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Transcritical organic Rankine vapor compression refrigeration system for intercity bus air-conditioning using engine exhaust heat
Energy xxx (2015) 1e10
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Transcritical organic Rankine vapor compression refrigeration system for intercity bus air-conditioning using engine exhaust heat Alper Yilmaz* Department of Automotive Engineering, Çukurova University, 01330 Adana, Turkey
a r t i c l e i n f o
a b s t r a c t
Article history: Received 16 May 2014 Received in revised form 8 December 2014 Accepted 2 February 2015 Available online xxx
T-ORVC (Transcritical organic Rankine vapor compression) refrigeration system for intercity bus airconditioning using engine exhaust heat is investigated. Wet refrigerant R134a and dry refrigerant R245fa are considered as refrigerants. Typical efficiencies for pump, turbine and compressor utilized in literature (namely LV (low values), MV (medium values) and HV (high values)) are compared for the influence on COP (coefficient of performance) of the system analyzed. The results are strongly dependent upon the efficiencies. COPH (coefficient of performance for heat energy) used for the heating of the refrigerant, coefficient of performance for electrical energy (COPW) needed for the pump and coefficient of performance for both heat energy and electrical energy (COPT) needed together for the heating of the refrigerant and for the pump (COPT) are calculated. Besides the isentropic efficiencies, mechanical, electromotor and electric energy production efficiencies are also taken into account. Using MV efficiencies, COPT values at 8 MPa pressure can be as high as 0.54 and 0.66 at 180 and 200 C temperatures for R134a and R245fa, respectively. COPW values are very high and therefore electrical energy need is very low as compared to exhaust heat energy need. Assuming 30 kW of cooling load for an intercity bus and under different engine loads, all heat exchanger properties are determined. The results for a cooling load of 30 kW show that using T-ORVC (organic Rankine vapor compression cycle) system, air-conditioning of the intercity buses can be realized using waste heat energy of engine exhaust gases. © 2015 Elsevier Ltd. All rights reserved.
Keywords: Transcritical cycle Vapor compression cycle Rankine cycle Intercity bus air-conditioning Engine exhaust heat
1. Introduction Air-conditioning is crucial for automobile industry. The need is huge especially for intercity busses. It takes 5e10 kW of mechanical energy from the main engine. Because it is very important to lower the use of fuels both for economical and environmental purposes, it is good to save this mechanical energy needed for air-conditioning. Instead, huge waste heat of exhaust gases should be used for airconditioning. There are different ways of air-conditioning using waste heat [1]. Systems applied to solar energy refrigeration [2e4] can also be used in automobile industry. Absorption and adsorption systems need two fluids or fluid þ solid phases. In mobile systems, these two phase cooling systems can be considered as not appropriate because they would increase
* Tel.: þ90 (322) 338 60 84. E-mail address: [email protected].
the maintenance and decrease the reliability. In the remaining two systems, namely ORVC (organic Rankine vapor compression) system and ER (ejector refrigeration) system, higher heat source temperatures must be utilized, such as exhaust gas temperatures, to obtain higher COP (coefficient of performance) values. Therefore, appropriate refrigerants must be selected for the subcritical cycles; otherwise, transcritical cycles must be considered. Transcritical cycles are normally used for vapor compression refrigeration using CO2 [5]. Many researchers have investigated transcritical CO2 refrigeration cycles [6e9]. Compared to ORVC, transcritical cycle has the advantage that wet refrigerants [2] can also be used without any condensation in turbines. For ORVC refrigeration systems, Little and Garimella [1] supplied results for 60 C and 120 C waste heat temperatures using R245fa as the refrigerant. Wang et al. [10] investigated experimentally ORVC using R245fa, where both power and refrigeration cycles were independent on each other. Wang et al. [11] presented a theoretical work with coupled expander and compressor using R245fa. They showed the influence of different parameters on COP.
http://dx.doi.org/10.1016/j.energy.2015.02.004 0360-5442/© 2015 Elsevier Ltd. All rights reserved.
Please cite this article in press as: Yilmaz A, Transcritical organic Rankine vapor compression refrigeration system for intercity bus airconditioning using engine exhaust heat, Energy (2015), http://dx.doi.org/10.1016/j.energy.2015.02.004
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Nomenclature
htm 4
COPH COPT COPW cp F h M_ P s T U w W
COP, eq. (14) COP, eq. (16) COP, eq. (15) specific heat, kJ kg1 K1 heat transfer area, m2 enthalpy, kJ kg1 mass flow rate, kg s1 pressure, MPa entropy, kJ kg1 K1 temperature, C overall heat transfer coefficient, Wm2K1 mass flow ratio, eq. (10) work, kW
Greek
Dhie DTSC DTSH DT m DTie h htc hci hcm he hpem hpi hpm hti
inlet-exit enthalpy difference subcooling temperature difference, C superheating temperature difference, C logarithmic mean temperature difference, C inlet-exit temperature difference, C efficiency turbine-compressor total isentropic-mechanical efficiency, eq. (11) compressor isentropic efficiency, eq. (4) compressor mechanical efficiency, eq. (9) electricity production efficiency, eq. (16) pump electromotor efficiency, eq. (15) pump isentropic efficiency, eq. (3) pump mechanical efficiency, eq. (16) turbine isentropic efficiency, eq. (6)
Aphornratana and Sriveerakul [12] used a piston-cylinder system for expanderecompressor combination and obtained COP values of 0.25 for generator, condenser and evaporator temperatures of 60 C, 35 C and 5 C, respectively. Jeong and Kang [13] analyzed coupled compressor-turbine system with different working fluids and recommended R245fa as the most promising refrigerant. Li et al. [14] investigated ORVC refrigeration system using hydrocarbons R290, R600, R600a and R1270 and determined COP's dependent upon evaporation and condensation temperatures. Saidur et al. [15] reviewed different technologies such as thermoelectric generators, organic Rankine cycle and turbocharger technology. They stated that 10% fuel can be saved using ORC (organic Rankine cycle). Wang et al. [16] reviewed thermal exhaust heat recovery using Rankine cycle and gave information about vehicle exhaust gas and refrigerant properties. Yu et al. [17] determined up to 5.8% increment in thermal efficiency of a diesel engine with ORC using R245fa. Shu et al. [18] analyzed a dual-loop (high temperature and low temperature) ORC system and showed that waste heat thermal efficiency can be as high as 20%. Weerasinghe et al. [19] showed that using ORC and turbocompounding, a fuel economy of 20% can be achieved. Zhu et al. [20] determined exhaust heat recovery till 14% using R113. Refrigerants which should be used in T-ORVC should have certain properties. These main criteria for the selections are chemical stability, environmental friendliness, non-toxicity and low flammability [16,21e27].
turbine mechanical efficiency, eq. (8) relative humidity, %
Subscripts a air c compressor con condenser e electricity production, exit em electromotor eva evaporator g exhaust gas i isentropic in indoor m mechanical, mean o oil or oil-refrigerant heat exchanger out outdoor p pump r refrigerant t turbine Abbreviations COP coefficient of performance ERC ejector refrigeration cycle go exhaust gas-oil heat exchanger HV high value LV low value MV medium value or oil-refrigerant heat exchanger ORC organic Rankine cycle ORVC organic Rankine vapor compression cycle T-ORVC transcritical ORVC
Dry refrigerants have positive slope of saturated vapor line in the T-s diagram. Therefore, there is no danger of condensation and turbine blade damage because of the erosion problem in using these refrigerants. Refrigerants with negative slope of vapor saturation line in the T-s diagram are named as wet refrigerants. If this slope is vertical, refrigerants are named as isentropic refrigerants. Fluids with positive slope are named as dry refrigerants [2,24,27]. R134a is one of the most used refrigerants. This refrigerant is also mostly used for bus air-conditioning. R134a should be named as wet refrigerant [26] because saturation line has negative slope; however, mostly it is named as isentropic refrigerant [2,16,23,27,28]. Refrigerant R134a entering into the turbine at saturation temperature would not condense, if the turbine has a LV (low values) isentropic efficiency. Therefore such kind of refrigerants can be named as near-isentropic refrigerants similar to near-azeotropic refrigerants. In this work, the temperature of the refrigerants are selected in such a way that no condensation occurs even at HV (high values) efficiencies. Velez et al. [29] pointed out that R134a and R245fa are good candidates for organic Rankine cycles for low temperature applications. Low temperature is considered for T < 230 C [30]. To obtain the highest efficiency for temperatures less than 200 C, organic Rankine cycles using R134a and R245fa should be preferred [31]. Decomposition temperature of organic fluids are greater than 200 C [18]. Decomposition temperature of R245fa exceeds 250 C [32] and it can be used till a temperature of 226.9 C [33].
Please cite this article in press as: Yilmaz A, Transcritical organic Rankine vapor compression refrigeration system for intercity bus airconditioning using engine exhaust heat, Energy (2015), http://dx.doi.org/10.1016/j.energy.2015.02.004
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3
From the above explanations, it can be inferred that most of the criteria for working fluids can be fulfilled by R134a and R245fa. Besides this, R134a is the most used refrigerant for vehicle airconditioning. There is not any investigation for refrigeration using exhaust waste heat in transcritical cycles. Here, especially the T-ORVC refrigeration system for wet (nearisentropic) refrigerant R134a and dry refrigerant R245fa are investigated.
2. Transcritical organic Rankine vapor compression refrigeration system T-ORVC is schematically shown in Fig. 1. The processes in this system are illustrated in Fig. 2 on a P-h diagram. Refrigerant is pumped to a pressure P11 which is higher than the critical pressure Pcr. Point 10 is arrived, when the pump compresses the refrigerant isentropically. In the oil-refrigerant heat exchanger, refrigerant is heated till the temperature T12 below the exhaust inlet gas temperature. Oil is heated in exhaust gas-oil heat exchanger and it is circulated via an oil pump. The heated high pressure refrigerant enters the turbine and leaves it at the condensation pressure P14. Point 13 is arrived when the refrigerant is expanded isentropically to the condenser pressure P4. At the same time, compressor that is coupled to the turbine, compresses the refrigerant leaving the evaporator at the evaporation pressure P1 to the condenser pressure P2 ¼ P14. Point 3 is arrived when the refrigerant is compressed to the condenser pressure isentropically. Both gases leaving the turbine and compressor are mixed at point 15 and transported to the condenser. After condensation at point 5, refrigerant is
Fig. 2. Schematical presentation of T-ORVC refrigeration system in p-h diagram.
subcooled till the temperature T6. Some part of the liquid refrigerant enters the expansion valve and afterwards enters the evaporator to get refrigeration at the evaporator temperature T7. Heat transfer at the expansion device can be neglected so that the pressure decrease process from point 6 to point 7 can be assumed isenthalpic. Remaining part of the refrigerant leaving the condenser enters the pump and the cycles are repeated.
