The energy and exergy analysis of a novel cogeneration organic Rankine power and twostage compression refrigeration cycle

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1 Energy Equip. Sys./ Vol. 5/No.3/September 2017/ Energy Equipment and Systems ttp://energyequipsys.ut.ac.ir Te energy and exergy analysis of a novel cogeneration organic Rankine power and twostage compression refrigeration cycle Autors Hamed Mortazavi Beni a Afsin Amadi Nadoosan a* Morteza Bayare a a Department of Mecanical Engineering, Sarekord University, Sarekord,Iran Article istory: Received : 13 February 2017 Accepted : 24 April 2017 ABSTRACT Te energy crisis in recent years as led to te use of termodynamic cycles tat work based on renewable energies. Lowtemperature cycles suc as organic cycles are suitable strategies for te application of renewable energies. Te present study proposes a novel cycle troug te integration of a two-stage compression refrigeration cycle wit a combined Rankine power and ejector refrigeration cycle by using te cascade condenser metod. Te fundamental idea of tis cycle is to obtain refrigeration production at lower temperatures, and to acieve iger termal and exergy efficiencies. Te results sowed tat te new cycle recorded an percent improvement in termal efficiency and a percent improvement in exergy efficiency compared to te basic cycle. Even toug te network output of te cycle is reduced, a significant increase in te refrigeration capacity of te cycle is observed. Keywords: Cogeneration Cycle, Exergy, Solar Energy, Ejector, Cascade Condenser. 1. Introduction In tis researc, te new cogeneration organic Rankine power and two-stage compression refrigeration cycle as been analyzed from te viewpoint of energy and exergy. In tis cycle, solar energy is used as te eat source. Nowadays, solar energy is considered more frequently because it is a clean and renewable energy. Wile te use of solar collectors and setting up te cycles wit solar energy as te eat source as iger initial costs, te current cycle cost is muc lower due to te lack of environmental costs and as fossil fuels are not required. Tis issue is also significant wit regard to te lack of environmental pollution in te solar cycles. Low-temperature cycles suc as cycles tat use organic fluids *Corresponding autor: Afsin Amadi Nadoosan Address: Department of Mecanical Engineering, Sarekord University, Sarekord,Iran address: amadi@eng.sku.ac.ir as refrigerants are te proper approac for using sustainable, clean, and renewable energies suc as geotermal and solar energy. Zang et al. [1] evaluated te performance of te double-compression flas intercooling refrigeration cycle in comparison wit te double-compression external intercooling cycle. Te results sowed tat te doublecompression flas intercooling refrigeration cycle and te double-compression external intercooling cycle yield te maximum coefficient of performance improvement of 23.18% and 11.03% over te basic cycle, respectively. Tere are numerous metods for improving te performance of te vaporcompression refrigeration cycle. One of tese metods involves te use of an ejector as an expander device. Using an ejector is preferred due to te ig efficiency and te lack of mecanical rotary components, wic reduces depreciation. Nowadays, te ejector

2 300 Hamed Mortazavi Beni et al./ Energy Equip. Sys. / Vol. 5/No.3/Sep is used in many mecanical cycles suc as power plants and compression refrigeration cycles. So far, many researc studies ave focused on ejector expansion compression refrigeration cycles [2 6]. In tis researc, te ejector is used as an expander device in te compression refrigeration cycle instead of using an expansion valve in order to avoid te work and energy losses in te expansion valve. In anoter study, Xing et al. [7] proposed an ejector subcooled vaporcompression refrigeration cycle. Te results sowed tat te performance of te ejector subcooled cycle is better tan tat of te conventional cycle. Wang et al. [8] compared te use of te ejector expander in te vaporcompression refrigeration cycles for applications in domestic refrigerator freezers and presented a novel modified ejector expansion vapor-compression refrigeration cycle. Zu and Jiang [9] developed a refrigeration system by combining te basic vapor-compression refrigeration cycle wit an ejector cooling cycle. In tis system, te ejector-cooling cycle is driven by te waste eat from te condenser in te vaporcompression refrigeration cycle. Te results sow tat te coefficient of performance is improved by 9.1% in te ybrid refrigeration system. Te ejector application domain is not limited to te vapor-compression refrigeration cycle. It is used in many cases, suc as in te petroleum and petrocemical industries for te purification and separation of crude oil, and also as a vacuum pump in devices suc as condensers. Te ejector is also used in te ejector refrigeration cycle. So far, many researces on ejector refrigeration cycles, as well as on combined Rankine power and ejector refrigeration cycles, ave been carried out. Cen et al. [10] studied te interactions and relationsips of various ejector parameters in te ejector refrigeration system to gain access to an optimum generator temperature tat obtains te maximum Carnot efficiency. Sorouradin et al. [11] investigated te performance of an ejector refrigeration cycle teoretically and experimentally. Te results indicate a decrease in te coefficient of performance wit increasing generator temperature and an increase in te second law of efficiency wit