Performance analysis of a combined organic Rankine cycle and vapor compression cycle for power and refrigeration cogeneration

Accepted Manuscript Performance Analysis of a Combined Organic Rankine Cycle and Vapor Compression Cycle for Power and Refrigeration Cogeneration Kyoung Hoon Kim, Horacio Perez-Blanco PII:

S1359-4311(15)00406-8

DOI:

10.1016/j.applthermaleng.2015.04.062

Reference:

ATE 6582

To appear in:

Applied Thermal Engineering

Received Date: 1 October 2014 Revised Date:

18 April 2015

Accepted Date: 20 April 2015

Please cite this article as: K.H. Kim, H. Perez-Blanco, Performance Analysis of a Combined Organic Rankine Cycle and Vapor Compression Cycle for Power and Refrigeration Cogeneration, Applied Thermal Engineering (2015), doi: 10.1016/j.applthermaleng.2015.04.062. This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.

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Performance Analysis of a Combined Organic Rankine Cycle and Vapor Compression Cycle for Power and Refrigeration Cogeneration Kyoung Hoon Kim a and Horacio Perez-Blanco b, * a

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Department of Mechanical Engineering, Kumoh National Institute of Technology, 61 Daehak-ro, Gumi, Gyeongbuk 730-701, Korea, b Department of Mechanical and Nuclear Engineering, The Pennsylvania State University, 204 Reber building, University Park, PA 16802-1412, USA

Abstract

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* Corresponding author. Tel: 1-814-865-7842, Fax: 1-814-863-4848, E-mail [email protected]

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A thermodynamic analysis of cogeneration of power and refrigeration activated by low-grade sensible energy is presented in this work. An organic Rankine cycle (ORC) for power production and a vapor compression cycle (VCC) for refrigeration using the same working fluid are linked in the analysis, including the limiting case of cold production without net electricity production. We investigate the effects of key parameters on system performance such as net power production, refrigeration, and thermal and exergy efficiencies. Characteristic indexes proportional to the cost of heat exchangers or of turbines, such as total number of transfer units (NTUtot), size parameter (SP) and isentropic volumetric flow ratio (VFR) are also examined. Three important system parameters are selected, namely turbine inlet temperature, turbine inlet pressure, and the flow division ratio. The analysis is conducted for several different working fluids. For a few special cases, isobutane is used for a sensitivity analysis due to its relatively high efficiencies. Our results show that the system has the potential to effectively use low grade thermal sources. System performance depends both on the adopted parameters and working fluid.

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Keywords: organic Rankine cycle, vapor compression cycle, low grade source, thermal efficiency, exergy efficiency

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Nomenclature isobaric specific heat of source fluid, kJ/kg·K exergy flow, kW flow of exergy input at evaporator, kW flow of exergy input of source fluid, kW specific enthalpy, kJ/kg mass flow rate, kg/s mass flow rate of source fluid, kg/s total number of transfer unit of heat exchangers pressure, bar exit turbine pressure, bar. Also, condenser pressure, bar. critical pressure, bar evaporator pressure, bar boiler pressure, bar heat flow addition at boiler, kW refrigeration capacity at evaporator, kW heat removal rate at condenser, kW flow division ratio critical flow division ratio mass flow ratio of working fluid to source specific entropy, kJ/kg·K size parameter, m temperature, °C condensing temperature, °C critical temperature, °C cooling space temperature, °C cooling water temperature, °C evaporating temperature, °C turbine inlet pressure, bar turbine inlet temperature, °C source temperature, °C outlet source temperature, °C total heat transfer capacity of heat exchangers, kW/°C isentropic volume flow ratio volume flow rate at expander inlet, m3/s volume flow rate after isentropic expansion, m3/s compressor power, kW net power production, kW pump power, kW turbine power, kW

∆hs ∆Tm ∆Tpp ηc

specific enthalpy drop due to isentropic expansion in the expander, J/kg logarithmic temperature difference of a heat exchanger, °C pinch temperature difference, °C compressor isentropic efficiency

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cps E EEV Ein h m ms NTUtot P Pcd Pcr Pe Ph QBO QEV QCD rp rpc rs s SP T Tcd Tcr Tcs Tcw Te TIP TIT Ts Tsout UAtot VFR Vin Vs,out Wc Wnet Wp Wt

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exergetic efficiency isentropic efficiency of pump isentropic efficiency of turbine thermal efficiency

Superscripts/Subscripts

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reference state for thermodynamic properties source

