Simulation study on performance of a dual-source hybrid heat pump unit with alternative refrigerants

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Simulation Study on Performance of a Dual-Source Hybrid Heat Pump Unit with Alternative Refrigerants Chenguang Bai , Zongwei Han , Haotian Wei , Xiaomei Ju , Xinwei Meng , Qi Fu PII: DOI: Reference:

S2666-1233(19)30004-2 https://doi.org/10.1016/j.enbenv.2019.08.004 ENBENV 4

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Energy and Built Environment

Received date: Revised date: Accepted date:

4 August 2019 23 August 2019 30 August 2019

Please cite this article as: Chenguang Bai , Zongwei Han , Haotian Wei , Xiaomei Ju , Xinwei Meng , Qi Fu , Simulation Study on Performance of a Dual-Source Hybrid Heat Pump Unit with Alternative Refrigerants, Energy and Built Environment (2019), doi: https://doi.org/10.1016/j.enbenv.2019.08.004

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Highlights 

A novel double-source hybrid heat pump unit is proposed.



The mathematical models of the unit with alternative refrigerants are established.



The unit characteristic of R134a is compared with those of R32, R290 and R600a.

1

Simulation Study on Performance of a Dual-Source Hybrid Heat Pump Unit with Alternative Refrigerants

Chenguang Baia, Zongwei Hana,*, Haotian Weia, Xiaomei Jua, Xinwei Menga, Qi Fua a SEP Key Laboratory of Eco-Industry, School of Metallurgy, Northeastern University, Shenyang 110819, China

Abstract To solve the problems of single heat source heat pump systems in severe cold regions, a dual-source hybrid heat pump unit (DSHHPU) is proposed. The mathematical models of the DSHHPU when charging R134a or its alternative refrigerants R32, R290 and R600a were established respectively, and the performance was simulated and analyzed. The results showed that the four refrigerants have different performance characteristics in different aspects. In heat pipe mode, the heating capacity and evaporating pressure of R32 are 36.94% and 59.94% higher than those of R134a. The heating capacity and evaporating pressure of R290 are 5.73% and 22.99% lower than those of R134a. The heating capacity and evaporating pressure of R600a are 43.29% and 68.08% lower than those of R134a. In vapour compression heating mode, the discharge temperature of R32, R290 and R600a are 184.88, 72.98 and 66.44% of that of R134a. The coefficient of performance (COP) of R32, R290 and R600a are 72.65, 111.59 and 117.94% of that of R134a. Finally, the effects of radiation intensity and ambient temperature on key performance parameters of the different refrigerants were analyzed. The research results provide a reference for research on refrigerant replacements for multi-heat source composite heat pump systems. Keywords: Heat pump; Performance; Simulation; Alternative refrigerants

1 Introduction In recent years, with the aggravation of air pollution, the problems of the low heating efficiency and serious pollution from traditional coal-fired boilers have attracted more and more attention. As an efficient and non-direct pollution heating technology, heat pumps are the preferred alternative to coal-fired boilers for building heating. Common heat pump systems mainly include air source heat pumps, solar heat pumps and ground source heat pumps. In low temperature environments, the performance of air source heat pumps is often affected by problems such as large compression ratios, low heat production and frosting [1, 2]. The low heat flux density and instability of solar mean that solar heat pumps have a large collector area and high initial investment, which limits their development [3, 4]. In the cold regions, the soil thermal imbalance caused by the imbalance between the building's cooling and heating loads affects the long-term performance of ground source heat pump systems [5-7]. To improve the performance of air source heat pumps operating in low temperature environments, two-stage compression heat pump systems were proposed and studied. Ko et al. [8] experimentally studied a two-stage compression heat pump system. The results showed that the coefficient of performance (COP) of the system is 36% higher than that of single-stage compression heat pump systems. In order to further improve the performance of the two-stage compression heat pump systems, Xu et al. [9] studied the optimal volume ratio of the high-and low-pressure cylinders of a heat pump system, which significantly improved the performance of the system. To reduce the impact of frosting on the performance of heat pumps, the commonly used defrosting methods include reverse cycle defrosting and hot gas bypass defrosting. Qu et al. [10] and Song et al. [11] experimentally studied the performance of air source heat pump systems using reverse cycle defrosting. Ji et al. [12] studied 2