increasing evaporator temperature and decreasing generator temperature. Adrian et al. [12] optimized an ejector refrigeration system wit different working fluids tat operates on waste eat provided by te exausted gas of an internal combustion engine by energy and exergy analysis. Lontsi et al. [13] proposed a multitemperature compression ejection refrigeration cycle by combining a compression refrigeration cycle and an ejector refrigeration cycle. Tey suggested using tis multi-temperature refrigeration cycle instead of te conventional two-stage vaporcompression refrigeration system. By adding a steam turbine before te ejector in ejector refrigeration, te new cycle can be based on combined power generation and ejector refrigeration. Many researc studies ave focused on suc cycles [14 18]. In tese researces, te researcers investigated te factors affecting te cycle, suc as te input and output temperature, te pressure of te turbine, te evaporator temperature, and te entrainment ratio of te ejector to analyze and optimize te combined Rankine power and ejector refrigeration cycle. In oter researces, Yang et al. [19 20] analyzed te combined power and refrigeration cycle by using a zeotropic refrigerant wit different mixture compositions. Te result sowed tat using a 50/50 composition of isobutane/pentane as te maximum exergy efficiency of 7.83%. In anoter study, Yang et al. [21] proposed a novel combined power and ejectorrefrigeration cycle by using te two-stage condensation of a zeotropic mixture. Also, a zeotropic mixture is divided into te power cycle and te ejector refrigeration cycle in different compositions. Te result revealed tat te cycle exergy efficiency acieves a maximum value of 10.29% wit te use of a 40/60 percent composition of isobutane/pentane, and te termal efficiency yields a maximum value of 10.77% wit te use of a 70/30 percent composition of an isobutane/pentane zeotropic mixture. In te current study, te novel cycle is proposed by te integration of a two-stage compression refrigeration cycle wit a combined Rankine power and ejector refrigeration cycle by using te cascade condenser metod. Since te ejector refrigeration cycle cannot produce refrigeration at very low temperatures, te fundamental idea of tis cycle is based on furter refrigeration production at lower temperatures by using te solar energy results in iger termal and exergy efficiencies. Nomenclature mass flow rate (kgs -1 ) entalpy (kjkg -1 )

3 Hamed Mortazavi Beni et al./ Energy Equip. Sys. / Vol. 5/No.3/Sep u P Greek Symbols work rate (kw) flow exergy rate (kw) eat transfer rate (kw) exergy destruction (kw) temperature (ᵒC) velocity (m/s) entropy (kjkg -1 K -1 ) pressure (kpa) exergy efficiency termal efficiency isentropic efficiency µ entrainment ratio Subscripts ambient conditions inlet e outlet out s surface des exergy destruction L cooled environment evap evaporator comp compressor cond condenser expan expansion valve Cas.Cond cascade condenser Se separator is isentropic T turbine B boiler P pump C1 Compressor 1 C2 Compressor 2 n primary nozzle m mixing camber d diffuser w water 2. Cycle description Te proposed cogeneration Rankine power and two-stage compression refrigeration cycle consists of a combined Rankine power and ejector refrigeration cycle and te two-stage compression refrigeration cycle. First, te refrigerant fluid, by taking eat in te steam generator troug te solar collector, and by increasing in temperature and pressure, canges into supereated steam. Usually from water or industrial oils used as primary fluid troug te solar collector absorber. In tis study, primary fluid from te industrial oil Terminol 66 is used in te absorbent collector. Terminol 66 is te world s most popular ig-temperature, liquid-pase eattransfer fluid. Terminol 66 is pumpable at low temperatures and offers termal stability at ig temperatures. Tis oil is used in a wide variety of systems, suc as te production of plastics and polymers, refining, syntetic fiber manufacturing, cemical processing, water purification, solar collectors, and organic Rankine cycle applications. Since tis oil as good stability even at te igest recommended temperature for oil in continuous use in te system, it also as considerable resistance against fouling. Fouling tends to reduce system efficiency and increase costs. Hence, tis oil is suggested for use in solar cycles. Using Terminol 66 as te primary fluid in te solar collectors can easily provide supereated vapor in te generator from te working fluid of te organic Rankine cycle even at temperatures up to 300 C. Te supereated vapor expands troug te turbine to produce power. Te turbine output tat is still supereated vapor is used as te primary flow in te ejector. Te inlet igpressure primary fluid of te ejector passes troug te converging diverging nozzle. Ten, by a vertical sock wave, te pressure energy of te fluid is converted to kinetic energy and te fluid velocity becomes supersonic. Te result of tis process creates a low-pressure area and produces te igvacuum area at te inlet of te ejector secondary fluid tat entrains te secondary fluid into te mixing camber of te ejector. So, te suction of te secondary fluid creates te required pressure drop in te evaporator to evaporate te refrigerant and produce a refrigeration effect. After mixing te primary and secondary flow in te mixing camber troug te constant diameter troat of te ejector, te fluid undergoes anoter sock wave; finally, te pressure of te fluid increases troug te divergent diffuser of te ejector. Te ejector outlet pressure is between te pressure of te primary and secondary flows. In te following step, te refrigerant fluid condensates troug te first condenser. Ten, to complete te processing part, te fluid is pumped into te vapor generator, and te rest of te refrigerant fluid passes troug te expansion valve and is driven to te

