- Email: [email protected]
Experimental investigation on the influence of internal heat exchanger in a transcritical CO2 heat pump water heater
Journal Pre-proofs Experimental investigation on the influence of internal heat exchanger in a transcritical CO2 heat pump water heater Feng Cao, Zuliang Ye, Yikai Wang PII: DOI: Reference:
S1359-4311(19)34690-3 https://doi.org/10.1016/j.applthermaleng.2019.114855 ATE 114855
To appear in:
Applied Thermal Engineering
Received Date: Revised Date: Accepted Date:
8 July 2019 25 November 2019 26 December 2019
Please cite this article as: F. Cao, Z. Ye, Y. Wang, Experimental investigation on the influence of internal heat exchanger in a transcritical CO2 heat pump water heater, Applied Thermal Engineering (2019), doi: https:// doi.org/10.1016/j.applthermaleng.2019.114855
This is a PDF file of an article that has undergone enhancements after acceptance, such as the addition of a cover page and metadata, and formatting for readability, but it is not yet the definitive version of record. This version will undergo additional copyediting, typesetting and review before it is published in its final form, but we are providing this version to give early visibility of the article. Please note that, during the production process, errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.
© 2019 Published by Elsevier Ltd.
Experimental investigation on the influence of internal heat exchanger in a transcritical CO2 heat pump water heater Feng Cao*, Zuliang Ye, Yikai Wang, School of Energy and Power Engineering, Xi’an Jiaotong University, 28 Xianning West Road, Xi’an 710049, China * Corresponding author, Email: [email protected]; Tel: 86-29-82663583; Fax: 86-029-82663583 Abstract In this paper, a transcritical CO2 heat pump water heater prototype was experimentally investigated to research the influence of internal heat exchanger (IHX). In addition, the exergy analysis was carried out based on the experimental data. The results showed that the optimal discharge pressure reduced with the application of IHX, and this reduction could be amplified by the decrease of the ambient temperature, the increase of the water inlet temperature and the increase of the water outlet temperature. From the energy point of view, under the optimal discharge pressure, the coefficient of performance (COP) increased by up to 6.65% and the total power consumption decreased by up to 6.22% via using IHX. Then the influence of IHX on the heat pump cycle was discussed. The results of the exergy analysis showed that the exergy destruction of compressor, gas cooler and evaporator increased, and the exergy destruction of expansion valve decreased with the application of IHX. Furthermore, the product exergy and the total exergy destruction increased, and the exergy efficiency of system decreased in all working conditions. In conclusion, the application of IHX enhanced the energy performance and impaired the exergy efficiency in this research.
Keywords transcritical CO2 heat pump; internal heat exchanger; experimental investigation; exergy analysis Nomenclature Latin symbols cp
Isobaric specific heat [kJ·(kg·K)-1]
E
Exergy [kJ·kg-1]
h
Specific enthalpy [kJ·kg-1]
m
Mass flow rate [kg·s-1]
P
Pressure [MPa]
Q
Heating capacity [kW]
s
Specific entropy [kJ·(kg·K)-1]
T
Temperature [℃]
W
Power consumption [kW]
w
Specific work [kJ·kg-1]
Abbreviations COP
Coefficient of performance
IHX
Internal heat exchanger
Subscripts amb
Ambient
comp
Compressor
d
Compressor discharge
D
Destruction
elec
Electronic devices
exp
Expansion valve
evap
Evaporator
f
Fans
gc
Gas cooler
gout
Gas cooler outlet
H
Heat source
HS
High pressure side
in
Inlet
LS
Low pressure side
out
Outlet
p
Product
r
Refrigerant
w
Water
0
Exergy reference environment
1. Introduction Resulting from worldwide attention to the environmental protection, natural refrigerants have become the prevailing choices for many refrigeration and heat pump applications. As a natural refrigerant, carbon dioxide (CO2) has superior environmental friendliness, outstanding thermodynamic properties and excellent safety performance, such as incombustibility, nontoxicity and chemical stability. CO2 is considered as a perfect refrigerant for heat pump water heaters and automobile air conditioning systems. As a consequence of efficient heat transfer of CO2 in the supercritical region, hot water (above 65℃) can easily be produced by transcritical CO2 heat pumps. Moreover, the temperature glide that happens in water-refrigerant heat exchangers reduces the heat transfer temperature difference and the irreversible loss, which improves the efficiency of system. Much work has been carried out on the transcritical CO2 heat pump water heater to validate the competent performance and improve the energy efficiency. Nawaz et al.
[1] conducted modeling analysis on a CO2 heat pump water heater. They confirmed that the performance of the CO2 heat pump water heater was comparable to the performance of an R134a heat pump water heater. Wang et al. [2] researched a prototype experimentally and established correlations for the optimal discharge pressure as the function of the ambient temperature and the water outlet temperature. They suggested that their correlations were suitable for only their prototype but the empirical fitting method was feasible for similar systems. Qi et al. [3] indicated that the optimal discharge pressure was largely influenced by the gas cooler outlet temperature and negligibly influenced by the ambient temperature. Furthermore, they obtained a correlation for the optimal discharge pressure in terms of the gas cooler temperature. Saikawa and Koyama [4] calculated the COP upper limits of single stage compression heat pump cycles for tap water heating with various refrigerants such as fluorocarbons and natural refrigerants, and they noted that the highest COP was achieved by CO2. Liu et al. [5] compared the primary energy ratio of the CO2 heat pump with those of three traditional heating methods, and they discovered the superiority of CO2 system. According to the review of Zhang et al. [6], with the technological development of compressors and water-CO2 heat exchangers, the transcritical CO2 heat pump water heaters for residential use have been extensively applied in Japan. The internal heat exchanger (IHX) is widely used to generate subcooling and improve the performance of refrigeration and heat pump systems [7]. A number of authors have explored the application of IHX. Klein et al. [8] defined the heat exchanger effectiveness of IHX based on temperature differences and analyzed the impact of IHX
on refrigeration systems with different refrigerants. Chen and Gu [9] defined another heat exchanger effectiveness of IHX based on enthalpy differences and studied the influence of IHX on the optimal discharge pressure by simulation. Kim et al. [10] researched the effect of IHX length on a water-source transcritical CO2 heat pump water heater, and they found that the optimal discharge pressure, the refrigerant mass flow rate and the compressor power decreased with the increase of IHX length. Tao et al. [11] experimentally evaluated the performance of a residential transcritical CO2 airconditioning system with IHX. Fernandez et al. [12] performed full tank heating tests in three typical scenarios for residential water heating. They reported that the performance enhancement of the cycle with IHX was up to 7.9% for initial tank water heating and approximately 3.4% for warm tank water reheating. Torrella et al. [13] analyzed the difference between transcritical CO2 refrigeration systems with and without IHX, and they found that the impact of IHX on the compressor power consumption was slight under the same discharge pressure. Zhang et al. [14] investigated the efficiencies of subcritical and transcritical CO2 cycles with and without IHX, and they proposed the concepts of transition discharge pressure and transition gas cooler outlet temperature. Sánchez et al. [15] studied a CO2 refrigerating plant with different IHX configurations, and they pointed out that the improvements in the COP and the cooling capacity were observed regardless of the IHX position. Regarding the exergy analysis based on the second law of thermodynamics, many studies concerning CO2 system have likewise been published. Fartaj et al. [16] and Sarkar et al. [17] evaluated the contributions of the components to the irreversibility of
transcritical CO2 system. Dai et al. [18] investigated a transcritical CO2 heat pump system integrated with mechanical subcooling by utilizing energy, exergy and economic methodologies. Based on their results, the exergy efficiency was apparently improved comparing with traditional CO2 heat pump due to the reduction in throttling loss. Purjam and Goudarzi [19] introduced subcritical cascade CO2 heat pumps with expansion valves or expanders to address the limitation of extreme cold climates. The results showed that the cycles performed properly in both energy and exergy points of view. Shariatzadeh et al. [20] noted that the COP and the exergy efficiency were reduced by using IHX in the cycle with expander while the COP and the exergy efficiency were improved by using IHX in the cycle with throttling valve. Similarly, Ghazizade-Ahsaee et al. [21] found that a slight increase in the COP and the exergy efficiency was obtained by using an additional IHX in a horizontal direct-expansion geothermal heat pump cycle with expansion valve. Through the references above, it can be seen that the influence of IHX on transcritical CO2 systems has been researched by many authors [8-10, 12-15, 20, 21] from the energy perspective. Whereas, the majority of the reported papers conducted analyses in refrigeration application. Among the limited publications about heat pump application, the air source heat pump water heater was a less mentioned research object. But compared with water and geothermal source systems, the air source systems have advantages such as convenient installation, low initial cost and freedom from the restriction of geological and water conditions, and thus are widely used in various applications. Hence, it is worthwhile to carry out work about the influence of IHX, in
order to give instruction of using IHX in air source transcritical CO2 heat pump water heater. In terms of the exergy analysis, the previous literature [16-22] mostly investigated the contributions of different components to the exergy destruction in certain system, and did not consider the difference between system with and without IHX. Although the effect of IHX on system exergy efficiency was researched by Shariatzadeh et al. [20] and Ghazizade-Ahsaee et al. [21], their work did not aim at air source heat pump application as well. Moreover, all authors conducted exergy analysis based on the results from theoretical simulation. Therefore, the discussion on the basis of experimental data is infrequent and significant. In this context, the objective of this work is to provide analysis based on experimental data and clarify the energy and exergy influence of IHX on an air source transcritical CO2 heat pump water heater. Therefore, a heat pump prototype was was experimentally studied with and without IHX. The influence of IHX on the performance parameters and the optimal discharge pressure was investigated. Moreover, the exergy analysis was performed to investigate the influence of IHX on the exergy destruction and exergy efficiency of system. 2. Experimental setup description 2.1 Transcritical CO2 heat pump water heater prototype The schematic diagram of the transcritical CO2 heat pump water heater prototype is shown in Fig. 1. The positions of the temperature and pressure sensors are shown as well. The prototype fundamentally consisted of a semi-hermetic reciprocating
compressor, a spiral tube-in-tube gas cooler, an electronic expansion valve, a fin-tube evaporator, a spiral tube-in-tube IHX, a liquid-vapor separator and two manual ball valves. The gas cooler was made of three same heat exchange tubes in parallel. In the heat exchange tubes of the gas cooler, water flowed in the inside tubes and CO2 circulated in the annular gaps between the inside and the outside tubes. In the heat exchange tube of IHX, the low pressure refrigerant flowed in the inside tube and the high pressure refrigerant flowed in the annular gap. The evaporator was composed of two same heat exchangers which presented a “V” shape, and the evaporator was equipped with two axial fans. The two manual ball valves were used to switch the connection status of IHX. In the experiments, the discharge pressure under a certain working condition was controlled through regulating the opening of the electronic expansion valve. The detailed characteristics of the components are given in Table 1.
