Applied Energy 110 (2013) 9–16

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Applied Energy

journal homepage: www.elsevier .com/locate /apenergy

Continuous heating of an air-source heat pump during defrostingand improvement of energy efficiency

0306-2619/$ - see front matter � 2013 Elsevier Ltd. All rights reserved.http://dx.doi.org/10.1016/j.apenergy.2013.04.030

⇑ Corresponding author. Tel.: +82 51 510 2440.E-mail address: myha@pusan.ac.kr (M.Y. Ha).

Ji Young Jang a, Heung Hee Bae a, Seung Jun Lee a, Man Yeong Ha b,⇑a SAC R&D Laboratory, LG Electronics, Seongsan-Dong, Changwon City, Gyeongnam, Republic of Koreab School of Mechanical Engineering, Pusan National University, Jang Jeon 2-Dong, Geum Jeong Gu, Busan 609-735, Republic of Korea

h i g h l i g h t s

� The newly designed dual hot gas spray defrosting method was examined.� Uninterrupted heating of an air source heat pump during defrosting.� We compared newly designed dual hot gas and traditional reverse cycle defrost.� Total energy efficiency was increased by 8% compared to traditional method.

a r t i c l e i n f o

Article history:Received 24 September 2012Received in revised form 9 April 2013Accepted 10 April 2013Available online 6 May 2013

Keywords:Heat pumpDefrostHot gas bypass defrostingFrost

a b s t r a c t

During winter operation, an air-source heat pump extracts heat from the cold outside air and releases theheat into the living space. At certain outside air conditions, when it operates in heating mode, frost canform on the air-cooled heat exchanger, decreasing the heating performance. Conventionally, reverse-cycle defrosting (RCD) has been the common method of frost removal. But this method requires the inter-ruption of heating during defrosting, as well as a period of time to reheat the cooled pipes of the indoorunits after defrosting. In this study, a new technology called continuous heating was developed, whichutilize only a hot gas bypass valve to remove the frost from the outdoor heat exchanger and thus enablingthe supply of hot air to indoors without any interruption. For this, a new high temperature and low pres-sure gas bypass method was designed, which is differentiated from the common high pressure hot gasbypass methods by its use of low pressure. Various refrigerant mass flow distributions were examined,and the most effective defrosting mass flow was 50% in this case. Heating capacity was increased by17% because of continuous heating, and the cumulated energy efficiency was increased by 8% comparedto the traditional reverse cycle defrosting over 4 h including two defrost operations. Also, cumulatedenergy efficiency was increased by 27% compared to electronic heaters that supply the same heatingcapacity during defrosting. By this new technology, it has been proved that continuous heating andenergy savings could be achieved without adopting expensive technologies.

� 2013 Elsevier Ltd. All rights reserved.

1. Introduction

In recent years, the variable refrigerant flow heat pump usingan air-source has been widely used for space cooling and heatingin residential and commercial buildings, because these systemshave precise capacity control and individualized thermal comfortcapabilities with very high energy efficiency and low initial instal-lation cost compared to any other air conditioning method [1]. Avariable refrigerant flow heat pump is a refrigerant system thatvaries the refrigerant flow rate with the help of the variable speed

compressor and electronic valves to match the capacity of thesystem to the space cooling or heating loads in order to maintainthe zone air temperature at the set temperature [2].

During winters, under certain outside weather conditions, theair-source heat pump often operates with substantial frost forma-tion on the outdoor heat exchanger, which is used as an evapora-tor, and this frost layer has to be melted away periodically. Thisunavoidable defrosting requirement causes an interruption of in-door heating, lowering energy efficiency and heating capacity[3,4]. Many studies have been undertaken in an attempt to over-come this weakness of the air-to-air heat pump, and to delay frostformation. Mei et al. [5] suggested that by adding a moderateamount of heat to the refrigerant steam in the accumulator, the

10 J.Y. Jang et al. / Applied Energy 110 (2013) 9–16

evaporator coil temperature can be raised by several degrees. Byunet al. [6] tried to retard the formation and propagation of frostusing alternative operation of multi-hot gas bypass line intoevaporator inlet. Silva et al. [7] studied the fan characteristics asa frost layer grows with various types of tube–fin type heatexchangers. Mader and Thybo [8] proposed an active distributionvalve that is able to feed parallel evaporators passed individually,and that would regularly shut single evaporator circuits off in orderto remain frost-free. Kwak and Bei [9] applied an electric heater infront of an outdoor heat exchanger instead of an indoor unit asusual to enhance the heating capacity and delay frost formation.Although many studies have been performed on avoiding froston evaporator heat exchangers and delaying the defrosting control,the need for defrosting is still one of the serious shortcomings of airto air heat pumps.

