DOI : 10.17577/IJERTV2IS120818
- Open Access
- Total Downloads : 28095
- Authors : P. Sarat Babu, Prof. N. Hari Babu
- Paper ID : IJERTV2IS120818
- Volume & Issue : Volume 02, Issue 12 (December 2013)
- Published (First Online): 18-12-2013
- ISSN (Online) : 2278-0181
- Publisher Name : IJERT
- License:
This work is licensed under a Creative Commons Attribution 4.0 International License
Experimental Study of A Domestic Refrigerator/Freezer Using Variable Condenser Length
P. Sarat Babu1, Prof. N. Hari Babu2
1M.E. (Thermal Engg) Student, Aditya Institute of Technology & Management, Tekkali.
2Prof & HOD., Department of Mechanical Engg. Aditya Institute of Technology & Management, Tekkali.
532201, Srikakulam Dist.
Abstract The condenser design plays a very important role in the performance of a vapour compression refrigeration system. Optimized design is possible through theoretical calculations, however may fail due to the reason that the uncertainties in the formulation of heat transfer from the refrigerant inside the condenser tubes to the ambient air. Hence experimental investigations are the best in terms of optimization of certain design parameters.
In my experimental work, it is proposed to optimize condenser length for domestic refrigerator of 165 litres capacity. It may give a chance to find a different length other than existing length will give better performance and concluded that the optimum length of coil is 7.01m
Keywords Vapour Compression Refrigeration System, Refrigerant, Optimized design.
-
INTRODUCTION
The first mechanically produced cooling system was developed in England in 1834. The process later became known as vapour compression. After availability of electricity automatic refrigeration system was developed in 1897. Basically a refrigeration or air conditioning is nothing more than a heat pump whose job is to remove heat from a lower temperature source and reject heat to high temperature sink. The Vapour Compression Refrigeration Cycle is a process that cools an enclosed space to a temperature lower than the surroundings. To accomplish this, heat must be removed from
the enclosed space and dissipated into the surroundings. However, heat tends to flow from an area of high temperature to that of a lower temperature.
During the cycle refrigerant circulates continuously through four stages. The first stage is called Evaporation and it is here that the refrigerant cools the enclosed space by absorbing heat. Next, during the Compression stage, the pressure of the refrigerant is increased, which raises the temperature above that of the surroundings. As this hot refrigerant moves through the next stage, Condensation, the natural direction of heat flow allows the release of energy into the surrounding air. Finally, during the Expansion phase, the refrigerant temperature is lowered by what is called the auto refrigeration effect. This cold refrigerant then begins The Evaporation stage again, removing more heat from the enclosed space. Each of the four stages will now be revisited in detail, explaining the physical changes that occur in the refrigerant and the devices used to accomplish these changes. A visual representation of the cycle is displayed below with the explanation of each stage.
A. Refrigeration System: Working Principle and Construction
Refrigeration system is based upon the Clausius statement of second law of thermodynamics. This statement shows, It is impossible to construct a device which, operating in a cycle, will produce no affect other than the transfer of heat from a cooler to a hotter body. The construction of
vapour compression refrigeration system is illustrated in figure 1. A vapour compression cycle is used in most household refrigerators, refrigeratorfreezers and freezers. In this cycle, a circulating refrigerant such as R134a enters a compressor as low-pressure vapour at or slightly above the temperature of the refrigerator interior. The vapour is compressed and exits the compressor as high-pressure superheated vapour. The superheated vapour travels under pressure through coils or tubes comprising "the condenser", which are passively cooled by exposure to air in the room. The condenser cools the vapour, which liquefies. As the refrigerant leaves the condenser, it is still under pressure but is now only slightly above room temperature. This liquid refrigerant is forced through a metering or throttling device, also known as an expansion valve (essentially a pin-hole sized constriction in the tubing) to an area of much lower pressure. The sudden decrease in pressure results in explosive-like flash evaporation of a portion (typically about half) of the liquid.
