Evaluation of Working Fluids for Organic Rankine Cycle Based on Exergy Analysis

Transcription

1 IOP Conference Series: Earth and Environmental Science PAPER OPEN ACCESS Evaluation of Working Fluids for Organic Rankine Cycle Based on Exergy Analysis To cite this article: D Setiawan et al 2018 IOP Conf. Ser.: Earth Environ. Sci View the article online for updates and enhancements. Related content - Exergy analysis of biomass organic Rankine cycle for power generation T B Nur and Sunoto - Thermodynamic analysis of organic Rankine cycle with Hydrofluoroethers as working fluids Suhas Upadhyaya and Veershetty Gumtapure - Working fluid selection for the Organic Rankine Cycle (ORC) exhaust heat recovery of an internal combustion engine power plant S Douvartzides and I Karmalis This content was downloaded from IP address on 01/10/2018 at 19:20

2 Evaluation of Working Fluids for Organic Rankine Cycle Based on Exergy Analysis D Setiawan 1, I D M Subrata 1, Y A Purwanto 1 and A H Tambunan 1 1 Department of Mechanical and Biosystem Engineering, Bogor Agricultural University, Indonesia ahtambun@yahoo.com Abstract One of the crucial aspects to determine the performance of Organic Rankine Cycle (ORC) is the selection of appropriate working fluids. This paper describes the simulative performance of several organic fluid and water as working fluid of an ORC based on exergy analysis with a heat source from waste heat recovery. The simulation was conducted by using Engineering Equation Solver (EES). The effect of several parameters and thermodynamic properties of working fluid was analyzed, and part of them was used as variables for the simulation in order to determine their sensitivity to the exergy efficiency changes. The results of this study showed that water is not appropriate to be used as working fluid at temperature lower than 130 C, because the expansion process falls in saturated area. It was also found that Benzene had the highest exergy efficiency, i.e. about 10.49%, among the dry type working fluid. The increasing turbine inlet temperature did not lead to the increase of exergy efficiency when using organic working fluids with critical temperature near heat source temperature. Meanwhile, exergy efficiency decreasing linearly with the increasing condenser inlet temperature. In addition, it was found that working fluid with high latent heat of vaporization and specific heat exert in high exergy efficiency. 1. Introduction In most of industrial processes, energy is often lost as waste heat and released to the atmosphere directly. Le et al. [1] reported that about 20-50% of energy to the industry is lost as waste heat. Furthermore, BCS Inc. [2] estimated that 60% of the industrial waste heat has temperatures below 230 o C. Palm oil plant, which is one of the most important industries in Indonesia, has several parts of it that potentially generate waste heat, such as boiler s flue gas with temperature around 290 o C. For palm oil plant with capacity of 5 tons fresh fruit bunch per hour, the flow rate of the flue gas is around 2.57 kg/s [3]. Conversion of the waste heat into electrical energy is indispensable in the midst of energy crisis and environmental problem. In practice, the most widely used technology for generating electrical energy from heat is the Rankine cycle (steam power cycle). Rankine cycle uses water as working fluid, so the ideal turbine inlet temperature has to be above 350 o C [4]. Using Rankine cycle is considered to be inefficient when using waste heat with lower temperature as heat source. Therefore, it is necessary to find working fluids that can give comparable performance at lower temperature. One of the working fluid group that may have comparable performance with water is organic fluid. Therefore, the thermodynamic cycle that used organic fluid is called Organic Rankine Cycle (ORC). Studies related to the determination of working fluid on ORC have been done by several researchers. Drescher and Bruggemann [5] reported that alkylbenzene had better thermal efficiency than Toluene Content from this work may be used under the terms of the Creative Commons Attribution 3.0 licence. Any further distribution of this work must maintain attribution to the author(s) and the title of the work, journal citation and DOI. Published under licence by Ltd 1