Defrosting can be carried out in a number of ways: compressorshut down, electric heating, reverse-cycle, hot-gas bypass. Of these,reverse-cycle defrosting (RCD) is the most common method offrost removal, and various papers have studied the dynamic char-acteristics during the defrosting process on the air to air heat pump[10–14]. Another widely-used method is hot gas bypass defrosting.Hot gas from compressor discharge can be a good resource to meltfrost easily. The advantages of hot gas defrosting are lower refrig-erant noise, smaller indoor temperature fluctuation, and no coldblowing. But because of its low quantity of heat, defrosting timeshould be much longer than the RCD time [15]. Stoecker [16] inves-tigated the correct pipe size for carrying hot gas to defrost evapo-rators. The correct size will ensure that the resulting increase inspecific volume of the hot gas that occurs at low pressure doesnot cause prohibitive pressure drops in the pipes. Hoffenbeckeret al. [17] developed an application of a transient model for pre-dicting the heat and mass transfer effects associated with an indus-trial air-cooling evaporator during a hot gas defrost cycle. Recently,Choi et al. [18] proposed a novel dual hot gas bypass method to re-duce the defrosting time.

Despite the large number of studies on the performance of airsource heat pumps during defrosting operations, defrosting neces-sarily causes the periodic interruption of indoor heating and degra-dation in winter heating efficiency. As a result, the heating ratio isusually 80–85% in the actual field. Therefore, the objective of thisstudy is to derive a way to improve the heating seasonal perfor-mance and supply uninterrupted heating of an air source heatpump that requires defrosting.

2. Non-stop heating ideas with air-source heat pump

There are two general methods of providing uninterruptedheating or increasing the heating ratio of air source heat pumps.One method is to use an electric heater and the other is the appli-cation of discharge hot gas bypass. Fig. 1 shows a general cycle

Fig. 1. General cycle diagram of a non-stop heating model.

diagram of a non-stop heating cycle. An electronic heater that isdesigned by considering evaporating capacity is generally installedin domain r, and the heat it supplies becomes deficient due tofrost. To remove frost from the heat exchanger, a hot gas bypassline that is extracted from the compressor discharge to heat ex-changer is installed in domain s. Both can be installed togetheror separately depending upon the situation and the designpurposes.

The application of an electronic heater is a very powerful, sim-ple and direct method for achieving non-stop heating and when aheater situated outdoors performing the role of an evaporator canbe more effective than directly locating indoors performing therole of a condenser. But this approach requires a relatively largesize and additional expenses if the system capacity is increased.Therefore, almost products supplying continuous heating with aheater are smaller home appliances of 5 kW or lower. Also, thereare some safety issues to consider, such as the potential for shortcircuit, electric shock and fire. Meanwhile, a hot gas bypass is sim-ple and relatively low cost compared to the heater application, butit reduces the indoor side heating capacity as well as the refriger-ant flux for defrosting. Furthermore, it requires more time to re-move frost compared to reverse cycle defrosting. This studyproposes a dual spray method using the existing hot-gas bypassmethod for a continuous cycle design that can also be applied tolarger air source heat pump of 30 kW or higher, as well as smallerones.

3. Applied hot gas defrost using dual spray

3.1. Schematics of dual spray hot gas defrosting (DSHG)

The most effective and economical non-stop heating cycle withdual hot gas spray method was designed as shown in Fig. 2. Thismethod is the same as the general hot gas bypass method that ex-tracts hot gas from the compressor discharge, but the heat exchan-ger was divided into two parts, each having its own extra valve.Also, the existing single electronic expansion valve was increasedto two. In normal heating operations, two bypass valves are closedand two expansion valves are working equally. When defrostingstarts, the upper expansion valve is closed and the upper hot gasvalve is opened to thaw the upper side frost and prevent the inflowof the refrigerant that must be evaporated. After the upper sidedefrosting is finished, the lower side defrosting is performed withcounter valve operations. When defrosting of one side is in pro-gress, the heat exchanger on the other side acts as a de-magnifiedevaporator to provide uninterrupted heating. The greatest differ-ence between this method and the existing RCD is that all sidescarry out continuous heating, without stopping. Therefore, throughRPM control the outdoor fan raises the low pressure of the systemon the side of evaporator operating at half-power.