The latent heat absorbed by this flash evaporation is drawn mostly from adjacent still-liquid refrigerant, a phenomenon known as "auto-refrigeration". This cold and partially vaporized refrigerant continues through the coils or tubes of the evaporator unit. A fan blows air from the refrigerator or freezer compartment ("box air") across these coils or tubes and the refrigerant completely vaporizes, drawing further latent heat from the box air. This cooled air is returned to the refrigerator or freezer compartment, and so keeps the box air cold. Note that the cool air in the refrigerator or freezer is still warmer than the refrigerant in the evaporator. Refrigerant leaves the evaporator, now fully vaporized and slightly heated, and returns to the compressor inlet to continue the cycle.
Figure 1: Basic cycle of domestic refrigeration system
Figure 2 : T-S Diagram for the Ideal Vapor Compression Refrigeration Cycle
Figure 3: Pressure-enthalpy graph for vapour compression refrigeration system
Process 12: Isentropic compression in compressor.
Process 23: Constant pressure heat rejection in condenser.
Process 34: Isenthalpic expansion in expansion device.
Process 41: Constant pressure heat absorption in evaporator.
-
LITERATURE REVIEW
R. Cabello, E. Torrella and J. Navarro-Esbri [1], have analyzed the performance of a vapour compression refrigeration system using three different working fluids (R134a, R407c and R22). The operating variables are the evaporating pressure, condensing pressure and degree of superheating at the compressor inlet. They analyzed that the power consumption decreases when compression ratio increases with R22 than using the other working fluids.
B.O. Bolaji et al[2] investigated experimentally the performances of three ozone friendly Hydrofluorocarbon (HFC) refrigerants R12, R152a and R134a. R152a refrigerant found as a drop in replacement for R134a in vapour compression system.
B.O. Bolaji[3] discussed the process of selecting environmental-friendly refrigerants that have zero ozone depletion potential and low global warming potential. R23 and R32 from methane derivatives and R152a, R143a, R134a and R125 from ethane derivatives are the emerging refrigerants that are non toxic, have low flammability and environmental-friendly. These refrigerants need theoretical and experimental analysis to investigate their performance in the system.
James M. Calm [4], has studied the emission and environmental impacts of R11, R123, R134a due to leakage from centrifugal chiller system. He also investigated the total impact in form of TEWI and change in system efficiency or performance due to charge loss. He also summarized the methods to reduce the refrigerant losses by the system like design modifications, improvement in preventive maintenance techniques, use of purge system for
refrigerant vapour recovery, servicing and lubricant changing in system.
Samira Benhadid-Dib and Ahmed Benzaoui [5], have showed that the uses of halogenated refrigerants are harmful for environment and the use of "natural" refrigerants become a possible solution. Here natural refrigerants are used as an alternative solution to replace halogenated refrigerants. The solutin to the environmental impacts of refrigerant gases by a gas which contains no chlorine no fluorine and does not reject any CO2 emissions in the atmosphere. The researchers showed that emissions have bad effects on our environment. They also concerned by a contribution to the reduction of greenhouse gases and by the replacement of the polluting cooling fluids (HCFC).
Eric Granryd [6], has enlisted the different hydrocarbons as working medium in refrigeration system. He studied the different safety standards related to these refrigerants. He showed the properties of hydrocarbons (i.e. no ODP and negligible GWP) that make them interesting refrigeratingalternatives for energy efficient and environmentally friendly. But safety precautions due to flammability must be seriously taken into account.
Y. S. Lee and C. C. Su [7], have studied the performance of VCRS with isobutene and compare the results with R12 and R22. They used R600a about 150 g and set the refrigeration temperature about 4 °C and -10
°C to maintain the situation of cold storage and freezing applications. They used 0.7 mm internal diameter and 4 to 4.5 m length of capillary tube for cold storage applications and 0.6 mm internal diameter and 4.5 to 5 m length of capillary tube for freezing applications.
