Document 7RN8QawNgQ5vKpe159E1yGjKE

International Journal of Refrigeration 90 (2018) 229-237 Contents lists available at ScienceDirect International Journal of Refrigeration journal homepage: www.elsevier.com/locate/ijrefrig Theoretical evaluation of different high-temperature heat pump 011) configurations for low-grade waste heat recovery Check for updates Carlos Mateu-Royo , Joaquin Navarro-Esbri, Adrian Mota-Babiloni, Marta Amat-Albuixech, Francisco Moles ISTENER Research Group, Department of Mechanical Engineering and Construction, Universitatfaume I, Campus de Riu Sec s/n, E12071 Castellon, Spain ARTICLE INFO Article history: Received 1 February 2018 Revised 17 April 2018 Accepted 18 April 2018 Available online 24 April 2018 Keywords: Waste heat recovery Low GWP refrigerants Industrial heat pump Energy efficiency Greenhouse gases Climate change ABSTRACT The introduction of high-temperature heat pumps for waste heat recovery with low GWP refrigerants can reduce the greenhouse gas emissions in the industrial sector. This article evaluates the energy performance and the volumetric heating capacity of five vapour compression system configurations using n-Pentane, Butane, HCFO-1233zd(E) and HFO-1336mzz(Z) as HFC-245fa low GWP alternative fluids for heating production at temperatures of 110, 130 and 150 C and different temperature lifts. The selected architectures and the equations are presented, and the most appropriate method to calculate the intermediate pressure is selected. The results of the simulation show that single-stage cycle with an internal heat exchanger (IHX) becomes the most efficient configuration at lower temperature lifts whereas two-stage cycle with IHX at higher lifts. While n-Pentane provides the highest energy performance values, Butane (only up to 130 C) and HCFO-1233zd(E) highlight in the heating volumetric capacities. HFO-1336mzz(Z) provides intermediate values in both parameters. Consequently, the working fluid selection is highly dependent on the specifications and the energetic and installation costs. 2018 Elsevier Ltd and IIR. All rights reserved. Evaluation theorique de differentes configurations de pompes a chaleur a haute temperature pour la recuperation de chaleur residuelle a basse temperature Mots-cles: Recuperation de chaleur perdue; Frigorigenes a faible GWP; Pompe a chaleur industrielle; Efficacite energetique; Gaza effet de serre; Changement climatique 1. Introduction The mitigation of the climate change is one of the greatest challenges that humanity faces in this century. To address it, greenhouse gas emissions (GHG), which is the main cause of the global warming, need to be controlled. Therefore, the efforts to improve the energy efficiency has been intensified, especially in the industrial sector as being one of the three main energy consuming sectors (Bruckner et al., 2015). Intergovernmental Panel on Climate Change considers the use of heat recovery technologies as a tool to improve the energy efficiency (Miro et al., 2015), decreasing the energy consumption of the industrial processes and therefore, the GHGs emissions. * Corresponding author. E-mail address: uji.es (C. Mateu-Royo). https://doi.org/10.1016/j.ijrefrig.2018.04.017 0140-7007/ 2018 Elsevier Ltd and IIR. All rights reserved. Many industrial processes reject a significant amount of waste heat below 100 C and hence, a substantial quantity of waste heat energy remain unused due to economic and technical barriers (Forman et al., 2016). However, a higher range of temperature is required by some industrial processes, especially between 100 and 130 C (Chua et al., 2010; Seck et al., 2015). Thus, the attempts to introduce high-temperature heat pumps (HTHPs) have attracted significant attention for waste heat recovery in industrial processes (Langan and O'Toole, 2017). In addition, the integration of HTHPs as a heat source for the industrial processes leads to a reduction of the fossil fuels dependence, promoting a decarbonised and sustainable industrial sector. At present, HFC-245fa, with a global warming potential (GWP) of 858, is widely used as a working fluid for HTHPs and Organic Rankine Cycles (Mounier et al., 2017). Nevertheless, the European Commission (2017) has adopted a proposal to ratify the Kigali 230 C. Mateu-Royo et al. / International Journal of Refrigeration 90 (2018) 229-237 Nomenclature Q thermal capacity (kW) V volumetric flow rate (m3 s-1) W electric power consumption (kW) m refrigerant mass flow rate (kg s-1) COP coefficient of performance h enthalpy (kJ kg-1) P pressure (MPa) q v volumetric heating capacity (kJ m-3) SC sub-cooling degree (K) SH superheat degree (K) T temperature (C) Greek symbols Efficiency (dimensionless) effectiveness (dimensionless) density (kg m-3) Subscripts c compressor em electromechanical HS high stage I intermediate in inlet is isentropic k condenser LS low stage o evaporator out outlet sec secondary circuit suc suction temp temperature vol volumetric Abbreviations CFC Chlorofluorocarbon GHG Greenhouse gas GWP Global warming potential HCFO Hydrochlorofluoroolefin HFC Hydrofluorocarbon