Fig. 1. Schematical presentation of T-ORVC refrigeration system.
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In this refrigeration system, recuperator after the turbine is not foreseen, because the exit of the turbine is near the saturation curve and therefore a recuperator is not needed [34]. Heat transfer from the exhaust gas to the system is achieved using an oil pump, two heat exchangers (namely gas-oil and oilrefrigerant) and heat transfer oil as presented in Fig. 1. These two cycles, namely the transcritical power and the subcritical refrigeration cycle, are modeled in the following thermodynamic analysis. 3. Thermodynamic analysis Evaporator refrigeration temperature is primarily determined by the inlet temperature of the fluid to be cooled. Depending on the evaporator and the fluid to be cooled, refrigerant evaporation temperature Te can be assumed 10e15 C below the inlet temperature of the fluid to be cooled. Because of this, Te ¼ T7 will be prescribed. Similarly, refrigerant condensation temperature depends primarily on the cooling fluid and this condensation temperature Tc can be assumed 10e15 C above the inlet temperature of this cooling fluid. Therefore, condensation temperature Tc ¼ T5 will also be prescribed. Parallel to Te and Tc, the subcooling temperature difference in the condenser DTSC and the superheating temperature difference in the evaporator DTSH will also be prescribed. The refrigerant pressure after the pump P11 and the refrigerant temperature after the heater T12 are also given. For the analysis, the isentropic and mechanical efficiencies of the turbine, compressor and pump must be given. Normally, the isentropic efficiencies are dependent at least on pressure ratio. However, they are assumed as constant for the sake of simplicity in this work. Besides them, one needs the electromotor efficiency of the pump and for the comparison of the mechanical work with heat, one needs the electricity production efficiency he using heat energy. Comparing different works in literature [1,5,10,11,18,20,35e41], different efficiencies given in literature are listed in Table 1. According to these values, three different groups of efficiencies are given which are classified as HV (high values), MV (medium values) and LV (low values). In this work, these three cases are considered for the calculations to see the influence of these different groups of efficiencies. These groups are given in Table 2. The pump pressure P11 ¼ P12 and the temperature T12 after the oil-refrigerant heat exchanger are varied to show the influence of these parameters on the COP. As mentioned above, evaporator and condenser temperatures are given. From these values, evaporator and condenser pressures are determined. Because DTSC and DTSH are also given, one can determine compressor inlet temperature T1 and condenser outlet
Table 2 High values (HV), medium values (MV) and low values (LV) of turbine, compressor and pump efficiencies.
HV MV LV
hti
hci
hpi
htm
hcm
hpm
hpe
he
0.90 0.80 0.70
0.80 0.75 0.65
0.90 0.80 0.75
1.00 0.97 0.95
1.00 0.97 0.95
1.00 0.97 0.95
0.95 0.92 0.90
0.35 0.30 0.25
temperature T6. From these values, one can easily calculate h6, s6 and h1, s1. Because of the isentropic compression process in the pump and in the compressor, one can write (point 3 and point 10 are the compressor and pump exits by assuming isentropic compressions):
s3 ¼ s1
(1)
s10 ¼ s6
(2)
Because evaporator pressure Pe and the pump exit pressure P11 are given, the enthalpies at the points 3 and 10 can be determined using the compressor and pump isentropic efficiencies. Enthalpies at the pump and compressor exit can be obtained using the following expressions:
h11 ¼ h6 þ
h2 ¼ h1 þ
h10 h6 hpi
h3 h1 hci
(3)
(4)
The temperature T12 after the oil-refrigerant heat exchanger is given. Pump exit pressure is also known. Enthalpy h13 at the point 13 is calculated with the assumption of isentropic expansion from the heater pressure to the condenser pressure and the following expression for entropy can be written at point 13:
s13 ¼ s12
(5)
Condenser pressure is given and therefore, h13 can be obtained. The enthalpy at the exit of the turbine is determined using the isentropic efficiency of the turbine:
h14 ¼ h12 hti ðh12 h13 Þ
(6)
In the expansion valve, the process is assumed as isenthalpic. Therefore,
h7 ¼ h6
(7)
Network produced by the turbine is
_ t ¼ M_ t ðh h Þh W 12 14 tm
(8)
The work needed by the compressor is
Table 1 Efficiencies used in literature. Literature
hti
hci
hpi
htm
hcm
hpm
hpe
he
[1] [10] [11] [41] [5] [18] [38] [37] [20] [35] [36] [40] [39]
0.90 0.75 0.75 0.80 0.85 0.75 0.85 0.70 0.80 0.85 0.80 0.70 0.80
0.65 e 0.80 e 0.80 e e 0.80 e e e e e
0.90 e e 0.75 e 0.80 0.85 0.80 0.80 e e 0.80 0.75
1.00 0.95 1.00 1.00 e e 0.95 e e e e e e
1.00 0.95 1.00 e e e e e e e e e e
1.00 e e 1.00 e e 0.95 e e e e e e
e e e e e e e e e e e e e
e e 0.25 e e e e e e e e e e
. _ c ¼ M_ c ðh h Þ h W 2 1 cm
(9)
htm and hcm are the mechanical efficiencies of the turbine and the compressor, respectively. With the definitions of mass flow rate ratio w and total efficiencies of turbine and compressor; w¼
M_ c M_ t
htc ¼ hti hci htm hcm
(10)
(11)
the following expression for mass flow rate ratio is obtained using eqs. ((4) and (6) and (8)e(11):
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w ¼ htc
h12 h13 h3 h1
(12)
Heat produced because of the mechanical efficiencies of the turbine and compressors are assumed to be transferred into ambient. Enthalpy of the fluid at the inlet of the condenser can be determined as:
h15 ¼
h14 þ wh2 1þw
(13)
It is possible to calculate COP of the T-ORVC using these equations. The following equation for the COP is obtained taking only the heat supplied to the refrigerant in the oil-refrigerant heat exchanger into consideration;
COPH ¼
wðh1 h7 Þ h12 h11
(14)
whereas the following equation is obtained taking only the work given to the pump into consideration:
COPW ¼
wðh1 h7 Þhpm hpem h11 h6
(15)
Here, hpm and hpem are mechanical efficiency of the pump and the electromotor efficiency of the pump, respectively. The best way for the definition of COP in this system would be accomplished if both heat supplied to the refrigerant and the heat equivalence of the electrical energy supplied to the pump are taken into consideration. The following equation is derived for overall COP (COPT);
COPT ¼
wðh1 h7 Þ ðh11 h6 Þ hpm hpem he
þ ðh12 h11 Þ
(16)
where he is the efficiency of the electricity production using heat energy.
5
Table 3 Comparison between the results given by Little and Garimella [1] and this work for the refrigerant R245fa (hti ¼ 0.90, hci ¼ 0.65, hpi ¼ 0.90, htm ¼ 1, hcm ¼ 1, hpm ¼ 1, hpe ¼ 1, he ¼ 1). P12 [kPa]
T12 [ C]
1519 419 2287
108.3 56.5 128.9
P5
142 333
T5 [ C]
P8
T8 [ C]
DTSC ¼ DTSH [ C]
COP [1].
COPH this work
Error [%]
47.3 23.6 48.8
66 63 53
4.7 3.8 0.0
3 1 3
0.413 0.676 0.392
0.426 0.695 0.414
3.0 2.7 5.3
Refrigerant-exhaust gas heat exchanger system consists of exhaust gaseoil (go) and oilerefrigerant (or) heat exchangers. In this system, a heat resistant heat transfer oil with an oil pump must be used. This exhaust gas heat exchanger system is also proposed by Yu et al. [17]. All the seven heat exchangers can be calculated assuming counter flow heat exchangers using the general heat exchanger equations:
Q_ ¼ ðUFÞDTm
(17)
_ p DT Q_ ¼ Mc ie
(18)
_ Q_ ¼ MDh ie
(19)
Q_ , (UF) and DTm are heat transferred, overall heat transfer coefficient-heat transferring surface area and logarithmic mean _ cp and DTie are mass flow rate, specific temperature difference.M, heat and inlet-exit temperature difference and Dhie inlet-exit enthalpy difference of the fluids. In this work, the data of a 6-cylinder, 4-stroke turbocharged diesel engine are used. The data used are given by Shu et al. [40]. Here, the components of the exhaust gas are also supplied for the engine loads between 50% and 100%, the mean values of the components of the exhaust gas are as follows:
CO2 ¼ 0:140; H2 O ¼ 0:055; O2 ¼ 0:0715; N2 ¼ 0:7335 (20)
4. Heat exchangers' analysis In the system analyzed, there are evaporator, condenser and exhaust gas heating system for heat transfer. All heat exchangers are assumed as counter flow heat exchangers. Condenser is cooled by outside air and in the evaporator, fresh outside air and indoor air mixture is cooled. Heat exchangers' temperature distributions are shown in Fig. 3. Evaporator can be calculated as heat exchangers eva 1 and eva 2. Similarly, condenser can be divided into subcooling, condensing and superheating parts which are designated as con 1, con 2 and con 3, respectively.
Using these data in eq. (20), cpg for different temperatures is calculated and the following equation for the specific heat of the exhaust gas is derived:
Tgm ¼ 30 C 530 C
:
cpg ¼ 1:043 þ 0:023 Tgm 100 (21)
cp values for different gases are used from the literature [42,43]. Here, cpg and Tgm are exhaust gas mean specific heat and mean temperature, respectively. A similar equation is given by Yang et al. [44].
Fig. 3. Temperature distributions in heat exchangers. a) Evaporator, b) Condenser, c) Exhaust gas heat exchanger system.
