Effectivene and Exergy Detruction Analyi of Evaporator in Organic Rankine Cycle Kyoung Hoon Kim, and Chul Ho Han Abtract---Thi paper carrie out a performance analyi baed on the firt and econd law of thermodynamic for ource heat exchanger in organic Rankine cycle when the heat ource i low-temperature energy in the form of enible heat. In the analyi, effect of the election of working fluid and turbine inlet preure are invetigated on the performance of the evaporator including the effectivene of heat tranfer, exergy detruction, pinch point, and econd law efficiency. The reult how that the heat exchanger effectivene and the exergy detruction increae but the pinch-point enthalpy increae a the turbine inlet preure increae. Keyword organic Rankine cycle, ource heat exchanger, effectivene, exergy detruction. E I. INTRODUCTION NERGY demand ha been rapidly increaing but the foil fuel to meet the demand i being drained, an efficient ue of low-grade energy ource uch a geothermal energy, exhaut ga from ga turbine ytem, bioma combution, or wate heat from variou indutrial procee become more and more important. The wate heat which i merely dicharged due to the lack of effective converion method contribute to the thermal pollution and environmental problem [1]-[2]. The organic Rankine cycle (ORC) and the power generating ytem uing binary mixture a a working fluid have attracted much attention a they are proven to be the mot feaible method to achieve high efficiency in converting the low-grade thermal energy to more ueful form of energy [3]-[7]. The ORC i commonly accepted a an etablihed technology to convert low grade heat ource into electricity. Furthermore, ORC are deigned for unmanned operation with little maintenance. Due to the excellent characteritic, everal ORC wate heat recovery plant are already in operation [8]. ORC i a Rankine cycle where an organic fluid i ued intead of water a working fluid. Since ORC ytem exhibit great flexibility, high afety and low maintenance requirement in recovering low-grade heat, it ha been widely accepted a a promiing technology for converion of low grade heat into ueful work and electricity. Drecher and Bruggemann [9] tudied the ORC in olid bioma power and heat plant and they propoed a method to find uitable thermodynamic fluid for ORC in Kyoung Hoon Kim i profeor in the Department of Mechanical Engineering, Kumoh National Intitute of Technology, Gumi, Gyeongbuk 39177, Korea Chul Ho Han i profeor in the Department of Mechanical Sytem Engineering, Kumoh National Intitute of Technology, Gumi, Gyeongbuk 39177, Korea (phone: 82-54-478-7393; fax: 82-54-478-7319; e-mail: chhan@kumoh.ac.kr), correponding author bioma plant and found that the family of alkylbenzene howed the highet efficiency. Heberle and Brueggemann [1] reported an analyi of a combined heat and power generation for geothermal re-ource with erie and parallel circuit of an ORC. Dai et al [11] identified and R236ea a efficient working fluid by uing a generic optimization algorithm. Hung et al. [12] examined Rankine cycle uing organic fluid which are categorized into three group of wet, dry and ientropic fluid. They pointed out that dry fluid have diadvantage of reduced net work due to uperheated vapor at turbine exit, and wet fluid of the moiture content at turbine inlet, o ientropic fluid are to be preferred. Kim and Perez-Blanco [13] preented a thermodynamic analyi of cogeneration of power and refrigeration baed on ORC activated by low-grade enible energy. Lee and Kim [14] reported energy and exergy analye for a combined cycle coniting of an ORC and a liquefied natural ga (LNG) Rankine cycle for the recovery of low-grade heat ource and LNG cold energy. In heat exchanger of power generation cycle for recovery of thermal energy ource in the form of enible heat energy, aement of pinch point i omewhat complex. Kim et al. [15] propoed efficient and novel method for pinch point aement in ource heat exchanger or condener in ammonia-water baed power generation cycle. They alo preented the firt and econd thermodynamic law analyi for heat recovery vapor generator (HRVG) of ammonia-water mixture when the heat ource i low- temperature energy in the form of enible heat [16]. Thi tudy preent a performance analyi of the evaporator in organic Rankine cycle baed on the firt and econd law of thermodynamic for the recovery of low-grade heat ource. Parametric tudy i carried out on the performance of the ource heat exchanger depending on the working fluid and turbine inlet preure. II. SYSTEM ANALYSIS The chematic diagram of ORC i hown in Fig. 1. The working fluid i compreed in pump from a aturated liquid tate (tate 1) to a compreed liquid tate (tate 2). The fluid i then heated in the evaporator uing the enible heat from the ource fluid to a aturated or uperheated vapor tate (tate 3). After the mechanical energy i obtained (tate 4) owing to the expanion proce in turbine, the flow enter the condener and exchange heat with the coolant, then return to tate 1. The ource fluid enter the evaporator with ma flow rate m and temperature T. On the other hand, working fluid enter the 23