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0 s

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ηex ηp ηt ηth

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1. Introduction

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The UNFCCC Copenhagen Accord and Cancun Agreements (UNFCCC 2009, 2010) on climate policy recognize that deep cuts in global greenhouse gas (GHG) emissions are needed so as prevent increases in global temperature greater than 2ºC above preindustrial levels. The required emission reductions are substantial in the 2050 time horizon, which will lead to a major transformation of the energy and economic systems worldwide [1]. Therefore, alternative energy sources have to be part of global energy solutions, and efficient use of lowgrade energy source such as geothermal energy, biomass combustion, or waste heat from various industrial processes will become increasingly important. Waste energy is also considered as carbon neutral, since it implies no additional emissions and it is energy that would otherwise be wasted. In recent years, the ORC and refrigeration systems using binary mixtures as working fluids have attracted much attention as they both harbor high probability of achieving high efficiency in converting low-grade thermal energy to more useful forms of energy [2-5]. ORC is a mature technology for low-temperature power generation. One major challenge for ORCs is temperature matching to the thermal energy source stream while heat is transferred to the ORC working fluid stream. Temperature matching to the source stream is important in minimizing the irreversibilities caused by heat transfer across a finite temperature difference. When the ORC is driven by a single phase-stream, temperature mismatching in the ORC evaporator is inevitable as the source exhibits a linear temperature profile, while the evaporating fluid exhibits constant temperature, or nearly so. Dai et al. [6] used a genetic optimization algorithm and identified isobutane and R236ea as efficient working fluids. Tranche et al. [7] investigated comparatively the performance of solar ORC using various working fluids. Volume flow rate, mass flow rate, power ratio as well as thermal efficiency were used for comparison. Hung et al. [8] examined Rankine cycles using organic fluids categorized into three groups of wet, dry and isentropic fluids. Kim and Han [9] carried out a thermodynamic performance analysis of transcritical organic Rankine cycles. Gao et al. [10] performed the analysis of a supercritical organic Rankine cycle system driven by exhaust heat using 18 organic working fluids. Li et al. [11] conducted an exergoeconomic analysis and performance optimization of a condenser for a binary mixture in ORC systems. Walraven et al. [12] investigated comparatively the performance of ORC and Kalina cycles. Wang et al. [13] proposed a theoretical model based on an ideal ORC to analyze the influence of working fluid properties on the thermal efficiency. Combined heat and power (CHP) systems and combined power and refrigeration systems are also becoming attractive due to the energy, economic, and environmental policies for pursuing stable electricity supply, sustainable development and environmental pollution mitigation [14]. Raj et al. [15] presented a review of renewable-activated cogeneration technologies. Feidt and Costea [16] presented a comparison of various CHP system configurations when different thermodynamic criteria are considered. The analysis confirms that the first-law efficiency criterion is only representative of the system thermal losses. Exergy efficiency, which takes into account the irreversibilities as well as quality of the thermal energy, allowed for a more precise optimization and comparison of performance of different thermal systems. Heberle and Brueggemann [17] analyzed the combined generation of heat and power

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generation from geothermal resources at temperatures below 450K using ORC in series and parallel circuits. Dai et al. [18], Li et al. [19], Habibzadeh et al. [20] presented a thermodynamic study focused on a thermal system combining the ORC and an ejector refrigeration cycle. Zhang and Lior [21], Pouraghaie et al. [22], Demirkaya et al. [23] carried out a thermodynamic analysis of another combined power and cooling cycle, combining the ORC and absorption cycles using ammonia-water mixture as the working fluid. Aphornratana and Sriveerakul [24] proposed a concept of an alternative heat-powered refrigeration cycle which combines an ORC and a VCC using a free-piston expander-compressor arrangement in which the compressor and expander are integrated in the same unit. The two systems would use the same working fluid and they would also share the same condenser. Wang et al. [25] proposed a thermally activated cooling cycle consisting of an ORC and a VCC. The system could be powered by solar thermal, geothermal or various waste heat streams. The shaft of the expander in the ORC and compressor in the VCC were directly coupled to reduce two-way energy conversion losses. Although there are thermally activated cooling technologies such as absorption cycles, they are generally used for large scale industrial applications as absorption chillers and the coefficient of performance (COP) is generally low for single- stage absorption cycles. The proposed ORC-VCC has some potential advantages over other thermally activated cooling systems. The total heat transfer requirement and the expander size are important technical and economic factors in ORC systems. To evaluate the expander size, Macchi [26] used two thermodynamic properties: the expander size parameter (SP) and the isentropic volume flow ratio (VFR). He reported that the physical significance of SP is given by its proportionality to actual turbine dimensions and VFR accounts for the compressibility effects in a more generalized way than other equivalent parameters of pressure ratio or Mach numbers. To evaluate the cost of heat exchangers, the total heat transfer capacity (UA)tot has been used, since it is considered to approximately reflect the heat transfer area of heat exchangers in the ORC system based on the hypothesis that the heat transfer coefficients for different ORC fluids tend to be fairly similar [10, 26-28]. In this work, cogeneration of power and refrigeration via combined ORC and VCC using various working fluids is analyzed. System performance as given by net power production, refrigeration, and the relevant indexes (SP, VFR and (UA)tot) already described. In addition, thermal and exergy efficiencies are projected. The most important design parameters are selected as turbine inlet temperature and pressure, and the flow division ratio defined as the ratio of mass flow rate in the ORC to that in the condenser. When the flow division ratio reaches its critical value, no net power is produced, namely all the power produced is internally consumed within the cycle. Hence, the system becomes in practice a heat-activated refrigeration cycle. 2. System Analysis The schematic diagram of the combined ORC and VCC is shown in Figure 1. Low-grade heat is supplied to the system as sensible energy in the boiler. The coolant enters the condenser at temperature Tcw and the working fluid leaves the condenser as saturated liquid at Tcd (state 1) at saturation pressure Pcd , namely the intermediate pressure of the system. The flow division ratio,

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rp, is defined as the ratio of mass flow rate in the ORC to that at the condenser. Then, out of 1 kg of the fluid exiting the condenser, rp kg of fluid flows to the ORC and the remaining (1 - rp) kg flows to the VCC. In the VCC, the fluid is throttled to the evaporator pressure (state 2), and leaves the evaporator as saturated vapor at temperature Te (state 3) and saturation pressure Pe, which is the low system pressure. Then, the fluid is compressed to the pressure Pcd by the compressor (state 4). In the ORC, the fluid pressure is increased by the pump to the turbine inlet pressure Ph which is the highest pressure of the system (state 6). The fluid flowing out from the recuperator enters the boiler (state 7). In the boiler, thermal energy is supplied at temperature Ts and the fluid is evaporated and superheated to a temperature Th at a pressure of Ph (state 8). The fluid is expanded in the turbine to a pressure Pcd and enters the recuperator (state 9). In the recuperator, the stream 9 temperature is reduced to that of state 10, preheating the heater inlet stream from 6 to 7. The fluid streams from the compressor and the recuperator, both at the same pressure Pcd, are mixed and then flow into the condenser (state 5). It is assumed that the minimum temperature differences between hot and cold streams in the heat exchangers should be greater than the prescribed value of the pinch temperature difference, ∆Tpp. The thermodynamic properties at points 4, 6 and 9 can be obtained in terms of the isentropic efficiencies of compressor, pump and turbine, ηc, ηp and ηt, respectively.

