DEVELOPING A PROTOTYPE FOR SOLAR
OPERATED CHILLER
By
RAGAB KASSEM MAHMOUD ALI
1
B. Sc. - Agric.“Agric. Engineering”, Fac. Agric., Al-Azhar
University, 2005.
THESIS
Submitted in Partial Fulfillment of the
Requirements for the Degree
Of
MASTER OF SCIENCE
In
AGRICULTURE
(Agricultural Engineering)
Agricultural Engineering Department
Faculty of Agriculture
Al-Azhar University
Assiut Branch
1434 A.H.
2013 A.D.
i
PPROVAL SHEET
NAME: RAGAB KASSEM MAHMOUD ALI
TITLE: DEVELOPING A PROTOTYPE FOR
SOLAR OPERATED CHILLER
THESIS
Submitted in Partial Fulfillment of the
Requirements for the Degree
Of
MASTER OF SCIENCE
In
AGRICULTURAL SCIENCES
(Agricultural Engineering)
Agricultural Engineering Department
Faculty of Agriculture
Al-Azhar University
Assiut Branch
1434 A.H.
2013 A.D.
Approved By:
Prof. Dr. MOHAMED NABIL EL AWADY…………...
Prof. Emerit. of Ag. Eng., Fac. of Ag., Ain Shams U.
Prof. Emt. Dr. SAMIR AHMED TAYEL……………...
Prof. Emerit. of Ag. Eng., Fac., of Ag. Eng., Al-Azhar U., Cairo.
Prof. Dr. AHMED M. EL- LITHY …………………….
Prof. of Ag. Eng., Fac. of Ag., Al-Azhar U., Assiut.
Dr. MAHMOUD ZAKKY .EL- ATTAR………………
Lec. of Ag. Eng., Fac. of Ag., Ain Shams U.
Date: 22/ 9/ 2013 A.D.
ii
2
TITLE: DEVELOPING A PROTOTYPE FOR
SOLAR OPERATED CHILLER
NAME: RAGAB KASSEM MAHMOUD ALI
3
THESIS
Submitted in Partial Fulfillment of the
Requirements for the degree
Of
MASTER OF SCIENCE
In
AGRICULTURE
(Agricultural Engineering)
Agricultural Engineering Department
Faculty of Agriculture
Al-Azhar University
Assiut Branch
1434 A.H.
2013 A.D.
Supervisory Committee:
Prof. Dr. HASSAN A. EL RAZIK A. MAWLA…………...
Prof. and Head of Ag. Eng. Dpt., Fac. of Ag., Al-Azhar U., Assiut.
Prof. Dr. AHMED M. EL- LITHY ………………………...
Prof. of Ag. Eng., Fac. of Ag., Al-Azhar U., Assiut.
Dr. MAHMOUD ZAKKY .EL- ATTAR……………………
Lec. of Ag. Eng., Fac. of Ag., Ain Shams U.
iii
2.1 Importance of post-harvest treatments ...................... 3
2.2 Postharvest agricultural products deterioration .......... 3
2.3 Classification of absorption refrigeration cycles ......... 5
2.3.1 Single-effect absorption system. ........................ 5
2.3.2 Multi-effect absorption refrigeration cycle. ........... 6
2.3.3 The ammonia-water hydrogen cycle ................... 7
2.3.4 Adsorption refrigeration cycle. .......................... 8
5.3.2 Desiccant refrigeration system. ........................ 10
2.4 Chilling solar assissted systems ............................. 11
2.4.1 Vapor compression refrigeration systems. ........... 11
2.4.2 Chiller solar assisted coefficient of performance ... 14
2.5 Chillers solar assissted systems ............................. 21
3.1 Principle and system description ........................... 24
3.2 System configuration: ......................................... 25
3.2.1 Generator: solar flat-plate collector FPC ............ 25
3.2.2 Generator: vessel .......................................... 32
3.2.3 Chiller condenser- Evaporator.......................... 36
iv
3.3 Chilling- solar assisted system overall efficiency ....... 38
3.4 Thermal load and chilling system .......................... 38
3.5 Experimental setup ............................................ 40
4.1 Generator performance ....................................... 44
4.1.1 Generator: FPC thermal performance ................. 44
4.1.2 Generator: vessel thermal performance ............... 49
4.2 Condeser thermal performance.............................. 49
4.3 Evaporatore thermal performance and COP ............. 50
v
LIST OF FIGURES
FIGURE 2-1: THE EINSTEIN REFRIGERATION CYCLE. .............................................................. 4
FIGURE 2-2: A SINGLE-EFFECT LIBR/WATER ABSORPTION REFRIGERATION SYSTEM. .......................... 5
FIGURE 2-3: A DOUBLE-EFFECT WATER/LIBR ABSORPTION CYCLE. ............................................. 7
FIGURE 2-4: SIMPLE DIAGRAM OF THE ABSORPTION REFRIGERATION SYSTEM. ................................ 8
FIGURE 2-5: (A) THE DESORPTION (REGENERATION) PROCESS AND (B) THE ADSORPTION (REFRIGERATION)
PROCESS. .................................................................................................. 9
FIGURE 2-6: DESICCANT COOLING SYSTEM. ..................................................................... 11
FIGURE 2-7: VAPOR COMPRESSION CYCLE....................................................................... 12
FIGURE 2-8: RANKINE REFRIGERATION SYSTEM. ................................................................ 13
FIGURE 2-9: A DOUBLE-EFFECT ABSORPTION CYCLE OPERATES WITH TWO PRESSURE LEVELS. ............... 20
FIGURE 3-1: BASIC VAPOR ABSORPTION REFRIGERATION CYCLE. .............................................. 25
FIGURE 3-2: FLAT-PLATE SOLAR COLLECTOR “FPC”. ........................................................... 26
FIGURE 3-3: FPC ABSORBER PLATE COMPONENTS AND OVERALL DIMENSIONS. .............................. 26
FIGURE 3-4: HEAT TRANSFER THROUGH LAYERS OF AIR, WOOD, GLASS WOOL AND ABSORBER PLATE. ..... 28
FIGURE 3-5: FPC CROSS SECTION OF ENERGY ABSORPTION PLATE AND ITS TWO TRANSPARENT
POLYETHYLENE COVERS. ................................................................................. 30
FIGURE 3-6: FPC ABSORBER PLATE AND PLASTIC DUAL COVERS. .............................................. 31
FIGURE 3-7: SOLAR INTENSITY MEASURENING PYRANOMETER MODEL PSP. ................................. 31
FIGURE 3-8: CUP-COUNTER TYPE ANEMOMETER. .............................................................. 32
FIGURE 3-9: GENERATOR STORAGE VESSEL ..................................................................... 33
FIGURE 3-10: GENERATOR STORAGE VESSEL DIMENSIONS. .................................................... 34
FIGURE 3-11: HEAT TRANSFER THROUGH LAYER OF GLASS WOOL, AND GENERATOR VESSEL SURFACE. ..... 35
FIGURE 3-12: CONDENSER –EVAPORATOR DIMENSIONS IN CM. .............................................. 37
FIGURE 3-13: THERMAL LOAD AND CHILLING SYSTEM CALCULATIONS......................................... 39
FIGURE 3-14: CONFIGURATION OF THE EXPERIMENTAL SETUP................................................. 41
FIGURE 3-15: THE SOLAR REFRIGERATOR SYSTEM:1) FLAT PLATE SOLAR COLLECTOR, 2) GENERATOR VESSEL,
3) PRESSURE GAUGES, 4) EXPANSION VALVE. 5) EVAPORATOR-CONDENSER(IMMERSED IN A TANK OF
WATER) 6) EVAPORATOR- CONDENSER (DRY) .......................................................... 42
FIGURE 3-17: THE SOLAR REFRIGERATION COMPONENT. ...................................................... 43
FIGURE 3-18: OVERVIEW OF THE SOLAR ASSISTED CHILLING SYSTEM. ......................................... 43
FIGURE 4-1: THE RELATIONSHIP BETWEEN AIR VELOCITY AND HEAT LOSSES................................... 44
FIGURE 4-2: FPC OPTICAL AND THERMAL EffiCIENCIES. ........................................................ 46
FIGURE 4-3: DALY HEAT TRANSFEAR FROM THE FPC TO THE WORKING FLUID IN CHILLING SYSTEM. ........ 47
FIGURE 4-4: THE RELATIONSHIP BETWEEN THE FPC ABOSRBER PLATE TEMPERATURE AND AMMONIA-WATER
TEMPERATURE. ........................................................................................... 47
FIGURE 4-5:THE RELATIONSHIP BETWEEN THE FPC ABSORBER PLATE TEMPERATURE AND THE FPC
EFFICIENCY. ............................................................................................... 48
vi
FIGURE 4-6: HEAT REMOVAL RATE IN RELATION TO THE FPC ABSORBER PLATE TEMPRATURE DURING THE
DAY........................................................................................................ 48
FIGURE 4-7: HEAT TRANSFER FROM THE GENERATOR FPC TO VESSEL IN RELATION TO TEMPRATURE
DIFFERENCE. .............................................................................................. 49
FIGURE 4-8: HEAT TRANSFER FROM GENERATOR VESSEL THROUGH CONDENSER-EVAPORATOR. ........... 50
FIGURE 4-9 :HEAT TRANSFER THROUGH CHILLING SYSTEM AT SOLAR RADIATION INTENSITIES. .............. 51
FIGURE 4-10: HEAT TRANSFER FLOW FROM FPC ABSORBER PLATE THROUGH CHILLING SYSTEM AT
DIFFERENT SYSTEM PRESSURE. .......................................................................... 51
FIGURE 4-11:CHILLING SOLAR ASSISETED SYSTEM EFFECIENCY AT SOLAR INTINSTIES DURING THE DAY TIME.
............................................................................................................ 52
FIGURE 4-12: CHILLING SOLAR ASSISETED SYSTEM ENERGY FLOW AT SOLAR INTENSITIES DURING THE DAY
TIME....................................................................................................... 52
FIGURE 4-13: ENERGY TRANSFER THROUGH GENERATOR FPC, GENERATOR VESSEL, AND CONDENSER. .... 53
FIGURE 4-14: THE COEFFICIENT OF PERFORMANCE FOR THE CHILLING SOLAR ASSISTED SYSTEM. ........... 53
FIGURE 4-15: REQUIRED TIME FOR HEAT REMOVAL FROM THE EVAPORATOR WITHOUT LOAD “EMPTY” AND
SYSTEM PRESSURE. ....................................................................................... 54
FIGURE 4-16: CALCULATED POTATOES HEAT LOAD AT DIFFERENT AMBIENT TEMPERATURE. ................ 55
FIGURE 4-17: POTATOES THERMAL LOAD ENERGY AND HEAT REMOVAL RESPONSE. ......................... 55
FIGURE 4-18: CALCULATED HEAT LOAD REMOVAL FROM KILOGRAM OF POTATOES AND ENERGY TRANSFER
FROM CONDENSING TO EVAPORATING PROCESSES. .................................................... 56
FIGURE 4-19: CHILLING SOLAR ASSISTED POTATOES HOLDING CAPACITIES AT DIFFERENT CHILLING
TEMPERATURE LEVELS.................................................................................... 56
FIGURE 4-20: REQUIRED TIME FOR HEAT REMOVAL FROM POTATOES AT DIFFERENT TEMPRATURE LEVELS. 57
FIGURE 7-1: OBSERVATION DURING REFRIGERATION TEST ON JUNE 25, 2011.............................. 78
FIGURE 7-2: ACTUAL AND THEORETICAL SYSTEM CYCLES FOR TEST ON JUNE 25, 2011. .................... 83
FIGURE 7-3.A: OBSERVATION DURING GENERATION TEST ON JUNE 26, 2011 .............................. 78
vii
1 INTRODUCTION
Postharvest treatments are crucial in maintaining agricultural
production quantity and quality. Egypt production of fruits and
vegetables were estimated by 11,750,975 and 21,236,320 billion
of Egyptian pounds annually, as stated in the Egyptian statistical
survey published in 2010-2011. Due to lack or absence of
postharvest treatments, experts evaluate 15.27% of fruit and
45.58% of vegetables drop in marketing value.
Kurian (2012) refers 30-40% of the total loss of fresh food
production to the inadequate of postharvest treatments and
storage conditions.
Eckert and Ogawa (1985) reviewed factors contribute to
postharvest losses in fresh fruits and vegetables. They concluded
that environmental conditions such as heat or drought,
mechanical damage during harvesting and handling, improper
postharvest sanitation, and poor cooling and environmental
control were the most affecting factors in food deterioration.
Harvested fresh vegetables and fruits undergo respiration,
which involves enzymatic oxidation of sugar in the produce into
CO2 and water accompanied by the release of energy. This causes
overspending, deterioration and ultimate destruction of the fruit
since low temperature retards the activity of enzymes, the
commonly accepted measure is to remove the field heat from the
product within a short period of time after harvesting (Tomkins
and Woreko-Brobby, 1982).
Energy supply to refrigeration and chilling systems constitutes
a significant role in postharvest treatments in remote areas with
poor energy distribution network. Utilizing solar energy in
1
postharvest treatments will reduced the total process costs and
production quality and quantity losses.
Application of solar assisted cooling saves electricity and thus
conventional primary energy sources. It is also leads to a
reduction of peak electricity demand, and additional cost savings.
Using solar energy technologies is environmentally sound
materials that have no ozone depletion and no (or very small)
global warming potential (IEA, 2011).
The aim of this study is to develop a prototype for dual purpose
solar operated chilling and heating postharvest treatment for
agricultural applications.
2
2 REVIEW OF LITERATURE
2.1 Importance of post-harvest treatments
Terms of the amount of production available to the consumer
of vegetables about 14970000 tons/year with the rate loss of the
estimated 3044000 tons/year and the amount of production of
fruit 6014000 tons/year with the rate loss by 868,000 tons/year
and produce citrus 2,656,000 tons/year rate loss was 371,000
tons/year, As well as losses of starches estimated at around
636000 tons/year of total production for the consumer 3,861,000
tons/year (Statistics 2010).