Meanwhile the control logic induces the compressor to operatein its maximum allowable frequency and higher pressure to pro-vide maximum heat supply. The magnitude of the indoor heatingsupply capacity in the defrosting range ultimately depends onthe frequency marginal rate of the compressor. The frequency mar-ginal rate is the difference between the present operation fre-quency and the maximum frequency. The high frequencymarginal rate attains when the outdoor temperature is high andleast indoor heating load. For this reason, it is advantageous formaintaining indoor heating capacity in the defrosting range. Onthe other hand, the proposed technique has the continuous heatingcycle without the frequency marginal rate, or it can restrict thecontinuous heating operation in the event that the indoor heatingcan be hardly supplied in the defrosting range, such as at a verylow outdoor temperature.

: Strainer

: Check valve

: Capillary tube

: Electronic expansion valve

: Electronic solenoid valve

: Pressure switch

: Pressure sensor

Service valve (Gas)

Service valve (Liquid)

Compressor 1

Oil Separator

Compressor 2

Heat Exchanger

Subcool unit

Four way valve

Accumulator

IDU1 IDU2 IDU3

Indoor Side: Condenser

Outdoor Side: Evaporator

IDU4

BLDC fan

Branch tube

Dual Spray Line

HEX1

HEX2

HEX3

: Temperature sensor

Fig. 2. Dual spray hot gas cycle diagram.

Fig. 3. Heating COP according to heating ratio.

J.Y. Jang et al. / Applied Energy 110 (2013) 9–16 11

3.2. Basic design of DSHG

Total system refrigerant fluxes are the sum of the indoor anddefrost bypass flow.

_msystem ¼ _mIndoor þ _mDefrost;bypass ð1Þ

Necessary total refrigerant flux can be calculated by systemcapacity and cycle parameters

Q system ¼ _msystem � Dh ð2Þ

Dh is the difference of inlet from outlet enthalpy of system. Whentotal mass flow is calculated, indoor mass flow can be determinedfor each design point. For example, 50% of total mass flow meansthis can provide half the heating capacity during the defrost. Afterbypass mass flow is determined, one can choose the valve orificesize that can pass a sufficient range of designed pressure conditions.The total energy efficiency including defrost can be expressed as,

COPdef ¼Q sys � a � a0 þ Q sys � bð2� a� a0Þ

Wsys � aþWdef ð1� aÞ ð3Þ

where a = nominal heating ratio, a0 = heating ratio under non-stopheating, b = heating supply ratio during defrost. A general air sourceheat pump system’s a = 0.8–0.85 and b = zero.

For example, 37.5 kW nominal heating capacity and 8.62 kWpower input model’s nominal COP is 4.35. (a ¼ 1;a0 ¼ 1; b ¼ 0).But energy efficiency would be decreased in a manner inverselyproportional to the heating ratio, a during the defrost condition(a < 1; a0 ¼ 1; b ¼ 0), as shown in Fig. 3. The energy efficiency willbe decreased by 15% at a = 0.8.

On the other hand, from the perspective of non-stop heating(a ¼ 1; a0 < 1; b–0), the energy efficiency will be a function of a0

and b. The estimated efficiency of non-stop heating was shownin Fig. 4. The a0 and b were increased along with the total efficiency.Assuming that b is 0.5, the non-stop heat pump can provide 50% ofnominal heating during frost with the same efficiency or higherunder the same heating ratio (a ¼ a0 ¼ 0:8).

3.3. The evaluation factors of continuous heating

There are some indexes to evaluate the durability of continuousheating performance. Fig. 5 shows normal and abnormal defrostcycle. Fig. 5a is the general cycle diagram of RCD and Fig. 5b and

Fig. 4. COP estimation during non-stop heating.