the effect of hot gas bypass defrosting on heat pump systems. To improve the performance of solar heat pumps and ground source heat pumps operating in cold regions, different forms of composite heat source heat pump systems including solar composite air source heat pump systems and solar composite ground source heat pump systems have been proposed. Lerch et al. [13] studied the performance of different types of solar composite air source heat pump systems, and found that the composite heat pump system provides a significant energy saving compared with the air source heat pump systems. Li and Liang et al. [14, 15] simulated the effects of different parameters on the performance of solar-assisted air source heat pump systems. The results showed that the circulation flow rate, solar collector area and initial water temperature have a greater impact on system performance. Rad and Verma et al. [16, 17] studied solar-assisted ground source heat pump systems. The results showed that solar collectors are used to absorb solar energy to replenish the soil, which can effectively reduce the length of the ground heat exchanger and improve the system COP. Dai and Si et al. [18, 19] studied the effects of operating modes and control strategies on the performance of solar-assisted ground source heat pump systems. Due to the low heat flux density of solar, the use of solar energy in cold regions to solve the problem of soil thermal imbalance caused by ground source heat pumps will lead to an excessive collector area, which increases the system volume and initial investment, and reduces the superiority of the system. To achieve complementary use of solar, air and geothermal energy, we proposed a multi-source hybrid heat pump system in which the refrigerant absorbs both air and solar energy in the collector/evaporator. This avoids the problems of increased initial investment and excessive system volume caused by the separate solar collectors. The simulation study found that the system effectively improved the performance of heat pumps operating in cold areas on the basis of maintaining the soil heat balance [20]. In this system, a dual-source hybrid heat pump unit (DSHHPU) is a core device. The DSHHPU operates the heat pipe mode during the non-heating period to obtain low-temperature heat for soil heat storage, and the vapour compression heating mode for heating during the heating period. In order to further study the influence of different parameters on the performance of the unit, R134a was used as the refrigerant, and the performance of the DSHHPU under typical meteorological conditions was studied experimentally. The results showed that the performance of the DSHHPU is better than that of conventional air source heat pump units [21]. In recent years, governments have gradually begun to restrict and prohibit the use of refrigerants with high global warming potential (GWP) such as R134a. Therefore, new environmentally friendly refrigerants with low GWP values are receiving more and more attention. Among them, R32, R290 and R600a are three alternative refrigerants which are widely used at present. In order to study the cycle performance of alternative refrigerants in heat pumps, Cho et al. [22] compared the heating and cooling performance of a heat pump system with R32 or R410a. Cheng et al. [23] replaced R22 and R410A with R32 and R290 as heat pump refrigerants respectively, and studied their cycle performance. Nawaz and Bengtsson et al. [24, 25] evaluated the cycle performance of R290, R600a and R134a in heat pumps. DSHHPU system is an innovative composite heat source heat pump system. Compared with conventional composite heat source heat pump system, DSHHPU system has the advantages of less initial investment, high energy efficiency and high system performance. However, the previous research on the performance of the system was based on the use of R134a refrigerant with high greenhouse effect coefficient (GWP), and the lack of research on alternative environment-friendly 3

refrigerant. Therefore, this paper takes the DSHHPU as the research object. By establishing mathematical models, the performance of the DSHHPU with R134a, R32, R290 or R600a is simulated, which provides a reference for determining the alternative refrigerants suitable for multi-heat source composite heat pumps. Nomenclature A

area (m2)

Greece letter 2

A'

total radiation heat transfer area (m )

C

dimensionless constant

α

heat transfer coefficient [W/(m2∙℃)]; void fraction

2

Gr

refrigerant mass flow density [kg/(m •s)]

η

radiation heat transfer efficiency

H

fin height (m)

ηv

volumetric efficiency

2

I

radiation intensity [W/m ]

ηs

isentropic efficiency

L

pipe length (m)

δ

fin thickness, m

L1

distance between the tubes (m)

v

specific volume (m3/kg)

M

occlusion factor

ρ

density (kg/m3)

Mr'

single-phase refrigerant charge amount (kg)

Mr''

two-phase refrigerant charge amount (kg)

a

air

N

number of fins

c

condenser

P

pressure (Pa)

com

compressor

Q

heat exchange amount (W)

con

water-cooled condenser

Vth

theoretical gas transmission volume of the

dis

compressor discharge

Subscripts

3

compressor (m /h) W

power (W)

d

diameter (m)

f

resistance coefficient

e eva

2

g

gravity acceleration (m/s )

h

enthalpy (J/kg)

h

'

evaporator collector/evaporator

g

gas

h

high-pressure

i

inner; inlet

int

intercooler

isentropic compression enthalpy (J/kg)

l

low-pressure; liquid; intercooler

m

mass flow rate (kg/s)

m

mean value

n1

number of tube passes in the heat exchanger

o

outside; outlet

t

temperature (℃)

og

saturated gas

x

dryness

ol

saturated liquid

height difference between the condenser

r

refrigeran

and the evaporator (m)

s

radiation heat transfer

∆H

Δx

length of the micro-segment (m)

suc

suction

sup

supplement gas

val

throttle valve

w

water; wall

2. DSHHPU operation control principle The DSHHPU is mainly composed of a collector/evaporator, water-cooled condenser, throttle valve, intercooler, high-pressure compressor, low-pressure compressor and other components, as shown in Figure 1. Unlike conventional air-cooled evaporators, 4

the outer surface of the collector/evaporator is sprayed with a solar-selective absorbing coating, which enables the collector/evaporator to absorb both air and solar energy simultaneously. By switching the valves, the unit can be operated in either heat pipe mode or vapour compression heating mode. During the non-heating period, when the outdoor temperature is high or the solar radiation is strong, the valves v1 and v2 are opened, v3, v4, and v5 are closed, and the unit is operated in the heat pipe mode, as shown in Fig. 1(a). The refrigerant absorbs air energy and solar energy in the heat collector/evaporator, and then continuously evaporates into a gaseous refrigerant, making the evaporation pressure higher than the condensation pressure. Driven by the differential pressure, the gaseous refrigerant flows upward into the condenser, then be cooled by water and condensed into liquid refrigerant. Under the action of gravity, the liquid refrigerant flows downward into the heat collector/evaporator to form the refrigerant circulation in the heat pipe loop. During the heating period, the unit runs the vapour compression heating mode. At this time, v3, v4, and v5 are opened, and v1 and v2 are closed, as shown in Fig. 1(b). The refrigerant absorbs heat in the collector/evaporator and evaporates. After two stages of compression, it enters the condenser. A part of the condensed high-pressure liquid refrigerant directly enters the intercooler, and the other part is throttled to the intermediate pressure by the throttle valve 1 and then enters the intercooler. The gaseous refrigerant which is flashed due to throttling is separated from the liquid refrigerant, mixed with the gaseous refrigerant discharged from the low-pressure compressor, and then enters the high-pressure compressor for secondary compression. The saturated liquid refrigerant is further throttled by the throttle valve 2 and then enters the collector/evaporator. Water-cooled condenser

Water-cooled condenser

Cooling water

High pressure compressor

V3

Cooling water

High pressure compressor

V3

Intercooler V1

Intercooler V1

V4

V5 Four way valve Low pressure compressor

Gas-liquid separator

Throttle valve 1

V4

V5 Four way valve

V2

Low pressure compressor

Throttle valve 2

Gas-liquid separator

Solar-collecting evaporator

Throttle valve 1

V2

Throttle valve 2

Solar-collecting evaporator

(a) Heat pipe mode (b) Vapour compression heating mode Fig. 1 Schematic diagram of the DSHHPU

3. DSHHPU mathematical model In order to simulate the performance of the DSHHPU under different conditions, the main components of the unit are first established, including the collector/evaporator, water-cooled condenser, compressor and intercooler. On this basis, the unit model was established and verified. 3.1 Collector/evaporator model The collector/evaporator is a finned coil heat exchanger, and the outer surfaces of the heat exchange tubes and fins are coated with a solar radiation absorbing coating. In the vapour compression heating mode, the collector/evaporator is divided into a 5

two-phase zone and a superheated zone according to the state of the refrigerant. In the heat pipe mode, it can be divided into a supercooled zone, a two-phase zone and a superheated zone, and the heat exchange model is shown in Fig. 2(a). Solar