Kader (2002) pointed that quality loss can be minimized by
using best harvest procedures, rapid cooling, refrigerated storage
and proper handling techniques during transportation and
distribution to market.
2.2 Postharvest agricultural products deterioration
Environmental conditions, mainly temperature, affect the
quality of the fresh horticultural produce. High temperature
increases produce respiration rate and water loss through
transpiration, causing loss in internal fresh quality, shriveling
and premature softening (Tanner and Smale, 2005).
Jorge (2006) stated that removing field heat can suppress
enzymatic degradation (softening) and respiratory activity; slow
down or inhibit water loss (wilting); slow down or inhibit the
growth of decay-producing microorganisms (molds and
bacteria); reduce the production of ethylene as a ripening agent.
Widely encountered in air, water, soil, living organisms, and
unprocessed food items. Microorganisms cause off-flavors and
3
odors, slime production, changes in the texture and appearances,
and the eventual spoilage of foods. Holding perishable foods at
warm temperatures is the primary cause of spoilage, and the
prevention of food spoilage and the premature degradation of
quality due to microorganisms is the largest application area of
refrigeration.
The Platen and Munters (1928) described a single pressure
cycle utilizes ammonia for the refrigerant and water for the
absorbent. The water separates the ammonia from the inert gas,
hydrogen. A recently uncovered U.S. patent by the famed Albert
Einstein (figure, 2-1) and Leo Szilard issued on 1930 discloses
another single pressure thermally driven refrigeration cycle
which uses butane, ammonia, and water. In the Einstein cycle
ammonia acts as an inert gas to lower the partial pressure over
the refrigerant, butane, and water later provides separation by
absorbing the ammonia.
Figure 2-1: The Einstein refrigeration cycle.
4
2.3 Classification of absorption refrigeration cycles
2.3.1 Single-effect absorption system.
A single-effect absorption refrigeration system is the simplest
and most commonly used design. There are two design
configurations depending on the working fluids used.
Figure (2-2) shows a single-effect system using non-volatility
absorbent such as LiBr/water.
Generator
Condenser
HX
Evaporator
Absorber
Figure 2-2: A single-effect LiBr/water absorption
refrigeration system.
High temperature heat supplied to the generator is used to
evaporate refrigerant out from the solution and is used to heat
the solution from the absorber temperature. Thus, is caused as
high temperature heat at the generator is wasted out at the
absorber and the condenser.
When volatility absorbent such as water/NH3 is used, the
system requires an extra component called “a rectifier”, which
will purify the refrigerant before entering the condenser. As the
absorbent used (water) is highly volatile, it will be evaporated
together with ammonia (refrigerant). Without the rectifier, this
5
water will be condensed and accumulate inside the evaporator,
causing the performance to drop (Aphornratana, 1995)..
2.3.2 Multi-effect absorption refrigeration cycle.
The main objective of a higher effect cycle is to increase
system performance when high temperature heat source is
available. By the term “multi-effect”, the cycle has to be
configured in a way that heat rejected from a high-temperature
stage is used as heat input in a low-temperature stage for
generation of additional cooling effect in the low-temperature
stage. Double-effect absorption refrigeration cycle was
introduced during 1956 and 1958 (Vliet, Lawson and Lithgow,
1982).
Figure (2-3) shows a system using LiBr/water. High
temperature heat from an external source supplies to the firsteffect generator. The vapor refrigerant generated is condensed at
high pressure in the second-effect generator. The heat rejected is
used to produce addition refrigerant vapor from the solution
coming from the first-effect generator. This system
configuration is considered as a series-flow-double-effect
absorption system.
If LiBr/water is replaced with water/NH3, maximum pressure
in the first-effect generator will be extremely high. Figure (2-2)
shows a double-effect absorption system.
6
Generator I
HX II
Generator II
Condenser
HX I
Absorber
Evaporator
Figure 2-3: A double-effect water/LiBr absorption cycle.
2.3.3 The ammonia-water hydrogen cycle
Watheq (2008) described the main components of the
absorption refrigeration system as an absorber, generator, a
condenser, an expansion valve, a heat exchanger and a pump. As
shown in figure (2-4), Watheq mentioned two kinds of working
medium used at the same time in refrigeration and absorption
processes.
The refrigerant vapor flows to the condenser passing through
a vapor-trap and condensed. Liquid refrigerant from the
condenser goes through an expansion valve while the pressure
is decreased to an evaporation pressure. At the evaporator,
cooling effect is achieved by the vaporization of the refrigerant
7
at a low temperature. Refrigerant vapor from the evaporator
continues to an absorber and dissolves in a weak refrigerant
solution, and it becomes a stronger refrigerant solution, which is
called “rich solution”. A pump is the only moving part in this
system.
Condenser
Generator
Expansion valve
pump
Evaporator
Absorber
Figure 2-4: Simple diagram of the absorption refrigeration
system.
2.3.4 Adsorption refrigeration cycle.
Watheq (2008) defined an adsorption is a preferential
partitioning of substances from a gaseous or liquid phase onto a
surface of a solid substrate. This process requires only thermal
energy. The principles of the adsorption process provide two
main processes, adsorption or refrigeration and desorption or
regeneration. In case zeolite and water, as an example, the
refrigerant (water) is vaporized by, the heat from cooling space
and the generator (absorbent tank) is cooled by ambient air. The
8
vapor from the cooling space is leaded to the generator tank and
absorbed by adsorbent (zeolite).
a
b
Figure 2-5: (a) The desorption (regeneration) process and
(b) the adsorption (refrigeration) process.
The rest of the water is cooled or frozen. In the regeneration
process, the zeolite is heated at a high temperature until the water
vapor in the zeolite is desorbed out, goes back and condenses in
the water tank, which is now acting as the condenser. For a
discontinuous process (figure, 2-5:a), desorption process can be
operated during daytime by solar energy, and the adsorption or
the refrigeration process can be operated during night-time
(figure, 2-5:b). The solar energy can be integrated with a
generator. The single adsorber is required for a basic cycle. The
number of adsorbers can be increased to enhance the efficiency,
which depends on the cycle. This process can also be adapted to
the continuous process.
9
2.3.5 Desiccant refrigeration system.
Daou et al. (2006) stated that as natural or synthetic
substances capable of absorbing or adsorbing water vapor due
the difference of water vapor pressure between the surrounding
air and the desiccant surface. Thus, in desiccant cooling systems
desiccants are used to dehumidify the inlet air and then this dry
air is cooled and humidified by evaporative cooling and
followed sometimes by a vapor compression system for sensible
cooling. Since desiccant cooling is utilized to handle latent
loads, the vapor compression system’s energy demand
decreases. Although the desiccant cooling system’s performance
is strongly dependent on weather conditions, energy savings
may reach up to 80% for dry climates. Energy savings decrease
as humidity ratio increases. A schematic of a desiccant cooling
system is presented in figure (2-6). The outside air stream at state
1 is passed through the desiccant wheel. Humidity of the air
stream at state 2 is significantly decreased. Since adsorption or
absorption of water vapor by the desiccant substance is an
exothermic reaction, temperature of the air stream increases.
Then by heat wheel the temperature of the air stream is
decreased. If heat wheel is not enough for required cooling,
vapor compression system should be used in conjunction with
heat wheel. Finally, the temperature of the air stream is further
decreased and humidity ratio of the air is increased by
evaporative cooling according to thermal comfort conditions. A
heater is used between state 7 and 8 in order to obtain high
temperatures required for regeneration of the desiccant material.
Solar energy or waste heat can be used for heating the exhaust
air stream. Assuming no need for a vapor compression system.
10
A small amount of electricity is required for rotating the
wheels. The desiccant materials for a solid-desiccant system are
usually silica gel or Zeolite. For a liquid desiccant system, the
desiccant dehumidifier’s hygroscopic aqueous solution can be
tri-ethylene glycol (TEG), CaCl2-H2O, LiBr-H2O, LiCl-H2O.
Figure 2-6: Desiccant cooling system.
2.4 Chilling solar assissted systems
2.4.1 Vapor compression refrigeration systems.
As seen from figure (2-7), the ideal vapor compression cycle
consists of four processes: isentropic compression in a
compressor; constant pressure heat rejection in a condenser;
throttling in an expansion valve; constant pressure heat
absorption in an evaporator.
11
Figure 2-7: Vapor compression cycle.
Wang (2000) summed up an adsorption ice-making system
driven by generation temperatures from 90 to 100 oC with
activated carbon–methanol as working pair. This system can
reach an evaporation temperature as low as 15.5 o C.
Assilzadeh et al. (2005) presented a H2O–LiBr absorption
unit using evacuated tube solar collectors. After the modeling
and simulation carried out with TRNSYS program, the author
concluded that the optimum system for Malaysia’s climate for a
3.5 kW system consists of 35 m2 evacuated tubes solar collector
sloped at 20 oC.
Wimolsiri Pridasawas (2006) stated that the Rankine
refrigeration system is the combination of the Rankine power
cycle and vapor compression refrigeration cycle. The turbine
work obtained from the Rankine cycle is used to drive the
compressor in the vapor compression cycle. A schematic of a
12
Rankine refrigeration cycle is given in figure (2-6) Rankine
refrigeration system’s COP is same as vapor compression cycle,
but the efficiency of Rankine power cycle is directly related to
the temperatures of the sink and source.
In order to increase the overall system’s performance, high
efficiency flat plate collectors, evacuated tube collectors or
parabolic trough collector can be used. fluids such as R114 that
give a positive slope of the saturated vapor line on a T-S
diagram, the outlet temperature from the turbine is significantly
higher than the condensation temperature gives the benefit to
preheat the working fluid before it enters the boiler.
Electricity
Production
Condenser
Boiler
pump
Expansion
valve
Turbine
Alienat
or
compressor
Condenser
Evaporator
Figure 2-8: Rankine refrigeration system.
13
In general Rankine refrigeration systems are complex and
suitable for large air conditioning applications system. Working
However R114 is not environmental friendly; it has an ozone
depleting potential due to a Chlorine atom the speed of the
turbine and the compressor should be analogous. The alternator
or other equipment that used to adjust the speed should be
installed with the system.
Abdellatif et al. (2009) the primary objectives of solar
heating system are to increase the solar radiation converted into
stored thermal energy and to investigate effective uses of that
stored energy. A solar collector which is continuously orientated
and tilted to maintain an incident solar angle of zero from sunrise
to sunset will allow maximum values of both; the effective
absorptance of the absorber surface and the effective
transmittance of the glass cover to be reached.
2.4.2 Chiller solar assisted coefficient of performance
Sierra et al. (1993) used a solar pond to power an intermittent
absorption refrigerator with NH3–H2O solution. It is reported
that generation temperatures as high as 73 o C and evaporation
temperatures as low as –2 oC could be obtained. The thermal
COP working under such conditions was in the range of 0.24–
.28.
Critop (1993) also built a laboratory scale activated carbon–
ammonia refrigerator. The generator, of exposed surface area
1.4 m2, consisted of an array of 15 stainless steel tubes, each of
2 m length, 42 mm outside diameter and 1.1 mm thick, rated to
30 bar pressure. About 17 kg of 208 oC activated carbon granules
were packed in the tubes. The condenser was a 4 m length of
14
12.5 mm diameter stainless steel tube coiled within a 100-liter
water tank. The evaporator was a 10 mm diameter stainless steel
coil immersed in 4 liters of water. The evaporator temperature
attained was up to - 1 oC and about 3 kg of ice was manufactured.
The peak collector temperature for the simulated day tests was
115 o C, and the solar COP was 0.04.
Critoph (1994) built a small solid adsorption solar
refrigerator in 1994. The collector was 1.4 m2 in area and
contains 17 kg of active carbon. It was possible to produce up to
4 kg of ice per day in a diurnal cycle.
Critoph(1996) studied a rapid cycling solar /biomasspowered adsorption refrigeration system with activated carbon
–ammonia as working pair. The thermal COP was about 0.3
when the initial generator temperature was about 50 o C and
evaporating temperature was about 0 oC.
Bansal et al. (1997) reported a unit of 1.5 kWh/day using
NH3 as refrigerant and IMPEX material (80% SrCl2 and 20%
Graphite) as absorbent. Theoretical maximum overall COP of
the unit is 0.143, and it depends upon the climatic conditions.
Oertel and Fischer (1998) modified a commercially
available low temperature (80–90oC) adsorption cooling system
for air conditioning application, using methanol/silica gel as
working pair for the cold storage of agricultural products at
temperatures of 2–4 oC in India. Calculation and test results
showed that the COP was about 0.30 when operating the system
at a chilled water temperature of -2 oC, a heating water
temperature of 85 oC and a condenser temperature of 30 oC.
15
Sumathy and Li (1999) operated a solar-powered ice-maker
with the solid adsorption pair of activated carbon and methanol,
using a flat-plate collector with an exposed area of
0.92 m2. This system could produce ice of about 4–5 kg/day with
a solar COP of about 0.1–0. 12.
Hammad and Habali (2000) designed a solar-powered
absorption refrigeration cycle using NH3–H2O solution to cool a
vaccine cabinet in the Middle East. A year round simulation
indicated that thermal COP ranged between 0.5 and 0.65 with
generation temperature at 100–120 oC and the cabinet inside
temperature at 0–8 oC.
Li et al. (2001) built a flat-plate solid-adsorption refrigeration
ice maker with activated carbon methanol as working pair for
demonstration purposes. The experimental results show that the
thermal COP is about 0.45 and solar COP is about 0.12–.14,
with approximately 5–6 kg of ice produced per m2 collector.
Sumathy et al. (2002) developed a new model of two-stage
H2O–LiBr absorption chiller. Test results have proved that the
two-stage chiller could be driven by low temperature hot water
ranging from 60 to 75 oC, which can be easily provided by
conventional solar hot water systems. Compared to the singlestage chiller, the two-stage chiller could achieve roughly the
same total COP as of the conventional system with a cost
reduction of about 50%.
De Francisco et al. (2002) developed and tested a prototype
of 2 kW NH3–H2O absorption system in Madrid for solar powered refrigeration in small rural operations. The test results
16
showed unsatisfactory operation of the equipment with COP
lower than 0.05.
Anyanwu and Ezekwe (2003) designed, constructed and
tested a solid adsorption solar refrigerator using activated
carbon-methanol as the working pair. Its flat-plate type
collector/generator/adsorber used clear plane glass sheet whose
effective exposed area was 1.2 m2 with the efficiencies of
11.6–16.4%. The steel condenser tube with a square plan view
was immersed in pool of stagnant water contained in a
reinforced sand tank. The evaporator is a spirally coiled copper
tube immersed in stagnant water. Ambient temperatures during
the adsorbate generation and adsorption process varied over
18.5–3 4 oC. The refrigerator yielded evaporator temperatures
ranging over 1.0–8.5 oC from water initially in the temperature
range 24–28 oC. Accordingly, the maximum daily useful cooling
produced was 266.8 kJ/m2 of collector area and the use full cycle
and the useful overall COP ranged over 0.056– 0.093 and 0.007–
0.015, respectively.
Khattab (2004) developed a solar-powered adsorption
refrigeration module with the solid adsorption pair of local
domestic type charcoal and methanol. The module consists of a
modified glass tube having a generator (sorption bed) a t one
end, a combined evaporator and condenser at the other end and
simple arrange men t o f plane reflectors to heat the generator.
Test results show that, the daily ice production is 6.9 and 9.4
kg/m2 and net solar COP is 0.136 and 0.159 for cold and hot
climate respectively.
17
Hildbrand e t al. (2004) the adsorption pair is silica gel–
water. Cylindrical tubes function as both the adsorber system
and the solar collector (flat-plate, 2 m2 double glazed); the
condenser is air-cooled (natural convection) and the evaporator
contains 40 Liters of water that can freeze. This ice functions as
a cold storage for the cabinet. This system has presented
performances with a solar COP of 0.16.
Li et al. (2004) developed a no valve solar icemaker. For this
system, there are no any reservoirs, connecting valves or
throttling valve. Experimental results showed that 6.0–7. 0 kg
ice can be obtained under indoor conditions when radiation
energy was about 17–20 MJ/m2. For these conditions, the solar
COP of this system was about 0.13–.15. In outdoor conditions,
the system could produce 4.0 kg ice and the solar COP was about
0.12 when the total insolation energy was about 16–18 M J/m2.
And then, a new solar ice maker developed can produce about
ice of 4–5 kg each sunny day under the condition of about 18–
22 MJ/m2 solar insolation.
Syeda et al. (2005) studied solar cooling system for typical
Spanish houses in Madrid. The system consisted of a flat -plate
collector array with a surface area of 49.9 m2, a 35 kW nominal
cooling capacity single-effect (H2O–LiBr) absorption chiller.
This machine operated within the generation and absorption
temperature ranges of 57–67 oC and 32–36 oC, respectively. The
measured maximum instantaneous, daily average and period
average COP were 0.60 (at maximum capacity), 0.42 and .3
respectively.
18
Lemmini and Errougani (2005) built and tested a solarpowered adsorption refrigerator using the pair AC35–methanol
(figure, 2-10). The system consists of a flat-plate collector, a
condenser and a cold chamber-evaporator. Experimental results
showed that the unit can produce cold air even for rainy and
cloudy days and the solar coefficient of performance (COP)
ranges between 0.05 and 0.08 for an irradiation between 12,000
and 27,000 kJ/m2, a daily mean ambient temperature between 14
and 18 oC and lowest temperature achieved by the evaporator
between 5 and 8 oC.
Using water/NH3. In contrast to the system for LiBr/water,
this system can be considered as a combination of two separated
single-effect cycles. The evaporator and the condensers of both
cycles are integrated together as a single unit.
19
Condenser
Evaporator
Rectifie
Generator I
HX I
Absorber I
Rectifi
Heat of
Absorption
Generator II
HX II
Absorber II
Figure 2-9: A double-effect absorption cycle operates with
two pressure levels.
20
Thus, there are only two pressures level in this system and
the maximum pressure can be limited to an acceptable level.
Heat from external source supplies to generator II only. As water
is an absorbent, there is no problem of crystallization in the
absorber.
Hence, absorber II can be operated at high temperature and
rejects heat to the generator I. This system configuration is
considered as a parallel-flow-double-effect absorption system
(Kaushik and Chandra 1985).
Yeung et al. (1992) designed and constructed a solarpowered absorption air-conditioning system to study the
feasibility of utilizing solar power for comfort cooling in Hong
Kong. The system consisted of a flat-plate collector array with a
surface area of 38.2 m, a 4.7 kW nominal cooling capacity H2O–
LiBr absorption chiller, a 2.75 m3 of hot-water storage tank, a
cooling tower, a fan-coil unit, and an electrical auxiliary heated.
I t had an annual system efficiency of 7.8% and an average solar
fraction of 55%.
2.5 Chillers solar assissted systems
Critoph and gong. (1992) In this version, two separate
adsorption cycles are operated out of phase such that when one
adsorber is being heated by the energy source, the other cools to
ambient temperature, reabsorbing its refrigerant and producing
useful cooling in the evaporator. The laboratory prototype rapid
cycling ice maker consisted of two adsorbers: each consisting of
seven 2 m long stainless steel tubes, with 1.04 kg of activated
carbon in each tube, packed in a hexagonal cell and manifold
together. Each hexagonal cell was contained in an outer copper
shell of 150 mm diameter that contained steam at 2 bar pressure
21
during the heating phase. Each individual tube had a smaller
diameter concentric tube that carried cooling water during the
cool down mode. The condenser was water cooled and the
evaporator consisted of a simple stainless steel coil soldered
around a copper box containing up to 5 liters of water. The
results of experiments conducted with this unit showed that the
half cycle times for optimum ice production varied from 16 min
with steam at 150 oC to 26 min with steam at 100 oC.
Abou Karima (1992) designed and tested of a refrigeration
system utilizing solar energy. He concluded that the intermittent
absorption refrigeration system used ammonia-water as a
working fluid by two methods for heating the working fluid in
the generator, the first is the kerosene burner and the second is
the solar energy.
Erickson DC (1994) illustrates the use of solar energy in
remote, non-grid-connected areas. The system consists of seven
double intermittent ammonia–water absorption cycle ice
makers, three of which are ground-mounted to generate ice for
processing fish and transporting it to market, and four, mounted
on the root of the storage building, to provide ice to cool a
storage tank. Each double intermittent ammonia–water
absorption cycle device is supplied with heat by a 12 m2 aperture
area parabolic trough concentrating solar collector and has
produced about 68 kg of ice daily since late 1992. The 84 m2
total parabolic trough solar collector area provides annually a
519 GJ heat input giving an annual 72.8 GJ of refrigeration.
Jianlin and Yanzhon. (2007) The theoretical analysis on the
performance characteristics was carried out for the novel cycle
22
with the refrigerant R141b. Compared with the conventional
cycle, the simulation results show that the coefficient of
performance (COP) of the novel cycle increases, respectively,
by from 9.3 to 12.1% when generating temperature is in a range
of 80 –160 oC, the condensing temperature is in a range of 35–
45 oC and the evaporating temperature is fixed at 10 oC.
Especially due to the enhanced regeneration with increasing the
pump outlet pressure, the improvement of COP of the novel
cycle is approached to 17.8% compared with that in the
conventional cycle under the operating condition that generating
temperature is 100 oC, condensing temperature is 40 oC and
evaporating temperature is 10 oC. Therefore, the characteristics
of the novel cycle performance show its promise in using low
grade thermal energy for the ejector refrigeration system.
Zhang and Noam (2007) proposed several novel systems,
based on ammonia–water working fluid. The proposed plants
operate in a fully-integrated combined cycle mode with
ammonia–water Rankine cycle(s) and an ammonia refrigeration
cycle, interconnected by absorption, separation and heat transfer
processes. It was found that the cogeneration systems have good
performance, with energy and energy efficiencies of 28% and
55-60%, respectively, for the base-case studied (at maximum
heat input temperature of 450 oC).
23
3 MATERIALS AND METHODS
3.1 Principle and system description
Vapor absorption refrigeration cycle as shown in
figure 3-1, consists in its basic configuration of a generator,
condenser, evaporative, and absorber.
A mixture of absorbent “water” and refrigerant
“ammonia” fluids concentrated at 50% used as a working
fluid. In the generator, heat is supplied from the solar flat- plate
collector “FPC” to the fluids mixture to drive off vapor
refrigerant and, as a result, the remaining mixture becomes
diluted, poor in refrigerant, and flows to the absorber
(appendix I).
High-pressure vapor refrigerant flows to the condenser
where it condenses to enter an expansion valve that reduces its
pressure. The outlet of the expansion valve leads to the
evaporator into which the liquid refrigerant flows and removes
heat at low pressure turning into vapor again.
In the absorber, the refrigerant vapor is absorbed into the
poor liquid mixture. Absorber feeds from its two inlets, one for
the refrigerant vapor that flows from the evaporator and the
other is for the poor mixture that flows from the generator after
passing through an expansion valve. Through the absorption
process, heat is released due to the exothermic absorption
reaction. The released heat is removed by a water-cooling
medium. Finally, absorption process causes the mixture to
become rich again in refrigerant.
The circulating control valve, then, raises the pressure of
the rich liquid mixture delivering it to the generator to repeat
the cycle.
24
Figure 3-1: Basic vapor absorption refrigeration cycle.
3.2 System configuration:
3.2.1 Generator: solar flat-plate collector FPC
The solar flat-plate collector has been built with major
purpose to collect as much solar energy as possible at lower
total cost using domestic materials. FPC performance tests
were conducted at Asyut governorates, Egypt. Latitude 27.19
and longitude 31.18, with 14.08 hours daylong and G t = 8200
Wh /m2 per day for solar declination angle of 23.41°.
3.2.1.1 Generator: FPC dimensions and constructions
The FPC main dimensions are illustrated in figure 3-2.
25
142 150
15
122
130
cm
DIM in cm
Figure 3-2: Flat-plate solar collector “FPC”.
3.2.1.2 Generator: FPC absorber plate
The energy absorber plate transfers collected energy to the
mixture fluids. The absorber plate was made of one millimeter
thick black-coated steel sheet for efficient heat transfer. Fluids
mixture were running through a steel tube, 2.5 cm in diameter,
forming ten rows bounded to the absorber sheet by means of
steel clips as shown in figure 3-3.
2.5 cm
0.25 cm
14.2 cm
Steel bonding clip
0.1 cm
FPC absorber dimensions length [Lp]:
FPC width [wp]:
FPC gross area [Ac]:
FPC absorber area [Ap]:
number of tubes:
tubes inner diameter:
tubes outer diameter:
tube-to-tube spacing:
142
122
1.95
1.73
10.0
2.0
2.5
14.2
cm
cm
m2
m2
cm
cm
cm
Figure 3-3: FPC absorber plate components and overall
dimensions.
26
System fluids and its mixture flow rate were measured by
the volume-time measuring method. Volumes were measured
in 500 ml glass cup accurate to ±4 at 20 °C, and time was
recorded with digital stop-watch accurate to 1/60 s. System
pressure was measured by three stainless steel 25 bar pressure
gauges.
3.2.1.3 Generator: FPC efficiency
The FPC was constructed according to the assumption that
the FPC performs under steady state conditions, thus, the
thermal performance analysis can be used to maximize system
efficiency 𝛈𝐅𝐏𝐂 (equation 3-1 and 19) using the system
analysis methods described by Duffie and Beckman (1991);
Kalogirou (2004); and Ashrea (2005) as follows:
𝜼𝑭𝑷𝑪 =
∫ 𝑸𝑼 𝒅𝒕
...................................................... Equation 3-1
𝑨 ∫ 𝑰𝒅𝒕
where 𝐐𝐮 is the useful energy gain in a solar collector “W”, I
is the solar radiation flux incident on the tilted surface of the
FPC “W/m2”,
𝑸𝒊 = 𝑰. (𝝉𝜶). 𝑨 ................................................... Equation 3-2
where Qi is the absorbed solar energy, and A is the surface area
of collector (m2), 𝝉 is the transmittance of the FPC covers, and
𝜶 is the absorptance of the FPC plate.
Overall heat transfer coefficient U, was determined by the
equations 3-7 to 10 (Awady, 1999).
𝑸𝒐 = 𝑼𝑨(𝑻𝒂 − 𝑻𝒎 )........................................... Equation 3-3
Q u = Q i − Q o = I(τα)A- UA(Ta − Tm )
𝑸𝒖 = 𝒎𝑪𝒑 (𝑻𝒉 − 𝑻𝒄 ) ........................................ Equation 3-4
where 𝐐𝐢 is the collector heat input “W”, 𝐐𝐨 is the solar
collector overall heat losses “W”, m is the fluid mass flow rate
27
(liters/s), 𝐓𝐜 is the temperature of fluid “oK”, 𝐓𝐡 is the
temperature of hot fluid “oK”, Ta is the FPC average
temperature “oK”, Tm is the temperature of ambient still air
“oK”, and 𝐜𝐩 is the fluid heat capacity “kJ/kg °K”.
3.2.1.4 Generator: FPC thermal losses
FPC average temperature is difficult to measure through the
different components and layers, which differs in its thermal
properties, as shown in figure 3-4, rather than evaluating the
collector heat removal factor FR. Equations 3-5, and 3-6
describe FR calculation (appendix A).
x1
x2 x3
x4
Tf
T1
T2
T3
T4
T5
Tm
Figure 3-4: Heat transfer through layers of air, wood, glass
wool and absorber plate.
𝑭𝑹 =
𝒎𝑪𝒑 (𝑻𝒉 −𝑻𝒄 )
𝑨(𝑰(𝝉𝜶)− 𝑼(𝑻𝒄 −𝑻𝒎 ))
.................................... Equation 3-5
𝑸𝒖 =𝑭𝑹 𝑨(𝑰(𝝉𝜶) − 𝑼(𝑻𝒄 − 𝑻𝒎 ))................. Equation 3-6
𝑹=∑
𝟏
𝑼𝑨
=
𝟏
𝑨𝟏 𝒉𝒂
+
𝑿𝟏
𝑨𝟏 𝑲 𝟏
+. .
𝑿𝒏
𝑨𝒏 𝑲 𝒏
+
𝟏
𝑨𝒏 𝒉𝒃
...... Equation 3-7
𝑹 = 𝑹𝒂 + 𝑹𝒃 + 𝑹𝟏 + 𝑹𝟐 +𝑹𝒏 ........................ Equation 3-8
R overall = R glass wool+ R wood ............................... Equation 3-9
𝑼=∑
𝟏
𝑨𝑹𝒐𝒗𝒆𝒓𝒂𝒍𝒍
...................................................... Equation 3-10
28
where R is the thermal resistance of insulation “°K/W”, R1 is
the thermal resistance of inner layer of insulation “°K/W”, R2
is the thermal resistance of second layer of insulation “°K/W”,
Rn is the thermal resistance of nth layer of insulation “°K/W”,
RS is the thermal resistance of outer surface of insulation
“°K/W”, hb is the surface coefficient of outer surface “W/m2
°K”, ha is the surface coefficient of inner surface “W/m2 °K”,
kI is the thermal conductivity of inner layer of insulation “W/m
°K”, k2 is the thermal conductivity of second layer of
insulation “W/m° K”, and kn is the thermal conductivity of nth
layer of insulation “W/m °K”.
3.2.1.5 Generator: FPC thermal gain
Egypt is one of countries with major potential of solar
power (appendix B). Estimating the average of total solar
radiation on the inclined surfaces of solar collector can be
found using equations 3-11 to 18.
𝑰 = 𝑰𝒃 𝑹𝒃 + 𝑰𝒅 𝑹𝒅 + (𝑰𝒃 + 𝑰𝒅 )𝑹𝒓 ................. Equation 3-11
𝑰𝒅 = 𝑪 𝑰𝒃𝒏 ............................................................. Equation 3-12
𝑰𝒃𝒏 = 𝑨 𝑬𝒙𝒑
(
−𝜷
)
𝒄𝒐𝒔 𝜽
.............................................. Equation 3-13
𝑰𝒃 = 𝑰𝒃𝒏 𝒄𝒐𝒔 𝜽 ................................................... Equation 3-14
𝑰𝒈 = 𝑰𝒃 + 𝑰𝒅 ....................................................... Equation 3-15
𝑹𝒃 =
𝝎𝒔𝒏 𝒔𝒊𝒏 𝜹 𝒔𝒊𝒏(𝝓−𝜷)+𝒄𝒐𝒔 𝜹 𝒔𝒊𝒏 𝝎𝒔 𝒄𝒐𝒔(𝝓−𝜷)
𝝎𝒔 𝒔𝒊𝒏 𝝓 𝒔𝒊𝒏 𝜹+𝒄𝒐𝒔 𝝓 𝒄𝒐𝒔 𝜹 𝒔𝒊𝒏 𝝎𝒔
. Equation 3-16
𝑹𝒅 = (𝟏 + 𝒄𝒐𝒔 𝜷)/𝟐 ....................................... Equation 3-17
𝟏−𝒄𝒐𝒔 𝜷
𝑹 𝒓 = 𝝆𝒈 (
𝟐
)............................................... Equation 3-18
where A, B, C constants dependent the number of day and
month, 𝜽 is zenith angle 𝑰𝒃𝒏 , 𝑰𝒅 and 𝑰𝒃 are the diffuse radiation
and direct solar radiation respectively, where 𝝎𝒔 is sun rise
29
angle , 𝜷 collector tilt angle, 𝝓 latitude angle and 𝜹 declination
angle, and 𝝆𝒈 is the reflectance of ground.
FPC efficiency, equation 3-18, varies depending on inlet
fluid temperature “Tc” relative to the ambient air temperature
“Tm” and was calculated according to the equation 3-18
𝜼𝑭𝑷𝑪 = 𝑭𝑹 (𝝉𝜶) − 𝑭𝑹 𝑼 (
𝑻𝒄 −𝑻𝒎
𝑰
) ..................... Equation 3-19
3.2.1.6 Generator: FPC thermal and optical losses
To minimize the heat losses from the FPC according to
previous equations, a wooden frame and glass-wool layers
were used to cover back and sides of the collector, figure 3-4.
Heat transfer overall thermal conductivity will be the sum of
the values of conductivity of air, wood glass wool, and
absorber layers showed in figure 3-6, and solved by equations
3-7 to 10 (Awady, 1999).
Figure 3-5: FPC cross section of energy absorption plate
and its two transparent polyethylene covers.
Polyethylene covers configuration and characteristics are
illustrated in figure 3-7.
30
Polyethylene cover # 1
Polyethylene cover # 2
Iron absorbance plate
Insulation material
Properties of cover material -Plastic (Polyethylene)
Solar spectrum refractive index:
Transmittance:
Long-wave absorbance:
Long-wave transmittance:
Number of covers:
Cover-plate air spacing:
Cover 1 – cover 2 air spacing:
Plate material plain carbon steels
conductivity:
Thickness:
Solar spectrum absorbance:
Long-wave emittance:
1.46
0.70
0.05
0.78
2.00
6.00 cm
2.50 cm
60.50 W/m K
0.10 cm
0.88
0.15
Figure 3-6: FPC absorber plate and plastic dual covers.
System temperatures were recorded by a glass-mercury
thermometer. Ranged from -10 to 200 oC, with accuracy of
1oC. A Pyranometer model PSP was used to measure the solar
radiation sensitive to 9 µV per W/m2. The Pyranometer
readings were linear to ±0.5% in measurement ranged from 0
to 2800 W/m2, figure 3-7.
Figure 3-7: Solar intensity measurening pyranometer
model PSP.
31
The system ambient air speed was measured by a cupcounter type anemometer with accuracy of 1 m/s (±5%) in the
measuring range 1- 67 m/s, as illustrated in figure 3-8.
Figure 3-8: Cup-counter type anemometer.
3.2.2 Generator: vessel
Generator vessel is the system component where chilling
liquid fluids gain thermal energy. Heat is supplied from the
FPC to the fluids mixture to drive off high-pressure vapor
refrigerant to the condenser. And, as a result, the remaining
mixture becomes diluted, poor in refrigerant, and flows back
to the absorber.
Since water–ammonia mixture chemically reacts with
traditional materials used to construct FPC systems, such as
copper or brass, the entire system was fabricated out of steel.
32
To the
condenser
The rectifying
Charging line
From the solar
collector
Pressure
gauge
Insulation
From
the evaporator
To the solar
collector
Figure 3-9: Generator storage vessel
Generator vessel was made of cylindrical steel pipe of
12.88 cm diameter, and 60 cm in height. The pipe was selected
with 0.4 cm thick to withstand high pressure resulted from
vapor expansion. Generator vessel -as shown figure 3-9 was
leveled above the top of the FPC for circulating the fluids by
gravitational force at desired flow rate.
Generator volume =
𝝅𝑫𝟐
𝟒
𝑳 ............................. Equation 3-20
3.2.2.1 Generator: vessel thermal losses
Thermal loss form natural convection represented as heat
transfer coefficient hc, depends on: 1) fluid thermal properties:
density𝝆, viscosity𝝁, conductivity k, specific heat at constant
pressure cp and coefficient of thermal expansion 𝜷.
2) generator dimensions: diameter D and length L.
3) environmental factors: generator and ambient air
temperature difference ∆T, and the gravitational acceleration
g.
33
Figure 3-10: Generator storage vessel dimensions.
Experimentally, convection heat transfer coefficient can be
described in factors grouped in dimensionless numbers
Nusselt number (𝑵𝒖) =
Prandtl number (𝑷𝒓) =
Grashof number (𝑮𝒓) =
𝒉𝒄 𝑫
𝑲
𝒄𝒑 𝝁
𝑲
......................... Equation 3-21
.......................... Equation 3-22
𝑫𝟑 𝝆𝟐 𝒈𝜷∆𝑻
𝝁𝟐
................ Equation 3-23
(Nu) = K(Pr)k(Gr)m(L/D)n ............................ Equation 3-24
The rate of convection-heat transfer can be obtained from
calculating Nusselt number, by evaluating of K, k, m, n,
(McAdams, 1954).The natural convection about vertical
cylinders can be found from equations
(Nu) = 0.53(Pr.Gr)0.25 ....................................... Equation 3-25
(Nu) = 0.12(Pr.Gr)0.33 ....................................... Equation 3-26
34
Equation 3-25 for the range 104 < (Pr.Gr) < 109 and
equation 3-26 for the range 109 < (Pr.Gr) < 1012
3.2.2.2 Generator: vessel efficiency
Generator vessel efficiency is a ratio of gained energy by
the vessel heat exchanger 𝐐𝐠𝐯 to the input thermal energy 𝐐𝐮
(equation 8) gained from the FPC. The generator vessel
efficiency ηgv , was calculated according to equation
𝜼𝒈𝒗 =
𝑸𝒈𝒗
𝑸𝒖
............................................................ Equation 3-27
The generator vessel net energy gain Qgv, was obtained by
equation 3-27.
𝑸𝒈𝒗 = 𝑸𝑼 − 𝑸𝒐 .................................................. Equation 3-28
𝑸𝒈𝒗 = 𝑼𝒈𝒗 𝑨 ∆𝑻 ................................................. Equation 3-29
𝑸𝑳 =
𝑻𝒊 −𝑻𝒐
∑𝑹
........................................................... Equation 3-30
r3, To Mineral wool
r2
r1, Ti Iron cylinder
T3 T2 T1
k2 k1
Figure 3-11: Heat transfer through layer of glass wool, and
generator vessel surface.
Σ R = R conv + R cyl + R insulation + R conv ............ quation 3-31
𝑹𝒄𝒐𝒗 𝒐 =
𝟏
𝑨𝒐 𝒉𝟎
.......................................................... Equation 3-32
𝑹𝒊𝒏𝒔𝒖𝒍𝒂𝒕𝒊𝒐𝒏 =
𝒓
𝑳𝒏 𝟑
𝒓𝟐
𝟐𝝅𝑲𝟐 𝑳
........................................... Equation 3-33
35
𝒓
𝑳𝒏 𝟐
𝒓𝟏
𝑹𝒄𝒚𝒍 =
𝟐𝝅𝑲𝟏 𝑳
𝑹𝒄𝒐𝒗 𝒊 =
𝑼𝒈𝒗 =
𝟏
𝑨𝒊 𝒉𝒊
........................................................ Equation 3-34
.......................................................... Equation 3-35
𝟏
𝒓
𝟏 𝒓𝒊 𝒍𝒏(𝒓𝒐 ⁄𝒓𝒑 ) 𝒓𝒊 𝒍𝒏(𝒓𝒑 ⁄𝒓𝒊 )
+
+
+ 𝒊 +𝑹𝒇
𝒉𝒊
𝒌𝒊𝒏𝒔𝒖𝒍
𝒌𝒑𝒊𝒑𝒆
𝒓𝒐 𝒉𝒐
............... Equation 3-36
where, mw is the mass of fluid in the storage tank “kg”, Tke
and Tkb are the fluid temperatures in the generator vessel at the
end and the beginning of each cycle “°K” respectively.
3.2.3 Chiller condenser- Evaporator
Assuming that the heat-transfer coefficient is independent
of the fluid constant temperature.
3.2.3.1 Condenser –Evaporator: dimension
Two overlapping cylinders represent the condenser and the
evaporator vessels as shown in figure (3-12). Outer cylinder is
30 cm in diameter and 40 cm of length. The inner cylinder is
10 cm in diameter and was cooled by running water in the outer
cylinder space in the generation process. While in refrigeration
process, the inner cylinder works as evaporator (without the
cooling process and the outer cylinder is empty of water).
36
Absorber vessel Generator vessel
Pressure
gauge
Expansion valve
Water outlet
Water cooling tank
Insulation
Water inlet
Figure 3-12: Condenser –Evaporator dimensions in cm.
3.2.3.2 Condenser –Evaporator: efficiency
In each one of the performed experiment the temperature of
the cold fluid (water) and the hot vapor of ammonia was
recorded and the efficiency of the condenser was estimated by
Helmut Wolf (1983) equation for heat transfer
𝜼𝒄 = (
𝑻𝒄𝒐 −𝑻𝒄𝒊
𝑻𝒉𝒊 −𝑻𝒄𝒊
)𝟏𝟎𝟎 .............................................. Equation 3-37
where Tco, Tci are the outlet and the inlet temperature of the
fluid respectively “°K”, Thi is inlet temperature of the “oK”.
According to Kassem et al. (1993) in order to calculate the
effective cooling produced, it is necessary first to measure the
maximum and minimum temperature of load and Qe can be
calculated from equation 3-48.
𝑸𝒆 = 𝑾𝑷 𝑪𝑷𝑳 (𝑻𝒎𝒂𝒙 − 𝑻𝒎𝒊𝒏 )........................... Equation 3-38
37
where Qe is the total cooling obtainable “kJ”, Wp is the thermal
mass load “kg”, CpL is the specific heat of load material
“kj/kg”, Tmax is initial temperature of thermal load at which the
cooling process starts °C, and Tmin is the minimum load
temperature that can be reached °C.