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Important cycle variables such as rates of heat addition in the boiler QBO, of heat rejection in the condenser QCD, refrigeration capacity QEV, turbine power Wt, pump power Wp, compression power Wc and net power production Wnet, can be obtained as follows: (1) QBO = rs ms (h8 − h7 ) = mCD rp ⋅ (h8 − h7 ) r (2) QCD = s ms (h5 − h1 ) rp QEV =

rs (1 − rp ) rp

ms (h3 − h2 )

Wt = rs ms (h8 − h9 )

W p = rs ms (h6 − h1 ) Wc =

rs (1 − rp )

(3) (4) (5)

ms (h4 − h3 )

(6)

Wnet = Wt − W p − Wc

(7)

rp

where h is the specific enthalpy, ms the mass flow rate of source fluid, and rs the mass flow ratio defined as the ratio of mass flow rate of working fluid in the ORC to that of the source fluid, namely: rs =

c ps (Ts − Ts ,out ) h8 − h7

(8)

where Ts,out is the source fluid boiler outlet temperature. The flow division ratio rp is defined as the ratio of mass flow rate at the pump to that at the condenser. Let us define the critical flow division ratio rpc as the value of rp when the ORC

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generates power exactly matching the power required to drive the pump and compressor so that Wnet = 0. Then rpc can be determined as follows; h4 − h3 (h1 + h4 + h8 ) − (h3 + h6 + h9 )

(9)

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rpc =

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Therefore, the flow division ratio should be greater than the critical value to generate both power and refrigeration, and there is no net power generation when rp equals the critical value rpc. In ORC systems the total heat transfer capacity UAtot can be regarded as proportional to the cost of heat exchangers, as already mentioned. The UAtot can be evaluated by the following equations

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 Q  Q Q Q Q UAtot = ∑  i  = BO + RC + CD + EV ∆Tm, HE ∆Tm, RC ∆Tm,CD ∆Tm, EV i  ∆Tm , i 

(10)

where RC indicates refrigeration and ΔTm represents the mean logarithmic temperature difference of the maximum and minimum temperature differences occurring in each heat exchanger, ΔTmax and ΔTmin as

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∆Tm =

∆Tmax − ∆Tmin ln(∆Tmax / ∆Tmin )

(11)

In this paper, the dimensionless total number of transfer units NTUtot is defined as the total heat capacity of the heat exchangers to the heat capacity of the source fluid as

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NTU tot =

UAtot ms c ps

(12)

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where cps is the isobaric specific heat of the source fluid. To evaluate the expander size, two thermodynamic properties, namely the size parameter SP and the isentropic volumetric flow ratio VFR have been used, defined as [26, 10, 28]

SP =

Vs ,out

∆hs V VFR = s ,out Vin

(13)

4

(14)

where Vin and Vs,out are the volume flow rates of the working fluid at the inlet and outlet of the expander after an isentropic expansion, respectively, and Δhs is the specific enthalpy drop calculated for an isentropic expander.

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The exergy is a property of a stream defined as the maximum useful work available when the stream evolves reversibly to reach equilibrium with the environment. The exergy of the working fluid can be evaluated as E = m[h − h0 − T0 (s − s0 )] (15) where m is the stream mass flow rate, s the entropy per unit mass of fluid, and subscript 0 denotes the reference dead state. Then, the rate of exergy input into the system by the source fluid Ein and the rate of exergy associated with the withdrawal of heat at the evaporator from the cooled space EEV, are evaluated as follows:

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  T  Ein = ms c ps Ts − T0 − T0 ln s   T0   T  EEV = QEV  0 − 1  Tcs 

(16)

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where cps is the isobaric specific heat of the source fluid, and Tcs the temperature of the cooled space. The thermal efficiency ηth is defined as the ratio of the usable energy of power and refrigeration to the heat input, and the exergy efficiency ηex is defined as the ratio of the usable exergy of power and refrigeration to the exergy input. Hence, the efficiencies are given by:

ηth =

W

ηex =

W

net

+ Q EV

Q BO

net

+ E EV

E in

(18) (19)

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Eight working fluids are considered in this study: R143a, R22, R134a, R152a, propane, ammonia, isobutane, and butane, which are sequenced by their critical temperature. The thermodynamic properties of the working fluids are calculated using the Patel-Teja equation of state [29-30]. As reference states, entropy and enthalpy were assigned null values at the condenser exit (saturated liquid, state 1). This convention was adopted for convenience and the chosen references do not influence the results of the analysis. What defines cycle performance are the differences in entropy and enthalpy, rather than their absolute values. The basic data of each fluid, needed to calculate the thermodynamic properties are given in Table 1, where M, Tcr, Pcr, and ω are molecular weight, critical temperature, critical pressure, and accentric factor, respectively [31]. The temperature-entropy diagrams for the working fluids are shown in Fig. 2. 3. Results and Discussion

The source fluid adopted here is air at Ts = 150oC with a mass flow rate of 1 kg/s. The boiler, recuperator and condenser are considered to be operated under the condition that the minimum temperature difference between the hot and cold streams in the heat exchangers reach the prescribed pinch value of ∆Tpp = 10oC. Other basic data for analysis are given in Table 2.

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3.1 Performance of the refrigeration mode

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In this section, it is considered that the working fluid enters the turbine as a superheated vapor. The flow division ratio equals the critical value, so only refrigeration is produced. The critical flow division ratios (rpc) are plotted against the turbine inlet pressure (TIP) in Fig. 3 for various working fluids. For a fixed value of condenser pressure, the critical flow division ratio decreases with increasing TIP, since the specific turbine work increases as TIP increases, and therefore a smaller mass flow rate is required in the ORC for the same compression work in the VCC. For a fixed TIP, the rpcs of R143a, ammonia or R22 are high, while those of butane or isobutane are low. It can be seen from the figure that working fluids with higher critical temperatures generally show lower rpcs, with the notable exception of ammonia.

