THERMAL SCIENCE: Year 0, Vol., No., pp. -4 PERFORMANCE IMPROVEMENT POTENTIALS OF LOW GLOBAL WARMING POTENTIAL REFRIGERANTS FOR INTERCITY BUS AIR CONDITIONING SYSTEM Saban UNAL a, Cagri KUTLU a, and M. Tair ERDINC a* a Department of Mecanical Engineering, Osmaniye Korkut Ata University, Osmaniye, Turkey Original scientific paper ttps://doi.org/0.9/tsci6070404u In tis study, teoretical investigation of two evaporator ejector refrigeration system was carried out based on energy and exergy analysis using R4a and low global warming potential refrigerants (namely R4yf and R4ze(E)). In order to perform te analyses, a simulation model was developed and ten te influence of different parameters on te COP, COP increase rate and exergy destructions were discussed for eac refrigerant. Te model was validated wit experimental data for R4a and later used to predict te beavior wit R4yf and R4ze(E). It was found tat te igest exergy destruction occurs in te R4yf system and te increase rate in te COP wit respect to te conventional system by using ejector as an expansion device are % for R4a, about 7% for R4yf and about % for R4ze(E), respectively. Key words: ejector, two-evaporator, COP, R4yf, exergy, low GWP Introduction Air-conditioning is important for te transportation industry. Te need is uge, especially for intercity buses. It takes about 0 kw of mecanical energy from te engine. Nearly all buses and minibuses are using air-conditioning systems to ensure passengers comfort recently. Standard bus refrigeration systems ave a significant effect on fuel consumption and exaust gas emissions to te environment because refrigeration system s compressor is connected to te engine saft by a pulley. Standard bus refrigeration systems generally work wit R4a and it is known tat mobile refrigeration system gas leakage plays a big role in global warming. According to te Kyoto protocol [] and te Directive of European Parliament [], mobile refrigeration system studies ave been focused on alternative refrigerants wic ave lower global warming potential. In order to meet global ecological goals, conventional refrigerants sould be replaced by more environmentally friendly ones. Researces ave focused on developing new refrigerants [].Te first candidate of te new refrigerants R4yf as been proposed as a replacement for R4a in mobile air conditioning systems [4]. It as comparable termodynamic properties to R4a, make it ideal as a replacement, as it requires for few or no alterations in order to replace R4a wit R4yf in a pre-designed system []. Researces on R4yf concluded tat te evaporation eat transfer coefficient is nearly same as R4a [6] and condensation eat transfer coefficient of R4yf was lower tan R4a by -% depending on te test conditions [7]. * Corresponding autor, e-mail: memettairerdinc@osmaniye.edu.tr
6 THERMAL SCIENCE: Year 0, Vol., No., pp. -4 Anoter candidate refrigerant R4ze(E) as been proposed in new systems of medium temperature applications [] because te vapor pressure of R4ze(E) is between % and 4% compared wit R4a. Yataganbaba et al. [9] presented exergy analysis of R4yf and R4ze(E) in a two-evaporator refrigeration system. Lastly, Qi [0] investigated te performance improvement potentials of R4yf in mobile air conditioning system under various operation conditions. Wile environment-friendly refrigerant researc is progressing, several studies ave been focused on te recover lost energy in expansion valve by using two-pase ejector instead of te expansion valve in refrigeration systems. Tus te performance of te conventional refrigeration cycle can be improved. Various ejector based refrigeration systems ave been studied wit different working substances. Te ejector-expansion tecnology of replacing te trottling valve was summarized by Sarkar [], Besagni et al. [], and Cen et al. []. A number of teoretical and experimental studies are given in te review papers. Besagni et al. [4] investigated working fluids wic used differently in models and compared to tese models wit a set of experimental data available in te literature, tey sow tat te working fluids ad a great impact on te ejector performance