4 302 Hamed Mortazavi Beni et al./ Energy Equip. Sys. / Vol. 5/No.3/Sep evaporator. Te cogeneration Rankine power and two-stage compression refrigeration cycle in tis study is started up and operated by solar energy. Figure 1 sows a scematic diagram of te discussed cycle. Te process from te evaporator of te combined power and te ejector refrigeration cycle used as te cascade condenser to te primary condensing of te vapor refrigerant in te two-stage compression refrigeration cycle is sown in Fig. 3. Here, te refrigerant of te two-stage vapor-compression refrigeration cycle passes troug Compressor 1 and increases te pressure driven to te cascade condenser were te primary condensate lies. Te cascade condenser output tat is still in te two-pase mixture of te refrigerant is entered into te vapor liquid separator. Ten, te cascade condenser output mixed wit Flow 16 in te separation and te final composition is te two-pase mixture of te refrigerant. Te saturated vapor refrigerant moves to Compressor 2 for ig-pressure compression. Ten, te secondary condensate in Condenser 2 transforms to te saturated liquid state and after passing troug te expansion valve returns again to te separator. Also, te saturated liquid part of te refrigerant mixture in te separator moves into Expansion Valve 3. It is ten driven to te evaporator for refrigeration production and completion of te cycle. Te use of te two-stage compression tecnique not only elps to reduce te power consumption of te compressors but also, because of te twostage condensation and expansion processes, allows te cycle to produce more refrigeration at lower temperatures. Te following assumptions are made for te simulation of te cycle: 1) All cycle processes are steady-state, steady-flow processes, and te effects of kinetic and potential energy are neglected. Te pressure drop and eat losses in te pipes and system components are negligible. 2) Te condensers and te evaporator outlet state are te intended saturated liquid and saturated vapor, respectively. Te condenser-saturated temperature is considered to be 20 C; according to te source [14], te ambient temperature and pressure are assumed to be 15 C and kpa, respectively. 3) Te condensers eat excangers are considered water cooling and te evaporator eat excanger is considered air cooling. 4) Te flow across te expansion valves is isentalpic. 5) Te ejector flow is steady-state and onedimensional. Te velocity of streams at te inlet and outlet of te ejector is negligible. 6) For simplicity, te effects of te stream losses in te frictional and mixing processes in te nozzle, mixing, and diffuser sections are considered to be te isentropic efficiency of te nozzle, te mixing camber, and te diffuser. Also, te ejector processes are assumed to be adiabatic and do not excange eat wit te environment. 7) Since in comparing te primary flow velocity, te secondary flow velocity is negligible, it is assumed tat te suction process is ideal and witout a pressure drop. Terefore, te mixing process in te mixing camber of te ejector occurs at te suction pressure. It complies wit te laws of conservation of energy and momentum. 3.Simulation Te equations governing on te cycle are conservation of te mass, energy and exergy. In discussed cycle we are dealing wit flow exergy and steady state steady flow processes. Te conservation of mass, energy and exergy are expressed as follows: (1) (2) ( ) (3) In tese equations, denotes entalpy, is te eat transfer rate, is work, and is irreversibility or exergy destruction. In te conservation of exergy, ( ) refers to te eat transfer exergy. Te eat transfer of te control surface is te maximum obtainable work from te transferred termal energy at temperature. Tis term of te exergy balance equation will apply only on devices tat involve eat transfer wit te environment (eat excanger), suc as in an air-cooled evaporator. In te adiabatic

5 Hamed Mortazavi Beni et al./ Energy Equip. Sys. / Vol. 5/No.3/Sep Parabolic Solar Dis Collector Energy Storage Tank 3 way plug valves Oil out Oil in Generator 6 Pump 1 2 Turbine Cooling Water Ejector Condenser 1 Expansion Valve Separator Cascade Condenser Expansion Compressor 1 Valve 2 Compressor Evaporator 18 Expansion Valve 3 Condenser 2 14 Cilled Air Cooling Water Fig 1. Scematic diagram of cogeneration rankine power and two stage compression refrigeration cycle processes, were te eat transfer rate is considered to be zero, suc as in compressors, tis term of te exergy balance equation will be removed. By taking te control volume for every single component of te termal cycle and applying te conservation of mass, energy, and exergy for te various components of te cycle, te equations for te exergy and energy analysis, and te calculation of te exergy destruction of te cycle components can be acieved. According to Fig. 1, tese equations are defined in Table 1 by using Eqs. (1) to (3). is te termo- pysical exergy flow rates of streams, given as: (4) Te isentropic efficiency of te compressors can be defined as a function of te compressor pressure ratio [23]. Te isentropic efficiency of te compressors can be expressed as: (5) were Pi and Po are te compressor suction and discarge pressures, respectively.

6 304 Hamed Mortazavi Beni et al./ Energy Equip. Sys. / Vol. 5/No.3/Sep Table 1. Energy and exergy destruction equations for te various components of te combined cycle Components Boiler Turbine Pump Ejector Condenser 1 Condenser 2 Energy balance,, m oil oil i oil o m W. m T T, s 2 2 3,s W P m 6 1, s Ps, m m m Q m cond Q m cond B Exergy destruction.... I Ex Ex Ex Ex oil, i 1 oil, o 2.. IT Ex 2 Ex 3 WT I Ex Ex W P Ej 6 1 I Ex Ex Ex P I Ex Ex Ex Ex cond 1 4 w, i, c1 5 w, o, c1 I Ex Ex Ex Ex cond 2 14 w,i,c 2 15 w,o,c 2 Cascade condenser Evaporator Separator Expansion valve 1 Expansion valve 2 Expansion valve 3 Compressor 1 Compressor m m Q m evap m m m m W W m C 1 10 m C , s 10 C1, s 14, s 13 C 2, s I Ex Ex Ex Ex CaCo I Ex Ex Q 0 evap evap( 1 ) T L I Ex Ex Ex Ex Se T I expan1 Ex 7 Ex 8 I expan 2 Ex 15 Ex 16 I expan3 Ex 17 Ex 18 I Ex W Ex C1 10 C1 11 I Ex W Ex C2 13 C2 14 Te termal efficiency of te cogeneration Rankine power and two-stage compression refrigeration cycle is expressed as: (6) Here,,, and denote te termal energy rate entered into te cycle, te refrigeration rate produced in te evaporator, and te net work output of te cycle, respectively. Te net work is calculated by subtracting te total power generated in te turbine from te power consumption by te pump and compressors. Te exergy efficiency of te cycle can be expressed as: ( ) (7) Here, is te total exergy input to te cycle. Tis is calculated by summation of te total exergy input into te cycle by te vapor generator and te total exergy entered into te cycle by te pump and te compressors. In fact, te difference between te first and second laws of termodynamics is tat te first law of termodynamic components, suc as pumps and compressors in te termal cycles, are considered as energy consumers, wile from te viewpoint of te second law of termodynamics, tese components are te entry points of te exergy into te cycle. So, te exergy efficiency of te cogeneration Rankine power and two-stage compression refrigeration cycle is calculated by te following equation: (8)

7 Hamed Mortazavi Beni et al./ Energy Equip. Sys. / Vol. 5/No.3/Sep Here, refers to te sum of exergy destruction of te cycle components tat is obtained using te equations given in Table 1. 4.Ejector simulation Te ejector is te key component in te cogeneration Rankine power and two-stage compression refrigeration cycles. So far, different models ave been provided to simulate te ejector. In tis study, te constant-pressure mixing model is used for simulating te ejector performance due to te better and more accurate prediction of tis model. Te basic principle of tis model was presented by Keenan et al. [24] based on gas dynamics. Ten te researc of Huang et al. [23] developed and generalized tis principle [14]. Te ejector simulation algoritm presented in tis study is inspired by te Sarkar algoritm [26]. According to Fig. 2, for te given primary and secondary termodynamic properties of te ejector and by applying te conservation of energy and momentum equations for different processes in te ejector, te output termodynamic properties of te ejector can be calculated. Considering te isentropic efficiency of te nozzle and applying te energy conservation to te ejector s primary nozzle, te entalpy and velocity of te flow can be calculated. (9) (10) Based on Assumption (7), te velocity of te secondary flow after te suction is negligible. Terefore, te entalpy of te secondary flow before and after te suction is equal, and te mixing process is done on te suction pressure, wic means te ejector pressure is equal to te suction pressure. So, te conservation of momentum for te mixing process can be expressed as follows: (11) Te mixing efficiency is: (12) By knowing te velocity of te flow after te mixing process, te energy-conservation equation for te mixing section can be written as: ( ) ( ) (13) Finally, at te diffuser outlet, it can be expressed as: (14) (15) One of te key parameters for te design and simulation of te ejector is te entrainment ratio. Te ejector entrainment ratio is defined as te mass flow ratio of te secondary flow to te primary flow and is given by: (16) Tus, according to te definition of te ejector entrainment ratio, Eq.(11), by using te isentropic efficiency of te mixing process, and Eq. (13) are rewritten as follows: (17) Fig 2. Scematic of te structure and processes in te ejector

8 306 Hamed Mortazavi Beni et al./ Energy Equip. Sys. / Vol. 5/No.3/Sep (18) Start 5.Simulation algoritm and validation A calculating program is written in EES. Te simulation procedure for te energy and exergy analysis of te cycle is as follows: First, Eqs. 5 to 8 are solved, followed by te equations in Table 1. Te termodynamic analysis of te cycle is performed in te EES code format. Te input information required in te simulation of te cycle is suc tat te equations are solved explicitly witout te need for trial and error. But in te ejector modeling, te simulation procedures will be sligtly different. Due to te equations given in te ejector simulation section, tere are two approaces to simulating te ejector. Te first approac is wen te entrainment ratio is given. Here, te equations are solved explicitly. Te second approac is wen te goal is to reac a given pressure in te output of te ejector. In tis case, te equations are solved by trial and error wit an initial value guess for te entrainment ratio to reac te required outlet pressure. Since te saturation temperature of te condenser is constant based on Assumption (2), te pressure of te ejector outlet stream must be equal to te saturation pressure corresponding to te saturation temperature of te condenser, so tat te evaporator back flow does not occur. Terefore, in tis study, te second approac is used. Fig. 3 sows te calculation flowcart used to solve te simulation algoritm proposed for te cogeneration cycle. To ensure te integrity of te software code written in EES and te accuracy of te calculation results obtained from te simulation outlet by EES, first a basic code is written to simulate te Rankine power and ejector refrigeration cycle. Next, te results of te simulation of tis cycle are validated by comparing tis wit te results of Dai et al. [14]. Te comparison results for te turbine power, refrigeration output, and exergy efficiency are sown in Fig. 4. Te results of te simulation matc te results of Dai et al. not only qualitatively [14] but also in numerical terms. Te minor differences in te results can be due to rounding-off errors in te computing process software. Tere may also be sligt differences in te resources and database used to calculate te properties of te refrigerants in te different software programs. No Select Refrigerant & input data: T 2,P 2,P 3,T c,t e,η n,η m,η d,etc. Evaluating flow and equipment properties using equations 1 to 5 & Table 1 Assume μ Solve ejector model equations 9 to 18 Output ejector properties T 4 & P 4 Is P 5-P 4 <ε? Yes Output result η T & η ex equations 6 & 8 Stop Fig 3. A simulation flowcart of te cycle analysis 6.Results and discussion Table 3 sows te results of te energy and exergy analysis of te cogeneration organic Rankine power and two-stage compression refrigeration cycle. In te simulation for a better analysis and understanding of te cycle performance, te input data and main assumptions of te cycle are taken to be te same as in te reference [14]. Tis allows comparing te results of te cogeneration Rankine power and two-stage compression refrigeration cycle wit te results of te conventional combined power and ejector refrigeration cycle. Some assumptions of te parameters are made as listed in Table 2. In te power section and two-stage compression refrigeration section of te cycle refrigerant, R123 and R134a are used respectively.

9 Hamed Mortazavi Beni et al./ Energy Equip. Sys. / Vol. 5/No.3/Sep Turbine power Refrigeration Exergy efficiency Present study Dai et al. [14] Qutput (kw) ex (%) T cond ( C) Fig 4. Validation te results of te turbine power, refrigeration output and exergy efficiency versus condenser temperature Table 2 Main assumptions for te combined cycle Parameter Heat source initial temperature (ºC) Cascade condenser temperature (ºC) Evaporator temperature (ºC) Turbine inlet pressure (kpa) Turbine inlet temperature (ºC) Turbine back pressure (kpa) Turbine isentropic efficiency (%) Pump isentropic efficiency (%) Nozzle efficiency of ejector (%) Mixing efficiency of ejector (%) Diffuser efficiency of ejector (%) Turbine mas flow rate (kg/s) Evaporator mas flow rate(kg/s) Value Table 3 Te results of energy and exergy analysis of te combined cycle Result Turbine work (kw) Pump work (kw) Compressor 1 work (kw) Compressor 2 work (kw) Refrigeration output (kw) Net power output (kw) Net power and refrigeration output (kw) Termal efficiency (%) Exergy efficiency (%) Present study Dai et al. [14]

10 308 Hamed Mortazavi Beni et al./ Energy Equip. Sys. / Vol. 5/No.3/Sep As sown in Table 2, te net work output of te Rankine power and two-stage compression refrigeration cycle is less tan te conventional combined power and ejector refrigeration cycle due to te lack of a compressor in te conventional cycle. Large increases are seen in te refrigeration capacity of te novel cycle, and a significant increase is observed in te termal efficiency and exergy efficiency compared to te conventional cycle. Since te increased refrigeration capacity is muc iger tan te decreased net power output, te termal efficiency as increased. Te increase in te exergy efficiency is due to increases in te net refrigeration exergy output of te cycle and reduction of te exergy destruction by te generator due to lack of a cimney in te generator owing to te use of solar energy. As a key component of te cycle, it is essential to investigate te effect of te ejector entrainment ratio on te termal and exergy efficiency. To do tis, by canging te temperature of te ejector s secondary flow, te ejector entrainment ratio can be canged. By increasing te temperature of te secondary flow, wic is in fact te temperature of te cascade condenser, te secondary flow rate will increase; tus, wit an increase in te secondary flow pressure, te entrainment ratio increases. Figs. 5 and 6 illustrate te effect of te variations of te entrainment ratio on te termal and exergy efficiency at various turbine inlet pressures. Te variation of te secondary flow temperature of te ejector is between 15 C to 5 C By increasing te turbine inlet pressure, te entalpy of te fluid increases and terefore te turbine power output increases. Tis causes an increase in te termal efficiency of te cycle. But by increasing te entrainment ratio in all te turbine inlet pressures, a reduction in te termal efficiency can be seen. Te reason for te decrease of te termal efficiency is tat it increases te Compressor 1 work and te cycle refrigeration capacity decreases as µ increases. Figure 6 sows tat te exergy efficiency increases as te turbine inlet pressure decreases. Tis beavior can be explained wit arguments similar to tose presented for termal efficiency. Also, it can be seen tat te exergy efficiency increases first to a maximum value and ten decreases as te entrainment ratio increases. By increasing te entrainment ratio, te ejector secondary mass flow rate increases. Increasing te secondary mass flow rate will increase te capacity of te cascade condenser, wic causes a reduction of te power consumption and te exergy destruction of te ig-pressure compressor. Terefore, at first, te exergy efficiency increases. But, on te oter side, increasing te secondary mass flow rate increases te exergy destruction by te cascade condenser. Also, by increasing te operating temperature of te cascade condenser, te power consumption and te exergy destruction of te ig-pressure compressor gradually increases. It is suc tat after a specific entrainment ratio, increasing te total exergy destruction as a predominant P Turbine =800 (kpa) P Turbine =810 (kpa) P Turbine =820 (kpa) P Turbine =830 (kpa) t (%) m Fig 5. Effect of te entrainment ratio on te termal efficiency at various turbine inlet pressure

11 Hamed Mortazavi Beni et al./ Energy Equip. Sys. / Vol. 5/No.3/Sep P Turbine =800 (kpa) P Turbine =810 (kpa) P Turbine =820 (kpa) P Turbine =830 (kpa) 39.4 ex (%) m Fig 6. Effect of te entrainment ratio on te exergy efficiency at various turbine inlet pressure effect on reducing te total exergy destruction since ten te exergy efficiency decreases. Figures 7 and 8 sow te effect of variations of te evaporator temperature on te termal and exergy efficiency at various entrainment ratios. Since in practical applications, te ejector refrigeration cycle cannot be used to produce refrigeration at very low temperatures, te ejector secondary flow temperature variations correspond to te considered entrainment ratio canges between 2 C to 5 C. As seen in Fig. 6, for an entrainment ratio more tan 0.1, wic approximately equates to a 5 C temperature for te ejector secondary flow, te exergy efficiency reduces as µ increases. Also, in Fig. 5, it is observed tat by increasing te entrainment ratio te termal efficiency decreases. Tis trend is also evident in Figs. 7 and 8. As sown in Figs. 7 and 8, generally by increasing te entrainment ratio, te termal and exergy efficiency decreases. But wit te increase of te evaporator temperature due to te reduction in te pressure ratio of Compressor 1, te power consumption of tis compressor reduces significantly. In addition, by increasing te evaporator temperature, te refrigeration capacity increases. So, in general, te termal efficiency of te cycle increases as te evaporator temperature increases. In te cogeneration organic Rankine power and two-stage compression refrigeration cycle, te evaporator temperature as a more important effect on te exergy efficiency. Wile te evaporator temperature increases, te low-pressure compressor power consumption and exergy destruction decrease.but tis decline in te power consumption of te low-pressure compressor decreases te net exergy input into te cycle. Also, increasing of te evaporator temperature despite be increased te refrigeration capacity, te quality of refrigeration output of te cycle yields because of te increasing te refrigeration temperature and approacing to te ambient temperature. However, as sown in Fig. 8, at te beginning of increasing te evaporator temperature, te reduction in te exergy destruction by te low-pressure compressor as a dominant influence in te exergy efficiency; so, at first, te exergy efficiency increases. But gradually wit a furter increase in te evaporator temperature, te effect of decreases in te net exergy input into te cycle and te yield of te quality of refrigeration ave a dominant effect on te exergy efficiency. Since ten, tere appears to be a decline in te exergy efficiency. Figure 9 illustrates te effect of te evaporator temperature on te exergy efficiency in comparison wit te termal efficiency at various eat source initial temperatures. Te exergy efficiency beavior can be explained qualitatively wit arguments similar to tose presented for Fig. 8. As sown in Fig. 9, te termal efficiency does not cange wit a variation of te eat source s initial temperature and only one

12 310 Hamed Mortazavi Beni et al./ Energy Equip. Sys. / Vol. 5/No.3/Sep grap is proposed for it. Te reason for tis beavior is because of te termal efficiency only Influenced by te eat flux input to te cycle and production of te net power and refrigeration. So, as long as te inlet eat flux input to te cycle troug te generator is fixed, te termal efficiency remains constant. Tis beavior sows te weakness of te first law of termodynamics in dealing wit te termal cycle. Te exergy efficiency decreases as te eat source s initial temperature increases. An increase in te inlet temperature of Terminol 66 in te generator wile te outlet temperature is fixed causes an increase in te generator temperature difference; terefore, te exergy destruction of te generator increases and te exergy efficiency decreases m=0.157 m=0.191 m=0.257 m=0.312 t (%) T evap ( C) Fig 7. Effect of te evaporator temperature on te termal efficiency at various entrainment ratio m=0.157 m=0.191 m=0.257 m=0.312 ex (%) T evap ( C) Fig 8. Effect of te evaporator temperature on te exergy efficiency at various entrainment ratio

13 Hamed Mortazavi Beni et al./ Energy Equip. Sys. / Vol. 5/No.3/Sep T B =150 ( C) T B =160 ( C) T B =170 ( C) ex (%) t (%) Termal efficiency T evap ( C) Fig 9. Effect of te evaporator temperature on te exergy efficiency in comparison wit te termal efficiency at various eat source initial temperature 7.Conclusion Te energy and exergy analysis of a novel cogeneration organic Rankine power and twostage compression refrigeration cycle was conducted by using EES software. Te results sowed tat te new cycle in te basic operation mode as percent improvement in te termal efficiency and percent improvement in te exergy efficiency compared to te conventional combined power and ejector refrigeration cycle. As observed in te results, because of te existence many components in te cycle, te exergy efficiency is more influenced by te variations of some parameters, suc as te eat source s initial temperature and te turbine inlet pressure, and less influenced by te variations of some oter parameters, suc as te evaporator temperature and te cascade condenser temperature. Even toug te net work output of te cycle is reduced, a significant increase occurs in te refrigeration capacity of te cycle. Te advantages of using tis new cogeneration organic Rankine power and two-stage compression refrigeration cycle includes te iger termal efficiency, te iger exergy efficiency, and increased access to more refrigeration capacity at te lower temperature compared to te conventional combined power and ejector refrigeration cycle. In te conventional ejector refrigeration cycle, refrigeration cannot be acieved at very low cooling temperatures. Te main idea of presenting tis new cycle is based on iger refrigeration production at lower temperatures. Since tis cycle as substantially improved te termal and exergy efficiency, it is recommended to use tis instead of te conventional combined power and ejector refrigeration cycle. References [1] Zang Z., Wang H., Tian L., Huang C., Termodynamic Analysis of Double- Compression Flas Intercooling Transcritical CO2 Refrigeration Cycle, Te Journal of Supercritical Fluids (2016)109: [2] Li H., Cao F., Bu X., Wang L., Wang X., Performance Caracteristics of R1234yf Ejector-Expansion Refrigeration Cycle, Applied Energy (2014)121: [3] Wang F., Li D.Y., Zou Y., Analysis for te Ejector Used as Expansion Valve in Vapor Compression Refrigeration Cycle, Applied Termal Engineering (2016) 96: [4] Lawrence N., Elbel S., Teoretical and Practical Comparison of Two-Pase Ejector Refrigeration Cycles Including First and Second Law analysis, International Journal of Refrigeration (2013)36: [5] Zang Z., Tong L., Cang L., Cen Y., Wang X., Energetic and Exergetic Analysis of an Ejector-Expansion Refrigeration Cycle Using te Working Fluid R32, Entropy (2015) 17: [6] Elgendy E., Parametric Study of a Vapor Compression Refrigeration Cycle Using a

14 312 Hamed Mortazavi Beni et al./ Energy Equip. Sys. / Vol. 5/No.3/Sep Two-Pase Constant Area Ejector, International Journal of Mecanical, Aerospace, Industrial, Mecatronic and Manufacturing Engineering (2013) 7. [7] Xing M., Yan G., Yu J., Performance Evaluation of an Ejector Subcooled Vapor-Compression Refrigeration Cycle, Energy Conversion and Management, (2015)92: [8] Wang X., Yu J., Zou M., Lv X., Comparative Studies of Ejector- Expansion Vapor Compression Refrigeration Cycles for Applications in Domestic Refrigerator-Freezers, Energy, (2014)70: [9] Zu Y., Jiang P., Hybrid Vapor Compression Refrigeration System wit an Integrated Ejector Cooling Cycle, Applied Energy (2012) 35: [10] Cen J., Havtun H., Palm B., Parametric Analysis of Ejector Working Caracteristics in te Refrigeration System, Applied Termal Engineering (2014)69: [11] Sorouradin A., Saberi A., Mamoudi S.M., Development of New Model to Prediction te Performance of Ejector Refrigeration Cycle, Modares Mecanical Engineering (2012)12: [12] Adrian G., Dobrovicescu A., Grosu L., Cerna E., Energy and Exergy Analysis of an Ejector Refrigeration System, U.P.B. Sci. Bull (2013)75. [13] Lontsi F., Hamandjoda O., Mayi O.T., Kemajou A., Development and Performance Analysis of a Multi- Temperature Combined Compression/Ejection Refrigeration Cycle Using Environment Friendly Refrigerants, International Journal of Refrigeration (2016)69: [14] Dai Y., Wang J., Gao L., Exergy Analysis, Parametric Analysis and Optimization for a Novel Combined Power and Ejector Refrigeration Cycle, Applied Termal Engineering (2005) 29: ,. [15] Zeng B., Weng Y.W., A Combined Power and Ejector Refrigeration Cycle for Low Temperature Heat Source, Solar Energy (2010) 84: [16] Habibzade A., Rasidi M.M., Galanis N., Analysis of a Combined Power and Ejector Refrigeration Cycle Using Low Temperature Heat, Energy Conversion and Management (2013) 65: [17] Wang J., Dai Y., Sun Z., A Teoretical Study on a Novel Combined Power and Ejector Refrigeration Cycle, International Journal of Refrigeration (2009)32: [18] Kim, K., Han C., Kim S., Ko H., Performance Analysis of a Combined Power and Ejector Refrigeration Cycle for Different Working Fluids ", International Journal of Mining, Metallurgy & Mecanical Engineering (2013) 1. [19] Yang X., Zao L., Li H., Yu Z., Teoretical Analysis of a Combined Power and Ejector Refrigeration Cycle Using Zeotropic Mixture, Applied Energy (2015) 160: [20] Yang X., Zao L., Termodynamic Analysis of a Combined Power and Ejector Refrigeration Cycle Using Zeotropic Mixtures, Energy Procedia (2015) 75: [21] Yang X., Zeng N., Zao L., Deng S., Li H., Yu Z., Analysis of a Novel Combined Power and Ejector- Refrigeration Cycle, Energy Conversion and Management (2016) 108: [22] Dincer I., Rosen M. A., Exergy, Energy, Environment and Sustainable Development, Elsevier (2007). [23] Brunin O., Feidt M., Hivet B., Comparison of te Working Domains of Some Compression Heat Pumps and a Compression-Absorption Heat Pump, International Journal of Refrigeration (1997)20: [24] Keenan H., Neumann E.P., Lustwerk F., An Investigation of Ejector Design by Analysis and Experiment, Journal of Applied Mecanics Transactions of te ASME (1950)72: [25] Huang B.J., Cang J.M., Wang C.P., Petrenko V.A., A 1-D Analysis of Ejector Performance, International Journal of Refrigeration (1999) 22: [26] Sarkar J., Performance Caracteristics of Natural-Refrigerants Based Ejector Expansion Refrigeration Cycles, Journal of Power and Energy (2009)223: