85 0.85 1.172 0.132 0.13 2.6532 0.5057 2.1475 3.023 840W
2---3 0.725 0.6 1.21 0.106 0.11 2.1306 0.4279 1.7027 2.488 691
3---4 0.55 0.45 1.22 0.07 0.78 1.407 0.30342 1.10358 1.649 458
4---5 0.35 0.25 1.4 0.03 0.04 0.603 0.1556 0.4474 0.7813 217

91
 Table 3.44 Storage Tank Temperature Variation with Radiation for Flat Plate Collector (June)

Specific Heat Transmittance Heat Overall Heat Storage Tilt Angle Average Latitude
Quantity Capacity of –Absorptance Removal Transfer Tank Heat  Daily 
Ammonia Factor Factor Coefficient Loss Radiation
Solution  Fr UL Coefficient H
Cp UA

Numerical
Value 3540J/KgK 0.55 0.605 8.04 W/m2K 7.838 W/mk 15.5 20.10 15.5
MJ/day/m2

Remarks See Table 6.1 From Tables (December) From

From Tables Design Design After Design From Tables
Applying a Tables
Safety Factor
of 2

Hourly Ambient Initial Final Tank Initial Tank Final Tank Initial Final
Quantity Radiation On Temperature Tank Temperature Temperature Temperature Tank Tank
Tilted Ground
IT
Ta Temperatu
re
Ts Ts Ts Temperatu
re
Temperat
-ure
Ts Ts Ts
Table 6.2 From Tables Storage Tank Mass Storage Tank Mass Storage Tank Mass
45 kg 15k 65kg
Remarks
7---8 27 27.8 27 29.6 27 27.6
Time of the Day

8---9 20.9 27.85 32.7 29.6 42.071 27.6 31.03

9--10 22.7 32.7 39.9 42.071 52.6 31.03 36.49
10-11 24.7 39.9 47.4 52.6 60.08 36.49 42.63
11-12 27 47.4 54.8 60.08 67.2 42.63 49.05
12--1 29.2 54.8 59.8 -- -- -- --
1---2 30.55 59.8 61.3 -- -- -- --
2---3 31.7 61.3 60.21 -- -- -- --
3---4 32.3 -- -- -- -- -- --

t
Ts  Ts  [Qu  (UA)(Ts  Ta )]
mC p

92
 Chapter 4
4. Construction and Performance Testing
This section describes the construction, assembly and reports the measured
performance of the absorption components.

4.1 Construction and Assembly

The components of the experimental unit were constructed at Khartoum University
Workshop. After construction, the unit was assembled and tested on the roof of
Engineering College at Khartoum University. The following figures (4.1&4.2) show
the refrigeration unit after assembly. The piping lines that join the system
components together were connected by threaded fittings, and a special pipe binding
material for high pressure pipe connection were used. All components, expect few
fittings and piping in very limited locations, are made of stainless steel. Electric
welding machine and stainless steel welding material were used in fabrication of
components. Among the components, generator and absorber experienced some
changes from their original designs due to technical and financial limitations. All
components fabricated at the University Workshop have been subjected to pressure
tests using an air compressor before assembly. Leakages revealed during the pressure
tests were treated using cold welding materials.

Fig. 4.1 The Absorption Refrigeration Unit after Assembly 93

Fig. 4.2 The Absorption Refrigeration Unit After Assembly

94
 4.2 Materials and Methods
4.2.1 Description of Experimental Set Up

Condenser Generator

Solution Heat
Exchanger

Evaporator Absorber

Dial Thermometer Direction of Flow

Pressure Gauge Hot Water Circulation
Pitot Tube Flow Meter Ammonia Solution Circulation
Non-return Valve Refrigerant Circulation
Circulation Pump Cooling Water
1,2….etc. State points as described in chapter 3 Solar Radiation
Fig. 4.3 Layout of the Experimental Unit and Positions of Testing Instruments

4.2.1.1 Overall Description of Experimental Unit

figure 4.3 illustrates the main components of the absorption refrigeration cycle and
positions of the measuring instruments. Rich ammonia solution in the storage of the
flat collector field is pumped to the generator(7) through the solution heat exchanger,
where it is heated by hot water from the evacuated collector field and then the
refrigerant in it is boiled off and rise to the condenser. High-pressure liquid
refrigerant (2) from the condenser passes into the evaporator (4) through an expansion
valve (3) that reduces the pressure of the refrigerant to the low pressure existing in the
evaporator. The liquid refrigerant (3) vaporizes in the evaporator by absorbing heat
from the material being cooled( water in a tank). The resulting low-pressure vapor (4)
passes to the absorber, where it is absorbed by the poor ammonia solution coming
from the generator (8). The poor ammonia flows to the absorber, after passing through

95
a solution heat exchanger (9), and giving off heat to the rich solution coming from the
flat plate collector field to the generator. At the absorber a rich ammonia solution is
thus formed (5). The rich solution (5) will accumulate in the absorber and ammonia
vapor from the evaporator continues to dissolve in the poor ammonia solution even
after the unit is shut down after six hours of operations. The next day at sun shine, the
rich solution in the absorber is allowed to flow to the storages of the flat plate
collector fields, where it starts to gain heat as a result of exposure to solar radiation.
When both the evacuated and flat plate collector fields have reached the unit operating
temperature, the cycle is powered on and the operation continues for 6 hours. The
following figures displays the flow streams in the heat exchangers. Table 4.1 and
photos (Fig. 4.4 to 4.9) describe the flow streams directions and the corresponding
state points.
Table 4.1 Direction and State of Flow Streams in the Heat Exchangers
Ammonia Vapor from generator to Hot water from generator to evacuated
condenser collector
Liquid Ammonia from condenser Rich ammonia solution from double tube heat
exchanger to generator
Hot Water from evacuated collector to Rich ammonia solution from flat plate
generator collectors to solution heat exchanger
Ammonia vapor from evaporator to Strong ammonia solution from absorber to flat
absorber plate collector
Poor ammonia from generator to Poor Ammonia solution from solution heat
solution heat exchanger exchanger to absorber

Fig.4.4 Flow of Rich Ammonia from

Absorber to Flat Plate Collectors(greater
storage capacity)at sun shine.

