Analysis of suitability ranges of high temperature heat pump working fluids

Accepted Manuscript Analysis of suitability ranges of high temperature heat pump working fluids Guido Francesco Frate, Lorenzo Ferrari, Umberto Desideri PII: DOI: Reference:

S1359-4311(18)35360-2 https://doi.org/10.1016/j.applthermaleng.2019.01.034 ATE 13214

To appear in:

Applied Thermal Engineering

Received Date: Revised Date: Accepted Date:

30 August 2018 10 January 2019 11 January 2019

Please cite this article as: G. Francesco Frate, L. Ferrari, U. Desideri, Analysis of suitability ranges of high temperature heat pump working fluids, Applied Thermal Engineering (2019), doi: https://doi.org/10.1016/ j.applthermaleng.2019.01.034

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Analysis of suitability ranges of high temperature heat pump working fluids Guido Francesco Fratea, Lorenzo Ferrarib and Umberto Desideric a

DESTEC – University of Pisa, Pisa, Italy, [email protected] a DESTEC – University of Pisa, Pisa, Italy, [email protected] a DESTEC – University of Pisa, Pisa, Italy, [email protected]

Abstract: Power generation systems recovering waste heat, such as Organic Rankine Cycles (ORC), cannot work efficiently when the highest temperature ranges between 100 °C and 85 °C. For this reason, upgrading and recycling the low-grade waste heat is very interesting in this temperature range. High temperature heat pumps can upgrade the waste heat raising the temperature up to 150 °C, which is adequate for some productive process in the food, paper, chemical and tobacco industries. Despite the intense research effort, even focusing solely on the vapour compression heat pumps, selecting the most suitable working fluid for these applications is not trivial. The selection of a fluid for a high temperature heat pump is limited by several technical constraints and a trade-of between a high efficiency and a large volume flow rate is necessary. Since the recommended fluid may change for different combinations of heat source and heat sink temperatures, a systematic investigation of the heat pump performance with different fluids working at different operational conditions has been carried out. The effect of different technical limitations has been discussed, and the most suited fluids both with respect to the highest efficiency and optimal volume flow rate have been recommended.

Keywords: High temperature heat pump; Waste heat upgrading; Working fluids; COP and VHC trade off; HTHP Compressors.

Nomenclature: Acronyms:  COP  CS  GWP  HP  IWH  NFPA  ODP  ORC  SS  VCHP  VHC  WHR

coefficient of performance carbon steel global warming potential heat pump industrial waste heat national fire protection agency ozone depletion potential organic Rankine cycle stainless steel vapour compression heat pump volumetric heating capacity waste heat recovery

superscripts and subscripts:  auto Autoignition  comp Compressor / Compression  crit Critical  dec Decomposition  in Inlet 1

        

is max min out pol sat sink source tot

Symbols:  a  h  NS  p  Q  T  u  V  W

Isentropic Maximum Minimum Outlet Polytropic Saturation Heat Sink Heat source Total

Acoustic velocity Enthalpy Number of stages Pressure Heat flow rate Temperature blade tip speed Volumetric flow rate electrical power

Greek letters:  η Efficiency  μ load coefficient

1. Introduction Energy saving is a well-established innovation paradigm in energy-intensive industrial activities. This is because the increase in energy utilization efficiency is beneficial for both users and utilities. The first benefit, for example, from a reduction of costs and CO2 emission, while the latter receive a direct economic benefit. In industry, one of the main pathways to save energy is by recovering the waste heat (WHR), which has received a significant attention in the last years thanks to the development of an efficient and simple WHR technology, i.e. the Organic Rankine Cycles (ORC). ORCs are suitable to recover the heat that would otherwise be rejected into the environment, and wasted, roughly between 100 °C and 400 °C. Several ORC manufacturers propose these temperatures as the applicability limits of their products [1]. In theory, the ORC technology can also be applied without any significant modification to recover heat at lower temperature, and several cases are described in the literature [2–4]. However, ORC applications, which use waste heat at a temperature lower than 85 °C, are rare. At that temperature level the ORC efficiency is low, and the profitability of a system must be carefully investigated case by case. This means that the waste heat at temperatures lower than 90 °C could be not suitable for the conversion into electric power. Despite this, such resource should not be disregarded, also in consideration of its high availability. As a matter of fact, on a world-wide basis the Industrial Waste Heat (IWH) is estimated to be available for 40% at temperatures lower than 100 °C [5]. An interesting perspective is to recover this low grade IWH by recycling it into the industrial process that produced it. This practice is often called waste heat upgrading, or recycling, and could be of an interest since the heat demand of many industrial processes is often at medium-low temperature. Entire industrial sectors, such as the tobacco and food industries, require 100% of their 2

heat input below 400 °C, and 50% below 100 °C. Furthermore, between 50% and 75% of the heat required by paper and chemicals industries is below 400 °C, while 25% is below 100 °C [6]. Cross-checking the low-grade IWH range of availability, the industries heat demand and the technological limits of the heat-upgrading technologies, an operative frame between 50 °C and 150 °C can be defined for the heat recycling. To upgrade the waste heat, some Heat Pump (HP) technologies has been proposed and investigated. Essentially, they can be divided into two main groups: Vapour-Compression HP (VCHP) and the Absorption HP (AHP). This last group is essentially made up of: classic absorption HP, Hybrid Absorption-Compression HP (HACHP) [7–9] and Absorption Heat Transformer (AHT) [10–13]. As far as it concerns the VCHP, the research focused both on the cycle architecture and on the working fluid selection. Many advanced cycle configuration can be found in literature, with different degree of complexity: simple, but effective, cycle modifications (internal heat exchanger, inter-cooled compression and vapor injection compression) can be found in [14–16]; cascaded heat pumps have been analysed in [17,18], and the integration of different heat sources at different temperature level is discussed in [19]. A combination of all these techniques can be found in [20], where the HP cycle architecture is optimized as a function of the available heat sources and their respective temperature level. Despite the efficiency improvement brought by these modifications, simple architectures are often preferred, as a compromise between the low cost and the improved performances. From this point of view, the regenerated architecture is of a particular interest and is extensively investigated in [15]. As far as working fluids are concerned, a systematic comparison was carried out by Bertinat [21]. Such analysis was based on some parameters, like the ideal COP and the Volumetric Heating Capacity (VHC) in MJ/m3, which have been extensively used also in more recent works, such as in [15] where a systematic comparison of promising working fluids (including R1233zd(E), R1234ze(Z), R600, R601, R1336mzz(Z) and R1224yd(Z)) has been carried out considering the trade-off between those two parameters, for a fixed HP temperature lift of 70 K. The current research trend is to focus on smaller fluid samples, and they generally consider also the environmental compatibility of the selected working fluids [22]. Due to the application temperature level, the pool of investigated fluids is often selected among the fluids considered suitable for Organic Rankine Cycle (ORC) applications [23–26]. Mixtures of fluids have also been investigated and compared with the pure working fluids [25,27]. Some authors, which are more concerned with the environmental issues related to the use of synthetic refrigerants, limit their analysis to a subset of environmentally friendly fluids, such as ultra-low GWP synthetic fluids, namely R1234ze(Z) and R1234ze(E) [28,29], or hydrocarbons, ammonia and water [30–32]. Using the water as refrigerant, a significant effort is dedicated to the analysis of configurations which are different from the classical heat pump. Direct steam compression, re-compression and flash evaporation can be found in [33] and [34]. Hybrid solutions with WHR combined with gas-fired boiler, electric heater, classical HP and HTHP which recover waste heat are analysed and compared from the environmental and techno-economic point of view in [35]. Finally, a review of the uses of water as a refrigerant is performed in [36]. Apart from the cases that investigate the performances of specific fluids, the choice of limiting the comparative analysis to a restricted pool of fluids can be justified in several ways. For example, by using well-known, or previously-proposed, fluid, the compatibility with the materials used for the thermal equipment and seals is guaranteed. In other cases, the analysis was restricted to a sample of fluids for which the VCHP compressor is commercially available [7,32]. In other words, in most cases an extensive comparison among the working fluids for a given operative temperature range is rightfully not the point of the research. Anyway, this has led to systematically neglect some fluids, which, as a matter of fact, have suitable thermophysical properties, the required environmental sustainability and that are compatible with the common materials used in the VCHP equipment. In order to extend the work of Bertinat [21], the size of the fluid samples and the monitored quantities have changed. Due to the evolution of computational capacities and of the thermophysical properties databases, the VCHP cycle can be simulated directly for several fluids, 3

whose performance can now be compared directly, instead of being extrapolated, e.g. from the critical temperature of the fluid. Instead of simulating an ideal VCHP cycle, valid only for the fluidcomparison purpose, it is possible to impose realistic features (temperature differences, compressor efficiencies, etc…) which allow to estimate the VCHP performance. In this way, several technical constraints can be evaluated (compressor discharge temperature, evaporation pressure, etc…), as it was done in [7,32]. In addition to the analysis of a restricted, although valid, pool of working fluids, another common feature is to fix a limited number of temperature raises (70 K and 50 K are common values [15]) and to test different fluids and cycle architectures with these assumptions. While this approach is valid for comparing different solutions, the general picture might get lost, as different fluids might be especially suited for some ranges of applications, while other, more efficient for some applications, might fall out of the technically viable operational framework, for other configurations. The present work aims to fill such gaps, by conducting a screening of the working fluids suitable for the VCHP for several combinations of heat source and heat sink temperatures. The performance of several fluids has been systematically compared in the temperature range of interest, i.e. from 50 °C to 150 °C. Furthermore, the environmental sustainability, the materials compatibility and several practical and techno-economic limitations of the equipment (particularly about the compressor technology) have been considered and discussed for the most promising fluids.

2. Preliminary fluid choice The thermo-physical properties of the investigated fluids have been calculated by using the CoolProp database [37]. Such database has been selected because it is free and opensource and it is validated against its well-known proprietary counter-part REFPROP. In addition to this, the amount of pure fluids available in the two software tools is very similar. Due to its non-commercial nature, the pool of fluids available in CoolProp is updated more rarely than that of REFPROP, so to investigate the two recently developed refrigerants R1336mzz(Z) and R1224yd(Z) the use of the last is mandatory. Therefore, for the calculation of the thermodynamic properties of these two working fluids, REFPROP, interfaced through CoolProp, which is one of the REFPROP official interfaces for Python and Matlab, has been used. To restrict the analysis to a realistic pool of working fluids, from the total amount of available fluids, i.e. 122, those which have a critical temperature Tcrit lower than 125 °C and a saturation pressure at 40 °C psat,40 °C lower than 0.05 bar have been discarded. The first limitation guarantees that the selected fluids can work at least in a part of the operational range far from the critical conditions. The second condition guarantees the exclusion of those fluids which would have shown an excessively low pressure at the HP evaporator. In addition to the two thermo-dynamic limitations, two additional environmental constraints have been enforced. Since CoolProp provides also the GWP100 and ODP of many fluids, all those with ODP > 0 and GWP100 > 103 have been discarded. The choice for the GWP100 threshold value is arbitrary, since there is not any specific regulation on the investigated application. Despite this, the current European regulation limits to 2500 the GWP100 for all kinds of applications, except for those which use recovered gases [38]. Therefore, 103 is chosen to limit the analysis to working fluids with a very low GWP, also in reference to the current regulations. After such limitations have been enforced, the total amount of available fluids drops to 35, excluding also some often proposed fluids for high temperature VCHP applications, such as the R245fa which has a GWP equal to 1030 [39]. Among those fluids, ten are not suitable for being used as VCHP working fluids, as explained below. Following the U.S. National Fire Protection Agency (NFPA) standard on the hazards of chemical substances (NFPA 704 [40]), three grades ranging between 0 and 4 have been assigned to each fluid, to quantify the flammability, the health-hazard and the instability of the compound. The instability grade fulfils an important role in this case, since it measures the chemical stability of the substance under high-temperature and high-pressure conditions, which are characteristic of the 4

VCHP applications. Among the 35 fluids which have been selected so far, six have an instability NFPA grade equal or higher than one and thus are discarded. The most unstable fluids of this group are the Propyne and the Ethylene Oxide which have a grade equal to three and cannot be heated or compressed without risks. Data concerning NFPA grades are taken from the two free online chemistry databases reported in [41,42]. A different approach must be followed for the flammability. As a matter of fact, the majority of the 35 initially selected fluids have a flammability grade higher or equal to 3 (the average is around 2.8). Differently from the chemical instability, which may pose uncontrollable and unexpected risks, the flammability-related issues can be contained with specific and well-known technical solutions [43]. This is demonstrated by the common use of isobutane (NFPA flammability grade 4) as working fluid for the household refrigerators. In any case, the flammability of the working fluid can become an issue due to potential leakages combined with the high temperature reached during the HP operation, which can potentially lead to the autoignition of the fluid. This is true especially for those fluids like ethanol and methanol, which can easily reach much higher temperatures than the maximum temperature at which the heat is upgraded (150 °C in this study). This is due to the characteristic shape of the fluid saturation curve, which may lead to a high degree of superheating, being equal the required operational conditions. To consider this potential source of risk, all the fluids, which have an autoignition temperature lower than 250 °C, were not considered in this study. The data about the autoignition temperature are found in [41,42]. Four fluids, among the initial 35, are discarded due to the limitation on the auto-ignition temperature. Among these there are: n-exane, Iso-Hexane, Cyclo-Hexane, DiethylEther and the n-pentane, which have been recently proposed and used as ORC working fluid [44]. The last parameter used in the NFPA classification is the health hazard. In similarity with the flammability index, the industrial experience teaches that it is possible to contain the risks originated by the use of toxic fluids, if there is an interest in using those particular compounds. This is, for example, the case of the Ammonia, which has a health-hazard grade of 3, and it is commonly used in the industrial or large-retail refrigeration systems. Since none of the initial 35 fluids has a health-hazard grade higher than the Ammonia one, for what concerns the toxicity they are all assumed to be suitable for the use in VCHPs. The case of the Ammonia is significant also for what concerns the material compatibility of the selected potential VCHP working fluids. The ammonia is not compatible with copper and brass thus forcing the use of steel for thermal equipment. Similar material compatibility constraints are also shown by several other potential working fluids. After the application of the limitations over the auto-ignition temperature and the chemical instability discussed before, 13 fluids out the initial 35 were ruled out. The remaining 27 fluids are all suitable for the use in conjunction with stainless steel thermal equipment and at least some of the materials commonly used for the seals. Furthermore, several investigated fluids do not specifically require the use of stainless steel; in this case, carbon steel equipment can be used to reduce the costs. The thermal equipment material does not play any role in the thermodynamic assessment of the VCHP performances, but it is important to calculate the system capital cost, and thus the profitability of the VCHP application. Data about materials compatibility can be easily found in several equipment vendors technical reports [45–47]. The last consideration is about the thermal degradation of the selected working fluids. While the instability grade of the NFPA 704 standard measures the potential of the fluid to decompose destructively at high-temperature and pressure, there is also the possibility that the fluid could decompose without incurring in dangerous reactions. This process alters the chemical composition of the working fluid, changing its saturation temperature and pressure, thus potentially worsening the VCHP performances. To the best of the author knowledge, this effect has been investigated only for those fluids which have been previously proposed for ORC [44,48–53], thus this information is incomplete, except for those rare cases in which the decomposition temperature is indicated in some vendor compound datasheets. Obviously, for well-known fluids such as ethanol and methanol the thermal decomposition characteristics are proven, although they are focused on the industrial 5

applications of such process with temperature ranges which are much higher than those analysed in the paper. Finally, the 27 working fluids which satisfy all the criteria introduced above, and were thus analysed further in this paper, are listed in Table 1.

3. Fluid comparison criterium Any VCHP operational condition is characterised by a COP value. Such parameter follows the classical definition (Eq. 1): COP 

Qsink  Wcomp

(1)

Table 1. Technical and physical properties of the selected working fluids. CS stands for carbon steel, while SS for stainless steel, and AHSRAE SC identifies the ASHRAE safety classification standard. Fluid

Tcrit [ °C]

psat,40 °C

Flammabil itya

ASHR AE SC

CS compatibili tyb

Tauto

Tdec,min

100

Healt ha

[ °C]

[ °C]

<1 <1 N/A

1 3 2

3 1 3

N/A B2L N/A

  

527 630 555

GWP

Acetone Ammonia Benzene

234.95 132.25 288.87

[bar] 0.57 15.55 0.24

Cyclo-Pentane

238.57

0.74

<6

1

3

N/A



320

Cyclo-Propane Cyclo-Hexane Dichloro-Ethane1 (DCE) Dimethyl-Carbonate (DMC) Ethanol Iso-Butane Iso-Hexane

125.15 280.45

10.64 0.25

N/A <6

1 1

4 3

N/A N/A

 

495 260

N/A N/A 315c 275e,f – 325d N/A N/A

288.45

0.21

<1

2

3

N/A

 - (SS)

440

300i

283.85

0.15

N/A

3

3

N/A

 - (SS)

458

N/A

241.56 134.67 224.55

0.18 5.31 0.51

1 3 <6

2 0 1

3 4 3

N/A A3 N/A

  

400 460 265

Iso-Pentane

187.20

1.52

2÷6

1

4

A3

 - (SS)

420

Methanol MM (Hexamethyldisiloxane)

239.35

0.35

2.8

1

3

N/A



440

N/A 300f N/A 275f – 290e N/A

245.60

0.11

N/A

1

3

N/A

N/A

310

300g

n-Butane

151.98

3.78

4

1

4

A3



405

n-pentane Neopentane Novec649 R1224yd(Z) R1233zd(E) R1234ze(Z) R13336mzz(Z) R245ca

196.55 160.59 168.66 155.5 166.45 150.12 171.3 174.42

1.16 2.70 0.73 2.45 2.16 2.90 1.28 1.73

5 N/A <1 <1 1 <1 2 726

1 2 N/A N/A N/A 1 N/A 1

4 4 0 0 0 0 0 1

A3 N/A N/A A1 A1 A2L A1 N/A

 - (SS) 

260 450 N/A N/A N/A 368 N/A 412

1

N/A

  N/A

 N/A

300f – 310e 280d 315c 300i 175i 200h N/A 200l 350i

Dichloro-Ethane (DCE) has very low ODP despite its chlorine content, due to its very short life-time in the atmosphere. DCE is part of the chlorinated Very Short-Lived Substances (VSLS), which are currently being investigated for their impact on the ozone layer. Despite the intense investigation, whether such substances should be regulated, or not, is still unclear, but latest assessments agree upon the fact that the VSLS impact is of orders of magnitude lower than that of classical ozone-depleting substances (CFC, HFC, etc…) [61].

6

R365MFC Sulfur-Dioxide

186.85 157.49

1.01 6.30

804 <1

3 3

0 0

A2 N/A

N/A

Toluene Water

318.60 0.08 3 2 3 N/A 480 315c,d 373.95 0.07 <1 0 0 A1 N/A N/A a NFPA 704 [40], b [45–47], c [48], d [44], e [50], f [51], g [52], h [53], i fluid datasheet,l [54]

 - (SS)   - (SS)

594 N/A

N/A N/A

Where Qsink (kW) is the heat flow rate rejected at the condenser and Wcomp (kW) is the power absorbed by the compressor. Eq. 1 can be rewritten considering that Qsink is also equal to the sum of Wcomp and the heat flow rate absorbed at the evaporator Qsource (kW), therefore the following holds (Eq. 2): Qsink 

COP  Qsource COP  1

(2)

In the case under consideration, Qsource is considered as waste heat. Such heat flow rate is considered to be provided by a medium which enters the VCHP evaporator at a temperature Tsource,in, and is cooled down to a temperature Tsource,out. Similarly, the heat sink, which absorbs the thermal energy upgraded by the VCHP, enters the VCHP condenser at a temperature Tsink,in, and is heated up to Tsink,out. If the temperature differences ΔTsource = Tsource,in − Tsource,out and ΔTsink = Tsink,out − Tsink,in are fixed, then each VCHP operational condition is uniquely determined by a couple of Tsource,in and Tsink,out. In other words, the VCHP operational conditions are conveniently defined by the temperatures at which the waste heat is provided (Tsource,in) and the temperature at which the waste heat is upgraded (Tsink,out). For each fluid selected in Section 2, the classical VCHP thermodynamic cycle, as reported in Figure 1 (a), is simulated in its steady state conditions for several combination of Tsource,in and Tsink,out. The parameters assumed in the simulations are summarized in the same picture and below:  The same temperature variations occur both in the heat source and in the heat sink. Therefore: ΔTsource = ΔTsink =10 K. The choice of different source / sink temperature glides influences the results. As a rule of thumb, larger sink temperature differences lead to higher COPs as the fluid subcooling phase during the condensation can be deeper exploited; on the contrary, larger source differences, keeping Tsource,in unchanged, are detrimental for the COP, since the evaporation temperature is pushed down, forcing the system to operate with higher gross temperature lifts. To avoid adding further complexity, the source / sink temperature differences are left unchanged in the analysis. The interested reader should refer to dedicated works (e.g. [32]) for a detailed analysis of the effect of different source / sink temperature variations.  The minimum temperature difference between the fluids achieved during the condensation and evaporation is 5 K. Therefore, for the condenser and evaporator, pinch points are: ΔTevap = ΔTcond =5 K;  The pressure losses along the heat exchangers are negligible;  The lamination process in the throttling valve is assumed to be adiabatic;  The compressor isentropic efficiency is 0.7. Further details about the choice of this value for the compressor are reported in Table 2.  The subcooling and the superheating at the exit of the condenser and the evaporator, respectively, are not assumed a priori. They are calculated in order to maximise the COP, once the condensation and evaporation conditions have been fixed in function of Tsource,in and Tsink,out. The considered VHCP layout is reported in Figure 1 (b). 7

The fluids are compared on the basis of the achieved COP value, which is uniquely determined for each couple of Tsource,in and Tsink,out. Therefore, once Qsource has been fixed, from Eq. 2 it follows that Qsink can be calculated for each combination of temperatures through the value of COP. The most efficient fluid for each temperature combination is that with the best COP. In this context, “most efficient fluid” is not a perfect synonym of “best fluid”. Firstly, only the fluids that comply with some technical limitations (e.g. on the compressor discharge temperature) can be considered realistic candidates for the VCHP applications. Secondly, the choice of the working fluid is also subjected to economic considerations, like those on the cost of the equipment required to handle the fluid. Particularly, in the case of the compressor, which is anyway the most critical piece of equipment for the VCHPs, a relevant parameter is the VHC (Volumetric Heating Capacity), defined as follows (Eq. 3): VHC 

Qsink Vcomp,in

(3)

Where Vcomp,in is the volumetric flow rate at the compressor suction. The VHC is a sort of economic parameter which accounts for the compressor size. In this context “the higher, the better”, so for any combination of Tsource,in and Tsink,out, the fluid with the higher VHC is likely to be the fluid with the smaller (i.e. cheaper) compressor. To consider this in the analysis, the VHC is used to compare the different fluids in conjunction with the COP. Since the two indicators are often in contrast [15], as the fluids with a high COP typically have a low VHC, and vice versa, the trade-off between the efficiency (higher operative revenues) and low volumetric flow rates (lower initial costs) is analysed and discussed.

Figure 1. (a) Temperature-relative heat transfer diagram for the evaporator and the condenser of the case with R1233zd(E) as working fluid and Tsink,out = 130 °C and Tsource,in = 70 °C. (b) Simulated VCHP layout. The fluids which can be considered realistic candidates for the high temperature VCHP applications are selected from the those reported in Table 1 applying the following constraints:  Evaporator pressure must be above 0.5 bar. This value is arbitrary, but it is considered reasonable to avoid air infiltration in the system. As a matter of fact, this problem is similar to that reported for ORC condenser, for which it is advisable to avoid sub-atmospheric 8



conditions [1]. Anyway, a certain degree of vacuum is acceptable, especially in small-sized systems which have moderate heat exchanger dimensions; The degree of superheating at the compressor exit must be as low as possible. Previous analysis set at 180 °C the maximum temperature of the working fluid, to avoid the thermal decomposition of the compressor lubricant [7,32]. Nonetheless, this limitation is subjected to the use of reciprocating compressors or to other compressor technologies which require lubrication. Other compressor technologies which are not affected by this issue could be used, like the oil-free centrifugal compressor. However, in this analysis the 180 °C limitation is maintained, in conformity with previous studies and since most of the refrigeration / heat pumping compressors require lubrication. As a matter of fact, even if oilfree centrifugal compressors are gaining relevance, not all of the dynamic machines use lubrication (chapter 38 of [55]). Therefore, assuming a maximum temperature equal to 180 °C can be seen as conservative hypothesis, which makes the analysis more general. Furthermore, the assumed maximum superheated temperature is the minimum between 180 °C and the current Tsink,out plus 50 °C (Eq. 4): Tmax  min(Tsink,out  50;180)



(4)

This constraint not only limits Tmax to 180 °C, but also sets a moving boundary that excludes all the fluids which do not comply with this specification. The rationale behind this choice is three-fold. First, heating a fluid to a much higher temperature than that required is not efficient, since a significant exergy destruction will occur in the heat exchanger. Second, a high degree of superheating is considered as the hint that the fluid is not suited for basic VCHP configurations, but requires alternative solutions, e.g. multi-stage intercooled compression, for being properly exploited. Third, large temperature differences across one side of the heat exchanger produces significant thermal stresses, which may damage the equipment. In the analysis, centrifugal compressors have been considered, due to the large volumetric flows that are encountered in the investigated applications. Even though reciprocating compressors have a wide application range in the refrigeration field, large-size chillers often use screw compressors and, in the last years, centrifugal compressors. In small-size high temperature VCHP applications other compressor technologies can also be found, such as scroll and, more rarely, rotary compressors [15,28]. For the sake of keeping the analysis realistic and rule out the fluids that would require too complex solutions for being employed, the maximum number of compressor stages is limited to 4. Two-stage centrifugal compressors are currently used for large chiller applications; therefore, a four-stage compressor should be considered as a trade-off between the necessity of using standard technologies and the flexibility required to include nonstandard applications, which might use components specifically designed or adapted. The number of stages NS is estimated through the following procedure outlined below. From the total isentropic enthalpy variation during the compression Δhis, which can be calculated directly from the VCHP cycle, the polytropic enthalpy variation Δhpol can be calculated as (Eq. 5): hpol 

his pol

(5)

is

Where ηpol and ηis are the polytropic and isentropic efficiencies of the compressor, respectively. If, as an approximation for a preliminary NS estimation, the ideal gas behaviour is assumed to describe the compression process with satisfactory accuracy, the following can be written (Eq. 6): 9

hpol

n pin   n  1 in

n 1    pin  n   1  p   out    

(6)

Which is the well-known relation of the polytropic specific work, where the subscripts in / out refer to the inlet and the outlet of the compressor, respectively. Eq. 6 can be numerically solved with respect to n, as all the other quantities are known from the cycle configuration and from Eq. 5. For the purpose of this simplified estimation, each stage of the compressor can be considered as characterized by the same of values of ηpol and μpol, as reported in chapter 38 of [55], where μpol is the polytropic load coefficient. Using this parameter, the polytropic enthalpy variation for each stage can be written as (Eq. 7): 2 hpol , st   pol  utip , st

(7)

where utip,st is the impeller tip-speed for the analysed stage. Mechanical limitations set the maximum tip-speed of refrigeration centrifugal compressors to about 400 m/s [55,56], whereas for compressor working with low molecular mass fluids (like the water) such limit can be brought up to 550 m/s [56,57]. Often, a stricter limit is imposed by the impeller pseudo-Mach number, defined as the ratio between utip,st and the speed of sound at the compressor stage inlet ain,st. Such dimensionless parameter is limited to 1.5 to avoid flow instabilities and inefficiencies [55,58]. Given this, Eq. 7 can be rewritten as (Eq. 8):





hpol , st   pol   min utip,max ;1.5  ain,st   

2

(8)

From Eq. 6 applied to the stage, the pressure ratio imposed by the stage can be estimated, since n is considered the same for all the stages. The stage inlet pressure is known, and the outlet pressure can be calculated (Eq. 9): n

pout , st

 hpol , st  in, st n  1  n 1  pin, st    1   pin, st n  

(9)

From the definition of polytropic efficiency, the actual enthalpy difference imposed by the stage to the fluid can be calculated and, from this, the stage outlet enthalpy can be calculated as well (Eq. 10): hout , st  hin, st 

hpol , st

(10)

 pol

At this point, if hout,st is higher than, or equal to, the enthalpy at the end of the whole compression hout,comp, which is calculated from the VCHP thermodynamic cycle, the procedure can be stopped. Otherwise, the procedure can be iterated until this condition is met. In this case, the quantities at the stage outlet are assumed to be equal to those at the following stage inlet and a new acoustic velocity can be calculated. The number of iterations needed to satisfy the condition (hout,st ≥ hout,comp) is equal to the number of stages (NS) needed to perform the compression. The values assumed for the parameters used in Equations (5), (7) and (8) are reported in Table 2. 10

Table 2. Parameters assumed for centrifugal compressor polytropic analysis. Sources: chapter 38 of [55] and [56,57].

Parameter: Value:

ηpol 0.76

ηis 0.70

μpol 0.55

umax [m/s] 400

umax,water [m/s] 550

4. Results and discussion The results presentation is divided in three parts: in Section 4.1 the decision process that leads to rule out the fluids which do not comply with technical limitations introduced in Section 3 is illustrated and discussed. In Section 4.2 the loss of efficiency due to a potential trade-off between COP and VHC is discussed. Finally, in Section 4.3, the effect of the system size, in terms of compressor technology applicability is investigated.

4.1

Effect of technical limitations on fluid selection

To allow a better comprehension of the fluid selection process, and to highlight the problems that the use of each fluid could produce, the most relevant operational parameters are provided for a selection of the investigated fluids. The results related to the end-of-compression temperature Tcomp, the evaporation pressure pevap and the number of compressor stages NS are reported in Figures 2, 3 and 4, respectively. Eight fluids for each parameter are reported; the first four are fluids which must be ruled out for the value assumed by that parameter, while the other four are the good ones, i.e. some of the best four according to that parameter. The upper limit for Tcomp has been set to 180 °C (see Section 3), based on technical consideration on the compressor lubricant oil degradation. Despite this, oil-free compressors are gaining relevance, thus this limitation might not hold anymore in the near future. The fluid that perform the worst is water, which achieves temperature values that are unacceptable even in the case of oil free compressors. This means that water might be used in alternative cycle layouts that could help to bring Tcomp to more reasonable values. This issue is related to the shape of saturation curve and the slope of the isobaric lines in the investigated operative frame. Given this, other fluids with similar saturation curve shape suffer from the same problem. This is true especially for Methanol that does not comply with the temperature limitation for every sink and source temperature values. For the “optimal” fluids, the difference between Tsink,out and Tcomp never exceeds 30 K. The best fluids, in this regard, are Cyclopentane and R1234ze(Z), which have a maximum Tsink,out − Tcomp difference around 22 K for the cases with high temperature lift (Tsource,in = 50 ÷ 60 °C, Tsink,out = 140 ÷ 150 °C), and R1224yd(Z) and R1336mzz(Z), which feature extremely low values of the Tsink,out − Tcomp difference, roughly between 10 K and 12 K for the first and always lower than 10 K for the second. Such low values are again due to the shape of the saturation curve and isobars trend, which allow for the compression to end near to the two-phase region, thus reducing the need for the desuperheating in the VCHP condenser.

11

Figure 2. Compressor discharge temperature in °C for a selection of eight working fluids. The results are reported in function of Tsink,out and Tsource,in. Examining the values of pevap, the results reported in Figure 3 suggest that, again, water is not a suitable fluid for high temperature VCHP applications. Some fluids, like MM (Hexamethyldisiloxane) or Novec649, do not have extremely low saturation pressures on their own (otherwise they would have been ruled out in the earlier stages of the fluid selection, as specified in Section 2) but the great saturation curve positive slope force it to lower the saturation temperature and increase the superheating to finish the compression without entering the two-phase zone. As shown in Figure 3, some fluids have vertical pevap lines, whereas others have positive slopes. Theoretically, the first is the expected behaviour, since the condensation temperature should not play any role into the evaporation conditions, which are dictated by the heat source thermal profile. When a deviation from this behaviour occurs, the given saturation curve shape imposes a correction to the evaporation temperature (and thus pressure) to avoid ending the compression process in the wet vapour state. When this happens, the heat source thermal profile sets only an upper limit to the evaporation temperature. As it can be seen, also promising fluids like R1336mzz(Z) show this behaviour, which is detrimental for the COP of the heat pump, since the actual operating temperature lift is increased, being equal the source and sink temperatures. To overcome this limitation, a cycle modification is needed, like the introduction of a regenerative heat exchanger in series with the evaporator. In such a way, the additional heating needed to perform a compression which ends with dry steam is performed in the second heat exchanger, where the hot liquid coming from the condenser heats up the refrigerants coming from the evaporator to the desired temperature. Since the analysis is focused on the classical heat pump cycle configuration, the regenerated architecture is not considered in the paper. This penalizes some otherwise promising fluids like R1336mzz(Z) and R1244yd(Z), which needed the regeneration to be used efficiently. In these cases, similarly to water, the paper results should not be understood that such fluids are not suited for HTHP applications, but that they require cycle modifications to be used at their best. Finally, the use of a fluid with the evaporation pressure that is not dependent on the conditions at the condenser could be beneficial in off-design operating conditions, since it simplifies the control of the heat pump. Apart from these considerations, a detailed analysis of the off-design behaviour of the VCHP is out of the scope of the present analysis and therefore was not investigated in detail. The operating pressures of the best performing fluids varies in a wide range, but all of these values are commonly encountered in the refrigeration and HVAC applications. Some concerns could be 12

arisen by the inclusion of acetone, which have sub-atmospheric pressures in a part of the investigated operational framework. As stated in Section 2, although sub-atmospheric conditions could lead to air-infiltration, a certain degree of vacuum can be handled, particularly if it is small as in this case.

Figure 3. Evaporation pressure pevap (bar) for a selection of eight working fluids. The results are reported in function of Tsink,out and Tsource,in. The last monitored parameter is the number of compressor stages (Figure 4). As previously discussed, centrifugal compressors have been considered. The number of stages is estimated with the iterative procedure outlined in Section 3 by considering the stage parameters reported in Table 2. This is a simplification, obviously, but it allows the identification of the approximate number of the stages required to perform the prescribed compression process. The results are a good tool to compare different solutions even though a more detailed procedure is required for the actual design of the VCHP compressor. The number of stages resulting from the calculation is an integer value, which leads to a step-wise contour map of the required number of stages. Each contour line represents the edge of the step and divide the Tsink,out and Tsource,in plane in portions. Those which lay under the contour line require the number of stages reported on the plane portion under the line, while those which lay over the contour line require the number of stages reported on the plane portion over the line. When only one contour line is reported, the fluid does not require any more stages than those indicated on the plane portion over the line, for any of the analysed Tsink,out and Tsource,in combinations. As can be seen, water, due to the very large enthalpy increase, requires a very large number of stages. This confirm that the VCHP basic configuration is not viable using such fluid, and, if water is to be used, cycle modifications aimed at reducing compressor work must be introduced (e.g. intercooling). Similar considerations can be drawn for ethanol, methanol and ammonia, even though the number of stages is smaller. From the number of stage point of view, these fluids can sometimes (or most of the time, like in the case of water) be ruled out due to the limitations to 4 stages, which is anyway arbitrary and could be loosened, or tightened, based on economic considerations. Other fluids, such as R1233zd(E), R1336mzz(Z) and cyclopentane, in most of the Tsink,out and Tsource,in combinations, require only a two-stages compressor and, for some of the applications with low temperature lift (Tsink,out < 120 °C and Tsource,in > 70 °C) only one stage. It is noteworthy that current applications of centrifugal compressors in refrigeration and heat-pumping field, as they are proposed by compressor manufacturers [59], are two-staged solutions. Therefore, such number of 13

stages should not be considered as high, but rather as a standard of the centrifugal compressor technology, when it is applied in the refrigeration and heat-pumps field. Finally, the result reported in Fig. 4 confirm the technical suitability of such compressor technology for this kind of applications.

Figure 4. Compressor number of stages NS for a selection of eight working fluids. The results are reported in function of Tsink,out and Tsource,in. After the application of the limitations introduced in Section 3, the fluid applicability range can be summarized as reported in Table 3.

Table 3. Summary of fluids applicability ranges. The analysed temperature ranges are divided into four quadrants: low-source / low sink temperature (Tsource,in < 65 °C - Tsink,out < 130 °C); high-source / low-sink temperatures (Tsource,in > 65 °C - Tsink,out < 130 °C); high-source / high-sink temperature (Tsource,in > 65 °C Tsink,out > 130 °C); low-source / high-sink temperatures (Tsource,in < 65 °C - Tsink,out > 130 °C). Fluid Acetone Ammonia Benzene Cyclo-Pentane Cyclo-Propane Cyclo-Hexane Dichloro-Ethane Dimethyl-Carbonate Ethanol Iso-Butane Iso-Hexane Iso-Pentane Methanol MM n-Butane n-pentane Neopentane

Tsource,in < 65 °C Tsink,out < 130 °C  a,d b   (partial) d b a,b b a,b  (partial) d  (partial) b  a b   

Tsource,in > 65 °C Tsink,out < 130 °C  a,d    (partial) d   (partial) a  (partial) b a  (partial) d   a b   

14

Tsource,in > 65 °C Tsink,out > 130 °C  (partial) a a,d   d   (partial) a  (partial) b a,b d   a b  (partial) d  

Tsource,in < 65 °C Tsink,out > 130 °C a a,d b  d b a,b b a,b d b  a,c b  (partial) d  

Novec649 R1224yd(Z) R1233zd(E) R1234ze(Z) R1336mzz(Z) R245ca R365MFC Sulfur-Dioxide Toluene Water

Violated constraint:

b       a b a,b,c

 (partial) b b b d   (partial)  (partial) d      (partial) d  (partial) d          a a a b b b a,b a,b,c   a,b,c a :Tcomp < 180 °C; b :pevap > 0.5 bar; c :NS < 4; d Tcond < Tcrit

Only 8 fluids, out of the 27 considered in this study, can be used in all the operative context without infringing one, or more, of the introduced limitations. Those fluids are: cyclo-pentane, iso-pentane, n-pentane, neopentane, R1233zd(E), R1336mzz(Z), R245ca, R365MFC. The fluids that collect more infringements against the constraints are water, ethanol and methanol. This does not rule out the use of such fluids from the high temperature VCHP applications, but the main message is that the basic VCHP architecture is not suited for those fluids. As a matter of fact, if with some layout modifications Tcomp is lowered, ethanol could be viable for Tsource,in > 65 °C and Tsink,out < 130 °C.

4.2

Most efficient fluids and trade-off between COP and VHC

After having excluded all the solutions that do not comply with the limitations imposed in Section 3, the fluids with the highest COP for each combination of Tsink,out and Tsource,in can be selected. The result of this selection process is reported in Figure 5 (a), where the most efficient fluid for each operational condition can be found. Figure 5 (a) shows that for high Tsource,in and low-to-medium Tsink,out, dichloro-ethane (DCE) is the recommended fluid. In the same range of heat source temperature, but at higher Tsink,out, benzene should be used. Moving to the lower side of Tsource,in, at lower Tsink,out acetone should be used, while at higher heat sink temperature cyclo-pentane is appropriate.

Figure 5. (a) Most efficient working fluids with the related COP values (solid lines). (b) Most efficient working fluids with the related VHC values (solid lines) (solid lines). The results are reported in function of Tsink,out and Tsource,in.

15

In figure 5 (b) the VHC values for the same fluids of Figure 5 (a) can be found. The VHC of the selected fluids is much lower than the values commonly found in the heat pump practice, which are between 3000 and 6000 kJ/m3 [15]. The reported values are even lower to the threshold value of 1000, often seen as the lower practical limit for positive displacement compressors [15]. This limitation is particularly pressing for reciprocating and screw compressors, but, since the VHC is related with the physical size of the compressor, also the centrifugal ones would benefit from high VHCs. A low VHC does not pose any direct technical limitation, as, for example, it is sufficient to use more than one compressor, to allow for larger VHC values. It is worth noting that this is a common practice in commercial refrigeration systems, which often are equipped with multiple compressors. In this case, the problem is not the low VHC, but the size of the systems, which would require compressors currently unavailable on the market at the moment. The VHC can be increased at the cost of the efficiency, selecting fluids not only based on the COP, but considering also the VHC for each operational configuration. Since the VHC is essentially an economic indicator, the trade-off between this and the COP should be based on economic considerations. A detailed economic analysis is out of the scope of the present work, then only an arbitrary trade-off criterium is presented here. Basically, for each combination of Tsource,in and Tsink,out two rankings of fluids are defined: the first is based on the COP, while the second on the VHC. In each ranking the fluids are ordered based on the relevant parameter (COP or VHC), and 25 points are assigned to the first fluid, 24 to the second, and so on, up to the last which receives one point. The scores obtained by each fluid in the two rankings are weighted with two arbitrary parameters α and β = 1- α and then summed to create a new ranking, which decides, based on both the COP and the VHC, which is the best fluid for that operational configuration. With this approach, it is easy to calculate how much efficiency is theoretically lost in favour of using high VHC fluids. In Figure 6 (a, b, c) three cases are reported: (a) α = 0.75 and β = 0.25; (b) α = 0.5 and β = 0.5; (c) α = 0 and β = 1. The case with α = 1 and β = 0 is reported in Figure 5 (a). For each case, the relative decrease of COP, compared to the fluids of Figure 5 (a), is reported. The COP variation contours are fragmented to consider the discontinuities due to the different fluid used in different zones of the operational envelope. The reasons why the COP variation contours are missing is that the same fluid of Figure 5 (a) is used. The same holds for Figure 7 (a, b, c), where the VHC relative increase can be read. As resulted in Figure 6 (a), as soon as the VHC becomes a decision variable, R1233zd(E) becomes a recommended fluid. Even losing between 8% and 12% of COP, the VHC relative gain, in Figure 7 (a), is roughly between 125% and 110% for low Tsource,in and medium-to-high Tsink,out, and between 135% and 175% for the same values of Tsource,in and low Tsink,out. More relevant VHC increments are achieved around Tsource,in = 65 °C, where, reducing the COP of a share between 14% and 17% the VHC can increase of about 3 – 4 times. When the COP and the VHC become equally relevant (Figures 6 (b) and 7 (b)), R1233zd(E) is the fluid that best combines efficiency and low volumetric flow. In all the operational conditions this fluid is the recommended one. Ideally, dividing the operative range in two (high and low Tsource,in) the greatest gains in terms of volumetric flow are encountered at high heat source temperatures, where the VHC increase is roughly included between 300% and 400%. In the same zones the COP reduction is between 7% and 14% for Tsink,out < 130 °C and between 15% and 20% for Tsink,out > 130 °C. When the fluid choice is dictated only by VHC, the advantages in terms of volumetric flow are significant, with a maximum increase of VHC around 14.5 times in case the of use of cyclopropane (Tsource,in > 70 °C and Tsink,out < 115 °C) which is near to its critical condition for those operating temperatures. In other cases, the VHC increase is more limited, but still very high. It ranges from a relative increase of 100% to a maximum one over 750% for Tsource,in < 65 °C, with a related loss of efficiency, which ranges between 13% and 27%. Similarly, for Tsource,in > 65 °C the VHC increase ranges from 300% to the above cited 1450%, with a related COP decrease which is always higher 16

than 10% and most of the times higher than 20% with a maximum value around 28% for Tsource,in around 70 °C and Tsink,out around 140 °C. Concluding, high advantages in terms of volumetric flows can be gained with a relatively modest efficiency reduction and therefore the use of less efficient fluids to reduce the compressor size should be considered as a viable option. This trade-off is dominated by economic considerations: on one hand there is the reduction of capital costs and even the possibility of using well-known compressor technologies such as the positive-displacement ones, which suffer from high volumetric flow-rates, and could become viable even for high capacity applications, if the proper refrigerant is used. On the other hand, giving up even a slight part of the COP should be carefully evaluated, because it directly penalizes the revenues for all the application lifetime, which can be easily equal to 15, or even 20, years. A good compromise could be the use of R1233zd(E), which has high VHC (it is the fluid of choice for Tsink,out > 140 °C, when the VHC is the only criterium) but also a decent COP, and could therefore be a suited fluid especially for the range characterized by Tsource,in < 65 °C.

Figure 6. Relative COP variation in respect to the values reported in Figure 5 (a) for three level of VHC relevance: (a) COP accounts for 75% in ranking the fluids, and VHC accounts for 25%; (b) COP and VHC are equally important; (c) COP accounts for 0% in ranking the fluids, and VHC accounts for 100%. The results are reported as a function of Tsink,out and Tsource,in.

Figure 7. Relative VHC variation in respect to the values reported in Figure 5 (b) for three level of VHC relevance: (a) COP accounts for 75% in ranking the fluids, and VHC accounts for 25%; (b) COP and VHC are equally important; (c) COP accounts for 0% in ranking the fluids, and VHC accounts for 100%. The results are reported as a function of Tsink,out and Tsource,in. 17

4.3

Effect of the size on the compressor choice

Both the VHC and the COP are specific indicators, so they do not depend on the size of the system. The relevant parameter which is affected by the heat source thermal size (or by the amount of heat flowrate to be provided at the heat sink) is the volumetric flow rate. This, together with the discharge pressure can give some hints about the compressor technology choice, as reported in Figure 8. The ranges of application of different compressor technologies are reported and compared to the operational field of the VCHP in four different significant cases: the fluids reported in Figure 5, which are those which maximise the COP, reported for two sizes of the systems (500 and 2000 kWth at the evaporator), and the R1233zd(E), which proved to be a good compromise between the COP and the VHC, for the same sizes of the system. The Single Stage (SS) reciprocating compressors are likely to not be the best suited for the HTHP applications of the proposed sizes (500 and 2000 kWth at the evaporator), due to the large volumetric flows, while the discharge pressure would not be a problem, and the same holds for their Multi Stage (MS) counterpart. For smaller size applications, reciprocating compressors, screw and twinscrew compressors are likely to be the most used. Finally, for very small applications (less than 10 kWth at the evaporator) scroll compressors are also used [15]. This leads to conclude that either multiple compressors or different compressor technologies (e.g. centrifugal ones) should be used for high capacity applications. The relatively low pressures needed to operate the fluids with high COP allow the use of screw compressors, although near to their technological limit for both maximum pressure and, in case of large thermal size, volumetric flow rate. Differently, for the R1233zd(E), although it has the highest VHC among the fluids reported in Fig. 8, the volumetric flow rates are too large for SS reciprocating compressors, while the pressures are too high for the SS screw compressors. In this case, the choice of centrifugal (both SS and MS) seems to be the most suited, even though the smaller size applications could be covered with MS reciprocating compressors. Concluding, the fluids selected in this analysis for the high temperature VCHP require MS centrifugal compressors, being such applications characterized by combinations of discharge pressure and volumetric flow which are likely to be well handled by such compressor technology. Depending on the VCHP thermal size, namely the heat flow rate supplied at the evaporator, other compressor technologies could become viable, like SS centrifugal, SS screw, and, especially if a major importance is attributed to the volumetric flow during the operating fluid choice, MS reciprocating. The data about the compressor applicability range are taken from chapter 2 of [60], but the SS reciprocating maximum pressure limit has been updated according to [7,32] from around 10 bar to around 30 bar.

18

Figure 8. Compressor technology application envelopes and high temperature VCHP operational conditions for the most efficient fluids and the R1233zd(E) (the fluid with good compromise between COP and VHC). The VCHP operational conditions are reported for 500 and 2000 kWth of thermal input at the evaporator. Adapted from [60], with SS reciprocating maximum pressure updated according to [7,32].

5. Conclusion A wide screening of the working fluids suitable for the VCHP has been carried out for several heat source and heat sink temperature combinations. The investigated VCHP operating field ranges from 50 °C (lower considered heat source temperature) up to 150 °C (higher considered heat sink temperature). By applying several environmental and technical limitations, the investigated pool of fluids has been reduced to 27, and, among those, the ones which showed the best COP for each combination of heat source and heat sink temperature have been selected. As a result, acetone is the recommended fluid for low heat source and high heat sink temperatures; benzene is the most suited for the case of high source and high sink temperatures; cyclopentane is for low source and high sink temperatures; finally, the dichloroethane is recommended for high source and low sink temperatures. All these fluids are flammable or highly flammable, and this could impair and discourage their use due to safety reasons. Nonetheless, the refrigeration practice showed that the use of toxic and flammable fluids can be justified in terms of economic advantages, when the efficiency gain is important. If less efficient fluids are preferred due to a higher VHC and lack of toxicity or flammability, synthetic refrigerants (namely the HFO R1336mzz(Z) and R1234ze(Z) and the HCFO R1233zd(E) and R1224yd(Z)) provide a very effective alternative. Furthermore, the use of safe fluids could also help to widespread the HTHP applications, being such fluids usable in all the contexts. That said, as the analysis demonstrated for the case of VHC, when a parameter other than the COP is pursued, the efficiency losses, can be easily higher than 10% and up to 20%, which is not negligible. Therefore, these trade-offs should be considered carefully. The fluids which resulted to be the most efficient ones have also low VHC, if compared with the classical heat pump fluids, therefore the efficiency loss due to their replacement with high VHC fluids has been discussed. It resulted that the R1233zd(E) can give a good compromise between COP and VHC, featuring COP relative decreases which can be as low as 7% and, in most of the 19

cases, are comprised between 10% and 20%, while bringing VHC relative increases which in some cases are over 400% and always higher than 110%. Finally, the use of different compressor technologies has been analysed as a function of the VCHP size. It resulted that multistage and single stage centrifugal compressor are likely to be the most suited, due to the size of the proposed application (500 ÷ 2000 kWth at the evaporator), even if in some cases screw compressors could handle the resulting volumetric flows. For smaller size applications, different compressor technologies, like the reciprocating one, could be used. In that case, different fluids could be the recommended ones, since the number of compression stages resulted to be a limiting factor for some of the investigated fluids. With different compressor technologies, this limitation does not hold anymore, and different fluids could comply with the remaining technical restrictions.

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The most efficient configuration of high temperature heat pumps is investigated Heat source/sink temperatures impact on working fluid applicability is assessed Compressor outlet temperature, stage number, inlet pressure effects are discussed The trade-off between COP and volumetric heating capacity is quantitively assessed The R1233zd(E) resulted as the best compromise fluid in all configurations

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