Fig. 1 Schematic diagram of the transcritical CO2 heat pump water heater prototype. Component
Characteristic
Compressor
Type: Semi-hermetic reciprocating Rated rotational speed: 1450 rpm Rated displacement: 11.69 m3·h-1 Gas cooler Type: Spiral tube-in-tube Flow type: Counter flow Outside tube: Φ28 mm×1.5 mm stainless steel tube Inside tube: Spiral copper tube Flow area ratio of inside tube to annular gap: 1.0 Total heat transfer area: 4.26 m2 Evaporator Type: Fin-tube Tube: Φ7 mm×0.7 mm copper tube Tube length: 1600 mm Surface of tube inside: Smooth Fin material: Aluminum Fins geometry: Wavy Fin spacing: 2.4 mm Fin thickness: 0.2 mm Number of rows per heat exchanger: 4 Number of copper tubes per row: 48 Number of refrigerant circuits per heat exchanger: 12 Internal heat exchanger Type: Spiral tube-in-tube Flow type: Counter flow Outside tube: Φ33 mm×1.5 mm stainless steel tube Inside tube: Spiral copper tube Flow area ratio of inside tube to annular gap: 1.2 Total heat transfer area: 0.27 m2 Electronic expansion valve Type: Stepper motor Axial fans Liquid-vapor separator
Rated power: 0.55 kW Rated rotational speed: 920 rpm Inside volume: 9.5 L
Table 1 Detailed characteristics of the components. 2.2 Environmental laboratory The environmental laboratory in this research consisted of five parts: water conditioning system, air conditioning system, electric control system, data acquisition system and environmental room. The ambient temperature could be controlled in a
range of -25℃ to 55℃, and the control precision was ±0.2℃. The water inlet temperature could be controlled in a range of 5℃ to 90℃ and the control precision was ±0.2℃. The water outlet temperature was controlled through regulating the variable frequency water pump by a PID controller in the water conditioning system. The type-T thermocouples and the thermal resistances (PT100) were adopted in the environmental laboratory. The measurement range of the temperature sensors was -200-350℃, and the accuracy was ±0.2℃. The CO2 pressure was measured by pressure transmitters with a measurement range of 0-15 MPa and an accuracy of ±2.5% of full scale. The water volumetric flow rate was measured by an electromagnetic flowmeter with a measurement range of 0-6 m3·h-1 and an accuracy of ±0.5% of full scale. The electric power consumption of the prototype was measured by an electric power analyzer with an accuracy of ±0.25% of reading. An Agilent HP34970A data acquisition unit was employed to collect the measurement data with a scan interval of 10 sec. To display and record data conveniently, a computer that installed BenchLink Data Logger 3 Software was used. Due to the measurement error, switching error and transducer conversion error, the accuracies of DC voltage, DC current, T-type thermocouple and thermal resistances were ±(0.0035% of reading + 0.0005% of range), ±(0.050% of reading + 0.005% of range), ± 1℃ and ± 0.06℃, respectively. Among them, the DC voltage and DC Current were respectively the outputs of pressure transmitters and electromagnetic flowmeter. 2.3 Performance parameters and error analysis
In this study, the heating capacity, the total power consumption and the coefficient of performance (COP) will be chosen to evaluate the performance of the prototype. The heating capacity Qgc is defined as
Qgc mwcpw Tw,out Tw,in
(1)
where mw is the water mass flow rate in kg·s-1; cpw is the isobaric specific heat in kJ·(kg·K)-1; Tw, out and Tw,in are the water outlet temperature and the water inlet temperature in ℃. The total power consumption W is defined as W Wcomp Wf Welec
(2)
where Wcomp , Wf and Welec are the power consumptions of the compressor, fans and electronic devices in kW, respectively. The COP is obtained by dividing the heating capacity by the total power consumption, as shown below.
COP =
Qgc
(3)
W
The measurement errors for the heating capacity and the COP can be calculated by using the error propagation method reported by Kline and McClintock as given below 1/2
n R wR ( wi ) 2 i 1 xi
(4)
where wR is the resultant uncertainty; R is the given function of the independent variable xi ; wi is the uncertainties of the independent variable xi . Using the above equation, the measurement errors for the heating capacity and the COP are 3.15% and
3.58%, respectively. 3. Exergy analysis model The exergy of a thermodynamic system is the maximum theoretical useful work [23]. But in practical systems, the exergy cannot be completely used, and the exergy destruction and loss are unavoidable. The exergy balance for a thermodynamic system: EF EP ED EL
(5)
Where EP denotes the product exergy which represents the desired result produced by system; EF denotes the fuel exergy which represents the resources expended to generate the product; ED denotes the exergy destruction due to irreversibilities within system; EL denotes the exergy loss engendered by interactions between system and environment. The exergy efficiency is a parameter that can provide a true measure of the performance of an energy system from the thermodynamic viewpoint [24]. The exergy efficiency E is the ratio between product and fuel exergy:
E
EP EP EF EP ED EL
(6)
Due to the wide-range application of thermodynamic systems, the definitions of product and fuel exergy are varied with the specific demand. In this research, the prototype was designed to convert the electric energy into thermal energy in water. Therefore, the exergy associated with the heat transfer in gas cooler was considered as the product, and the power input of compressor was considered as the fuel. Fig. 2 shows the combined system and boundary in the exergy analysis. The heat pump water heater prototype was a closed system, in which the heat and work
interactions between the system and the environment existed. By combining the prototype and part of the environment, the boundary of the combined system was located outside the prototype where the temperature corresponded to the ambient temperature. The ambient temperature was taken here as the temperature of the exergy reference environment. In this case, the boundary of the combined system allowed energy transfer across it by heat transfer with the water and by electrical work in the compressor. With the boundary mentioned above, neglecting the heat dissipation in other components, the gas cooler was the only component where heat interaction between the combined system and the environment existed. Because the exergy associated with the heat transfer in the gas cooler was defined as product, the exergy loss vanished accordingly. Then, the exergy destruction E D in equation (5) accounted for the total exergy destruction within the combined system.
Fig. 2 The combined system and boundary in the exergy analysis In this research, the product exergy of system E p is defined as
Ep
h
gc ,in
T hgc ,out 1 0 TH
(7)
where hgc ,in and hgc ,out are the specific enthalpy of CO2 at the inlet and outlet of the gas cooler; T0 is the temperature of the exergy reference environment, namely the ambient temperature; the heat source temperature TH is the thermodynamic average temperature of water and is defined as TH
hw,out hw,in sw,out sw,in
(8)
The exergy balance of the compressor can be written as
comp ,in wcomp comp ,out ED ,comp
(9)
where comp ,in and comp ,out are the specific flow exergy at the inlet and outlet of the compressor. The specific flow exergy
is calculated by
h h0 T（ 0 s s0）
(10)
Considering the compression as adiabatic, the specific work of compressor wcomp can be calculated by wcomp hcomp ,out hcomp ,in
(11)
By substituting equation (10) and (11) into equation (9), the exergy destruction of the compressor E D ,comp is expressed as E D ,comp T0 scomp ,out scomp ,in
(12)
where scomp ,in and scomp ,out are the specific entropy of CO2 at the inlet and outlet of the compressor. Similarly, the exergy destruction of other components can be deduced. The exergy destruction of the gas cooler E D , gc is defined as E D , gc hgc ,in hgc ,out T0 sgc ,in sgc ,out E p
(13)
where s gc ,in and s gc ,out are the specific entropy of CO2 at the inlet and outlet of the gas cooler. The exergy destruction of the expansion valve E D ,exp is defined as E D ,exp hexp,in hexp,out T0 sexp,in sexp,out
(14)
where hexp,in and hexp,out are the specific enthalpy of CO2 at the inlet and outlet of the expansion valve; sexp,in and sexp,out are the specific entropy of CO2 at the inlet and outlet of the expansion valve. The exergy destruction of the evaporator E D ,evap is defined as
ED,evap hevap,in hevap,out T0 sevap,in sevap,out
(15)
where hevap ,in and hevap ,out are the specific enthalpy of CO2 at the inlet and outlet of the evaporator; sevap ,in and sevap ,out are the specific entropy of CO2 at the inlet and outlet of the evaporator. The exergy destruction of the IHX E D , IHX is defined as
ED , IHX T0 sIHX ,out ,HS sIHX ,in ,HS sIHX ,out ,LS sIHX ,in ,LS
(16)
where s IHX ,in , HS , s IHX ,out , HS , sIHX ,in ,LS and sIHX ,out ,LS are the specific entropy of CO2 at the high pressure side inlet, the high pressure side outlet, the low pressure side inlet and the low pressure side outlet of the IHX, respectively. The total exergy destruction of system E D ,total is defined as E D ,total E D ,comp E D , gc E D ,exp E D ,evap E D , IHX
(17)
4. Results and discussion An air-source heat pump water heater absorbs heat from the ambient air and releases heat to the water. The affecting factors on the system performance are mainly
the ambient temperature, the water inlet temperature and the water outlet temperature. In this research, the working condition of the prototype was determined by the ambient temperature, the water inlet temperature and the water outlet temperature. 4.1 Influence of IHX on the performance parameters and the optimal discharge pressure Fig. 3 shows the variations of the performance parameters with variable discharge pressure under different ambient temperature. As shown in Fig. 3, the heating capacity and the total power consumption increased with the increase of discharge pressure. From Fig. 3(c), it could be noted that the maximum COP occurred whether or not the IHX was connected. The growth of the heating capacity was higher than that of the total power consumption when the discharge pressure was relatively small, and it led to the increase of the COP. Conversely, the COP decreased when the discharge pressure was high, because the growth of the total power consumption was higher than that of the heating capacity. It could also be found from Fig. 3(c) that the maximum COP was improved by using IHX under the same working condition. Moreover, the optimal discharge pressure corresponding to the maximum COP was reduced by the application of IHX.