Fig. 4.5 Flow of Rich Ammonia

Solution to Flat Plate Collector
(smaller storage capacity) at Sunshine.

96
 Fig. 4.7 Flow of Rich &Poor Ammonia Solutions to and from Solution
Fig.4.6 Flow of Poor Ammonia Solution , Ammonia Vapor and Hot Heat Exchanger and Absorber.
Water(from evacuated collectors) to and from Generator.

97
Fig.4.8 Circulation of Hot Water from Evacuated Collector to Fig.4.9 Flow of Rich Ammonia Solution From Absorber to Flat Plate
Generator and back to Collector. Collectors
4.2.1.2 Measuring Instruments
The positions of the instruments used in testing of the absorption unit are shown in
figure 4.3, and instruments are shown in figure 4.10 , all measuring instruments are
inserted in the pipe lines except the digital thermometers are inserted in evaporator
and absorber water tanks:

Digital Stem Thermometer

Liquid Filled Pressure Gauge
Pitot Tube Insertion Flow Meter Digital Thermometer

Fig. 4.10 Measuring Instruments of the Experimental Unit

4.2.1.3 Detailed Description of the Test Components

(i) Generator
The generator is the main component of this absorption unit, in which rich ammonia
solution is heated until ammonia vapor is driven off to the condenser. A shell and
tube heat exchanger has been designed and constructed to undertake the generation
process. The tube bundle (see Fig.4.13 ) consists of 5 Stainless tubes (3/4"OD) and
25 baffles (see Fig. 4.11). The generator has double tube sheets so that any leaks can
be detected immediately. The tube sheets(6mm thickness) are made of stainless steel
round plates(see Fig.4.12 ). The rich ammonia flows in the tubes and the hot water
flows in the shell (see Fig.14). Note: Section 3.3 contains more details about heat
exchangers specifications.

(ii) Condenser and Receiver

In the condenser, the ammonia vapor coming from the generator is condensed into
liquid by cooling water in a tank(Fig16). The condenser coil(see Fig.4.15) is made up
of stainless steel pipe 13 m long (1.25"OD). After condensation the ammonia liquid
passes to the receiver(see Fig.4.17) m long ("OD).

98
 Fig. 4.11 Baffles
Fig. 4.12 Tube Sheet

Fig. 4.13 Tube Bundle

Fig. 4.14 Shell and Tube Heat Exchanger

Fig.4.15 Condenser Coil in Fig.4.16 Water Tank Fig. 4.17 Ammonia Receiver
Tank

(iii) Evaporator
In the evaporator the liquid ammonia expands through a throttle valve(see Fig.4.18)
changing from liquid to vapor , in doing so it removes heat from the water tank (see.
Fig.4.18). The evaporator coil(see Fig.4.19) is made of a stainless steel U shape pipe
(0.75"OD).

99
 Fig.4.19 Evaporator coil

Fig.4.18 Evaporator coil in the insulated water Tank

(iv) Solution Heat Exchanger

The solution heat exchanger preheats the rich ammonia solution flowing to the
generator by cooling the weak ammonia solution exiting it. This heat exchanger
reduces the generator heat duty requirement by preheating the rich ammonia solution,
and the absorber heat duty by reducing the poor ammonia solution temperature
entering the absorber. A Double tube exchanger is used as the solution heat
exchanger(see Fig4.20). The rich ammonia solution flow in the inner tube(1"OD) and
the weak ammonia solution flows in the outer tube(1.5"OD).

100
Fig. 4.20 Solution Heat Exchanger
 (v) Absorber
In the absorber ammonia vapor flowing from the evaporator is absorbed by the poor
solution from the generator, rejecting its heat to cold stagnant water in a tank(see
Fig.4.21). A header-riser tube arrangement is used as an absorber (see Fig.4.21 –a ). It
consists of two rows (upper &lower) of tubes (4"OD). Poor ammonia solution flows
from the solution heat exchanger to the header of the upper row(see Fig.4.21-a)and
then flows by gravity to the tubes of lower rows. Ammonia vapor from the evaporator
enters the header of the upper and lower rows simultaneously through nozzles fixed to
tubes (see Fig.4.21-c). The nozzles distribute the ammonia equally among the risers
(see Fig.4.21-b) of the absorber. Then the rich ammonia solution flows to the flat
plate collectors through opening in the lower header of the absorber. Stream Flows to
and from the absorber are shown in figure. 4.22.