2 Pump 1 1 T L T C Evaporator ORC Condener 3 T S P H Turbine Fig. 1. Schematic diagram of ORC and ource heat exchanger evaporator with ma flow rate m w temperature T 2, and turbine inlet preure P H, i heated by the hot ource tream, and then leave the heat exchanger with turbine inlet temperature T H =T 3. The important aumption in the preent cycle analyi are a follow [16]: 1) The fluid flow and heat tranfer rate are teady. 2) Heat lo between the ytem and environment i negligible o that heat tranfer occur only between hot and cold fluid tream in the heat exchanger. 3) Preure drop due to flow inide heat exchanger i negligible o that the preure inide a heat exchanger i maintained contant. 4) The kinetic energy and the potential energy change of the fluid in and out of the heat exchanger are negligible. 5) The longitudinal heat conduction in the tube wall i negligible. 6) The working fluid at the turbine inlet i a aturated vapor. For a pecified ma flow rate of the ource fluid m, the ma flow rate of the working fluid m w, the ma flow rate of coolant m c, and heat tranfer at evaporator Q can be determined from the energy balance at the evaporator and condener and the condition of pinch temperature difference a follow [12]: Q 4 G mc p T Tout mw (1) h 3 h 2 mwh4 h1 mc (2) c T T pc cout c h h m c T T m (3) w 3 2 p out where h i the pecific enthalpy of the mixture, T out i the ource exhaut temperature, T cout i the coolant outlet temperature, c p i the iobaric pecific heat, and ubcript, w, and c denote the ource, working fluid, and coolant, repectively. And the condition of pinch temperature difference can be determined a the temperature difference between hot and cold tream in a heat exchanger i a ame a the precribed value of ΔT pp [2], [15]: T hot Tcold Tpp min (4) Let u define a the dimenionle enthalpy at pinch point H pp and enthalpy ratio at pinch point x pp a follow: x H h h pp f pp (5) hg h f h h pp 2 pp (6) h3 h2 where h pp, h g, and h f repreent the enthalpie of working fluid at pinch point, aturated liquid and aturated vapor, repectively. Exergy i a meaure of the maximum capacity of a ytem to perform ueful work a it proceed to a pecified final tate in equilibrium with it urrounding. Exergy i generally not conerved a energy, but detructed in the ytem. The pecific exergy i defined a [17]: e h h (7) T where i the pecific entropy and the ubcript denote the dead tate. The pecific exergy of ource fluid at temperature T can be evaluated approximately in accordance to the following equation: T ln T T e c T T / (8) The effectivene ε, econd law efficiency η, and exergy detruction D of evaporator are defined a follow: T T (9) T w e out T 2 out m e3 e2 (1) m e e e m D m out w e 3 e 2 (11) In thi work, the thermodynamic propertie of the working fluid are calculated by the Patel Teja equation of tate [18], [19]. III. RESULTS AND DISCUSSIONS In the preent tudy, thermodynamic analyi of evaporator in ORC i carried out. Eight fluid of,,,,,, R123, and are conidered. The baic thermodynamic data for the working fluid are lited in acending order of critical temperature in Table 1 [2]. The ource fluid i conidered a water at T = 15 C with m = 1 kg/. The baic data of the ytem variable are a follow: condenation temperature T L = 4 C, coolant temperature T c = 25 C, dead tate temperature T = 298.15 K, pinch temperature difference ΔT pp = 1 C, ientropic pump efficiency η p =.85, ientropic turbine efficiency η t =. The key parameter in thi tudy i the turbine inlet preure, P H. The ma flow ratio i plotted againt the turbine inlet preure in Fig. 2 for variou working fluid. The ma flow 24
TABLE I BASIC THERMODYNAMIC DATA OF WORKING FLUIDS. Subtance M(kg/kmol) T cr(k) P cr(bar) 12.31 38. 36.9.239 66.51 386.6 44.99.263 44.96 396.82 42.49.152 58.123 48.14 36.48.177 58.123 425.18 37.97.199 134.48 427.2 36.4.372 R123 136.467 456.9 36.74.282 72.15 46.43 33.81.228 i defined a the ratio of the ma flow rate of working fluid to the ource fluid. The ma flow ratio decreae with increaing turbine inlet preure. It i becaue a the turbine inlet preure increae, the correponding evaporation temperature of working fluid increae, which lead to higher temperature of ource exhaut and conequently to lower heat tranfer in the evaporator. It can be een from the figure that there exit an upper limit of turbine inlet preure for the working fluid with high critical temperature uch a R123 and, where the heat tranfer i reduced to zero. Fig. 3 how the variation of enthalpy ratio at pinch point in evaporator. A i een in the Eq. (5), the enthalpy ratio i ame a the quality of the fluid when it i higher than zero and lower than unity. When the enthalpy ratio i lower than zero, the lower enthalpy ratio become, the greater degree of ubcooling of the working fluid become. When the enthalpy ratio i greater than unity, the higher the enthalpy ratio become, the greater the degree of uperheating of the working fluid become. A i een in the figure that