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Fig. 4. Variation of COPH with temperature for different pressures P12 and medium efficiencies for R134a.
For air, the following equation is derived;
Tam ¼ 0 C 100 C
:
cpa ¼ 1:006 þ 0:006ðTam =100Þ1:5 (22)
where cpa and Tam are mean specific heat and mean temperature of air. The data given in Ref. [43] are used for cpa. In this work, “Dowtherm G” is foreseen for heat transfer oil and its specific heat cpo is described as a function of its temperature as follows:
Tom ¼ 0 C 360 C
:
cpo ¼ 1:530 þ 0:233ðTom =100Þ
(23)
For the derivation of the above equation, cpo values given in Ref. [43] are used. For real fluids such as refrigerant and humid air, the following formula for the calculation of the specific heat is used:
Fig. 6. Variation of COPW with temperature for different pressures P12 and medium efficiencies for R134a.
subcritical systems if T12 is assumed greater than the saturation temperature for given values of pump exit pressure. However, only in the work given by Little and Garimella [1] are given all the parameters necessary for a comparison. They used R245fa as the refrigerant and used the same COP definition as the COPH definition by eq. (14) in this work. The same parameters used by Little and Garimella [1] are used for the calculations in this work for comparison. The parameters used and the results are presented in Table 3. As seen from this table, the difference between the COP given by Little and Garimella and the COPH determined in this work are between 2.7 and 5.3%. These differences can be considered as acceptable and it is expected that the calculations for T-ORVC should also produce acceptable results.
6. Results and discussions for the refrigerant cycle
There are not any investigation to compare with the system described in this work. But, there are some investigations for subcritical systems. The system described here is also valid for
For summer air-conditioning in intercity buses, indoor temperature can be assumed as 25 C [45,46]. Because of the dehumidifying duty, it would be appropriate to select refrigerant evaporation temperature of Te ¼ 10 C. Outdoor design conditions can be assumed as 35 C for vehicle design [11]. The highest design condition is 42 C [46]. Therefore, adequate refrigerant condensation temperature is selected as Tc ¼ 50 C. Both superheating and subcooling are assumed as DTSC ¼ DTSH ¼ 3 C, which can be considered as appropriate. In Figs. 4e6, COPH, COPT and COPW for the refrigerant R134a are demonstrated as a function of temperature using MV efficiencies. Pump exit pressure P12 is given as the parameter. Both subcritical and supercritical pressures are considered. There is an optimum
Fig. 5. Variation of COPT with temperature for different pressures P12 and medium efficiencies for R134a.
Fig. 7. Variation of COPH with temperature for different pressures P12 and medium efficiencies for R245fa.
cp ¼
hi he Ti Te
(24)
where hi, he, Ti and Te are inlet, exit enthalpies and inlet and exit temperatures of the fluids. In the above equations, the temperatures are in C and specific heats are in kJ/kg C. 5. Validation of the calculations
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Fig. 8. Variation of COPT with temperature for different pressures P12 and medium efficiencies for R245fa.
Fig. 10. Variation of maximum COPH and COPT values with pressure P12 for R134a and R245fa for medium efficiencies.
value for COPH and COPT at a certain temperature T12 for a given pump exit pressure. Beyond 6 MPa pressure, there is not any optimum point. COPH and COPT values increase very slightly after 180 C, so that optimum values can be approximately assumed as the values at 180 C. Optimum COPT values decrease when the pressure is greater than 8 MPa. The absolute highest COPT determined is 0.54 at 8 MPa. COPW values increase with temperature, but inverse to COPH and COPT, COPW decreases with increasing pressure, because the work needed for the pump increases with increasing pressure. COPW values are so high that mechanical energy needed for the pump is very low as compared to the heat energy supplied to the refrigerant. Results obtained for the refrigerant R245fa are depicted in Figs. 7e9. The results for R245fa are similar to the results obtained for R134a. For subcritical pump exit pressures, there are not any optimum values. COPH and COPT values slightly decrease with temperature. Beyond the critical pressure, COPH and COPT values slightly increase with temperature. Optimum values are determined for pressures between 4 and 6 MPa. COP values in these figures are greater than that are given in Figs. 4e6. Highest COPT value obtained using R245fa is 0.66 at 8 MPa pressure and 200 C temperature. Maximum values of COPH and COPT are illustrated in Fig. 10 dependent on pressure for both R134a and R245fa. Highest value of COPT ¼ 0.54 is obtained around 8 MPa for R134a and highest value of COPT till 200 C is obtained around 8 MPa for R245fa. There is no maximum COPT value for R245fa, but as the highest value of COPT ¼ 0.66 can be assumed as optimum. Because condensation temperature is assumed as 50 C, COPH and COPT values are zero at the saturation pressures of this temperature. The
difference between COPH and COPT values increases with increasing pressure. In Figs. 11 and 12, optimum temperatures at a given pressure for the optimum values of COPH and COPT are depicted. In these figures, saturation temperatures at given pressures are also drawn. Optimum temperatures are a little bit higher than saturation pressure till critical pressure. Beyond 8 MPa, there is little change in optimum temperature, which is approximately 180 C for R134a and 200 C for R245fa, respectively. COPT values for R134a and R245fa are also calculated using low (LV) and high (HV) efficiencies according to Table 2. From these calculations, one can conclude that COPT values increase with increasing efficiencies (LV, MV and HV) as expected. But even for LV efficiencies, the results can be considered as satisfactory. The results for HV (high values) and LV (low values) are approximately 30% higher and 30% lower than the results obtained for MV (medium values) for both refrigerants R134a and R245fa. Highest COPT values are 0.38, 0.54 and 0.71 for R134a using LV, MV and HV efficiencies, respectively. Highest COPT values are 0.47, 0.66 and 0.86 for R245fa using LV, MV and HV efficiencies, respectively. Optimum values for COPT can be assumed approximately 8 MPa pressure for all types of efficiencies. Optimum temperatures for LV and HV efficiencies are nearly the same as the values demonstrated in Figs. 11 and 12 for MV efficiencies. Exhaust temperatures of diesel engines of vehicles are around 400e700 C and waste thermal energy ranges from 20 kW to 400 kW [16,20,22]. It is clear that even with LV efficiencies, T-ORVC refrigeration system is capable of air-conditioning of intercity buses using exhaust heat.
Fig. 9. Variation of COPW with temperature for different pressures P12 and medium efficiencies for R245fa.
Fig. 11. Variation of optimum temperatures with pressure P12 for R134a for medium efficiencies.
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A. Yilmaz / Energy xxx (2015) 1e10 Table 6 Engine data for different engine loads [40]. Engine load [%] 50 75 90 100 Power output [kW] 117.7 176.2 211.6 235.8 Exhaust gas temperature [ C] 420 474 498 519 Exhaust gas mass flow rate [kg/s] 0.1697 0.2235 0.2586 0.2752
Table 7 The properties of the condenser for R134a (M_ a ¼ 8:038 kg=s, M_ r;con ¼ 0:4220 kg=s, Q_ con ¼ 80:98 kW).
Fig. 12. Variation of optimum temperatures with pressure P12 for R245fa for medium efficiencies.
Ta,i [ C]
Ta,e [ C]
Tr,i [ C]
Tr,e [ C]
DTm [ C] UF [W/ C]
1.98 64.06 14.94
35 35.24 43.15
35.24 43.15 45.00
50.0 50.0 81.0
47.0 50.0 50.0
13.35 10.30 17.57
148.5 6217.0 850.3
Table 8 The properties of the condenser for R245fa (M_ a ¼ 7:114 kg=s, M_ r;con ¼ 0:3657 kg=s, Q_ con ¼ 71:68 kW).
7. Results and discussions for heat exchangers Seven heat exchangers' properties are calculated for the optimum conditions P12 ¼ 80 bar and T12 ¼ 180 C for R134a and P12 ¼ 80 bar and T12 ¼ 200 C for R245fa. Evaporator load is assumed as Q_ eva ¼ 30 kW [45]. Under these assumptions, the temperatures, the enthalpies and other values are presented in Tables 4 and 5. In addition, engine exhaust gas temperature and exhaust gas mass flow rate changes by different engine loads are given in Table 6. Evaporator and condenser dimensions are independent of the engine load, because Q_ eva and the temperatures at the condenser and at the evaporator are prescribed. However, exhaust heat exchanger system changes with the engine load. Results for condenser are given in Tables 7 and 8 for R134a and R245fa, respectively. Outside air temperature increases from 35 C to 45 C as foreseen. It is seen from these tables that there are not much difference between the results for the two refrigerants. For the determination of the evaporator, indoor and outdoor air temperatures and humidities must be prescribed. They are taken as [45].
Con 1 Con 2 Con 3
Q_ con ½kW
Tin ¼ 25 C; 4in ¼ 0:50; Tout ¼ 35 C; 4out ¼ 0:40
(25)
Besides them, one needs fresh air mass flow rate, latent heat production in the bus and air exit temperature from the evaporator. They are assumed as [45]. V_ fresh ¼ 190l=s; Q_ l ¼ 3 kW; Ta;e ¼ 15 C
(26)
With these data and Q_ eva ¼ 30 kW, one can obtain the following equation using psychometric equations, mass balances for dry air and water vapor and the first law of thermodynamics:
Con 1 Con 2 Con 3
Q_ con ½kW
Ta,i [ C]
Ta,e [ C]
Tr,i [ C]
Tr,e [ C]
DTm [ C] UF [W/ C]
1.50 63.89 6.29
35.0 35.21 44.12
35.21 44.12 45.00
50.0 50.0 66.54
47.0 50.0 50.0
13.35 9.66 12.06
112.3 6616.0 521.5
M_ a ¼ 2:197 kg=s; Ta;i ¼ 25:98 C; ha;i ¼ 52:39 kJ=kg;
ha;e ¼ 38:74 kJ=kg
(27)
Here M_ a is dry air mass flow rate in the evaporator. Ta,i is air inlet temperature in the evaporator after mixing of the fresh air and indoor air. The properties of the evaporators eva 1 and eva 2 are given in Tables 9 and 10. It is seen that superheating part (eva 1) is very small. The evaporator is smaller than the condenser as expected. There are not important differences between the results for the refrigerants R134a and R245fa. For the determination of exhaust gas-oil (go) and oil-refrigerant (or) heat exchangers at different engine loads, it is assumed that oil temperatures at the inlet and exit of the heat exchangers are arithmetic mean temperatures of exhaust gas and refrigerant. Under this assumption, the properties of the exhaust gas-oil (go) and oil-refrigerant (or) heat exchangers are presented in Tables 11 and 12 for the refrigerants R134a and R245fa, respectively. From the tables, one can conclude that, there is no problem for air-conditioning even at 50% engine load because exhaust gas exit temperatures are high enough for both refrigerants. The UF values, even at 50% engine load, are very low as compared with the values of evaporator and especially condenser because of high logarithmic mean temperature differences.