(a) (b) (c)

Fig. 4.21 Absorber

101
Fig. 4.22 Stream Flows in the Absorber
 (vi) Flat Plate Collectors
In the flat plate collectors field (solar warm up cycle), the rich ammonia, after being
discharged from the absorber at sunshine, would be left to warm up in the collectors
before the system operation. The flat plate collector field is riser-header type and
consists of two (21m) collectors. Each collector consists of single-glazed absorber
plate coated with dull-black commercial paint. The flat-plate absorber was inclined at
15˚to the horizontal and facing due south (latitude of Khartoum=15˚). It was made
from 2.7 mm thick iron sheet and has 14 tubes ( each 1m long (1.25"OD). The
absorber plate is encased in a metallic box insulated by 5cm thick fiber wool on the
bottom and sides. One collector has 110 liter capacity storage tank (see Fig. 4.23),
while the other has 36 (see Fig. 4.24) liter storage capacity. The storage tanks are
insulated and fitted with insulated pipeline connections for the circulation of the
working fluid from the collector to the solution heat exchanger, and back from the
absorber to the collectors. A pump (see Fig.4.25) is used to drive the rich ammonia
solution from the collectors to the solution heat exchanger. The ammonia solution will
be withdrawn from the collector with smaller capacity (36 liter) first and then from
the collector with bigger (110 liter), since it would have higher temperature due to the
less quantity of ammonia solution.

Fig. 4.23 Flat Plate Collector 110 Liter Storage Fig.4.24 Flat Plate Collector 36 liter Storage

102
Fig. 4.25 Rich Solution Pump
 (vii) Evacuated Tube Collectors.
In order to evaporate ammonia from the rich ammonia solution flowing in the
generator, hot water in the range of 80-88 ˚C is passed through the shell of the
generator in a counter current flow. The hot water comes from the evacuated tube
field (Dewar Flask type) which consists of two collectors with storage capacity of 90
liter each. The maximum operating temperature of the collector is 100˚C. The two
collectors are connected in series (parallel connection is possible too). The hot water
flowing out of the generator discharges to a upper hot water tank (see Fig.4.26). Then
the hot water flows by gravity from the tank to the evacuated collector 1, then from
the evacuated collector 1 to evacuated collector 2. Next it flows by gravity to the
lower storage tank (see Fig.4.26). From there it is pumped up to the generator shell by
hot water circulation pump (see Fig.4.27) and then to the hot water tank.

Fig.4.26 Evacuated Collector

Field

Fig. 4.27 Hot Water Pumping from Storage to Generator

103
4.2.2 Test Procedure and Result
4.2.2.1 Initial Observed Operation of the System Components
(i) Evacuated Collector Test
Pre-design performance test was conducted on the evacuated collector which had
been purchased from local market. The test objective was to study the variation of
temperature of the storage tank through the day under no load conditions. The test
result was used to predict the long term performance and to decide the number of
collectors needed to operate the system(see Table 4.2 and Fig.4.28).
Table 4.2 No Load Test Result of Evacuated Collector
Time Ambient Temperature ( ˚C) Outlet Water Temperature( ˚C) Date:
9.00 28.5 73 5/4/2005
10:00 29 72
11:00 30.1 72 Weather
12:00 32.4 80 Condition:
1:00 35.2 81.2 Wind
2:00 35.7 89 &Dust
3:00 37.8 91
4:00 37.5 93.3

Time Ambient Temperature ( ˚C) Outlet Water Temperature( ˚C) Date:

9.00 31.7 75 10/4/2005
10:00 32.9 76.9
11:00 36.0 82.7 Weather
12:00 38.0 86.7 Condition:
1:00 37.5 90.4 Sunny
2:00 39.5 93.4
3:00 41.7 96.5
4:00 43.1 98

Time Ambient Temperature ( ˚C) Outlet Water Temperature( ˚C) Date:

9.00 25.4 77.4 19/4/2005
10:00 27.4 76.4
11:00 29.8 72.4 Weather
12:00 32 83 Condition:
1:00 33 85 Sunny
2:00 32.8 90
3:00 34 95
4:00 34.6 98.6

Time Ambient Temperature ( ˚C) Outlet Water Temperature( ˚C) Date:

9.00 33.2 75 24/4/2005
10:00 31.8 75.4
11:00 29 76 Weather
12:00 28.7 79.9 Condition:
1:00 42.2 80.4 Cloudy
2:00 34.3 83.5
3:00 33.3 85
4:00 41 85

104
 Temp ˚C

Time of the Day

Fig.4.28 Variation of Tank Temperature with Time Under No-Load Condition

(ii) Flat Plate Collector Test
The performance of flat plate solar collectors field (the preheating solar components)
was observed (from December to March) and it was found that:
 The flat plate collector with 36 liter capacity recorded 100˚ C temperature at 3
0'clock (time for maximum yield for systems with storage) and the designed
temperature was 67˚C with 15 liters.
 The flat plate collector with 110 liter capacity recorded a temperature of 80˚
at 3 0'clock and the designed temperature was 60 ˚C with 65 litre capacity
(iii) Testing of the Experimental Unit.
After assembly of the experimental setup , tests had been conducted using
compressed air to detect leakages and water to test the proper functioning of pumps. .

4.2.2.2 Test Procedure of Experimental Unit and Result

It was decided to carry out experiments on the heat exchangers of the prototype and
the solar collector fields individually , so the absorption machine will be used as a test
bank to validate the design procedure and to evaluate the impact of the system
components performance on the overall performance of the absorption machine. The
arrangement of components of the experimental setup was relatively flexible and
some minor modifications had to be applied to carry out experiments on the different

105
components separately. These special arrangements would be described for each
component separately. The main characteristics of the different cases under study are
shown in Table 4.3 below. The suggested test procedure for each component and
result obtained would be presented in the following sections.
Table 4.3 Overview of Cases under Study.
Components under Study
Shell and Double Flat Plate Evacuated
Cases Parameters Tested Tube Heat Tube Heat Collector Collector
under Exchanger Exchanger Field Field
Study
Case 1 Optimum Flow Rate   