12 J.Y. Jang et al. / Applied Energy 110 (2013) 9–16

c are bad examples for DSHG. The definition of theta, h; is the slopeof the same event during the cycle. For example, as in Fig. 5, theslope of the defrost starting point is a very good choice. If h has anegative value, that means the cycle will collapse gradually andthe heating capacity will decline as well, as shown in Fig. 5b. Theother index s is the ratio of heating operation. If s is less thanone, that means the defrost period will occur too frequently andthe cumulated heating capacity could decrease, as shown in (c).So h must be greater than zero and s must be greater than one tomaintain the steady heating cycle.

s1 ¼ t2=t1; s2 ¼ t3=t2; � � � ð4Þ

The index of each cycle shown in Fig. 5 is (a) h : 0:18; s : 0:99,(b) h : �6:31; s : 1:02 and (c) h : 0:24; s : 0:71: In conclusion, this in-dex should be fully considered before the evaluation of continuousheating.

4. Experimental setup and test procedure

4.1. Test model and facility

As shown in Fig. 2, the test heat pump has two scroll-type com-pressors, which are composed of one inverter and one constantspeed type. A wide louver fin-type heat exchanger, two electronicexpansion valves, an accumulator and a reversing valve are the

Fig. 5. New evaluation fact

other main components of the test heat pump, and two newly-added solenoid valves control the dual-spray hot gas flow. The testheat pump has a 33.5 kW nominal cooling and 37.5 kW nominalheating capacity (12 hp). The nominal power input is 9.11 kW forcooling and 8.62 kW for heating. Accordingly, the coefficient ofperformance of the tested heat pump is 3.68 for cooling and 4.35for heating. Four cassette-type indoor units, each with a nominalheating capacity of 10 kW, were connected. Each connecting tubelength to indoor unit from branch was 30 m and total length ofpipe was 150 m.

The test heat pump was installed in a psychrometric calorime-ter as described in Fig. 6. The test facility is composed of two cham-bers: an indoor and an outdoor chamber as well. Both chamber canmeasure cooling capacity up to 87 kW and able to control the tem-perature from 60 �C to �20 �C. To achieve this, it has three coolingcoils of 87 kW in total for normal temperature and a cooling coil ofthe same capacity for low temperature.

In addition, it has a large heater of the same capacity and ahumidifying heater to control the relative humidity within 5–95%. In the indoor side, a large capacity duct of nozzle type is in-stalled to measure the amount of wind. It can measure up to130CMM. When measuring the specimen capacity, its allowableerror was calibrated to be made under 2%. The tested psychromet-ric type calorimeter satisfied the ISO standard. The data acquisitionsystem automatically collected all experiment data, including tem-perature, humidity, wind amount and power consumption, andstored them continuously during the test in order to avoid any hu-man error in data reading. The data recording instrument provided64 data channels, which scanned all data channels consecutivelyevery 2 s. To reduce experimental uncertainties, Null points of alltemperature and pressure sensors were adjusted with the calibra-tions. The temperature sensors were calibrated with a standardwater bath over a temperature range of �30 �C to 120 �C with anerror of ±0.2 �C. The relative error of the pressure sensors overthe effective measurement range of 0–25 bars was ±0.2%. The rel-ative error of the power meter was ±0.1% of full scale. The time sig-nal had a relative error of ±0.1%.

4.2. Test procedure

The outdoor test condition was 2 �C dry-bulb and 1 �C wet-bulbtemperature, and indoor was 20 �C dry-bulb and 15 �C wet-bulbtemperature, and this test condition meets the Korean defrostingstandard KS B SIO 15042 [19]. The test was carried out with theabove-mentioned specimen. The specimen was designed to controlboth RCD and DSHG during the defrosting operation. To accurately

ors during defrosting.

Indoor Unit

Outdoor Unit

FAN

CoolingCoil

Heater

Humidifier

Air-sampler

Nozzle chamber

InductedFan

Temp. & Humid.Measuring sensor

Pre-Heater

Mass flowmeter

Connectingtube

Fig. 6. Psychrometric test facility.

Fig. 7. RCD cycle.

J.Y. Jang et al. / Applied Energy 110 (2013) 9–16 13

compare the two different types of defrosting performance, opera-tions were performed on the same specimen.

In the operation state of 2 �C/1 �C (DB/WB), under full load con-dition, the heat evaporating exchanger fin was covered with frostdue to moisture in the air about 80–90 min after the systemstarted. The defrosting operation was determined by the tempera-ture difference between the evaporator and the outdoor air. Thelow pressure sensor data was another basis for defrost starting.In both RCD and DSHG, the temperature sensor to check thedefrosting start was the temperature sensor HEX1 of evaporatorside in Fig. 2. The escape condition of the defrosting operationwas checked using the temperature of the outdoor heat exchanger.At this time, both have different directions of entrance and exit ofoutdoor heat exchanger during the defrosting operation and thedefrosting end should be judged based on the exit temperature,RCD used HEX1 and DSHG used HEX2, 3. In DSHG, HEX 2 judgesthe defrosting end of upper-side heat exchanger, while HEX 3judges that of lower-side heat exchanger. RCD was first performedunder the control logic mentioned above. As a result, it took about4 h for three defrosting cycles. The defrosting started when the dif-ference in temperature between the evaporator and the outdoor airwas 10 �C, and ended when the outdoor heat exchanger tempera-ture was 25 �C. To compare with RCD, the newly designed DSHGwas performed with same defrost judgment logic and test time,but a different temperature due to its heat limit. After finding theoptimized DSHG operation, starting and ending temperatures weremaintained as 8 �C and5 �C respectively over 3 min. This earlierstage defrosting concept is referred to in the next chapter.

5. Test results

5.1. RCD method results

The key factor of RCD defrosting is to change the reverse valvedirection from heating to cooling. Once the defrosting operationstarts, it is necessary to reduce the rotation speed of the invertercompressor. If a constant-speed compressor is in operation, it shutsdown to prevent a drastic pressure change. When the reverse valveis switched to defrosting, the discharge pressure is dropped sharplyand the suction pressure is increased. Finally, the high pressureand the low pressure meet to achieve a balance, and it is an inev-itable phenomenon that an outdoor heat exchanger which used asa low pressure evaporator will be switched to a high pressure con-denser. On the other hand, the indoor heat exchanger that operatedas a high pressure condenser changes its role to function as a low

pressure evaporator. At this time, the indoor fan shuts down to cutoff the supply of cool air into the room. The outdoor fan is also usu-ally turned off to achieve a rapid pressure increase. Fig. 7 shows thedynamic characteristics of the RCD test cycle. The temperature ofthe HEX sharply increased due to the incoming superheated hotgas from the compressor discharge. During the defrosting stage,the discharge pressure increased slightly, but the suction pressuredecreased noticeably, from 800 kPa to 300 kPa. As mentionedabove, when the fan of the indoor evaporator is not operating, bothstart and end of defrosting operation will be judged by the temper-ature sensor of HEX1, the defrosting range temperature of HEX1rises rapidly to 0 �C and stays there for a certain period of time be-fore it goes up again rapidly. This is because the frost melts intowater at the temperature 0 �C. Once it reaches 25 �C, the set-uptemperature for ending defrost, it enters the defrosting end step,lowering the inverter frequency and switching off the 4-way valveto stop the whole defrosting operation. The defrosting time is usu-ally counted up to the switching-off point of the 4-way valve. Inthe case of this specimen, the defrosting operation was carriedout for 5 min, from about 400–700 s, as shown in Fig. 7. Thisdefrosting time is much shorter than that of hot gas defrosting,and it is one of the greatest advantages of the RCD method, in thatdefrosting can be achieved without much additional cost.

Fig. 9. Heating capacity results.

14 J.Y. Jang et al. / Applied Energy 110 (2013) 9–16

The RCD method uses two energy sources for defrosting. One isthe power input of the compressor, and the other is the evapora-tion heat of the indoor side. Most of defrosting depends on thepower input of compressor, but some of the indoor evaporationheat caused by natural convection brings about the cooling of in-door air that has been deprived of heat source in the defrostingoperation range, lowering the indoor temperature about 2–5 �C.Once the defrosting is completed and the heating operation is car-ried out again, a certain period of time is needed to heat the indoorheat exchanger and the ducts. As shown in Fig. 7, the high pressureof the system is stabilized about 700 s after the start of re-heating,and it takes more than 10 min until it reaches 2500 kPa or higher inthe normal cycle before defrosting. So, we can see that the greatestdefect of RCD type is the duration of a certain time from the stop ofindoor heating supply in the defrosting range until the lowering ofindoor temperature and the supply of normal heating after defrost-ing. The duration time for three cycles of defrosting operation wasabout 4 h in total. For fair comparison and analysis, the outdoorchamber temperature was sufficiently raised to dry the outdoorheat exchanger and the measuring temperature was reset; then,the test was carried out using the DSHG method for 4 h.

5.2. DSHG method results

Fig. 8 shows the cycle characteristics of DSHG defrosting range.Like RCD, its defrosting start also judged based on the value givenby temperature sensor HEX1 on the evaporator side. But, a differ-ent logic is used for the starting condition to enter the defrostingoperation earlier than RCD. This is because the DSHG method islimited in terms of its capacity to melt the frost layer due to thefact that less energy is provided to the defrosting heat source thanin the RCD method, and so the defrosting process must be carriedout before the frost layer gets too thick. Actually, the enthalpy dif-ference in the defrosting range estimated from the temperatureand the pressure is so great that DSHG method is about 25% ofRCD method. If this early stage defrosting method is not used butDSHG method is applied in the same way for defrosting as RCD,the frost will not be completely removed, and it will fail to achievethe continuous defrosting cycle in (b) or (c) of Fig. 5. Therefore,according to the preliminary checkup on the optimum start condi-tion, the logic was adopted to start the defrosting when the differ-ence between evaporator and outdoor temperatures became 8 �C,2 �C lower than that of RCD method, which usually starts thedefrosting at thermal difference of 10 �C. This early stage

Fig. 8. DSHG cycle.

defrosting results in more frequent defrosting operations thanRCD during the same time duration.

When the DSHG method is used, the circulation flux of refriger-ant is, as shown in Eq. (1), divided into the indoor flux for the ac-tual heating and the defrosting flux. The balance of each flux canbe adjusted according to the designed orifice size of dual spray lineand the expansion valve of outdoor heat exchanger. The increaseof indoor flux will naturally be better for the indoor heating supplyin the defrosting range, but it causes a decrease in defrosting fluxand elongates the defrosting time, increasing the possibility ofdefrosting failure. On the other hand, if the indoor flux is reduced,it can shorten the defrosting time and raise the defrosting capacity,but it lowers the indoor heating capacity in the defrosting range,diminishing the benefit of continuous heating. The optimum fluxratio designed through the preliminary checkup is to keep the in-door heating capacity in the defrosting range to be 50–55% of nor-mal operation right before the start of defrosting, and thus todivide the refrigerant flux into about 5:5. The defrosting processbegins with opening the hot gas valve on the top and closing theexpansion valve on the top to remove the frost on the upper heatexchanger. The defrosting end of the upper side is judged by thesensor HEX2 that checks the outlet temperature of the upper side,and the optimum temperature was set as about 5 �C. The reasonwhy the defrosting temperature is relatively lower than that ofRCD is the same as the reason why its defrosting capacity is lower:

Fig. 10. Heating COP results.

Fig. 11. Cumulated performance results (a) capacity and (b) power input.

Fig. 12. Cumulated COP results.

J.Y. Jang et al. / Applied Energy 110 (2013) 9–16 15

the enthalpy difference that was mentioned above. Once the upperdefrosting is completed, the lower side defrosting is carried outwith the opposite valves. Likewise, the defrosting end of the lowerside is judged by sensor HEX2, which checks the outlet tempera-ture of the lower side. In the case of DSHG, the high pressure ofthe system in the defrosting range drops from 2800 kPa G to2200 kPa G, as shown in Fig. 8. This corresponds to the case inwhich the heating effusion temperature of the indoor side goesdown about 10 �C, from 43 �C to 33 �C. It is a natural consequenceof the fact that the indoor heating flux dropped down to about 50%of regular state, but it may cause unpleasantness for users. So, anadditional logic is necessary to reduce the RPM of indoor fan in thedefrosting range and raise the outlet temperature. Total time re-quired to finish defrosting was about 10 min, while the reverse cy-cle defrost was about 5 min with the approach of 4-way valveswitching. As mentioned above, in RCD, when the defrosting oper-ation is carried out for 5 min from the point of 4-way valve switch-ing, the indoor heating supply is stopped, and it takes about10 min to resume the normal heating state after defrosting. But,in DSHG, while the defrosting is carried out for about 10 min,50–55% of heating is supplied to the indoors and the high temper-ature of the system is, as shown in Fig. 8, stabilized immediatelyafter the defrosting and recovers the normal heating supply. Ofcourse, its defrosting capacity is lower than that of RCD, and thusthe condition to use DSHG method can be restricted.

5.3. Performance comparison and discussion

Fig. 9 shows the heating capacity results between common RCDand DSHG that we designed. The instant COP results are alsoshown in Fig. 10. The RCD method should shut down the indoorheating periodically to defrost. The designed DSHG method couldprovide about 50% of nominal heating during defrost.

So, total heating capacity was increased by 17% compared to thereverse cycle defrosting, as shown in Fig. 9. There were two reversecycle defrosting and four hot gas defrosting over approximately4 h. The index noted was h : 0:24; s : 0:96; this means that thenon-stop heating cycle would remain stable. As we have seen inFigs. 9 and 10, there were more defrost operations using DSHGcompared to RCD defrost. The hot gas method takes a long timeto remove frost with a relatively low defrost heat because it shares

heat energy with defrost and indoor heating supply. Therefore, touse the hot gas method, as mentioned previously, earlier stage de-frost is required before a full growth of frost, and this is why ahigher hot gas defrost frequency appears under the same condi-tions. The cumulated performance indicators are shown inFig. 11. The slope of the graph was almost the same between thetwo methods, but the slope of the reverse cycle defrosting was zeroduring defrost because of the indoor heating interruption. On theother hand, the slope of the hot gas defrosting that was designedto supply 50% of nominal heating capacity during defrost was uni-form, without variations. The cumulated performance of DSHG wasincreased by 17% compared to the reverse cycle defrosting during4 h. The cumulated power input was increased by 7.8% in the samemechanism. So, total cumulated energy efficiency was increased by8%, and compared with using an electronic heater that covered thelack of heating capacity, and total efficiency was increased by 27%,as shown in Fig. 12.

Fig. 13 shows the photographs of dual hot gas spray defrost.When heat exchanger frost has grown and the signal has reachedthe designed condition, one-side defrosting begins. Generally,upper side defrosting should be started first, because the coldwater that has been melted from the upper side heat exchanger

(a) Defrost start (b) Half progress (c) Defrost ending

Fig. 13. Dual spray defrosting results.

16 J.Y. Jang et al. / Applied Energy 110 (2013) 9–16

may freeze again at the lower side. But repeated upper sidedefrosting can have deleterious effects on the lower side heat ex-changer, such as causing frost to remain on the lower side. In somecases, lower side defrosting should be started first, as shown inFig. 13b.

6. Conclusions

In this study, a newly designed dual hot gas spray defrostingmethod was examined, and the defrost performance was com-pared to a traditional reverse cycle defrosting method. The resultscan be summarized as follows:

(1) A simple and low cost dual hot gas defrosting cycle structurewas introduced and the most effective bypassed hot gas flowrate achieved was about 50% of the total refrigerant flow rateduring the defrost.

(2) The total heating capacity was increased by 17% and theinput power was increased by 7.8% over 4 h test period.Finally, the total energy efficiency was increased by 8% com-pared to reverse cycle defrosting.

(3) The total efficiency could be increased by 27% comparedwith system using an electronic heater that covered the lackof heating capacity under the reverse cycle defrostingmethod.

The main achievement of this study is the discovery of a meth-od to supply heating continuously in the large capacity cycle of30 kW or higher without any additional equipment, through theoptimum control of hot gases, overcoming the problem of the needto cut off the heating supply in the defrosting stage, which hasbeen the greatest shortcoming of air source heat pump. But, asthe indoor heating supply was just about 50% of normal operationin the defrosting range when the outdoor temperature was 2 �Cand the heating is provided in all rooms, it may cause complaintsby users due to the lowered outlet temperature of the indoor facil-ity. Furthermore, if the outdoor temperature drops lower than 2 �C,the indoor heating supply in the defrosting range may drop below50%. Therefore, for the application to the actual product, it may re-quire the control of diverse boundary conditions. For example, itsuse may be confined to cases when it is not the whole room indoor

operation, or when there is some margin in the frequency of the in-verter compressor in normal operation.

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