3 and OMTS as working fluid based on biomass combustion (temperature o C) as heat source. Liu et al. [6] found that n-pentane had the highest thermal efficiency for heat source at temperature range of o C. Lakew and Bolland [7] concluded that R227ea produced the highest power for heat source temperature range of o C, while R245fa gave higher work output for temperature greater than 160 o C. Although there have been many studies on the selection of working fluids, no single fluid has been identified as the most optimal for the ORC. It is due to the large number of working fluid variations, the difference of heat sources temperature and working conditions, and the difference in indicator performance [8]. Many of the studies evaluating the appropriate working fluids for ORC were based on thermal efficiency. However, exergy efficiency is a more rational criterion to evaluate the working fluid performance based on waste heat recovery. Therefore, the purposes of this study are: (1) to simulate the performance of several organic fluids, including water, based on exergy analysis with heat source from waste heat recovery (2) looking for the parameters that influence the exergy efficiency level, so it can be useful as guidelines for selecting working fluids. 2. Methods This research is a simulative study for investigating the properties of working fluid by using Engineering Equation Solver (EES) System description The ORC system being studied is a simple cycle, which consists of a pump, evaporator, turbine and condenser (figure 1). The working fluid absorbs heat from the heat source (waste heat recovery), evaporates in the evaporator (process 2-3). The vapor from the outlet of the evaporator will expands in the turbine for doing useful work (process 3-4). The vapor enters the turbine at saturated vapor condition (quality = 1). Then, the working fluid enters condenser to dissipate heat out (process 4-1) before returning to the pump to complete the cycle (process 1-2). The phase of working fluid is saturated liquid at the exit of the condenser. The characteristics of the heat sources (flue gas) and cooling water are preset by fixing the inlet temperature, inlet pressure, mass flow rate. Performance of the heat transfer process is determined by fixing the pinch point temperature. Pinch point temperature is the minimum temperature difference between the saturation temperature of the working fluid and the temperature of the flue gas or cooling water. Temperature of the evaporator and the condenser in this research is the working fluid temperature at saturation condition on the evaporator and the condenser. Assumptions used in this research are: the system operates under steady state, there are no pressure drop and heat loss. The Temperature-Entropy diagram (T-S) of this cycle is shown in figure 2. Figure 1. Simple Configuration of ORC 2

4 Figure 2. T-S Diagram 2.2. Thermodynamic analysis Using the First and Second Law of Thermodynamics, the performance of working fluid for ORC can be evaluated under different condition. Based on figure 1 and 2, the mathematical expression for each of the components are given as follows: a. Pump (process 1-2) The work consumption in the pump is given in equation 1. w p,act = w p,ideal η p = m f (h 2s h 1 ) η p (1) where W p,act is the actual power of the pump (W), W p,ideal is the ideal power of the pump (W), m f is the working fluid mass flow rate (kg/s), η p is the isentropic efficiency of the pump (%), while h 1 and h 2s are the enthalpy of the working fluid at the inlet and at isentropic condition of the pump s outlet (J/kg), respectively. The exergy destruction rate (irreversibility) for uniform flow conditions can be expressed as in equation 2. ds I = T 0 = T dt 0m f [ s out s in + ( ds sistem ) + ( q j dt j )] (2) T j Here, ds sistem = 0 for steady state conditions, T 0 is ambient temperature ( o C). So, the exergy dt destruction in the pump can be expressed as in equation 3. I p = T 0 m f(s 2 s 1 ) (3) Here, s 1 and s 2 are the specific entropy of the working fluid at the inlet and outlet of the pump (J/kg.K), respectively. 3

5 b. Evaporator (process 2-3) Heat transfer in the evaporator can be expressed as in equation 4. Q e = m f (h 3 h 2 ) (4) Where h 3 and h 2 are the enthalpy of the working fluid at the outlet and inlet of the evaporator (J/kg), respectively. Using equation 2, the exergy destruction in the evaporator can be expressed as in equation 5. I e = T 0 [m f (s 3 s 2 ) m hf(s hf,o s hf,i )] (5) Where s 3, m hf, s hf,o and s hf,i are the specific entropy of the working fluid at the inlet of evaporator (J/kg.K), mass flowrate of heat source (kg/s), specific entropy of the heat source at outlet and inlet of evaporator (J/kg.K), respectively. c. Turbine (process 3-4) The turbine power is mathematically described by equation 6. w t,act = w t,ideal η t = m f (h 3 h 4s )η t (6) Here, W t,act is the actual power produced by turbine (W), W t,ideal is the ideal power of turbine (W), h 4s and η t are the enthalpy of the working fluid at the outlet isentropic condition of turbine (J/kg) and isentropic turbine efficiency (%), respectively. Meanwhile, the exergy destruction of the turbine can be expressed as in equation 7. I t = m f T 0 (s 4 s 3 ) (7) where s 4 is specific entropy at the outlet of the turbine (J/kg.K). d. Condenser (process 4-1) Heat transfer in the condenser can be expressed as in equation 8. Q k = m f (h 4 h 1 ) (8) The exergy destruction in condenser can be expressed as in equation 9. I k = T 0 [m f (s 4 s 1 ) m cf(s cf,o s cf,i )] (9) Where h 4, m cf, s cf,o dan s cf,i are enthalpy of working fluid at inlet of condenser (J/kg), mass flow rate of cooling fluid (kg/s), specific entropy of the cooling fluid at outlet and inlet of condenser (J/kg.K), respectively. The total of exergy destruction can be expressed as in equation 10. I total = I p + I e + I t + I k (10) 4

6 The net output work of the cycle: W net = m f (w t w p ) (11) Meanwhile, the cycle efficiency can be expressed as: η th = w t,act w p,act Q e (12) The exergy efficiency is defined as: η ex = w t,act w p,act Ex hf,i (13) Where the inlet exergy of the heat source is: Ex hf,i = m hf[(h hf,i h o ) T 0 (s hf,i s o )] (14) As shown in figure 3, working fluids can be classified into wet, isentropic, or dry, with respect to the slope of saturation curve in T-S diagram, i.e negative, infinite, or positive. The type of the working fluid, then, can be determined by using equation 15. ds dt = S 1 S 2 T 1 T 2 (15) where T 1 and T 2 are saturation temperature at evaporator and condenser, respectively. Meanwhile, s 1 and s 2 are specific entropy at 1 and 2 conditions, as shown in figure 3. Saturated vapor Saturated vapor Saturated vapor Saturated liquid Saturated liquid Saturated liquid (a) (b) (c) Figure 3. T s diagram for (a) wet, (b) isentropic and (c) dry working fluids 2.3. Operation parameters The operation parameters used for the simulation process are presented in table 1. In order to know the effect of parameter changes on exergy efficiency and exergy destruction, the turbine and condenser inlet temperature were used as variables. 5

7 3. Results and discussion Table 1. Operation Parameters Parameters Value Note Environmental pressure [kpa] Environmental temperature [ o C] 25 Working fluid mass flowrate [kg/s] 0.5 Evaporator temperature [ o C] 130 Condenser temperature [ o C] 30 Cooling fluid pressure [MPa] 0.2 Cooling fluid inlet temperature [ o C] 18 Water cooling mass flowrate [kg/s] 5 Pinch point evaporator [K] 6 Pinch point condenser [K] 6 Overheating [ o C] 0 Sub cool [ o C] 5 Flue gas inlet pressure [MPa] 0.4 [3] Flue gas inlet temperature [ o C] 170 Flue gas mass flowrate [kg/s] 2.57 [3] Isentropic efficiency of pump [%] 80 Isentropic efficiency of turbine [%] Analysis of working fluid performance In order to meet the objectives of this study, 33 organic fluid were evaluated, and their effect on the performance of the ORC are summarized in table 2. Based on net output work, cycle efficiency, and exergy efficiency, the use of water as working fluid by utilizing waste heat recovery has better performance than organic working fluids. The significant difference between water and organic fluids is the highest latent heat, where at the evaporator temperature, latent heat of evaporation of water is 2174 kj/kg while the organic fluids are ranged in kj/kg. Accordingly, water needs more thermal energy to change its phase from saturated liquid to saturated vapor, yields in lower mass flow rate required to absorb the same amount of thermal power from heat source than organic fluids. A higher mass flow rate leads to higher power consumption by pump and a higher piping system diameter required to overcome pressure losses. In real conditions, the use of conventional Rankine cycle by utilizing waste heat recovery is not efficient both in terms of thermodynamics and economic view. It can be explained through the T-S diagram of the simulation result presented in figure 4. The expansion process occurs in the saturated area (figure 4(a)), which will produce droplets in the turbine. It does not only make the expansion process fails to occur but the droplets could also be corrosive to the turbine [9]. In case of organic fluid, in as shown in figure 4(b), the expansion process occurs in superheated region. In addition, condensation process of organic fluids occurs at saturation pressure above atmospheric, while water occurs below atmospheric pressure. Condensation above atmospheric pressure is advantageous to avoid air intrusion to the system which can decrease the cycle s efficiency. Previous works recommend the minimum condensing pressure of ORC between MPa [5], [9], [10]. In terms of evaporator pressure, most of organic fluids have higher pressure than water, which will require more safety concern. Some studies suggest that the maximum pressure on the evaporator is between Mpa [5], [11], [12]. The two parameters can be limited factor to determine the best working fluid. 6

8 Table 2. Simulation result of working fluid performance Name Tc ( o C) Pc (Mpa) Wt (Watt) Wp (Watt) W_net (Watt) η th (%) η ex (%) I total (Watt) Pcd (Mpa) Pev (Mpa) Pratio Type Acetone Wet Benzene Dry Cis-2-butene Dry Cyclohexane Dry Diethylether Dry Dimethylcarbonate Dry HFE Dry HFE Dry Isobutene Isentropic Isopentane Dry mm Dry M-Xylene Dry N-butane Dry N-heptane Dry N-hexane Dry N-nonane Dry N-octane Dry N-pentane Dry R Isentropic R Dry R Dry R Dry R1233ZD(E) Dry R141B Dry R236EA Dry R245Fa Dry R365MFC Dry R600A Dry Toluene Dry Trans-2-butene Dry Isohexane Dry N-decane E Dry N-dodecane E Dry Water Wet Note: Tc = critical temperature, Pc = critical pressure, Pcd = condenser pressure, Pev = evaporator pressure, Pratio = P ev P cd 7

9 (a) (b) Figure 4. Process line of ORC using (a) water and (b) R245fa on T-S diagram 3.2. The effect of operation parameters Figure 5 shows the variation of the exergy efficiency of the system with the turbine inlet temperature, using selected organic fluids to represent wet, isentropic, and dry types. It was found that highest exergy efficiency for dry fluid type is given by Benzene, while for wet type organic fluid it is given by Acetone. Wet fluids like Water and Acetone need to be superheated, while dry or isentropic type of organic fluids does not need superheating [13]. The figure shows that the exergy efficiency of organic fluids can be related to the evaporation temperature in polynomial relationship until it approaches the critical temperature. Organic fluids with low critical temperature, such as HFE7100, R11, and R245fa, reveals the polynomial relationship clearly in the figure. Thereby, the selected organic fluids to be used in an ORC should not be operated at evaporation temperature near its critical temperature. Figure 5. Variation of the exergy efficiency with the turbine inlet temperature Figure 6 shows the variation of exergy destruction with the turbine inlet temperature. It can be observed that exergy destruction of several organic fluids increases with the increment of the turbine inlet temperature. The largest gradient of exergy destruction occurs in the n-octane, while acetone, R141b, R11, and isobutene show a decreasing trend of exergy destruction to the turbine inlet temperature. It can be concluded that turbine inlet temperature does not significantly contribute to the 8

10 changing of exergy destruction for certain organic fluids, such as Acetone, R141b, R11, and Isobutene. Figure 6. Variation of exergy destruction with the turbine inlet temperature Condensation temperature is one of the important parameters that affects ORC performance. The limitation of condenser temperature depends on the condenser pressure and freezing point of the working fluid. figure 7 shows the variation of exergy efficiency with the condenser inlet temperature. It can be observed that for all of the organic fluids, exergy efficiency decreases linearly with the increasing condenser inlet temperature. These results indicate that ORC will be more beneficial if operated at low ambient temperatures. The trend observed in this figure is consistent with the results shown in other figures where benzene shows the best exergy efficiency while HFE7100 the worse among the eight evaluated dry type organic fluids. Figure 7. Variation of the exergy efficiency with the condenser inlet temperature 9

11 3.3. The influence of selected thermodynamic properties to exergy efficiency Organic fluids that have high latent heat of vaporization, high density, and low specific heat are preferable. It absorbs more energy from heat sources, thereby reducing flow rate, pump power consumption, and the installation size [11]. However, Yamamoto et al. [14] suggested that low latent heat of vaporization provides better performance because the working fluid is in superheating conditions when entering the turbine. Based on exergy efficiency, latent heat of evaporation (turbine inlet temperature 130 o C at saturation pressure) has significant effect on the exergy efficiency, as shown in figure 8. It can be observed that the exergy efficiency of organic fluids increases with the latent heat. In addition, figure 9 illustrates the relationship between specific heat (Temperature =25 o C and Pressure = 1 atm) with exergy efficiency. The higher specific heat of the organic working fluid, there is a tendency to increase its exergy efficiency. Figure 8. Variation of the exergy efficiency with the latent heat Figure 9. Variation of the exergy efficiency with the specific heat 10

12 A good working fluid should have low vapor and low liquid specific volumes. These properties affect the rates of heat transfer in the heat exchangers. The vapor specific volumes relate directly to the size and cost of the cycle components. Moreover, a high vapor specific volume leads to larger volumetric flows. The specific volume of the liquid at the condenser pressure should be as small as possible to minimize the required pump work [15]. While the relationship of specific volume (Temperature: 130 o C at saturation vapor) with the exergy efficiency does not show good relationship, which is the correlation value of R 2 = It can be seen clearly in figure 10. Figure 10. Variation of the exergy efficiency with the specific volume Any fluid with temperature and pressure above its critical point shows no distinct phases of liquid and gas. When the organic fluid operates above the critical point, it will have a very high work potential due to increased pressure and temperature. Baik et al. [16] found that the output power produced by transcritical ORC using R125 as working fluid is 10% greater than subcritical ORC using R134a, R245fa, and R152a as working fluids. The study by Chen et al. [13] found that transcritical ORC using CO 2 as working fluid has higher cycle efficiency than using subcritical ORC using R123 as working fluid. Meanwhile, the selection of working fluids based on critical temperature shows a positive tendency, even though the correlation is quite low, i.e. R 2 =0.2765, as shown in figure 11. Figure 11. Variation of the exergy efficiency with the critical temperature 11

13 4. Conclusions Based on the analysis that has been done, it can be concluded that the utilization of heat source with temperature lower than 130 o C was inappropriate for a Conventional Rankine cycle since the expansion process takes place in the saturated area. Higher exergy efficiency was found in wet type organic working fluids, but it should be noted that the use of wet working fluids is not considerable for ORC. It was also found that Benzene had the highest exergy efficiency of about 10.49% for dry type organic working fluid. Meanwhile the highest exergy destruction of dry type organic working fluid was N- dodecane. The increasing turbine inlet temperature did not lead to the increasing exergy efficiency of organic working fluids with critical temperature near a heat source temperature. Exergy efficiency decreases linearly with the increase of condenser inlet temperature. In addition, it was also found that working fluid with high latent heat of vaporization and specific heat exert in high exergy efficiency. Meanwhile the parameters such as specific volume at saturation vapor and critical temperature do not show good relationship to the exergy efficiency. References [1] Le V L, Kheiri A, Feidt M and Pelloux-prayer S 2014 Thermodynamic and economic optimizations of a waste heat to power plant driven by a subcritical ORC (Organic Rankine Cycle) using pure or zeotropic working fluid Energy [2] BCS Inc 2008 Waste Heat Recovery: Technology and Opportunities in U.S Industry (US) [3] Situmorang H 2007 Utilization of exhaust gas from palm oil mill as heating feedwater woth shell and tube heat exchanger (Medan: North Sumatera University) [4] Bajaj S S, Patil H B, Kudal G B and Shisode S P 2016 Organic rankine cycle and tts working fluid selection-a review 4(4): [5] Drescher U and Bru D 2007 Fluid selection for the Organic Rankine Cycle (ORC) in biomass power and heat plants [6] Liu H, Liu Y and Li J 2011 A biomass-fired micro-scale CHP system with Organic Rankine Cycle (ORC) - Thermodynamic modelling studies Biomass and Bioenergy 35(9): [7] Lakew A A and Bolland O 2010 Working fluids for low-temperature heat source Appl. Therm. Eng. 30(10): [8] Bao J and Zhao L 2013 A review of working fluid and expander selections for Organic Rankine Cycle Renew. Sustain. Energy Rev [9] Tchanche B F, Lambrinos G, Frangoudakis A and Papadakis G 2011 Low-grade heat conversion into power using Organic Rankine Cycles - A review of various applications Renew. Sustain. Energy Rev. 15(8): [10] Katsanos C O, Hountalas D T and Pariotis E G 2012 Thermodynamic analysis of a Rankine cycle applied on a diesel truck engine using steam and organic medium Energy Convers. Manag [11] Maizza V and Maizza A 1996 Working fluids in non-steady flows for waste energy recovery systems Appl. Therm. Eng. 16(7): [12] Lai N A, Wendland M and Fischer J 2011 Working fluids for high-temperature organic Rankine cycles Energy 36(1): [13] Chen H, Goswami D Y and Stefanakos E K 2010 A review of thermodynamic cycles and working fluids for the conversion of low-grade heat Renew. Sustain. Energy Rev. 14(9): [14] Yamamoto T, Furuhata T, Arai N and Mori K 2001 Design and testing of the organic rankine cycle Energy 26(3): [15] Badr O, Probert S D and O Callaghan P W 1985 Selecting a working fluid for a Rankine-cycle engine Appl. Energy 21(1): 1 42 [16] Baik Y, Kim M, Chang K, Lee Y and Yoon H 2013 A comparative study of power optimization in low-temperature geothermal heat source driven R125 transcritical cycle and HFC organic Rankine cycles Renew. Energy