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Figure 4 displays the effects of TIP on the mass flow ratio rs for the various working fluids. The mass flow ratio (rs) is defined as the ratio of the mass flow rate of working fluid in the ORC to that of the source fluid and is evaluated from the energy balance boiler (Fig.1, Eq. (8)). Several factors are at work in Eq.8: when the inlet and outlet temperatures of the working fluid and the inlet temperature and mass flow rate of the source fluid are specified, its outlet temperature decreases with increasing mass flow rate of the working fluid. When the temperature difference between the streams in the heat exchanger reaches the prescribed pinch value ∆Tpp, the mass flow rate of the working fluid reaches a maximum. Then, as the TIP increases (Fig.4), the mass flow ratio decreases, reaching a minimum that depends on the working fluid. The existence of the minimum stems from the following considerations: when the TIP is low, the refrigerant at the boiler inlet (Fig.1, state 7) is in saturated liquid-vapor state. As the TIP increases, the saturation temperature of the working fluid and thus the outlet temperature of the source fluid increase, decreasing the mass flow ratio. When the TIP is high, the refrigerant at the boiler inlet (Fig.1, state 7) becomes a compressed liquid and the outlet temperature of the source fluid decreases with increasing TIP in the counter flow heat exchanger, in such way that the mass flow ratio increases. This two opposing trends result on a minimum boiler mass flow ratio as TIP increases. For a fixed TIP, the mass flow ratios of R143a or R22 are high, while the ratios of ammonia or butane are quite low.

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The refrigeration capacity QEV varies with TIP as shown in Fig. 5 for the various working fluids. With increasing TIP, QEV increases monotonically for R143a, R22, R134a, R152a, and ammonia. The profiles for butane, isobutane, and propane exhibit a maximum, due to the effects of the mass flow ratios. Generally speaking, the refrigeration capacities are large for low TIPs in the case of working fluids of high critical temperatures such as butane or isobutene. Conversely, fluids of low critical temperature such as R134a or R152a exhibit large refrigeration capacities at high TIPs. The explanation for these diverging behaviors can be traced to Fig. 3. This figure shows that that the flow division ratio, rp (rpc in this case) behaves like a convex parabolic curve with respect to TIP for the working fluids with relatively high critical temperature such as butane, but it decreases monotonically for the working fluids with relatively low critical temperature such as R134a. The mass flow ratio rs increases with increasing TIP as the subcooled boiler load increases. These opposing trends result in the complex nonlinear behaviors of the refrigeration capacity of Fig. 5. Qev generally increases with increasing TIP, mainly due to the increased

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critical flow division ratio. For working fluids with relatively high critical temperatures (e.g. butane) it behaves like a parabolic curve, because the critical flow division ratio behaves like a convex parabolic curve. The refrigeration capacity for R134a shows an inflexion, mainly due to the increased boiler load to bring the liquid to saturation for high TIPs.

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Variations of the thermal efficiency vs. TIP are shown in Fig. 6 for various working fluids. Generally, the thermal efficiency increases monotonically with increasing TIPs in the subcritical pressure region. However, for working fluids of high critical temperature such as butane or isobutane, the thermal efficiency reaches a maximum in the subcritical pressure region. For a fixed TIP, the higher the critical temperature, the higher thermal efficiencies are, except for ammonia. It is remarkable that the range of possible efficiencies is large, varying from 40 to 70% for the given TIP range. In all cases, the thermal efficiencies are well below that assigned by Carnot to a reversible, heat-activated cycle, (about 2 in this case).

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For a fixed source temperature and no power production, the exergy efficiency is proportional to the refrigeration capacity, QEV. Figure 7 shows the how the thermal and exergy efficiencies are related. In general, thermal end exergetic efficiencies are proportional to each other, to a point. When the thermal efficiency rises above 40%, (i.e. at high TIPs) the exergy efficiency of R134a or R152a increases faster than the thermal efficiency, while those of isobutane or butane actually decrease. For butane and isobutane, the TIP for the maximum exergy efficiency is lower than that for the maximum thermal efficiency. For these two refrigerants, it can be shown that as TIP increases, the refrigeration capacity decreases. Yet, the boiler heat input decreases even more steeply, and the thermal efficiency keeps on gaining. Since the exergy output decreases with refrigeration capacity, the net effect for these refrigerants is reflected in Fig.7: the thermal efficiency increases but the exrgetic efficiency decreases with TIP. For all the other refrigerants, the refrigeration capacity keeps on gaining with TIP, and so does the exergetic efficiency. The thermal efficiencies are much higher than the exergetic ones for these cases (by a factor of about 10), which can be interpreted as this cycle being capable of extracting large quantities of available energy from the relatively low temperature waste heat, while converting it to refrigeration that, whereas of considerable societal value, has relatively low availability. To appraise the overall viability of an ORC system, resort is had to a number of parameters generally regarded as acceptable indicators. The total heat transfer capacity UAtot, has been used in the past as an index of the relative cost of similar heat exchangers constructed of the same materials [28]. Variations of the dimensionless total number of transfer units NTUtot , defined as the ratio of UAtot to the thermal capacity of source fluid ms·cps, are plotted against the TIP in Fig. 8 for all the working fluids. Since the product ms·cps is nearly constant, UAtot is essentially proportional to NTUtot for the purposes of this analysis. When the working fluid is butane, isobutane, ammonia or propane, the number of transfer unit decreases with increasing TIPs. However, the number of transfer units stays constant or actually increases for R134a, R22 or R143a. Butane offers the possibility of comparatively low NTUstot. As the reader might recall from the preceding sections, the size parameter SP accounts for the

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3.2. Performance of Cogeneration, Superheated Vapor

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actual turbine dimensions and the isentropic volume flow ratio VFR accounts for the effects of the compressibility through the expansion [27]. Figure 9 displays the variations of the size parameter and of the volume flow ratio for various working fluids. Small values of SP and of VFR are preferred, in that they point to a small turbine with an expansion close to reversibility. It can be observed from the figure that the VFR decreases with increasing SPs, which also coincides with decreasing TIPs. So, cycles closer to reversible expansion call for larger compressors, although the effect is less intense for R143a and for R22. Generally, fluids of high critical temperatures show large SPs at high TIPs, while ammonia, R152a, or propane exhibit smaller SPs at at high TIPs. Also, working fluids of low critical temperature exhibit low VFRs. The fluids associated with high thermal efficiencies and small NTUtot, namely butane and isobutane, tend to result on large SPs and VFRs. By comparison, ammonia shows lower VFRs than isobutane or butane for the same SPs.