(a)
(b)
(c) Fig. 3 Variations of the performance parameters with variable discharge pressure under different ambient temperature. (a) Heating capacity. (b) Total power consumption. (c) COP. Fig. 4 shows the variations of the performance parameters with variable discharge pressure under different water inlet temperature. The heating capacity increased consistently with the increase of the discharge pressure when the water inlet temperature was 40℃. However, when the water inlet temperature was 10℃, the heating capacity first increased and then decreased with the increase of the discharge pressure. Therefore, as shown in Fig. 4(c), the COP decreased more rapidly with the increase of the discharge pressure when the water inlet temperature was 10℃ in comparison with that when the water inlet temperature was 40℃. Whether the IHX was applied, the optimal discharge pressure increased when the water inlet temperature changed from 10℃ to 40℃.
（a）
（b）
(c) Fig. 4 Variations of the performance parameters with variable discharge pressure under different water inlet temperature. (a) Heating capacity. (b) Total power consumption. (c) COP. Fig. 5 shows the variations of the performance parameters with variable discharge pressure under different water outlet temperature. The optimal discharge pressure increased when the water outlet temperature changed from 70℃ to 90℃.
(a)
(b)
(c) Fig. 5 Variations of the performance parameters with variable discharge pressure under different water outlet temperature. (a) Heating capacity. (b) Total power consumption. (c) COP. Table 2 summarizes the optimal discharge pressure and the performance parameters under the optimal discharge pressure in several working conditions. As Table 2 shows, the optimal discharge pressure decreased with the application of IHX in all working conditions, and the highest decreasing rate could be up to 6.92%. As for the performance parameters, under the optimal discharge pressure, the total power consumption decreased by up to 6.22%; the COP increased by up to 6.65%; but there is no obvious trend in the variations of the heating capacity.
Moreover, the reduction in the optimal discharge pressure caused by using IHX was influenced by the ambient temperature, the water inlet temperature and the water outlet temperature. It can be seen from Table 2 that the reduction in the optimal discharge pressure could be enlarged by the decrease of the ambient temperature, the increase of the water inlet temperature and the increase of the water outlet temperature. Optimal discharge pressure [MPa]
Heating capacity [kW]
Total power consumption [kW]
COP [-]
Without IHX
10.18
71.90
19.10
3.76
With IHX
9.90
70.16
18.42
3.81
Variation
-2.75%
-2.41%
-3.53%
1.15%
Without IHX
11.76
60.36
21.91
2.76
With IHX
11.41
60.50
20.73
2.92
Variation
-2.98%
0.23%
-5.40%
5.95%
Without IHX
11.40
50.42
20.77
2.43
With IHX
10.81
50.59
19.54
2.59
Variation
-5.18%
0.33%
-5.92%
6.65%
Without IHX
9.81
53.79
17.91
3.00
With IHX
9.52
54.99
17.46
3.15
Variation
-2.96%
2.24%
-2.53%
4.89%
Without IHX
11.89
55.83
20.30
2.75
With IHX
11.25
55.85
19.19
2.91
Variation
-5.38%
0.05%
-5.44%
5.81%
Without IHX
11.13
44.84
20.01
2.24
With IHX
10.36
44.36
18.76
2.36
Variation
-6.92%
-1.06%
-6.22%
5.51%
Working condition Tamb [℃]
16
16
7
2
2
2
Status of IHX Tw,in [℃]Tw,out [℃]
10
40
40
10
10
40
70
70
70
70
90
70
Table 2 Optimal discharge pressure and the performance parameters under the optimal discharge pressure in several working conditions. 4.2 Influence of IHX on the transcritical CO2 heat pump cycle In this section, the influence of IHX on the transcritical CO2 heat pump cycle was discussed based on experimental data. Fig. 6 depicts the transcritical CO2 heat pump
cycles with and without IHX under the optimal discharge pressure in the P-H diagram, in which 1-2-3-4-1 represents the cycle without IHX and 1’-2’-3’-4’-5’-6’-1’ represents the cycle with IHX. Because the quality of refrigerant at the inlet of the evaporator cannot be determined by measuring temperature and pressure, we assumed that the throttling in the expansion valve was isenthalpic.
Fig. 6 Transcritical CO2 heat pump cycles with and without IHX under the optimal discharge pressure in the P-H diagram. The heat pump cycles can be divided into the following processes: Process 1-2 (1’-2’): Compression in the compressor Process 2-3 (2’-3’): Heat release without phase change in the gas cooler Process 3-4 (4’-5’): Throttling process in the expansion valve Process 4-1 (5’-6’): Heat absorb with phase change in the evaporator
Process 3’-4’ and 6’-1’: Heat transfer in the IHX The influence of IHX on the cycle pressure mainly occurred on the high pressure side. As shown in Fig. 6, the gas cooler pressure in the cycles with IHX was significantly lower when compared with that in the cycles without IHX. In terms of the low pressure side, it can be seen from the figure that the evaporator pressure was barely affected by using IHX. Due to the heat transfer in the IHX, the cycle temperature was influenced. Firstly, the suction temperature and discharge temperature of the compressor in the cycles with IHX were higher when compared with those in the cycles without IHX. However, the coking of lubricating oil could occur when the discharge temperature is extremely high. In a heat pump cycle with IHX, the discharge temperature increases when the ambient temperature decreases, because of the reduction in evaporation temperature and the increase of compressor pressure ratio. Besides, the increase of the discharge temperature was larger when the water inlet temperature was higher. Hence, in order to avoid the adverse effect caused by extremely high discharge temperature, careful consideration should be made when the IHX is applied in the working conditions with low ambient temperature and high water inlet temperature. For instance, the heat exchanger effectiveness of IHX should be maintained under a low level. Secondly, the refrigerant status at the outlet of the evaporator was marginally influenced by using IHX. This can be evidenced by the situations of point 1 and 6’. As a consequence of the subcooling and the superheat generated by IHX, the evaporator inlet enthalpy decreased, and the suction temperature increased. Then the refrigerant
mass flow rate reduced with the increase of the suction temperature. In this case, under the constant heat transfer area of the evaporator, the enthalpy difference between the inlet and outlet of the evaporator increased with the decrease of the refrigerant mass flow rate. Therefore, the refrigerant status at the outlet of the evaporator had no significant change as a net result. Thirdly, the gas cooler outlet temperature was influenced by using IHX. When the water inlet temperature was 10℃, the gas cooler outlet temperature slightly decreased with the application of IHX. However, when the water inlet temperature was 40℃, there was little change in the gas cooler outlet temperature. The reason can be explained by the heat transfer area of the gas cooler. The heat transfer area was suitable when the water inlet temperature was low, and surplus when the water inlet temperature was high, because the gas cooler in the prototype was basically designed for tap water heating. Hence, when the water inlet temperature was 10℃, the decrease of the refrigerant mass flow rate caused by using IHX could make the heat release more complete and reduce the temperature approach at the outlet of gas cooler. Owing to the influence of IHX on the cycle temperature, as shown in Fig. 6, the enthalpy differences between the inlet and outlet of the compressor and between the inlet and outlet of the gas cooler increased. In general, the heating capacity and the compressor power consumption were increased by the increase of the enthalpy differences, and decreased by the decrease of the refrigerant mass flow rate. Also, the compressor power consumption decreased with the decrease of the optimal discharge pressure. Therefore, the variations on the heating capacity and the compressor power
consumption were the net results of the above multiple effects. 4.3 Influence of IHX on the system exergy In this section, the exergy analysis was conducted to analyze the influence of IHX based on the heat pump cycles under the optimal discharge pressure. Table 3 summarizes the exergy destruction of the components and heat pump system in different working conditions. As shown in Table 3, the total exergy destruction increased by using IHX. Regarding the exergy destruction of the components, by the application of IHX, the exergy destruction of the compressor, the gas cooler and the evaporator increased, the exergy destruction of the expansion valve decreased. Exergy destruction [kJ·kg-1] Tamb/Tw,in/Tw,out Tamb/Tw,in/Tw,out Tamb/Tw,in/Tw,out Tamb/Tw,in/Tw,out Tamb/Tw,in/Tw,out
Component
=16/10/70℃
=16/40/70℃
=2/10/70℃
=2/40/70℃
=2/10/90℃
Without With Without With Without With Without With Without With IHX
IHX
IHX
IHX
IHX
IHX
IHX
IHX
IHX
IHX
Compressor
1.71
4.72
4.52
8.31
3.31
11.89
7.54
12.67
7.06
9.31
Gas cooler
7.70
9.12
5.75
9.99
7.55
10.44
5.78
11.22
7.41
9.94
Expansion valve
7.31
6.29
13.57
10.94
8.52
6.77
18.07
13.74
10.13
8.79
Evaporator
8.90
10.12
5.14
5.66
6.78
7.87
3.57
3.58
7.40
7.85
IHX
-
0.54
-
1.87
-
0.88
-
3.64
-
0.88
Total
25.62
30.78
28.98
36.77
26.17
37.84
34.96
44.85
32.00
36.77
Table 3 Exergy destruction of the components and the heat pump system in different working conditions. Fig. 7 shows the exergy destruction percentages of the components in different working conditions. By using IHX, the exergy destruction percentage of the compressor
increased; the exergy destruction percentage of the gas cooler increased when the water inlet temperature was 40℃ or the water outlet temperature was 90℃; but the exergy destruction percentage of the gas cooler slightly decreased when the water inlet temperature and the water outlet temperature were 10℃ and 70℃, respectively; the exergy destruction percentage of the expansion valve and evaporator decreased; with regard to the IHX, the exergy destruction percentage of IHX was significantly higher when the water inlet temperature was 40℃ than that when the water inlet temperature was 10℃.
Fig. 7 Exergy destruction percentages of the components in different working conditions. Fig. 8 shows the product exergy, the total exergy destruction and the exergy efficiency of system in different working conditions. It can be seen that the product exergy and the total exergy destruction increased with the application of IHX. Moreover, the exergy efficiency of system decreased by using IHX in all working conditions.
Fig. 8 Product exergy, total exergy destruction and exergy efficiency of system in different working conditions. 5. Conclusions In this paper, the influence of IHX on the air source transcritical CO2 heat pump water heater prototype was experimentally investigated from the energy and exergy viewpoints. The results showed that the optimal discharge pressure reduced with the application of IHX, and this reduction could be enlarged by the decrease of the ambient temperature, the increase of the water inlet temperature and the increase of the water outlet temperature. Under the optimal discharge pressure, the total power consumption decreased by up to 6.22% and the COP increased by up to 6.65% by using IHX. The evaporator pressure in heat pump cycles under the optimal discharge pressure was barely affected by using IHX, and the refrigerant status at the outlet of evaporator
was nearly constant. With the application of IHX, the suction temperature and the discharge temperature of the compressor increased. When the water inlet temperature was 10℃, the gas cooler outlet temperature decreased slightly by using IHX. However, when the water inlet temperature was 40℃, there was little change in the gas cooler outlet temperature. By using IHX, the exergy destruction of compressor, gas cooler and evaporator increased and the exergy destruction of expansion valve decreased. The total exergy destruction and the product exergy increased with the application of IHX, and the exergy efficiency of system reduced in all working conditions. Through this research, the improvement in the energy performance that was attributed to using IHX was verified in different working conditions. Besides, with the application of IHX, the reliability of the heat pump system might also be enhanced by the reduced optimal discharge pressure. Hence, using IHX could be recommended in case that the increased discharge temperature does not exceed the upper limit value associated with compressor and lubricating oil. Actually, in terms of the studied prototype, when the ambient temperature was higher than 2℃, the IHX could be unconditionally applied and the discharge temperature did not exceed 140℃ which was the recommended upper limit from the manufacturer. This result could be used as a reference for similar systems. Notably, the reduction in the system exergy efficiency caused by using IHX was also discovered, which was different from the theoretical results in previous publications [20, 21]. This difference might result from the refrigerant pressure drop in
the heat exchangers, change of compressor efficiency with increased suction and discharge temperature, etc. Considering the difficulties in studying these factors experimentally, the detailed theoretical model should be developed in further study, and then the analysis about potential influential parameters could be conducted based on the model. Acknowledgement The research presented was supported by the National Natural Science Foundations of China (51976153) and the Fundamental Research Funds for the Central Universities. References [1] K. Nawaz, B. Shen, A. Elatar, V. Baxter, O. Abdelaziz, Performance optimization of CO2 heat pump water heater, International Journal of Refrigeration, 85 (2018) 213228. [2] S. Wang, H. Tuo, F. Cao, Z. Xing, Experimental investigation on air-source transcritical CO2 heat pump water heater system at a fixed water inlet temperature, International Journal of Refrigeration, 36 (2013) 701-716. [3] P.-C. Qi, Y.-L. He, X.-L. Wang, X.-Z. Meng, Experimental investigation of the optimal heat rejection pressure for a transcritical CO2 heat pump water heater, Applied Thermal Engineering, 56 (2013) 120-125. [4] M. Saikawa, S. Koyama, Thermodynamic analysis of vapor compression heat pump cycle for tap water heating and development of CO2 heat pump water heater for residential use, Applied Thermal Engineering, 106 (2016) 1236-1243. [5] S. Liu, Z. Li, B. Dai, Z. Zhong, H. Li, M. Song, Z. Sun, Energetic, economic and environmental analysis of air source transcritical CO2 heat pump system for residential heating in China, Applied Thermal Engineering, 148 (2019) 1425-1439. [6] J.-F. Zhang, Y. Qin, C.-C. Wang, Review on CO2 heat pump water heater for residential use in Japan, Renewable and Sustainable Energy Reviews, 50 (2015) 13831391. [7] R. Llopis, L. Nebot-Andrés, D. Sánchez, J. Catalán-Gil, R. Cabello, Subcooling methods for CO2 refrigeration cycles: A review, International Journal of Refrigeration, 93 (2018) 85-107. [8] S.A. Klein , D.T. Reindl, K. Brownell, Refrigeration system performance using liquid-suction heat exchangers, International Journal of Refrigeration, (2000). [9] Y. Chen, J. Gu, The optimum high pressure for CO2 transcritical refrigeration
systems with internal heat exchangers, International Journal of Refrigeration, 28 (2005) 1238-1249. [10] S.G. Kim, Y.J. Kim, G. Lee, M.S. Kim, The performance of a transcritical CO2 cycle with an internal heat exchanger for hot water heating, International Journal of Refrigeration, 28 (2005) 1064-1072. [11] Y.B. Tao, Y.L. He, W.Q. Tao, Z.G. Wu, Experimental study on the performance of CO2 residential air-conditioning system with an internal heat exchanger, Energy Conversion and Management, 51 (2010) 64-70. [12] N. Fernandez, Y. Hwang, R. Radermacher, Comparison of CO2 heat pump water heater performance with baseline cycle and two high COP cycles, International Journal of Refrigeration, 33 (2010) 635-644. [13] E. Torrella, D. Sánchez, R. Llopis, R. Cabello, Energetic evaluation of an internal heat exchanger in a CO2 transcritical refrigeration plant using experimental data, International Journal of Refrigeration, 34 (2011) 40-49. [14] F.Z. Zhang, P.X. Jiang, Y.S. Lin, Y.W. Zhang, Efficiencies of subcritical and transcritical CO2 inverse cycles with and without an internal heat exchanger, Applied Thermal Engineering, 31 (2011) 432-438. [15] D. Sánchez, J. Patiño, R. Llopis, R. Cabello, E. Torrella, F.V. Fuentes, New positions for an internal heat exchanger in a CO2 supercritical refrigeration plant. Experimental analysis and energetic evaluation, Applied Thermal Engineering, 63 (2014) 129-139. [16] A. Fartaj, D.S.K. Ting, W.W. Yang, Second law analysis of the transcritical CO2 refrigeration cycle, Energy Conversion and Management, 45 (2004) 2269-2281. [17] J. Sarkar, S. Bhattacharyya, M. Ram Gopal, Transcritical CO2 heat pump systems: exergy analysis including heat transfer and fluid flow effects, Energy Conversion and Management, 46 (2005) 2053-2067. [18] B. Dai, H. Qi, S. Liu, M. Ma, Z. Zhong, H. Li, M. Song, Z. Sun, Evaluation of transcritical CO2 heat pump system integrated with mechanical subcooling by utilizing energy, exergy and economic methodologies for residential heating, Energy Conversion and Management, 192 (2019) 202-220. [19] M. Purjam, K. Goudarzi, High efficiency sub-critical carbon dioxide supplementary heat pump for low temperature climates (energy and exergy analysis), Renewable Energy, 133 (2019) 166-176. [20] O. Joneydi Shariatzadeh, S.S. Abolhassani, M. Rahmani, M. Ziaee Nejad, Comparison of transcritical CO2 refrigeration cycle with expander and throttling valve including/excluding internal heat exchanger: Exergy and energy points of view, Applied Thermal Engineering, 93 (2016) 779-787. [21] H. Ghazizade-Ahsaee, M. Ameri, Effects of using expander and internal heat exchanger on carbon dioxide direct-expansion geothermal heat pump, Applied Thermal Engineering, 136 (2018) 389-407. [22] D. Wang, Y. Liu, Z. Kou, L. Yao, Y. Lu, L. Tao, P. Xia, Energy and exergy analysis of an air-source heat pump water heater system using CO2/R170 mixture as an azeotropy refrigerant for sustainable development, International Journal of Refrigeration, 106 (2019) 628-638.
[23] G. Tsatsaronis, Definitions and nomenclature in exergy analysis and exergoeconomics, Energy, 32 (2007) 249-253. [24] A. Bejan, G. Tsatsaronis, M. Moran, Thermal design and optimization, Wiley, New York (1996).
Highlights
•
The energy and exergy influence of internal heat exchanger was studied.
•
Influence of internal heat exchanger on optimal discharge pressure was presented.
•
Influence of internal heat exchanger on system cycle was discussed in P-H diagram.
•
Exergy destruction and efficiency was analyzed based on the experimental data.
Declaration of interests
☒ The authors declare that they have no known competing financial interests or personal relationships that could have appeared to influence the work reported in this paper.
☐The authors declare the following financial interests/personal relationships which may be considered as potential competing interests:
S1359-4311(19)34690-3 https://doi.org/10.1016/j.applthermaleng.2019.114855 ATE 114855
To appear in:
Applied Thermal Engineering
Received Date: Revised Date: Accepted Date:
8 July 2019 25 November 2019 26 December 2019
Please cite this article as: F. Cao, Z. Ye, Y. Wang, Experimental investigation on the influence of internal heat exchanger in a transcritical CO2 heat pump water heater, Applied Thermal Engineering (2019), doi: https:// doi.org/10.1016/j.applthermaleng.2019.114855
This is a PDF file of an article that has undergone enhancements after acceptance, such as the addition of a cover page and metadata, and formatting for readability, but it is not yet the definitive version of record. This version will undergo additional copyediting, typesetting and review before it is published in its final form, but we are providing this version to give early visibility of the article. Please note that, during the production process, errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.
© 2019 Published by Elsevier Ltd.
Experimental investigation on the influence of internal heat exchanger in a transcritical CO2 heat pump water heater Feng Cao*, Zuliang Ye, Yikai Wang, School of Energy and Power Engineering, Xi’an Jiaotong University, 28 Xianning West Road, Xi’an 710049, China * Corresponding author, Email: [email protected]; Tel: 86-29-82663583; Fax: 86-029-82663583 Abstract In this paper, a transcritical CO2 heat pump water heater prototype was experimentally investigated to research the influence of internal heat exchanger (IHX). In addition, the exergy analysis was carried out based on the experimental data. The results showed that the optimal discharge pressure reduced with the application of IHX, and this reduction could be amplified by the decrease of the ambient temperature, the increase of the water inlet temperature and the increase of the water outlet temperature. From the energy point of view, under the optimal discharge pressure, the coefficient of performance (COP) increased by up to 6.65% and the total power consumption decreased by up to 6.22% via using IHX. Then the influence of IHX on the heat pump cycle was discussed. The results of the exergy analysis showed that the exergy destruction of compressor, gas cooler and evaporator increased, and the exergy destruction of expansion valve decreased with the application of IHX. Furthermore, the product exergy and the total exergy destruction increased, and the exergy efficiency of system decreased in all working conditions. In conclusion, the application of IHX enhanced the energy performance and impaired the exergy efficiency in this research.
Keywords transcritical CO2 heat pump; internal heat exchanger; experimental investigation; exergy analysis Nomenclature Latin symbols cp
Isobaric specific heat [kJ·(kg·K)-1]
E
Exergy [kJ·kg-1]
h
Specific enthalpy [kJ·kg-1]
m
Mass flow rate [kg·s-1]
P
Pressure [MPa]
Q
Heating capacity [kW]
s
Specific entropy [kJ·(kg·K)-1]
T
Temperature [℃]
W
Power consumption [kW]
w
Specific work [kJ·kg-1]
Abbreviations COP
Coefficient of performance
IHX
Internal heat exchanger
Subscripts amb
Ambient
comp
Compressor
d
Compressor discharge
D
Destruction
elec
Electronic devices
exp
Expansion valve
evap
Evaporator
f
Fans
gc
Gas cooler
gout
Gas cooler outlet
H
Heat source
HS
High pressure side
in
Inlet
LS
Low pressure side
out
Outlet
p
Product
r
Refrigerant
w
Water
0
Exergy reference environment
1. Introduction Resulting from worldwide attention to the environmental protection, natural refrigerants have become the prevailing choices for many refrigeration and heat pump applications. As a natural refrigerant, carbon dioxide (CO2) has superior environmental friendliness, outstanding thermodynamic properties and excellent safety performance, such as incombustibility, nontoxicity and chemical stability. CO2 is considered as a perfect refrigerant for heat pump water heaters and automobile air conditioning systems. As a consequence of efficient heat transfer of CO2 in the supercritical region, hot water (above 65℃) can easily be produced by transcritical CO2 heat pumps. Moreover, the temperature glide that happens in water-refrigerant heat exchangers reduces the heat transfer temperature difference and the irreversible loss, which improves the efficiency of system. Much work has been carried out on the transcritical CO2 heat pump water heater to validate the competent performance and improve the energy efficiency. Nawaz et al.
[1] conducted modeling analysis on a CO2 heat pump water heater. They confirmed that the performance of the CO2 heat pump water heater was comparable to the performance of an R134a heat pump water heater. Wang et al. [2] researched a prototype experimentally and established correlations for the optimal discharge pressure as the function of the ambient temperature and the water outlet temperature. They suggested that their correlations were suitable for only their prototype but the empirical fitting method was feasible for similar systems. Qi et al. [3] indicated that the optimal discharge pressure was largely influenced by the gas cooler outlet temperature and negligibly influenced by the ambient temperature. Furthermore, they obtained a correlation for the optimal discharge pressure in terms of the gas cooler temperature. Saikawa and Koyama [4] calculated the COP upper limits of single stage compression heat pump cycles for tap water heating with various refrigerants such as fluorocarbons and natural refrigerants, and they noted that the highest COP was achieved by CO2. Liu et al. [5] compared the primary energy ratio of the CO2 heat pump with those of three traditional heating methods, and they discovered the superiority of CO2 system. According to the review of Zhang et al. [6], with the technological development of compressors and water-CO2 heat exchangers, the transcritical CO2 heat pump water heaters for residential use have been extensively applied in Japan. The internal heat exchanger (IHX) is widely used to generate subcooling and improve the performance of refrigeration and heat pump systems [7]. A number of authors have explored the application of IHX. Klein et al. [8] defined the heat exchanger effectiveness of IHX based on temperature differences and analyzed the impact of IHX
on refrigeration systems with different refrigerants. Chen and Gu [9] defined another heat exchanger effectiveness of IHX based on enthalpy differences and studied the influence of IHX on the optimal discharge pressure by simulation. Kim et al. [10] researched the effect of IHX length on a water-source transcritical CO2 heat pump water heater, and they found that the optimal discharge pressure, the refrigerant mass flow rate and the compressor power decreased with the increase of IHX length. Tao et al. [11] experimentally evaluated the performance of a residential transcritical CO2 airconditioning system with IHX. Fernandez et al. [12] performed full tank heating tests in three typical scenarios for residential water heating. They reported that the performance enhancement of the cycle with IHX was up to 7.9% for initial tank water heating and approximately 3.4% for warm tank water reheating. Torrella et al. [13] analyzed the difference between transcritical CO2 refrigeration systems with and without IHX, and they found that the impact of IHX on the compressor power consumption was slight under the same discharge pressure. Zhang et al. [14] investigated the efficiencies of subcritical and transcritical CO2 cycles with and without IHX, and they proposed the concepts of transition discharge pressure and transition gas cooler outlet temperature. Sánchez et al. [15] studied a CO2 refrigerating plant with different IHX configurations, and they pointed out that the improvements in the COP and the cooling capacity were observed regardless of the IHX position. Regarding the exergy analysis based on the second law of thermodynamics, many studies concerning CO2 system have likewise been published. Fartaj et al. [16] and Sarkar et al. [17] evaluated the contributions of the components to the irreversibility of
transcritical CO2 system. Dai et al. [18] investigated a transcritical CO2 heat pump system integrated with mechanical subcooling by utilizing energy, exergy and economic methodologies. Based on their results, the exergy efficiency was apparently improved comparing with traditional CO2 heat pump due to the reduction in throttling loss. Purjam and Goudarzi [19] introduced subcritical cascade CO2 heat pumps with expansion valves or expanders to address the limitation of extreme cold climates. The results showed that the cycles performed properly in both energy and exergy points of view. Shariatzadeh et al. [20] noted that the COP and the exergy efficiency were reduced by using IHX in the cycle with expander while the COP and the exergy efficiency were improved by using IHX in the cycle with throttling valve. Similarly, Ghazizade-Ahsaee et al. [21] found that a slight increase in the COP and the exergy efficiency was obtained by using an additional IHX in a horizontal direct-expansion geothermal heat pump cycle with expansion valve. Through the references above, it can be seen that the influence of IHX on transcritical CO2 systems has been researched by many authors [8-10, 12-15, 20, 21] from the energy perspective. Whereas, the majority of the reported papers conducted analyses in refrigeration application. Among the limited publications about heat pump application, the air source heat pump water heater was a less mentioned research object. But compared with water and geothermal source systems, the air source systems have advantages such as convenient installation, low initial cost and freedom from the restriction of geological and water conditions, and thus are widely used in various applications. Hence, it is worthwhile to carry out work about the influence of IHX, in
order to give instruction of using IHX in air source transcritical CO2 heat pump water heater. In terms of the exergy analysis, the previous literature [16-22] mostly investigated the contributions of different components to the exergy destruction in certain system, and did not consider the difference between system with and without IHX. Although the effect of IHX on system exergy efficiency was researched by Shariatzadeh et al. [20] and Ghazizade-Ahsaee et al. [21], their work did not aim at air source heat pump application as well. Moreover, all authors conducted exergy analysis based on the results from theoretical simulation. Therefore, the discussion on the basis of experimental data is infrequent and significant. In this context, the objective of this work is to provide analysis based on experimental data and clarify the energy and exergy influence of IHX on an air source transcritical CO2 heat pump water heater. Therefore, a heat pump prototype was was experimentally studied with and without IHX. The influence of IHX on the performance parameters and the optimal discharge pressure was investigated. Moreover, the exergy analysis was performed to investigate the influence of IHX on the exergy destruction and exergy efficiency of system. 2. Experimental setup description 2.1 Transcritical CO2 heat pump water heater prototype The schematic diagram of the transcritical CO2 heat pump water heater prototype is shown in Fig. 1. The positions of the temperature and pressure sensors are shown as well. The prototype fundamentally consisted of a semi-hermetic reciprocating
compressor, a spiral tube-in-tube gas cooler, an electronic expansion valve, a fin-tube evaporator, a spiral tube-in-tube IHX, a liquid-vapor separator and two manual ball valves. The gas cooler was made of three same heat exchange tubes in parallel. In the heat exchange tubes of the gas cooler, water flowed in the inside tubes and CO2 circulated in the annular gaps between the inside and the outside tubes. In the heat exchange tube of IHX, the low pressure refrigerant flowed in the inside tube and the high pressure refrigerant flowed in the annular gap. The evaporator was composed of two same heat exchangers which presented a “V” shape, and the evaporator was equipped with two axial fans. The two manual ball valves were used to switch the connection status of IHX. In the experiments, the discharge pressure under a certain working condition was controlled through regulating the opening of the electronic expansion valve. The detailed characteristics of the components are given in Table 1.
Fig. 1 Schematic diagram of the transcritical CO2 heat pump water heater prototype. Component
Characteristic
Compressor
Type: Semi-hermetic reciprocating Rated rotational speed: 1450 rpm Rated displacement: 11.69 m3·h-1 Gas cooler Type: Spiral tube-in-tube Flow type: Counter flow Outside tube: Φ28 mm×1.5 mm stainless steel tube Inside tube: Spiral copper tube Flow area ratio of inside tube to annular gap: 1.0 Total heat transfer area: 4.26 m2 Evaporator Type: Fin-tube Tube: Φ7 mm×0.7 mm copper tube Tube length: 1600 mm Surface of tube inside: Smooth Fin material: Aluminum Fins geometry: Wavy Fin spacing: 2.4 mm Fin thickness: 0.2 mm Number of rows per heat exchanger: 4 Number of copper tubes per row: 48 Number of refrigerant circuits per heat exchanger: 12 Internal heat exchanger Type: Spiral tube-in-tube Flow type: Counter flow Outside tube: Φ33 mm×1.5 mm stainless steel tube Inside tube: Spiral copper tube Flow area ratio of inside tube to annular gap: 1.2 Total heat transfer area: 0.27 m2 Electronic expansion valve Type: Stepper motor Axial fans Liquid-vapor separator
Rated power: 0.55 kW Rated rotational speed: 920 rpm Inside volume: 9.5 L
Table 1 Detailed characteristics of the components. 2.2 Environmental laboratory The environmental laboratory in this research consisted of five parts: water conditioning system, air conditioning system, electric control system, data acquisition system and environmental room. The ambient temperature could be controlled in a
range of -25℃ to 55℃, and the control precision was ±0.2℃. The water inlet temperature could be controlled in a range of 5℃ to 90℃ and the control precision was ±0.2℃. The water outlet temperature was controlled through regulating the variable frequency water pump by a PID controller in the water conditioning system. The type-T thermocouples and the thermal resistances (PT100) were adopted in the environmental laboratory. The measurement range of the temperature sensors was -200-350℃, and the accuracy was ±0.2℃. The CO2 pressure was measured by pressure transmitters with a measurement range of 0-15 MPa and an accuracy of ±2.5% of full scale. The water volumetric flow rate was measured by an electromagnetic flowmeter with a measurement range of 0-6 m3·h-1 and an accuracy of ±0.5% of full scale. The electric power consumption of the prototype was measured by an electric power analyzer with an accuracy of ±0.25% of reading. An Agilent HP34970A data acquisition unit was employed to collect the measurement data with a scan interval of 10 sec. To display and record data conveniently, a computer that installed BenchLink Data Logger 3 Software was used. Due to the measurement error, switching error and transducer conversion error, the accuracies of DC voltage, DC current, T-type thermocouple and thermal resistances were ±(0.0035% of reading + 0.0005% of range), ±(0.050% of reading + 0.005% of range), ± 1℃ and ± 0.06℃, respectively. Among them, the DC voltage and DC Current were respectively the outputs of pressure transmitters and electromagnetic flowmeter. 2.3 Performance parameters and error analysis
In this study, the heating capacity, the total power consumption and the coefficient of performance (COP) will be chosen to evaluate the performance of the prototype. The heating capacity Qgc is defined as
Qgc mwcpw Tw,out Tw,in
(1)
where mw is the water mass flow rate in kg·s-1; cpw is the isobaric specific heat in kJ·(kg·K)-1; Tw, out and Tw,in are the water outlet temperature and the water inlet temperature in ℃. The total power consumption W is defined as W Wcomp Wf Welec
(2)
where Wcomp , Wf and Welec are the power consumptions of the compressor, fans and electronic devices in kW, respectively. The COP is obtained by dividing the heating capacity by the total power consumption, as shown below.
COP =
Qgc
(3)
W
The measurement errors for the heating capacity and the COP can be calculated by using the error propagation method reported by Kline and McClintock as given below 1/2
n R wR ( wi ) 2 i 1 xi
(4)
where wR is the resultant uncertainty; R is the given function of the independent variable xi ; wi is the uncertainties of the independent variable xi . Using the above equation, the measurement errors for the heating capacity and the COP are 3.15% and
3.58%, respectively. 3. Exergy analysis model The exergy of a thermodynamic system is the maximum theoretical useful work [23]. But in practical systems, the exergy cannot be completely used, and the exergy destruction and loss are unavoidable. The exergy balance for a thermodynamic system: EF EP ED EL
(5)
Where EP denotes the product exergy which represents the desired result produced by system; EF denotes the fuel exergy which represents the resources expended to generate the product; ED denotes the exergy destruction due to irreversibilities within system; EL denotes the exergy loss engendered by interactions between system and environment. The exergy efficiency is a parameter that can provide a true measure of the performance of an energy system from the thermodynamic viewpoint [24]. The exergy efficiency E is the ratio between product and fuel exergy:
E
EP EP EF EP ED EL
(6)
Due to the wide-range application of thermodynamic systems, the definitions of product and fuel exergy are varied with the specific demand. In this research, the prototype was designed to convert the electric energy into thermal energy in water. Therefore, the exergy associated with the heat transfer in gas cooler was considered as the product, and the power input of compressor was considered as the fuel. Fig. 2 shows the combined system and boundary in the exergy analysis. The heat pump water heater prototype was a closed system, in which the heat and work
interactions between the system and the environment existed. By combining the prototype and part of the environment, the boundary of the combined system was located outside the prototype where the temperature corresponded to the ambient temperature. The ambient temperature was taken here as the temperature of the exergy reference environment. In this case, the boundary of the combined system allowed energy transfer across it by heat transfer with the water and by electrical work in the compressor. With the boundary mentioned above, neglecting the heat dissipation in other components, the gas cooler was the only component where heat interaction between the combined system and the environment existed. Because the exergy associated with the heat transfer in the gas cooler was defined as product, the exergy loss vanished accordingly. Then, the exergy destruction E D in equation (5) accounted for the total exergy destruction within the combined system.
Fig. 2 The combined system and boundary in the exergy analysis In this research, the product exergy of system E p is defined as
Ep
h
gc ,in
T hgc ,out 1 0 TH
(7)
where hgc ,in and hgc ,out are the specific enthalpy of CO2 at the inlet and outlet of the gas cooler; T0 is the temperature of the exergy reference environment, namely the ambient temperature; the heat source temperature TH is the thermodynamic average temperature of water and is defined as TH
hw,out hw,in sw,out sw,in
(8)
The exergy balance of the compressor can be written as
comp ,in wcomp comp ,out ED ,comp
(9)
where comp ,in and comp ,out are the specific flow exergy at the inlet and outlet of the compressor. The specific flow exergy
is calculated by
h h0 T（ 0 s s0）
(10)
Considering the compression as adiabatic, the specific work of compressor wcomp can be calculated by wcomp hcomp ,out hcomp ,in
(11)
By substituting equation (10) and (11) into equation (9), the exergy destruction of the compressor E D ,comp is expressed as E D ,comp T0 scomp ,out scomp ,in
(12)
where scomp ,in and scomp ,out are the specific entropy of CO2 at the inlet and outlet of the compressor. Similarly, the exergy destruction of other components can be deduced. The exergy destruction of the gas cooler E D , gc is defined as E D , gc hgc ,in hgc ,out T0 sgc ,in sgc ,out E p
(13)
where s gc ,in and s gc ,out are the specific entropy of CO2 at the inlet and outlet of the gas cooler. The exergy destruction of the expansion valve E D ,exp is defined as E D ,exp hexp,in hexp,out T0 sexp,in sexp,out
(14)
where hexp,in and hexp,out are the specific enthalpy of CO2 at the inlet and outlet of the expansion valve; sexp,in and sexp,out are the specific entropy of CO2 at the inlet and outlet of the expansion valve. The exergy destruction of the evaporator E D ,evap is defined as
ED,evap hevap,in hevap,out T0 sevap,in sevap,out
(15)
where hevap ,in and hevap ,out are the specific enthalpy of CO2 at the inlet and outlet of the evaporator; sevap ,in and sevap ,out are the specific entropy of CO2 at the inlet and outlet of the evaporator. The exergy destruction of the IHX E D , IHX is defined as
ED , IHX T0 sIHX ,out ,HS sIHX ,in ,HS sIHX ,out ,LS sIHX ,in ,LS
(16)
where s IHX ,in , HS , s IHX ,out , HS , sIHX ,in ,LS and sIHX ,out ,LS are the specific entropy of CO2 at the high pressure side inlet, the high pressure side outlet, the low pressure side inlet and the low pressure side outlet of the IHX, respectively. The total exergy destruction of system E D ,total is defined as E D ,total E D ,comp E D , gc E D ,exp E D ,evap E D , IHX
(17)
4. Results and discussion An air-source heat pump water heater absorbs heat from the ambient air and releases heat to the water. The affecting factors on the system performance are mainly
the ambient temperature, the water inlet temperature and the water outlet temperature. In this research, the working condition of the prototype was determined by the ambient temperature, the water inlet temperature and the water outlet temperature. 4.1 Influence of IHX on the performance parameters and the optimal discharge pressure Fig. 3 shows the variations of the performance parameters with variable discharge pressure under different ambient temperature. As shown in Fig. 3, the heating capacity and the total power consumption increased with the increase of discharge pressure. From Fig. 3(c), it could be noted that the maximum COP occurred whether or not the IHX was connected. The growth of the heating capacity was higher than that of the total power consumption when the discharge pressure was relatively small, and it led to the increase of the COP. Conversely, the COP decreased when the discharge pressure was high, because the growth of the total power consumption was higher than that of the heating capacity. It could also be found from Fig. 3(c) that the maximum COP was improved by using IHX under the same working condition. Moreover, the optimal discharge pressure corresponding to the maximum COP was reduced by the application of IHX.
(a)
(b)
(c) Fig. 3 Variations of the performance parameters with variable discharge pressure under different ambient temperature. (a) Heating capacity. (b) Total power consumption. (c) COP. Fig. 4 shows the variations of the performance parameters with variable discharge pressure under different water inlet temperature. The heating capacity increased consistently with the increase of the discharge pressure when the water inlet temperature was 40℃. However, when the water inlet temperature was 10℃, the heating capacity first increased and then decreased with the increase of the discharge pressure. Therefore, as shown in Fig. 4(c), the COP decreased more rapidly with the increase of the discharge pressure when the water inlet temperature was 10℃ in comparison with that when the water inlet temperature was 40℃. Whether the IHX was applied, the optimal discharge pressure increased when the water inlet temperature changed from 10℃ to 40℃.
（a）
（b）
(c) Fig. 4 Variations of the performance parameters with variable discharge pressure under different water inlet temperature. (a) Heating capacity. (b) Total power consumption. (c) COP. Fig. 5 shows the variations of the performance parameters with variable discharge pressure under different water outlet temperature. The optimal discharge pressure increased when the water outlet temperature changed from 70℃ to 90℃.
(a)
(b)
(c) Fig. 5 Variations of the performance parameters with variable discharge pressure under different water outlet temperature. (a) Heating capacity. (b) Total power consumption. (c) COP. Table 2 summarizes the optimal discharge pressure and the performance parameters under the optimal discharge pressure in several working conditions. As Table 2 shows, the optimal discharge pressure decreased with the application of IHX in all working conditions, and the highest decreasing rate could be up to 6.92%. As for the performance parameters, under the optimal discharge pressure, the total power consumption decreased by up to 6.22%; the COP increased by up to 6.65%; but there is no obvious trend in the variations of the heating capacity.
Moreover, the reduction in the optimal discharge pressure caused by using IHX was influenced by the ambient temperature, the water inlet temperature and the water outlet temperature. It can be seen from Table 2 that the reduction in the optimal discharge pressure could be enlarged by the decrease of the ambient temperature, the increase of the water inlet temperature and the increase of the water outlet temperature. Optimal discharge pressure [MPa]
Heating capacity [kW]
Total power consumption [kW]
COP [-]
Without IHX
10.18
71.90
19.10
3.76
With IHX
9.90
70.16
18.42
3.81
Variation
-2.75%
-2.41%
-3.53%
1.15%
Without IHX
11.76
60.36
21.91
2.76
With IHX
11.41
60.50
20.73
2.92
Variation
-2.98%
0.23%
-5.40%
5.95%
Without IHX
11.40
50.42
20.77
2.43
With IHX
10.81
50.59
19.54
2.59
Variation
-5.18%
0.33%
-5.92%
6.65%
Without IHX
9.81
53.79
17.91
3.00
With IHX
9.52
54.99
17.46
3.15
Variation
-2.96%
2.24%
-2.53%
4.89%
Without IHX
11.89
55.83
20.30
2.75
With IHX
11.25
55.85
19.19
2.91
Variation
-5.38%
0.05%
-5.44%
5.81%
Without IHX
11.13
44.84
20.01
2.24
With IHX
10.36
44.36
18.76
2.36
Variation
-6.92%
-1.06%
-6.22%
5.51%
Working condition Tamb [℃]
16
16
7
2
2
2
Status of IHX Tw,in [℃]Tw,out [℃]
10
40
40
10
10
40
70
70
70
70
90
70
Table 2 Optimal discharge pressure and the performance parameters under the optimal discharge pressure in several working conditions. 4.2 Influence of IHX on the transcritical CO2 heat pump cycle In this section, the influence of IHX on the transcritical CO2 heat pump cycle was discussed based on experimental data. Fig. 6 depicts the transcritical CO2 heat pump
cycles with and without IHX under the optimal discharge pressure in the P-H diagram, in which 1-2-3-4-1 represents the cycle without IHX and 1’-2’-3’-4’-5’-6’-1’ represents the cycle with IHX. Because the quality of refrigerant at the inlet of the evaporator cannot be determined by measuring temperature and pressure, we assumed that the throttling in the expansion valve was isenthalpic.
Fig. 6 Transcritical CO2 heat pump cycles with and without IHX under the optimal discharge pressure in the P-H diagram. The heat pump cycles can be divided into the following processes: Process 1-2 (1’-2’): Compression in the compressor Process 2-3 (2’-3’): Heat release without phase change in the gas cooler Process 3-4 (4’-5’): Throttling process in the expansion valve Process 4-1 (5’-6’): Heat absorb with phase change in the evaporator
Process 3’-4’ and 6’-1’: Heat transfer in the IHX The influence of IHX on the cycle pressure mainly occurred on the high pressure side. As shown in Fig. 6, the gas cooler pressure in the cycles with IHX was significantly lower when compared with that in the cycles without IHX. In terms of the low pressure side, it can be seen from the figure that the evaporator pressure was barely affected by using IHX. Due to the heat transfer in the IHX, the cycle temperature was influenced. Firstly, the suction temperature and discharge temperature of the compressor in the cycles with IHX were higher when compared with those in the cycles without IHX. However, the coking of lubricating oil could occur when the discharge temperature is extremely high. In a heat pump cycle with IHX, the discharge temperature increases when the ambient temperature decreases, because of the reduction in evaporation temperature and the increase of compressor pressure ratio. Besides, the increase of the discharge temperature was larger when the water inlet temperature was higher. Hence, in order to avoid the adverse effect caused by extremely high discharge temperature, careful consideration should be made when the IHX is applied in the working conditions with low ambient temperature and high water inlet temperature. For instance, the heat exchanger effectiveness of IHX should be maintained under a low level. Secondly, the refrigerant status at the outlet of the evaporator was marginally influenced by using IHX. This can be evidenced by the situations of point 1 and 6’. As a consequence of the subcooling and the superheat generated by IHX, the evaporator inlet enthalpy decreased, and the suction temperature increased. Then the refrigerant
mass flow rate reduced with the increase of the suction temperature. In this case, under the constant heat transfer area of the evaporator, the enthalpy difference between the inlet and outlet of the evaporator increased with the decrease of the refrigerant mass flow rate. Therefore, the refrigerant status at the outlet of the evaporator had no significant change as a net result. Thirdly, the gas cooler outlet temperature was influenced by using IHX. When the water inlet temperature was 10℃, the gas cooler outlet temperature slightly decreased with the application of IHX. However, when the water inlet temperature was 40℃, there was little change in the gas cooler outlet temperature. The reason can be explained by the heat transfer area of the gas cooler. The heat transfer area was suitable when the water inlet temperature was low, and surplus when the water inlet temperature was high, because the gas cooler in the prototype was basically designed for tap water heating. Hence, when the water inlet temperature was 10℃, the decrease of the refrigerant mass flow rate caused by using IHX could make the heat release more complete and reduce the temperature approach at the outlet of gas cooler. Owing to the influence of IHX on the cycle temperature, as shown in Fig. 6, the enthalpy differences between the inlet and outlet of the compressor and between the inlet and outlet of the gas cooler increased. In general, the heating capacity and the compressor power consumption were increased by the increase of the enthalpy differences, and decreased by the decrease of the refrigerant mass flow rate. Also, the compressor power consumption decreased with the decrease of the optimal discharge pressure. Therefore, the variations on the heating capacity and the compressor power
consumption were the net results of the above multiple effects. 4.3 Influence of IHX on the system exergy In this section, the exergy analysis was conducted to analyze the influence of IHX based on the heat pump cycles under the optimal discharge pressure. Table 3 summarizes the exergy destruction of the components and heat pump system in different working conditions. As shown in Table 3, the total exergy destruction increased by using IHX. Regarding the exergy destruction of the components, by the application of IHX, the exergy destruction of the compressor, the gas cooler and the evaporator increased, the exergy destruction of the expansion valve decreased. Exergy destruction [kJ·kg-1] Tamb/Tw,in/Tw,out Tamb/Tw,in/Tw,out Tamb/Tw,in/Tw,out Tamb/Tw,in/Tw,out Tamb/Tw,in/Tw,out
Component
=16/10/70℃
=16/40/70℃
=2/10/70℃
=2/40/70℃
=2/10/90℃
Without With Without With Without With Without With Without With IHX
IHX
IHX
IHX
IHX
IHX
IHX
IHX
IHX
IHX
Compressor
1.71
4.72
4.52
8.31
3.31
11.89
7.54
12.67
7.06
9.31
Gas cooler
7.70
9.12
5.75
9.99
7.55
10.44
5.78
11.22
7.41
9.94
Expansion valve
7.31
6.29
13.57
10.94
8.52
6.77
18.07
13.74
10.13
8.79
Evaporator
8.90
10.12
5.14
5.66
6.78
7.87
3.57
3.58
7.40
7.85
IHX
-
0.54
-
1.87
-
0.88
-
3.64
-
0.88
Total
25.62
30.78
28.98
36.77
26.17
37.84
34.96
44.85
32.00
36.77
Table 3 Exergy destruction of the components and the heat pump system in different working conditions. Fig. 7 shows the exergy destruction percentages of the components in different working conditions. By using IHX, the exergy destruction percentage of the compressor
increased; the exergy destruction percentage of the gas cooler increased when the water inlet temperature was 40℃ or the water outlet temperature was 90℃; but the exergy destruction percentage of the gas cooler slightly decreased when the water inlet temperature and the water outlet temperature were 10℃ and 70℃, respectively; the exergy destruction percentage of the expansion valve and evaporator decreased; with regard to the IHX, the exergy destruction percentage of IHX was significantly higher when the water inlet temperature was 40℃ than that when the water inlet temperature was 10℃.
Fig. 7 Exergy destruction percentages of the components in different working conditions. Fig. 8 shows the product exergy, the total exergy destruction and the exergy efficiency of system in different working conditions. It can be seen that the product exergy and the total exergy destruction increased with the application of IHX. Moreover, the exergy efficiency of system decreased by using IHX in all working conditions.
Fig. 8 Product exergy, total exergy destruction and exergy efficiency of system in different working conditions. 5. Conclusions In this paper, the influence of IHX on the air source transcritical CO2 heat pump water heater prototype was experimentally investigated from the energy and exergy viewpoints. The results showed that the optimal discharge pressure reduced with the application of IHX, and this reduction could be enlarged by the decrease of the ambient temperature, the increase of the water inlet temperature and the increase of the water outlet temperature. Under the optimal discharge pressure, the total power consumption decreased by up to 6.22% and the COP increased by up to 6.65% by using IHX. The evaporator pressure in heat pump cycles under the optimal discharge pressure was barely affected by using IHX, and the refrigerant status at the outlet of evaporator
was nearly constant. With the application of IHX, the suction temperature and the discharge temperature of the compressor increased. When the water inlet temperature was 10℃, the gas cooler outlet temperature decreased slightly by using IHX. However, when the water inlet temperature was 40℃, there was little change in the gas cooler outlet temperature. By using IHX, the exergy destruction of compressor, gas cooler and evaporator increased and the exergy destruction of expansion valve decreased. The total exergy destruction and the product exergy increased with the application of IHX, and the exergy efficiency of system reduced in all working conditions. Through this research, the improvement in the energy performance that was attributed to using IHX was verified in different working conditions. Besides, with the application of IHX, the reliability of the heat pump system might also be enhanced by the reduced optimal discharge pressure. Hence, using IHX could be recommended in case that the increased discharge temperature does not exceed the upper limit value associated with compressor and lubricating oil. Actually, in terms of the studied prototype, when the ambient temperature was higher than 2℃, the IHX could be unconditionally applied and the discharge temperature did not exceed 140℃ which was the recommended upper limit from the manufacturer. This result could be used as a reference for similar systems. Notably, the reduction in the system exergy efficiency caused by using IHX was also discovered, which was different from the theoretical results in previous publications [20, 21]. This difference might result from the refrigerant pressure drop in
the heat exchangers, change of compressor efficiency with increased suction and discharge temperature, etc. Considering the difficulties in studying these factors experimentally, the detailed theoretical model should be developed in further study, and then the analysis about potential influential parameters could be conducted based on the model. Acknowledgement The research presented was supported by the National Natural Science Foundations of China (51976153) and the Fundamental Research Funds for the Central Universities. References [1] K. Nawaz, B. Shen, A. Elatar, V. Baxter, O. Abdelaziz, Performance optimization of CO2 heat pump water heater, International Journal of Refrigeration, 85 (2018) 213228. [2] S. Wang, H. Tuo, F. Cao, Z. Xing, Experimental investigation on air-source transcritical CO2 heat pump water heater system at a fixed water inlet temperature, International Journal of Refrigeration, 36 (2013) 701-716. [3] P.-C. Qi, Y.-L. He, X.-L. Wang, X.-Z. Meng, Experimental investigation of the optimal heat rejection pressure for a transcritical CO2 heat pump water heater, Applied Thermal Engineering, 56 (2013) 120-125. [4] M. Saikawa, S. Koyama, Thermodynamic analysis of vapor compression heat pump cycle for tap water heating and development of CO2 heat pump water heater for residential use, Applied Thermal Engineering, 106 (2016) 1236-1243. [5] S. Liu, Z. Li, B. Dai, Z. Zhong, H. Li, M. Song, Z. Sun, Energetic, economic and environmental analysis of air source transcritical CO2 heat pump system for residential heating in China, Applied Thermal Engineering, 148 (2019) 1425-1439. [6] J.-F. Zhang, Y. Qin, C.-C. Wang, Review on CO2 heat pump water heater for residential use in Japan, Renewable and Sustainable Energy Reviews, 50 (2015) 13831391. [7] R. Llopis, L. Nebot-Andrés, D. Sánchez, J. Catalán-Gil, R. Cabello, Subcooling methods for CO2 refrigeration cycles: A review, International Journal of Refrigeration, 93 (2018) 85-107. [8] S.A. Klein , D.T. Reindl, K. Brownell, Refrigeration system performance using liquid-suction heat exchangers, International Journal of Refrigeration, (2000). [9] Y. Chen, J. Gu, The optimum high pressure for CO2 transcritical refrigeration
systems with internal heat exchangers, International Journal of Refrigeration, 28 (2005) 1238-1249. [10] S.G. Kim, Y.J. Kim, G. Lee, M.S. Kim, The performance of a transcritical CO2 cycle with an internal heat exchanger for hot water heating, International Journal of Refrigeration, 28 (2005) 1064-1072. [11] Y.B. Tao, Y.L. He, W.Q. Tao, Z.G. Wu, Experimental study on the performance of CO2 residential air-conditioning system with an internal heat exchanger, Energy Conversion and Management, 51 (2010) 64-70. [12] N. Fernandez, Y. Hwang, R. Radermacher, Comparison of CO2 heat pump water heater performance with baseline cycle and two high COP cycles, International Journal of Refrigeration, 33 (2010) 635-644. [13] E. Torrella, D. Sánchez, R. Llopis, R. Cabello, Energetic evaluation of an internal heat exchanger in a CO2 transcritical refrigeration plant using experimental data, International Journal of Refrigeration, 34 (2011) 40-49. [14] F.Z. Zhang, P.X. Jiang, Y.S. Lin, Y.W. Zhang, Efficiencies of subcritical and transcritical CO2 inverse cycles with and without an internal heat exchanger, Applied Thermal Engineering, 31 (2011) 432-438. [15] D. Sánchez, J. Patiño, R. Llopis, R. Cabello, E. Torrella, F.V. Fuentes, New positions for an internal heat exchanger in a CO2 supercritical refrigeration plant. Experimental analysis and energetic evaluation, Applied Thermal Engineering, 63 (2014) 129-139. [16] A. Fartaj, D.S.K. Ting, W.W. Yang, Second law analysis of the transcritical CO2 refrigeration cycle, Energy Conversion and Management, 45 (2004) 2269-2281. [17] J. Sarkar, S. Bhattacharyya, M. Ram Gopal, Transcritical CO2 heat pump systems: exergy analysis including heat transfer and fluid flow effects, Energy Conversion and Management, 46 (2005) 2053-2067. [18] B. Dai, H. Qi, S. Liu, M. Ma, Z. Zhong, H. Li, M. Song, Z. Sun, Evaluation of transcritical CO2 heat pump system integrated with mechanical subcooling by utilizing energy, exergy and economic methodologies for residential heating, Energy Conversion and Management, 192 (2019) 202-220. [19] M. Purjam, K. Goudarzi, High efficiency sub-critical carbon dioxide supplementary heat pump for low temperature climates (energy and exergy analysis), Renewable Energy, 133 (2019) 166-176. [20] O. Joneydi Shariatzadeh, S.S. Abolhassani, M. Rahmani, M. Ziaee Nejad, Comparison of transcritical CO2 refrigeration cycle with expander and throttling valve including/excluding internal heat exchanger: Exergy and energy points of view, Applied Thermal Engineering, 93 (2016) 779-787. [21] H. Ghazizade-Ahsaee, M. Ameri, Effects of using expander and internal heat exchanger on carbon dioxide direct-expansion geothermal heat pump, Applied Thermal Engineering, 136 (2018) 389-407. [22] D. Wang, Y. Liu, Z. Kou, L. Yao, Y. Lu, L. Tao, P. Xia, Energy and exergy analysis of an air-source heat pump water heater system using CO2/R170 mixture as an azeotropy refrigerant for sustainable development, International Journal of Refrigeration, 106 (2019) 628-638.
[23] G. Tsatsaronis, Definitions and nomenclature in exergy analysis and exergoeconomics, Energy, 32 (2007) 249-253. [24] A. Bejan, G. Tsatsaronis, M. Moran, Thermal design and optimization, Wiley, New York (1996).
Highlights
•
The energy and exergy influence of internal heat exchanger was studied.
•
Influence of internal heat exchanger on optimal discharge pressure was presented.
•
Influence of internal heat exchanger on system cycle was discussed in P-H diagram.
•
Exergy destruction and efficiency was analyzed based on the experimental data.
Declaration of interests
☒ The authors declare that they have no known competing financial interests or personal relationships that could have appeared to influence the work reported in this paper.
☐The authors declare the following financial interests/personal relationships which may be considered as potential competing interests: