A novel defrosting method using heat energy dissipated by the compressor of an air source heat pump

Applied Energy 133 (2014) 101–111

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A novel defrosting method using heat energy dissipated by the compressor of an air source heat pump Zhang Long, Dong Jiankai, Jiang Yiqiang ⇑, Yao Yang Department of Building Thermal Energy Engineering, Harbin Institute of Technology, Harbin, China

h i g h l i g h t s  A novel defrosting method using heat energy dissipated by the compressor was developed and examined.  During the defrosting process, mean temperature difference between the air entering and leaving the indoor coil reached 4.1 °C.  Compared to the traditional method, defrosting time was shortened by 65% while the resuming heating period disappeared.  For the defrosting operation, the total energy consumption was less than 27.9% compared to the traditional method.

a r t i c l e

i n f o

Article history: Received 15 January 2014 Received in revised form 21 June 2014 Accepted 10 July 2014

Keywords: Air source heat pump Defrosting Continuous heating Dissipated heat of compressor Thermal energy storage

a b s t r a c t When an air source heat pump (ASHP) unit is used for space heating at low ambient temperatures in winter, frost may form on its outdoor coil surface. Since the accumulated frost adversely affects its performance and energy efficiency, periodic defrosting of the outdoor coil is necessary. Currently, the reverse-cycle defrosting (RCD) method is widely used for the defrosting of ASHP. However, this operation interrupts space heating during the defrosting process. A time lag occurs to resume heating at end of the defrosting cycle. Moreover, frequent reversing of the 4-way valve may cause mass leakage of the refrigerant, even make the system unsafe. Furthermore, some amount of heat is dissipated to the atmosphere through the compressor casing. To improve the defrosting process and use this waste heat, a novel ASHP unit is developed. The space is heated during the defrosting process using the heat dissipated by the compressor. Experiments using both the RCD method and the novel reverse cycle defrosting (NRCD) method developed in this study are conducted on an ASHP unit of 8.9 kW nominal heating capacity. The experimental results indicated that in the NRCD method, the discharge and suction pressures are increased by 0.33 MPa and 0.14 MPa, respectively, the defrosting time is shortened by 65% while the resuming heating period vanished with the NRCD method, and that the total energy consumption in comparison to RCD method is reduced by 27.9% during the period which is composed of defrosting period and resuming heating period. Moreover, the NRCD method ensured continuous heating during defrosting. The mean temperature difference between the air entering and leaving the indoor coil reaches 4.1 °C during defrosting. Over a test period of 125 min, compared to RCD method, the total heating capacity and input power are increased by 14.2% and 12.6%, respectively. The increase in the system COP is 1.4%. Ó 2014 Elsevier Ltd. All rights reserved.

1. Introduction Due to their significant energy potential, the air source heat pump (ASHP) units are used worldwide [1]. However, when an ASHP unit operates for the space heating at low ambient temperatures in winter, frost that forms on its outdoor coil presents several problems. The frost reduces the airflow and acts as a thermal insulator, leading to deterioration in the performance of the outdoor ⇑ Corresponding author. Tel./fax: +86 451 86282123. E-mail address: [email protected] (J. Yiqiang). http://dx.doi.org/10.1016/j.apenergy.2014.07.039 0306-2619/Ó 2014 Elsevier Ltd. All rights reserved.

coil, and even shutdown of the ASHP unit, lowering the energy efficiency and heating capacity [2–4]. To prevent this, the frost layer requires periodic removal. Many studies have been undertaken to solve the problems caused by frosting of outdoor coil which include improving the defrosting performance of the ASHP unit [5–7], delaying the frost formation [8–11] and the application of more effective defrosting methods [12–16]. Currently, the RCD [17] is the most widely used method for defrosting ASHP units. There are several factors affecting the dynamic characteristics of the RCD operation in an ASHP unit. The studies reported on the RCD method are mainly dynamic