𝑸𝒆 = 𝒉𝒗 − 𝒉𝑳 ...................................................... Equation 3-39
where hL is the enthalpy of liquefied ammonia at condensation
temperature and pressure (kJ), and hv is the enthalpy of
ammonia vapor at vaporization temperature and pressure (kJ).
The cooling ratio of the cycle measures the performance of
the system and is defined in equation 3-48.
𝑪𝒐𝒐𝒍𝒊𝒈 𝒓𝒂𝒕𝒊𝒐 =
𝑸𝒆
𝑸𝑼
..................................... Equation 3-40
where Qe is the cooling available during refrigeration period,
and Qa is the heat absorbed by ammonia-water in collector
during regeneration process.
3.3 Chilling- solar assisted system overall efficiency
The performance of the solar refrigerator is defined as
mentioned by in equation (3-41) Mansoori and Patel (1979).
𝑪𝑶𝑷 = 𝜼𝑭𝑷𝑪 𝜼𝒈 𝜼𝒄 ........................................... Equation 3-41
3.4 Thermal load and chilling system
Thermal load calculations summary is illustrated in figure
(3-13). For experimental setup, the calculations of thermal
load will cover field heat, heat of respiration, and heat leakage
for potatoes crop.
38
Figure 3-13: Thermal load and chilling system calculations.
39
3.5 Experimental setup
Experiments were conducted in Agricultural Engineering
Department, Faculty of Agriculture, Al-Azhar University,
Assiut branch. At the commencement of the generation
process, valves I. III and V are closed and valves I is opened.
The solar flat plate collector gain causes the ammonia-water
mixture to heat up. Circulation of the fluid is up the pipes, and
into the header than to the generation storage vessel. Ammonia
vapor rises out of solution, than rectifying column and into
condenser.
At the condenser, the ammonia vapor condenses. At the end
of the generation period, the condenser is isolated from the
generator by closing valve I and the temperature the system
allowed cooling to ambient conditions. The refrigeration
period is started by opening valve III so liquid ammonia in the
evaporator vaporizes back to the absorber, where it is
reabsorbed by the weak solution. The evaporation of the
ammonia extracts heat from cooling load in the box
surrounding the evaporator sufficient heat is removed to chill
the products.
A photograph of the solar refrigeration is shown in figure
(3-14), and the engineering drawing as shown in figure 3-15.
40
a-Generation
b- Refrigeration
Figure 3-14: Configuration of the experimental setup.
41
Figure 3-15: The solar refrigerator system:1) flat plate
solar collector, 2) generator vessel, 3) pressure gauges, 4)
expansion valve. 5) evaporator-condenser(immersed in a
tank of water) 6) evaporator- condenser (dry)
The detailed description of the essential parts of the solar
refrigeration is shown in figure (3-17 and 18). These parts are
flat plate solar collector, Generator storage vessel, Evaporator-
42
condenser (immersed in a tank of water) and Evaporatorcondenser (dry).
Figure 3-16: The solar refrigeration component.
Figure 3-17: Overview of the solar assisted chilling system.
43
4 RESULTS AND DISCUSSIONS
4.1 Generator performance
4.1.1 Generator: FPC thermal performance
4.1.1.1 Air speed thermal removal rate
In FPC energy losses occur by natural and forced
convection, as well as radiation. The FPC thermal losses were
minimized by using two insulation materials. A wooden frame
-4 cm thick- with thermal conductivity of 0.16 W/m°K. A glass
wool layer -4 cm thick-with thermal conductivity of 0.0052
W/m°K. The FPC overall thermal condactivity at sides and
bottom was 0.17 W/m°K , as illustrated in equations 3-7, 8, 9.
Thermal losses (kWh)
Figure (4-1) shows measured thermal losses from air flow
around the FPC. The thermal losses were estimated by 2.0, 3.0,
and 1.2 kWh for air speeds ranged between 1-4, 4-21, and
21-30 km/h respectively. Appendix A shows the Egypt
climate graph, and air speed according to month of the year.
0.45
0.40
0.35
0.30
0.25
0.20
0.15
0.10
0.05
0.00
0.4
heat removal rate= 1,2 k Wh/ km/h
0.3
0.2
heat removal rate= 3.0 k Wh/ km/h
0.1
heat removal rate= 2.0 k Wh/ km/h
0
5
10
15
20
Air speed (km/h)
25
30
Figure 4-1: The relationship between air velocity and heat
losses.
44
4.1.1.2 Generator: FPC: optical efficiency
To maintain FPC top thermal loss at minimum level. FPC
was covered by a dual polyethylene-plastic covers. The
amount of absorbed solar energy by the FPC covers depended
on the effective transmittance (solar spectrum τ= 0.7, and
long-wave τ= 0.78), reflective index (n= 1.46), and
absorptance (α =0.05) of the dual polyethylene plastic covers.
Two transparent polyethylene-covers were used to reduce
convection losses from the FPC through the restraint of the
stagnant air layer between the absorber plate and the
transparent covers. It also reduces radiation losses from the
collector as the covers are transparent to the short wave
radiation received by the sun but it is nearly opaque to longwave thermal radiation emitted by the absorber plate.
4.1.1.3 Generator: FPC overall heat removal coefficient
The generator FPC overall heat removal coefficient U was
8.775 according to equation 3-10. FPC overall heat removal
factor was 8.755 W/m2 °C, measured at solar intensity of 1002
W/m2, air ambient temperature 34 °C, air speed
5.7 m/s, and relative humidity 70%.
4.1.1.4 Generator: FPC overall heat removal factor FR
FPC overall heat removal factor was used instead of FPC
average temperature which is difficult to measure. FR
importance is to measure effectiveness of thermal insulation
materials as shown from the efficiency curve slope in
figure (4-2). FPC overall heat removal factor FR was 0.3971
according to equation (3-5).
45
1.00
 = −1,8215
0.80
 = −5,3411
𝛈 𝐅𝐏𝐂
0.60
(𝑇𝑎𝑣 −𝑇𝑎 ) 2
−
𝐺𝑡
5,2294
(𝑇𝑎𝑣 −𝑇𝑎 )
𝐺𝑡
+ 0,2591
(𝑇𝑎𝑣 − 𝑇𝑎 )
+ 0,2606
𝐺𝑡
0.40
0.20
theraml loss = 5.3411
0.00
0
0.01
0.02
0.03
0.04
0.05
0.06
Figure 4-2: FPC optical and thermal efficiencies.
FPC is characterized by the intercept value of the efficiency
curve, which reperesent optical efficiency of the FPC top cover
materials (0= FR (𝜏𝛼) = 0.26). Typical value of the 0 for
standard FPC type ranges between 0.65-0.80.
Thermal insulation effectivness was found by the slope of
the efficieny curve represeneted by FRU and equals to
5.3 W/m2 °C. Typical value of the FRUL for standard FPC type
ranges between 3-8 W/m2 °C.
Figure (4-3) illustrates the relationship of gained
temperature of the working fluid measured at different time of
the day. Figure (4-4) illustrates the relationship between the
FPC absorber plate temperature TP, and ammonia-water
mixture temperature Taw.
46
Temperature o C
100
90
80
70
60
50
40
30
9 AM
10 AM
11 AM
12 PM
1 PM
2 PM
Time "hours"
Figure 4-3: Daly heat transfear from the FPC to the
working fluid in chilling system.
Ammonia -water
temperature oC
100
80
60
40
20
0
50
70
90
Plate temperature oC
110
Figure 4-4: The relationship between the FPC abosrber
plate temperature and Ammonia-water temperature.
4.1.1.5 Generator: FPC thermal performance
The generator FPC efficiency curver intercept with FPC
absorber plate temprature at temprature euillpruim point.
Temprature equilibruim means that usefule energy is no longer
removed by the working fluid. That state of equilibrium can be
reached when the working fluid is stoped to circulate through
47
the system. The equilibruim point was emerged when FPC
absorber plate temprature was 93 °C, as shwon in figure (4-5).
30
25
20
15
10
5
0
44
54
64
74
84
94
FPC absorber plate temperature in °C
30
𝛈 𝐅𝐏𝐂
25
20
15
10
5
0
9 AM 10 AM 11 AM 12 PM 1 PM
Time of the day.
100
90
80
70
60
50
40
30
20
2 PM
Figure 4-6: Heat removal rate in relation to the FPC
absorber plate temprature during the day.
48
Plate temperature oC
Figure 4-5:The relationship between the FPC absorber
plate temperature and the FPC efficiency.
4.1.2 Generator: vessel thermal performance
To minimize thermal loss form generator vessel, a glass
wool layer was cover the exposed surface of the generator
vessel. Measurements were taken to calculate the amount of
energy passed through the generator. Figure (4-7), showed the
energy delivered to the generator vessel from the generator
FPC. It was found that increased temperature difference will
decrease the amount of useful energy passed through the
generator vessel. The generator vessel heat transfer was
decreased from 98% to 22.8% with increase temperature
difference from 9 to 44 °C.
generator FPC
ηgenerator
300
1
250
0.8
generator vessel
Energy W/h
generator vessel
200
0.6
150
0.4
100
0.2
50
0
0
0
10
20
30
40
Tgenerator - T ambient in °C
50
Figure 4-7: Heat transfer from the generator FPC to vessel
in relation to temprature difference.
4.2 Condeser thermal performance
results of this experiment that the efficiency were varied
between 55 to 25% through the day hours (appendix C).
Figure (4-8) illustrates the energy gain from the generator
49
vessel to the condenser as a function of the temprature
difference of fluid.
300
Energy W/h
250
200
150
100
50
0
1.0
2.3
2.0
1.7
Two - Twi
2.5
3.0
Figure 4-8: Heat transfer from generator vessel through
Condenser-Evaporator.
Energy losses depends on the difference between inlet and
outlet temperature of the cold fluid and the difference between
inlet temperature to each of cold and hot fluid during the
process. Figure (4-9) showed the system temprature changed
according to solar intensity during one day of operation (one
cycle of genration and evaporation, appendices J and k). The
system pressure changed as a function of system energy (figure
4-10).
4.3 Evaporatore thermal performance and COP
Solar energy flows through the chilling system at different
rates. System construction materials, system thermal
insulations, fluids, and environmental factors are crucial in
determining the system overall efficiency. Figure (4-11)
demonstrate an efficiency comparison for the generator FPC,
generator vessel, and for the condenser-evaporator, at different
solar intensities (appendices E through H).
50
1200
90
80
900
Temperature °C
70
60
600
50
40
300
Solar radiation W/m2
100
30
20
0
9AM
10AM 11AM
12PM
1PM
2PM
Time
12
80
10
Temperature oC
100
8
60
6
40
4
20
2
0
Pressure kg/cm2
Figure 4-9 :Heat transfer through chilling system at solar
radiation intensities.
0
9 AM
10 AM 11 AM 12 PM
1 PM
2 PM
Time
Figure 4-10: Heat transfer flow from FPC absorber plate
through chilling system at different system pressure.
At average performance, the generator FPC delivered a
14.3% of the total available solar energy in one day of tests.
While generator vessel passes 98% from the generator FPC
energy or 13% from the total solar intensity to the condenserevaporator section. Net energy of the condenser unit reached
51
1.20
1.00
Efficiency %
0.80
0.60
0.40
0.20
0.00
8 AM
9 AM 10 AM 12 PM 1 PM
Time
1800
1600
1400
1200
1000
800
600
400
200
0
2 PM
Solar intensity (Gt) W/m2
55% form the energy delivered form generator vessel, and
equals to 4.43 % from the total solar intensity (figure 4-12 and
4-13).
Enrgy gain "W"
300
2000
250
1500
200
150
1000
100
500
50
0
8 AM
Solar intensty (Gt) W/m2
Figure 4-11:Chilling solar assiseted system effeciency at
solar intinsties during the day time.
0
9 AM
11 AM 12 PM
Time
1 PM
Figure 4-12: Chilling solar assiseted system energy flow at
solar intensities during the day time.
52
9027.262872
1130.29094
1288.84395
399.782342
Figure 4-13: Energy transfer through generator FPC,
generator vessel, and condenser.
solar intensity "W/m2"
1800
0.14
1600
0.12
1400
0.1
1200
1000
0.08
800
0.06
600
0.04
400
0.02
200
0
0
COP, and ton of refrigeration
The chilling-solar assisted system coefficient of
performance COP relationship as found in figure (4-14)
reached its maximum value at 0.21 (equivalent to 0.03 ton of
refrigeration) and decreased as system temperature increased.
9:00 AM 10:00 AM 11:00 AM 12:00 PM 1:00 PM 2:00 PM
Time
Figure 4-14: The coefficient of performance for the chilling
solar assisted system.
53
35
30
25
20
15
10
5
0
8
7
6
5
4
3
2
1
0
0
50
100
150
200
Chilling time "min"
Pressure kg/cm2
Temprature °C
Chilling solar assisted system was able to reduce evaporator
temperature form 30 °C to 4 °C within 100 minutes without
load as illustrated in figure (4-15).
250
Figure 4-15: Required time for heat removal from the
evaporator without load “empty” and system pressure.
Field growing crops suffered from heat gain of respiration,
and from the surrounding environment. Calculated heat load
for one kilogram of potatoes showed increase of energy load
form 25 W to 35.4 W with increasing ambient air temperature
form 5 °C to 25 °C (figure 4-16).
Chilling solar assisted system was constructed for fast
removal of crops’ field heat. The system capacities will depend
on the desired final temperature for safe transport of the crop.
Figure (4-17) compares the net energy to remove form one
kilogram of potatoes crop to reduce its temperature from 30 °C
to 4 °C. Summing up needed reduction of potatoes energy,
cooling energy lost for surroundings, and potatoes respiration
energies form integration of the area under curve in figure (418). It was found that cooling energy of one kilogram of
potatoes will consume 119.65 W of the 399.8 W- evaporator
54
Potatoes heat load "W"
available energy or 1.33% of the total solar intensity under the
experimental conditions.