HFO Hydrofluoroolefin HTHP High-temperature heat pump IHX Internal heat exchanger ODP Ozone depletion potential POE Polyolester amendment to the Montreal Protocol to gradually limit the HFCs production and consumption to mitigate the global warming. The use of renewable energy sources for heating becomes a possible solution for this challenge (Esen and Yuksel, 2013). Thus, the use of HTHPs for waste heat recovery using working fluids with low GWP becomes a solution to climate change mitigation by improving the industrial processes energy efficiency, using environmental friendly working fluids. A few studies realised a fundamental performance analysis of different low GWP working fluids for HTHPs (Fukuda et al., 2014; Kondou and Koyama, 2015). Moreover, vapour compression systems can use different configurations alternatives to basic cycle to improve the energy performance (Mols et al., 2014; Mota-Babiloni et al., 2014). HTHPs can also use these variations to increase the temperature of the heat source to a higher and more useful temperature. These heat sources can come from the ground (Esen et al., 2006) or from the combination of ground source with solar energy (Esen et al., 2017; Esen, 2000). The use of ground as a heat source requires a detailed study of the temperature distribution (Balbay and Esen, 2013) and especial attention in cold climates (Balbay and Esen, 2010). However, waste heat can used as a proper heat source for HTHP. Hence, Cao et al. (2014) studied different heat pump systems configurations for using waste heat recovery with an average temperature of 45 C as a heat source and heating production temperature up to 95 C. Similar system configurations than Cao et al. (2014) are analysed by Antonijevic et al. (2012) for groundwater HTHPs. Nevertheless, studies that analyse HTHP configurations with low GWP working fluids at different heating production temperatures are not found in the literature. Therefore, the aim of this article is to compare theoretically the energy performance and the volumetric heating capacity of five vapour compression system configurations using n-Pentane, Butane, HCFO-1233zd(E) and HFO-1336mzz(Z) as low GWP alternative fluids to HFC-245fa for heating production temperatures of 110 and 130 C. Moreover, this article evaluates the proper low GWP working fluid and system configuration for heating production temperature up to 150 C, which is a situation not reachable by the traditional refrigerant HFC-245fa. 2. Refrigerants under consideration Despite hydrofluorocarbons (HFCs) have been commonly used as refrigerants for HTHPs, the European Commission is gradually limiting the production and use of these chemicals. Hence, hydrofluoroolefins (HFO), hydrochlorofluoroolefin (HCFO) and natural refrigerants become potential alternatives. In addition, all the potential replacements must have comparable or higher critical temperature than HFC-245fa, as is the parameter that defines the possibility to achieve high heating production temperatures in a subcritical cycle. HCFO-1233zd(E) and HFO-1336mzz(Z) are synthetic fluids widely used for organic Rankine cycle applications (Mols et al., 2016; Navarro-Esbr et al., 2017) and, moreover, those refrigerants are being considered in HTHP applications (Juhasz, 2017). Both fluids are considered non-flammable and non-toxic. Butane (R-600) is an commonly known hydrocarbon limited by its critical temperature to HTHP applications with heating production temperatures below 130 C (Moisi and Rieberer, 2017; Pan et al., 2011), whereas n-Pentane (R-601), is only considered by Yamazaki and Kubo (1985) and Jakobs et al. (2010). However, the higher critical temperature of n-Pentane compared with the other refrigerants makes it an interesting candidate. Both hydrocarbons are flammable but non-toxic refrigerants. Table 1 provides a general overview of several relevant characteristics of these HTHP refrigerants. As it can be seen, the molar mass of the hydrocarbons is much lower than that of synthetic fluids (HFC, HFO and HCFO). From this initial comparison, the molecular weight difference can be highlighted, because of the influence of the density that can affect the HTHPs components design. 3. System configurations In this paper, five vapour-compression architectures are considered to provide an adapted solution to provide useful heat from different grades of industrial waste heat: single-stage cycle (SS), single-stage cycle with internal heat exchanger (S-S IHX), twostage cycle (T-S), two-stage cycle with IHX (T-S IHX) and two-stage cycle with intermediate-IHX (T-S I-IHX). The first system configuration presented is a single-stage cycle, Fig. 1(a). This basic configuration is composed of the following elements: compressor, condenser, expansion valve and evaporator. This cycle is used as a reference for the system configuration analysis. Then, adding a heat exchanger between the suction and the liquid line to this basic configuration the single-stage cycle C. Mateu-Royo et al. / International Journal of Refrigeration 90 (2018) 229-237 Table 1 Characteristics of