Thermodynamic performance assessment of carbon dioxide blends with
low-global warming potential (GWP) working fluids for a heat pump water
heater☆Baomin Daia, Chaobin Dangb, Minxia Li*a, Hua Tianc, Yitai Maaa

Key Laboratory of Efficient Utilization of Low and Medium Grade Energy, MOE, Tianjin University,
Tianjin 300072, Chinab
Department of Human and Engineered Environmental Studies, The University of Tokyo, Chiba
277-8563 Japanc
State Key Laboratory of Engines, Tianjin University, Tianjin 300072, China*

Corresponding author. Tel.: +86 022 27406040; fax: +86 022 27404741.
E-mail address: tjmxli@tju.edu.cn (M. Li); dbm@tju.edu.cn (B. Dai)

Abstract: Blends of CO2 with ten low-global warming potential (GWP) working fluids are evaluatedfor use in a heat pump water heater. The effects that the discharge pressure, component ratio (XCO2),
hot-water outlet temperature and chilled water inlet temperature have on the coefficient of performance(COP) of heat pump are analyzed when the pinch point of the heat exchange is considered . It is foundthat the temperature glide of zeotropic mixture has a good thermal match with the temperature changeof water as two pinch points appear in the gas cooler/condenser . The good thermal match in the heatexchanger promotes COP of heat pump. Addition of low-GWP working fluids to pure CO2 can reducehigh-side pressure. The results show that CO2/R41 and CO2/R32 are suitable candidates for heat pumpwater heaters because of their high COP and low high-side pressure in comparison with those of a pure
CO2 cycle.
Keywords: CO2; low-GWP refrigerant; zeotropic; heat pump water heater; pinch point; thermal match☆

The original paper was presented on the 11th IIR Gustav Lorentzen Conference on Natural

Refrigerants (GL2014), Aug. 31 to Sept. 2, 2014, Hangzhou, China.

1Nomenclature
E
h
I
m
s
T
P
Q
W
XCO2
Subscripts
0
1, 2, 2s, 3, 4bC
Carnot
CW
Com
Cri
E
e
Exe
Exp
HW
In
m
Opt
Pinch
r
Tot
TP
W
Greek symbols
ηAcronyms
COP
CP
GWP
HC
HCFC
HFC
HFO
HVAC&R
LEL
MAC
ODP
OEL
PP