Solar Overcooled area

Two-phase area

I

tao hao

Overheated area Refrigerant

Refrigerant

tri hri

tro hro

Air

tai hai

Air

dL

(a) (b) Fig. 2 Schematic diagram of collector/evaporator model

Firstly, the following assumptions are made: 1) refrigerant flow is one-dimensional and uniform; 2) air flow is one-dimensional and uniform; 3) heat loss from the heat exchanger can be ignored; 4) the axial heat conduction and friction heat of the refrigerant and the thermal resistance of the tube wall can be ignored. On this basis, a parameter distribution model of the collector/evaporator is established, where a micro-segment of the collector/evaporator is as shown in Fig. 2(b). 1) Refrigerant side heat transfer equations

dQr,e  mr,e  hro,e  hri,e    r,edAi  tw,e  trm,e 

(1)

In the single-phase zone and the two-phase zone, the refrigerant-side heat transfer coefficients are calculated by the heat transfer correlations of Dittus-Boelter [26] and Shah [27], respectively. 2) Air side heat transfer equations dQa  ma  hai  hao   dAo'  I   odAo  tam  tw,e 

(2) 2

Where αo= αa+ αs, αa is the air convection heat transfer coefficient, W/m ∙℃, which is calculated by the correlation in the literature [28]; αs is the equivalent convective heat transfer coefficient of radiation heat transfer, W/m2∙℃. s  Ao' I  Ao  tam  tw,e  (3) Ao'  n1  HL1   do2 / 8  M  N  1  1   do n1  L  N  / 2   2do  n1  1 L1 / 4 (4)

3) Refrigerant pressure drop 1 P  4dLGr2 f  i di   Gr2  o1  i1 

 64 Re1  f  0.3164 Re0.25 0.0054  0.3964 Re0.3 

Re  2320 2320  Re  8 104 Re  8 104

(5)

(6)

3.2 Water-cooled condenser mathematical model The condenser is a sleeve type heat exchanger, and is divided into a superheated zone, a two-phase zone and a supercooled zone according to the state of the refrigerant, as shown in Fig. 3(a).

6

hwi twi

Water Refrigerant

Overcooled area

Two-phase area

Water

hwo two

CO2

hro tro

Refrigerant

hri tri

dL

Overheated area

(a) (b) Fig. 3 Schematic diagram of water-cooled condenser model

Firstly, the following assumptions are made: 1) refrigerant flow is one-dimensional and uniform; 2) water flow is one-dimensional and uniform; 3) the axial heat conduction and friction heat of the refrigerant and the thermal resistance of the tube wall can be ignored. On this basis, a parameter distribution model of the water-cooled condenser is established, where a micro-segment of the condenser is shown in Fig. 3(b). 1) Refrigerant side heat transfer equations

dQr,c  mr,c  hri,c  hro,c  = r,cdAi trm,c  tw,c 

2) Cooling water side heat transfer equations dQw  mw  hwo  hwi  = w dAo  tw,c  twm 

(7) (8)

3.3 Compressor model 1) Compressor displacement 2) Compressor power

mcom  v Vth vsuc

' W  mcom  hdis  hsuc  s

(9) (10)

3) Refrigerant outlet enthalpy value

' hdis  hsuc   hdis  hsuc  s

(11)

4) Intermediate pressure of two-stage compression

Pm  Pe Pc

(12)

3.4 Intercooler model A parameter concentration model is established for the intercooler, and the flow heat transfer process of the refrigerant in the intercooler is solved according to the following control equations: mrl  hri,int  hol   msup  hog  hri,int  (13) msup =mrh  mrl

(14)

3.5 Throttle valve model The throttling process is considered as an isenthalpic process. hval,i  hval,o

(15)

Refrigerant mass flow rate of throttle valve: mr,val  C i ( Pi  Po )

(16)

3.6 Refrigerant charge model The refrigerant is mainly present in the collector/evaporator, water-cooled condenser, compressor cavities, and intercooler. Therefore, refrigerant in the connection line can be ignored, and the refrigerant charge can be calculated by the following formula. 7

M  M eva  M con  M int  M com,h  M com,l

(17) The single-phase refrigerant charge amount Mr' and the two-phase refrigerant charge amount Mr'' are calculated by the following formulas: (18) M 'r    Ax M ''r  [g  (1   ) l ]Ax

 1      1   -1  g    x  l 

(19)

1

(20)

3.7 Heat pipe cycle power The cycle power of heat pipe loop is related to the height difference between condenser and evaporator and the density difference between gaseous refrigerant and liquid refrigerant. The calculation formula of the cycle power of heat pipe loop is as follows: P  ( l  g )  H  g (21) 3.8 Solution to the system mathematical model The flow charts of two modes algorithm are shown in Fig. 4. The internal flow parameters are unknown therefore three iterations are both needed for simulation of the heat pipe mode and vapour compression heating mode. Thermodynamic parameters of refrigerant were calculated by RefProp8.0.

(a) Heat pipe mode (b) Vapour compression heating mode Fig. 4 Flow chart of two modes algorithm

8

3.9 Model verification To verify the accuracy of the mathematical models, R134a was chosen as the refrigerant. The evaporating pressure and condensing pressure of the unit in the heat pipe mode and the evaporating temperature and compressor discharge temperature of the unit in the vapour compression heating mode were calculated by MATLAB software. The calculation results were compared with the experimental results [21]. When the unit is running in the heat pipe mode, the ambient temperature is 24 ~28 °C, the radiation intensity is 600 W/m2, the inlet water temperature is 11 °C, and the cooling water flow rate is 10.8 m3/h. When the unit is operating in the vapour compression heating mode, the radiation intensity is 300~700 W/m2, the ambient temperature is 3 °C, the inlet water temperature is 4 °C, and the cooling water flow rate is 0.9 m3/h. It can be seen from Fig. 5 that in the heat pipe mode, the error between the simulated value and the experimental value of the evaporating pressure is kept within 5%, and the error between the simulated value and the experimental value of the condensing pressure is between 10% and 20%. As shown in Fig. 6, in the vapour compression heating mode, the error between the simulated value and the experimental value of the evaporating temperature is within -10%, and the error of the discharge temperature is between -5% and 5%. The calculation error of the unit model is small, thus the simulation results can meet the performance prediction requirements. Condensation pressure/MPa

Evaporation pressure/MPa

0.95

Experimental value Simulation value +5% error line

0.93

0.90

0.87

0.84

Experimental value Simulation value

+10% error line +20% error line

0.90 0.85 0.80 0.75

0.81

0.70 24

25

26

27

24

28

Outdoor temperature/℃

25 26 27 Outdoor temperature/℃

28

Fig. 5 Curves of evaporation pressure (a) and condensation pressure (b) under the heat pipe mode

Discharge temperature/℃

Evaporation temperature/℃

Experimental value Simulation value +5% error line -5% error line

90

Experimental value Simulation value -10% error line

-3

-4

-5

85

80

75 -6 300

400

500 600 Radiation intensity/W/m2

700

300

400

500 600 Radiation intensity/W/m2

700

Fig. 6 Curves of evaporation temperature (a) and discharge temperature (b) under the vapour compression heating mode

4. Simulation results and analysis To compare the performance of the DSHHPU when charging different refrigerants, the performance of the unit with R134a, R32, R600a or R290 was calculated by MATLAB software. The physical properties of the refrigerant under different conditions are obtained by REFPROP software. 9

4.1 Performance of the DSHHPU in heat pipe mode The performance of the unit was simulated under the conditions of ambient temperature of 27 °C, radiation intensity of 0~800 W/m2, and radiation intensity of 700 W/m2, and ambient temperature of 24~30 °C. In the two conditions, the cooling water inlet temperature is 10 °C, and the water flow rate is 10.8 m3/h. R134a R290

1800

R32 R600a

1500

Evaporation pressure / kPa

Evaporation pressure / kPa

1800

1200 900 600

R134a R290

R32 R600a

1500 1200 900 600

300

300 0

200 400 600 Solar radiation intensity/ W/m2

800

24

26

28

30

Outdoor ambient temperature/℃

(a) (b) Fig. 7 Variation of evaporation pressure of the DSHHPU with four refrigerants

In the heat pipe mode, the refrigerant is naturally circulated, driven by the density difference, and the flow rate is low, thus the values of the evaporation pressure and the condensing pressure are very close. In order to compare the operating pressures of different refrigerants in the heat pipe mode, the evaporation pressure of each refrigerant under different conditions was calculated. Figure 7(a) shows the variation of evaporation pressure of different refrigerants with the intensity of solar radiation. When the solar radiation intensity increases from 0 to 800 W/m2, the heat transfer performance of the unit is improved. The evaporation pressures of R134a, R32, R290 and R600a are increased by 2.33, 5.96, 1.47 and 3.29%, respectively. As shown in Fig. 7(b), when the ambient temperature is raised from 24 °C to 30 °C, the evaporation pressures of R134a, R32, R290, and R600a are increased by 3.88, 2.65, 2.37, and 7.25%, respectively. The average evaporation pressures of R134a, R32, R290, and R600a are 887.43, 1419.33, 772.12, and 283.28 kPa, respectively. After comparison, it is found that the average evaporation pressure of R32 is 59.94% larger than that of R134a because the standard evaporation temperature of R32 is lower than that of R32. The average evaporation pressures of R290 and R600a are 12.99% and 68.08%, respectively, smaller than that of R134a. 12

R134a R290

12

R32 R600a

R32 R600a

10

Heating capacity/ kW

Heating capacity/ kW

10

R134a R290

8

6

4

8

6

4 0

200 400 600 Solar radiation intensity/ W/m2

800

24

26 28 Outdoor ambient temperature/℃

30

(a) (b) Fig. 8 Variation of heating capacity of the DSHHPU with four refrigerants

Figure 8(a) shows the variation of the heat capacity of the DSHHPU with the intensity of solar radiation. When the solar radiation intensity increases from 0 to 800 10

W/m2, the heat capacities of R134a, R32, R290, and R600a are increased by 9.88, 7.50, 13.13, and 16.41%, respectively. As shown in Fig. 8(b), when the ambient temperature is raised from 24 °C to 30 °C, the heating capacities of R134a, R32, R290, and R600a are increased by 15.76, 7.97, 17.55, and 22.24%, respectively. The average heating capacities of R134a, R32, R290, and R600a are 7.09, 9.71, 6.68, and 4.02 kW, respectively. The heating capacity of each refrigerant is mainly determined by its latent heat of vaporization and mass flow rate. In heat pipe mode, the refrigerant mass flow rate is determined by the difference in gas–liquid density. Although the latent heats of vaporization of R32 and R134a are less than those of R290 and R600a, the mass flow rates of R32 and R134a are much larger than those of R290 and R600a due to the large difference in gas–liquid density. Therefore, the heat capacities of R32 and R134a are larger than those of R290 and R600a. At the same time, since the latent heat of vaporization of R32 is larger than that of R134a, the heat capacity of R32 is higher than those of the other three refrigerants. It is found that the average heating capacity of R32 is 36.94% larger than that of R134a, while the average heating capacities of R290 and R600a are 5.73% and 43.29% smaller than that of R134a. In addition, since the energy consumption of the DSHHPU in the heat pipe mode is mainly composed of the energy consumptions of the water pumps and fans, the total energy consumptions of the DSHHPU with different refrigerants are basically the same, and the COP of each refrigerant is mainly determined by its heat capacity. Therefore, the COP of R32 is the largest in heat pipe mode, and the COPs of R290 and R600a are lower than that of R134a. 4.2 Performance of the DSHHPU in vapour compression heating mode In order to study the performance of the unit with different refrigerants in the vapour compression heating mode, the DSHHPU's evaporation pressure, compressor discharge temperature, heating capacity and COP were calculated when the ambient temperature is between -16 °C and 2 °C. The radiation intensity is 200, 400 and 600 W/m2, respectively, the cooling water inlet temperature is 40 °C, and the cooling water flow rate is 0.9 m3/h. The performance of the DSHHPU in vapour compression heating mode is affected by various operating parameters, where the operating pressure is a key influence parameter. Therefore, taking the evaporation pressure as an example, the operating pressure of each refrigerant under different working conditions was calculated. Figure 9 shows the variation of the evaporation pressure of the DSHHPU with R134a, R32, R290 and R600a. When the ambient temperature is raised from -16 °C to 2 °C, the average evaporation pressures of R134a, R32, R290, and R600a are increased by 83.71, 75.35, 75.65, and 94.63%, respectively, due to the improvement of heat transfer performance. When the solar radiation intensity is increased from 200 W/m2 to 600 W/m2, the average evaporation pressures of R134a, R32, R290 and R600a are increased by 7.49, 8.01, 7.04 and 8.05%, respectively. Due to the difference in standard evaporation temperatures, the average evaporation pressures of R134a, R32, R290, and R600a are 313.52, 468.51, 308.69, and 96.31 kPa, respectively. After comparison, it is found that the average evaporation pressure of R32 is 49.44% higher than that of R134a, and the average evaporation pressures of R290 and R600a are 1.54% and 69.28% lower than that of R134a, respectively.

11

700

600W/m2 400W/m2 200W/m2

400

650

R32 Evaporation pressure / kPa

R134a Evaporation pressure / kPa

450

350

300

250

200

600 550 500 450 400 350 300

-16

-14

-12

450

-10 -8 -6 -4 Outdoor temperature /℃

-2

0

2

140

600W/m2 400W/m2 200W/m2

400

-16

R600a Evaporation pressure / kPa

R290 Evaporation pressure / kPa

600W/m2 400W/m2 200W/m2

350

300

250

-14

-12

-10

-8

-6

-4

-2

0

-12

-10 -8 -6 -4 Outdoor temperature /℃

-2

0

2

-10

-2

0

2

600W/m2 400W/m2 200W/m2

120

100

200 -16

-14

80

60 -16

2

-14

-12

-8

-6

-4

Outdoor temperature /℃

Outdoor temperature /℃

Fig. 9 Evaporation pressure of the DSHHPU with four refrigerants

The compressor discharge temperature has an important influence on the performance of the unit, especially on the unit reliability. Therefore, the compressor discharge temperature of the unit with different refrigerants under different conditions is calculated. Figure 10 shows the variation of the compressor discharge temperature of the DSHHPU with four refrigerants. When the ambient temperature is raised from -16 °C to 2 °C, the average discharge temperatures of R134a, R32, R290, and R600a are decreased by 10.39, 17.33, 4.19, and 7.60%, respectively. When the solar radiation intensity is increased from 200 W/m2 to 600 W/m2, the average discharge temperatures of R134a, R32, R290 and R600a are decreased by 1.97, 2.72, 0.80, and 1.12%, respectively. The average discharge temperatures of R134a, R32, R290, and R600a are 86.13, 159.25, 62.86, and 57.23 °C, respectively. The discharge temperature is mainly determined by the adiabatic index and compression ratio of the refrigerant. Since the R32 has a higher adiabatic index and compression ratio than the other three refrigerants, the discharge temperature of R32 is much larger than those of the other three refrigerants. After comparison, it is found that the average discharge temperature of R32 is 84.88% higher than that of R134a. The average discharge temperatures of R600a and R290 are 33.56% and 27.02% lower than that of R134a, respectively.

12

175

R32 Discharge temperature /℃

R134a Discharge temperature /℃

92 90 88 86 84

600W/m2 2

400W/m 200W/m2

82 80

161 154

600W/m2 400W/m2 200W/m2

147 140

-16

-14

-12

-10 -8 -6 -4 Outdoor temperature /℃

-2

0

2

64.50

63.75

63.00

600W/m2 400W/m2 200W/m2

62.25

-16

-14

-12

62

R600a Discharge temperature /℃

R290 Discharge temperature /℃

168

-10 -8 -6 -4 Outdoor temperature /℃

-2

0

2

-10

-2

0

2

60

58

56

600W/m2 400W/m2 200W/m2

54

61.50 52 -16

-14

-12

-10

-8

-6

-4

-2

0

2

-16

Outdoor temperature /℃

-14

-12

-8

-6

-4

Outdoor temperature /℃

Fig. 10 Compressor discharge temperature of the DSHHPU with four refrigerants

Figure 11 shows the variation of heating capacity of the DSHHPU with four refrigerants in vapour compression heating mode. When the ambient temperature is raised from -16 °C to 2 °C, the average heating capacities of R134a, R32, R290, and R600a are increased by 22.12, 21.91, 22.72, and 32.22%, respectively. When the solar radiation intensity is increased from 200 W/m2 to 600 W/m2, the average heating capacities of R134a, R32, R290, and R600a are increased by 2.45, 2.66, 2.84, and 3.43%, respectively. The average heating capacities of R134a, R32, R290, and R600a are 11.95, 19.63, 11.18, and 5.61 kW, respectively. In the vapour compression heating mode, the heating capacity is proportional to the latent heat of vaporization and mass flow rate of the refrigerant. The mass flow rate in this mode is primarily determined by the refrigerant density at the compressor inlet. Since R32 and R134a have higher inlet densities, which make their mass flow rates greater than those of R290 and R600a, R32 and R134a have higher heating capacities than those of R290 and R600a. After comparison, it is found that the average heating capacity of R32 is 64.27% higher than that of R134a, and the average heating capacities of R290 and R600a are 6.42% and 53.01% lower than that of R134a, respectively.

13

13.5 22 2

2

600W/m

600W/m

2

2

400W/m

R32 Heating capacity / kW

R134a Heating capacity / kW

13.0

2

12.5

200W/m

12.0 11.5 11.0

400W/m

21

2

200W/m 20

19

18

10.5

17 -16

-14

-12

-10 -8 -6 -4 Outdoor temperature /℃

-2

0

2

-16

-14

-12

-10

-8

-6

-4

-2

0

2

-2

0

2

-2

0

2

-2

0

2

Outdoor temperature /℃

7.0

12.5 2

R600a Heating capacity / kW

R290 Heating capacity / kW

2

600W/m

12.0

2

400W/m

2

200W/m

11.5 11.0 10.5

600W/m

6.5

2

400W/m

2

200W/m 6.0

5.5

5.0

10.0 4.5 -16

-14

-12

-10

-8

-6

-4

-2

0

-16

2

-14

-12

-10

-8

-6

-4

Outdoor temperature /℃

Outdoor temperature /℃

Fig. 11 Heating capacity of the DSHHPU with four refrigerants 3.4 2.3

2

2

600W/m 3.2

600W/m

2

2

400W/m

2.2

2

200W/m

2

200W/m

3.0

R32 COP

R134a COP

400W/m

2.8

2.6

2.1

2.0

1.9

2.4

1.8 -16

-14

-12

-10

-8

-6

-4

-2

0

2

-16

Outdoor temperature /℃

-14

-12

-10

-8

-6

-4

Outdoor temperature /℃

4.2

3.8 4.0

2

600W/m 3.6

2

400W/m

2

600W/m

2

3.8

400W/m

2

3.4

200W/m

3.6

R600a COP

R290 COP

2

200W/m

3.2 3.0

3.4 3.2 3.0

2.8 2.8 2.6 2.6 -16

-14

-12

-10

-8

-6

-4

-2

0

-16

2

-14

-12

-10

-8

-6

-4

Outdoor temperature /℃

Outdoor temperature /℃

Fig. 12 COP of the DSHHPU with four refrigerants

Figure 12 shows the variation of COP of the DSHHPU with four refrigerants in vapour compression heating mode. When the ambient temperature is raised from -16 °C to 2 °C, the average COPs of R134a, R32, R290, and R600a are increased by 30.65, 23.45, 38.43, and 42.91%, respectively. When the solar radiation intensity is increased from 200 W/m2 to 600 W/m2, the average COPs of R134a, R32, R290, and 14

R600a are increased by 4.65, 3.96, 4.62, and 4.52%, respectively. The average COPs of R134a, R32, R290, and R600a are 2.85, 2.07, 3.18, and 3.36, respectively. Although R32 has a higher heating capacity than the other three refrigerants, its discharge temperature and compression ratio are relatively high, which makes its performance worse, so its COP is lower. After comparison, the average COP of R32 is 27.35% lower than that of R134a, while the average COPs of R600a and R290 are 17.94% and 11.59% higher than that of R134a, respectively.

5 Conclusion In order to compare the performance of the DSHHPU with different refrigerants, the mathematical model of the unit was established, and its performance was simulated and analyzed. Through analysis, the main conclusions are as follows: 1. In the heat pipe mode, the average evaporation pressure and average heating capacity of the unit are R32, R134a, R290 and R600a respectively from high to low order. Taking R134a as the comparison object, the average COP of R32 was 36.94% larger than that of R134a, while the average COP of R290 and R600a was 5.73% and 43.29% smaller than that of R134a, respectively. Besides, the increase of radiation intensity is helpful to improve the evaporation temperature and heating capacity of the system, so as to increase the COP of the system. 2. In the vapour compression heating mode, the average evaporation pressure, average discharge temperature and average heating capacity of the unit are R32, R134a, R290 and R600a respectively in the order from high to low. Taking R134a as the comparison object, the average COP of R32 was 27.35% lower than that of R134a, while the average COP of R290 and R600a was 17.94 and 11.59% higher than that of R134a, respectively. Besides, the increase of radiation intensity is helpful to improve the evaporation temperature and heating capacity of the system, reduce the discharge temperature, so as to increase the COP of the system.

Acknowledgements The authors gratefully acknowledge the support from the Natural Science Foundation of China (grant No.51778115) and the Fundamental Research Funds for the Central Universities (grant No.N182502043).

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