refrigerant selected (Klein, 2006). ASHRAE Standards 34 (ASHRAE, 2017) HFC-245fa B1 HCFO-1233zd(E) A1 HFO-1336mzz(Z) A1 Butane A3 n-Pentane A3 Molar mass [g mol-1] 134.05 130.50 164.06 58.12 72.15 Pcrit [MPa] 3.65 3.77 2.90 3.80 3.37 Normal boiling point [ C] 15.18 18.32 33.47 -0.52 35.87 Tcrit [C] 154.01 165.60 171.27 151.98 196.54 231 ODP (ASHRAE, 2017) 0 0 0 0 0 100-yr GWP (ASHRAE, 2017) 858 1 2 20 20 Fig. 1. Schematic and P-h diagram of single-stages configurations: (a) single-stage cycle and (b) single-stage cycle with IHX. with IHX configuration is achieved Fig. 1(b). The IHX reduces the liquid temperature before the expansion valve (sub-cooling) by means of heating the suction vapour (superheat). In contrast to refrigeration systems (Mols et al., 2014), the interest of IHX for HTHP lies more in heating this suction gas to increase the gas discharge temperature instead of increasing the sub-cooling degree of the system. Higher discharge temperature increases the performance of heat pump applications, considering the components temperature limits. Two-stage cycles can be obtained by the division of the compression stage in two stages. The introduction of two-stage cycles, Fig. 2(a), reports an increase in the cycle performance through a reduction of the pressure ratio of compression stages. In refriger- ation applications, this system configuration has an inter-cooler or liquid injection between the compressors with the purpose of decrease the gas discharge temperature (Nemati et al., 2017). Nevertheless, in heating applications, it is interesting to increase the discharge temperature to enhance the useful heat transfer to the secondary fluid (Tk,sec,out). Therefore, two modifications of the initial two-stage configuration are proposed for the analysis of the HTHPs performance. IHX can be added between the evaporator and the low stage (LS) compressor and hence two-stage cycle with IHX configuration is obtained, Fig. 2(b). This configuration, similar to single-stage cycle with IHX, increases the superheat before the LS compressor at the expense of cooling the liquid before the expansion valve. It produces an increase of the gas discharge tempera- 232 C. Mateu-Royo et al. / International Journal of Refrigeration 90 (2018) 229-237 Fig. 2. Schematic and P-h diagram of two-stages configurations: (a) two-stage cycle, (b) two-stage cycle with IHX and (c) two-stage cycle with intermediate-IHX. ture of the LS compressor and consequently, an increase of the gas discharge pressure of the high stage (HS) compressor. Finally, the IHX could be located between the LS compressor and the HS compressor, giving rise to the last configuration two-stage cycle with intermediate-IHX. In contrast to single-stage cycle with IHX, this configuration increases the vapour suction temperature of the HS compressor instead of the LS compressor, as shown in Fig. 2(c). Hence, the gas discharge temperature of LS compressor remains similar to the two-stage cycle configuration, but the gas discharge temperature of the HS compressor increases. It can result interesting for the analysis because HS gas discharge temperature is a parameter relevant to the useful heat production. Nevertheless, a comprehensive analysis is required to find the proper system configuration and the appropriate working fluid. C. Mateu-Royo et al. / International Journal of Refrigeration 90 (2018) 229-237 233 4. Operating conditions and model description The heat absorbed by the evaporator (Q o) is an input of the model and it is considered constant for all the operating condi- tions. Considering the high-temperature applications of this per- formance analysis and the low-grade waste heat sources, two op- erating parameters are combined to simulate the typical working conditions: outlet condenser temperature of the secondary fluid or heating production temperature (Tk,sec,out) and inlet evapora- tor temperature of the secondary fluid or waste heat temperature (To,sec,in ). Firstly, heating production temperatures of 110 and 130 C are proposed to select an alternative working fluid and proper con- figuration to HFC-245fa in HTHP applications. Additionally, heat- ing production temperature of 150 C is considered due to the in- dustrial development interest of higher temperatures in heat pump applications. The sub-cooling degree is assumed to be 10 K and the temperature difference in the condenser of the secondary fluid be- tween the inlet (Tk,sec,in) and the outlet (Tk,sec,out) is considered to be 15 K. Secondly, inlet evaporator secondary fluid temperatures of 50, 70 and 90 C are considered in the simulations due to the interest of low-grade waste heat recovery. The superheat degree is consid- ered to be 15 K and the temperature difference in the evaporator of the secondary fluid between the inlet (To,sec,in) and the outlet (To,sec,out) is assumed to be 10 K. For both condenser and evapora- tor, the pitch point is assumed to be 5 K. The refrigerant thermodynamic properties have been evaluated using software Engineering Equation Solver (EES) (Klein, 2006). Moreover, isenthalpic process is considered at the expansion valve and heat transfer to the surroundings and pressure drops are ne- glected. Isentropic and volumetric efficiency of the compressors are cal- culated using Pierre's correlations for "good" reciprocating com- pressors (Granryd et al., 1999). In this way, volumetric efficiency, vol, is obtained from Eq. (1): t2k - 18 p1 vol = k1 1 + ks 100 exp k2 p2 (1) where t2k is the inlet temperature to the compressor and p1/p2 is the pressure ratio. The remaining symbols k1, ks and k2 are con- stants and the values of these are 1.04, 0.15, and -0.07, respec- tively. Following Pierre's correlations, isentropic efficiency, is, is cal- culated with Eq. (2) and the volumetric efficiency calculated with Eq. (1). vol t2k - 18 T1 is = 1 + ke 100 exp a T2 + b (2) where T1/T2 is the ratio of the absolute temperatures (in kelvin) of the condensation and evaporation corresponding to the discharge and the suction compressor pressures. The constants values con- sidered are those given by Pierre for CFC-12, and therefore, ke, a and b are -0.1, -2.40 and 2.88, respectively. This constant values are taken from the literature (Granryd et al., 1999). The electrome- chanical efficiency, em, is assumed to be 0.95. In two-stage system configurations, it is necessary to select the appropriate intermediate pressure to achieve competitive system efficiencies. Thus, the equations proposed by Baumann and Blass (1961) for theoretical cycles and assuming vapour as a perfect gas, Eq. (3), De Lepeleire (1973), Eq. (4), and Domanski (1995), Eq. (5), has been analysed for HTHP applications (Purohit et al., 2016). PI = PkPo (3) PI = PkPo + 0.35 [bar] (4) = TI - To 0.5 (5) Tk - To Fig. 3(a) shows that the resulting intermediate pressures are different depending on the configuration analysed and the equation used. On the other hand, the system performance remains almost constant in TS and TS IHX configurations whereas for the TS I-IHX configuration the equation proposed by Baumann and Blass (1961) achieves higher efficiencies than the others, as shown in Fig. 3(b). Consequently, Eq. (3) is selected as an equation to calculate the intermediate pressure in all two-stage system configurations. The refrigerant mass flow rate for all system configurations is calculated using Eq. (6). m = Q o (6) (ho,out - ho,in ) The required volumetric flow rate at the compressor suction, V suc, is obtained from the mass flow rate, the suction density and the volumetric efficiency, Eq. (7), V suc = m (7) vol suc The IHX effectiveness (IHX) is limited to the maximum dis- charge temperature Tdisc at 180 C. Degradation of HFO fluids and POE lubricants can be observed above 200 C (Kontomaris, 2012; Navarro-Esbr et al., 2015). The discharge temperature is dependent on the enthalpy after the compressor, hc, out, and the condenser pressure, PK. The vapour suction temperature Tsuc is calculated using Eq. (8). IHX = Tsuc - To,out (8) Tk,out - To,out The electric power consumption for the single-stage compressor, W C, is expressed in Eq. (9) as the product of the mass flow rate and the isentropic enthalpy increase at the compressor divided by the isentropic and electromechanical efficiencies. W C = m his,c (9) is em For the two-stage compressors, the electric power consumption is calculated using Eq. (10). W C = m his,c, HS + his,c,LS em is,HS is,LS (10) The heating capacity, Q k, is calculated using the product of the mass flow rate and the enthalpy difference between the inlet and outlet in the condenser, Eq. (11). Q k = m hk,in - hk,out (11) To properly compare the influence of the volumetric flow rate at the compressor suction and the heating capacity, volumetric heating capacity (qv) is calculate using Eq. (12). qv = Q k V suc (12) Finally, the Coefficient of Performance, COP, is calculated from the heating capacity and the compressor electric power consumption, using Eq. (13). COP = Q k (13) W C 234 C. Mateu-Royo et al. / International Journal of Refrigeration 90 (2018) 229-237 Fig. 3. Intermediate pressure and COP comparison analysis for different two-stage system configurations. 5. Results and discussion The main results of the simulations are obtained considering five configurations and five working fluids at different heating production temperature (Tk,sec,out) and different waste heat temperatures (To,sec,in). As exposed in Section 2, refrigerants HFO1336mzz(Z), HCFO-1233zd(E), Butane and n-Pentane have been selected as low GWP alternatives for HFC-245fa in high-temperature applications with Tk,sec,out below 130 C. Then, for applications with Tk,sec,out of 150 C, HFC-245fa and Butane have been discarded by the limitation of their critical temperature, and HFO-1336mzz(Z), HCFO-1233zd(E) and n-Pentane have been analysed. The parameters chosen in this study to evaluate the appropriate configuration and working fluid are the COP and the volumetric heating capacity. COP is commonly used to evaluate the energy performance of the system and the volumetric heating capacity is used to compare the relationship between the heating capacity and compressor and installation size. 5.1. Comparison of low GWP alternatives to HFC-245fa in moderate-high temperature applications For moderate and high-temperature applications, Tk,sec,out of 110 and 130 C, respectively, the results for alternative low GWP working fluids and configurations are shown as a rela- tive difference (%COP and %qv) between each pair refrigerant- configuration taken and HFC-245fa single-stage configuration, as indicated Eqs. (15) and (16). %COP = COPcon f,alt. f luid - COPSS,HFC-245 f a COPSS,HFC-245 f a (15) %qv = qvcon f,alt. f luid - qvSS,HFC-245 f a qvSS,HFC-245 f a (16) Fig. 4 presents the results for %COP at three inlet secondary fluid temperatures on the heat source, two outlet secondary fluid temperatures on the heat sink (110 and 130 C), five different configurations and five refrigerants. The difference between the evaporation and condensing temperature is known as temperature lift (Eq. (17)), which is commonly used to compare the heat pump system performance. Hence, all simulation results are grouped into four different temperature lifts in order to select and analyse the proper configuration depending on each parameter. Temperature li f t = Tcondensation - Tevaporation (17) On the one hand, for lift temperatures of 40 and 60 K, singlestage cycle IHX has higher performance (COP) than the other configurations. At lower temperature lifts, two-stage cycles have smaller pressure ratios than single-stage cycles, causing a decrease of the combined isentropic compressor efficiency. It increases the electric power consumption, and therefore, decrements the COP until the pressure ratio reaches higher values. On the other hand, for lift temperatures of 80 and 100 K, twostage cycle IHX has higher performance than the other configurations, as shown in Fig. 4. Two-stage cycle IHX achieves higher discharge temperature, heating capacity and isentropic compressor efficiency than other configurations. Therefore, higher the temperature lift, higher the performance improvement difference between two-stage cycle IHX and the other configurations. Besides, two-stage cycle IHX and two-stage cycle intermediate IHX have similar isentropic HS compressor efficiency, but the position of the IHX produces that two-stage cycle IHX has higher isentropic LS compressor efficiency than the other configuration. Thus, two-stage cycle IHX has lower overall electric power consumption and higher heating capacity, achieving the maximum COP. Attending to the COP results, single-stage with IHX is selected for analysis of the alternative working fluid in applications with temperature lifts of 40 and 60 K, whereas two-stage cycle IHX is selected for applications with temperature lifts of 80 and 100 K. Similar results are presented by Antonijevic et al. (2012) and Cao et al. (2014). These studies concluded that two-stage configuration with modifications is the proper system solution for high temperature lifts. Nevertheless, none study has been found for proper architecture of high-temperature heat pumps operating at low temperature lift. Fig. 5(a) presents the performance comparison of the different working fluids for a single-stage cycle with IHX configuration with temperature lifts of 40 and 60 K. All the alternative refrigerants perform better than HFC-245fa, given the positive differences. For the temperature lift of 40 K, higher %COP values are obtained by n-Pentane, with a performance increment of 7% compared to HFC-245fa single-stage cycle. Then, HFO-1336mzz(Z) and HCFO-1233zd(E) show an improvement of 5% and 4%, respectively. As the temperature lift increases, the %COP increases in the same way and for the temperature lift of 60 K, n-Pentane obtain a COP increase of 13-15% whereas HCFO-1233zd(E) achieves an improvement of 9-11%, being the HFO-1336mzz(Z) lower than those. The lower critical temperature of Butane causes that at higher evaporation temperatures it has worse improvements than the other refrigerants. Attending to the qv, Fig. 5(b) shows a reduction of 45% and 35% when n-Pentane and HFO-1336mzz(Z) are used instead of HFC245fa, whereas the reduction of the volumetric heating capacity using HCFO-1233zd(E) is between 5% and 15%. Nevertheless, Butane is the unique refrigerant analysed that increases the volumetric heating capacity between 15-30%, therefore, lower compressor displacement and installations size is required than using HFC245fa. C. Mateu-Royo et al. / International Journal of Refrigeration 90 (2018) 229-237 235 Fig. 4. Simulations results of %COP compared with single-stage cycle using HFC-245fa for different temperature lifts. Fig. 5. Simulation relative results of COP and volumetric heating capacity for temperature lifts of 40 and 60 K. Fig. 6. Simulation relative results of COP and volumetric heating capacity for temperature lifts of 80 and 100 K. Fig. 6(a) illustrates the COP variation for temperature lift of 80 and 100 K in two-stage cycle with IHX. A higher temperature lift will increase the COP difference compared with the reference system. The %COP for all the analysed refrigerants is at a similar level but again, the highest COP is obtained by n-Pentane, being 23-35% than that of HFC-245fa single-stage cycle. However, Fig. 6(b) shows a great variation of the volumetric heating capacity between the different working fluids simulated. While n-Pentane and HFO-1336mzz(Z) has a reduction of the volumetric heating capacity, Butane and HCFO-1233zd(E) presents an increment. Concretely, 50-70% for Butane and 2-15% for HCFO1233zd(E). As a result, with similar variations of COP and providing the same heating capacity, HCFO-1233zd(E) requires a slightly similar compressor and installation size whereas Butane requires a lower compressor and installation size comparing with HFC-245fa. It can be concluded that HCFO-1233zd(E) presents promising results in all the simulations carried out, achieving remarkable increments of COP with a slight variation of the volumetric heating capacity compared to the reference HFC-245fa (positive for higher temperature lifts and negative for lower). Although the highest increment of COP is achieved by n-Pentane, this working fluid obtained the most significant reduction of the volumetric heating capacity in all the cases and would require a significantly higher investment cost than HFC-245fa (in addition to the already needed because of the security measures). Butane has the largest increment of volumetric heating capacity, reducing the installation cost and proving higher heating capacity than HFC-245fa. Moreover, Butane has an increment of COP in all the simulations realised, nonetheless, when the heat source temperature is bringing closer to its critical temperature, this increment of COP is reduced, as it 236 C. Mateu-Royo et al. / International Journal of Refrigeration 90 (2018) 229-237 1233zd(E). This means that higher displacement compressor and installation size are needed for n-Pentane and HFO-1336mzz(Z) than for HCFO-1233zd(E). 6. Conclusions Fig. 7. Two-stage cycle with IHX simulations results of COP for heating production temperature of 150 C. Fig. 8. Two-stage cycle with IHX simulations results of volumetric heating capacity for heating production. can be seen in Fig. 5(a). Hence, Butane resulted in a proper solution if the heat source temperature does not come close to its critical temperature. Finally, HFO-1336mzz(Z) has a general increment of COP, most of the cases lower than the other fluids, and in addition, it has a highly decrement of the volumetric heating capacity. Thus, HFO-1336mzz(Z) is not a proper solution when the production temperature is below 130 C. 5.2. Analysis for high-temperature application solutions This section presents the results for the heating production temperature of 150 C and heat source temperatures of 50, 70 and 90 C. HFC-245fa and Butane have a critical temperature below that value, therefore, these refrigerants are not considered for these conditions. Therefore, the results cannot be compared with HFC-245fa and are shown as HFO-1336mzz(Z), HCFO-1233zd(E) and n-Pentane absolute values of COP and qv, Figs. 7 and F8. With a heat source temperature of 90 C, Fig. 7 shows that nPentane and HFO-1336mzz(Z) obtain a COP of 3.85 and 3.59, respectively. Nevertheless, Fig. 8 illustrates the volumetric heating capacity of these simulations, where HFO-1336mzz(Z) has slightly higher volumetric heating capacity than n-Pentane. Note that nPentane requires security requirements because of its flammability and slight higher investment cost compared with the other two candidates due to its low volumetric heating capacity. For heat source temperature of 70 C, Fig. 7 shows that, like in the previous application, n-Pentane and HFO-1336mzz(Z) are the potential candidates with a COP of 2.98 and 2.74, respectively. Nevertheless, HCFO-1233zd(E) achieves a COP of 2.50 with higher volumetric heating capacity than the other two candidates, as shown in Fig. 8. For heat source temperature of 50 C, similar behaviour than that for heat source temperature of 70 C is observed. Similar results are obtained by Juhasz (2017), who compares the energy performance of HFO-1336mzz(Z) with other low GWP, obtaining better COP for HFO-1336mzz(Z) than for HCFO-1233zd(E). In general, higher COPs are achieved by n-Pentane in all the options simulated, followed by HFO-1336mzz(Z). Nevertheless, both refrigerants obtain lower volumetric heating capacity than HCFO- Due to the concern for the effects of greenhouse gasses emissions on the global environment there is large interest in Europe and elsewhere for the use of high-temperature heat pumps for waste heat recovery with low GWP refrigerants. This paper has investigated the energy performance of five different system configurations for high-temperature heat pumps using low GWP alternatives to HFC-245fa. n-Pentane, Butane, HFO-1336mzz(Z) and HCFO-1233zd(E) are the proposed replacements because of the high critical temperatures and promising thermodynamic and transport properties. The proposed configurations are single-stage and two-stage cycles with and without internal heat exchanger in different positions. The optimal intermediate pressure has been calculated for each situation. Heating production and waste heat source temperatures are highly sensitive to the configuration and working fluid selection. Therefore, temperature lift is a key parameter to analyse the proper configuration. For temperature lift of 40 K, single-stage with IHX presents higher performance than the other configuration, but as temperature lift increases, two-stage with IHX becomes the proper system configuration. For moderate-high temperature applications with lowtemperature lift, single-stage cycle with IHX using n-Pentane achieves a COP improvement up to 15% compared with a singlestage cycle with HFC-245fa. However, n-Pentane has a reduction of 45% of the volumetric heating capacity whereas HCFO-1233zd(E) keeps this reduction up to 5% with a COP improvement up to 11%. For higher temperature lifts, the configuration with highest energy performance is a two-stage cycle with IHX. The COP variation is slightly different for each refrigerant; therefore, volumetric heating capacity becomes the significative parameter. Butane and HCFO-1233zd(E) obtain an increment of the volumetric heating capacity of 70% and 15%, respectively, whereas n-Pentane and HFO-1336mzz(Z) decrease this parameter in 40% and 35%. For high-temperature applications with production heating temperature up to 150 C, two-stage cycle with IHX using nPentane achieves the highest COP, followed by HFO-1336mzz(Z) (up to 3.85 and 3.59, respectively). Otherwise, n-Pentane and HFO1336mzz(Z) obtain up to 50% lower volumetric heating capacity than HCFO-1233zd(E), which has a COP of 3.42. Hence, depends on the energy or the installation cost, most proper working fluid would be selected. In conclusion, it seems likely that single-stage and two-stage cycles with IHX will be used as system configurations in hightemperature heat pumps. While the system configuration candidates are clear, working fluid selection is highly dependent on the energy and installation costs aside from the regulation and safety measures of flammable fluid in the installations. Acknowledgements The authors acknowledge the Spanish Government for the financial support under projects ENE2015-70610-R and RTC-20154193-3. Furthermore, the authors acknowledge the Universitat Jaume I (Castelln de la Plana, Spain) for the financial support under grant POSDOC/2016/23 and project P1-1B2015-38. References Antonijevic , D.L., Manic , D.J., Komatina, M.S., Rudonja, N.R., 2012. Groundwater heat pump selection for high temperature heating retrofit. Energy Build 49, 294-299. https://doi.org/10.1016/j.enbuild.2012.02.028 . C. Mateu-Royo et al. / International Journal of Refrigeration 90 (2018) 229-237 237 ASHRAE, 2017. ASHRAE Handbook Fundamentals. American Society of Heating, Refrigerating and Air-Conditioning Engineers. Balbay, A., Esen, M., 2013. Temperature distributions in pavement and bridge slabs heated by using vertical ground-source heat pump systems. Acta Sci. Technol. 35, 677-685. Balbay, A., Esen, M., 2010. Experimental investigation of using ground source heat pump system for snow melting on pavements and bridge decks. Sci. Res. Essays 5, 3955-3966. Baumann, K., Blass, E., 1961. Beitrag zur Ermittlung des Optimalen Mitteldruckes bei zweistufigen Kaltdampf Verdichter-Kaltemaschinen. Ktetechnik 13, 210-216. Brckner, S., Liu, S., Mir, L., Radspieler, M., Cabeza, L.F., Lvemann, E., 2015. Industrial waste heat recovery technologies: an economic analysis of heat transformation technologies. Appl. Energy 151, 157-167. https://doi.org/10.1016/j. apenergy.2015.01.147 . Cao, X.-Q., Yang, W.-W., Zhou, F., He, Y.-L., 2014. Performance analysis of different high-temperature heat pump systems for low-grade waste heat recovery. Appl. Therm. Eng. 71, 291-300. https://doi.org/10.1016/j.applthermaleng.2014.06.049. Chua, K.J., Chou, S.K., Yang, W.M., 2010. Advances in heat pump systems: a review. Appl. Energy 87, 3611-3624. https://doi.org/10.1016/j.apenergy.2010.06.014. De Lepeleire, G., 1973. Une nouvelle faon d'apprciation et de slection des compresseurs frigorifiques bitags. In: XIII International Congress of Refrigeration Washington, pp. 39-48. Domanski P.A., Theoretical evaluation of the vapor compression cycle with a liquidline/suction-line heat exchanger, economizer, and ejector, NIST Interagency Report 5606, National Institute of Standards and Technology, Gaithersburg, MD, 1995 https://www.nist.gov/publications/theoretical- evaluation- vaporcompression- cycle- liquid- linesuction- line- heat- exchanger. Esen, H., Esen, M., Ozsolak, O., 2017. Modelling and experimental performance analysis of solar-assisted ground source heat pump system. J. Exp. Theor. Artif. Intell. 29, 1-17. Esen, H., Inalli, M., Esen, M., 2006. Technoeconomic appraisal of a ground source heat pump system for a heating season in eastern Turkey. Energy Convers. Manag. 47, 1281-1297. Esen, M., 2000. Thermal performance of a solar-aided latent heat store used for space heating by heat pump. Sol. Energy 69, 15-25. Esen, M., Yuksel, T., 2013. Experimental evaluation of using various renewable energy sources for heating a greenhouse. Energy Build 65, 340-351. European Commission, 2017. Proposal for a council decision on the conclusion of the agreement to amend the montreal protocol on substances that deplete the ozone layer adopted in Kigali. Brussels, Belgium. https://eur-lex.europa. eu/resource.html?uri=cellar:c4917fa1- e93f- 11e6- ad7c- 01aa75ed71a1.0002.02/ DOC_1&format=PDF. 5 pages. Forman, C., Muritala, I.K., Pardemann, R., Meyer, B., 2016. Estimating the global waste heat potential. Renew. Sustain. Energy Rev. 57, 1568-1579. https://doi.org/ 10.1016/j.rser.2015.12.192 . Fukuda, S., Kondou, C., Takata, N., Koyama, S., 2014. Low GWP refrigerants R1234ze(E) and R1234ze(Z) for high temperature heat pumps. Int. J. Refrig. 40, 161-173. https://doi.org/10.1016/j.ijrefrig.2013.10.014. Granryd, E., Ekroth, I., Lundqvist, P., Melinder, A., Palm, B., Rohlin, P., 1999. Refrigeration Engineering. KTH. Jakobs, R., Cibis, D., Laue, H.., 2010. Status and outlook: industrial heat pumps. In: Proceedings of the International Refrigeration and Air Conditioning Conference. Juhasz, J.R., 2017. Novel working fluid, HFO-1336mzz (E), for use in waste heat recovery application. In: Proceedings of the Twelfth IEA Heat Pump Conference. 2017. Rotterdam. Klein, S., 2006. Engineering Equation Solver (EES) V10.2. Fchart software, Madison, USA www.fchart.com. Kondou, C., Koyama, S., 2015. Thermodynamic assessment of high-temperature heat pumps using low-GWP HFO refrigerants for heat recovery. Int. J. Refrig. 53, 126- 141. https://doi.org/10.1016/j.ijrefrig.2014.09.018. Kontomaris, K., 2012. A zero-ODP, low GWP working fluid for high temperature heating and power generation from low temperature heat: DR-2. In: Proceedings of the JRAIA International Symposium 2012, pp. 212-216. Langan, M., O'Toole, K., 2017. A new technology for cost effective low grade waste heat recovery. Energy Procedia 123, 188-195. https://doi.org/10.1016/j.egypro. 2017.07.261 . Mir, L., Brckner, S., Cabeza, L.F., 2015. Mapping and discussing Industrial Waste Heat (IWH) potentials for different countries. Renew. Sustain. Energy Rev. 51, 847-855. https://doi.org/10.1016/j.rser.2015.06.035. Moisi, H., Rieberer, R., 2017. Refrigerant selection and cycle development for a high temperature vapor compression heat pump. In: Proceedings of the Twelfth IEA Heat Pump Conference. 2017. Rotterdam, pp. 1-10. Mols, F., Navarro-Esbr, J., Peris, B., Mota-Babiloni, A., 2016. Experimental evaluation of HCFO-1233zd-E as HFC-245fa replacement in an organic rankine cycle system for low temperature heat sources. Appl. Therm. Eng. 98, 954-961. https://doi. org/10.1016/j.applthermaleng.2016.01.011 . Mols, F., Navarro-Esbr, J., Peris, B., Mota-Babiloni, A., Barragn-Cervera, Angel, 2014. Theoretical energy performance evaluation of different single stage vapour compression refrigeration configurations using R1234yf and R1234ze(E) as working fluids. Int. J. Refrig. 44, 141-150. https://doi.org/10.1016/j.ijrefrig.2014. 04.025. Mota-Babiloni, A., Navarro-Esbr, J., Barragn, ., Mols, F., Peris, B., 2014. Theoretical comparison of low GWP alternatives for different refrigeration configurations taking R404A as baseline. Int. J. Refrig. 44, 81-90. https://doi.org/10.1016/ j.ijrefrig.2014.04.015. Mounier, V., Mendoza, L.C., Schiffmann, J., 2017. Thermo-economic optimization of an ORC driven heat pump based on small scale turbomachinery and comparison with absorption heat pumps. Int. J. Refrig.. https://doi.org/10.1016/j.ijrefrig.2017. 05.021 . Navarro-Esbr, J., Mols, F., Peris, B., Mota-Babiloni, A., 2015. Small scale orc design for a cogeneration solar biomass suported application. In: Proceedings of the Third International Seminar on ORC Power Systems. Brussels, Belgium. Oct. 12-14 Navarro-Esbr, J., Mols, F., Peris, B., Mota-Babiloni, A., Kontomaris, K., 2017. Experimental study of an Organic Rankine Cycle with HFO-1336mzz-Z as a low global warming potential working fluid for micro-scale low temperature applications. Energy 133, 79-89. https://doi.org/10.1016/j.energy.2017.05.092. Nemati, A., Nami, H., Yari, M., 2017. Comparaison des frigorignes dans un cycle frigorifique transcritique jecteur-dtente bi-tag, bas sur une analyse environnementale et exergo-conomique. Int. J. Refrig. 84, 139-150. https://doi.org/ 10.1016/j.ijrefrig.2017.09.002. Pan, L., Wang, H., Chen, Q., Chen, C., 2011. Theoretical and experimental study on several refrigerants of moderately high temperature heat pump. Appl. Therm. Eng. 31, 1886-1893. https://doi.org/10.1016/j.applthermaleng.2011.02.035. Purohit, N., Gupta, D.K., Dasgupta, M.S., 2016. Effect of inter-stage pressure on the performance of a two stage refrigeration cycle using inter cooler. Energy Procedia 90, 171-178. https://doi.org/10.1016/j.egypro.2016.11.182. Seck, G.S., Guerassimoff, G., Mazi, N., 2015. Heat recovery using heat pumps in non-energy intensive industry: are energy saving certificates a solution for the food and drink industry in France? Appl. Energy 156, 374-389. https://doi.org/ 10.1016/j.apenergy.2015.07.048. Yamazaki, T., Kubo, Y., 1985. Development of a high-temperature heat pump. IEA Heat Pump Cent. Newsl. 3, 18-21.