and eac refrigerant ad its own range of operating conditions. On te oter and, it sould be noted tat te ejectors in te mentioned studies operated wit only one evaporation temperature, wic was known as te single-evaporator refrigeration system. In fact, introducing te ejector into te multi-evaporator refrigeration system as been an attractive option since te ejector could be used to maintain te required pressure differences between te ig temperature and low-temperature evaporators and obviously lower te power consumption by its pressure recovery effect []. Li et al. [6] experimentally investigated entrainment and pressure recovery performances of te variable area-ratio ejector applied in te multi-evaporator refrigeration system in detail. Kutlu et al. [7] studied bi-evaporator transcritical ejector refrigeration system and tey reported % COP improvement can be obtained wit using ejector in te R744 transcritical cycle. Unal and Yılmaz [] performed te termodynamic analysis of a bus air-conditioning system enanced wit a two-pase ejector. Te analysis sowed tat te system COP improvement could reac % for proper design. A survey of te literature sows tat te ejector cycle becomes complicated due to te pase separator from its mass balance point of view. Practically tis may give a lower performance during separation. So tere is a clear need for furter investigations of te bi-evaporator system in wic instead of te pase separator, an auxiliary evaporator is applied for te case of low GWP refrigerants as a working fluid. In tis paper, termodynamic and exergetic analysis of two evaporator ejector refrigeration systems working wit low GWP refrigerants in buses is performed. In calculations, refrigerant velocities at te ejector inlets are taken into consideration in energy and momentum equations. Cange of COP of te conventional cycle and ejector refrigeration cycle is compared and sown grapically. Exergy destruction rates are also investigated. Two evaporator ejector bus refrigeration systems In buses, air conditioning systems are vapor compression cycles driven by te engine troug te clutc. Bus refrigeration systems components are a compressor, condenser, liquid tank, evaporator and expansion valve. Te layouts of te ejector bus refrigeration system s components are settled on te bus roof, except te compressor. Figure gives a scematic view of te system. Two-pase ejector system is one of te alternate cycles for expansion work recovery. Te liquid at te outlet of te condenser is split into two streams. One of tese streams
THERMAL SCIENCE: Year 0, Vol., No., pp. -4 7 Expansion valve (LH) 7 Evaporator (LH) Expansion valve (LH) 7 6 Evaporator- (LH) Evaporator- (LH) Expansion valve (RH) Condenser is isentropically expanded troug te ejector, and te oter stream is isentalpically trottled, sent troug an evaporator, and ten sent to te suction nozzle of te ejector. Te two streams are combined in te ejector mixing section, and te two-pase fluid at te ejector outlet is sent troug a secondary evaporator before entering te compressor. Main advantages of tis system tat because te two-pase fluid exiting te ejector are evaporated before entering te compressor, tis cycle does not require a liquid-vapor separator, and because of te increase in saturation pressure and temperature provided by te ejector, tis cycle yields te possibility for two different evaporation temperatures. Pressure-entalpy (P-) diagram of te system is sown in fig.. Te refrigerant wic comes from te first evaporator enters te compressor. Ten te fluid is compressed to te desired level. After te refrigerant is cooled in condenser ten it is divided into to be sent to te ejector and expansion valve at te condenser exit. Te fluid, wic leaves te expansion valve, goes te second evaporator ten te refrigerant flows into te ejector as te secondary flow. Primary and secondary flows are mixed in te mixing section. After te process, te refrigerant enters diffuser section of te ejector. Ten it enters to te first evaporator. Te refrigerant wic comes from te first evaporator goes to te compressor, tus te cycle is completed. Compressor Expansion 6 Evaporator- (RH) 