performance of an air-to-air heat pump under frosting and defrosting conditions

ELSEVIER

Applied Energy 48 (1994) 225 241 © 1994 Elsevier Science Limited

Printed in Great Britain. All rights reserved 0306-2619/94/$7.00

Performance of an Air-to-Air Heat Pump Under Frosting and Defrosting Conditions

Kamil Kaygusuz

Karadeniz Technical University, Department of Chemistry, Trabzon, 61080, Turkey

A BSTRA CT

In order to investigate the perjormance of air-to-air heat pumps for residential heating in the Black Sea region of Turkey, an experimental set-up was constructed. An electrically driven air-to-air heat pump was used ./'or heating a laboratory building with a floor of 75 m 2. In this study, a high- efficiency air-to-air residential heat pump was instrumented and tested in the heating mode under laboratory conditions. The coefficient of performance and heating capacity of the system were measured during steady- state, dehumidifying, and frosting~defrosting conditions. The s'tudy encompassed an evaluation of system and component performance jor ambient temperature levels" between 10 and -4.0°C and for discrete relative-humidity levels ranging from 65 to 80%.

A COP (DD)m hi 7l PLF

Q~on QL SPF Ta Tb Tind

T,

NOTATION

Area (m 2) coefficient of performance Monthly degree-days Hours of occurrence of each temperature bin Number of days in the month Past-load factor Condenser heat output (kWh) Average seasonal heating load (kWh) Seasonal performance factor of the heat pump Ambient-air temperature (K) Base temperature (291.4 K) Indoor-air temperature (K) Inlet refrigerant temperature of condenser (K)

225

226 Kamil Kaygusuz

U WsH

Outlet refrigerant temperature of condenser (K) Overall thermal loss coefficient (kWh/m 2 K day) Supplementary heat in each temperature bin (kcal)

INTRODUCTION

The heat pump is a natural candidate for efficient space heating because of its ability to augment input energy with heat extracted from the envir- onment. Its capability also to provide summer cooling, needed in many parts of Turkey (i.e. Adana, Antalya, Izmir, etc.), is an added attraction. Despite these advantages, the viability of heat pumps has been ques- tioned due to the need for information on such factors as source energy losses in the conversion of thermal to electrical energy, reduced capacity and efficiency at low outdoor-air temperatures, first cost relative to gas furnaces, records of reliability, quality of service, and comfort. While some criticisms of heat pumps may have had a justifiable basis in the past, a number of recent developments have suggested that heat pumps are a viable alternative to fossil-fuel heating systems. 14 These developments include changing economic factors, such as cost and availability of fuels of various types; improved heat-pump designs, which result in higher efficiencies and improved reliability; and the general maturation of the manufacturing and service industries.

On the other hand, for many situations, electric heat pumps would be an obvious choice for space conditioning of buildings, since they can op- erate at higher heating efficiencies than conventional gas-fired furnaces. Because the heating output of a heat pump declines, however, supplementary heating is required when the ambient air is very cold. As a result, an all-electric heat pump is usually installed with an electric resistance heater which serves as a back-up when the building's heating load exceeds the heat-pump's capacity.

In recent years, residential heat pumps have received increasing atten- tion in the Japanese, UK and US markets from the viewpoints of energy saving and the effective use of facilities, and they have been installed in many homes. 5 8 However, conventional air-source heat-pump systems have disadvantages inherent in the nature of the heat source. For example: (1) decreasing outdoor temperature lowers the heating capacity which means that heat demands of the building are not satisfied; (2) to prevent the efficiency loss due to frosting of the outdoor coils during the heating operation, periodic defrosting operations are required (especially in cold climates such as Kars, Ankara, Sivas and Erzurum in Turkey), spoiling the thermal comfort. Moreover, under weather conditions in which

Air-to-air heat pump under frosting~defrosting conditions 227

the heating capacity exceeds the heat demand of the building, partial- load operation causes some loss in the on/off cycle. 9 13

In this study, a higher efficiency air-to-air residential heat pump was instrumented and tested in the heating mode under laboratory conditions. The coefficient of performance (COP) and heating capacity of the system were measured during steady-state dehumidifying and frosting/ defrosting conditions. The seasonal heating performance and required electrical energy demand for the heating of the laboratory building were- calculated according to the experimental results. Also, seasonal analyses were conducted for the determination of the magnitude of frosting losses and defrosting losses for the heat pump.

E X P E R I M E N T A L SET-UP

The air-to-air heat-pump system connected to the laboratory building is shown in Fig. 1. It consists of a compressor four-way valve, outdoor heat-exchanger and associated measuring and controlling equipment. The hermetic type compressor was driven by a 1490 W electrical motor. The heat-pump's compressor characteristics are given in Table 1.

The split-system heat pump used in the experimental study was installed in two separate air loops; one loop housing the indoor unit and the other housing the outdoor unit. Thermocouple grids were used to measure the average air temperature entering and leaving each individual

TABLE 1 Main Heat-Pump Compressor Characteristics 15

Manufacturer Model Nominal voltage Capacity rating Power consumption

evaporating temperature condensing temperature suction-gas temperature liquid-subcooling ambient-air temperature

R -22 Displacement 2900 rpm Oil charge Refrigerant charge Maximum working-pressure Net weight

D.W.M Copeland (Europe) DCRD 1-0200 220 240/1/50 V 5820 W (under following conditions) 2.07 kW (under following conditions) 7.2oc 54.4°C 35.0°C 8-3 K 35.0°C 50 Hz 7.16 m3/h 1.5 litre 2.7 kg (recommended maximum) 27 bar 28 kg

228 Kamil Kaygusuz

ITll ~ w

g:>

. _ J

l ~ ~ o ~ °

oc;~

t

~5

I ~ m00 ~

0

> 0

o r ~

;.r,,


TABLE 2 Structural Properties of the Building

Window area (single glazing, U = 4.8 W/m2°C) Wall area (single brick, U = 1.6 W/m2°C) Floor area (concrete, U -- 2-5 W/m2°C) Ceiling area (concrete + metal, U = 2-0 W/m-~°C) Effective UA (kWh/°C day) Comfort temperature Average degree-days for heating season Average total heating load in heating season

75 m 2 60 m 2 75 m 2 75 m 2 0.800 20°C 1750 21500 kWh

unit. Relative humidity was measured by two hygrosensors. One sensor, used for humidity control, was positioned at the entrance of the outdoor unit, and another sensor was placed just in the stream of the outdoor unit. The airflow across each unit was measured by using orifice meter plates installed in both test loops. The refrigerant flowrate was measured using a flowmeter. Refrigerant temperatures and pressures were measured respectively with thermocouples and pressure transducers connected at various strategic locations in the refrigerant circuit.

HEATING-LOAD CALCULATION

In order to evaluate the performance of the conventional air-to-air heat pump and the other heating systems, the heating requirements of the laboratory building have been assessed. The laboratory building has a reinforced-concrete construction without insulation and its floor area is 75 m 2. The structural properties of the building are given in Table 2. The climatic conditions for Trabzon are given in Table 3.

One of the most widely used procedures for calculating the heating-

TABLE 3 Climatic Conditions of Trabzon During the Heating Season

Average outdoor temperature (°C) Minimum outdoor temperature (°C) Average minimum outdoor temperature (°C) Maximum outdoor temperature (°C) Average maximum outdoor temperature (°C) Average relative humidity (%) Average wind speed (m/s) Average solar insolation (MJ/m 2 day)

9"35 - 3"40

2"30 29-10 22"91 74"85

2"50 5"10

230 Kamil Kaygusuz

energy requirements of buildings is the degree-day method. It is based on a combination of theory and empirical observation and assumes that fuel consumption for heating will be proportional to the difference between the mean daily temperature and 18.3°C. Thus, the degree-day can be ex- pressed in simple form a s : 14

(DD)m = n (T b - Ta) (1)

and then the monthly heating load is given by:

QL = ( UA)h (DD)m (2)

Thus, the average monthly space heating load, QL, was calculated using eqns (1) and (2); the results for the heating season being given in Table 2.

HEAT-PUMP PERFORMANCE

The steady-state COP and the capacity of the heat-pump system chosen for this study are shown in Fig. 2 according to the experimental results. Both the COP and capacity are assumed to be linear functions of the outdoor temperature. The performance of a heat pump in an actual installation, however, may deviate considerably from the steady-state

==

9

5

3

I

0 -" - -4 -2 0

~ PUMP CAPACITY

HEATING jLOAD,

4 6 8 I0 12 14 16

2

Fig. 2. OUTDOOR TEMPERATURE (C)

Variations of the heating load, heat-pump capacity and its COP with outdoor temperature.

Air-to-air heat pump under frosting/deJrosting conditions 231

performance due to the transient losses at start-up, the losses caused by frosting and defrosting of the outdoor coil, and the addition of supplementary heat.

The actual performance data obtained from experimental results are used to generate third-order polynomials relating the heat-pump's COP and the condenser output to the outdoor-air temperature. The steady- state COP of the heat pump is given by: ~5

COP=-27.99+O.121Ta+ 1-601 × 104T~ - 7.035 × 107T~ (3)

Also, the-condenser output is given by:

Qcon = 18.45 - 0-1019T~ + 6.508 × 10 5T,~ + 5.044 × 10 7T~ (4)

Cycling losses

These occur in a heat pump because, when switched on, its output does not reach steady-state conditions instantaneously but follows a transient response with a time constant, which depends on the design of the system. The laboratory air-to-air heat pump, that was used in this study, reached steady-state conditions within 20 minutes of being switched on.

Figure 2 shows the variation of the heating load and the heat-pump's capacity with outdoor-air temperature. At temperatures below the balance-point temperature, where the heat-pump capacity matches the heating load, the output of the unit is less than the load and therefore the system operates continuously with no on/off cycling losses. The addi- tional heat requirement is generally provided by an auxiliary heating source (i.e. electrical resistance heaters or oil gas furnace). At temperatures above the balance-point temperature, the capacity of the heat pump exceeds the heat requirement, and therefore the unit cycles on and off to match the load. The frequency of the on/off cycling and the resulting efficiency losses are functions of the steady capacity of the heat pump and the building's heating load.

Frosting and defrosting

When an air-source heat pump is operated at temperatures below 4°C, the surface temperature of the outdoor coil falls below 0°C and frost is formed on the surface of the coil. The frost formation causes a decrease in the heating capacity, and periodical defrosting is needed. The indoor coil fan is stopped during defrosting, and no supplementary heat is used. An air-source heat pump heating unit takes about five minutes to defrost the outdoor coil, and during this period the heating is stopped. The aver-

232 Kamil Kaygusuz

age heating capacity is also reduced due to the interruption of heating during defrost.

SEASONAL P E R F O R M A N C E

This factor for the heat pump is defined as:

Total energy delivered during the heating season SPF =

Total energy input during the heating season

By determining the energy delivered and the energy input for each hour of operation of the system, and summing these over the whole heating season, the SPF becomes: ~5

Z~i':, QLhi S P F : (5)

i N, QLh i/COP PLF + Z i N, Wsahi

E X P E R I M E N T A L P R O C E D U R E

Steady-state experiments

Dry-coil testing started once the heat pump had achieved steady-state indoor and outdoor ambient-air conditions. The data acquisition system (DAS) then monitored and recorded refrigerant (R-22) temperatures and pressures, plus power consumptions and refrigerant mass-flow rates. These were recorded as average values within the time period of a number of oscillations of the refrigerant flowrate.

With outdoor air temperatures of 3 and -3°C, and relative humidities greater than 65%, frosting of the outdoor heat exchanger began before data collection could start. Thus, direct measurement of steady-state performance at these temperature and humidity levels was not possible. The steady-state performance was extrapolated from the start of each experi- ment by using the slope of curves representing reduced test data per respective temperature and relative humidity test run.

Frosting experiments

These were performed to observe the effect of relative humidity and temperature of the outdoor air on the heat-pump's COP and heating capacity. The time of frost initiation and the duration of each frosting test were determined visually and noted for each test run. The DAS was


again used to monitor and record the performance parameters, as was done in the steady-state tests. The airflow across the outdoor unit was also recorded and used as the criterion for defrost initiation of the heat exchanger in the outdoor unit.

Defrosting test-procedure

Defrosting tests were begun at the termination of each frosting test. The time required for defrosting was noted, and instantaneous values of the compressor power consumption, compressor high-and-low pressures and temperatures, and the temperature of the refrigerant exiting the outdoor heat exchanger were monitored and recorded at 10 s intervals by the DAS. Upon completion of the defrost cycle, the heat pump was manu- ally restored to the heating-mode operation.

RESULTS AND DISCUSSION

The heating mode performance of a heat pump is usually characterized by the COP and heating capacity. These parameters are affected by the outdoor heat-exchanger capacity, which is, in turn, dependent on the driving potentials of ambient-air temperature and humidity levels. The system's performance was examined in terms of these driving potentials for ambient conditions that (i) allow for a dry evaporator surface, (ii) produce a wet surface, and (iii) produce frosting on the coil.

The COP and outdoor heat-exchanger capacity increased linearly with

Fig. 3a.