Table 4 Data for the calculation of heat exchangers at different points of the cycles for the refrigerant R134a (P12 ¼ 8 MPa, T12 ¼ 180 C, Q_ eva ¼ 30 kW, Q_ con ¼ 80:98 kW, Q_ or ¼ 49:88 kW, M_ r;eva ¼ 0:2140 kg=s, M_ r;con ¼ 0:4220 kg=s, M_ r;or ¼ 0:2080 kg=s). Points in the cycles Temperature [ C] Enthalpy [kJ/kg]
1 13.0 259.0
4 50.0 275.3
5 50.0 123.5
6 47.0 118.8
7 10.0 118.8
8 10.0 256.2
11 52.75 126.2
12 180.0 366.1
15 81.0 310.7
Table 5 Data for the calculation of heat exchangers at different points of the cycles for the refrigerant R245fa (P12 ¼ 8 MPa, T12 ¼ 200 C, Q_ eva ¼ 30 kW, Q_ con ¼ 71:68 kW, Q_ or ¼ 40:83 kW, M_ r;eva ¼ 0:1970 kg=s, M_ r;con ¼ 0:3657 kg=s, M_ r;or ¼ 0:1687 kg=s). Points in the cycles Temperature [ C] Enthalpy [kJ/kg]
1 13.0 414.5
4 50.0 441.0
5 50.0 266.3
6 47.0 262.2
7 10.0 262.2
8 10.0 411.8
11 51.17 269.6
12 200.0 511.6
15 66.54 458.2
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Table 9 The properties of the evaporator for R134a (M_ a ¼ 2:197 kg=s, M_ r;eva ¼ 0:1970 kg=s, Q_ eva ¼ 30 kW).
Eva 1 Eva 2
Q_ eva ½kW
Tr,i [ C]
Tr,e [ C]
hr,i [kJ/kg]
hr,e [kJ/kg]
Ta,i [ C]
Ta,e [ C]
ha,i [kJ/kg]
ha,e [kJ/kg]
DTm [ C]
UF [W/ C]
0.53 29.47
10.0 10.0
13.0 10.0
411.8 262.2
414.5 411.8
25.98 25.78
25.78 15.00
52.39 52.14
52.14 38.74
14.34 9.59
37.10 3072.0
Table 10 The properties of the evaporator for R245fa (M_ a ¼ 2:197 kg=s, M_ r;eva ¼ 0:2140 kg=s, Q_ eva ¼ 30 kW).
Eva 1 Eva 2
Q_ eva ½kW
Tr,i [ C]
Tr,e [ C]
hr,i [kJ/kg]
hr,e [kJ/kg]
Ta,i [ C]
Ta,e [ C]
ha,i [kJ/kg]
ha,e [kJ/kg]
DTm [ C]
UF [W/ C]
0.60 29.40
10.0 10.0
13.0 13.0
256.2 118.8
259.0 256.2
25.98 25.76
25.76 25.00
52.39 52.12
52.12 38.74
14.33 9.37
41.83 3139.0
Table 11 The properties of the exhaust gas-oil (go) heat exchanger and oil-refrigerant (or) heat exchanger for R134a (M_ r;or ¼ 0:2079 kg=s, Q_ ¼ 49:88 kW, Tr,i ¼ 52.75 C, Tr,e ¼ 180.0 C, DTm,go ¼ DTm,or, (UF)go ¼ (UF)or). Engine load [%]
M_ g ½kg=s
M_ o ½kg=s
Tg,i [ C]
Tg,e [ C]
To,i [ C]
To,e [ C]
DTm [ C]
UF [W/ C]
100 90 75 50
0.2752 0.2586 0.2235 0.1697
0.1603 0.1566 0.1460 0.1271
519.0 498.0 474.0 420.0
360.6 328.5 276.4 155.0
349.5 339.0 327.0 300.0
206.7 190.6 164.6 103.9
161.6 148.2 128.6 103.9
308.7 336.6 387.8 617.8
From the results of the heat exchangers, it is seen that 30 kW cooling load for an intercity bus can be produced by optimum sizing of the cycle and appropriate choice of UF values of all heat exchangers by engine loads down to 50%. 8. Conclusions Transcritical ORVC refrigeration system is investigated for bus air-conditioning using exhaust heat as primary energy source. The following conclusions can be drawn out of this research: COPH and COPT values increase till the supercritical pressure of 8 MPa for both R134a and R245fa. COPH and COPT values increase till 180 C and 200 C for R134a and R245fa, respectively. The higher the efficiencies of the pump, the turbine and the compressor, the higher are the COPH and the COPT values COPW values are generally very high; therefore, electrical energy needed for the pump is very low compared to the heat energy needed. At optimum conditions, there is not any condensation in the turbine; therefore, no danger of turbine blade damages also. Optimum values for COPT for R134a are obtained at 180 C and 8 MPa. They are 0.38, 0.54 and 0.71 for LV, MV and HV efficiencies, respectively. Optimum values for COPT for R245fa are obtained at 200 C and 8 MPa. They are 0.47, 0.66 and 0.86 for LV, MV and HV efficiencies, respectively.
Table 12 The properties of the exhaust gas-oil (go) heat exchanger and oil-refrigerant (or) heat exchanger for R245fa (M_ r;or ¼ 0:1687 kg=s, Q_ ¼ 40:83 kW, Tr,i ¼ 51.17 C, Tr,e ¼ 200.0 C, DTm,go ¼ DTm,or, (UF)go ¼ (UF)or). Engine M_ g ½kg=s M_ o ½kg=s Tg,i load [%] [ C]
Tg,e [ C]
To,i [ C]
To,e [ C]
100 90 75 50
389.7 359.7 312.8 204.2
359.5 349.0 337.0 310.0
220.4 164.3 28.4 205.4 151.6 269.3 182.0 133.9 304.9 127.7 92.25 442.6
0.2752 0.2586 0.2235 0.1697
0.1331 0.1307 0.1234 0.1098
519.0 498.0 474.0 420.0
DT m [ C]
UF [W/ C]
Optimum COPT values for LV and HV efficiencies are approximately 30% lower and higher than the optimum values for MV efficiencies for both refrigerants R134a and R245fa. Optimum values for COPT of R245fa are approximately 23% higher than the optimum COPT values for R134a. Assuming 30 kW of cooling demand for an intercity bus, airconditioning is possible even at 50% engine load using both refrigerants R134a and R245fa. The UF values of exhaust gas-oil and oil-refrigerant heat exchangers are much lower than the values of evaporators and condensers. References [1] Little AB, Garimella S. Comparative assessment of alternative cycles for waste heat recovery and upgrade. Energy 2011;36:4492e504. [2] Pridasawas W. Solar-driven refrigeration systems with focus on the ejector cycle. Doctoral Thesis. Stockholm: Royal Institute of Technology; 2006. http:// www.diva-portal.org/. [3] Kim DS, Infante Ferreira CA. Solar refrigeration options-a state of the art review. Int J Refrig 2008;31:3e15. [4] Abdulateef JM, Sopian K, Alghoul MA, Sulaiman MY. Review of solar-driven ejector refrigeration technologies. Renew Sustain Energy Rev 2009;13: 1338e49. [5] Wang J, Zhao P, Niu X, Dai Y. Parametric analysis of a new combined cooling, heating and power system with transcritical CO2 driven by solar energy. Appl Energy 2012;94:58e64. [6] Wang H, Tian J, Hou X. On the coupled system performance of transcritical CO2 heat pump and Rankine cycle. Heat Mass Transf 2013;49:1733e40. [7] Sarkar J, Bhattacharyya S, Gopal MR. Transcritical CO2 heat pump systems: exergy analysis including heat transfer and fluid flow effects. Energy Convers Manag 2005;46:2053e67. [8] Elbel S. Historical and present developments of ejector refrigeration systems with emphasis on transcritical carbon dioxide air-conditioning applications. Int J Refrig 2011;34:1545e61. [9] Deng J, Jiang P, Lu T, Lu W. Particular characteristics of transcritical CO2 refrigeration cycle with an ejector. Appl Therm Eng 2007;27:381e8. [10] Wang H, Peterson R, Harada K, Miller E, Ingram-Gable R, Fisher L, et al. Performance of a combined organic Rankine cycle and vapor compression cycle for heat activated cooling. Energy 2011;36:447e58. [11] Wang H, Peterson R, Herron T. Design study of configurations on system COP for a combined ORC (organic Rankine cycle) and VCC (vapor compression cycle). Energy 2011;36:4809e20. [12] Aphornratana S, Sriveerakul T. Analysis of a combined Rankine-vaporcompression refrigeration cycle. Energy Convers Manag 2010;51:2557e64. [13] Jeong J, Kang YT. Analysis of a refrigeration cycle driven by refrigerant steam turbine. Int J Refrig 2004;27:33e41. [14] Li H, Bu X, Wang L, Long Z, Lian Y. Hydrocarbon working fluids for a Rankine cycle powered vapor compression refrigeration system using low-grade thermal energy. Energy Build 2013;65:167e72. [15] Saidur R, Rezaei M, Muzammil WK, Hassan MH, Paria S, Hasanuzzaman M. Technologies to recover exhaust heat from internal combustion engines. Renew Sustain Energy Rev 2012;16:5649e59. [16] Wang T, Zhang Y, Peng Z, Shu G. A review of researches on thermal exhaust heat recovery with Rankine cycle. Renew Sustain Energy Rev 2011;15:2862e71. [17] Yu G, Shu G, Tian H, Wei H, Liu L. Simulation and thermodynamic analysis of a bottoming organic Rankine cycle (ORC) of diesel engine (DE). Energy 2013;51: 281e90. [18] Shu G, Liu L, Tian H, Wei H, Yu G. Parametric and working fluid analysis of a dual- loop organic Rankine cycle (DORC) used in engine waste heat recovery. Appl Energy 2014;112:1188e98.