7 Evaporator (RH) valve (RH) 7 Evaporator- (RH) In order to conduct termodynamic analyses of te cycle, we accepted te following assumptions: condensing and evaporation temperatures were constant, pressure losses of te wole system are neglected, nozzle and diffuser isentropic efficiencies are known, and te process in te mixing section takes place at constant pressure and constant cross-sectional area and efficiency of te mixing section of te ejector are known. Termodynamic properties of point () can be determined if te first evaporator temperature is known. For calculating te termodynamic properties at te compressor exit, compressor isentropic efficiency expression can be used: η comp = Ejector s Condenser Figure. Scematic view of te conventional and ejector refrigeration systems P [kpa] 0 4 0 4 7 6 Compressor 0 0 7 00 0 7 00 0 7 00 [kj kg ] Figure. Te P- diagram of te conventional and ejector refrigeration systems ()
THERMAL SCIENCE: Year 0, Vol., No., pp. -4 Isentropic efficiency of compressor canges wit compression ratio and compressor rpm value. However, as te general assumption, can be used empirical formulation given in reference []. Te s is determined by using eq. () as a function of te P s and s s. P s and s s indicate pressure and entropy after isentropic compression, respectively. As known, s s is equal to s : = f( s, P ) () s s s Te entalpy of te refrigerant at te compressor discarge can be found from eq. () by using te compressor isentropic efficiency. Since te condenser temperature is known, termodynamic properties at te condenser exit are calculated. Te ejector consists of tree main sections tat are nozzle, mixing, and diffuser. Points (), (4), () and (6) indicate te nozzle inlet, nozzle exit, diffuser inlet and diffuser exit, respectively. Termodynamic properties at te point (4) can be calculated by using te energy equation between points () and (4) given in Eq. () and eq. (4). Due to conservation of mass principle, it can be considered tat and velocity of te refrigerant at te nozzle inlet is neglected in eq. (). V4 = 4 + () 4 ηn = (4) 4s In te mixing section, constant pressure-constant cross-sectional area model is used. Pressure difference produced by te entrainment process of te ejector is neglected. However, inlet velocity of te secondary flow is taken into consideration in mixing section energy equation. Here, conservation of mass principle is applicable as given in eq. () and te termodynamic properties of te refrigerant at te point () can be calculated by using energy and momentum equations wic are given in eqs. (6) and (7). m + m = m () 4 V 4 V V 4 + + ω + = ( + ω) + ( + ω) V ηm = V 4 + ωv ω is defined as te entrainment ratio, wic sows te ratio of mass flow rates of primary and secondary fluids tat enter te ejector, given in eq. (): m ω = m In order to determine termodynamic properties at te diffuser exit, diffuser isentropic efficiency, and energy equation can be used as given eqs. (9) and (0), respectively. (6) (7) () V V6 + = 6 + (9) ηd = 6s 6 (0)
THERMAL SCIENCE: Year 0, Vol., No., pp. -4 9 In order to provide proper oil return, te minimum refrigerant velocity is recommended as -7 m/s in te compressor suction line [9]. However, tere is an evaporator between te compressor and diffuser outlet in tis work. So, velocity of te fluid at te ejector diffuser exit is considered as V 6 = m/s for te sake of safe oil return. Te refrigerant goes to te expansion valve at te point () and exits from tere at point (7). Te expansion process in te expansion valve is constant entalpy process. Primary mass flow rate (ṁ ) can be calculated from te cooling capacity of te system wic is given in eq. (): Q = m ω( ) + ( + ω)( ) () [ ] 7 6 Te coefficient of performance of te system can be calculated by te equation below: Q e+ Q e ω( 7) + ( + ω)( ) COP = = W ( + ω)( ) COP of te ejector refrigeration system given in eq. () is compared to te COP of te conventional refrigeration system (COP conv ), and te COP increase rate (COP * ) is determined: COP COP = COPconv conv COP 00 Te secondary evaporation temperature is used for te conventional cycle computation. Terefore, COP of te conventional refrigeration system can be calculated by eq (4): COP conv = 7 () () (4) Exergy analysis Exergy analysis can overcome many limitations of energy analysis. It is useful to identify te locations, magnitudes, and causes of process inefficiencies [0]. Exergy analysis is a powerful tool for designing, optimization, and performance evaluation of energy systems and it is aimed to determine te maximum performance of te system wit overcoming exergy destructions [, ]. Te system s exergy models are establised as follows [-]. According to te definition of exergy and exergy balance at steady operation, te exergy at any point and exergy destruction in a component can be expressed as follows: [( ) ( )] Ex = m T s s () i i 0 0 i 0 T0 T0 Ex d = Ex in Ex out + Q Q + Win Wout T T (6) in Were T 0 is as reference temperature maintained at 7 o C and te reference pressure sets at 0. kpa trougout tis study for reference entalpy and entropy. Subscript i denotes to eac point sown in fig.. Te temperature T was assumed to be o C greater tan te evaporation temperature in te evaporators and o C less tan te condensation temperature in te condenser. out
0 THERMAL SCIENCE: Year 0, Vol., No., pp. -4 Results and discussion In tis study, under different operating conditions, a variation of COP, entrainment ratio and exergy analysis of a bus air conditioning system tat as two separate evaporators using two-pase ejector are investigated. To evaluate termodynamic properties and solve te equations Klein [6] is used. Air conditioning systems wit te cooling capacity of 4 kw are widely used on midi-buses. Te ambient temperature can be assumed to be o C for te design condition of te bus air conditioning system for te Mediterranean climate zone [7]. For practical applications, evaporator temperatures were taken as and as te primary and secondary evaporators, respectively []. In te calculations, nozzle, mixing, and diffuser section efficiencies were taken as 0.9, 0. and, 0.9, respectively, []. Te model to be used in tis system was validated by Unal and Yılmaz []. In addition, R4a refrigerant system, Unal [9] prepared an experimental set-up for te bus ejector air conditioning system wic is depicted in fig.. An air conditioner currently used on te midibuses was turned into an ejector air conditioning system. It as 4 kw cooling capacity, and experimental studies were conducted on tis system. Te range and accuracy of te pressure transmitter are 0- bar and ±.0%, respectively. K-type termocouples are calibrated to provide an accuracy of ±0.. According to termodynamic analysis, te pressure difference between te refrigerant at diffuser outlet and te secondary evaporator outlet value was 9.4 kpa and temperature difference was 4 o C. According to te experimental results, pressure difference was measure around 0 kpa and temperature difference was. o C after te system reaces steady state. COP increase rates were also calculated % and measured %. According to tese results, it can be said tat differences between te teoretical and experimental results were witin acceptable limits and terefore, te calculation metod for te ejector refrigeration system can be considered as validated. Figure. Ejector refrigeration system for bus Te effect of condensing and evaporation temperatures Mobile air conditioning systems can operate under wide range ambient conditions. To conclude tat reason, te performance of air conditional system sould be investigated wit a variation of condensing temperature. Figure 4 sows a variation of COP, COP*, and ω wit condensing temperature for tree refrigerants. Similar to standard systems, in ejector refrigeration system also beave similar trend. As sown in fig. 4(a), increasing condenser temperature results wit decrease in performance. Te significant result of te figure tat te system wic is working wit R4yf, as lower performance tan oters about %. In stark contrast, owever, COP increase rate for R4yf is better. So, potential of te system performance of te R4yf system can be improved more by using two-pase ejector. As seen in fig. 4(b), entrainment ratio is inversely proportional to COP.
THERMAL SCIENCE: Year 0, Vol., No., pp. -4 COP 4 COP R4a COP R4yf COP R4ze(E) COP* R4a COP* R4yf COP* R4ze(E) 0 COP* [%] ω. 0.9 R4a R4yf R4ze(E) 0 0.6 0 0. 40 4 0 60 T c [ C] 0.0 40 4 0 60 T c [ C] Figure 4. (a) effect of condensing temperature on COP and COP *, (b) effect of condensing temperature on entrainment ratio COP.40..0..0..0.0.00 COP R4a COP R4yf COP R4ze(E) COP* R4a COP* R4yf COP* R4ze(E) 0 4 6 0 Te [ C] 40 0 0 0 COP* [%] COP.6.4..0..6 R4a R4yf R4ze(E) COP* R4a COP* R4yf COP* R4ze(E) 0 4 Te [ C] Figure. (a) effect of second evaporator temperature on COP and COP *, (b) effect of first evaporator temperature on COP and COP * Te effect of te second evaporator temperature on COP is given in fig. (a). COP of te standard system increases wit second evaporator temperature, owever, te ejector system COP decreases sligtly. Since cooling load is kept as constant, COP depend on te compressor work only. Wen second evaporator temperature gets iger, its saturation pressure increases. In conventional systems, as te second evaporator pressure increases, compression ratio decreases, and terefore COP increases. Second evaporator temperature does not affect compression ratio in ejector refrigeration system. Because of te total refrigerant mass flow rate rises, COP decreases sligtly. Te COP * decreases because of te conventional system performance increase but ejector system performance is nearly stable. Altoug R4yf as lower ejector refrigeration system performance, it as iger COP increase rate potential. Te effect of te first evaporator temperature on COP and COP * is given in fig. (b). In a conventional refrigeration system, tere is no first evaporator, so it does not effect on conventional COP. In ejector refrigeration system, COP depends on te first evaporator pressure and temperature because te compressor is placed on te first evaporator line. So, COP of te ejector refrigeration system increases wit first evaporator temperature. Te effect of compressor efficiency Te compressor plays a key role in vapor compression refrigeration system performance. In addition to tis, mobile air conditioning system suc as bus refrigeration systems can operate under a wide range of ambient conditions. As mentioned above, te compressor driven 40 0 0 0 0 COP* [%]
THERMAL SCIENCE: Year 0, Vol., No., pp. -4 COP 6 4 R4a ejector R4a conv R4yf ejector R4yf conv R4ze(E) ejector R4ze(E) conv 0.4 0. 0.6 0.7 0. η c Figure 6. Effect of compressor efficiency on COP by te engine and engine speed is canged according to road conditions (stopping at stations or traffic ligts, sifting gear, etc.) as a result of tese situations, te speed of air conditioning compressor is canged, te isentropic efficiency of te compressor does not remain constant. Tus, COP of te air conditioning system is affected. In fig. 6, te benefit from compressor isentropic efficiency improvement was te system COP increasing, wic resulted from compressor power consumption reduction. Exergy destruction rates Exergy calculations are carried out for bot conventional and ejector based cycles to ave a clear view of losses. Exergy destruction rates are estimated at condensing temperature 4 o C, cooling load is 4 kw, first evaporator and second evaporator temperatures are o C and o C, respectively. For te compressor, isentropic efficiency empiric equation in reference [] is used. Figure 7(a) sows a comparison of exergy destruction rates in components for te conventional refrigeration system and ejector refrigeration system as an R4yf refrigerant. Exergy destruction rate in te evaporator and condenser of bot te cycles are almost te same for te given operating conditions and cooling capacity. As seen in te figure tat te compressor as te biggest exergy destruction for bot systems. Since compressor isentropic efficiency is low in bus compressors due to te transmission components, a large amount of energy is lost in te compressor. And ejector system as lower exergy destruction tan te conventional system as. kw and.6 kw, respectively. Wen te cycle is turned into ejector refrigeration system, destruction in compressor and expansion valve decreases because of te falling in compression ratio. So te improvement in COP is acieved wit tis reduction. Figure 7(b) sows te exergy destruction rates for all components in ejector refrigeration system wit substances suc as R4a, R4yf, and R4ze(E). Te biggest exergy destruction occurs in te R4yf system, it explains wy te performance is lower tan oter refrigerant systems. Total exergy destructions are.6 kw,. kw, and. kw for R4ze(E), R4a, and R4yf, respectively. Te important difference of te destruction rate occurs in te compressor, oter components cause a similar amount of exergy destruction for all refrigerants.. Exergy destruction [kw].. R4yf conv R4yf ejector Exergy destruction [kw]. R4ze(E) R4a R4yf 0. 0. (a) 0 Compressor Condenser Expansion Evaporator Ejector valve (b) Figure 7. (a) Comparison of exergy destruction rates for conventional and ejector system, (b) exergy destruction rates for all components 0 Compressor Condenser Expansion Evaporator Ejector valve
THERMAL SCIENCE: Year 0, Vol., No., pp. -4 Conclusions In tis paper, te termodynamic and exergetic analysis of low GWP refrigerants R4yf and R4ze(E) intercity bus air condition system were analyzed under variable operating conditions. Te performance improvement by using ejector was mainly focused on new refrigerants. Te following conclusions for teoretical study can be drawn. y Te analysis results revealed tat altoug R4yf ejector refrigeration system COP was lower by 6% tan te R4a ejector system, R4yf more perfectly meets environmental concerns. Tat makes it ideal as a replacement, as it may be possible tat few or no alterations are required in order to replace R4a wit R4yf in a pre-designed system. y Te igest exergy destruction occurs in te R4yf system. y Te igest exergy destruction occurs in te compressor for eac refrigerant. y Te increase rates in te COP wit respect to te conventional system by using ejector as an expansion device are % for R4a, about 7% for R4yf, and about % for R4ze(E), respectively. y Entrainment ratio is inversely proportional to COP values. It sould be noted tat te aim of tis study is to investigate te performance improvement potential. Future studies sould be performed in order to validate experimentally for alternative refrigerants on real road conditions. Nomenclature COP coefficient of performance, [ ] COP* increase rate, [%] Ėx exergy, [W] entalpy [Jkg ] ṁ mass flow rate, [kgs ] P pressure, [Nm ] Q cooling capacity, [W] s entropy, [Jkg K ] T temperature, [ o C] V refrigerant velocity, [ms ] Ẇ work [W] References Greek symbols η efficiency, [ ] ω entrainment ratio, [ ] Subscripts c condenser comp compressor conv conventional d diffuser e first evaporator e second evaporator m mixing section n nozzle s isentropic state [] ***, Kyoto Protocol, Report of te Conference of te Parties, United Nations Framework Convention on Climate Cange (UNFCCC), 997, ttp://www.unfccc.int [] ***, Regulation (EU) No. 7/04 of te European Parliament and of te Council on Fluorinated Greenouse Gases and Repealing Regulation (EC) No. 4/006, ttp://eur-lex.europa.eu [] Jaranejad, M., New Low GWP Syntetic Refrigerants, M. Sc. tesis, KTH Scool of Industrial Engineering and Management Energy Tecnology, Stockolm, 0 [4] Minor, M., Spatz, M., HFO-4yf Low GWP Refrigerant Update, Proceedings, International Refrigeration and Air Conditioning Conference, Purdue, West Lafayette, Ind., USA, 00, Paper No. 49 [] Reasor, P., et al., Refrigerant R4yf Performance Comparison Investigation, Proceedings, International Refrigeration and Air Conditioning Conference, Purdue, West Lafayette, Ind., USA, 00, Paper No. 0 [6] Lu, M.C., et al., Investigation of te Two-Pase Convective Boiling of HFO-4yf in a.9 mm Diameter Tube, Int. J. Heat. Mass Transf., 6 (0), Oct., pp. 4- [7] Longo, G. A., Zilio, C., Condensation of te Low GWP Refrigerant HFC4yf Inside a Brazed Plate Heat Excanger, Int. J. Refrigeration, 6 (0), Nov., pp. 6-6 [] Mota-Babiloni, A., et al., A Review of Refrigerant R4ze(E) Recent Investigations, Appl. Term. Eng., 9 (06), pp. -
4 THERMAL SCIENCE: Year 0, Vol., No., pp. -4 [9] Yataganbaba, A., et al., Exergy Analysis of R4yf and R4ze as R4a Replacements in a Two Evaporator Vapour Compression Refrigeration System, Int. J. Refrig., 60 (0), Dec., pp. 6-7 [0] Qi, Z., Performance Improvement Potentials of R4yf Mobile Air Conditioning System, Int. J. Refrig., (0), Oct., pp. -40 [] Sarkar, J., Ejector Enanced Vapor Compression Refrigeration and Heat Pump Systems A Review, Renew Sustain Energy Rev., 6 (0), 9, pp. 6647-669 [] Besagni, G., et al., Ejector Refrigeration: A Compreensive Review, Renew Sustain Energy Rev., (06), Jan., pp. 7-407 [] Cen, J. Y., et al., A Review on Versatile Ejector Applications in Refrigeration Systems, Renew Sustain Energy Rev., 49 (0), Sept., pp. 67-90 [4] Besagni, G., et al., A Study of Working Fluids for Heat Driven Ejector Refrigeration Using Lumped Parameter Models, Int. J. Refrig., (0), Oct., pp. 4-7 [] Wang, X., Yu, J., Experimental Investigation on Two-Pase Driven Ejector Performance in a Novel Ejector Enanced Refrigeration System, Energy Conversion and Management, (06), Mar., pp. 9-400 [6] Li, C., et al., Analysis on Performance Caracteristics of Ejector wit Variable Area-Ratio for Multi-Evaporator Refrigeration System Based on Experimental Data, Appl. Term. Eng., 6 (04), -, pp. - [7] Kutlu, C., et al., Termodynamic Analysis of Bi-Evaporator Ejector Refrigeration Cycle Using R744 as Natural Refrigerant, Journal of Termal Engineering, (06),, pp. 7-740 [] Unal, S., Yilmaz, T., Termodynamic Analysis of te Two-Pase Ejector Air Conditioning System for Buses, Appl. Term. Eng., 79 (0), Mar., pp. 0-6 [9] Trott, A. R., Welc, T., Refrigeration and Air-Conditioning, t ed., Butterwort-Heinemann, Oxford, UK, 000 [0] Dincer, I., Rosen, M. A., Exergy, Energy, Environment and Sustainable Development, st ed., Elsevier, Oxford, UK, 007 [] Zao, H., et al., Termodynamic Investigation of a Booster-Assisted Ejector Refrigeration System, Appl. Term. Eng., 04 (06), July, pp. 74- [] Yan, G., et al., Energy and Exergy Analysis of a New Ejector Enanced Auto-Cascade Refrigeration Cycle, Energy Conversion and Management, 0 (0), Nov., pp. 09-7 [] Fazelpour, F., Energetic and Exergetic Analyses of Carbon Dioxide Transcritical Refrigeration Systems for Hot Climates, Termal Science, 9 (0),, pp. 90-94 [4] Dai, Y., et al., Exergy Analysis, Parametric Analysis and Optimization for a Novel Combined Power and Ejector Refrigeration Cycle, Appl. Term. Eng., 9 (009), 0, pp. 9-990 [] Cian, E., Kavasogulları, B., Energy and Exergy Analysis of a Combined Refrigeration and Waste Heat Driven Organic Rankine Cycle System, Termal Science, (06), 6A, pp. 6-6 [6] Klein, S. A., Engineering Equation Solver Academic Commercial V9, F-Cart Software, 0 [7] ***, Climatic Design Information, ASHRAE Handbook, Fundamentals, Atlanta, Gea., USA, 00 [] Liu, F., Groll, E. A., Study of Ejector Efficiencies in Refrigeration Cycles, Appl. Term. Eng., (0),, pp. 60-70 [9] Unal, S., Determination of te Ejector Dimensions of a Bus Air-Conditioning System Using Analytical and Numerical Metods, Appl. Term. Eng., 90 (0), Nov., pp. 0-9 Paper submitted: October 4, 06 Paper revised: Marc 7, 07 Paper accepted: April, 07 0 Society of Termal Engineers of Serbia Publised by te Vinča Institute of Nuclear Sciences, Belgrade, Serbia. Tis is an open access article distributed under te CC BY-NC-ND 4.0 terms and conditions