Theoretical investigations on combined power and ejector cooling system powered by low-grade energy source

Transcription

1 Theoretical investigations on combined power and ejector cooling system powered by low-grade energy source... Xiangjie Chen *, Yuehong Su, Siddig Omer and Saffa Riffat Department of Architecture and Built Environment, University of Nottingham, Nottingham NG7 2RD, UK... Abstract A combined power and cooling system is proposed for cogeneration, which integrates the ejector cooling cycle with the Rankine cycle. Low-temperature heat source such as industrial waste heat or solar energy can be used to drive the Rankine cycle. This system will provide electricity and cooling effect simultaneously without consuming primary energy. The partially expanded vapor (from low-grade energy) will be bleed off and enter into ejector s primary nozzle, which achieves cooling effect. Simulations have been carried out to analyze the effects of various working conditions on the overall system performance, on ejector entrainment ratio and turbine power output. Five different refrigerants HFE7100, HFE7000, methanol, ethanol and water have been selected, and the above three parameters were compared, respectively. The simulation results indicated that turbine expansion ratio, heat source temperature, condenser temperature and evaporator temperature play significant roles on the turbine power output, ejector entrainment ratio and the overall thermal efficiency of the system. At a heat source temperature of 1208C, evaporator temperature of 108C and condenser temperature of 358C, methanol showed the highest thermal efficiency (0.195), followed by ethanol and water (0.173). It is recommended that the evaporator temperature and the appropriate working fluid should be selected according to the different working cooling requirements, and the turbine power output can then be determined accordingly. *Corresponding author: xiangjie.chen@nottingham. ac.uk Keywords: combined power and cooling; ejector; low-grade energy; thermodynamic simulation; turbine Received 22 September 2014; revised 11 January 2015; accepted 29 May INTRODUCTION In the era of rapid global economic development, the electricity consumption is increasingly significant. According to European Community report [1], the EU electricity imports increased by 28% from 2002 to Currently most of the electricity is generated by the combustion of primary energy (coals and fossil fuels). The rising electricity demand will result in greater consumption of primary energy and larger emission of green-house gases. To generate electricity by renewable energy source (such as solar energy) or industrial waste, heat is therefore a promising, cost-effective and sustainable option. Meanwhile, the heating and cooling consumption for buildings is at a critical stake, accounts for 40% of energy demand in Europe [2]. The energy demand for cooling increases dramatically these years. According to [3], in Europe, 6% of office, commercial and industry buildings are air-conditioned, with a total volume of 20 million cubic meters. Hence, developing more energy efficient and environmental benign cooling systems is the key point for scientific researches. As a kind of low-grade energy, solar energy is intensively studied as the driving force for thermal power plants [4 6]. The thermal efficiency is uneconomically low when the vapor temperature drops below 3708C. Low-grade energy cannot be utilized directly in thermal power [7]. Exploitation of using low-grade energy (temperature of,2508c [8]) to generate electricity therefore becomes a promising research topic. With the advantages of simplicity in construction and the possibility of harnessing low-grade energy (solar energy and industrial waste heat), ejector cycle is recognized as a promising # The Author Published by Oxford University Press. This is an Open Access article distributed under the terms of the Creative Commons Attribution License ( which permits unrestricted reuse, distribution, and reproduction in any medium, provided the original work is properly cited. doi: /ijlct/ctv015 Advance Access Publication 2 July