They observed that the COP lies between
1.2 and 4.5 in cold storage applications and between 0.8 and 3.5 in freezing applications. They also observed that the system with two
capillary tubes in parallel performs better in the cold storage and air conditioning applications, whereas that with a single tube is suitable in the freezing applications.
-
EXPERIMENTAL STUDY
-
Calculations and Analysis with Existing Dimensions
Condenser Sizes
Length 6.1m (20feet)
Diameter 6.4mm Evaporator Sizes Length 7.26m
Diameter 6.4mm Capillary tube Sizes Length 3.35m
Voltage = 230 Volts
From pressure enthalpy Chart for r 134a, enthalpy values at state points 1, 2, 3, 4. The state points are fixed using pressure and temperature and each point.
h 1 = 426.510 KJ/Kg p = 453.88 KJ/Kg p = 263.284 KJ/Kg h4 = 199.38 KJ/Kg
Calculations Performance Parameters
-
Net Refrigerating Effect (NRE)
= (p h4)
Diameter 0.8mm Ambient Temperature = 31.50
Temperatures
Compressor suction temperature T
C
= 29.50 C
= (426.51 199.38)
= 227.13 KJ/Kg
-
Circulating rate to obtain one tone of Refrigeration, kg/min.
1
Compressor Discharge Temperature T2 = 72.2 0 C
Condensing Temperature T3 = 39.4 0 C Evaporator Temperature T4 = 0.10 C Pressure
Compressor Suction pressure P1= 16.5 psi Compressor Discharge Pressure p2= 165 psi Condensing Pressure P3= 156 psi Evaporator pressure P4= 17 psi
Convert all the pressure in to Bar Conversion pressure Unit -1.psi = 0.069 bar P1 =16.5 x 0.06 = 1.13 bar
P2 =165 x 0.069 = 11.38 bar
P3 = 156 x 0.069 = 10.76 bar
P4 = 17 x 0.069 = 1.173 bar
Current and Voltage Current = 1.1 Amps
m r= 210/NRE = 210/227
= 0.924Kg/min
-
Heat of compression = (p p )
= (453.88 426.51)
= 27.37 KJ/Kg
-
Heat Equivalent of work of compressor = m r x ( p p )
= 0.924x (27.37)
=25.31kJ/min
-
Compressor power =(25.31/60 )
=0.422 kW
-
Coefficient of performance (COP) Net refrigerating Effect
= ———————————
Heat of Compression
= (227.13/27.37) =8.298
-
Heat rejected in condenser
= (p p )
= (453.88-263.284) =190.59
-
Heat rejection Rate
= (210/227.13) x 190.59 =176.105
-
Heat rejection factor = (176.105/210) =0.839
-
Specific volume of suction gas Vs = 0.19 m3/Kg
-
Volume of refrigerant to be handled bycompressor V=mrxVs=0.927×0.19=0.176 m3/min
-
Compression Pressure Ratio=
Pd/Ps
=11.38/1.13=10 bar
-
-
Calculations and Analysis with Varying Dimensions (Decreased Condenser Length) Condenser Sizes
Length 5.1m (17feet)
Diameter 6.4mm Evaporator Sizes Length 7.26m
Diameter 6.4mm Capillary tube Sizes Length 3.35m
Diameter 0.8mm
Ambient Temperature = 31.50 C
Temperatures
Compressor suction temperature T 1= 29.10 C Compressor Discharge Temperature T2 = 70.2 0 C Condensing Temperature T3 = 41.2 0 C Evaporator Temperature T4 = 0.80 C Pressure
Compressor Suction pressure P1= 16 psi Compressor Discharge Pressure p2= 162 psi Condensing Pressure P3= 157 psi Evaporator pressure P4= 21 psi
Convert all the pressure in to Bar Conversion pressure Unit -1.psi = 0.069 bar P1 =16 x 0.06 = 1.104 bar
P2 =162 x 0.069 = 11.17 bar
P3 = 157 x 0.069 = 10.83 bar
P4 = 21 x 0.069 = 1.44 bar
Current and Voltage Current = 1.1 Amps Voltage = 230 Volts
From pressure enthalpy Chart for r 134a, enthalpy values at state points 1, 2, 3, 4. The state points are fixed using pressure and temperature and each point.
h 1 = 426.190 KJ/Kg p = 451.567 KJ/Kg p = 260.560 KJ/Kg h4 = 200.566 KJ/Kg
Calculations Performance Parameters
-
Net Refrigerating Effect (NRE)
= (p h4)
= (426.19 200.566)
= 225.624 KJ/Kg
-
Circulating rate to obtain one tone of Refrigeration, kg/min.
m r= 210/NRE = 210/225.624
= 0.931Kg/min
-
Heat of compression = (p p )
= (451.567 426.19)
= 25.377 KJ/Kg
-
Heat Equivalent of work of compressor = m r x ( p p )
= 0.931x (25.377)
=23.619kJ/min
-
Compressor power =(23.619/60 )
=0.394 kW
-
Coefficient of performance (COP) Net refrigerating Effect
= ———————————
Heat of Compression
= (225.624/25.377) =8.89
-
Heat rejected in condenser
= (p p )
=(451.567-260.565) =191.007 kJ/Kg
-
Heat rejection Rate
= (210/225.624) x 191.007 =177.78
-
Heat rejection factor = (177.78/210) =0.846
-
Specific volume of suction gas Vs = 0.19 m3/Kg
-
Volume of refrigerant to be handled bycompressor V=mrxVs=0.931×0.19=0.177 m3/min
-
Compression Pressure Ratio=
Pd/Ps
=11.17/1.104=10.117 bar
-
-
Calculations and Analysis with Varying Dimensions (Increased Condenser Length)
Condenser Sizes
Length 7.01m (23feet)
Diameter 6.4mm Evaporator Sizes Length 7.26m
Diameter 6.4mm Capillary tube Sizes Length 3.35m
Diameter 0.8mm
Ambient Temperature = 31.50 C
Temperatures
Compressor suction temperature T 1= 36.50 C Compressor Discharge Temperature T2 = 75.6 0 C Condensing Temperature T3 = 38.5 0 C Evaporator Temperature T4 = -4.50 C Pressure
Compressor Suction pressure P1= 9 psi Compressor Discharge Pressure p2= 157 psi Condensing Pressure P3= 149 psi Evaporator pressure P4= 13 psi
Convert all the pressure in to Bar Conversion pressure Unit -1.psi = 0.069 bar P1 =9 x 0.069 = 0.621 bar
P2 =157 x 0.069 = 10.83 bar
P3 = 149 x 0.069 = 10.30 bar
P4 = 13 x 0.069 = 0.897 bar
Current and Voltage Current = 1.1 Amps Voltage = 230 Volts
From pressure enthalpy Chart for r 134a, enthalpy values at state points 1, 2, 3, 4. The state points are fixed using pressure and temperature and each point.
h 1 = 430.77 KJ/Kg p = 458.676 KJ/Kg p = 258.6 KJ/Kg
h4 = 193.23 KJ/Kg
Calculations Performance Parameters
-
Net Refrigerating Effect (NRE)
= (p h4)
= (430.77 193.23)
= 237.54 KJ/Kg
-
Circulating rate to obtain one tone of Refrigeration, kg/min.
m r= 210/NRE = 210/237.54
= 0.884Kg/min
-
Heat of compression = (p p )
= (458.676 430.77)
= 27.906 KJ/Kg
-
Heat Equivalent of work of compression = m r x ( p p )
= 0.884x (2.906)
=24.667 kJ/min
-
Compressor power =(24.667 /60 )
=0.411 kW
-
Coefficient of performance (COP) Net refrigerating Effect
= ———————————
Heat of Compression
= (237.54/27.906) =8.51
-
Heat rejected in condenser
= (p p )
= (458.676 -258.6) =200.076
-
Heat rejection Rate
= (210/237.54) x 200.076 =176.88
-
Heat rejection factor = (176.88/210) =0.842
-
Specific volume of suction gas Vs = 0.2 m3/Kg
-
Volume of refrigerant to be handled bycompressor V=mrxVs=0.884×0.2=0.177 m3/min
-
Compression Pressure Ratio=
-
P /P
t a = 1.30 C
LMTD = 9.030 C
Air side heat transfer coefficient (h0)
Normally the air velocities over air- cooled condenser are between 2 to 6 m/sec depending upon the application.
Volume of flow = face area x Velocity of flow
Assuming a blower Capacity = 2 m3/ sec Face area = 2/6 = 0.333 m2
Face diameter= [0.333/ (/4)]0.5 = 0.651m Mean temperature of air = 31. 5 + ta / 2
= 31.5+(1.3/2) = 32.150C
At 32.50C properties air at atmospheric
d s
=10 /
17.43 bar
pressure
.83 0.621=
-
-
THEORITICAL STUDY Theoretical calculations for determining of length of condenser for air cooling cross flow type
Size of Condenser Tube
Outer diameter of tube = 6.4 mm Inner diameter of tube = 4.6 mm Condensing temperature= 39.40 C Ambient temperature = 31.50 C For air
= 1.13 kg/m3
µ = 18.93 x10-3 NS / m 2
Cp = 1.009 k J/ Kg
K = 27.04 w/mK
Re =(VD/ µ)air
(1.13 x 6 x 0.651)
= ————————
18.93 x 10 -6
µCp Pr= ———-
Specific heat of air, CP = 1.007 KJ/KgK Density of air = 1.1614 m3/kg
K
18.93 x 10
-6 x 1.009
Condenser design load
= ————————— = 0.706
q condenser = m r
(h 2
p)
27.04 x 10 -3
= (0.924/60) x (453.88 263.284)
= 2.94 = 3 kW
(h0 D/k) = Nu= 0.3 Re
= ( 0.6 x (
0.6 Pr 0.333
0.333
Log mean temperature difference
i – 0
LMTD = ————-
i
In———
0
Temperature rise of air
But m = v
q condenser = mc p t a
3 x 103 = 1.1614 x 2 x 1.007 x ta
0.3 239642) 0.706)
= 440.36
440.36 x 27.04 x 10-3
h 0 = ———————————
0.651
h0 = 18.29 w/m2 k
Condensing heat transfer coefficient (hi)
Flow are of refrigerant through the tube
= (II/4) x (4.6/1000)2
= 1.66 x 10 -5 m2
Re = (VD/µ)
m r Di
= ————-
µi Ai
Refrigerant at properties at 420 C
µ = 0.574 x 10-3 NS/m2
Cp = 1510 J/kg k
K= 73.9 x 10-3 w/m K
(0.924/60) x (4.6/1000)
Re = ——————————— (0.574 x 10-3) x (1.66 x 10-5)
Calculations for Condenser Length
Case: 1
Q = UA (LMTD)
Surface area of fins:
Q 224
= ————- = ——————- = 1.85 m2 Uf (LMTD) 13.408 x (9.03)
Coil outside bare surface area of tube (As) As = (1.85/15) = 0.1233 m2
As 0.1233
Length of tube (L) = ——– = ————-
IID II (6.4/1000)
= 7434
µCp 0.574 x 10-3 x 1510
Case: 2
= 6.13 m
Pr = —— = ——————– = 11.73 K 73.9 x 10-3
(hi D/k) = Nu = 0.026 x Re 0.8 pr 0.4
= 0.026 x (7434)0.8 x (11.73)0.4
= 87
87 x (73.9 x 10-3) hi= ————————
(4.6/1000)
hi = 1398 w/m2 k
Overall heat transfer coefficient
Neglect metal resistance, and scale effect. Since refrigerant to condensing insider the pipe, denoting subscript f for fin;
1 1 1
————– = ———— + ————-
U f A f h 0 A f h f A i
Finned surface to outside bare are a ratio 20 (let us consider)
(1/u f)= (1/ h0) + (1/ h I) x (A f/ A0) (D 0/ Di)
(1/ uf) = (1/18.29)+(1/1398) x (20) x (6.4/4.6)
u f = (1/ 0.0745) = 13.408 w/m2 k
Q = UA (LMTD)
Surface area of fins:
Q 225
= ————- = ——————- = 1.858 m2 Uf (LMTD) 13.408 x (9.03)
Coil outside bare surface area of tube (As) As = (1.858/15) = 0.123 m2
As 0.123
Length of tube (L) = ——– = ————-
IID II (6.4/1000)
= 6.16 m
Case: 3
Q = UA (LMTD)
Surface area of fins:
Q 227
= ————- = ——————- = 1.874 m2 Uf (LMTD) 13.408 x (9.03)
Coil outside bare surface area of tube (As) As = (1.874/15) = 0.125 m2
As 0.124
Length of tube (L) = ——– = ————-
IID II (6.4/1000)
= 6.16 m
Case: 4
Q = UA (LMTD)
Surface area of fins:
Q 237
= ————- = ——————- = 1.957 m2 Uf (LMTD) 13.408 x (9.03)
Coil outside bare surface area of tube (As) As = (1.957/15) = 0.1304 m2
As 0.133
Length of tube (L) = ——– = ————-
IID II (6.4/1000)
= 6.49 m
Case: 5
Q = UA (LMTD)
Surface area of fins:
Q 246
= ————- = —————- = 2.0318 m2 Uf (LMTD) 13.408 x (9.03)
Coil outside bare surface area of tube (As) As = (2.0318/15) = 0.135 m2
As 0.135
Length of tube (L) = ——– = ————-
IID II (6.4/1000)
= 6.74 m
-
RESULTS AND DISCUSSIONS The performance of Vapour
Compression Refrigeration Cycle, operating the experimental domestic refrigerator varies considerably with the length of condenser. The results are plotted on graphs. The relationship between length of condenser and performance parameters have been compared are shown in the following Graphs.
2
2
1 2 3 4 5 6 7 8 9
Length of Condensers
1 2 3 4 5 6 7 8 9
Length of Condensers
Condenser Length Vs COP
Condenser Length Vs COP
C O
P
C O
P
12
12
7
7
Graph 6.1. Lengths of Condenser Vs COP The length of condenser increases COP is
gradually decreases. It starts to increase at
6.1 meter length of condenser.
Net Refrigerating Effect slightly increases as the of condenser length increases [Graph 6.2].
Heat Rejection in Condenser increases as the length of condenser increases. It starts to slightly increase at 6.1 meters length of condenser [Graph 6.3].
Compressor work increases as the length of condenser increases [Graph 6.4].
The mass flow rate of refrigerant decreases as the length of condenser increases. [Graph 6.5].
Heat rejection condenser increases as the length of condenser increases. It decreases at
6.1 meters of condenser [Graph 6.6].
For lower condensing temperature, the length of condenser required is more. As the condensing temperature increases, the requirement of condenser length decreases sharply. Hence condensing temperature influences the length of the condenser [Graph 6.7].