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Nomenclature

a c Cl Cs I M P1 P2 Q1 Q2 Q3

effective exothermic coefficient of PCM specific heat of air (J/kg °C)) specific heat of liquid PCM (J/kg °C) specific heat of solid PCM (J/kg °C) current (A) the quality of PCM (kg) input power (W) heating capacity (W) energy consumption during defrosting (kJ) space heating energy during defrosting (kJ) energy supplied by the PCM-HE during defrosting (kJ)

simulations for improving the operational characteristics of ASHP system. O’Neal et al. [18] experimentally investigated the transient defrosting performance of a 3-ton (or 3.514 kW) residential ASHP unit using a thermostatic expansion valve (TEV) and found that both the accumulator and the TEV impacted significantly on the system’s dynamic responses. In addition, the cycle performance during frosting in an ASHP unit with either a scroll or reciprocating compressor was experimentally studied and compared, in accordance with ANSI/ASHRAE Standard 116-1983 [19]. The results suggested that an ASHP unit using scroll compressor achieved a slightly higher integrated COP and a lower compressor discharge temperature during the frosting and defrosting. There are other reported studies aiming to decrease the period of defrosting and the associated energy losses. For example, O’Neal et al. [20] considered that increasing the orifice diameter would contribute to decrease the defrosting period. Huang et al. [21] found that at the end of RCD operation, a discharge pressure of 742.3 kPa in an ASHP unit using a fan pre-start method which is lower than that using the normal fan-start method. Also, the results from dynamic simulation of the RCD operation of an ASHP are reported [22,23]. However, there are still a series of problems in the RCD. First, during the defrosting process, the indoor supply air fan is normally switched off, and the energy used for melting the frost mainly comes from the input power of the compressor (and the indoor coil metal) which is usually insufficient for the defrosting. This will prolong the defrosting duration [24]. Second, the heating is interrupted during defrosting, and certain time is taken to reheat the cooled pipes of the indoor coil after the defrosting, which further delays resuming the space heating [25]. Third, the 4-way valve needs to be reversed twice in a single defrosting cycle, and such frequent reversals may cause mass leakage of refrigerant, and even affect the system’s safety [26,27]. To minimize the damage to the 4-way valve, prior to its reversal, the compressor discharge and suction pressures need to be balanced which however will prolong the defrosting time. On the other hand, a certain amount of heat from operation of compressor is usually dissipated to the ambient environment. Ooi and Wong [28] estimated that 10–20% of the total input energy is dissipated to the ambient environment through the compressor casing by convection and conduction. Park [29] indicated that the heat dissipated from the compressor is approximately 6.3% of the input power. In other words, over a period of time, considerable quantity of heat energy is dissipated to the surrounding environment through the compressor casing. Therefore, a need exists to study how to utilize the heat dissipated by the compressor to speed up the defrosting and further improve the indoor environment. In order to solve the fundamental problems, such as nonavailability of sufficient heat and interruption of heating during defrosting, a RCD method using thermal energy storage of the heat

T Ti Tm Tf t U V_

q PF

air temperature (°C) the initial temperature of PCM (°C) the melting point of PCM (°C) the final temperature of PCM (°C) defrosting time duration (s) voltage (V) airflow rate (m3/s) density of air (kg/m3) power factor

dissipated by the compressor of an ASHP is developed. In this method, the heat dissipated by the compressor is stored in a phase change material-heat exchanger (PCM-HE) during space heating, and later the stored energy is used to melt the frost off the outdoor coil. For the purpose of comparison, experiments using both RCD and NRCD methods are also conducted. The presentation of detailed description of the experimental ASHP unit integrated with the PCM-HE, experimental procedure and a detailed discussion on the experimental results follows. 2. Experimentation 2.1. The experimental setup and instrumentation As can be seen from Fig. 1, the experimental setup was made up of a psychrometric room and a test heat pump. The psychrometric room was of size of 3.12 m (L)  3.10 m (W)  2.14 m (H), in which frosting outdoor environment could be simulated, by using air conditioning system consisting of 6.5 kW electric heaters, a humidifier of 3 kg/h water evaporation and a cooling coil with a capacity of 9.8 kW. The psychrometric room and the test heat pump were located inside a large enclosed space which was provided with a separate central space heating system. Since the central heating which was thermostatically controlled was used for general heating purpose, the air temperature inside the space was maintained at its design value, whether the experimental ASHP unit was in operation or not. Hence, the simulated environmental conditions of the ASHP indoor and outdoor coils, such as air temperature and air humidity, could be kept constant. Fig. 2 shows the schematic of the experimental ASHP unit in which the novel thermal energy storage (TES) based reverse-cycle hot gas defrosting method was applied and tested. Compared to the conventional ASHP unit, the experimental unit includes a phase change material-heat exchanger (PCM-HE) and three solenoid modulating valves (F1, F2 and F3 shown in Fig. 2). The PCM-HE was wrapped around the compressor tightly to store the heat dissipated by the compressor while in operation and release it for defrosting. During the standard heating and cooling operations, the valves F1 and F3 were kept in closed position and the valve F2 in open position. The specifications of the experimental ASHP unit are shown in Table 1. The PCM-HE is a key component in the successful development of this novel defrosting method. Following is the stepwise procedure for its design. First, the amount of energy that could be stored by the PCM-HE is estimated. According to Park’s experimental data [29], the energy dissipated through the compressor casing is 6.3% of its input power. Therefore, during a heating cycle of 50 min, the amount of energy that can be stored by the PCM-HE would

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Large environmental room Psychrometric room Test heat pump Heaters

M

F1 M F2

Four way valve

Outdoor coil

Indoor coil

Humidifier PCM-HE Liquid-gas separator

Compressor M

F3 Electronic expansion valve

Capillary tube

Cooling coil

Fig. 1. The schematic of the experimental setup.

F1

T

F2

M

M

Outdoor coil

T

Four way valve

Indoor coil

T W

T P

T W

PCM-HE

T P

T

Liquid-gas separator Electronic expansion valve

F3

T P Capillary tube P T

T T

Temperature sensor

T

Compressor M

T P

T

T

P

Pressure sensor

W

Humidity sensor

Fig. 2. A schematic showing the operation of the experimental AHSP unit.

reach 558 kJ. Second, an appropriate PCM is selected. Its melting point requires careful consideration. If it is too high, it would decrease heat transfer between the compressor casing and the PCM-HE during heating and over heat the suction gas during defrosting. If the melting point of PCM is too low, it would reduce the casing temperature and the performance of compressor during heating and decrease the defrosting efficiency. In addition, the PCM should have a high latent heat, less volume change between liquid and solid phases and high thermal stability. Third, on the considerations mentioned the PCM-HE is designed. It was designed as to cover three fourths of the compressor casing and was in shape of a circular tank to wrap around the compressor tightly, and then the PCM-HE was covered with heat insulating material. Finally, the PCM selected was a mixture of 35% mollauric acid and 65% molcapric acid. The thermal properties were tested using

Differential Scanning Calorimetry (DSC). The test results are shown in Table 2. A finned tube was used as PCM-HE for enhancing the heat transfer between PCM and the refrigerant. The constructional details and structural parameters of the PCM-HE are shown in Fig. 3 and Table 3, respectively. During the experiments, the temperatures of the refrigerant, PCM, indoor and outdoor air were measured with the temperature sensors (PT1000, Class A). The relative humidity of air at both the inlet and outlet of outdoor coil were measured using a relative humidity sensor (of ±1.0% accuracy). The refrigerant pressures were measured using pressure transmitters (of ±0.05 bar accuracy) with a measuring range of 0–60 bar. The input power was measured with a power module (of ±0.5%accuracy). All sensors and measuring devices used in the experimental ASHP unit were able to output direct current signals of 1–5 V and

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Table 1 The specifications of the experimental ASHP unit.

Cooling capacity Heating capacity Air flow rate

High Middle Low Indoor unit Outdoor unit

Noise Electrical characteristics

Power Input power Operating current

Specifications

Dimensions Net weight Compressor Fan EEV Capillary tube Valves F1/F2/F3 Refrigerant

Value

Units

19.8 133.2 0.143 900 895

°C kJ/kg W/m K kg/m3 kg/m3

1-220-50 2130 2950 10.0 13.8

mm (W  H  D) mm (W  H  D) kg kg Rotary type Axial flow DPF(Q)1.8C-23 mm (u  L) FDF8AB R410A

2.2. Comparative experiments: modes and procedures The comparative experiments were carried out in the following two operational modes:

530  310  1810 860  308  730 53 60

1.4  100

2.2.1. Mode A: standard heating and RCD operation mode During the standard heating (or frosting) operation, the air temperature and relative humidity inside the large environmental room were maintained at 20.0 ± 0.2 °C (dry-bulb) and 38.0% ± 5.0%, respectively. The inside conditions in the psychrometric room were at 2.0 ± 0.1 °C (dry-bulb) and 83.0% ± 2.0%. The temperature and relative humidity were maintained with the experimental ASHP unit, the existing heaters, cooling coil and humidifier. The stated conditions meet the Chinese defrosting standard GBT 77252004 [30]. The experimental ASHP unit was initially operated for space heating (or outdoor coil frosting) for 50 min, with the valves F1 and F3 closed and valve F2 open. The refrigerant was discharged from the compressor and passed through the 4-way valve, then, it flowed into the indoor coil and condensed by the indoor air. Later, it was throttled by the capillary tube and an electronic expansion valve and flowed into the outdoor coil and evaporated by the out-

PCM-HE

270mm

Circular ring tank

u-V-Hz

Indoor unit Outdoor unit Indoor unit Outdoor unit Type Type Type

0–1 V, which were transferred to a data acquisition unit for logging and recording. The experimental data were recorded every 5 s.

R106mm R58mm

7200 8900 1200 900 650 41 56

W W A A

Table 2 The thermal properties of the mixture of 35% mollauric acid and 65% molcapric acid.

Melting point Latent heat Thermal conductivity Density (solid) Density (liquid)

W W m3/h m3/h m3/h dB dB

Cooling Heating Cooling Heating

Type Type

Property

Value

250mm

Rated performance

Units

Finned refrigerant tube

Fig. 3. The constructional details and structural parameters of the PCM-HE.

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Z. Long et al. / Applied Energy 133 (2014) 101–111 Table 3 The structural parameters of the PCM-HE.

Circular ring tank

Finned refrigerant tank

PCM

Parameters

Valve

Unit

Material Height Inner ring radius/length Outer ring radius/length Volume

Copper 270 58/314 106/573 5750

mm mm mm ml

Fin material Spacing of fins Refrigerant tube material Diameter of refrigerant tube Height Total length of refrigerant tube

Copper 4 Copper 10 250 7

Net volume

4750

mm mm m ml

door air. Finally, the refrigerant passed through the liquid-gas separator and returned to the compressor. In this manner, the cyclic process repeated. The RCD operation consisted of two stages, first, the defrosting operation and the second, resuming space heating operation. During the whole process, both the valves F1 and F3 were closed and valve F2 was kept open. In the first stage, at the end of a heating operation, the compressor speed was reduced to balance the system pressure to prevent the 4-way valve from damage when it was reversed. At the same time, both the indoor and outdoor air fans were turned off, and the electronic expansion valve was set for maximum opening. After the 4-way valve was reversed, the compressor speed was increased and the defrosting operation was started. The refrigerant was discharged from the compressor and passed through the 4-way valve, then, it was sent to the outdoor coil and used for melting the frost accumulated on the outdoor coil surface. Subsequently, the refrigerant was throttled by the capillary tube and flowed into the indoor coil, where it gained some heat energy. Finally, the refrigerant passed through the liquid-gas separator and returned to the compressor. The defrosting operation was manually terminated when the tube wall temperature of the lowest refrigerant circuit of the outdoor coil reached 15 °C, at which the frost was completely melted and drained or evaporated. In the second stage, the compressor speed was first reduced, and then the 4-way valve was reversed. Thereafter, the compressor speed was increased to the original speed and the outdoor air fan was turned on. In this stage the refrigerant flow was same as that in the frosting operation. The resuming space heating operation was manually terminated when the tube wall temperature of indoor coil reached 25 °C, and at the same time indoor air fan was turned on. The indoor environment heating was started again. The experimental results obtained in this mode of operation provided the basis for comparison of RCD with NRCD.

2.2.2. Mode B: storing the heat dissipated by the compressor during the standard heating and NRCD operation modes The frosting operation for this mode was the same as that of the mode A. The heat dissipated by the compressor was stored by the PCM automatically when it was operated. For proper comparison, the NRCD operation was also divided into the defrosting operation and resuming space heating operation. For the defrosting operation, at the end of a heating operation, the valves F1 and F3 were switched on and the valve F2 was switched off. The indoor air fans peed was reduced to low. However, there was no need to reduce the compressor speed for not reversing the 4-way valve. The refrigerant was discharged from the compressor and passed through the 4-way valve, then, it was divided into two flows according to the pressure difference

between the indoor and outdoor sides. One refrigerant flow was sent to the indoor coil, condensed by the indoor air, then throttled by the capillary tube; the other refrigerant flow after passing through the valve F1 was diverted to the outdoor coil, for melting the accumulated frost, and then throttled by the electronic expansion valve. The two refrigerant flows were mixed and passed through the valve F3, then, the refrigerant flowed into the PCMHE and evaporated by the PCM. Finally, the refrigerant passed through the liquid-gas separator and returned to the compressor. When the tube wall temperature of the lowest refrigerant circuit of the outdoor coil reached 15 °C, the defrosting operation was terminated. It was followed by the resuming space heating stage, in which both the valves F1 and F3 were switched off and the valve F2 was switched on, and the 4-way valve remained out of operation. The outdoor air fan was also switched on. In this stage the refrigerant flow was the same as that in the frosting operation. The resuming space heating stage was terminated when the liquid tube wall temperature of the indoor coil reached 25 °C, and at the same time, the indoor air fans peed was changed from low to high. 2.3. Data reduction During the defrosting process, the total energy consumption and the input power to the compressor and fans were measured to determine the overall performance. The defrosting duration was considered as the time interval between the defrosting initiation and its termination. The resuming space heating duration was the time interval between the defrosting termination and heating initiation. The input power of the compressor was evaluated by the following relationship:

P1 ¼ U  I  PF

ð1Þ

The actual amount of energy consumption can be evaluated by:

Q1 ¼

Z

t

P1 dt ¼

X

P 1  Dt

ð2Þ

0

The heating capacity can be evaluated by:

P2 ¼ c  q  V_  DT

ð3Þ

The actual amount of space heating energy can be evaluated by:

Q2 ¼

Z

t

P2 dt ¼

X

P 2  Dt

ð4Þ

0

The actual amount of energy supplied by the PCM-HE can be evaluated by:

  Q 3 ¼ a  M C l ðTi  Tm Þ þ L þ C s ðTm  Tf Þ

ð5Þ

where Dt = 5 s, is the time interval used for measuring the operating parameters. 3. Experimental results and discussion Under similar experimental conditions, the defrosting and resuming space heating performances with the NRCD method were different from those obtained with the RCD method, as shown in Figs. 4–13. In all these figures, the origin of horizontal (time) axis is the actual starting point of the defrosting operation. Fig. 4 shows the measured values of tube wall temperature of the lowest refrigerant circuit of the outdoor coil during the defrosting operation in both the modes. It can be used to determine the time for the termination of the defrosting operation. As seen, in the mode A, initially the temperature increased rapidly due to adequate energy transfer from the indoor coil and system interconnecting copper piping. Thereafter, it reached 0 °C after70 s and

Z. Long et al. / Applied Energy 133 (2014) 101–111

o

Temperature ( C)

106

20 16 12 8 4 0 -4 -8 -12 -16 -20

Mode A

0

20

40

60

80

100 120 140 160 180 200 220 240 260 280 300 320 340

o

Temperature ( C)

Times (s) 20 16 12 8 4 0 -4 -8 -12 -16 -20

Mode B

10

0

20

30

40

50

60

70

80

90

100

110

120

Times (s) Fig. 4. The measured tube wall temperatures of the lowest refrigerant circuit of the outdoor coil during the defrosting operation in the modes A and B.

2.4

Pressure (MPa)

2.0

Suction pressure in mode A

Discharge pressure in mode A

1.6 1.2 0.8 0.4 0

20

40

60

80

100 120 140 160 180 200 220 240 260 280 300 320 340

Times (s) 2.4

Pressure (MPa)

Suction pressure in mode B

Dischagre pressure in mode B

2.0 1.6 1.2 0.8 0.4 0

10

20

30

40

50

60

70

80

90

100

110

120

Times (s) Fig. 5. The measured compressor suction and discharge pressures during the defrosting operation in the modes A and B.

remained at this temperature for 150 s. The defrosting started at 220 s and 120 s later the temperature steadily increased to 16.3 °C with the frost decreasing on the outdoor coil surface. In the mode B, the temperature increased rapidly during the entire defrosting operation due to the high temperature of discharge gas and large refrigerant flow through the outside coil. It can be seen from Fig. 4, that the change in temperature from 17.2 °C to 18.5 °C in 120 s is linear. It is evident that the defrosting duration was 340 s in the mode A and 120 s in the mode B. This amounted to shortening the defrosting duration by 65%. This reduction in the defrosting duration was due to the additional energy provided by the PCM-HE.

This was because more energy was provided from the PCM-HE, which stored the compressor’s waste heat energy during heating, compared with heat energy transferred from the indoor coil. Fig. 5 shows the compressor suction and discharge pressures measured during defrosting operation in both the modes. In both these modes, when the defrosting operation started, the discharge pressure sharply decreased and the suction pressure increased rapidly. The trend in variation of pressures in the mode B was sharper than that in the mode A. The discharge pressures decreased due to the reduction of the compressor speed in the mode A and the same in the Mode B was due to closing and opening of the valves. In the mode A, at the time instant of 60 s, the discharge and suction pres-

107

o

Temperature ( C)

Z. Long et al. / Applied Energy 133 (2014) 101–111

24 20 16 12 8 4 0 -4 -8 -12

Mode A

0

20

40

60

80 100 120 140 160 180 200 220 240 260 280 300 320 340

Times (s) 12

Temperature ( C)

10

Mode B

o

8 6 4 2 0 -2 -4

0

10

20

30

40

50

60

70

80

90

100

110

120

Times (s) Fig. 6. The DS at suction point during the defrosting operation in the modes A and B.

38

36

36

34 Outlet air temperature of indoor unit in mode B Inlet air temperature of indoor unit in mode B

30 28 26 24

PCM temperature

34 32

Temperature (oC)

Temperature (oC)

32

30 28 26 24 22 20

22

18

20 18

16

0

10

20

30

40

50

60

70

80

90

100 110 120

Times (s)

14

0

10

20

30

40

50

60

70

80

90

100 110 120

Times (s)

Fig. 7. The measured air temperatures at the inlet and outlet of the indoor coil during the defrosting operation in the mode B.

Fig. 8. The measured PCM temperatures during the defrosting operation in the mode B.

sures were 1.04 MPa and 1.00 MPa, respectively, which confirms achieving pressure balance when the 4-way valve was reversed. At the end of defrosting operation, in the mode A, the discharge and suction pressures reached 1.68 MPa and 0.46 MPa, respectively, while they reached 1.87 MPa and 0.71 MPa respectively in the mode B. During the entire period of defrosting operation, the average discharge and suction pressures for NRCD were higher by 0.33 MPa and 0.14 MPa, than those for the RCD method. With a higher suction pressure, the risk of ASHP unit turning off due to low suction pressure could be avoided, and this resulted in higher operational reliability of the ASHP unit. Fig. 6 shows the degree of refrigerant superheat (DS) at suction point during the defrosting operation in the modes A and B. In mode A, from 0 s to 75 s the DS at suction point was below 0 °C, which was caused by the system reaching a balance in pressure, from 75 s to 120 s, it increased due to the energy transfer from the indoor coil and system interconnecting copper piping, thereafter, with the energy provided by the indoor coil decreasing

gradually, it dropped to 0 °C. In the mode B, at the initiation of defrosting, the DS at suction point sharply increased to 10.5 °C, then decreased slowly and reached 1 °C at the time instant of 50 s, thereafter, it remained at about 1 °C until the defrosting operation was completed. It can be seen that, compared to the RCD method, even though the refrigerant mass flow rate was larger with the NRCD method, the DS at suction point could be maintained above 0 °C for the energy provided by the PCM-HE during defrosting. Fig. 7 shows the air temperatures measured at the inlet and outlet of the indoor coil during the defrosting operation in the mode B. It can be seen that, indoor space heating during defrosting was realized with the NRCD method. At the initiation of the defrosting, the outlet air temperature decreased rapidly for the reason that the refrigerant pressure inside the indoor coil was higher than the discharge pressure of the compressor as a consequence of switching on/off the valves, due to which the hot refrigerant could enter the indoor coil with the indoor air fan in operation. As the time progressed, the refrigerant pressure in the indoor coil declined

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28

Mode A

20

o

Temperature ( C)

24 16

Defrosting Resume heating

12 8 4 0 -4

0

40

80

120

160

200

240

280

320

360

400

440

480

Times (s) 28

Defrosting

o

Temperature ( C)

26

Mode B

24 22 20 18 16 14

0

10

20

30

40

50

60

70

80

90

100

110

120

Times (s) Fig. 9. The measured liquid tube wall temperatures of the indoor coil during the whole operation in the modes A and B.

2500

Mode A Power (W)

2000

Defrosting Resume heating 1500 1000 500 0

0

40

80

120

160

200

240

280

320

360

400

440

480

Times (s) 2600

Power (W)

Defrosting

Mode B

2400 2200 2000 1800 1600 0

10

20

30

40

50

60

70

80

90

100

110

120

Times (s) Fig. 10. The measured input power to the compressor during the whole operation in the modes A and B.

while the compressor discharge pressure increased. This allowed more and more hot refrigerant gas to enter the indoor coil. At time instant of 30 s, the outlet air temperature reached the minimum which was equal to the inlet air temperature. The lowest outlet air temperature could be increased by lowering the pressure of the indoor coil before beginning the defrosting operation, in this way the hot gas could enter the indoor coil. Afterwards, the outlet air temperature continued to increase. At the end of defrosting, the outlet air temperature reached 26.5 °C while the inlet temperature remained at 20 °C. During the entire period of defrosting, the mean difference between the outlet and inlet air temperatures reached 4.1 °C, with

the air flow of indoor air fan manually adjusted to 0.18 m3/s. The energy supplied to the indoor space during defrosting was determined using Eq. (4) as 116.7 kJ. On the other hand, with the NRCD method, an additional electronic heater could be used to enhance space heating during the defrosting operation. Fig. 8 shows the variation of PCM temperature during the defrosting operation in the mode B. The temperature of PCM was at 37 °C as a liquid at the beginning of defrosting. Thereafter, the temperature decreased sharply to reach the solidifying temperature (19.8 °C) within 50 s, and then leveled off as the PCM changed state releasing the latent heat. However, the solidifying temperature plateau was not obvious, because the PCM was of commercial

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Z. Long et al. / Applied Energy 133 (2014) 101–111

p

Mode A Mode B

ķ

Ļ

ĸ

ķ

ĸ Ĺ

Ĺ´´

Ĺ´

ĺ

Ĺ

ĺ´

ĺ

h Fig. 11. The vapour compression cycles in the modes A and B.

11

Mode A Mode B

10

3.0

9 8

2.5

7

COP

Heating Capacity (kW)

Mode A Mode B

3.5

6 5

2.0 1.5

4 1.0

3 2

0.5

1 0

0.0 0

25

50

75

100

125

Times (min) Fig. 12. The heating capacity over a test period of 125 min in the modes A and B.

grade (purity between 98% and 99%), and the temperature sensor was enclosed by a copper pipe, which influenced actual temperature of the test point. At about 90 s of defrosting, the latent heat of the PCM was completely released, making the PCM into a solid. The PCM temperature decreased rapidly again due to the sensible heat transfer between the PCM and refrigerant. At the end of defrosting, the PCM reached a temperature of 16.1 °C. In the actual experiment, it was found that about 20% of the PCM was distant from the fin-tube and did not transfer energy during the defrosting operation. The energy supplied by the PCM during defrosting was determined using Eq. (5) as 618.3 kJ. Fig. 9 shows the refrigerant liquid tube wall temperature of the indoor coil during the whole operation in both the modes. This temperature can be used to determine the finishing point of the resuming space heating operation. Generally, there will be a period of time for resuming space heating due to the indoor coil attaining a low temperature after the defrosting operation. It can be seen

0

25

50

75

100

125

Times (min) Fig. 13. The system COP over a test period of 125 min in the modes A and B.

from this figure, that the time for resuming space heating in mode A was 130 s while no time was taken in the mode B. In the mode A, from 0 s to 340 s during the defrosting, the liquid tube wall temperature decreased from 24.3 °C to 2.4 °C due to the reversing operation. In the next 10 s, the temperature rapidly decreased from 2.4 °C to 1.1 °C due to the reduced compressor speed. In the next step of 50 s, it increased slowly because of the system reaching a balance in pressure. At time instant of 400 s, the tube wall temperature reached 2.9 °C, mean while the 4-way valve was reversed and the compressor speed was increased. At 470 s, the temperature reached 25.0 °C, the indoor air fan was turned on, and the experimental ASHP unit commenced heating the indoor space. In the mode B, during defrosting operation, the temperature decreased rapidly from 26.2 °C to 17.6 °C during the first 20 s for the reason that the hot refrigerant could not enter the indoor coil with the indoor air fan in operation. As time progressed, the temperature increased steadily. When the temperature reached 25.1 °C

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at 120 s, the defrosting operation ended, and the ASHP unit started to reheat the indoor space immediately after defrosting was completed. Therefore, the resuming space heating duration was 0 s. Fig. 10 shows the input power of the compressor for the operations in the modes A and B. At the beginning of defrosting, the compressor speed reduction which lasted for 60 s caused the rapid decreasing of the input power compared to mode A. With the operation of valves’ on and off which lasted for 10 s, same result was obtained for the mode B. At 60 s of defrosting in the mode A, the 4-way valve was reversed and the input power of the compressor increased steadily. At the end of defrosting operation at 340 s, the input power reached 1438 W, which was much lower than that provided during heating operation (at the end of heating, it was 2300 W) and then the system went into resuming the space heating operation. In the next 60 s, the input power decreased due to reduction of compressor speed. The time from 400 s to 470 s was the resuming space heating operation, during which the compressor speed was increased. This caused the input power of the compressor to increase from 143 W to 1157 W. In the mode B, from 10 s to 60 s during the defrosting operation, the input power of the compressor leveled at 2000 W. In the next 60 s, it increased slowly and finally reached 2338 W. It is obvious that the input power of the compressor in the mode B was much higher than that in the mode A. The defrosting duration in the mode B was much shorter than that of the mode A. In the mode A, the energy consumption during the defrosting operation and resuming space heating operation were 293.9 kJ and 70.8 kJ, respectively. In the mode B, the energy consumption during the defrosting operation was 262.8 kJ. These values were obtained using Eq. (2). Therefore, the total energy consumption in the mode B was reduced by 27.9%, compared to the mode A. Fig. 11 shows the vapour compression cycles in the modes A and B. In the mode A, the processes 1–2, 2–3, 3–4, 4–40 and 40 –1 took place in the outdoor coil, throttle mechanism (including EEV and capillary tube), indoor coil, liquid–gas separator and compressor, respectively. As pressure and temperature sensors were not installed at the inlet of the compressor, the point 40 was assumed. In the mode B, the processes 1–2, 2–30 , 3–4, 4–1, 1–5 and 5–300 took place in the outdoor coil, EEV, PCM-HE, compressor, indoor coil and capillary tube, respectively. Among them, the points 30 and 300 were assumed and converged on point 3. According to the experimental results of Dong [24], who studied on defrosting heat supplies and energy consumptions for a defrosting operation with ample defrosting energy, 59.4% of total energy consumption was used to melt frost during defrosting. In this experiment, the weight of thaw dripping was 1.07 kg in the mode B, so the total calculated defrosting energy consumption was 601.6 kJ. On the other hand, the energy provided by the compressor and PCM-HE was 184.0 kJ (electrical efficiency of compressor was 0.7) and 618.3 kJ, respectively, and the energy used to the space heating was 116.7 kJ. Therefore, the actual total defrosting energy consumption was 685.6 kJ. Compared with the calculated estimate, the relative error was less than 14%. Fig. 12 shows the heating capacity and Fig. 13 shows the dynamic COP during the whole period of operation of 125 min for both the modes. It can be seen from the figures, that during heating (frosting) operation, the trends of heating capacity and dynamic system COP had no significant differences in both the modes. The only difference noticed was that at the beginning of the heating operation, the increasing trend of the dynamic COP with the NRCD method was slower than that with the RCD method. This trend was caused by the PCM-HE, which could store some liquid refrigerant at low pressure after defrosting. The average system COP was 2.84 in the mode A and 2.82 in the mode B during the frosting operation. Therefore, the PCM-HE, which enclosed the compressor, had no significant effect on the system COP.

Compared to the RCD method, the total heating capacity and input power were increased by 14.2% and 12.6% respectively over a test period of 125 min. Further, the system COP was increased by 1.4%. Although there was little improvement in total energy efficiency, the indoor thermal comfort improved obviously with the NRCD method. 4. Conclusions A novel thermal energy storage based RCD method for the continuous heating during defrosting was developed and its performance was evaluated. The PCM-HE was integrated into an ASHP unit, for storing the heat dissipated by the compressor during heating operation and to serve as a heat source to solve the problem of insufficient energy for the defrosting. Comparative experiments using both the NRCD method and RCD method were carried out. The summary of results is as follows: 1. Compared to the RCD method, the defrosting duration was shortened by 65% with the NRCD method. This resulted from availability of more energy to melt the frost off the outdoor coil and to reduce the time for reversing 4-way valve during the defrosting operation. In addition, due to some of the hot gas flowing into the indoor coil during the defrosting, the resuming space heating duration was 0 s. 2. The average discharge and suction pressures for the NRCD method increased by 0.33 MPa and 0.14 MPa, respectively, compared to those for the RCD method. Therefore, the risk of shutting down the ASHP unit due to low suction pressure could be avoided, and make the ASHP unit to operate with higher reliability. 3. A novel defrosting method was demonstrated for the continuous space heating during defrosting. It operated with the lowest outlet air temperature equal to the inlet air temperature and an average temperature difference of 4.1 °C between the outlet and inlet air. 4. Compared to the RCD method, the input power was higher for the NRCD method. However, the operations of the defrosting and resuming space heating were shorter. Consequently, the total energy consumption for the defrosting and resuming space heating operations was reduced by 27.9%. 5. Compared to the RCD method, the total heating capacity and input power were increased by 14.2% and 12.6% respectively over the 125 min test period. The system COP was increased by 1.4%. A significant achievement of this study is the development of a method to heat the space continuously during defrosting; taking advantage of the compressor’s dissipated heat. While conducting this research few problems were encountered, which may be solved in a future study. The future tests may include methods for the effective integration of the PCM-HE with the compressor casing, achieving a higher outlet air temperature from the indoor coil during defrosting, and studying the defrosting period, which may impact the effect of defrosting. Acknowledgement The authors gratefully acknowledge the financial support provided by the National Key Technology R&D Program in the 12th Five Year Plan of China (No. 2012BAJ06B02). References [1] Aynur TN. Variable refrigerant flow system: a review. Energy Build 2010;42:1106–12.

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