38
36
34
32
30
28
26
24
5
10
15
20
Ambiant air temprature °C
25
Figure 4-16: Calculated potatoes heat load at different
ambient temperature.
50
45
40
35
30
25
20
15
10
5
0
30
Energy load "W"
20
15
10
5
respiration heat load
25
50
75
100
Time "min"
0
125
Figure 4-17: Potatoes thermal load energy and heat
removal response.
55
Temprature "°C"
25
500
25
400
20
300
15
200
10
5
100
Energy "W"
Temprature "°C"
30
heat load of kg potatoes "W"
0
0
25
50
75
100
Time "min"
125
Figure 4-18: Calculated heat load removal from kilogram of
potatoes and energy transfer from condensing to
evaporating processes.
Chiller capacity in kg
From figure (4-19) the evaporator capacity will be three
kilograms of potatoes for temperature reduction form 30 °C to
5 °C, and 18 kilograms for temperature reduction form 30 °C
to 20 °C.
20
18
15
10
10
7
3
5
0
5
10
15
Temprature °C
20
Figure 4-19: Chilling solar assisted potatoes holding
capacities at different chilling temperature levels.
56
Chilling will take 45 minutes to reduce kilogram of
potatoes from 30 °C to 20 °C and 120 minutes to reach 4 °C as
seen in figure (4-20).
5 °C
10 °C
15 °C
20 °C
Energy "W"
25
20
15
10
5
0
0
25
50
75
100
Chilling time "min"
125
Figure 4-20: Required time for heat removal from potatoes
at different temprature levels.
57
5 SUMMARY AND CONCLUSION
Small, remotes, and reclaimed agricultural lands sufferer
from economical potential and poor or absent of power
networks for post-harvest treatments. Solar assisted chiller was
constructed for fast removal of field heat which causes rapid
deteriorations in field crops. The system consisted of generator
FPC and vessel, and a condenser- evaporator unit. And woks
as intermittent chilling system.
A mixture of absorbent “water” and refrigerant “ammonia”
fluids concentrated at 50% was used as a working fluid. The
power source was a solar flat plate collector FPC with gross
area of 1.90 m2. The FPC accumulate the collected solar
energy in the generator steel vessel by the ammonia-water
mixture. Two overlapping steel cylinders represent the
condenser and the evaporator vessels. The outer cylinder woks
as a water cooler for the inner cylinder which works as a
condenser in the condensing process. In the evaporating
process (chilling) the outer cylinder is empty of water and the
inner cylinder works as evaporator.
Experiments of the solar assisted chiller was carried for
determination of the system components and the overall
thermal performance. It was found that the maximum system
COP was 0.21 equals to 0.04 ton of refrigeration load. And can
be used to reduce three kilograms of potatoes crop from 30 °C
to 4 °C under the experiment conditions. And 18 kilograms of
potatoes to 20 °C under the same conditions.
Chilling system generator, consisted of FPC and generator
vessel. FPC optical efficiency (0= FR (𝜏𝛼)) was 26% and
varied according to ambient air temperature and the rate of
fluid flow. The FPC average efficiency under the experimental
58
conditions was 13% at total solar intensity of 9027.26
W/m2/day and air speed of 5.3 m/s. The lowest obtained
temperature in evaporator was 3.2 °C.
FPC energy losses occur by natural and forced convection,
as well as by radiation. To minimize these loses the collector
was thermally insulated on its sides and bottom. While top of
the collector was covered by a dual polyethylene-plastic
covers.
Measured thermal losses from air flow around the collector
were estimated by 2.0, 3.0, and 1.2 kWh for air speeds ranged
between 1-4, 4-21, and 21-30 km/h respectively. Installation
of the system nearby agricultural facilities or vegetative
canopies barrier should minimize energy losses alongside the
collector’s perimeter. The barrier reduces the air velocities
over the solar collector by creating a recirculation zone behind
it.
Polyethylene covers are limited in the temperatures they can
sustain without deteriorating or undergoing dimensional
changes, and the ability of polyethylene to withstand the sun’s
ultraviolet radiation for long periods. These drawbacks of
using polyethylene as FPC covers can be recovered by its low
weight and cost with its ability to withstand shocks without
being broken.
The generator FPC overall heat removal coefficient U was
8.775 W/m2 °C and the overall heat removal factor FR was
0.3971. Typical value of the FRUL for standard FPC type
ranges between 3-8 W/m2 °C. Due to the lower efficiency of
the FPC and energy delivered by it. Further enhancement on
optical efficiency must be mad, by changing the plastic cover
59
to non-iron glass cover. Focused and solar concentrator
collectors are essential for large capacities of chilling load.
Generator vessel, maximum thermal efficiency was 98%
with ability to deliver 12% of the total solar intensity per day.
The steel generator vessel was able to withstand system
pressure (3.4 to 12.5 kg/cm2) and corrosions at all experiments.
A rectifying column was fitted at the top of the generator
storage vessel to prevent water from being carried over to the
condenser.
Condenser-evaporator, with 55% thermal efficiency was
able to deliver 4.4% of the total solar intensity energy as
chilling energy per cycle. The ammonia-water mixture has
excellent thermodynamic and physical properties for chilling
purposes. Toxicity might be the reason behind its limited
usage, which can be vanished in open field applications.
Solar assisted chilling system is a simple and cheap system
to build. It has a dual usage in post-harvest treatments, and can
work as chiller or as a heater in many agricultural applications.
60
6 REFERENCES
Abdellatif, S. M.; El-Adl, M. A. and Silim, A. H., 2009,
Solar energy utilization for providing and maintaining
optimal microclimatic conditions of greenhouse
cucumber. Part I: Thermal performance analysis of
solar water heat, Miser J. Agric. Eng., 26(3): 13771396.
Abou Karima, 1992, Design and testing of a refrigeration
system utilizing solar energy, M.Sc., Faculty of
Agric., Alex. Univ.
Anyanwu, E.E and Ezekwe, C.I., 2003, Design,
construction and test run of a solid adsorption solar
refrigerator using activated carbon/methanol, as
adsorbent/adsorbate pair. J. Energy Converse
Manage. 44:2879–2892.
Ashrae, 2005, Handbook of fundamentals, American Soci.
of Heating Refrig. Air Cond. Eng. N.Y.,
Assilzadeh, F., Kalogirou, S.A. Alia, Y. and Sopian K.,
2005, Simulation and optimization of a LiBr solar
absorption cooling system with evacuated tube
collectors, J. Renewable Ene., 30:1143–1159.
Awady, M. N., 1999, Engineering of agricultural materials
processing, Memos, Col. Ag., Ain-Shams Univ.:138.
Bansal, N. K., Blumenberg, J. Kavasch, H. J and
Roettinger, T., 1997, Performance testing and
evaluation of solid absorption solar cooling unit,
Solar Ene., 61:127–140.
61
Bargach, M. N., Tadili, R.; Dahmam, A. S.; and
Boukallouch, M., 2000, Survey of thermal
performances of a solar system used for heating
agricultural structures in Morocco, J. Renewable
Energy, 20:133-415.
Critoph, R.E and Gong, F. 1992, A rapid cycling ice
maker for use in developing countries. Second World
Renewable Energy Congress, vol. 2. p. 708–713.
Critoph, R.E. 1994. An ammonia carbon solar refrigerator
for vaccine cooling. Renewable Energy 5:502–508.
Critoph, R.E. 1996, Towards a one tone per day solar ice
maker. Renewable Energy.9:626–631.
De Francisco, A., Illanes, R. Torres, J.L. Castillo, M.
De Blas, M., and Prieto, E., 2002, Development and
testing of a prototype of low-power water–ammonia
absorption equipment for solar energy applications. J.
Renewable Energy: 25.
Duffie, J.A and Beckman, W. A., 1991, Solar engineering
of thermal processes. N. Y.: Wiley.
Ackermann, H., 2004, On the correlations between
collector efficiency factor and material content of
parallel flow flat-plate solar collectors, J. Solar
Energy, 76:381-387.
Erickson DC ,1994, Solar icemakers in Maruata. Solar
Today, July/August.
Eze, C.L., 2004, Alternate energy resources, 1st Ed. Mac.
Nigeria Pub. L.: 159 – 177.
62
Hammad, M and Habali. S., 2000, Design and
performance study of a solar energy powered vaccine
cabinet, J. App. Therm. Eng., 20:1785–1798.
Hildbrand, C., Dind, P. Pons, M and Buchter, F., 2004,
A new solar powered adsorption refrigerator with
high performance, J. Solar Energy, 77:311–319.
International Energy Agency (IEA), 2011, Solar Airconditioning and refrigeration solar cooling position.
Jianlin, Yu and Yanzhong, L., 2007, A theoretical study
of a novel regenerative ejector refrigeration cycle.
Inter. J. Refrig., 30:464 – 470.
Jorge; E. Lozano, 2006, Fruit manufacturing, scientific
basis, engineering properties, and deteriorative
reactions to technological importance, Food Eng.
Ser., Springer Sci.: 8.
Kader; A.A., 2002, Post-harvest technology of
horticultural crops, 3rd Ed., Univ. Calif. Agric. Nat.
Reso., Oakland, Pub. 3311.
Kalogirou, S. A. 2004, Solar thermal collectors and
applications. Progress in Energy and Combustion
Science, 30 (2004):231-295.
Kassem, A.; A.Shoker and A.Bassuony 1993,
performance of an intermittent absorption solar
refrigeration system. Misr J. Agric. Eng, 10 (2): 267285
Khattab, N.M. 2004, A novel solar-powered adsorption
refrigeration module, App. Thermal Eng., 24: 2747–
2760.
63
Lemmini, F and Errougani, A., 2005, Building and
experimentation of a solar powered adsorption
refrigerator, J. Renewable Energy ,.30:1989–2003.
Li, M., Sun, C. J. Wang, R. Z and CAI, W. D., 2004,
Development of no valve solar ice maker. J. App.
Thermal Eng., 24:865-872.
Zhang, Na., and Noam, Lior, 2007, Methodology for
thermal
design
of
novel
combined
refrigeration/power binary fluid systems, Inter. J.
Refrig., 30: 1072 –1085.
Oertel, K., and Fischer M., 1998, Adsorption cooling
system for cold storage using methanol/silica-gel. J.
App. Thermal Eng.:18.
Pridasawas W., 2006, Solar-driven refrigeration systems
with focus on the ejector cycle, PhD thesis, Royal
Institute of Technology:
Sierra, F.Z., Best, R. and Holland, F.A., 1993,
Experiments on an absorption refrigeration system
powered by a solar pond, J. Heat Recov. Sys. CHP.
13:401-408.
Sumathy, K, and Li, Z.F., 1999, Experiments with solarpowered
adsorption
ice-maker.
Renewable
Energ.,16:704–707.
Sumathy, K., Huang. Z.C, and Li, Z.F., 2002, Solar
absorption cooling with low grade heat source—a
strategy of development in south China, J. Solar
Energy, 72:155–165.
64
Syeda, A., Izquierdo.M, Rodrı´gue.P, Maidment. G,
Missenden. J, and Lecuona. A, 2005, A novel
experimental investigation of a solar cooling system
in Madrid. Inter. J Refrig., 28:859–871.
Wang, R.Z., 2000, Adsorption refrigeration research in
Shanghai Jiao Tong University. J. Renewable
Sustainable Energy Rev., 5:1–37.
Watheq, 2008, Solar energy refrigeration by liquid-solid
adsorption technique, M.Sc. thesis, Dept. of Sc. in
Clean Energy and Energy Conserv. Eng., Fac.
Graduate Stud., An-Najah University, Nablus –
Palestine
Wimolsiri Pridasawas, 2006, Solar-driven refrigeration
systems with focus on the ejector cycle, PhD thesis,
Royal Inst. Tech., Denmark:
Yeung, M. R., P.K. Yuen, A. Dunn, and Cornish, L.S.,
1992, Performance of a solar-powered air
conditioning system in Hong Kong, J. Solar Energy,
48:309–319.
65
7 APPENDICES
APPENDIX .A
The relationship between ambient air flow and
heat losses from the FPC
Day time
09:00 AM
09:30 AM
10:00 AM
10:30 AM
11:00 AM
11:30 AM
12:00 PM
12:30 PM
01:00 PM
01:30 PM
02:00 PM
Air flow
(km/h)
Heat losses
(kWh)
FPC temp.
°C
Air flow temp.
°C
18.36
0.121
38
30
18.36
0.211
43
30
24.12
0.333
51
32
24.12
0.391
56
32
25.92
0.499
63
34
25.92
0.557
66
34
29.52
0.575
68
35
29.52
0.610
70
35
27.72
0.645
73
36
27.72
0.679
75
36
25.92
0.732
78
36
66
APPENDIX .B
Egypt Climate Graph
Published on http://www.climatetemp.info/egypt/
67
APPENDIX .C
Condenser efficiency.
16/6/2011
9:00 AM
9:30 AM
10:00 AM
10:30 AM
11:00 AM
11:30 AM
12:00 PM
12:30 PM
1:00 PM
1:30 PM
2:00 PM
η
0.556
0.5
0.477
0.417
0.333
0.27
0.236
0.25
0.257
0.263
0.272
Tho
26.6
27.5
28.5
29
29.7
30
30.8
31
32
32.4
33
Thi
27.8
29.2
31
32.5
33
34
35
36
38
38
40
Two
27
27.6
28.5
29
29
29.3
29.5
30
30.8
31
32
Twi
26
26
26.2
26.5
27
27.5
27.8
28
28.3
28.5
29
where Tco, Tci are outlet and inlet temperature of the water
respectively ”oK”, Thi inlet temperature of the refrigerant
fluid.
68
APPENDIX .D
Relationship between plate temperature and water
temperature day 16/6/2011
Day hours
FPC temp.
°C
Air flow
(km/h)
T-plate
o
C
T-Win
o
C
T-out
o
C
9:00 AM
30
18.36
44
38
40
9:30 AM
30
18.36
51
43
46
10:00 AM
31
24.12
59
51
54
10:30 AM
33
24.12
64
56
58
11:00 AM
34
25.92
70
63
65
11:30 AM
34
25.92
74
66
68
12:00 PM
35
29.52
79
68
73
12:30 PM
35
29.52
82
70
77
1:00 PM
36
27.72
86
73
80
1:30 PM
36
27.72
90
75
83
2:00 PM
36
25.92
95
78
87
Twi, Two: Interring and exit cold water temperature respectively, oC
69
APPENDIX .E
Observation during Generation and Refrigeration Test on June 16, 2011
Generation
Pressure T-conde
Kg/cm2
Co
5.5
26.6
6.2
27.5
7.4
28.5
8.9
29
10.6
29.7
11
30
11.4
30.8
11.4
31
11.5
32
11.6
32.4
11.7
33
10.34
30
T-water
out Co
27
27.6
28.5
29
29
29.3
29.5
30
30.8
31
32
29.4
Refrigeration
Time min
Evaporator-temp. Co
Absorber- temp. Co
Absorber –press. kg/cm2
0
36
36
6.7
T-water
in Co
26
26
26.2
26.5
27
27.5
27.8
28
28.3
28.5
29
27.345
15
32
36
5.8
30
29
36
5.6
T-solution
Co
40
46
54
58
65
68
73
77
80
83
87
68.16
45
18
36
5.2
60
11
34
4.6
Fi Co
38
43
51
56
63
66
70
74
77
80
78.5
64.95
75
8.2
34
4.2
90
7.5
34
3.8
- 70 -
T-plate
Co
44
51
59
64
70
74
79
82
86
90
95
74
105
6.2
34
3.5
120
4.8
33
3.6
wind
m/s
5.1
5.1
6.7
6.7
7.2
7.2
8.2
8.2
7.7
7.7
7.2
7.016
135
4.2
33
4
Amb temper
Co
30
30
31
33
34
34
35
35
36
36
36
33.9
150
4.4
33
4.3
165
4.8
33
4.8
180
5
31
5
radiation
W/m2
720.97
802.465
883.96
913.745
943.53
973.32
1003.11
973.32
943.53
913.745
883.96
896.51
195
5.3
31
5.6
210
5.7
31
5.9
Day hours
00 AM
9:30 AM
10:00 AM
10:30 AM
11:00 AM
11:30 AM
12:00 PM
12:30 PM
1:00 PM
1:30 PM
2:00 PM
Ave
225
6.3
31
5.9
240
6.7
31
6.2
APPENDIX .F
Observation during Generation and Refrigeration Test on June 18, 2011
Generation
Pressure T-conde
Kg/cm2
Co
5.1
27
5.5
27.8
7.2
28.3
8.2
29
10.3
30
10.8
30
11.3
30.8
11.5
31
11.4
31.7
11.6
32.4
11.6
32.8
10.125
30.1
Refrigeration
Time min
Evaporr-temp. Co
Abso- temp. Co
Abso –pres kg/cm2
T-water
out Co
27
28.2
29
29.5
30
30
30.2
30.7
31
31.5
32
29.9
0
32
32
6.2
15
28
32
5.5
T-wa in
Co
26
26
26.4
26.7
27.2
27.8
28
28.5
29
29.2
29.5
27.664
30
24
32
5.2
45
16
32
5.2
T-solu
Co
38
44
55
57
63
66
71
75
78
84
86
66.9166
60
10
31
4.6
75
7.5
31
4.2
Fi Co
35
41
50
52
61
63
68
70
74
81
83
63.4166
90
6.4
31
3.8
105
6
31
3.4
71
T-pl
Co
45
50
61
63
68
73
75
80
85
90
92
72.83
120
5.2
29
3.9
wind
m/s
6.7
6.7
7.2
7.2
8.2
8.2
8.2
8.2
9.3
9.3
9.3
8.15
135
4.5
29
4.2
150
4.8
29
4.5
Amb tem
Co
26
27
28
29
30
31
32
32
33
33
33
30.583
165
4.9
29
4.9
180
5.3
28
5.3
radiation
W/m2
721.07
802.57
884.07
913.83
943.59
973.355
1003.12
973.355
943.59
913.83
884.07
896.585
195
5.8
28
5.5
210
6
28
5.5
Day hours
9:00 AM
9:30 AM
10:00 AM
10:30 AM
11:00 AM
11:30 AM
12:00 PM
12:30 PM
1:00 PM
1:30 PM
2:00 PM
Ave
225
5.8
28
6.3
240
5.9
28
6.9
APPENDIX .G
Observation during Generation and Refrigeration Test on June 22, 2011
Generation
Pressure
Kg/cm2
5.7
7.1
8.8
11.2
11.4
11.4
11.5
11.5
11.6
11.7
12.5
10.6667
T-cond
Co
28
28.7
29.3
29.8
30
30.5
31
31.5
32
32.6
33
30.6
Refrigeration
Time min
Evaporator-temp. Co
Absorber- temp. Co
Abso –press. kg/cm2
T-out
Co
28.5
29.2
30
30.5
31
31.6
32
32.8
33.2
33.8
34
31.5
0
35
39
6.5
15
30
39
6.1
T- in
Co
28
28
28.2
28.5
28.8
29.5
30
30.5
30.5
30.8
31
29.436
30
25
39
5.8
45
18
39
5.2
T-solu
Co
45
50
58
70
74
77
80
84
86
86
91
74.25
60
10.5
37
4.6
Fi Co
T-p Co
42
47
55
67
70
74
78
82
84
84
86
71.3333
52
56
63
76
81
83
86
89
93
96
95
80.4167
75
7.5
37
4.2
90
6.2
37
3.8
72
105
5
37
3.4
120
4.3
35
3.6
wind
m/s
4.1
4.1
4.6
4.6
5.7
5.7
6.7
6.7
5.7
5.7
4.6
5.23333
135
3.6
35
4.3
Amb temp
Co
29
29
31
32
33
34
35
36
37
37
38
34.0833
150
3.8
35
4.9
165
4.1
35
5.2
180
4.4
33
5.2
radiation
W/m2
721.14
802.63
884.12
913.88
943.64
973.41
1003.18
973.41
943.64
913.88
884.12
896.64
195
4.9
33
5.4
210
5.3
33
5.6
Day hours
9:00 AM
9:30 AM
10:00 AM
10:30 AM
11:00 AM
11:30 AM
12:00 PM
12:30 PM
1:00 PM
1:30 PM
2:00 PM
Ave
225
5.7
33
5.9
240
6.3
32
6
APPENDIX .H
Observation during Generation and Refrigeration Test on June 26, 2011
Generation
Pressure T-cond oC
Kg/cm2
5.5
28
7.1
28.8
9
29.5
10.8
29.8
11.4
30.6
11.4
31.3
11.6
32
11.7
32.2
11.8
32.5
12
32.8
11.5
34.4
10.7667
31.1
T- out
o
C
28
29
29.8
30
30.8
31.6
32
32.6
33
34
35
31.4
Refrigeration
Time min
Evaporator-temp. Co
Absorber- temp. Co
Abso –press. kg/cm2
0
34
38
6.7
15
28
38
6.4
T- in
o
C
26
26
26.6
26.8
27.2
27.6
27.8
28
28.4
28.6
29.4
27.491
30
21
38
5.9
T-solu
o
C
45
51
60
68
75
81
84
87
92
91
90
75.9167
45
14
38
5.5
60
9
37
4.8
Fi Co
T-p oC
41
58
57
64
72
78
81
84
88
89
87
73.6667
51
56
65
74
82
86
90
93
95
95
94
81.0833
75
6.2
37
4.4
90
5.1
37
5
73
105
4.7
37
4.5
wind
m/s
2.1
2.1
3.1
4.1
4.1
5.7
5.7
5.1
5.1
4.1
4.1
4.11667
120
4
35
3.9
135
3.2
35
3.5
Amb temp
Co
31
32
33
34
35
35
35
36
36
37
37
34.9167
150
3.4
35
3.5
165
3.7
35
3.8
180
3.9
33
4.3
radiation
W/m2
720.8
802.33
883.86
913.665
943.47
973.275
1003.08
973.275
943.47
913.665
883.86
896.4233
195
4.2
33
4.8
210
4.5
33
5.2
Day hours
9:00 AM
9:30 AM
10:00 AM
10:30 AM
11:00 AM
11:30 AM
12:00 PM
12:30 PM
1:00 PM
1:30 PM
2:00 PM
Ave
225
4.8
33
5.5
240
5.5
30
5.5
APPENDIX .I
Diagram of the mixture ammonia-water properties
74
APPENDIX .J
Observation during Refrigeration Test on June 16, 2011
Absorber Temperature
Absorber pressure
8
7
6
5
4
3
2
1
0
44
40
36
32
28
24
20
16
12
8
4
0
0
pressure kg/cm2
Temperature °C
Evaporator Temperature
30 60 90 120 150 180 210 240
Time "min"
Observation during Generation Test on June 22, 2011
T-condenser
Pressure
T-generator
120
12
100
10
80
8
60
6
40
4
20
2
0
0
9 AM
10 AM 11 AM 12 PM
Time
75
1 PM
2 PM
Pressuer kg/m2
Temprature °C
T-plate
Observation during Generation Test on June 22, 2011
T-plate
T-condenser
solar total
1200
1000
800
600
400
200
0
Temperature Co
120
100
80
60
40
20
0
9 AM
10 AM 11 AM 12 PM
1 PM
solar radiation W/m2
T-generator
2 PM
Time
Observation during Refrigeration Test on June 22, 2011
240 210 180 150 120 90 60 30
Time "min"
76
0
o
44
40
36
32
28
24
20
16
12
8
4
0
C
Absorber pressure
Pressure kg/cm2
Absorber Temperature
Temperature
Evaporator Temperature
7
6
5
4
3
2
1
0
Observation during Generation Test on June 25, 2011
120
T-generator
T-plate
T-condenser
Pressure
14
12
10
8
6
4
2
0
Pressure kg/cm2
100
Temperature Co
80
60
40
20
0
9 AM 10 AM 11 AM 12 PM 1 PM
Time
2 PM
Observation during Generation Test on June 25, 2011
T-plate
T-condenser
solar total
1200
100
1000
80
800
60
600
40
400
20
200
Temperature Co
120
0
0
9AM 10AM 11AM 12PM
Time
77
1PM
2PM
solar radiation W/m2
T-generator
Observation during Refrigeration Test on June 25, 2011
7
Absorber Temperature
Absorber pressure
44
40
36
32
28
24
20
16
12
8
4
0
Pressure kg/cm2
6
5
4
3
2
1
0
240 210 180 150 120 90
Time - min
60
30
0
Observation during Generation Test on June 26, 2011
T-plate
T-condenser
Pressure
Temperature Co
100
14
12
10
8
6
4
2
0
80
60
40
20
0
9AM
10AM
11AM
12PM
Time
78
1PM
2PM
Pressure kg/cm2
T-generator
Temperature Co
Evaporator Temperature
Observation during Generation Test on June 26, 2011
T-plate
T-condenser
solar total
100
90
80
70
60
50
40
30
20
10
0
1200
Temperature Co
1000
800
600
400
200
solar radiation W/m2
T-generator
0
9 AM 10 AM 11 AM 12 PM
1 PM
2 PM
Time
Observation during Refrigeration Test on June 26, 2011
Absorber Temperature
Absorber pressure
44
40
36
32
28
24
20
16
12
8
4
0
Pressure kg/cm2
6
4
2
0
240 210 180 150 120 90
Time - min
79
60
30
0
Temperature Co
Evaporator Temperature
8
Observation during Generation Test on June 28, 2011
T-generator
T-plate
T-condenser
Pressure
12
10
8
6
4
2
0
Pressure kg/cm2
Temperature oC
100
80
60
40
20
0
9 AM
10 AM 11 AM 12 PM
1 PM
2 PM
Time hours
Observation during Generation Test on June 28, 2011
80
T-plate
T-condenser
solar total
1200
100
90
80
70
60
50
40
30
20
10
0
Temperature oC
1000
800
600
400
200
0
9 AM
10 AM
11 AM 12 PM
Time
1 PM
2 PM
Observation during Refrigeration Test on June 28, 2011
Absorber Temperature
Absorber44
pressure
40
36
32
28
24
20
16
12
8
4
0
6
5
4
3
2
1
0
240 210 180 150 120 90 60 30
Time - min
81
0
Temperature Co
Pressure kg/cm2
Evaporator
Temperature
7
solar radiation W/m2
T-generator
APPENDIX .K
Actual and Theoretical system cycles for test on June 26, 2011
Temperature °C
11 k
g/cm
²
12 k
g/cm
²
Actual Cycle
4.8
kg
/cm
²
3.4
kg
/cm
²
Theoretical Cycle
Concentration, kg (NH3)/kg(NH3 + H2O)
82
Actual and Theoretical system cycles for test on June 25, 2011.
Temperature °C
11 k
g/cm
²
12.5
kg/c
m²
Actual Cycle
3.5
kg
/cm
²
4.8
kg
/cm
²
Theoretical Cycle
Concentration, kg (NH3)/kg(NH3 + H2O)
83
Actual and Theoretical system cycles for test on June 22,
2011
Temperature °C
11 k
g/cm
²
12 k
g/cm
²
Actual Cycle
3.4
kg
/cm
²
4.8
kg
/cm
²
Theoretical Cycle
Concentration, kg (NH3)/kg(NH3 + H2O)
84
Actual and Theoretical system cycles for test on June 28, 2011
Temperature °C
11 k
g/cm
²
12 k
g/cm
²
Actual Cycle
3.4
4.8
kg
/cm
²
kg
/cm
²
Theoretical Cycle
Concentration, kg (NH3)/kg(NH3 + H2O)
85
‫تطوير نموذج للتبريد بالطاقة الشمسية‬
‫ملخص الدراسة‬
‫تتعرض المحاصيل الزراعية للتلف بعد حصادها‪ ،‬لنقص االهتمام باإلجراءات‬
‫المناسبة من معامالت ما بعد الحصاد مثل التبريد المبدئى وخفض درجة ح اررتها‬
‫بعد حصادها مباشرة‪ .‬مما يتسبب في خسارة في إنتاجية الخضروات والفاكهة‬
‫والموالح بجمهورية مصر العربية‪ ،‬قدر بنحو ‪ 32،33‬مليار جنيه سنويا طبقا‬
‫إلحصائيات العام ‪.5122‬‬
‫وقد أرجع نقص االهتمام بإجراءات ما بعد الحصاد الح اررية الرتفاع تكلفة‬
‫العمليات المرتبطة بها من أجهزة ومتطلبات الطاقة وما يتبعه من زيادة تكاليف‬
‫اإلنتاج أو لغياب مصادر الطاقة المالئمة بالمناطق الزراعية وخصوصا النائية‬
‫منها‪ ،‬مما يحد من خطط التوسع فى الزراعات المستقبلية‪.‬‬
‫لذا عمدت الدراسة على تطوير نموذج للمعاملة الح اررية بالتبريد يستمد قدرته‬
‫من الطاقة الشمسية بواسطة مجمع شمسى مسطح‪ ،‬بحيث يتميز بسهولة اإلنشاء‬
‫والصيانة مع توافر المواد والتكنولوجيات الخاصة بها محليا‪ ،‬واستخدام الطاقة‬
‫الشمسية (الطاقة الخضراء) كمصدر للطاقة والرتباطها على نحو وثيق بمتطلبات‬
‫طاقة التبريد المطلوبة‪ .‬فكلما زادت طاقة اإلشعاع الشمسى وح اررة المنتج الز ارعى‪،‬‬
‫زاد االحتياج للتبريد وزادت كفاءة تشغيل النموذج‪ ،‬وفى ظروف انخفاض معدالت‬
‫التعرض الشمسى يقل االحتياج للتبريد‪ ،‬وما يتبعه من انخفاض كفاءة النموذج‬
‫بما ال يؤثر على كفاءة عمليات ما بعد الحصاد الح اررية إجماالً‪.‬‬
‫‪1‬‬
‫بلغ معامل كفاءة المبرد الشمسى القصوى ‪ 1.20‬بما يعادل ‪ 1.13‬طن تبريد‪.‬‬
‫تحت شدة إشعاع شمسى بلغ ‪ 6157.59‬وات‪/‬م‪ 5‬لليوم وحركة هواء ‪ 2.3‬م‪/‬ث‪.‬‬
‫وقد تم تخفيض درجة ح اررة المبخر إلى ثالثة درجات ونصف الدرجة عند نفس‬
‫ظروف التشغيل‪.‬‬
‫والتي تمكن نظام التبريد وحسب المتبع من الحسابات المرجعية لتبريد‬
‫الحاصالت الزراعية‪ ،‬من تبريد ثالثة كيلوجرامات من درنات البطاطس من ‪°31‬م‬
‫إلى ‪°0‬م‪ ،‬أو ما يعادل ثمانية عشرة كيلوجرامات من درنات البطاطس إلى ‪°51‬م‬
‫عند نفس ظروف التشغيل‪.‬‬
‫تكون نظام التشغيل من مولد حرارى يتكون من جزيئين هما المجمع الشمسى‪،‬‬
‫ووحدة استيعاب سائل التشغيل المسخن باإلضافة إلى أنبوبتين معدنيتين‬
‫متداخلتين تعمالن كمكثف ومبخر على التوالى‪ .‬حيث تعمل األنبوبة الداخلية‬
‫عمل المكثف وتملئ األنبوبة الخارجية بالماء الجارى لتتم عملية التبريد‪ .‬وفى‬
‫مرحلة التبخير‪ ،‬يتم منع دخول الماء إليقاف عمليات التبريد المائية إتاحة الفرصة‬
‫لعمل المبخر في خفض درجة ح اررة المنتج الغذائي‪.‬‬
‫تكون سائل التشغيل من مخلوط من األمونيا‪-‬ماء بتركيز ‪ .%21‬وعلى الرغم‬
‫من المميزات المرغوبة لهذا الخليط‪ ،‬إال انه يعاب عليه سميته واحداث تأكل في‬
‫معظم المواد المنشأة لوحدات التبريد التقليدية المصنعة من النحاس‪ .‬وباستخدام‬
‫الوحدة في ظروف الحقل المفتوح وانشاء المكونات من معدن الحديد تم حل‬
‫مشكالت التشغيل‪.‬‬
‫‪2‬‬
‫استخدمت وحدة تجميع شمسى من نوع المجمعات المسطحة بمساحة‬
‫‪2،62‬م‪ ،5‬وقد بلغ متوسط كفاءة الوحدة الشمسية البصرية ‪ ،%59‬ومتوسط الكفاءة‬
‫اليومية نحو ‪( %23‬تحت الظروف االختبارية)‪ .‬بلغ معامل فقد الح اررة من المجمع‬
‫الشمسى المسطح ‪ 3.722‬وات‪/‬م‪° 5‬م‪ .‬ومعامل السخان الشمسى ‪1.3672‬‬
‫بحاصل يبلغ ‪ 3.2‬وات‪/‬م‪° 5‬م‪ ،‬في حين تقع النسبة العيارية لهذه النوعية من‬
‫المجمعات الشمسية في المدى بين ‪ 3-3‬وات‪/‬م‪° 5‬م‪.‬‬
‫نظ اًر لتأثير سرعة الهواء حول المجمع الشمسى على كفاءته في استخالص‬
‫الطاقة الح اررية من الطاقة الشمسية ‪-‬وبالتالى على كفاءة عملية التبريد‪-‬حددت‬
‫ثالثة معدالت للفقد الحراري من المجمع والذى سجل فقد مقداره ‪ 5.1‬ك‪.‬وات‪.‬س‬
‫للهواء للسرعة التى تقع في المدى من ‪ 2‬إلى ‪ 0‬كم‪/‬س‪ ،‬وزاد معدل الفقد الحرارى‬
‫ليبلغ ‪ 3.1‬ك‪.‬وات‪.‬س بزيادة سرعة الهواء في المدى من ‪ 0‬إلى ‪ 52‬كم‪/‬س‪،‬‬
‫وبمعدل فقد حرارى ‪ 2.5‬ك‪.‬وات‪.‬س للهواء المتحرك بسرعة واقعة فى المدى من‬
‫‪ 52‬إلى ‪ 31‬كم‪/‬س من خالل مسطح المجمع الشمسى بالمساحة المعرضة‬
‫‪ 2.62‬م‪ .5‬وطبقا للنتائج‪ ،‬ينصح باختيار موقع وحدة التبريد فى مكان مناسب ‪-‬‬
‫لقابلية النموذج للنقل – على مسافة من الغطاء النباتى أو المنشآت الزراعية حتى‬
‫تعمل كمصد للرياح‪ ،‬األمر الذي يعمل على اتزان عمل النموذج خالل إجراء‬
‫المعامالت الح اررية‪.‬‬
‫كما روعى أضفاء القدرة على نقل النموذج أثناء عمله بالحقل اإلنتاجى‪ ،‬مما‬
‫يضمن له الحصول على الطاقة الشمسية واجراء المعاملة الح اررية آنياً و حصاد‬
‫المنتج الزراعى‪ .‬مكن استخدام طبقتين من األغطية البالستيكية الشفافة أعلى‬
‫‪3‬‬
‫سطح المجمع الشمسى من خفض تكاليف اإلنشاء‪ ،‬وتقليل أخطار تعرضه للكسر‬
‫أو الشرخ عند صيانة أو نقل النموذج‪ .‬أال أن النموذج المختبر كان أقل كفاءة‪،‬‬
‫وتدنت نسبة الطاقة الشمسية المستخلصة‪ .‬ويجب أن يراعى التحول إلى أنظمة‬
‫التركيز الشمسية المختلفة عند الرغبة في زيادة الكفاءة أو زيادة السعة الكلية‬
‫للمبرد‪.‬‬
‫بلغت كفاءة وحدة االستيعاب الحديدية لسائل التشغيل بالمولد نحو ‪،%63‬‬
‫بما يساوى ‪ %25‬من إجمالي طاقة األشعة الشمسية في اليوم‪ .‬وقد أمكن لوحدة‬
‫االستيعاب من تحمل ضغوط التشغيل الواقعة في المدى ‪ 3.0‬إلى‬
‫‪ 25.2‬كيلوجرام‪/‬سم‪ ،5‬وتالشى أخطار التآكل الناجمة عن تفاعل األمونيا مع‬
‫معدن النحاس‪ .‬كما استخدمت وحدة تصحيح أعلى وحدة االستيعاب لمنع المياه‬
‫من الوصول إلى المكثف‪.‬‬
‫بلغت كفاءة وحدة التكثيف – تبخير الح اررية ‪ %22‬من الطاقة الواردة إليها‪،‬‬
‫وما يساوى ‪ % 0.0‬من إجمالي الطاقة الشمسية المجمعة في اليوم الواحد‪.‬‬
‫يساعد النظام المختبر على الحد من الخسائر الناجمة لتدهور المنتجات‬
‫الزراعية نتيجة غياب عمليات ما بعد الحصاد وخصوصا إزالة الح اررة الحقلية‪،‬‬
‫كما يمكن االستفادة المزدوجة من النظام للحصول على الماء الساخن وظروف‬
‫التبريد للتطبيقات الزراعية المختلفة‪.‬‬
‫‪4‬‬
‫تطوير نموذج للتبريد بالطاقة الشمسية‬
‫رسالة مقدمة من‬
‫رجب قاسم محمود على‬
‫بكالوريوس العلوم الزراعية – كلية الزراعة ‪ -‬جامعة األزهر‪ -‬أسيوط – ‪5002‬م‬
‫استيفاء لمتطلبات الحصول على درجة‬
‫التخصص (الماجستير)‬
‫فى العلوم الزراعية (الهندسة الزراعية)‬
‫قسم الهندسة الزراعية‬
‫كلية الزراعة ‪ -‬جامعة األزهر‬
‫فرع أسيوط‬
‫‪1131 8‬هـ‬
‫‪ 2113‬مـ‬
‫‪5‬‬
‫تطوير نموذج للتبريد بالطاقة الشمسية‬
‫رسالة مقدمة من‬
‫رجب قاسم محمود علي‬
‫بكالوريوس في العلوم الزراعية (هندسة زراعية) كلية الزراعة بأسيوط ‪ -‬جامعة األزهر ‪5002‬م‬
‫إستيفاء لمتطلبات الحصول على درجة‬
‫التخصص (الماجستير)‬
‫فى العلوم الزراعية (هندسة زراعية)‬
‫قسم الهندسة الزراعية‬
‫كلية الزراعة بأسيوط ‪ -‬جامعة األزهر‬
‫‪1131‬ه‬
‫‪2113‬م‬
‫أجــــازهــا‪:‬‬
‫أ‪.‬د‪.‬م‪ /‬محمد نبيل محمد عبد العظيم العوضى‪.......................................‬‬
‫أستاذ الهندســة الزراعيــة غير المتفرغ ‪ -‬كليـة الزراعة ‪ -‬جامعـة عين شمس‪.‬‬
‫أ‪.‬د‪ /‬سـمير أحمد محمد طايل‪...........................................................‬‬
‫أستاذ الهندســة الزراعيــة المتفرغ ‪ -‬كليـة الهندسة الزراعية ‪ -‬جامعـة األزهر بالقاهرة‪.‬‬
‫أ‪.‬د‪ /‬أحمد ماهر محمد الليثى‪...........................................................‬‬
‫أستاذ الهندســة الزراعيــة ‪ -‬كليـة الزراعة ‪ -‬جامعـة األزهر بأسيوط‬
‫د‪/‬محمود زكى العطار‪...................................................................‬‬
‫مدرس الهندســة الزراعيــة ‪-‬كليـة الزراعة – جامعـة عين شمس‪-‬القاهرة‬
‫تاريخ المناقشة‪ 2113 / 9/ 22 :‬م‬
‫‪6‬‬
‫‪ 1131 / 11 / 11‬هـ‬
‫تطوير نموذج للتبريد بالطاقة الشمسية‬
‫رسالة مقدمة من‬
‫رجب قاسم محمود على‬
‫بكالوريوس العلوم الزراعية– كلية الزراعة ‪ -‬جامعة األزهر‪ -‬أسيوط – ‪5002‬م‬
‫استيفاء لمتطلبات الحصول على درجة‬
‫التخصص (الماجستير)‬
‫فى العلوم الزراعية (الهندسة الزراعية)‬
‫قسم الهندسة الزراعية‬
‫كلية الزراعة‪ -‬جامعة األزهر‬
‫فرع أسيوط‬
‫‪1131‬ه‬
‫‪2113‬م‬
‫لجنة اإلشراف‪:‬‬
‫ا‪.‬د‪ /‬حسن عبد الرازق عبد المولى‪..................................................‬‬
‫أستاذ ورئيس قسم الهندســة الزراعية ‪ -‬كلية الزراعة – جامعـة األزهر‪ -‬أسيوط‪.‬‬
‫ا‪.‬د‪/‬أحمد ماهر محمد الليثي‪.............................................................‬‬
‫‪.L‬‬
‫أستاذ الهندســة الزراعيــة ‪ -‬كليـة الزراعة – جامعـة األزهر‪ -‬أسيوط‪.‬‬
‫د‪/‬محمود زكى العطار‪...................................................................‬‬
‫مدرس الهندســة الزراعيــة‪ -‬كليـة الزراعة – جامعـة عين شمس‪-‬القاهرة‪.‬‬
‫‪7‬‬