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This section of the paper focuses only on isobutene, , that generally exhibits high efficiencies. This fluid enters the turbine as a superheated vapor with fixed value of TIT = 130oC (Fig.1, state 8). The temperature-entropy diagrams for the combined ORC-VCC are shown in Fig. 10 for various TIPs. It can be seen from the figure that for a fixed TIT, the entropy at the turbine inlet increases with decreasing TIP, resulting on a turbine exit temperature that increases with decreasing TIPs. Thus, at the recuperator inlet (Fig.1, state 6), the working fluid temperature increases with decreasing TIP (states 9a, 9b, 9c and 9d)) and the fluid from the pump (state 6) upon exiting the recuperator (state 7, all) may be a saturated mixture when TIP is rather low, for example, when TIP equals 10 bar (state 7a).

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Figures 11and 12 show the effects of TIP on the refrigeration capacity QEV and the net power production Wnet respectively, for various flow division ratios. For a fixed flow division ratio, QEV decreases with increasing TIP due to the decrease in the mass flow ratio rs as shown in Fig. 4. For the condition of critical flow division ratio, QEV has a maximum value at a TIP of about 20 bar. For a fixed TIP, QEV decreases with increasing flow division ratio due to the decreasing evaporator to condenser mass flow ratio. Concerning Fig.12, a peak value exists for Wnet with respect to TIP for a fixed flow division ratio. The magnitude of the peak increases with increasing flow division ratios and reaches the maximum value when the flow division ratio becomes unity. The effect of TIP on the thermal and exergy efficiencies is shown in Figs 13-14, with the flow division ratio as parameter. The thermal efficiency increases with TIP for a fixed flow division ratio. Whereas the specific refrigeration, specific net work, and specific heat input decrease with increasing TIP for a fixed flow division ratio, the decrease of heat input is dominant among them, which explains the thermal efficiency increase. The exergy efficiency (Fig.14) has a well-defined peak with respect to TIP for a fixed flow division ratio. For a fixed TIP, the thermal efficiency decreases whereas the exergetic efficiency increases with increasing flow division ratios. It appears that for this cycle, a TIP in the order of 20 bar in conjunction with large flow division ratios leads to optimal exergy use. At the peak, the system outputs the largest amount of exergy for the given input, which is constant. The exergy peak occurs substantially at 20 bars. This can be explained because the

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thermal efficiency defined as Eq. (18) is proportional to QEV and inversely proportional to the heat flow addition at the boiler, QBO. For the condition of superheated vapor at the turbine inlet, Figs. 11-12 show that the refrigeration capacity decreases slightly with increasing TIP, however, the net power production also has a peak value with respect to TIP substantially at 20 bar. The latter effect is more dominant than the former. Therefore, the exergy efficiency has a peak with respect to TIP as is shown in Fig. 14.

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Fig. 15 shows the effects of TIP on the total number of transfer units NTUtot for various flow division ratios. The total number of transfer units decrease with increasing TIP or with increasing flow division ratios. Therefore, for a fixed flow division ratio, high TIPs lead to high thermal efficiencies and require a small total number of transfer units. For a fixed turbine inlet pressure, a low flow division ratio leads to high thermal efficiency but requires a large number of transfer units. The preceding observations show that producing refrigeration (decreasing rp values) leads to increased thermal efficiencies but requires additional NTUs. Remarkably, the optimal exergetic efficiency requires a TIP of about 20 bar, resulting on a well-defined limit for the NTUs.

3.3 Performance of Cogeneration, Saturated Vapor

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In this section, we consider cases where isobutane enters the turbine as a saturated vapor. The temperature-entropy diagrams are shown in Fig. 16 for various turbine inlet temperatures (TIT’s) for the critical flow division ratios. Recall in this case that the ORC produces power just to meet the requirements of the pump and of the compressor. The turbine exit temperature (state 9) increases with TIT, however, the increase is not remarkable. Because of this behavior, preheating of the working fluid in the recuperator is insignificant, and the working fluid consequently enters the boiler at a relatively low temperature. The effects of TIT on the QEV for various flow division ratios are shown in Fig. 17. For a fixed flow division ratio, QEV decreases with increasing TIT, since at the evaporator, (h3 - h2) remains constant but the mass flow rate of the working fluid per unit mass flow rate of the source, rs, decreases with increasing TIT. For a fixed TIT, the refrigeration capacity decreases with increasing flow division ratio due to the decrease of mass flow rate at the evaporator. It is to be noted that when the flow division ratio equals unity, there is no refrigeration, since the evaporator receives no refrigerant. For the critical flow division ratio, the refrigeration capacity (Eq. 3) has a peak value with respect to TIT, since an increasing contribution due to (1/rp - 1), is counterbalanced by a decreasing contribution due to rs, as given by Eq.(3). Fig. 18 shows the effects of TIT on the net power Wnet. For a fixed flow division ratio, Wnet exhibits a peak value with respect to TIT, since it has an increasing contribution of turbine power as TIT increases, and a decreasing contribution due to the decreasing amount of input thermal energy by the waste heat source fluid. For a fixed TIT, Wnet becomes zero when the flow division ratio reaches the critical value, and increases with increasing flow division ratio, reaching a maximum value when the flow division ratio becomes unity. At that point, only power is produced, with no accompanying refrigeration.