2 Theoretical investigations on combined power and ejector cooling system and sustainable cooling system with less energy consumptions compared with traditional vapor compression system. The working principle and various applications of ejector cooling system can be found in [9 12]. The proposed combined power and ejector cooling system can not only meet the cooling requirements in summer times but also generate power in the meantime. During winter season when cooling effect is not demanded, the great amount of low-grade energy could still be converted into electricity for end-users. If the combined system could be further designed to match with the electricity grid, the tight power supply during peak hours could be substantially alleviated. Maidmenta and Tozer [13] conducted a combined cooling heat and power system in supermarket application in UK. The energy saving and carbon reduction were calculated compared with conventional combined cooling and power generation system. Roy et al. [14] carried out analysis of a Rankine cycle with ammonia absorption refrigeration system. The thermodynamic simulations based on heat source of 1008C and a heat sink of 108C revealed that the maximum thermal efficiency was 11 12%. Godefroy et al. [15] presented a CHP-ejector trigeneration system, which utilized the heat output from the CHP unit to drive an ejector cooling system. Four different scenarios were analyzed and compared, and an optimum system efficiency of 50% with cooling capacity of 2 kw was achieved according to the simulation. A similar combined ejector refrigeration and power cogeneration system was studied by Alexis [16]. The Rankine cycle produced electrical output of 2 MW, and part of the vapor was bled off from the turbine at 7bar to drive a vapor ejector cycle. The parametric studies showed that the ratio between electrical power and heat was in the range of with turbine inlet temperature of 3608C and pressure of 20bar, condenser temperature of 508C and evaporator temperature of 108C. However, the turbine inlet temperature and pressure were relatively above the quality of low-grade energy. Khaliq [17] combined a Libr-H2O absorption system with power and ejector cooling system using R141b as refrigerant. The results of first and second law investigation showed that the proposed cogeneration cycle yielded better thermal and exergy efficiencies than the cycle without absorption system. An ammonia-water absorption cycle combined with power generation system was proposed by Martin and Goswami [18] and Sadrameli and Goswami [19]. In order to optimize the cooling capacity with respect to the work production, an effective COP was defined. The simulation results indicated that the effective COP optimum was influenced by higher expander exhausttemperatures but with increased vapor flow. Zheng et al. [20] carried out a combined power and cooling system with kalina cycle. The maximum effective COP was reported near 1.1 under the heat source temperature in the range of C. The overall thermal efficiency was achieved at 24.3%. The authors concluded that the refrigeration cycle can help to improve the thermodynamic performance of the overall system. In recent years, researches are mostly focused on the combination between absorption refrigeration system with Rankine cycle to provide cooling effect and generate electricity. This paper will conduct theoretical investigations in to a combined power and ejector cooling system. The properties of different working fluids would be of importance to the system performance, which are hardly revealed by researchers. This paper presents thermodynamic study of a combined power and ejector cooling cycle. The performance of the combined system with different working fluids (water, ethanol, methanol, HFE7000 and HFE7100) is compared and analyzed. Turbine power output and system thermal efficiency under various low-grade source temperature, turbine expansion ratio and condenser temperature are also reported. 2 THERMODYNAMIC SIMULATIONS OF COMBINED POWER AND EJECTOR COOLING SYSTEM 2.1 System description The schematic diagram of this modified system is shown in Figure 1. The saturated liquid from pump outlet (state point 1) goes into the vapor generator where the working fluid absorbs heat from the low-temperature heat source (industrial waste heat or solar energy). The liquid working fluid vaporizes and comes out from the vapor generator as high-pressure vapor (state point 2) and enters the vapor turbine, where the thermal energy is converted into mechanical energy. The partially expanded vapor (state point 3) acts as the motive (primary) fluids and enters into the ejector s convergent-divergent nozzle (as shown in Figure 2), where it expands to supersonic flow creating a lowpressure region. The partial vacuum created by the supersonic primary flow is fed by a secondary flow consisting of entrained refrigerant vapor coming from the evaporator (state point 4). The primary and secondary fluids combine in the mixing chamber of the ejector and discharge through a diffuser to the condenser at the ambient temperature (state point 5). Figure 1. Schematic diagram of the combined power and ejector cooling system. 467

3 X. Chen et al. Figure 2. Schematic representation of an ejector (Section i: primary flow inlet; Section 0: secondary flow inlet; Section x: nozzle exit plane; Section 1 2: mixing chamber; Section 3: diffuser). The remaining fully expanded vapor (state point 6) from the turbine outlet together with the combined stream from ejector are all condensed in the condenser by rejecting heat to the chilled water. The liquid working fluid leaves the condenser as saturated liquid (state point 8). A proportion of the condensate is returned to the evaporator through an expansion valve at state point 9, whereas the remainder is re-circulated to the vapor generator by the pump. For the detailed ejector working mechanism and governing equations, the reader is referred to [9, 10]. 2.2 Assumptions Following assumption are made to simplify the simulation of the combined cycle: (1) System operates in steady-state condition. (2) The turbine and ejector are at adiabatic conditions. No heat exchange with the ambient conditions is considered for the vapor generator, evaporator and condenser. (3) The potential, kinetic energies and friction losses are neglected, hence: P 4 ¼ P 9 ; P 2 ¼ P 1 ; P 5 ¼ P 6 ¼ P 7 ¼ P 8 ; (4) Isenthalpic expansion process through expansion valve: h 8 ¼ h 9 ; (5) The isentropic efficiencies of turbine and pump are assumed to be 0.8 [19]. (6) Each calculated result equates to a different machine designpoint condition in every case. This is different from a real machine, which can be assumed to operate at steady state but at part-load or off-design conditions. 2.3 Governing equations In this paper, the ejector simulation is based on a 1-D constantpressure mixing model. The basic principle of the model was introduced by Keenan et al. [11], and developed by Eames et al. ð1þ ð2þ [12]. All the model assumptions and the detailed equations were described by Chen et al. [9]. As shown in Figure 2, when high-pressure primary stream enters primary (supersonic) nozzle, it expands to produce a low pressure at the exit plane. The high-velocity primary stream entrains the secondary fluid into the mixing chamber. Applying the energy equation between i-x and o-x to primary flow and secondary flow, respectively, gives the primary and secondary flow velocities at Section x-x: qffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi v px ¼ 2h p ðh p h px Þ; ð3þ p v sx ¼ ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 2ðh s h sx Þ: ð4þ Applying the energy equation, the relation between the pressure ratio across the nozzle and Mach number at the exit of the nozzle is given as: vffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi u 2h p P ðk 1Þ=k i M px ¼ t 1! ; ð5þ ðk 1Þ where h p is an isentropic efficiency of the primary nozzle considering fiction loss and k is the isentropic index of compression and expansion. Applying the momentum equation with ideal lossless mixing between primary nozzle exit plane (Section x) and the exit of mixing chamber (Section 2): P n A n þ _m p v px þ _m s v sx ¼ P 2 A 2 þðm p þ m s Þv 2 : Thus, the Mach number of the mixed fluid at Section 1 can be expressed as follows: P x ð6þ M1 ¼ h mðmpx þ pffiffi vm sx t Þ p ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi ; ð7þ ð1 þ vtþð1 þ vþ where v is ejector entrainment ratio, t is defined as the ratio of 468

4 Theoretical investigations on combined power and ejector cooling system inlet stagnation temperatures and t ¼ T 0 : ð8þ T i At some distance in the mixing chamber, a transverse shock wave is induced creating a compression effect. Across this point, the velocity of mixed stream suddenly drops to subsonic value [12]. The Mach number of the mixed fluid after the shock wave is obtained from: M 2 ¼ M2 1 þ 2=ðk 1Þ ½2k=k 1ŠM1 2 1 : Further compression of the mixed fluid is achieved as it passes through the subsonic diffuser. It is assumed that the flow speed is reduced to zero at the end of the diffuser (Section 3-3). The pressure lift ratio across the diffuser can be obtained from: ð9þ P 3 ¼ ðk 1Þh dm 2 k=k 1 2 þ 1 ; ð10þ P 2 2 where h d is the isentropic efficiency of the diffuser. The critical pressure lift ratio across the ejector is then found from Equation (11). P 3 ¼ P 3 P2 Px ¼ Ns P 0 P 2 P x P : ð11þ 0 Applying the principle of mass flow continuity at different states identified in Figure 1 gives the following relationships: _m 1 ¼ _m 2 ¼ _m 3 þ _m 6 ; _m 5 ¼ _m 3 þ _m 4 ; _m 8 ¼ _m 7 ¼ _m 5 þ _m 6 : ð12þ ð13þ ð14þ Apply the principle of energy conservation to each of the components in the system. Thus, for the heat rate input at the vapor generator: _Q g ¼ _m 2 ðh 2 h 8 Þ: Turbine expansion ratio was defined by: a ¼ P 2 P 3 : ð15þ ð16þ The high-pressure vapor in the turbine is expanded from state point 2 to state point 3, and the losses were taken into account by assuming an isentropic efficiency of h tu ¼ 0.8 [19]. The following equations could be obtained: h 3 isen ¼ enthalpy(p ¼ P 3 ; s ¼ s 2 Þ; ð17þ The same process can be applied to obtain the properties at state point 6. The turbine power output rate is calculated by: _W tu ¼ _m 2 ðh 2 h 3 Þþ _m 6 ðh 6 h 6 Þ: ð19þ The cooling effect rate obtained from the evaporator can be calculated by: _Q e ¼ _m 4 ðh 4 h 8 Þ: The energy balance at condenser is as follows: ð _m 2 _m 6 Þh 6 þð_m 4 þ _m 3 Þh 5 ¼ _m 7 h 7 : ð20þ ð21þ The heat rate rejected to the chilled water from the condenser is determined by: _Q c ¼ _m 7 ðh 7 h 8 Þ: ð22þ The remaining condensate from the condenser (state point 8) is compressed with the isentropic efficiency of 0.8 (as stated in Table 2) by the pump to the operating pressure of the boiler. The pump mechanical work rate is calculated by: _W pum ¼ _m 2 ðh 1 h 8 Þ: Then, the enthalpy at the state point 1 is calculated by: The net work output is as follows: ð23þ h 1 ¼ h 8 þ _ W pum _m 2 : ð24þ _W net ¼ _W tu _W pum : ð25þ All the ejector working mechanism and governing equations can be found in [9, 10]. The energy balance over the ejector is as follows: _m 3 h 3 þ _m 4 h 4 ¼ _m 5 h 5 : ð26þ Ejector entrainment ratio is defined as the mass flow ratio of the secondary flow to the primary flow as shown in the following equation. v ¼ _m 4 _m 3 : ð27þ Entrainment ratio is a crucial parameter closely related to the ejector performance. While the coefficient of performance for ejector is considered as the ratio between the evaporator cooling capacity and the heat input to the generator, which can be calculated as in the following equation: h tu ¼ h 2 h 3 h 2 h 3 isen : ð18þ COP ej ¼ _ Q e _Q g ¼ v h 4 h 8 h 2 h 8 : ð28þ 469

5 X. Chen et al. Based on the first law of thermodynamics, the thermal efficiency of the system, defined as the ratio of the useful energy output to the total energy input, can be given by: h ¼ _ W net þ _Q e _Q g ; ð29þ where W net is the net mechanical work rate defined in Equation (25), _Q e is the cooling rate output and _Q g is the total heat rate absorbed from the low-temperature heat source in the vapor generator. 2.4 Working fluids selection The choice of the appropriate working fluid plays an indispensable role in the design of combined power and ejector cooling system. The ozone crisis has caused a major stir in the cooling industry and has triggered a critical look at the refrigerants in use. Commercial widely used working fluids, such as CFCs, HCFCs and HFCs, allow more ultraviolet radiation into the earth s atmosphere by destroying the protective ozone layer and thus contributing to the greenhouse effect and global warming. Hence, many working fluids such as R11, R12, R113 or R114, which were suggested in previous work for ejector, are now forbidden, due to their environmental effects. Hence, the working fluids selected for our proposed system are based on the following criteria: high working performance for both organic Rankine cycle and ejector cooling cycle, high latent heat, low critical pressure, cost-effective and environmental friendly. As described by Huang [21], the refrigerants can be categorized into two types: dry fluid and wet fluid. Dry fluid has a positive slope of saturated vapor line, whereas wet fluid has a negative slope. Due to the effect that the liquid droplets carryover for wet fluids can be detrimental to the expansion process for power generation and ejector cycle, dry fluid is more favorable and selected for this combined power and ejector cooling system. Regarding the environmental aspects and the working performance of ejectors, researchers [22, 23] carried out simulations with a great number of working fluids (including R11, R113, R12, R500, R141b, R134a, R290, R717, etc.), and refrigerants such as methanol, ethanol, HFE7000 and HFE7100 are hardly studied. Based on the above-mentioned consideration, also considering the thermophysical properties of these fluids (as shown in Table 1), three natural fluids (water, methanol and ethanol) and two synergic fluids (HFE 7100 and HFE7000) were selected for the proposed cycle, as natural fluids, water, methanol and ethanol, have very low GWP (,3) and methanol is cheap and easily available. However, the volumetric refrigerating capacities of methanol, ethanol and water are lower than HFE fluids. Hence, the ejector cooling performances of water, methanol and ethanol are not competitive than that of HFE fluids. It would be interesting to investigate the working performances for organic Rankine cycle and ejector cooling cycle based on these different fluids and select the best candidate for various applications. 2.5 Simulation procedure Based on the governing equations in Section 2.3, computer simulation was carried out in EES software, from whose database the thermodynamic properties of selected working fluids could be obtained. The input parameters for the simulation are summarized in Table 2. The cooling capacity of the ejector was chosen to match the capacity (0.5 kw) of a small experimental test rig built at the University of Nottingham [13]. Future work will be carried out to compare experimental results with the ones reported in this paper. The future works will be to modify the rig and carry out experimental testing in the existing rig and compare the results with these reported in this paper. The corresponding temperature, pressure, enthalpy and entropy at all the state points are identified in Table 3. Five different working fluids, methanol, ethanol, water, HFE7000 and HFE7100, were selected and compared in the simulation. 2.6 Simulation results and discussions The thermodynamic properties and mass flow rates at each of the state points for water as the working fluid are indicated in Table 3 and the T-S diagram is shown in Figure 3. It should be pointed out that the values of mass flow rates and thermodynamic properties have been rounded off. Hence, the deviations between the two sides of Equations (1 17) are 0.5% of their average value in some cases Effect of heat source temperature The heat source temperature T 2 is varied between 110 and 1708C, whereas other input parameters are at fixed value as Table 1. Thermophysical properties of the selected refrigerants and their environmental impact. H 2 O Methanol Ethanol HFE7000 HFE7100 Molecular weight (g/mol) Melting point (8C) Boiling point (8C) Critical point temperature (8C) Critical Point Pressure (bar) Explosive limits No 6 36% % No No Flammable No Highly yes No No Toxic No Highly If consumed in large amounts No No GWP

6 Theoretical investigations on combined power and ejector cooling system Table 2. Simulation input parameters. Turbine expansion ratio a ¼ P 2 /P 3 ¼ 2.5 [19] Turbine extraction ratio l ¼ _m 3 = _m 2 ¼ 0:3[19] Turbine/pump isentropic efficiency h tu ¼ h pum ¼ 0.8 Evaporator temperature T 4 ¼ 108C Condenser temperature T 8 ¼ 358C Evaporator cooling capacity 0.5 kw Table 3. Simulation results with water as the working fluid. State point Pressure (kpa) Specific entropy (kj/kg K) Specific enthalpy (kj/kg) Temperature (8C) Mass flow rate (g/s) Figure 3. Variations of entrainment ratio with respect to heat source shown in Table 1. According to Equation (16), the turbine inlet pressure increases as heat source temperature increases. Since the evaporator and the condenser temperature is fixed at 10 and 358C, respectively, the thermodynamic conditions at state point 1, 8 and 9 remain constant. Hence, the mechanical work consumed by the pump stays constant, whereas the turbine power output increases (according to Equation (19)). Consequentially, the net work increases with the rise of heat source temperature. As the evaporator cooling output remains constant, the mass flow rate through the evaporator m 4 remains constant (as h 4 and h 9 do not change). Therefore, as shown in Figure 4, the ejector entrainment ratio v increases as heat source temperature changes from 110 to 1708C. For instance, at heat source temperature of 1408C, the entrainment ratio of the ejector with Figure 4. Variations of thermal efficiency with respect to heat source HEF7000 as working fluid is 0.47, 0.41 for HFE 7100, 0.31 for methanol, 0.28 for ethanol and 0.20 for water. Due to the better thermal physical properties of synergetic fluids, HFE 7000 and HEF7100 show higher entrainment ratio than methanol, ethanol and water, at the same heat source temperature. In other words, with the same heat source temperature, the COP and cooling capacity of ejector can be improved with HFE7000 and HFE7100 as refrigerants. Figure 4 shows the variations of thermal efficiency with respect to heat source temperature for different working fluids. For instance, when heat source temperature increases from 140 to 1508C, the percentage increase of thermal performance for methanol is 8.7%, whereas that for ethanol and water is 9.6%, HFE7000 is 5.6% and HFE7100 is 7.3%. In general, the thermal efficiency increases with the heat source temperature from 110 to 1708C. As discussed above, together with Equation (25), it could be noted that as the heat source temperature increases, the increased net work is proportional larger than the increased vapor generator energy input; therefore, the overall thermal efficiency is improved. It is evident that the increase rate of thermal efficiency is highest for methanol, followed by water and ethanol, with HFE7000 and HFE7100 as the lowest. Figure 5 shows that the turbine power output increases as the heat source temperature increases from 110 to 1708C. For instance, at the heat source temperature of 1408C, the turbine power output was 802 W for methanol and HFE7000, 913 W for HFE 7100, 1001 W for ethanol and 1168 W for water as refrigerant. Among all the compared working fluids, water demonstrates the highest turbine power output, followed by ethanol, whereas methanol, HFE7000 and HFE7100 share similar power output values. It can be seen from the diagram that with heat source temperature of 1208C, the turbine mechanical work can reach up to 1.2 kw with water as working fluid Effect of turbine expansion ratio In order to analyze the effect of turbine expansion ratio on the system performance, the expansion ratio a ¼ p 2 /p 3 is varied between 1.5 and 7.5. Since the generator outlet pressure is fixed 471

7 X. Chen et al. Figure 5. Variations of turbine power output with respect to heat source Figure 7. Variations of thermal efficiency with respect to expansion ratio for five working fluids. Figure 6. Variations of entrainment ratio with respect to expansion ratio for five working fluids. Figure 8. Variations of turbine power output with respect to expansion ratio for five working fluids. at 0.6 MPa, the pressure at state point 3 decreases from 0.4 to 0.08 MPa accordingly, which reduces the pressure and temperature of the primary flow entering the ejector. As the evaporator cooling capacity is fixed at constant value, the secondary flow remains constant. Thus, the increased expansion ratio will lead to the decrease of entrainment ratio, as shown in Figure 6. Figure 7 shows that the thermal efficiency increases as the turbine expansion ratio increases. The reason is that as expansion ratio increases, turbine outlet pressure and temperature decreases, resulting in decreased entrainment ratio in the ejector. Since the evaporator cooling capacity is constant and the temperature at state point 4 and 9 remain constant, the mass flow rate m 4 across the evaporator remains constant. Thus, the reduced entrainment ratio will be accompanied by the increase in the mass flow rate m 3 through the turbine. According to Equation (19), the turbine mechanical work will increase in the meantime. As shown in Figure 7, as the heat source temperature remains at 1208C, the heat input to the generator and the work consumed by pump remain constant. Therefore, the thermal efficiency (according to Equation (25)) is improved as the turbine expansion ratio increases. At the same expansion ratio of 4, methanol demonstrates the highest thermal efficiency of 0.22, followed by water and ethanol, with HFE7000 and HFE7100 at last. Similar trend could be found in the variations of turbine power output with respect to expansion ratio. As shown in Figure 8, water demonstrates highest turbine power output, followed by ethanol and methanol, with HFE7100 and HFE7000 as the lowest. The reason is that the increased expansion ratio could be translated into larger pressure difference across turbine inlet and outlet, which virtually produces higher turbine mechanical work. When we consider the effect of increasing heat source temperature on both ejector entrainment ratio (Figure 6) and turbine power output (Figure 8), it can be found that while the turbine expansion ratio decreases, the ejector entrainment ratio increases as a result of the higher primary flow inlet pressure. In this case, ejector works at higher COP, but the turbine power output decreases. In real application, the turbine and ejector have to be carefully designed in order to meet the climatic conditions and customer requirements. 472

8 Theoretical investigations on combined power and ejector cooling system Figure 9. Variations of turbine power output with respect to condenser Figure 11. Variations of entrainment ratio with respect to condenser Figure 10. Variations of thermal efficiency with respect to condenser Effect of condenser temperature In order to analyze the effect of condenser temperature on system performance, condenser temperature is varied from 30 to 408C with heat source temperature of 1208C, evaporator temperature of 108C and the other input parameters at fixed value as shown in Table 1. The variations of turbine power output with respect to the condenser temperature are shown in Figure 9. It can be seen that the turbine power output decreases with the increase of condenser temperature. The reason is that the turbine power output is influenced by the turbine back pressure, which is affected by the increase of condenser temperature. When condenser temperature is in the range of 30 to 358C, the turbine power output is higher for ethanol and methanol, followed by water, with HFE7100 and HFE7000 at last. When the condenser temperature is beyond 358C, the turbine power output values for various working fluids are much close to each other than those when the condenser temperature is,358c. For instance, if the condenser temperature is at 358C, the turbine power output for water as working fluid is 1112 W, whereas that for ethanol is 1007 W, HFE7000 is 978 W, methanol is 872 W and HFE7100 is 825 W. Due to the global warming effect in the world, the summer peak outdoor temperature in southern Europe can be in the range of C. If this combined power and ejector system is applied in these areas, the simulation results summarized here can be regarded as an important guidance for the system refrigerants selection. Figure 10 indicates that the thermal efficiency decreases with the increasing of condenser temperature. The reason is that increasing the condenser temperature will lead to the increase of ejector back pressure. This will result in increased compression ratio (ratio of condenser pressure to evaporator). Thus, less secondary flow will be entrained from the ejector, leading to reduced entrainment ratio (as illustrated in Figure 11). Since the evaporator cooling capacity is defined as constant, according to Equation (20), the secondary mass flow rate is always constant. Thus, the decreasing of entrainment ratio will result in the increase of primary mass flow rate, which finally leads to the increase of the heat input in the vapor generator Equation (15). Hence, the thermal efficiency reduces as more heat input is required in order to maintain the same cooling capacity when condenser temperature increases. The variations of thermal efficiency with respect to different working fluids could be found in Figure 10. Methanol demonstrates the highest thermal efficiency, followed by ethanol and water, with HFE7000 and HFE 7100 at the last. It should be noted that if the condenser temperature is further increased beyond the critical condenser pressure, the thermodynamic shock wave of the mixed primary and secondary flow will be moved backward into the mixing chamber. This will prevent secondary flow from reaching sonic velocity, and the ejector will lose its cooling function [24]. As the condenser temperature is closely related to the summer outdoor temperature, the combined ejector and cooling system will have to be designed specifically according to the local climatic conditions with relatively large condenser temperature range Effect of evaporator temperature Figures 12 and 13 show the effect of evaporator temperature on the system thermal efficiency and turbine power output for the five different working fluids. It was found that the thermal 473

9 X. Chen et al. Figure 12. Variations of thermal efficiency with respect to evaporator Figure 14. Variations of entrainment ratio with respect to evaporator lead to the decrease of turbine power output. For instance, at lower evaporator temperatures (5 88C), water and ethanol can produce higher turbine power output compared with methanol, HFE7000 and HFE7100. At higher evaporator temperate (9 158C), the turbine power outputs produced by these five refrigerants are very limited and tend to approach each other. 3 CONCLUSIONS Figure 13. Variations of turbine power output with respect to evaporator efficiency increases with the rise of evaporator temperature, whereas the turbine power output decreases with the increase of evaporator temperature. The reason is that, as the evaporator cooling capacity is fixed at 0.5 kw, the mass flow rate m 4 across the evaporator is constant. According to Equation (27), the increasing of evaporator temperature will lead to the improvement of entrainment ratio (as shown in Figure 14). This will accompany by the decreasing of the primary mass flow rate m 3 entering the ejector. Hence, the turbine power output and the vapor generator heat input will decrease accordingly, which results in the increase of the overall thermal efficiency. However, this will have to sacrifice with the desired cooling temperature. Normally, the cooling temperature should be fixed below 108C for retailing food conservation and in the range of C to satisfy with people s thermal comfort in summer time. Hence, the evaporator temperature and cooling capacity should be designed according to customers requirements and the turbine power output can then be determined accordingly. Figure 12 also demonstrates that the methanol shows the highest thermal efficiency, followed by ethanol and water, with HFE7000 and HFE 7100 at last. The overall thermal efficiency is 0.16 at evaporator temperature of 88C for water as refrigerant. Figure 13 shows that increasing the evaporator temperature will A combined power and ejector cooling system is proposed, which combines ejector with the Rankine cycle, providing electricity and cooling effect simultaneously. In order to explore the possibility of utilizing the low-grade energy (temperature in the range of C) as the heat source to drive turbine and ejector cooling system, a novel configuration was presented and the computer simulation was carried out with the evaporator temperature of 108C and the condenser temperature of 358C. Five different working fluids, water, methanol, ethanol, HFE7100 and HFE7000, were compared in the simulation. The following conclusions could be drawn from the simulation results: The system performance of the five different working fluids were compared at heat source temperature of 1208C, evaporator temperature of 108C and condenser temperature of 358C. Methanol showed the highest thermal efficiency (0.195), followed by ethanol and water (0.173), and HFE fluids the lowest (0.145). Water demonstrated the highest turbine power output (1463 W), followed by ethanol (1217 W) and HFE fluids (850 W). With water as the working fluid, the thermal efficiency of the combined system improved from 0.15 to 0.25 and the turbine power output increased from 1200 to 1400 W when the heat source temperature increased from 110 to 1708C. For all the five working fluids, an average 10% thermal efficiency improvement could be achieved when the turbine expansion ratio increased from 1.5 to 7. In the meantime, 474

10 Theoretical investigations on combined power and ejector cooling system significant improvements (3 3.5 times) of turbine power output could be notified with water, methanol and ethanol as working fluids. However, ejector entrainment ratios decreased about 50% as the turbine expansion ratio raised from 1.5 to 7. It could be concluded that the increasing of turbine expansion ratio is beneficial to the turbine power output, but not helpful for the improvement of ejector system performance. When the condenser temperature increased from 30 to 408C, the thermal efficiency of the combined system reduced approximately by 40% and the turbine power output decreased about 60%. This means that the condenser temperature plays a crucial role in the overall system performance. If the ambient temperature goes beyond the design conditions (358C), the performance of the combined system drops dramatically. When the evaporator temperature increased from 5 to 158C, the thermal efficiency of the combined system improved significantly (average 60% for five working fluids). However, the turbine power output decreased about 50% as the evaporator temperature increased from 10 to 158C. In this case, the evaporator temperature has to be selected properly in order to guarantee the thermal efficiency and the turbine power output at thedesignedconditions. One could optimize the working performance of combined power and ejector cooling system based on the choice of the expansion ratio and evaporator temperature depending on the cooling capacity and power generation requirements. This could be an interesting topic for further numerical research. Other future works will be concentrated on experimental studies, and the results will be compared with the simulation outcomes presented in this paper. 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