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Transcritical organic Rankine vapor compression refrigeration system for intercity bus air-conditioning using engine exhaust heat Alper Yilmaz* Department of Automotive Engineering, Çukurova University, 01330 Adana, Turkey
a r t i c l e i n f o
a b s t r a c t
Article history: Received 16 May 2014 Received in revised form 8 December 2014 Accepted 2 February 2015 Available online xxx
T-ORVC (Transcritical organic Rankine vapor compression) refrigeration system for intercity bus airconditioning using engine exhaust heat is investigated. Wet refrigerant R134a and dry refrigerant R245fa are considered as refrigerants. Typical efficiencies for pump, turbine and compressor utilized in literature (namely LV (low values), MV (medium values) and HV (high values)) are compared for the influence on COP (coefficient of performance) of the system analyzed. The results are strongly dependent upon the efficiencies. COPH (coefficient of performance for heat energy) used for the heating of the refrigerant, coefficient of performance for electrical energy (COPW) needed for the pump and coefficient of performance for both heat energy and electrical energy (COPT) needed together for the heating of the refrigerant and for the pump (COPT) are calculated. Besides the isentropic efficiencies, mechanical, electromotor and electric energy production efficiencies are also taken into account. Using MV efficiencies, COPT values at 8 MPa pressure can be as high as 0.54 and 0.66 at 180 and 200 C temperatures for R134a and R245fa, respectively. COPW values are very high and therefore electrical energy need is very low as compared to exhaust heat energy need. Assuming 30 kW of cooling load for an intercity bus and under different engine loads, all heat exchanger properties are determined. The results for a cooling load of 30 kW show that using T-ORVC (organic Rankine vapor compression cycle) system, air-conditioning of the intercity buses can be realized using waste heat energy of engine exhaust gases. © 2015 Elsevier Ltd. All rights reserved.
Keywords: Transcritical cycle Vapor compression cycle Rankine cycle Intercity bus air-conditioning Engine exhaust heat
1. Introduction Air-conditioning is crucial for automobile industry. The need is huge especially for intercity busses. It takes 5e10 kW of mechanical energy from the main engine. Because it is very important to lower the use of fuels both for economical and environmental purposes, it is good to save this mechanical energy needed for air-conditioning. Instead, huge waste heat of exhaust gases should be used for airconditioning. There are different ways of air-conditioning using waste heat [1]. Systems applied to solar energy refrigeration [2e4] can also be used in automobile industry. Absorption and adsorption systems need two fluids or fluid þ solid phases. In mobile systems, these two phase cooling systems can be considered as not appropriate because they would increase
* Tel.: þ90 (322) 338 60 84. E-mail address: [email protected].
the maintenance and decrease the reliability. In the remaining two systems, namely ORVC (organic Rankine vapor compression) system and ER (ejector refrigeration) system, higher heat source temperatures must be utilized, such as exhaust gas temperatures, to obtain higher COP (coefficient of performance) values. Therefore, appropriate refrigerants must be selected for the subcritical cycles; otherwise, transcritical cycles must be considered. Transcritical cycles are normally used for vapor compression refrigeration using CO2 [5]. Many researchers have investigated transcritical CO2 refrigeration cycles [6e9]. Compared to ORVC, transcritical cycle has the advantage that wet refrigerants [2] can also be used without any condensation in turbines. For ORVC refrigeration systems, Little and Garimella [1] supplied results for 60 C and 120 C waste heat temperatures using R245fa as the refrigerant. Wang et al. [10] investigated experimentally ORVC using R245fa, where both power and refrigeration cycles were independent on each other. Wang et al. [11] presented a theoretical work with coupled expander and compressor using R245fa. They showed the influence of different parameters on COP.
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Nomenclature
htm 4
COPH COPT COPW cp F h M_ P s T U w W
COP, eq. (14) COP, eq. (16) COP, eq. (15) specific heat, kJ kg1 K1 heat transfer area, m2 enthalpy, kJ kg1 mass flow rate, kg s1 pressure, MPa entropy, kJ kg1 K1 temperature, C overall heat transfer coefficient, Wm2K1 mass flow ratio, eq. (10) work, kW
Greek
Dhie DTSC DTSH DT m DTie h htc hci hcm he hpem hpi hpm hti
inlet-exit enthalpy difference subcooling temperature difference, C superheating temperature difference, C logarithmic mean temperature difference, C inlet-exit temperature difference, C efficiency turbine-compressor total isentropic-mechanical efficiency, eq. (11) compressor isentropic efficiency, eq. (4) compressor mechanical efficiency, eq. (9) electricity production efficiency, eq. (16) pump electromotor efficiency, eq. (15) pump isentropic efficiency, eq. (3) pump mechanical efficiency, eq. (16) turbine isentropic efficiency, eq. (6)
Aphornratana and Sriveerakul [12] used a piston-cylinder system for expanderecompressor combination and obtained COP values of 0.25 for generator, condenser and evaporator temperatures of 60 C, 35 C and 5 C, respectively. Jeong and Kang [13] analyzed coupled compressor-turbine system with different working fluids and recommended R245fa as the most promising refrigerant. Li et al. [14] investigated ORVC refrigeration system using hydrocarbons R290, R600, R600a and R1270 and determined COP's dependent upon evaporation and condensation temperatures. Saidur et al. [15] reviewed different technologies such as thermoelectric generators, organic Rankine cycle and turbocharger technology. They stated that 10% fuel can be saved using ORC (organic Rankine cycle). Wang et al. [16] reviewed thermal exhaust heat recovery using Rankine cycle and gave information about vehicle exhaust gas and refrigerant properties. Yu et al. [17] determined up to 5.8% increment in thermal efficiency of a diesel engine with ORC using R245fa. Shu et al. [18] analyzed a dual-loop (high temperature and low temperature) ORC system and showed that waste heat thermal efficiency can be as high as 20%. Weerasinghe et al. [19] showed that using ORC and turbocompounding, a fuel economy of 20% can be achieved. Zhu et al. [20] determined exhaust heat recovery till 14% using R113. Refrigerants which should be used in T-ORVC should have certain properties. These main criteria for the selections are chemical stability, environmental friendliness, non-toxicity and low flammability [16,21e27].
turbine mechanical efficiency, eq. (8) relative humidity, %
Subscripts a air c compressor con condenser e electricity production, exit em electromotor eva evaporator g exhaust gas i isentropic in indoor m mechanical, mean o oil or oil-refrigerant heat exchanger out outdoor p pump r refrigerant t turbine Abbreviations COP coefficient of performance ERC ejector refrigeration cycle go exhaust gas-oil heat exchanger HV high value LV low value MV medium value or oil-refrigerant heat exchanger ORC organic Rankine cycle ORVC organic Rankine vapor compression cycle T-ORVC transcritical ORVC
Dry refrigerants have positive slope of saturated vapor line in the T-s diagram. Therefore, there is no danger of condensation and turbine blade damage because of the erosion problem in using these refrigerants. Refrigerants with negative slope of vapor saturation line in the T-s diagram are named as wet refrigerants. If this slope is vertical, refrigerants are named as isentropic refrigerants. Fluids with positive slope are named as dry refrigerants [2,24,27]. R134a is one of the most used refrigerants. This refrigerant is also mostly used for bus air-conditioning. R134a should be named as wet refrigerant [26] because saturation line has negative slope; however, mostly it is named as isentropic refrigerant [2,16,23,27,28]. Refrigerant R134a entering into the turbine at saturation temperature would not condense, if the turbine has a LV (low values) isentropic efficiency. Therefore such kind of refrigerants can be named as near-isentropic refrigerants similar to near-azeotropic refrigerants. In this work, the temperature of the refrigerants are selected in such a way that no condensation occurs even at HV (high values) efficiencies. Velez et al. [29] pointed out that R134a and R245fa are good candidates for organic Rankine cycles for low temperature applications. Low temperature is considered for T < 230 C [30]. To obtain the highest efficiency for temperatures less than 200 C, organic Rankine cycles using R134a and R245fa should be preferred [31]. Decomposition temperature of organic fluids are greater than 200 C [18]. Decomposition temperature of R245fa exceeds 250 C [32] and it can be used till a temperature of 226.9 C [33].
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From the above explanations, it can be inferred that most of the criteria for working fluids can be fulfilled by R134a and R245fa. Besides this, R134a is the most used refrigerant for vehicle airconditioning. There is not any investigation for refrigeration using exhaust waste heat in transcritical cycles. Here, especially the T-ORVC refrigeration system for wet (nearisentropic) refrigerant R134a and dry refrigerant R245fa are investigated.
2. Transcritical organic Rankine vapor compression refrigeration system T-ORVC is schematically shown in Fig. 1. The processes in this system are illustrated in Fig. 2 on a P-h diagram. Refrigerant is pumped to a pressure P11 which is higher than the critical pressure Pcr. Point 10 is arrived, when the pump compresses the refrigerant isentropically. In the oil-refrigerant heat exchanger, refrigerant is heated till the temperature T12 below the exhaust inlet gas temperature. Oil is heated in exhaust gas-oil heat exchanger and it is circulated via an oil pump. The heated high pressure refrigerant enters the turbine and leaves it at the condensation pressure P14. Point 13 is arrived when the refrigerant is expanded isentropically to the condenser pressure P4. At the same time, compressor that is coupled to the turbine, compresses the refrigerant leaving the evaporator at the evaporation pressure P1 to the condenser pressure P2 ¼ P14. Point 3 is arrived when the refrigerant is compressed to the condenser pressure isentropically. Both gases leaving the turbine and compressor are mixed at point 15 and transported to the condenser. After condensation at point 5, refrigerant is
Fig. 2. Schematical presentation of T-ORVC refrigeration system in p-h diagram.
subcooled till the temperature T6. Some part of the liquid refrigerant enters the expansion valve and afterwards enters the evaporator to get refrigeration at the evaporator temperature T7. Heat transfer at the expansion device can be neglected so that the pressure decrease process from point 6 to point 7 can be assumed isenthalpic. Remaining part of the refrigerant leaving the condenser enters the pump and the cycles are repeated.
Fig. 1. Schematical presentation of T-ORVC refrigeration system.
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In this refrigeration system, recuperator after the turbine is not foreseen, because the exit of the turbine is near the saturation curve and therefore a recuperator is not needed [34]. Heat transfer from the exhaust gas to the system is achieved using an oil pump, two heat exchangers (namely gas-oil and oilrefrigerant) and heat transfer oil as presented in Fig. 1. These two cycles, namely the transcritical power and the subcritical refrigeration cycle, are modeled in the following thermodynamic analysis. 3. Thermodynamic analysis Evaporator refrigeration temperature is primarily determined by the inlet temperature of the fluid to be cooled. Depending on the evaporator and the fluid to be cooled, refrigerant evaporation temperature Te can be assumed 10e15 C below the inlet temperature of the fluid to be cooled. Because of this, Te ¼ T7 will be prescribed. Similarly, refrigerant condensation temperature depends primarily on the cooling fluid and this condensation temperature Tc can be assumed 10e15 C above the inlet temperature of this cooling fluid. Therefore, condensation temperature Tc ¼ T5 will also be prescribed. Parallel to Te and Tc, the subcooling temperature difference in the condenser DTSC and the superheating temperature difference in the evaporator DTSH will also be prescribed. The refrigerant pressure after the pump P11 and the refrigerant temperature after the heater T12 are also given. For the analysis, the isentropic and mechanical efficiencies of the turbine, compressor and pump must be given. Normally, the isentropic efficiencies are dependent at least on pressure ratio. However, they are assumed as constant for the sake of simplicity in this work. Besides them, one needs the electromotor efficiency of the pump and for the comparison of the mechanical work with heat, one needs the electricity production efficiency he using heat energy. Comparing different works in literature [1,5,10,11,18,20,35e41], different efficiencies given in literature are listed in Table 1. According to these values, three different groups of efficiencies are given which are classified as HV (high values), MV (medium values) and LV (low values). In this work, these three cases are considered for the calculations to see the influence of these different groups of efficiencies. These groups are given in Table 2. The pump pressure P11 ¼ P12 and the temperature T12 after the oil-refrigerant heat exchanger are varied to show the influence of these parameters on the COP. As mentioned above, evaporator and condenser temperatures are given. From these values, evaporator and condenser pressures are determined. Because DTSC and DTSH are also given, one can determine compressor inlet temperature T1 and condenser outlet
Table 2 High values (HV), medium values (MV) and low values (LV) of turbine, compressor and pump efficiencies.
HV MV LV
hti
hci
hpi
htm
hcm
hpm
hpe
he
0.90 0.80 0.70
0.80 0.75 0.65
0.90 0.80 0.75
1.00 0.97 0.95
1.00 0.97 0.95
1.00 0.97 0.95
0.95 0.92 0.90
0.35 0.30 0.25
temperature T6. From these values, one can easily calculate h6, s6 and h1, s1. Because of the isentropic compression process in the pump and in the compressor, one can write (point 3 and point 10 are the compressor and pump exits by assuming isentropic compressions):
s3 ¼ s1
(1)
s10 ¼ s6
(2)
Because evaporator pressure Pe and the pump exit pressure P11 are given, the enthalpies at the points 3 and 10 can be determined using the compressor and pump isentropic efficiencies. Enthalpies at the pump and compressor exit can be obtained using the following expressions:
h11 ¼ h6 þ
h2 ¼ h1 þ
h10 h6 hpi
h3 h1 hci
(3)
(4)
The temperature T12 after the oil-refrigerant heat exchanger is given. Pump exit pressure is also known. Enthalpy h13 at the point 13 is calculated with the assumption of isentropic expansion from the heater pressure to the condenser pressure and the following expression for entropy can be written at point 13:
s13 ¼ s12
(5)
Condenser pressure is given and therefore, h13 can be obtained. The enthalpy at the exit of the turbine is determined using the isentropic efficiency of the turbine:
h14 ¼ h12 hti ðh12 h13 Þ
(6)
In the expansion valve, the process is assumed as isenthalpic. Therefore,
h7 ¼ h6
(7)
Network produced by the turbine is
_ t ¼ M_ t ðh h Þh W 12 14 tm
(8)
The work needed by the compressor is
Table 1 Efficiencies used in literature. Literature
hti
hci
hpi
htm
hcm
hpm
hpe
he
[1] [10] [11] [41] [5] [18] [38] [37] [20] [35] [36] [40] [39]
0.90 0.75 0.75 0.80 0.85 0.75 0.85 0.70 0.80 0.85 0.80 0.70 0.80
0.65 e 0.80 e 0.80 e e 0.80 e e e e e
0.90 e e 0.75 e 0.80 0.85 0.80 0.80 e e 0.80 0.75
1.00 0.95 1.00 1.00 e e 0.95 e e e e e e
1.00 0.95 1.00 e e e e e e e e e e
1.00 e e 1.00 e e 0.95 e e e e e e
e e e e e e e e e e e e e
e e 0.25 e e e e e e e e e e
. _ c ¼ M_ c ðh h Þ h W 2 1 cm
(9)
htm and hcm are the mechanical efficiencies of the turbine and the compressor, respectively. With the definitions of mass flow rate ratio w and total efficiencies of turbine and compressor; w¼
M_ c M_ t
htc ¼ hti hci htm hcm
(10)
(11)
the following expression for mass flow rate ratio is obtained using eqs. ((4) and (6) and (8)e(11):
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w ¼ htc
h12 h13 h3 h1
(12)
Heat produced because of the mechanical efficiencies of the turbine and compressors are assumed to be transferred into ambient. Enthalpy of the fluid at the inlet of the condenser can be determined as:
h15 ¼
h14 þ wh2 1þw
(13)
It is possible to calculate COP of the T-ORVC using these equations. The following equation for the COP is obtained taking only the heat supplied to the refrigerant in the oil-refrigerant heat exchanger into consideration;
COPH ¼
wðh1 h7 Þ h12 h11
(14)
whereas the following equation is obtained taking only the work given to the pump into consideration:
COPW ¼
wðh1 h7 Þhpm hpem h11 h6
(15)
Here, hpm and hpem are mechanical efficiency of the pump and the electromotor efficiency of the pump, respectively. The best way for the definition of COP in this system would be accomplished if both heat supplied to the refrigerant and the heat equivalence of the electrical energy supplied to the pump are taken into consideration. The following equation is derived for overall COP (COPT);
COPT ¼
wðh1 h7 Þ ðh11 h6 Þ hpm hpem he
þ ðh12 h11 Þ
(16)
where he is the efficiency of the electricity production using heat energy.
5
Table 3 Comparison between the results given by Little and Garimella [1] and this work for the refrigerant R245fa (hti ¼ 0.90, hci ¼ 0.65, hpi ¼ 0.90, htm ¼ 1, hcm ¼ 1, hpm ¼ 1, hpe ¼ 1, he ¼ 1). P12 [kPa]
T12 [ C]
1519 419 2287
108.3 56.5 128.9
P5
142 333
T5 [ C]
P8
T8 [ C]
DTSC ¼ DTSH [ C]
COP [1].
COPH this work
Error [%]
47.3 23.6 48.8
66 63 53
4.7 3.8 0.0
3 1 3
0.413 0.676 0.392
0.426 0.695 0.414
3.0 2.7 5.3
Refrigerant-exhaust gas heat exchanger system consists of exhaust gaseoil (go) and oilerefrigerant (or) heat exchangers. In this system, a heat resistant heat transfer oil with an oil pump must be used. This exhaust gas heat exchanger system is also proposed by Yu et al. [17]. All the seven heat exchangers can be calculated assuming counter flow heat exchangers using the general heat exchanger equations:
Q_ ¼ ðUFÞDTm
(17)
_ p DT Q_ ¼ Mc ie
(18)
_ Q_ ¼ MDh ie
(19)
Q_ , (UF) and DTm are heat transferred, overall heat transfer coefficient-heat transferring surface area and logarithmic mean _ cp and DTie are mass flow rate, specific temperature difference.M, heat and inlet-exit temperature difference and Dhie inlet-exit enthalpy difference of the fluids. In this work, the data of a 6-cylinder, 4-stroke turbocharged diesel engine are used. The data used are given by Shu et al. [40]. Here, the components of the exhaust gas are also supplied for the engine loads between 50% and 100%, the mean values of the components of the exhaust gas are as follows:
CO2 ¼ 0:140; H2 O ¼ 0:055; O2 ¼ 0:0715; N2 ¼ 0:7335 (20)
4. Heat exchangers' analysis In the system analyzed, there are evaporator, condenser and exhaust gas heating system for heat transfer. All heat exchangers are assumed as counter flow heat exchangers. Condenser is cooled by outside air and in the evaporator, fresh outside air and indoor air mixture is cooled. Heat exchangers' temperature distributions are shown in Fig. 3. Evaporator can be calculated as heat exchangers eva 1 and eva 2. Similarly, condenser can be divided into subcooling, condensing and superheating parts which are designated as con 1, con 2 and con 3, respectively.
Using these data in eq. (20), cpg for different temperatures is calculated and the following equation for the specific heat of the exhaust gas is derived:
Tgm ¼ 30 C 530 C
:
cpg ¼ 1:043 þ 0:023 Tgm 100 (21)
cp values for different gases are used from the literature [42,43]. Here, cpg and Tgm are exhaust gas mean specific heat and mean temperature, respectively. A similar equation is given by Yang et al. [44].
Fig. 3. Temperature distributions in heat exchangers. a) Evaporator, b) Condenser, c) Exhaust gas heat exchanger system.
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Fig. 4. Variation of COPH with temperature for different pressures P12 and medium efficiencies for R134a.
For air, the following equation is derived;
Tam ¼ 0 C 100 C
:
cpa ¼ 1:006 þ 0:006ðTam =100Þ1:5 (22)
where cpa and Tam are mean specific heat and mean temperature of air. The data given in Ref. [43] are used for cpa. In this work, “Dowtherm G” is foreseen for heat transfer oil and its specific heat cpo is described as a function of its temperature as follows:
Tom ¼ 0 C 360 C
:
cpo ¼ 1:530 þ 0:233ðTom =100Þ
(23)
For the derivation of the above equation, cpo values given in Ref. [43] are used. For real fluids such as refrigerant and humid air, the following formula for the calculation of the specific heat is used:
Fig. 6. Variation of COPW with temperature for different pressures P12 and medium efficiencies for R134a.
subcritical systems if T12 is assumed greater than the saturation temperature for given values of pump exit pressure. However, only in the work given by Little and Garimella [1] are given all the parameters necessary for a comparison. They used R245fa as the refrigerant and used the same COP definition as the COPH definition by eq. (14) in this work. The same parameters used by Little and Garimella [1] are used for the calculations in this work for comparison. The parameters used and the results are presented in Table 3. As seen from this table, the difference between the COP given by Little and Garimella and the COPH determined in this work are between 2.7 and 5.3%. These differences can be considered as acceptable and it is expected that the calculations for T-ORVC should also produce acceptable results.
6. Results and discussions for the refrigerant cycle
There are not any investigation to compare with the system described in this work. But, there are some investigations for subcritical systems. The system described here is also valid for
For summer air-conditioning in intercity buses, indoor temperature can be assumed as 25 C [45,46]. Because of the dehumidifying duty, it would be appropriate to select refrigerant evaporation temperature of Te ¼ 10 C. Outdoor design conditions can be assumed as 35 C for vehicle design [11]. The highest design condition is 42 C [46]. Therefore, adequate refrigerant condensation temperature is selected as Tc ¼ 50 C. Both superheating and subcooling are assumed as DTSC ¼ DTSH ¼ 3 C, which can be considered as appropriate. In Figs. 4e6, COPH, COPT and COPW for the refrigerant R134a are demonstrated as a function of temperature using MV efficiencies. Pump exit pressure P12 is given as the parameter. Both subcritical and supercritical pressures are considered. There is an optimum
Fig. 5. Variation of COPT with temperature for different pressures P12 and medium efficiencies for R134a.
Fig. 7. Variation of COPH with temperature for different pressures P12 and medium efficiencies for R245fa.
cp ¼
hi he Ti Te
(24)
where hi, he, Ti and Te are inlet, exit enthalpies and inlet and exit temperatures of the fluids. In the above equations, the temperatures are in C and specific heats are in kJ/kg C. 5. Validation of the calculations
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Fig. 8. Variation of COPT with temperature for different pressures P12 and medium efficiencies for R245fa.
Fig. 10. Variation of maximum COPH and COPT values with pressure P12 for R134a and R245fa for medium efficiencies.
value for COPH and COPT at a certain temperature T12 for a given pump exit pressure. Beyond 6 MPa pressure, there is not any optimum point. COPH and COPT values increase very slightly after 180 C, so that optimum values can be approximately assumed as the values at 180 C. Optimum COPT values decrease when the pressure is greater than 8 MPa. The absolute highest COPT determined is 0.54 at 8 MPa. COPW values increase with temperature, but inverse to COPH and COPT, COPW decreases with increasing pressure, because the work needed for the pump increases with increasing pressure. COPW values are so high that mechanical energy needed for the pump is very low as compared to the heat energy supplied to the refrigerant. Results obtained for the refrigerant R245fa are depicted in Figs. 7e9. The results for R245fa are similar to the results obtained for R134a. For subcritical pump exit pressures, there are not any optimum values. COPH and COPT values slightly decrease with temperature. Beyond the critical pressure, COPH and COPT values slightly increase with temperature. Optimum values are determined for pressures between 4 and 6 MPa. COP values in these figures are greater than that are given in Figs. 4e6. Highest COPT value obtained using R245fa is 0.66 at 8 MPa pressure and 200 C temperature. Maximum values of COPH and COPT are illustrated in Fig. 10 dependent on pressure for both R134a and R245fa. Highest value of COPT ¼ 0.54 is obtained around 8 MPa for R134a and highest value of COPT till 200 C is obtained around 8 MPa for R245fa. There is no maximum COPT value for R245fa, but as the highest value of COPT ¼ 0.66 can be assumed as optimum. Because condensation temperature is assumed as 50 C, COPH and COPT values are zero at the saturation pressures of this temperature. The
difference between COPH and COPT values increases with increasing pressure. In Figs. 11 and 12, optimum temperatures at a given pressure for the optimum values of COPH and COPT are depicted. In these figures, saturation temperatures at given pressures are also drawn. Optimum temperatures are a little bit higher than saturation pressure till critical pressure. Beyond 8 MPa, there is little change in optimum temperature, which is approximately 180 C for R134a and 200 C for R245fa, respectively. COPT values for R134a and R245fa are also calculated using low (LV) and high (HV) efficiencies according to Table 2. From these calculations, one can conclude that COPT values increase with increasing efficiencies (LV, MV and HV) as expected. But even for LV efficiencies, the results can be considered as satisfactory. The results for HV (high values) and LV (low values) are approximately 30% higher and 30% lower than the results obtained for MV (medium values) for both refrigerants R134a and R245fa. Highest COPT values are 0.38, 0.54 and 0.71 for R134a using LV, MV and HV efficiencies, respectively. Highest COPT values are 0.47, 0.66 and 0.86 for R245fa using LV, MV and HV efficiencies, respectively. Optimum values for COPT can be assumed approximately 8 MPa pressure for all types of efficiencies. Optimum temperatures for LV and HV efficiencies are nearly the same as the values demonstrated in Figs. 11 and 12 for MV efficiencies. Exhaust temperatures of diesel engines of vehicles are around 400e700 C and waste thermal energy ranges from 20 kW to 400 kW [16,20,22]. It is clear that even with LV efficiencies, T-ORVC refrigeration system is capable of air-conditioning of intercity buses using exhaust heat.
Fig. 9. Variation of COPW with temperature for different pressures P12 and medium efficiencies for R245fa.
Fig. 11. Variation of optimum temperatures with pressure P12 for R134a for medium efficiencies.
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A. Yilmaz / Energy xxx (2015) 1e10 Table 6 Engine data for different engine loads [40]. Engine load [%] 50 75 90 100 Power output [kW] 117.7 176.2 211.6 235.8 Exhaust gas temperature [ C] 420 474 498 519 Exhaust gas mass flow rate [kg/s] 0.1697 0.2235 0.2586 0.2752
Table 7 The properties of the condenser for R134a (M_ a ¼ 8:038 kg=s, M_ r;con ¼ 0:4220 kg=s, Q_ con ¼ 80:98 kW).
Fig. 12. Variation of optimum temperatures with pressure P12 for R245fa for medium efficiencies.
Ta,i [ C]
Ta,e [ C]
Tr,i [ C]
Tr,e [ C]
DTm [ C] UF [W/ C]
1.98 64.06 14.94
35 35.24 43.15
35.24 43.15 45.00
50.0 50.0 81.0
47.0 50.0 50.0
13.35 10.30 17.57
148.5 6217.0 850.3
Table 8 The properties of the condenser for R245fa (M_ a ¼ 7:114 kg=s, M_ r;con ¼ 0:3657 kg=s, Q_ con ¼ 71:68 kW).
7. Results and discussions for heat exchangers Seven heat exchangers' properties are calculated for the optimum conditions P12 ¼ 80 bar and T12 ¼ 180 C for R134a and P12 ¼ 80 bar and T12 ¼ 200 C for R245fa. Evaporator load is assumed as Q_ eva ¼ 30 kW [45]. Under these assumptions, the temperatures, the enthalpies and other values are presented in Tables 4 and 5. In addition, engine exhaust gas temperature and exhaust gas mass flow rate changes by different engine loads are given in Table 6. Evaporator and condenser dimensions are independent of the engine load, because Q_ eva and the temperatures at the condenser and at the evaporator are prescribed. However, exhaust heat exchanger system changes with the engine load. Results for condenser are given in Tables 7 and 8 for R134a and R245fa, respectively. Outside air temperature increases from 35 C to 45 C as foreseen. It is seen from these tables that there are not much difference between the results for the two refrigerants. For the determination of the evaporator, indoor and outdoor air temperatures and humidities must be prescribed. They are taken as [45].
Con 1 Con 2 Con 3
Q_ con ½kW
Tin ¼ 25 C; 4in ¼ 0:50; Tout ¼ 35 C; 4out ¼ 0:40
(25)
Besides them, one needs fresh air mass flow rate, latent heat production in the bus and air exit temperature from the evaporator. They are assumed as [45]. V_ fresh ¼ 190l=s; Q_ l ¼ 3 kW; Ta;e ¼ 15 C
(26)
With these data and Q_ eva ¼ 30 kW, one can obtain the following equation using psychometric equations, mass balances for dry air and water vapor and the first law of thermodynamics:
Con 1 Con 2 Con 3
Q_ con ½kW
Ta,i [ C]
Ta,e [ C]
Tr,i [ C]
Tr,e [ C]
DTm [ C] UF [W/ C]
1.50 63.89 6.29
35.0 35.21 44.12
35.21 44.12 45.00
50.0 50.0 66.54
47.0 50.0 50.0
13.35 9.66 12.06
112.3 6616.0 521.5
M_ a ¼ 2:197 kg=s; Ta;i ¼ 25:98 C; ha;i ¼ 52:39 kJ=kg;
ha;e ¼ 38:74 kJ=kg
(27)
Here M_ a is dry air mass flow rate in the evaporator. Ta,i is air inlet temperature in the evaporator after mixing of the fresh air and indoor air. The properties of the evaporators eva 1 and eva 2 are given in Tables 9 and 10. It is seen that superheating part (eva 1) is very small. The evaporator is smaller than the condenser as expected. There are not important differences between the results for the refrigerants R134a and R245fa. For the determination of exhaust gas-oil (go) and oil-refrigerant (or) heat exchangers at different engine loads, it is assumed that oil temperatures at the inlet and exit of the heat exchangers are arithmetic mean temperatures of exhaust gas and refrigerant. Under this assumption, the properties of the exhaust gas-oil (go) and oil-refrigerant (or) heat exchangers are presented in Tables 11 and 12 for the refrigerants R134a and R245fa, respectively. From the tables, one can conclude that, there is no problem for air-conditioning even at 50% engine load because exhaust gas exit temperatures are high enough for both refrigerants. The UF values, even at 50% engine load, are very low as compared with the values of evaporator and especially condenser because of high logarithmic mean temperature differences.
Table 4 Data for the calculation of heat exchangers at different points of the cycles for the refrigerant R134a (P12 ¼ 8 MPa, T12 ¼ 180 C, Q_ eva ¼ 30 kW, Q_ con ¼ 80:98 kW, Q_ or ¼ 49:88 kW, M_ r;eva ¼ 0:2140 kg=s, M_ r;con ¼ 0:4220 kg=s, M_ r;or ¼ 0:2080 kg=s). Points in the cycles Temperature [ C] Enthalpy [kJ/kg]
1 13.0 259.0
4 50.0 275.3
5 50.0 123.5
6 47.0 118.8
7 10.0 118.8
8 10.0 256.2
11 52.75 126.2
12 180.0 366.1
15 81.0 310.7
Table 5 Data for the calculation of heat exchangers at different points of the cycles for the refrigerant R245fa (P12 ¼ 8 MPa, T12 ¼ 200 C, Q_ eva ¼ 30 kW, Q_ con ¼ 71:68 kW, Q_ or ¼ 40:83 kW, M_ r;eva ¼ 0:1970 kg=s, M_ r;con ¼ 0:3657 kg=s, M_ r;or ¼ 0:1687 kg=s). Points in the cycles Temperature [ C] Enthalpy [kJ/kg]
1 13.0 414.5
4 50.0 441.0
5 50.0 266.3
6 47.0 262.2
7 10.0 262.2
8 10.0 411.8
11 51.17 269.6
12 200.0 511.6
15 66.54 458.2
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Table 9 The properties of the evaporator for R134a (M_ a ¼ 2:197 kg=s, M_ r;eva ¼ 0:1970 kg=s, Q_ eva ¼ 30 kW).
Eva 1 Eva 2
Q_ eva ½kW
Tr,i [ C]
Tr,e [ C]
hr,i [kJ/kg]
hr,e [kJ/kg]
Ta,i [ C]
Ta,e [ C]
ha,i [kJ/kg]
ha,e [kJ/kg]
DTm [ C]
UF [W/ C]
0.53 29.47
10.0 10.0
13.0 10.0
411.8 262.2
414.5 411.8
25.98 25.78
25.78 15.00
52.39 52.14
52.14 38.74
14.34 9.59
37.10 3072.0
Table 10 The properties of the evaporator for R245fa (M_ a ¼ 2:197 kg=s, M_ r;eva ¼ 0:2140 kg=s, Q_ eva ¼ 30 kW).
Eva 1 Eva 2
Q_ eva ½kW
Tr,i [ C]
Tr,e [ C]
hr,i [kJ/kg]
hr,e [kJ/kg]
Ta,i [ C]
Ta,e [ C]
ha,i [kJ/kg]
ha,e [kJ/kg]
DTm [ C]
UF [W/ C]
0.60 29.40
10.0 10.0
13.0 13.0
256.2 118.8
259.0 256.2
25.98 25.76
25.76 25.00
52.39 52.12
52.12 38.74
14.33 9.37
41.83 3139.0
Table 11 The properties of the exhaust gas-oil (go) heat exchanger and oil-refrigerant (or) heat exchanger for R134a (M_ r;or ¼ 0:2079 kg=s, Q_ ¼ 49:88 kW, Tr,i ¼ 52.75 C, Tr,e ¼ 180.0 C, DTm,go ¼ DTm,or, (UF)go ¼ (UF)or). Engine load [%]
M_ g ½kg=s
M_ o ½kg=s
Tg,i [ C]
Tg,e [ C]
To,i [ C]
To,e [ C]
DTm [ C]
UF [W/ C]
100 90 75 50
0.2752 0.2586 0.2235 0.1697
0.1603 0.1566 0.1460 0.1271
519.0 498.0 474.0 420.0
360.6 328.5 276.4 155.0
349.5 339.0 327.0 300.0
206.7 190.6 164.6 103.9
161.6 148.2 128.6 103.9
308.7 336.6 387.8 617.8
From the results of the heat exchangers, it is seen that 30 kW cooling load for an intercity bus can be produced by optimum sizing of the cycle and appropriate choice of UF values of all heat exchangers by engine loads down to 50%. 8. Conclusions Transcritical ORVC refrigeration system is investigated for bus air-conditioning using exhaust heat as primary energy source. The following conclusions can be drawn out of this research: COPH and COPT values increase till the supercritical pressure of 8 MPa for both R134a and R245fa. COPH and COPT values increase till 180 C and 200 C for R134a and R245fa, respectively. The higher the efficiencies of the pump, the turbine and the compressor, the higher are the COPH and the COPT values COPW values are generally very high; therefore, electrical energy needed for the pump is very low compared to the heat energy needed. At optimum conditions, there is not any condensation in the turbine; therefore, no danger of turbine blade damages also. Optimum values for COPT for R134a are obtained at 180 C and 8 MPa. They are 0.38, 0.54 and 0.71 for LV, MV and HV efficiencies, respectively. Optimum values for COPT for R245fa are obtained at 200 C and 8 MPa. They are 0.47, 0.66 and 0.86 for LV, MV and HV efficiencies, respectively.
Table 12 The properties of the exhaust gas-oil (go) heat exchanger and oil-refrigerant (or) heat exchanger for R245fa (M_ r;or ¼ 0:1687 kg=s, Q_ ¼ 40:83 kW, Tr,i ¼ 51.17 C, Tr,e ¼ 200.0 C, DTm,go ¼ DTm,or, (UF)go ¼ (UF)or). Engine M_ g ½kg=s M_ o ½kg=s Tg,i load [%] [ C]
Tg,e [ C]
To,i [ C]
To,e [ C]
100 90 75 50
389.7 359.7 312.8 204.2
359.5 349.0 337.0 310.0
220.4 164.3 28.4 205.4 151.6 269.3 182.0 133.9 304.9 127.7 92.25 442.6
0.2752 0.2586 0.2235 0.1697
0.1331 0.1307 0.1234 0.1098
519.0 498.0 474.0 420.0
DT m [ C]
UF [W/ C]
Optimum COPT values for LV and HV efficiencies are approximately 30% lower and higher than the optimum values for MV efficiencies for both refrigerants R134a and R245fa. Optimum values for COPT of R245fa are approximately 23% higher than the optimum COPT values for R134a. Assuming 30 kW of cooling demand for an intercity bus, airconditioning is possible even at 50% engine load using both refrigerants R134a and R245fa. The UF values of exhaust gas-oil and oil-refrigerant heat exchangers are much lower than the values of evaporators and condensers. References [1] Little AB, Garimella S. Comparative assessment of alternative cycles for waste heat recovery and upgrade. Energy 2011;36:4492e504. [2] Pridasawas W. Solar-driven refrigeration systems with focus on the ejector cycle. Doctoral Thesis. Stockholm: Royal Institute of Technology; 2006. http:// www.diva-portal.org/. [3] Kim DS, Infante Ferreira CA. Solar refrigeration options-a state of the art review. Int J Refrig 2008;31:3e15. [4] Abdulateef JM, Sopian K, Alghoul MA, Sulaiman MY. Review of solar-driven ejector refrigeration technologies. Renew Sustain Energy Rev 2009;13: 1338e49. [5] Wang J, Zhao P, Niu X, Dai Y. Parametric analysis of a new combined cooling, heating and power system with transcritical CO2 driven by solar energy. Appl Energy 2012;94:58e64. [6] Wang H, Tian J, Hou X. On the coupled system performance of transcritical CO2 heat pump and Rankine cycle. Heat Mass Transf 2013;49:1733e40. [7] Sarkar J, Bhattacharyya S, Gopal MR. Transcritical CO2 heat pump systems: exergy analysis including heat transfer and fluid flow effects. Energy Convers Manag 2005;46:2053e67. [8] Elbel S. Historical and present developments of ejector refrigeration systems with emphasis on transcritical carbon dioxide air-conditioning applications. Int J Refrig 2011;34:1545e61. [9] Deng J, Jiang P, Lu T, Lu W. Particular characteristics of transcritical CO2 refrigeration cycle with an ejector. Appl Therm Eng 2007;27:381e8. [10] Wang H, Peterson R, Harada K, Miller E, Ingram-Gable R, Fisher L, et al. Performance of a combined organic Rankine cycle and vapor compression cycle for heat activated cooling. Energy 2011;36:447e58. [11] Wang H, Peterson R, Herron T. Design study of configurations on system COP for a combined ORC (organic Rankine cycle) and VCC (vapor compression cycle). Energy 2011;36:4809e20. [12] Aphornratana S, Sriveerakul T. Analysis of a combined Rankine-vaporcompression refrigeration cycle. Energy Convers Manag 2010;51:2557e64. [13] Jeong J, Kang YT. Analysis of a refrigeration cycle driven by refrigerant steam turbine. Int J Refrig 2004;27:33e41. [14] Li H, Bu X, Wang L, Long Z, Lian Y. Hydrocarbon working fluids for a Rankine cycle powered vapor compression refrigeration system using low-grade thermal energy. Energy Build 2013;65:167e72. [15] Saidur R, Rezaei M, Muzammil WK, Hassan MH, Paria S, Hasanuzzaman M. Technologies to recover exhaust heat from internal combustion engines. Renew Sustain Energy Rev 2012;16:5649e59. [16] Wang T, Zhang Y, Peng Z, Shu G. A review of researches on thermal exhaust heat recovery with Rankine cycle. Renew Sustain Energy Rev 2011;15:2862e71. [17] Yu G, Shu G, Tian H, Wei H, Liu L. Simulation and thermodynamic analysis of a bottoming organic Rankine cycle (ORC) of diesel engine (DE). Energy 2013;51: 281e90. [18] Shu G, Liu L, Tian H, Wei H, Yu G. Parametric and working fluid analysis of a dual- loop organic Rankine cycle (DORC) used in engine waste heat recovery. Appl Energy 2014;112:1188e98.
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Please cite this article in press as: Yilmaz A, Transcritical organic Rankine vapor compression refrigeration system for intercity bus airconditioning using engine exhaust heat, Energy (2015), http://dx.doi.org/10.1016/j.energy.2015.02.004