exergy (kJ/kg)
specific enthalpy (kJ/kg)
irreversibility (kJ/kg)
mass flow rate (kg/s)
specified entropy (kJ/(kg K))
temperature (°C)
pressure (MPa)heat exchange capacity (kW)power input (kW)mass fraction of CO2 (-)
dead state
state point
boiling
gas cooler/condenser
Carnotchilled watercompressor
criticalevaporator
electricalexergyexpansion valve
hot water
inletmechanicaloptimalpinch point
refrigeration
total
thermodynamic perfectibility
water
efficiency
coefficient of performancecritical point
global warming potential
hydrocarbonhydrochlorofluorocarbonhydrofluorocarbonhydrofluoroolefin
heating, ventilation, air-conditioning and refrigeration
lower explosive limit
mobile air-conditioningozone depletion potential
occupational exposure limit
pinch point21. Introduction
The working fluid selection for an air-conditioning and refrigeration system is not only determinedby thermodynamic properties, but is also largely related to environmental constraints. The use ofchlorinated substances, such as chlorofluorocarbons (CFCs) and hydrochlorofluorocarbons (HCFCs),
has been successfully reduced by the Montreal Protocol (UNEP, 1987). The Kyoto Protocol (UNFCCC,
1997) depicted the reduction of the emission of refrigerants with high global warming potentials(GWPs). The European Union fluorinated greenhouse gases (F-gas) regulation (Regulation (EU) No
517/2014) demands a phase-out of the use of F-gas with GWP above 150 in domestic refrigerators andfreezers by January 1, 2015, and in new types of mobile air-conditioning (MAC) systems by January 1,
2017. In order to respect the environmental issues and follow the new proposed regulations, great effort
has been made to identify the next generation of candidate refrigerants. Consequently, several
approaches are under consideration, including low-GWP refrigerants, natural refrigerants, andrefrigerant mixtures (Miyara et al., 2012).
Among the low-GWP refrigerants, new synthetic refrigerants hydrofluoroolefins (HFOs) such as
R1234yf and R1234ze, which have zero ozone depletion potential (ODP) and a low GWP of 1, aregetting increasing attention as alternatives to R134a. However, these two HFOs have more complex
molecule structures than most HFCs, and will be more costly (McLinden et al., 2014). Moreover, they
are classified as “A2L” in the ASHRAE standard 34 safety group (ANSI/ASHRAE Standard 34-2010),
because of their mild flammability. In addition, because of their relatively low operation pressure, thepressure drops that occur in the pipelines and the heat exchanger have a significant influence on system
performance. Apart from the HFOs, several HFCs (R152a, R161, and R41) and hydrocarbons (R290,
R1270, R600, R600a, R601, and R601a) that have a GWP below 150 are also promising candidates for
HVAC&R applications. Nevertheless, some of them possess moderate flammability or require a highoperating pressure.
In the class of natural refrigerants including CO2, NH3, and HCs, CO2 is considered to be the mostpromising refrigerant because of its advantages in practical applications (Lorentzen, 1994, 1995;
Lorentzen and Pettersen, 1993), such as zero ODP, GWP of 1, non-toxicity, non-flammability, easy
availability, compatibility with lubricants, and excellent thermophysical and transport properties. CO2
can be used in various refrigeration, air-conditioning, and heat pump systems, especially domestic
water heating systems (Kim et al., 2004). A domestic CO2 heat pump water heater was firstcommercialized in 2001, and by 2008 the system had shipped over 500,000 units (Miyara et al., 2012).
However, the following two main factors restrict the widespread use of the CO2 system: The system operates at a significantly high pressure, which leads to higher initial cost because of
the requirement of high-pressure bearings. Moreover, the safety during operation should also be
carefully considered.
 Large throttling losses and irreversibility occur during the expansion process, which decrease thewhole system performance.
In order to reduce the high-side pressure and improve the energy efficiency of a CO2 system, someresearchers proposed the use of zeotropic refrigerant mixtures of CO2 and eco-friendly refrigerants. The
previous studies are listed as follows:
Kim and Kim (2002) investigated the performance of zeotropic mixtures of CO2/R134a and
CO2/R290 used in an autocascade refrigeration system. Both simulation and experimental studies werecarried out, and the results indicate that as the CO2 concentration decreases, the COP tends to increase
while the high-side pressure declines. Di Nicola et al. (2005) analyzed blends of CO2 with several
HFCs (including R125, R41, R32, and R23) for a low-temperature-stage working fluid used in acascade refrigeration cycle. They concluded that the CO2 blends are preferable for cascade systems
with a low-temperature approaching 200 K. They also studied the characteristics of blends of CO2 withnatural refrigerants (R170, R290, R1150, R1270, and RE170) used in a cascade refrigeration cycle (DiNicola et al., 2011). Kim et al. (2008) studied blends of CO2/R290 and indicated that the dischargepressure of the mixtures decreases with an increase of the R290 fraction. Furthermore, adding R290 to
CO2 improves the system energy efficiency, but reduces the cooling capacity. They also indicated theimportance of appropriate temperature matching of the zeotropic refrigerant mixture with the heattransfer fluid, in order to improve the refrigeration system efficiency. Sarkar and Bhattacharyya (2009)
proposed blends of CO2/R600 and CO2/R600a as working fluids for medium- and high-temperature
heating applications. Simulation results showed that the blends could be employed in heat pumps for
simultaneous cooling and heating applications, with effective performance. The CO2/R600a blend isrecommended as the best alternative refrigerant to R114 for a high-temperate heat pump, because of its
superior COP and lower high-side pressure in comparison with CO2 systems. Niu and Zhang (2007)
tested a binary mixture of CO2/R290 (71/29 mole fraction) for a cascade refrigeration cycle. The results
revealed that the COP and cooling capacity of the mixture are both higher than those of R13, especially3when the evaporation temperature is above 201 K. Onaka et al. (2010) analyzed the heat pump
performance using a zeotropic mixture of CO2/RE170 for a hot water supply system, with the CO2mass concentration ranging from 0 to 1. They concluded that the system obtains a maximum COP at a
certain heat-rejection pressure for each mixture, and the COPs of the mixtures and the pure RE170 are
higher than that of pure CO2. The irreversibility of each component of the system was also discussed.
Zhang et al. (2013) theoretically and experimentally investigated a transcritical cycle operating with
CO2/R290 with a CO2 mass fraction larger than 0.78. They proposed a correlation for the prediction of
optimal heat rejection pressure and verified the simulation results with experimental data.
Hakkaki-Fard et al. (2014) developed a numerical model to simulate refrigerant mixtures used for acold-climate air-source heat pump. They concluded that the heat pump using the mixture of CO2/R32(20/80) shows the best improvement in heating capacity. Dai et al. (2014) also studied blends of CO2
with low-GWP working fluids used in a transcritical Rankine cycle for low-grade heat recovery. The
binary mixtures show better thermodynamic performance than the pure CO2. As can be seen from theabove studies, the addition of low-GWP refrigerants to CO2 can improve the energy performance andreduce the high-side pressure of a CO2 heat pump/refrigeration system. Moreover, the flammability of
the added component, such as HFCs and HCs, can be reduced because of the dilution effect of pureCO2 (Chen et al., 2009; Kondo et al., 2006).
Yang et al. (2005) performed an exergy analysis of the transcritical carbon dioxide refrigeration
cycles with a throttling valve and with an expander, and concluded that the main exergy losses occur inthe gas cooler and the compressor. Then, a further study (Yang et al., 2007) compared three different
variations of transcritical carbon dioxide two-stage compression cycles with expanders, and the exergy
efficiencies were achieved. Tao et al. (2010) experimentally and theoretically analyzed a CO2
residential air-conditioning with an internal heat exchanger. They concluded that the exergy loss in thegas cooler is the highest one, accounting about 30.7% of the total exergy loss in the system. Sun and
Ma (2011) investigated the exergy loss in a transcritical carbon dioxide refrigeration cycle with ejector
and with throttling valve. They found that ejector instead of throttling valve can reduce 25% exergy
loss or more.
There is growing awareness that global warming is becoming a more significant and challenging
environmental issue. Thus, the use of a blend of low-GWP refrigerants with pure CO2 is promisingbecause of the following advantages: Reducing the high operation pressure of a CO2 system, improving
the system energy efficiency, and alleviating the flammability while maintaining a low GWP.
However, in the previous studies, only a few CO2 binary mixtures were evaluated without
comparison with each other. Thus, in this study the blends of CO2 with ten low-GWP refrigerants areevaluated, in order to screen the promising alternative mixtures for application to a heat pump water
heater.
2. System modeling
2.1 System description and working fluid selection
The water-source heat pump water heater system using the CO2 binary mixtures with low-GWP
refrigerants is shown in Fig. 1. The system consists of a compressor, gas cooler/condenser, expansion
valve, and evaporator.
The terms “high GWP” and “low GWP” are comparative in nature. United Nations Environment
Programme (2010) proposed an explicit definition of low, moderate, and high GWP for greenhousegases:
 “Low” if GWP < 300;
 “Moderate” if 300 < GWP < 1000;
 “High” if GWP >1000.
In the field of HVAC&R, R134a is a widely used refrigerant that is considered to beenvironmentally safe refrigerant and essentially non-toxic. However use of R134a must be significantly
reduced according to the provisions of Kyoto protocol (UNEP, 1997), because its GWP is as high as1370 (Calm and Hourahan, 2011). Although R134a belongs to the high-GWP category, it is still
selected in this study in respect of its extensive application. The refrigerants selected as potential
candidates are shown in Table 1. As can be seen, most of the fluids are low-GWP refrigerants according
to the above classification, except for R32 (moderate GWP) and R134a (high GWP). The physical,
safety, and environmental data are described in Table 1, and the refrigerants are sorted by normal
boiling point (Tb).
2.2 Assumptions for the model
The analysis was conducted based on the following assumptions:(1) The system runs under steady working condition.
(2) The pressure drop and heat loss of fluid flowing inside the heat exchangers (gas cooler/condenser
4and evaporator) and connecting pipes are ignored.
(3) The refrigerant is at saturated state at the evaporator outlet.
(4) The gas cooler/condenser and evaporator are counter-flow heat exchangers.
(5) The heat transfer fluids of both the gas cooler/condenser and evaporator are water at an ambientpressure of 0.1 MPa.
2.3 Energy and exergy mathematical model
For the specified application of a heat pump, the most important factors are the inlet and outlet
temperatures of the heat sink and heat source, also called the external conditions. In this work, the aim
is to assess suitable working fluids used for water heater. Thus, the heat pump system in this study is
specified based on the rating condition provided in the Chinese standards GB/T 21362-2008 (2008) and
GB/T 23137-2008 (2008), which is shown in Table 2.
The status points of the refrigerant are determined from the fixed approach temperature differenceat the pinch point in the heat exchangers. The pinch point is defined as the location where the minimum
temperature difference between the working fluid and secondary fluid appears in the heat exchanger.
The pinch temperature differences at the gas cooler/condenser and evaporator are both set to 5 °C inthis study. The mass flow rate of working fluid is set as 1 kg/s for simplicity. The modeling algorithm isshown in Fig. 2. The status points of the cycle are determined by means of iteration, with a 0.01 °C
tolerance of the pinch temperature difference in the exchangers and a 1 kPa pressure tolerance. The
properties of the mixtures are from REFPROP 9.0 (Lemmon et al., 2010). The energy and exergy
analysis are .described as the following equations (Lior and Zhang, 2007): For the compressor:
WCom  mr  h2  h1  me (1)where ηm and ηe are the mechanical and electrical efficiency, respectively, both assumed to be 0.9.
I Com  T0 mr  s2  s1 (2)where ICom is the irreversibility at the compression process, and T0 is dead state temperature specified as
25 °C (298.15 K)
Com 

h2S  h1h2  h1(3)

where ηCom is compressor isentropic efficiency, which is set as 0.8 in this study.
 For the gas cooler/condenser:

QC  mr  h2  h3 

mHW  mr  h2  h3   hHW , out  hHW , in 

I C  T0  mr  s3  s2   mHW  sHW , out  sHW , in  

(4)
(5)
(6)

where QC and IC are the heat exchange capacity and irreversibility inside the gas cooler/condenser, andmHW is the mass flow rate of hot water.
 For the expansion valve:h4  h3(7)

I Exp  T0 mr  s4  s3 (8)where IExp is irreversibility at the expansion process.
 For the evaporator:

QE  mr  h1  h4 

(9)EE  mr  h4  h1  T0  s4  s1  

(10)mCW  mr  h1  h4   hCW , in  hCW , out 

I E  T0  mr  s1  s4   mCW  sCW , out  sCW , in  

(11)
(12)

where QE, EE, and IE are the heat exchange capacity, exergy, and irreversibility inside the evaporator,and mCW is the mass flow rate of chilled water.
 For the total system:Q
COP  C
(13)
WCom

ITot  I Com  I C  I Exp  I E
The exergy efficiency (the second law efficiency) for the overall system (Lior and Zhang, 2007):
5

(14)

I
EE

 Exe  1  Tot(15)

3. Results and discussion
3.1 Analysis of energy efficiency
3.1.1 Effect of discharge pressure
The previous studies of Kauf (1999) and Liao et al. (2000) indicated that there exists an optimalhigh pressure for a transcritical CO2 cycle. However, those analyses were all performed based on theinternal status points (i.e., the working fluid side) without considering the external conditions (i.e., theheat transfer fluid side). In practical applications, the external conditions are more significant. Thus, theheat pump water heater system was analyzed based on the inlet and outlet temperatures of the chilled
and cooling water in this study. The COP variations with the high pressure using CO2 and CO2/R32(50/50) are shown in Fig. 3(a) and (b), respectively. As can be seen from the figures, the COP firstincreases significantly and then decreases gradually as the high pressure increases, and a maximum
COP occurs at a certain high pressure. The optimal high pressure for the pure CO2 system is 9.187 MPa,
while that of the CO2/R32 (50/50) system is 5.638 MPa. Whereas in the studies by Kauf (1999) andLiao et al. (2000) the COP curve had a smooth shape, in this study the curve of COP changes sharply atthe optimal high pressure. This is because in the previous analyses (Kauf, 1999; Liao et al., 2000), thegas cooler outlet temperature of the refrigerant side was set as a constant. In this study, the outlet
temperature of the gas cooler (T3) first increases slightly from 8 to 8.4 MPa, then drops suddenly with
the increase in gas cooler pressure until 9.2 MPa, to meet the requirements of the constant pinchtemperature difference and the water side temperature profile. The outlet temperature then remains at
20 °C with further increase in the pressure, as shown in Fig. 3. As indicated in the study by Liao et al.
(2000), the outlet temperature of the gas cooler has a significant influence on the system performance
because the heating capacity (Qh) is very sensitive to its value, and the higher the outlet temperature,
the lower the COP becomes. Fig. 3 shows that Qh increases substantially until the high pressure reaches
the optimal value, and turns to a moderate increase. However, the compression work (W) increases
linearly, with a low slope. The COP variation tendency of the CO2/R32 (50/50) system shown in
Fig. 3(b) is similar to that with pure CO2, indicating that an optimal high pressure also exits for thesystem using a mixture.
The T-s diagrams of the two working fluids at different high-side pressures are shown in Figs. 4(a)and (b), respectively. As shown in Fig. 4(a), in the case of PC = 8 MPa, the status of gas cooler outlet(point 3) is to the right side of the critical point with the temperature being as high as 35.6 °C. Thisimplies that the refrigerant is throttled too early, leading to a relatively low heating capacity and highvapor quality at the inlet of the evaporator (point 4). It can also be seen that the pinch point of the gas
cooler (PPC) appears inside the heat exchanger. With an increase in high-side pressure to PC = 9 MPa,
the CO2 is further cooled inside the gas cooler to a lower outlet temperature. When the high pressure
increases to the optimal value of 9.187 MPa, two pinch points appear in the gas cooler, one inside andone at the outlet of the gas cooler. With a further increase of the high pressure to PC = 10 MPa, there is
only one pinch point again, which is at the outlet of the gas cooler. For the mixture CO2/R32 (50/50),
there are also two pinch points at the optimal high pressure of 5.638 MPa during the heat rejection
process. However, the heat rejection process for the CO2/R32 (50/50) system is below the critical point(i.e., a subcritical cycle). Thus, the heat transfer takes place in the condenser rather than the gas cooler
for the CO2/R32 (50/50) system. Consequently, it can be found that the maximum COP of the heatpump water heater system is achieved at the condition where two pinch points appear in the gas
cooler/condenser.
However, in the process of evaporation, the pinch point (PPE) is located at the inlet of theevaporator for pure CO2, whereas for CO2/R32 it is situated at the outlet. This difference is attributed tothe difference between the internal and external temperature glides. The temperature glide of CO2
blends with ten low-GWP and moderate-GWP refrigerants at various mass fractions is shown in Fig. 5.
The temperature glide of CO2/R32 (50/50) is 14.5 °C at the bubble point temperature of 10 °C.
However, for pure CO2 the evaporation temperature is constant, leading to zero internal temperature
glide. At the rating condition specified in Table 2, the external temperature glide in the evaporator is
5 °C. Consequently, for the evaporation process of pure CO2 shown in Fig. 4(a), when the internal
temperature glide is smaller than the external one, the pinch point is situated on the inlet side of therefrigerant. For the zeotropic mixture of CO2/R32 (50/50), shown in Fig. 4(b), the situation is reversed,
with the pinch point on the outlet side when the internal temperature glide is smaller.
3.1.2 Effect of mixture fraction
Fig. 5 presents the variation of temperature glide with CO2 mass fraction (XCO2). For each mixture
pair, a maximum temperature glide appears at a certain XCO2. As indicated in our previous study (Dai etal., 2014), the maximum temperature glide is dependent on the normal boiling point (Tb). For the pure6

refrigerants studied in this work CO2 has the lowest Tb, as shown in Table 1, and the larger thedifference of Tb between the two components of the zeotropic blends, the higher the maximum
temperature glide. As indicated in Fig. 5, the blends of CO2/R41 and CO2/R32 show relatively low
temperature glide, those of CO2/R1270 and CO2/R161 are moderate, and the mixture CO2/R1234ze has
the highest glide temperature.
The analysis in Section 3.1.1 explains that there exists a maximum COP and corresponding
optimal high pressure for each working fluid (pure or zeotropic fluid). Thus, to reasonably compare thesystem performance with different mixtures used in a heat pump water heater, the following
comparisons are performed under the optimal working condition with maximum COP. Fig. 6 shows the
maximum COP of ten kinds of CO2 blends with various CO2 mass fractions at the optimal highpressure. It can be seen from Fig. 6 (a) that except for CO2/R41, the COP of the mixtures show asaddle-shaped change with CO2 mass fraction. For CO2/R41, the COP first increases and then decreases,
and a maximum COP of 4.65 is found at XCO2 = 0.4. In addition, the COP of both the CO2/R41 mixturesand pure R41 are higher than that of pure CO2 (COP = 4.47). However, for the other mixtures, there
exists a minimum COP in the range of XCO2 = 0.5–0.8, which is lower than that of pure CO2. In
particular, CO2/R1234ze (60/40) achieves the lowest COP of 3.45, which is 22.8% lower than that ofpure CO2. Although most blends (except CO2/R41) present a much lower COP than that of pure CO2,
the mixtures CO2/R32, CO2/R161, CO2/RE170, CO2/R1270, CO2/R152a, and CO2/R290 all givepromising results. CO2/R32 has good performance, with the COP ranging from 4.38 to 4.67 for XCO2changes from 0 to 1, which is comparable to the COP of pure CO2. In particular, for the cases
XCO2 = 0.1 and XCO2 = 0.8, it operates with the highest COP of 4.67. Other blends (CO2/R161,
CO2/RE170, CO2/R1270, CO2/R152a, and CO2/R290) perform better than pure CO2 at XCO2 = 0.1–0.2
or XCO2 = 0.9. It can also be seen from Fig. 6(a) that the pure fluids (XCO2 = 0) all present a lower COP
than that of pure CO2, except for R41.
In Fig. 6(b) the corresponding optimal high pressure (solid line) and critical pressure (dashed line)
are depicted. It can be seen that the optimal high pressure of the mixtures and pure fluids are all lower
than that of pure CO2. Furthermore, the high pressure decreases with the reduction of CO2 massfraction. Moreover, the higher the normal boiling point of the refrigerant component is, the more thehigh pressure can decrease. The mixtures with the three lowest high pressures are CO2/RE170,
CO2/R152a, and CO2/R1234ze, whereas CO2/R41 and CO2/R32 have the highest high pressures.
It can also be found that under the optimal working condition, the systems using the CO2/R41
mixture are in a transcritical cycle, because the high pressure is above the corresponding criticalpressure at any CO2 mass fraction. However, for other mixtures, at high CO2 mass concentration (XCO2= 0.8) they are transcritical cycles, and are subcritical cycles at lower XCO2.
The change in shape of the COP curves, as depicted in Fig. 6(a), is mainly caused by thetemperature matching in the evaporator. The T-s diagrams using CO2/R32 at different CO2 massfractions under the optimal high pressures are shown in Fig. 7. It can be seen that for CO2/R32 (80/20),
the mixture in the evaporation process (4-1) shows a good thermal match with the chilled water, withthe temperature profiles of the two sides parallel to each other. Furthermore, the heat rejection process(2-3) presents a smooth curve, which approximately passes through the critical point. For the case of
CO2/R32 (20/80), although it is a subcritical cycle, the zeotropic mixture condensation process causes atemperature glide in the condenser, which also shows a proper thermal match with the cooling water. In
the above two cases (XCO2 = 0.8 and XCO2 = 0.2), the temperature profiles of the internal and external
fluids are almost parallel with each other. Because the temperatures of the chilled water are fixed, agood temperature match leads to a relatively high mean evaporation temperature, leading to a muchhigher COP.
However, for the cases of CO2/R32 (50/50) and pure R32, the temperature match is inferior to theabove two cases. In particular, when XCO2 = 0.5 the internal temperature glide at the evaporationpressure is as high as 13.5 °C, compared with the external temperature glide of 5 °C. Thus, a poor
thermal matching is observed in the evaporator. The temperature difference at the inlet of therefrigerant side (11.2 °C) is much larger than that of the outlet (5 °C). Although a good thermal
performance is achieved in the heat rejection process, a negative temperature match in the evaporator
causes a relatively low evaporation temperature, which reduces the performance of the whole system.
Consequently, high energy efficiency can be achieved based on the condition that there are goodthermal matches in both the gas cooler/condenser and evaporator.
Except for R41, the COPs of systems using pure fluids are all lower than with pure CO2, forexample the pure R32 shown in Fig. 7. This is because the critical temperatures of these pure fluids arevery high, and the heat rejection takes place in the condenser under isothermal conditions during phasechange. This leads to a poor thermal match because of the restriction of the pinch point at the dew point
under condensation pressure, resulting in a higher condensation temperature. Therefore, relatively low
7COPs are achieved for the pure working fluids. However, for pure R41, the critical thermophysicalcharacteristics are close to that of CO2, and the maximum temperature glide is 1.8 °C (shown in Fig. 5),
which results in a better thermal match in the evaporator. Moreover, as shown in Fig. 6(b), the optimalhigh pressure of R41 is 5.886 MPa, which is 3.3 MPa lower than that of pure CO2, leading to a smaller
expansion loss. Thus, the performance of pure R41 is preferable to that of pure CO2.
3.1.3 Effect of hot water outlet temperature
If the outlet temperature of water heater is relatively low, then the tank of the water heater may
become a breeding ground for Legionella. Thus, it is necessary to increase the supply watertemperature to kill the bacteria. However, the regulations for the outlet temperature of a heat pumpwater heater are not consistent across different countries. As shown in Table 3, it is the lowest for China
with an outlet temperature of 55 °C, and is as high as 65 °C in the Japanese standards.
Thus, the system performances with a wide range of hot water outlet temperature are shown in
Fig. 8. As can be seen from the figure, the maximum COP decreases as the hot water outlet temperature
increases. In addition, for CO2/R41 the COP values are all higher than those of pure CO2 with the outlet
temperature ranging from 45 to 65 °C. However, for CO2/R32 when XCO2 = 0 or XCO2 = 0.5, the system
performance is marginally inferior to that using pure CO2, but at XCO2 = 0.2 or XCO2 = 0.8, the system
operates with a rather high COP that is also better than that of CO2/R41. Although the performance of
the CO2/R41 system is more robust than CO2/R32 across the whole range of mass concentrations(XCO2 = 0–1), at some concentration ratios (XCO2 = 0.2 or XCO2 = 0.8) CO2/R32 is superior to CO2/R41,
with higher COP and lower high pressure.
3.1.4 Effect of chilled water inlet temperature
As indicated by the above analysis, the thermal match in the evaporator has a significant influenceon the whole system performance. However, this conclusion was achieved on the basis of the rated
condition with external temperature glide of 5 °C in the evaporator. In contrast, in practical applications,the external temperature glide varies with the seasonal and climatic conditions. Thus, the following
results are achieved from analyses with the chilled water outlet temperature of 10 °C as specified inTable 1, but with the inlet water temperature changing from 12.5 to 30 °C. The results are shown in
Fig. 9. For CO2/R41, the COP is almost constant, due to the low value of temperature glide in theevaporator. For CO2/R32, the COP of the blends first increases and then remains constant. For the casesof CO2/R1270 and CO2/R161, the shapes of the COP curves are similar. In particular, at XCO2 = 0.5, theCOPs increase monotonically with the inlet temperature, and reach maximum values of 5.90 and 6.06,
respectively, when the inlet temperature is 30 °C for the two mixtures. This can also be attributed to thetemperature match in the evaporator, as shown in Fig. 10.
Fig. 10 illustrates the T-s diagrams of a CO2/R161 (80/20) system under different chilled water
inlet temperatures (TW, E, in). The temperature glide of CO2/R161 (80/20) is 19.5 °C at the dew point of
10 °C. It can be found that in the case of a chilled water inlet temperature of 15 °C, as shown in
Fig. 10(a), the evaporator presents a bad thermal match: At the evaporator inlet on the refrigerant side,
the approach temperature difference between the internal and external fluids is extremely high(17.1 °C), leading to a large irreversibility and reducing the mean evaporation temperature. With theincrease of the chilled water inlet temperature, as shown in Fig. 10(b), the two temperature profiles
gradually become parallel to each other, and the COP also increases, as shown in Fig. 9(d). When theinlet temperature increases to 26.5 °C, as shown in Fig. 10(c), two pinch points appear at both the inlet
and outlet of the evaporator. However, with a further increase in inlet temperature (TW, E, in = 30 °C),shown in Fig. 10(d), the COP will not change because under this working condition the external
temperature glide (20 °C) is larger than the internal one (16.5 °C), and the pinch point shifts to the inlet
side of the evaporator on the refrigerant side. Consequently, it can be concluded from the above
analysis that a proper temperature match in the evaporator leads to a good system thermal performance.
The pure fluids are suitable for a working condition with relatively low external temperature glide, but
for high external temperature glide, the zeotropic mixtures are superior for dealing with the temperaturechange that occurs in the process of phase change.
From the above analysis, it can be found that CO2/R41 at various concentration ratios is a good
choice to substitute for pure CO2, for use in a heat pump water heater, because of its robust system
performance, relatively low GWP, and low rejection pressure. Furthermore, the flammability can be
reduced with the addition of CO2. CO2/R32 is also a suitable candidate because of its rather high COP
and mild temperature glide. Therefore, it is more suitable for use in a heat pump water heater withmoderate temperature glide (~10 °C) on the chilled water side. Moreover, the high pressure can also be
distinctly reduced. However, the other binary mixtures are not recommended because their temperature
glides are extremely large. The internal temperature glide of these binaries can hardly match with theexternal temperature glide, except at some special concentration ratio with low or moderatetemperature glide, such as XCO2 = 0.1.
8

3.2 Analysis of exergy efficiency
The exergy efficiency is equivalent to the thermodynamic perfectibility (Ma et al., 2012).
Thermodynamic perfectibility is also called Second Law efficiency (Lior and Zhang, 2007), which isdefined as the ratio of the system COP to that of the corresponding Carnot cycle (Yuan et al., 2011):

TP 

COP
COPCarnot(16)

The Carnot cycle COP is related to the temperatures of the heat sink and source. In this study, thosetemperatures are specified in Table 2. Thus, COPCarnot is a constant under the rated working condition,and there also exists a maximum exergy efficiency for each component, which is shown in Fig. 11. It
can be seen that the exergy efficiency curves have the same variation tendency as those of the
maximum COP shown in Fig. 6(a).
The irreversibility per unit heating capacity (i) can directly represent the performance of thesystem, and is defined as:
I
i
(17)
QC
The values of i for each component are depicted in Fig. 12. The compressor performs the largest
irreversibility among the four major components for all of the fluids studied in this paper. For pure CO2,
the irreversible loss in the throttle valve (iT) is the second largest, because of the huge pressure
difference between the high- and low-side pressures, but for pure low-GWP refrigerants, or for thebinaries with low XCO2, iT is the lowest of the four major components of the heat pump system. This isbecause a lower pressure drop in the throttle process leads to a small irreversibility. It can also be foundthat the curves of iC and iT show an inverse trend with the variation of XCO2. The phenomenon is
attributed to the fact that with the decrease in XCO2 the pressure difference between high- and low-side
pressures reduces, but the thermal match deteriorates due to the system changing from transcritical tosubcritical. Thus, the irreversibilities in the two components complement each other. It can also be
noted from the figure that the variation tendency of iTot is the same as that of iE. Therefore, it can beconcluded that the irreversibility occurring in the evaporator is the depending factor to the performanceof system. Thus, a proper thermal match in the evaporator dominates the whole heat pump system
performance.
4. Conclusions
In this study, carbon dioxide blends with ten low-GWP working fluids were systematicallyevaluated for use in a heat pump water heater, based on the method of pinch point with the specified
temperatures of the heat transfer fluids. The effects that the discharge pressure, component ratio,
hot-water outlet temperature, and chilled water inlet temperature had on the thermal performance of thesystem were analyzed. The following conclusions were drawn from this study:(1) A maximum COP was obtained at a corresponding optimal high pressure for each component ratio,
when two pinch points appear in the gas cooler/condenser. The addition of low-GWP working fluids topure CO2 can reduce the high pressure, which decreases as the CO2 mass fraction decreases.
(2) The system with mixture CO2/R41 is a transcritical cycle and operates with a high COP. However,for the other binary mixtures, the system changes to a subcritical cycle at lower XCO2. The temperature
glide of the mixtures that occurs in the gas cooler/condenser can be well matched to that of the hotwater.
(3) The temperature match in the evaporator plays a leading role for the whole system performance. A
proper temperature match in the evaporator promoted high COP and reduced the irreversibility.
(4) CO2 mixtures are recommended on the base of the working condition of chilled water temperature
glide:
 CO2/R41 is recommended because of the robust performance at various XCO2 under the conditionof low external temperature glide.
 CO2/R32 is more suitable for its fairly high COP and lower high-side pressure in the case ofmoderate to high external temperature glide (~10 °C).
 The other binary mixtures are not recommended for use in a heat pump water heater, except atsome special concentration ratio with low or moderate temperature glide.
Acknowledgement
This work was supported by the National Natural Science Foundation of China (No. 51176133)
and the Tianjin Municipal Science and Technology Commission Foundation (No. 12JCYBJC13800).
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11

Figure captions
Fig. 1. Schematic of heat pump water heater system.
Fig. 2. Flow chart of the modeling.
Fig. 3. Thermal performance of heat pump water heater system, variation with discharge pressure. (a)Pure CO2. (b) CO2/R32 (50/50).
Fig. 4. T-s diagrams of heat pump water heater system under different discharge pressures. (a) Pure
CO2. (b) CO2/R32 (50/50).
Fig. 5. Temperature glide for blends of CO2 with low- and moderate-GWP working fluids at bubble
point temperature of 10°C.
Fig. 6. Thermal performance of the heat pump system under optimal working condition using various
CO2 blends. (a) Maximum COP. (b) Optimal high pressure and corresponding critical pressure.
Fig. 7. T-s diagrams of heat pump water heater system using CO2/R32 as refrigerants under optimal
working condition at different concentration ratios.
Fig. 8. Variation of maximum COP with hot water outlet temperature. (a) CO2/R41. (b) CO2/R32.
Fig. 9. Variation of maximum COP with chilled water inlet temperature. (a) CO2/R41. (b) CO2/R32. (c)CO2/R1270. (d) CO2/R161.
Fig. 10. T-s diagrams of heat pump water heater system using CO2/R161 (80/20) under optimal
working condition at different chilled water inlet temperatures. (a) TW, E, in = 15 °C, (b) TW, E, in = 25 °C,(c) TW, E, in = 26.5 °C, (d) TW, E, in = 30 °C.
Fig. 11. Maximum exergy efficiency variation with mass fraction of CO2 for different mixtures.
Fig. 12. Irreversibility per unit heating capacity of every component of the system under the conditionof maximum exergy efficiency. (a) CO2/R41. (b) CO2/R32. (c) CO2/R1270. (d) CO2/R290.

12

Fig. 1. Schematic of heat pump water heater system.

13

Fig. 2. Flow chart of the modeling.

14

38
36

4.5

34

200

32

4.0

160

3.5
Qh
W
COP 3.0
T3

120
80

COP

30

26
24
22

2.5

40

28

20
0

8

9

10

11

2.0
12

18

5.0

40

Gas cooler outlet temperature (C)

240

Heat capacity or power (kW)

5.0

CO2

Gas cooler pressure (MPa)

CO2/R32 (50/50)

300

38
4.5

250

36
34

4.0

200

100

3.5
Qh
W
COP 3.0
T3

50

2.5

0
5.0

2.0
7.0

150

32

COP

Heat capacity or power (kW)

350

30
28
26
24
22
20

5.5

6.0

6.5

Gas cooler outlet temperature (C)

(a)

18

Gas cooler pressure (MPa)(b)
Fig. 3. Thermal performance of heat pump water heater system, variation with discharge pressure. (a)Pure CO2. (b) CO2/R32 (50/50).

15

100

Temperature (C)

80

CO2PC = 8 MPa

CO2
PC = 9 MPa

60

2S

PPC

40

PPC

CP 3

CP
3

20
1

4

0
-20
100

PPE

CO2
PC = 10 MPa

CO2
PC = 9.187 MPa

80

1

4

PPE

2

2S

2

2S

Temperature (C)

2

2S

2

60
40
3

20

CP

PPC

PPC

3
1

4

0

1.2

1

4
PPE

PPE
-20
1.0

CP

1.4

1.6

1.8

2.01.01.2

1.4

Entropy (kJ/kg K)

1.6

1.8

2.0

Entropy (kJ/kg K)

(a)
100

Temperature (C)80

CO2/R32 (50/50)
PC = 5 MPa

2

40

3

PPC

3

PPE

PPE
1

4

0

Temperature (C)

CP

PPC

20

80

2
2S

2S
CP

60

-20
100

CO2/R32 (50/50)
PC = 5.5 MPa

1
42

CO2/R32 (50/50)
PC = 5.638 MPa

2S

CO2/R32 (50/50)
PC = 6 MPa

40

PPC

20

3

0

4

PPC
PPE

PPE

3

1

1

-20

1.2

2

CP

CP

60

2S

4
1.4

1.6

1.8

2.0

2.2

Entropy (kJ/kg K)

1.2

1.4

1.6

1.8

2.0

2.2

Entropy (kJ/kg C)(b)Fig. 4. T-s diagrams of heat pump water heater system under different discharge pressures. (a) Pure
CO2. (b) CO2/R32 (50/50).

16

60

Temperature glide (C)50

CO2/R1234yf
CO2/R134a
CO2/RE170
CO2/R152a
CO2/R1234ze

CO2/R41
CO2/R32
CO2/R1270
CO2/R290
CO2/R161

40
30
20
10
0
0.0

0.2

0.4

0.6

0.8

1.0

Mass fraction of CO2
Fig. 5. Temperature glide for blends of CO2 with low- and moderate-GWP working fluids at bubble
point temperature of 10°C.

17

5.2
5.0

Maximum COP

4.8

CO2/R41
CO2/R32
CO2/R1270
CO2/R290
CO2/R161

CO2/R1234yf
CO2/R134a
CO2/RE170
CO2/R152a
CO2/R1234ze

0.2

0.6

4.6
4.4
4.2
4.0
3.8
3.6
3.4
0.0

0.4

0.8

1.0

0.8

1.0

Mass fraction of CO2(a)

10High pressure (MPa)

8
6
4
2
0
0.0

PC, opt Pcri


0.2

0.4

0.6

Mass fraction of CO2
PC, opt Pcri

CO2/R41
CO2/R32
CO2/R1270
CO2/R290
CO2/R161

CO2/R1234yf
CO2/R134a
CO2/RE170
CO2/R152a
CO2/R1234ze(b)Fig. 6. Thermal performance of heat pump system under optimal working condition using various CO2
blends. (a) Maximum COP. (b) Optimal high pressure and corresponding critical pressure.

18

100

CO2

Temperature (C)

80 PC = 9.187 MPa

CO2/R32 (80/20)
PC = 7.629 MPa

2S 2

60

3

0

4

PPC

PPC

3

1

4

1
1.2

1.4

1.6

1.8

2.0

1.2

Entropy (kJ/kg K)
CO2/R32 (20/80)

1.4

1.6

2

R32
PC = 3.207 MPa

40

2S

1

4

1.2 1.4 1.6 1.8 2.0 2.2

Entropy (kJ/kg K)
2

PPC

PPC
3

20 3
4

2.0

PPE

3

CP

CP

60

1.8

Entropy (kJ/kg K)
2S

80 PC = 4.065 MPa

Temperature (C)

2

PPC
PPE

PPE

-20
1.0

-20

2S

CP

CP

20

0

CO2/R32 (50/50)
PC = 5.638 MPa

CP

40

100

2S 2

PPE

1

4

1
PPE

1.2 1.4 1.6 1.8 2.0 2.2

1.0 1.2 1.4 1.6 1.8 2.0 2.2

Entropy (kJ/kg K)

Entropy (kJ/kg K)

Fig. 7. T-s diagrams of heat pump water heater system using CO2/R32 as refrigerants under optimal
working condition at different concentration ratios.

19

5.4

5.0
4.8

5.0

4.6
4.4
4.2
4.0
45

CO2/R32
100/0
80/20
50/50
20/80
0/100

5.2

Maximum COP

5.2

Maximum COP

5.4

CO2/R41
100/0
80/20
50/50
20/80
0/100

4.8
4.6
4.4
4.2

50

55

60

4.0
45

65

50

55

60

Hot water outlet temperature (C)

Hot water outlet temperature (C)

(a)

(b)

Fig. 8. Variation of maximum COP with hot water outlet temperature. (a) CO2/R41. (b) CO2/R32.

20

65

6.0

Maximum COP

5.6
5.2

CO2/R41
100/0
80/20
50/50
20/80
0/100

CO2/R32
100/0
80/20
50/50
20/80
0/100

4.8
4.4
4.0

(a)

3.6
6.0

Maximum COP

5.6
5.2

(b)
CO2/R161
100/0
80/20
50/50
20/80
0/100

CO2/R1270
100/0
80/20
50/50
20/80
0/100

4.8
4.4
4.0
3.6

(c)
15

20

25

30

Chilled water inlet temperature (C)

(d)
15

20

25

30

Chilled water inlet temperature (C)

Fig. 9. Variation of maximum COP with chilled water inlet temperature. (a) CO2/R41. (b) CO2/R32. (c)CO2/R1270. (d) CO2/R161.

21

100
80

Temperature (C)2

CO2/R161 (80/20)
Tw, E, in = 15 C
PC = 6.278 MPa

2S

60

2S

40

PPE
PPE

PPC

3

20

-20
100

2

CP

CP

1

4

4

(a)

PPC

3

1

0

(b)

CO2/R161 (80/20)
Tw, E, in = 26.5 C80
PC = 6.546 MPa

Temperature (C)

CO2/R161 (80/20)
Tw, E, in = 25 C
PC = 6.508 MPa

2S

60

CO2/R161 (80/20)
Tw, E, in = 30 C
PC = 6.546 MPa

2

CP

2S

2

CP

40
20

3

0

4

PPC
PPE

4

(c)

-20
1.0

PPC

3

1

1

PPE

(d)
1.2

1.4

1.6

1.8

2.0

2.2

Entropy (kJ/kg K)

1.0

1.2

1.4

1.6

1.8

2.0

2.2

Entropy (kJ/kg K)

Fig. 10. T-s diagrams of heat pump water heater system using CO2/R161 (80/20) under optimal

working condition at different chilled water inlet temperatures. (a) TW, E, in = 15 °C, (b) TW, E, in = 25 °C,(c) TW, E, in = 26.5 °C, (d) TW, E, in = 30 °C.

22

Maximum exergy efficiency (%)

36
34
32

CO2/R41
CO2/R32
CO2/R1270
CO2/R290
CO2/R161

CO2/R1234yf
CO2/R134a
CO2/RE170
CO2/R152a
CO2/R1234ze

30
28
26
24
22
20
0.0

0.2

0.4

0.6

0.8

1.0

Mass fraction of CO2
Fig. 11. Maximum exergy efficiency variation with mass fraction of CO2 for different mixtures.

23

0.20

CO2/R41

CO2/R32(a)(b)

CO2/R1270

CO2/R290

(c)

(d)

iCom

Irreversibility per unit
heating capacity (kW/kW)

iC
0.16

iE
iT

0.12

iTot

0.08
0.04
0.00

Irreversibility per unit
heating capacity (kW/kW)

0.20
0.16
0.12
0.08
0.04
0.00
0.0

0.2

0.4

0.6

0.8

1.0 0.0

Mass fraction of CO2

0.2

0.4

0.6

0.8

1.0

Mass fraction of CO2

Fig. 12. Irreversibility per unit heating capacity of every component of the system under the conditionof maximum exergy efficiency. (a) CO2/R41. (b) CO2/R32. (c) CO2/R1270. (d) CO2/R290.

24

Table 1. Physical, safety, and environmental data for refrigerants (Calm and Hourahan, 2011).

Substance

Safety data

Physical data

Environmental data

Molecular

Tb

Tcri

Pcri

OEL

LEL

ASHRAE

Atmospheric

mass

(°C)

(°C)

(MPa)

(PPMv)(%)

34 safety

life (yr)

ODP

GWP

group
R744

44.01

−78.4

31.1

7.38

5000

None

A1

>50

0

1

R41

34.03

−78.3

44.1

5.9

-

-

A2

2.8

0

107

R32

52.02

−51.7

78.1

5.78

1000

14.1

A2

4.9

0

675

R1270

42.08

−47.7

92.4

4.66

660

2.0

A3

0.001

0

~20

R290

44.10

−42.1

96.7

4.25

1000

2.1

A3

0.041

0

~20

R161

48.06

−37.6

102.2

5.09

-

3.8

-

0.21

0

12

R1234yf

114.04

−29.5

94.7

3.38

500

6.2

A2L

0.029

0

<4.4

R134a

102.03

−26.1

101.1

4.06

1000

none

A1

14.0

0

1370

RE170

46.07

−24.8

127.2

5.34

1000

3.4

A3

0.015

0

1

R152a

66.05

−24.0

113.3

4.52

1000

4.8

A2

1.4

0

124

R1234ze

114.04

−19.0

109.4

3.64

1000

7.6

A2L

0.045

0

6

Table 2. Standard rating condition of water source heat pump water heater (GB/T 21362-2008; GB/T

23137-2008).
Hot water at gas cooler/condenser side

Chilled water at evaporator side

Inlet temperature (°C)

Outlet temperature (°C)

Inlet temperature (°C)

Outlet temperature (°C)

15

55

15

10

Table 3. Standards of heat pump water heater of different countries.

Specified outlet
Standard

Country

Application field

temperature of hot
water (°C)

GB/T 23137-2008

China

Residential

GB/T 21362-2008

China

ANSI/AHRI 470-2006

United States of America-

60

EN 16147-2011

European Union

Residential

60a

JRA 4060-2009

Japan

Residential

65

JIS C 9220 -2011

Japan

Industrial

65

Commercial andIndustrial

a. Equivalent hot water temperature.

25

5555