Case 2 Compliance with

Designed Temperature    
range
Case 3 Efficiency 
Case 4 Ability to meet Designed
Load    
Case 5 Effect on Overall System
Performance    

(i) Testing of the Shell and Tube Heat Exchanger

Experimental Set up: Experiments were conducted on the shell and tube heat
exchanger described in section 4.2.1.3(i) above. Hot water was supplied by the
evacuated collector field (see Fig.4.27) and colder water was supplied by the flat plate
collector with 110 liter capacity (see Fig.4.23)
Experimental Procedure:
The solution pump (see Fig.4.25 above ) delivered water from the flat plate collector
to the heat exchanger tube bundle , while the hot water circulation pump(see Fig.4.27
above) delivered hot water from the evacuated collector field to the shell of the heat
exchanger. It was waited until the steady state had been reached. At steady state, all
the four temperatures and flow rates of cold and hot fluid do not change. The readings
recorded are shown below (see Table 4.4).
Operational Problems and Special Arrangements:
 Dial thermometers malfunction occurred in 4 of the measuring points, so a
portable digital thermometer was used to measure temperatures at those
positions.
 The flow meter inserted between the flat plate collector field and the tank was
damaged, and a measuring jar and stop watch was used to measure the flow
rate from a discharge valve.

106
  The flow of water from the flat plate collector could not be adjusted due to
inability of the pump to deliver liquid at the designed flow rate, so the solution
pump in figure 4.25 was replaced by a hot water circulation pump similar to
the evacuated collector field pump (see Fig. 4.27)
 A portable submersible pump was used to supply water at ambient
temperature (37-39) since the flat plate collector could not deliver water at this
temperature.
 According to the previous arrangement, an additional valve was inserted in
the delivery line of the solution pump so as to supply the shell and tube heat
exchanger with hot water from the flat plate collector field (without passing
through the double tube heat exchanger). A flexible tube was used to connect
this valve to a side valve that led the flow to the inner tube bundle (see
Fig.4.30) A flexible tube was connected to the discharge line of the tube
bundle so as to redirect flow to the flat plate collector storage tank (see
Fig.3.29) instead of the outer tube of the double heat exchanger.
Note: the first four above arrangements were also applied in the following test.

Fig. 4.29 Connection of Tube Bundle Fig. 4.30 Connection of Solution

Discharge Line to Flat Plate Collector. Pump Delivery Line to Tube Bundle
Inlet.
. Table 4.4 Hot Water –Water Test Result of Shell and Tube Heat Exchanger
Date of Volumetric Volumetric Fluid Temperatures (˚C)
Experiment Flow Rate Flow Rate Cold Cold Hot Hot
of Hot Water of Cold Water Fluid Fluid Fluid Fluid
(lpm) (lpm) Inlet Outlet Inlet Outlet
23/7/2009 0.5/3 0.5 51 59 66 62
23/7/2009 4 1.6 50 55 67 48
27/7/2009 2 0.67 45 52 64 59
13/8/2009 0.63 0.74 38 46 60 50
18/8/2009 0.4 0.55 39 50 71 54
19/8/2009 1.36 0.5 58 57 65 61

107
 (ii) Testing of Double Tube Heat Exchanger
Experimental Set up: Experiments were conducted on the Double Tube Heat
Exchanger described in section 4.2.1.3(iv) above. Hot water was supplied by the
evacuated collector field (see Fig.4.27) and colder water was supplied by the flat
plate collector with 110 liter capacity (see Fig.4.23)
Experimental Procedure:
The solution pump (see Fig.4.25) delivered water from the flat plate collector to the
heat exchanger inner tube , while the hot water circulation pump(see Fig.4.27)
delivered hot water from the evacuated collector field to the outer tube of the heat
exchanger. It was waited until the steady state had been reached.. The readings
recorded are shown (see Table 4.5).
Operational Problems and Special Arrangements:
 The first four special arrangement in section 4.2.2.2.(i) above were applied in
addition to the following arrangements.
 In the original experimental set up the inner tube of the double tube heat
exchanger admitted fluid flow from flat plate collector and discharge it to the
generator. In the modified experimental set up the flow from the inner tube of
the double heat exchanger was directed to discharge in the flat plate
collector(see Fig.4.32).
 A flexible tube was made to connect the hot pump delivery line to the inlet of
outer tube of the double heat exchanger (see Fig.4.31). In this way the
evacuated collector field can supply double tube heat exchanger in addition to
the shell and tube heat exchanger. Then the fluid coming out of the outlet of
the outer tube , would flow to the evacuated collector (see Fig.4.33)

Fig. 4.31Connection of Evacuated Collector

Fig. 4.32 Connection of Outlet of Inner Tube of
Field to Outer Tube Double Tube Heat
Double Heat Exchanger to Flat Plate Collector
Exchanger.
108
 Fig. 4.33Connection of Outlet Tube of Double Heat
Exchanger to Evacuated Collector

Table 4.5 Hot Water –Water Test Result of Double Tube Heat Exchanger
Day of the Volumetric Volumetric Fluid Temperatures (˚C)
Experiment Flow Rate Flow Rate Cold Cold Hot Hot
of Hot Water of Cold Water Fluid Fluid Fluid Fluid
(lpm) (lpm) Inlet Outlet Inlet Outlet
23/7/2009 0.57 0.9 51 57 66 51
26/7/2009 1.8 2.4 50 54 67 56
27/7/2009 1.5 2.7 45 49 64 45
10/8/2009 3.3 2.5 34 55 72 59
10/8/2009 2.5 0.8 37 50 64 58
19/8/2009 0.625 0.7 38 46 60 50
29/7/2009 0.6 0.7 47 57 71 49
13/8/2009 0.55 0.6 38 55 69 41
13/8/2009 0.625 0.5 38 50 61 42
18/8/2009 0.65 0.41 37 59 71 42
19/8/2009 0.5 0.65 58 58 65 55.6

(iii) Testing of Flat Plate Collector

a) Test of Absorber Plate of Flat Plate Collector
Experimental Set up: Experiments were conducted on the flat plate collector
(with 36 liter capacity) described in section 4.2.1.3(vi) above. The storage tank
was isolated from the absorber plate(see Fig. 4.34). Stream flows in and out of the
collector absorber plate are shown in figure 4.34. Water was supplied from the
lower hot water storage tank of the evacuated collector field and passed through
flexible tubes to the inlet of the lower header of the absorber plate(see Fig.4.35).
Experimental Procedure: The standard procedure for collector testing is to
operate the collector under conditions in which operation is nearly steady i.e. the

109
 radiation and other conditions are essentially constant for a time long enough for
the outlet temperature and useful gain to become steady. In order to achieve the
above conditions water at a constant temperature was supplied to the collector
absorber plate at constant flow rate temperature . Tests were made with a range of
inlet temperature conditions. To minimize effects of heat capacity of collectors,
tests are usually made in nearly symmetrical pairs, one before and one after solar
noon, with results of pairs averaged. Instantaneous efficiencies are determined
from  i  m
 C p (T0  Ti ) Ac GT for the averaged pairs. The flow rate of water at

outlet of the collector was measured using a stop watch and a measuring cylinder.
A digital thermometer was used to measure the inlet and outlet temperature (see
Table 4.6).

Operational Problems and Special Arrangements:

 In order to obtain different inlet temperatures at constant flow rate, the lower
hot water tank of the evacuated field was used to supply water at different inlet
temperatures (see Fig.4.35). That was accomplished by mixing water at
ambient temperature with hot water from the collector field inside the tank.
Then the mixed water would delivered to the collector absorber plate through
the hot water circulation pump (see Fig.4.35).

Table 4.6 Efficiency Test of Flat Plate Collector

Time of Ambient Radiati Flow Inlet and Outlet
the Day Temperature on Rate Heat Transfer Fluid
˚C W/m2 
m Temperature ˚C
Ta G Ti To

1:00 33 709 2liters/min 40 50

1:20 37 47 53
1:30 785
1:35 37 54 59
2:00 739

110
 Fig. 4.34 Inlet and Outlet flow through the
Collector Plate

Fig. 4.35 Mixing of Hot Water with cold Water

to Obtain Water with Different Temperature

b) Observation of Storage Tank Variation

Experimental Set up and Procedure: The storage tank temperature variation with
weather conditions of the collector field described in section 4.2.1.3.(vi) was
observed at different times of the year and result is shown in Table 4.7

111
 Table 4.7 Variation of Tank Temperature with Weather Conditions of Flat Plate Collectors
Temperature(˚C) at Different Times of the Day

Date
Experimental 10:00 10:30 11:00 11:30 12:00 12:30 1:00 1:30 2:00 2: 3:00 3:30 4:00 4:30
Device 30

Test
no.1
34 46 48 63 72
Collector Tank with 150 liter Storage Tank

17/8/2009
Test
no.2
65 67 701 70 68

March
Test 76
no.3
50 58 62 70 78 78

2/8/2009
Test
no.4
44 47 56 60 66 68

2
5/8/2009
Test
no.5
56 58 64 70

Test
no.6
42 44 52 58
C
Collect

e Tank
Storag

Test
38 38 42 96 100
Liter
with

no.1
36
or

2/8/2009
without Storage

Test 100
41 58 81 89 97
Absorber Plate

no.1

5/8/2009
Test
Tank

no.1
51 66 81 87

(iv) Testing of Evacuated Collector Field

a) Effect of Hot Water Withdrawal on Storage Tank Temperature
Experimental Set up: Experiments were conducted on the evacuated collector field
which consists of two collectors connected in series(parallel connection is also
possible). The stream flows in and out of the collectors field are described in figure
4.26 &4.27 above. Hot Water from the shell and tube heat exchanger enters the upper
storage tank and flows by gravity to collector 1, through collector 2 and finally
discharges through the upper tank port of collector 2 to the lower storage tank (see
Fig.4.26 &4.27).
Experimental Procedure: To study the effect of hot water withdrawal rate on the
collector field average delivery temperature, hot water was withdrawn at a fixed flow
rate from the collector field to supply the load of shell and tube heat exchanger and
then returned back the collector through the upper storage tank. The flow rate and
return tank temperature were measured at the inlet of the upper storage tank, while the

112
outlet temperature of the field was measured at the inlet of the lower storage tank.
Ambient temperature and general weather conditions were also recorded (see Table
4.8)
Table 4.8 Effect of Hot Water Withrawal on Storage Tank Temperature of Evacuated Collector.
Date: 17/8/2009 17/8/2009 Evacuated Ambient Temperature of
Time Collector Tank Temperature Water Inlet to
Temperature ˚C Tank of
Evacuated
10:00 81 39
Withdrawal
Before Hot

11:00 82 38
Water

12:00 85 40

1:00) 88 32 45
Water of 0.5L/min

2:00 87 39.6 55
Start of Hot

Flow Rate

3:00 86 41 48

4:00 84 42 49

b) Observation of Storage Tank Variation

Experimental Set up and Procedure: The storage tank temperature variation with
weather conditions of the collector field described in section 4.2.1.3.(vii) was
observed at different times of the year and result is shown in Table. 4.9
Table 4.9 Variation of Tank Temperature with Weather Conditions of Evacuated Collector
Temperature(˚C) at Different Times of the Day

Date
Experimental 10:00 10:30 11:00 11:30 12:00 12:30 1:00 1:30 2:00 2: 3:00 3:30 4:00 4:30
Device 30

17/8/
2009
Evacuated Collector Tank Storage

Test
81 82 85 88
no.1
Test
March

no.2
55 65 72 79 8 81 83
0
Test 75
Tank

2/8/
2009

no.3
79 80 84 85
Test
5/8/
2009

no.4
70 73 77 81
2

Test
5/8
/2009

no.5
72 72 75 77
C

113
 4.3 Analysis of Result
Several tests have been carried out on the absorption machine components in
order to investigate a range of operating conditions. The experimental data has been
compared with the design parameters. The conclusions reported will lead to the future
revisions of design and improvements to be implemented in order to achieve a better
performance and reliability.

4.3.1 Shell and Tube Heat Exchanger

Operating parameters under investigation are:
 Energy delivered or gained by flow streams
 Temperature ranges
 Flow Rates
 Heat Transfer Coefficient
Table 4.10 & 4.11 below shows the test result of the shell and tube heat exchanger and
the designed operating parameters respectively.

Table 4.10 Experimental Data of Shell and Tube Heat Exchanger

Flow Flow Cold Water Cold Water Hot Water Hot Water Log Mean Heat Heat Heat
Rate Rate Inlet Outlet Inlet Outlet Temperature Energy Energy Transfer
of Hot of Temperature Temperature Temperature Temperature Difference Lost By Gained
Water Cold (˚C) (˚C) (˚C) (˚C) Hot By Coefficient
(lpm) Water Stream Cold (W/m2 k)
(lpm) (Kw) Stream
(Kw)
2 0.67 45 52 64 59 12.97 0.704 0.330 48.83
0.63 0.74 38 46 60 50 12.97 0.443 0.416 61.64
0.4 0.55 39 50 71 54 17.82 0.4789 0.426 45.83

Table 4.11 Designed Operating Parameters of Shell and Tube Heat Exchanger

Flow Flow Cold Stream Cold Stream Hot Stream Hot Stream Log Mean Heat Heat Heat
Rate of Rate Inlet Outlet Inlet Outlet Temperature Energy Energy Transfer
Hot of Temperature Temperature Temperature Temperature Difference Lost By Gained Coefficient
Stream Cold (˚C) (˚C) (˚C) (˚C) Hot By (W/m2 k)
(lpm) Stream Stream Cold
(lpm) (Kw) Stream
(Kw)
0.5964 0.396 72 ˚C 83C 88˚C 80˚C 0.3328 0.3328
6.382 KW KW 130

4.3.1.1 Energy Gain and Loss between Cold and Hot Streams

Referring to table 4.10 above there is a difference between the heat gained by cold
stream and heat lost by hot stream which shows that there is heat loss to the
surrounding (see Fig.4.36). This difference between heat gained and loss means that

114
the insulation was insufficient and more insulating material is needed to isolate the
shell and the connecting pipes from surroundings.

Comparison Between Heat Gained by Cold Stream and Heat

lost by Hot Stream

0.8
0.7
0.6

Heat Energy(KW)
0.5
Qhot
0.4
Qcold
0.3
0.2
0.1
0
3 2 1

Fig. 4.36 Heat Gain and Loss by Cold and Hot Streams in Shell &Tube Heat Exchanger

4.3.1.2 Temperature Ranges

The test was conducted in July and August and weather was always cloudy. Due to
the almost permanent cloud cover and heat losses in piping and lower storage tank of
the evacuated collector field(see Fig. 4.26), the evacuated collector failed to attain the
designed operating conditions during the above mentioned testing period. The shell
&tube heat exchanger was supplied by a hot stream of lower temperature. That
necessitated the use of cold water of lower temperature (ambient temperature) so as to
test the ability of the heat exchanger to provide the rise in temperature required in the
cold stream.
Referring to table 4.10 (2nd and 3rd rows), it appears that the heat exchanger can
deliver the required heat capacity at the designed flow rate(approximately) only at a
minimum log mean temperature of 12.97 which is twice the designed value. This
result indicates that the temperature difference between inlet and outlet streams
should be increased accordingly. Figure 4.37 illustrate the effect of difference
between inlet temperatures of the two streams on the increase in temperature of cold
stream.

4.3.1.3 Flow Rates

The system was designed to work at relatively low flow rates so as to minimize the
quantity of hot water required and consequently the size of the evacuated collector
field . Difficulties were encountered in adjusting the control valves to deliver the

115
designed flow rate required. The solution pump (see Fig 4.25) was replaced by a
variable speed circulation pump in order to operate the system at the designed flow
rate. Due to the above mentioned difficulties, the flow rates were adjusted as close as
possible to the designed values (but not exactly equivalent to them). Fewer operational
problems could have been met, if the system was designed to operate at higher flow
rates.

Effect of Increase of Difference Between Inlet Temperatures of

Cold and Hot Streams on Increase of Cold Stream Temperature

35

30

Temperature in Celcuis
25
Difference in Inlet Temperature
of Hot and Cold Streams 20
Increase in Cold Water Inlet 15

10

0
3 2 1

Fig. 4.37 Effect of Flow Streams Inlet Temperature on Temperature Rise of Cold Stream.

4.3.1.4 Heat Transfer Coefficient

It is evident from tables, 4.10 &4.11 that the experimental heat transfer coefficient
was lower than the designed value. It is worth mentioning that the experimental
values of the heat transfer coefficient were obtained from calculations based on heat
gained by cold stream not heat lost by hot stream, which explains the low values
obtained. In addition, the designed value was based on two-phase flow in the cold
stream and not a single phase flow.

4.3.2 Double Tube Heat Exchanger

Operating parameters under investigation are:
 Energy delivered or gained by flow streams
 Temperature ranges
 Flow Rates
 Heat Transfer Coefficient
Table 4.12 &4.13 below shows the test result of the double tube heat exchanger and
the designed operating parameters respectively.

116
 4.3.2.1Energy Gain and Loss between Cold and Hot Streams
Referring to Table 4.12 below, it appears that there is a difference between heat
gained by cold water and heat lost by hot water, which necessitates the use of more
insulating material around the heat exchanger and pipe lines. It can be noticed that
heat losses are more serious than at the shell and tube heat exchanger (see Fig.4.38)

Table 4.12 Experimental Data of Double Heat Exchanger

Flow Flow Cold Water Cold Water Hot Water Hot Water Log Mean Heat Heat Heat
Rate of Rate Inlet Outlet Inlet Outlet Temperat Energy Energy Transfer
Hot of Temperature Temperature Temperature Temperature ure Lost By Hot Gained Coefficient
Water Cold (˚C) (˚C) (˚C) (˚C) Difference Stream By (W/m2 k)
(lpm) Water (Kw) Cold
(lpm) Stream
(Kw)
1.8 2.4 50 54 67 56 9.05 1.3947 0.6762 99.78
2.5 0.8 37 50 64 58 17.26 1.0566 0.7325 56.67
0.625 0.7 38 46 60 50 12.97 0.4402 0.3944 40.61
0.6 0.7 47 57 71 49 6.16 0.9298 0.493 106.76
0.55 0.6 38 55 69 41 7.14 1.0847 0.7184 134.36
0.625 0.5 38 50 61 42 6.91 0.8364 0.4226 81.57
0.65 0.41 37 59 71 42 7.99 1.3277 0.6353 106.11

Table 4.13 Designed Operating Parameters of Double Tube Heat Exchanger

Flow Flow Rate Cold Cold Hot Hot Stream Log Heat Heat Heat
Rate of of Cold Stream Stream Stream Outlet Mean Energy Energy Transfer
Hot Stream Inlet Outlet Inlet Temperature Tempera Lost By Gained By
Stream (lpm) Temperatur Temperat Temperat- (˚C) -ture Hot Cold Coefficient
(lpm) -e ure ure Differenc Stream Stream (W/m2 k)
(˚C) (˚C) (˚C) e (Kw) (Kw)
0.396 0.342 59 73 83 66 8.4 0.426KW 0.426Kw 47.6

Comparision between Heat gained by Cold Water Stream and

Heat Lost by Hot Water Stream

1.6
1.4
1.2
Heat Energy(KW)

1
Heat Aquired by cold Stream
0.8
Heat Lost by Hot Stream
0.6
0.4
0.2
0
7 6 5 4 3 2 1

Fig. 4.38 Heat Gain and Loss by Cold and Hot Streams in Double Tube Heat Exchanger

117
4.3.2.2 Temperature Ranges
Again, as in the case of shell & tube heat exchanger, the tests were conducted at off
designed operating parameters due to unfavorable weather conditions. As it is shown
in Table 4.12 above, the heat exchanger can meet the designed load at log mean
temperature difference lower than the designed log mean temperature difference. But
the experimental flow rates are slightly higher than designed flow rates. The effect of
difference of inlet stream temperatures on the increase of cold water temperature can
be clearly shown by figure 4.39.

Effect of Temperature Difference Between Inlet Temperature of

Hot and Cold Stream on the Increase of Temperature of Cold
Stream

Temperature in Degree Celcius

40
35
30
Difference between Inlet Hot and 25
Cold Water Temperature
20
Increase in Cold Water
Temperature 15
10
5
0
7 6 5 4 3 2 1

Fig. 4.39 Effect of Flow Streams Inlet Temperature on Temperature Rise of Cold Stream.

4.3.2.3 Flow Rates

Difficulties were met at trying to operate the exchanger at the designed flow rates, so
tests were conducted with flow rates as close as possible to the designed flow rates.

4.3.2.4 Heat Transfer Coefficient

As shown by table 4.12 and 4.13 the values of the experimental heat transfer
coefficient are higher than the designed values, which means that the heat exchanger
can meet greater load than the designed load.

4.3.3 Flat Plate Collector Field

Operating parameters under investigation are:
 Efficiency of the absorber plate of the flat plate collector.
 Storage Tank Temperature variation with weather condition.
Table 4.14 shows the test result of the flat plate collector.

118
 4.3.3.1 Efficiency Test of Flat Plate Collector
Referring to table 4.14 below, the average collector efficiency was found to be 65.9%
which is remarkable for a flat plate collector without selective surface. It appears also
that the instantaneous efficiency is highest when the temperature difference between
inlet temperature and ambient temperature is the lowest. This result is in agreement
with that reported in literature.
Table 4.14 Efficiency Test of Flat Plate Collector
Ambient Radiation Inlet Outlet Average
efficienc
Temperat Intensity Tempera Tempe Ti-Ta Efficienc
y
ure I ture rature y
2 Heat Gain(W) Q
Ta (W/m ) Ti To 
(˚C) (˚C) (˚C) Qm
 c (T0  Ti ) AI

33 709 40 50 1407.258 0.94194 7

37 747 47 53 844.3548 0.565164 10 0.659
37 762 54 59 703.629 0.47097 17

4.3.3.2 Variation of Storage Tank Temperature with Ambient Conditions

(i) Flat Plate Collector with 110 liter Storage
The designed operating parameters and the experimental data (summarized from
Table 4.7) are shown in table 4.15.

Table 4.15 Comparison between Design and Experimental Data for Flat Plate Collector with 110
Liter Capacity
No. Parameters Designed Experimental Values
Values
1 Quantity of liquid to be preheated 65 liter 110 liter
2 Temperature to be attained prior to operation 59˚C 58* up to70

*Reading obtained at very cloudy weather condition

Table 4.15 shows the flat plate collector can preheat almost twice the quantity of fluid
to temperatures greater than the designed temperatures by up to 10 degrees and it can
fulfill the preheating duty even at the worst weather conditions. At clear sky
conditions the storage tank temperature can reach 80 ˚C at 3 O'clock p.m.

(ii) Flat Plate Collector with 36 liter Capacity

The purpose of the flat plate collector was to preheat the strong ammonia solution to
temperature higher than that of the flat plate collector with 110 liter capacity to
provide the system with preheated fluid for the first hour of operation till operation
stability was attained. The performance of the collector was observed from December
till March and it was evident that the storage tank temperature could reach 100 ˚C at
3O'clock p.m. Table 4.16 compare the designed parameter with the experimental data.
119
It shows that the collector can preheat more than double the heat transfer fluid to
temperatures higher than up to 30 degrees.

Table 4.16 Comparison between Design and Experimental Data for Flat Plate Collector with 36
Liter Capacity

No. Parameters Designed Experimental Values

Values
1 Quantity of liquid to be preheated 15 liter 36 liter
2 Temperature to be attained prior to operation i.e. 67˚C 80 up to 96 (from
one O'clock local time December 2008 to
March2009)

4.3.4 Evacuated Collector Field

Operating parameters under investigation are:
 Variation of storage tank temperature under constant load withdrawal.
 Attaining operating temperature under different weather conditions.

4.3.4.1 Effect of Constant Load Withdrawal On Storage Tank

Temperature Variation.
Table 4.3.4.1. compares the designed parameters to the experimental data. As shown
by the table, there is a slight difference between the experimental and designed
delivery temperature but there is a big difference between the experimental and
designed temperature. This difference is attributed to the following:
 Heat losses in pipe lines and upper &lower storage tanks (see Fig. 4.26)
 Difficulties in controlling the heat exchange process between the hot and cold
streams at the shell and tube heat exchanger which was used as the load of the
experiment.
 Presence of cloud covers during the testing period prevented the collector
field from receiving adequate radiation to meet the load

4.3.4.2 Variation of Storage Tank Temperature under Different Weather

Conditions
As shown by Fig. 3.16, failure of the collector field to meet the load was predicted to
occur at 3 and 4 O'clock from April till September. Although observations of the
collector field temperature showed that it can attain temperatures higher than the
designed delivery temperature during period from Decemeber2008 to March 2009,

120
the storage tank temperature observation during July and August showed that the
collector field would not attain the operating temperature at time of system
operation(see Table 4.8)

Table 4.17 Comparison of Experimental Data with Design Parameters of Effect of Constant
Load withdrawal on the Evacuated Collector Field

Parameters Design
Time Experimental Data Parameters
Hot Water Delivery (p.m.)
Temperature 1:00 88 88
2:00 87 88
3:00 86 88
4:00 84 88
Time
1:00 45 80
Hot Water Return 2:00 55 80
Temperature 3:00 48 80
4:00 49 80
Mass Flow Rate of Hot Water(lpm) 0.55 0.59

121
 Chapter 5
5. Conclusion and Recommendation
5.1 General Conclusion and Recommendation
Within the course of this study, a continuous absorption refrigeration machine has
been designed and constructed. Performance tests were conducted on the components
of the experimental unit. In view of the previous result analysis, the following can be
concluded:
 The performance of the shell and tube heat exchanger and double tube heat
exchanger could be improved, if greater log mean temperature difference and
higher flow rates could be used.
 The temperature of the hot water delivered by the evacuated collector field
need to be increased by at least 10 to 15 ˚C. This rise in temperature is
necessary to compensate for pipe line losses and unfavorable weather
conditions. In addition, it will increase the value of the log mean temperature
difference of the heat exchangers. Since the water in glass collector operating
temperature is below this temperature range, a different type of evacuated
collector should be inserted between the evacuated collector and the shell&
tube heat exchanger. The duty of this suggested evacuated collector is to
provide the rise in temperature necessary to improve system performance.
 The performance of the flat plate collector field was excellent and obviously
it can accomplish the preheating duty even under unfavorable weather
condition.

5.2 Further Work

 Testing of heat exchangers using higher flow rates and different temperature
ranges to define the maximum power capacity of the system.
 Subjecting the evacuated collector field to withdrawal different mass flow
rates to determine the variation of storage tank temperatures under different
load patterns.

122
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123
124
 Appendix
Thermophysical Properties of Saturated Water
Refrigerant 717 (Ammonia) Properties of Liquids and Saturated Vapor
Specific Volume of Saturated Ammonia Solution

Specific Volume of Saturated Ammonia Solution

Viscosities and Densities of Liquids
Latent Heats of Vaporization
Specific Heats at Constant Pressure of Gases and Vapors at (at 101.kN/m2)
Thermal Conductivities of Gases and Liquids