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The profiles of thermal and exergy efficiency (Eq.18) vs. TIT are shown in Figs. 19-20. For a given flow division ratio, the thermal efficiency increases with TIT, Fig.19. This can be explained as follows. For a fixed flow division ratio, the refrigeration capacity QEV decreases with increasing TIT (Fig.17) while the net power production Wnet has a peak value with respect to TIT (Fig.18). Hence, one would expect either a decreasing efficiency or a peaking one with TIT. However, the boiler heat input decreases with increasing TIT, lending a dominant positive slope to the plot. The exergy efficiency, Eq.19, exhibits a peak value with respect to TIT for a fixed flow division ratio (Fig.20). The peak occurs for TITs in the range of 75-100 C. The peaks of Fig. 20 occur for saturated vapor at the turbine inlet, because as shown in Figs. 17-18 the refrigeration capacity decreases slightly with increasing turbine inlet temperature (TIT), but the net power production has a peak value with respect to TIT (Fig.18). The latter effect being more dominant than the former, the exergy efficiency has a resulting peak with respect to TIT as is shown in Fig. 20. This means that increasing the refrigeration duty is more advantageous from the viewpoint of energy, but increasing the power generation duty is more advantageous from the viewpoint of exergy, irrespective of the economic viewpoint.

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Figure 21 shows the effects of TIT on the total number of transfer units NTUtot for various flow division ratios. The total number of transfer units decreases with increasing TIT or increasing flow division ratios. Therefore, for a fixed flow division ratio, high turbine inlet temperatures lead to high thermal efficiencies and require small NTUs. For a fixed turbine inlet temperature, small flow division ratios lead to high thermal efficiencies but require large NTUs.

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4. Conclusions The thermodynamic performance of the cogeneration of power and refrigeration by an organic Rankine cycle (ORC) and vapor compression cycle (VCC) utilizing low grade sensible energy is analyzed. The cases of pure refrigeration and cogeneration, including the cases of saturated and superheated vapor at the turbine inlet are investigated in the subcritical pressure regions. The eight working fluids considered are: R143a, R22, R134a, R152a, propane, ammonia, isobutane, and butane. A systematic investigation of the effects of system parameters on performance such as specific refrigeration, net work production, thermal and exergy efficiencies, total number of transfer units, and size parameter and volume flow ratio for the turbine of the system led to a selection of dominant parameters. The three important parameters are the turbine inlet temperature, the turbine inlet pressure, and the flow division ratio. In the case of pure refrigeration, our results show that the working fluids with higher critical temperature, except ammonia, show higher thermal efficiency, and the refrigeration per unit mass of source fluid is large for working fluids of high critical temperature. In that regard, butane or isobutane are good candidates for refrigeration at low turbine inlet pressures. Generally, flammability is not regarded as a desirable characteristic for some working fluids, which arises the consideration of low critical temperatures such as those of R134a or R152a. Those compounds tend to exhibit better refrigeration performance at high turbine inlet pressures. The NTUs have high values for R134a, R143a or R22. Whereas the former two may yet find acceptance for heat recovery cycles suitable configured, the latter is in the process of being eliminated as a viable option due to ozone layer considerations. In the case of cogeneration, results show that the exergy efficiency has a peak value with respect to both TIT and TIP. The Second Law trends are met in that TITs or TIPs lead to high

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thermal efficiencies while requiring small NTUs. Small flow division ratios lead to high thermal efficiencies, but require large NTUs. Isobutane offers good thermal efficiencies, and we carried out a sensitivity analysis of performance versus TIP and TIT, further elucidating how those variables influence performance and size of the ORC-VCC. The work presented here covers the behavior of the cycle of Fig.1 from sole power generation (rp=1) to sole refrigeration (rp=rpc). Generally, the production of power from waste heat preserves exergy more effectively than the production of refrigeration. Whereas the topic is not investigated here, it can be speculated that power production via thermal cycles requires heat rejection, and that matching high-grade heat energy to power production and waste heat to refrigeration could very well be a preferred option from an exergy viewpoint. Significantly, the exergy efficiency peaks at TIPs of 20 bar and at TITs of about 100 C for all the flow division ratios of this study. The combined ORC-VCC cycle exhibits the potential to efficiently utilize low-grade thermal sources. The present study provides relevant information towards judicious selection of working fluid and operational conditions. Such selection is bound to lead to energy savings over other fluids and conditions, and reduced investment costs.

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Acknowledgements

References

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This research was supported by Basic Science Research Program through the National Research Foundation of Korea (NRF) funded by the Ministry of Education, Science and Technology (No. 2010-0007355).