Condenser Length Vs Mass Flow Rate
Condenser Length Vs Mass Flow Rate
0.95
0.94
0.93
0.92
0.91
0.9
0.89
0.88
0.87
0.86
0.85
0.84
1
2Con3den4ser L5eng6th (m7 ts)8
9
0.95
0.94
0.93
0.92
0.91
0.9
0.89
0.88
0.87
0.86
0.85
0.84
1
2Con3den4ser L5eng6th (m7 ts)8
9
Condensser Length Vs NRE
260
240
220
N 200
R 180
E 160
140
120
100
1 2 3 4 5 6 7 8 9
Condenser Length
Condensser Length Vs NRE
260
240
220
N 200
R 180
E 160
140
120
100
1 2 3 4 5 6 7 8 9
Condenser Length
Graph 6.2. Length of Codenser Vs Net RefrigeratingEffect
Condenser Length Vs Heat Rejection in Condenser
Condenser Length Vs Heat Rejection in Condenser
Mass Flow Rate
Mass Flow Rate
Graph 6.5. Length of Condenser Vs Mass flow rate of Refrigerant
Condenser Length Vs Heat Rejection Rate in Condenser
Heat Rejection rater in
Condenser
Heat Rejection rater in
Condenser
179
178
177
176
175
174
173
172
171
170
210
200
190
180
170
160
150
140
130
120
110
100
210
200
190
180
170
160
150
140
130
120
110
100
Heat Rejection in Condenser
Heat Rejection in Condenser
1 2 3 4 5 6 7 8 9
Condenser Length
1 2
1 2
Condenser Length (mts)
Condenser Length (mts)
3 4 5 6 7 8 9
3 4 5 6 7 8 9
10
10
Graph 6.6. Length of Condenser Vs Heat Rejection in Condenser:
Graph 6.3. Length of Condenser Vs Heat Rejection in Condenser
CondenserLength Vs Compressor Work
Condenser Length Vs Condenser Temperature
30
Compressor Work
Compressor Work
25
20
15
10
5
0
1 2 3 4 5 6 7 8 9
Condenser Length (mts)
42
Condenser Temperature
Condenser Temperature
40
38
36
1 2 3 4 5 6 7 8 9
Condenser Length
Graph 6.4. Length of Condenser Vs compressor work.
Graph 6.7. Length of Condenser Vs Condensing Temperature:
-
CONCLUSION
In the present work, the length of the condenser is optimized for a vapour compression refrigeration system used for a domestic refrigeration of 165 Litres capacities, through experimental investigation.
Theoretical computation are also made and compared and found that the optimum length of coil is 7.01 m instead of standard value 6.1m.
-
REFERENCES
-
R. Cabello, E. Torrella, J. Navarro-Esbri, Experimental evaluation of a vapour compression plant performance using R134a, RR407C and R22 as working fluids, Applied Therma Engineering 24 (2004) 1905-1917.
-
B.O.Bolaji, M.A. Akintunde, T.O. Falade, Comparative analysis of performance of three ozone-friends HFC refrigerants in a vapour compression refrigerator, Journal of Sustainable Energy and Environment 2 (2011) 61-64.
-
B.O.Bolaji, Selection of environment- friendly refrigerants and the current alternatives in vapour compression refrigeration systems, Journal of Science and Management, Vol 1, No. 1 (2011) 22-26.
-
James M. Calm, Emissions and environmental impacts from air- conditioning and refrigeration systems, International Journal of Refrigeration 25, pp. 293305, 2002.
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Samira Benhadid-Dib, and Ahmed Benzaoui, Refrigerants and their impact in the environment. Use of the solar energy as the source of energy, Energy Procedia 6, pp. 347352, 2011.
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Eric Granryd, Hydrocarbons as refrigerants – an overview, International Journal of Refrigeration 24, pp. 15-24, 2001.
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Y.S. Lee, and C.C. Su, Experimental studies of isobutene (R600a) as the refrigerant in domestic refrigeration system, Applied Thermal Engineering 22, pp. 507519, 2002.
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Mao-Gang He, Tie-Chen Li, Zhi-Gang Liu, and Ying Zhang, Testing of the mixing refrigerants HFC152a/HFC125 in domestic refrigerator, Applied Thermal Engineering 25, pp. 11691181, 2005.
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