for mot working fluid the pinch point occur enthalpy ratio (x pp ). -.1 -.2 -.3 -.4 -.5 -.6 -.7 Fig. 3. Variation of enthalpy ratio at pinch point with turbine inlet preure for variou working fluid. at which the enthalpy ratio i unity o when the working fluid i aturated vapor. But for working fluid with lower critical Figure temperature 1. Plot uch of ma a R152 flow rate or, of mixture the againt enthalpy ammonia ratio decreae concentration with for increaing variou temperature turbine inlet of ource preure. airvariation Thi mean of exergy that the lo pinch ratio point due to occur exhaut when with repect the working to preure fluid ratio i for ubcooled. variou water injection ratio. The dimenionle enthalpy at pinch point which wa defined in Eq. (6) i illutrated in Fig. 4 for variou working fluid. It Figure 2. can be een from the figure that the dimenionle enthalpy increae with increaing turbine inlet preure. It i becaue a the turbine inlet preure increae, the correponding evaporation become high, o the potion of heating for ubcooled liquid in the total heat tranfer in the evaporator increae. 2.5.7 ma flow ratio 2. 1.5 1..5. dimenionle enthalpy (H pp ).6.5.4.3.2.1. Fig. 2. Variation of ma flow ratio with turbine inlet preure for variou working fluid. Fig. 4. Variation of dimenionle enthalpy at pinch point with turbine inlet preure for variou working fluid. Figure 3. Plot of ma flow rate of mixture againt ammonia 25 concentration for variou temperature of ource airvariation of exergy lo ratio due to exhaut with repect to preure ratio for variou water
1 8 9 heat exchanger effectivene, % 8 7 6 5 4 3 2 1 econd law efficiency, % 7 6 5 4 3 Fig. 5. Variation of heat exchanger effectivene of evaporator with turbine inlet preure for variou working fluid. Figure The 5. effectivene Plot of ma of evaporator flow rate of i plotted mixture againt the ammonia turbine inlet concentration preure for in variou Fig. 5 for temperature variou of working ource airvariation fluid. It can of exergy be een from lo ratio the figure due to exhaut that the with effectivene repect to preure decreae ratio for with variou increaing water injection ratio. turbine inlet preure, becaue the vaporization heat of working fluid decreae with increaing turbine inlet preure. For a Figure 6. pecified turbine inlet preure, the effectivene become lower a the critical temperature of working fluid increae, ince a high critical temperature of working fluid reult in a low heat tranfer in the evaporator and conequently a low effectivene of evaporator. Fig. 6 how the exergy detruction at evaporator a a function of turbine inlet preure for variou working fluid. It i worthy to note that in thi cae the exergy detruction i equal to the product of dead tate temperature T and the entropy exergy detruction, kw 6 5 4 3 2 1 Fig. 6. Variation of exergy detruction with turbine inlet preure for variou working fluid. Fig. 7. Variation of econd law efficiency with turbine inlet preure for variou working fluid. generation in the evaporator. The exergy detruction decreae Figure 7. Plot of ma flow rate of mixture againt ammonia with increaing turbine inlet preure. Therefore, the higher concentration for variou temperature of ource airvariation of exergy turbine lo inlet ratio due preure to exhaut become, with repect the to lower preure irreveribility ratio for variou of water the evaporator injection become. ratio. It can be alo een that for a pecified value of turbine inlet preure, the exergy detruction at the evaporator Figure how 8. a decreaing tendency with repect to the critical temperature of working fluid. However, the exergy detruction of i higher than that of. Fig. 7 how the econd law efficiency of evaporator with repect to turbine inlet preure for variou working fluid. In the range of imulation, the efficiency increae with increaing turbine inlet preure for the working fluid with low critical temperature uch a,, and. However, for the working fluid with high critical temperature uch a,, R123, and, the efficiency firtly increae and reache a maximum value and then decreae again a the turbine inlet preure increae. However, it can be oberved from the figure that the maximum efficiency value of the working fluid are imilar to each other. The optimum turbine inlet preure for the econd-law efficiency become higher a the critical temperature of working fluid become lower. IV. CONCLUSIONS A performance analyi of the evaporator in organic Rankine cycle with variou working fluid i carried out baed on the firt and econd law of thermodynamic for the recovery of low grade finite heat ource. The reult how that the election of the working fluid and the turbine inlet preure exhibit ignificant effect on the thermodynamic performance of the evaporator. Reult how that the heat exchanger effectivene of evaporator or exergy detruction decreae with increaing turbine inlet preure or critical temperature of working fluid. However, the econd law efficiency increae with turbine inlet 26
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