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[1] Ciscar J., Saveyn B., Soria A., Szabo L., Regemorter D. V., Ierland T. V., A Comparability analysis of global burden sharing GHG reduction scenarios. Energy Policy 55 (2013) 73-81. [2] Dresher U., Brueggemann D., Fluid selection for the Organic Rankine Cycle (ORC) in biomass power and heat plants. Applied Thermal Engineering 27 (2007) 223-228. [3] He Y. L., Mei D. H., Tao W. Q., Yang W. W., Liu H. L., Simulation of the parabolic trough solar energy generation system ith organic Rankine cycle, Applied Energy 97 (2012) 630-641. [4] Kim K. H., Han C. H., Kim K., Effects of ammonia concentration on the thermodynamic performances of ammonia-water based power cycles. Thermochimica Acta 530 (2012) 7-16. [5] Kim K. H., Han C. H., Kim K. Comparative exergy analysis of ammonia-water based Rankine cycles with and without regeneration. Int. J. Exergy (2013) in press. [6] Dai Y., Wang J., Gao L., Parametric optimization and comparative study of organic Rankine cycle (ORC) for low grade waste heat recovery. Energy Conversion and Management 50 (2009) 576-582. [7] Tranche B. F., Papadakis G., Lambrinos G., Frangoudakis A., Fluid selection for a lowtemperature solar organic Rankine cycle. Applied Thermal Engineering 29 (2009) 2468-2476. [8] Hung T. C., Wang S. K., Kuo C. H., Pei B. S., Tsai K. F., A study of organic working fluids on system efficiency of an ORC using low-grade energy sources. Energy 35 (2010) 1403-1411. [9] Kim K. H., Han C. H., Analysis of Transcritical organic Rankine cycles for low-grade heat conversion. Advanced Science Letters 8 (2012) 216-221. [10] Gao H., Liu C., He C., Xu X., Wu S., Li Y., Performance analysis and working fluid selection of a supercritical organic Rankine cycle for low grade waste heat recovery. Energies 5 (2012) 3233-3247. [11] Li Y. R., Du M. T., Wu S. Y., Peng L., Liu C., Exergoeconomic analysis and optimization of a condenser for a binary mixture of vapors in organic Rankine cycle, Energy 40 (2012) 341-347. [12] Walraven D., Laenen B., D’haeseleerW., Comparison of thermodynamic cycles for power production from low-temperature geothermal heat sources, Energy Converse and Management 66 (2013) 220-233. [13] Wang D., Ling X., Peng H., Liu L., Tao L. L., Efficiency and optimal performance evaluation of organic Rankine cycle for low grade waste heat power generation, Energy 50 (2013) 343352. [14] Tsai W. T., Regulatory compliance and environmental benefit analysis of combined heat and power (CHP) systems in Taiwan. Energies 6 (2013) 557-565. [15] Raj N. T., Iniyan S., Goic R., A review of renewable energy based cogeneration technologies. Renewable and Sustainable Energy Reviews 15 (2011) 3640-3643. [16] Feidt M., Costea M., Energy and exergy analysis and optimization of combined heat and power systems. Comparison of various systems. Energies 5 (2012) 3701-3722. [17] Heberle F., Brueggemann D., Exergy based fluid selection for a geothermal organic Rankine

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List of Tables Table 1. Basic data of working fluids.

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Table 2. Basic calculation conditions for the system

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Table 1. Basic data of working fluids

M(kg/kmol)

Tcrc(K)

Pcr(bar)

ω

R143a

84.041

346.25

37.58

0.253

R22

86.468

369.30

49.71

R134a

102.031

380.00

36.90

R152a

66.051

386.60

44.99

propane

44.096

396.82

ammonia

17.031

isobutane butane

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Substance

0.219

0.239

0.263

0.152

405.65

112.78

0.252

58.123

408.14

36.48

0.177

58.123

425.18

37.97

0.199

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42.49

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Table 2. Basic calculation conditions for the system symbol

description

data unit

Ts

source temperature

150 °C

Tcw

cooling water temperature

Tcd

condenser temperature

T0

reference dead state temperature

∆Tpp

pinch temperature difference

Tcs

cooling space temperature

Te

evaporator temperature

ηp

isentropic efficiency of pump

ηt

isentropic efficiency of turbine

80 %

ηc

isentropic efficiency of compressor

80 %

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25 °C

40 °C 25 °C

10 °C

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15 °C 5 °C

80 %

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List of Figures

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Fig. 1 Schematic of the combined ORC-VCC system. Fig. 2 Temperature-entropy diagrams for the working fluids. Fig. 3 Variation of the critical flow division ratios with turbine inlet pressure. Fig. 4 Variation of the mass flow ratio with turbine inlet pressure. Fig. 5 Variation of the refrigeration capacity with turbine inlet pressure. Fig. 6 Variation of the thermal efficiency with turbine inlet pressure. Fig. 7 Exergy efficiency vs. thermal efficiency for all refrigerants. Fig. 8 Variation of the total number of transfer units with turbine inlet pressure. Fig. 9 Volume flow ratio vs. size parameter. Fig. 10 ORV-VCC state points in T-S chart for various turbine inlet pressures, isobutane. Fig. 11 Refrigeration capacity vs. turbine inlet pressure, isobutane.Fig. 12 Net power production vs. turbine inlet pressure, isobutane. Fig. 13 Thermal efficiency vs turbine inlet pressure, isobutane. Fig. 14 Exergy efficiency vs. turbine inlet pressure, isobutane. Fig. 15 Total number of transfer unit vs. turbine inlet pressure, isobutane. Fig. 16 ORV-VCC state points in T-S chart for various turbine inlet temperatures, isobutane. Fig. 17 Refrigeration capacity vs turbine inlet temperature, isobutane. Fig. 18 Net power production vs. turbine inlet temperature, isobutane. Fig. 19 Thermal efficiency with turbine inlet temperature, isobutane. Fig. 20 Exergy efficiency vs. turbine inlet temperature, isobutane. Fig. 21 Total number of transfer units vs. turbine inlet temperature, isobutane.

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8

Boiler

Turbine

QBO

9 Recuperator

10 4

5

6 Condenser

Evaporator

QEV

2

Expansion valve

Pump

rp 1-rp

QCD

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3

Compressor

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7

Receiver

1

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Fig. 1 Schematic of the combined ORC-VCC system.

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160

120

o

Temperature [ C]

140

100

8 7 5

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1: R143a 2: R22 3: R134a 4: R152a 5: propane 6: ammonia 7: isobutane 3 8: butane 2

4

6

80 1

40 0

20

40

60

80

100

120

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60

140

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Fig. 2 Temperature-entropy diagrams for the working fluids.

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R143a R22 R134a R152a propane ammonia isobutane butane

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0.9

0.8

0.7

0.6

0.5 10

15

20

25

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Critical flow division ratio, rpc

1.0

30

35

40

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Turbine inlet pressure [bar]

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Fig. 3 Variation of the critical flow division ratios with turbine inlet pressure.

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3.0 R143a R22 R134a R152a

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2.0 1.5

1.0

0.5

0.0 10

15

20

25

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Mass flow ratio at boiler, rs

2.5

propane ammonia isobutane butane

30

35

40

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Turbine inlet pressure [bar]

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Fig. 4 Variation of the mass flow ratio with turbine inlet pressure.

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150

100

R143a R22 R134a R152a propane ammonia isobutane butane

50

0 10

15

20

25

30

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Refrigeration capacity, QEV [kW]

200

35

40

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Turbine inlet pressure [bar]

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Fig. 5 Variation of the refrigeration capacity with turbine inlet pressure.

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80

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60 50 40 30 20

R143a R22 R134a R152a

10 0 10

15

20

25

30

propane ammonia isobutane butane

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Thermal efficiency, ηth [%]

70

35

40

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Turbine inlet pressure [bar]

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Fig. 6 Variation of the thermal efficiency with turbine inlet pressure.

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8

R143a R22 R134a R152a propane ammonia isobutane butane

5 4

High TIP

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6

3 2 1

Low TIP

0 0

10

20

30

40

50

60

70

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SC

Exergy efficiency, ηex [%]

7

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Fig. 7 Exergy efficiency vs. thermal efficiency for all refrigerants.

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18 16

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12 R143a R22 R134a R152a propane ammonia isobutane butane

10 8 6 4 10

15

20

25

30

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NTUtot

14

35

40

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Turbine inlet pressure [bar]

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Fig. 8 Variation of the total number of transfer units with turbine inlet pressure.

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10

8 7

VFR

6 5 4 3 2

0.010

0.015

0.020

0.025

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SP [m]

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low TIP

1

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R143a R22 R134a R152a propane ammonia isobutane butane

high TIP

9

0.030

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Fig. 9 Volume flow ratio vs. size parameter.

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120

o

TIT = 130 C o Tcd = 40 C

100

9a 9b

o

Te = 5 C

80

7d

60

7c

9d

7a

Organic Rankine Cycle (ORC)

6

40

9c

7b

1

20

2

4

Vapor Compression Cycle (VCC)

3

0 0.0

0.2

0.4

0.6

0.8

1.0 o

TIP 10 bar(a) 15 bar(b) 20 bar(c) 25 bar(d)

1.2

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10 5

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Temperature [ C]

8d 8c 8b 8a

combined ORC-VCC

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isobutane o Ts = 150 C

140

1.4

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Fig. 10 ORV-VCC state points in T-S chart for various turbine inlet pressures, isobutane.

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rpc

rp

0.8 0.9 1.0

0.6 0.7

140

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120 100 80 60 40 20 0 10

15

20

25

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Refrigeration capacity, QEV[kW]

160

30

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Fig. 11 Refrigeration capacity vs. turbine inlet pressure, isobutane.

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rpc

rp

0.8 0.9 1.0

0.6 0.7

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25 20 15 10 5 0 10

15

20

25

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Net power production, Wnet [kW]

30

30

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Turbine inlet pressure [bar]

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Fig. 12 Net power production vs. turbine inlet pressure, isobutane.

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80 rpc

60

0.8 0.9 1.0

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rp

40

20

0 10

15

20

25

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Thermal effficiency, η th [%]

0.6 0.7

30

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Fig. 13 Thermal efficiency vs turbine inlet pressure, isobutane.

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35

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25 20 15 10 5

rp 0 10

15

rpc 0.6 20

0.9 1.0

0.7 0.8 25

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Exergy efficiency, η ex [%]

30

30

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Turbine inlet pressure [bar]

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Fig. 14 Exergy efficiency vs. turbine inlet pressure, isobutane.

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16 rpc

rp

14

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12

NTUtot

0.8 0.9 1.0

0.6 0.7

10 8

4 10

15

20

25

SC

6

30

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Fig. 15 Total number of transfer unit vs. turbine inlet pressure, isobutane.

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o

Tcd = 40 C o

100

Te = 5 C

80

combined ORC-VCC

o

Organic Rankine Cycle (ORC)

60

6 1

40

8d 8c 8b 8a

7

4 Vapor Compression Cycle (VCC)

20 0

3

2 0.0

9 510

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Temperature [ C]

120

TIT o 90 C(a) o 100 C(b) o 110 C(c) o 120 C(d)

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isobutane o Ts = 150 C

140

0.2

0.4

0.6

0.8 o

1.0

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1.2

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Fig. 16 ORV-VCC state points in T-S chart for various turbine inlet temperatures, isobutane.

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200 rpc

0.8 0.9 1.0

0.6 0.7

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150

100

50

0 80

90

100

110

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Refrigeration capacity, QEV[kW]

rp

120

o

130

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Fig. 17 Refrigeration capacity vs turbine inlet temperature, isobutane.

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40 rpc

0.8 0.9 1.0

0.6 0.7

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30

20

10

0 80

90

100

110

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Net power production, Wnet [kW]

rp

120

130

o

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Fig. 18 Net power production vs. turbine inlet temperature, isobutane.

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0.6 0.7 rp 0.8 0.9 1.0

40 30 20 10 0 80

90

100

110

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Thermal efficiency, ηth [%]

60

120

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o

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Fig. 19 Thermal efficiency with turbine inlet temperature, isobutane.

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25 20 15 10 5

rpc

rp 0 80

90

0.9 1.0

0.7 0.8

0.6 100

110

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Exergy efficiency [%]

30

120

130

o

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Fig. 20 Exergy efficiency vs. turbine inlet temperature, isobutane.

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rp

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10 NTUtot

0.8 0.9 1.0

0.6 0.7

8

4 80

90

100

110

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130

o

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Fig. 21 Total number of transfer units vs. turbine inlet temperature, isobutane.

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Highlights

1. Waste heat utilization can reduce emissions of carbon dioxide.

3. Efficiencies and size parameters are used for cycle evaluation.

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2. The ORC/ VCC cycle can deliver power and/or refrigeration using waste heat.

4. The cycle performance is studied for eight suitable refrigerants. Isobutane is used for a sensitivity analysis.

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5. The work shows that the isobutene cycle is quite promising.