3

t j

2

• Refrigerant side measurement

0 Air side measurement

o , I , I , I , I , I , I 4 -2 0 2 4 6 8

OUTDOOR AIR TEMPERATURE (C)

Steady-state system performance measured with a dry outdoor heat-exchanger.

234 Kamil Kaygusuz

Fig. 3b.

4 butdoor heat exchanger U capacity

-4

• Refriger~nt side measurement

0 Air side measuremed

, i , I , I , I , i , I -2 0 2 4 6 B

OUTDOOR AIR TEMPERATURE (C)

Steady-state heating capacity measured with a dry outdoor heat-exchanger.

increasing ambient temperature under dry coil conditions, as shown in Figs 3a and 3b. The system's COP increased from 1-84 at -4 .0°C to 2.82 at 8.0°C; and the rate of heating increased from 3.80 kW at -4 .0°C to 6.35 kW at 8.0°C.

The system's COP and heating capability improved slightly under conditions of high relative humidity in the non-frosting range. With a 10°C ambient air temperature and a relative humidity of 65% or greater, the outdoor coil's surface temperature is below the dew point of the air but above the freezing point of water. Thus, the coil is wet but without frost

3

O_ O t )

2

Jr 85 % ambient-air relative humidity • 75 % ambient--~ir relative humidity

0 65 % ambient-~ir relative humidity

30 60 90 120 150 OPERATION TIME (rain)

I 1~0

Fig. 4. COP for a wet outdoor heat-exchanger at an ambient temperature of IO°C.


8,0

~ 6.0

~ 4.0

2.0 -- -~- 85 % ambient-air relative humidity

• 75~ ambient-air relative humidity

0 65 % ambient_air relative humidity

o.o , I , I , i , I , I , I 0 30 60 90 120 150 180

OPERATING TIME (rain)

Fig . 5. Heating capacity measured at 10°C.

deposition. As shown in Figs 4 and 5, the average COP at this temperature was observed to increase from 2.85 to 3.25 as the relative humidity was increased from 65 to 80%. The corresponding heating capacities were 9.60 and 10.3 kW.

The improvements in COP and heating capacity mentioned above are the result of an increased outdoor heat-exchanger capacity. For a constant dry-bulb temperature, an increase in the relative humidity yields higher mass-transfer rates and an increased latent-heat contribution to the capacity of the outdoor heat-exchanger.

3.0

2.5

cl

O o

2.0

1.5

1.0

• 65 ;~ ambient~ir rela[ive humidity

0 75 7, ambient-air relative humidity

~-85 ~ ambient-~ir re]at[ve humidity

, I , I , I , I , J , 30 60 90 120 150


Fig . 6. C O P measured at 4 ° C .

1 IBO

236 Kamil Kaygusuz

4

u

2 -- • 65 ~'i ambient-air relative humidity

0 75 f ambient_air relative humidity

Jr 85 7 ambient-air relative humidity

o , i , 1 , I , I , [ , I 30 60 90 120 150 180


Fig. 7. Heating capacity measured at 4°C.

At an ambient temperature of 4-0°C, a temperature favourable to frosting, the system COP and heating capacity decreased for relative humidity levels greater than 65% (see Figs 6 and 7). The increased moisture content in the air increased the rate of performance degradation because of the increased frosting rates on the outdoor heat-exchanger.

The COP at 65% humidity decreased from an initial value of 2-70 to 2.45, a 0.9% reduction within 180 min of operation. At the higher humidity of 75%, the COP decreased from 2.55 to 2.25, i.e., an 11% reduc-

3.0 r

2 . 5 ~ . • _

"' / ~ o----------o~ z° F

1.0 0

• 65 % ambient~air relative humidity

0 75 % ambient-air relative humidity

-~ 85 % ambient-air relative humidity

, L , J , i , I I J 30 60 90 120 150


Fig. 8. C O P measured at - 2 ° C .

, I 180


8.0

6.0

v

n~4,0 <

2.0

0.0

• 65 % ambient-air relative humidity

O 75 % ambient-air relative humidity

q- 85 % ambient-,l~r l-elative h u m i d i t , '

, I , I L l , I J J I 30 60 90 120 180

OPERATING TIME

Fig. 9.

I 150

(rain)

Heating capacity measured at - 2 ° C .

tion within 180 min of operation. The COP degraded by 20% of the steady-state value within 180 min of the start of the heat-pump operation for the 85% relative-humidity test conducted at 4.0°C. The heating capacity observed during the 65, 75 and 85% relative humidity tests decreased by 29, 40 and 49%, respectively (see Fig. 7).

Figures 8 and 9 show the variations of COP and heating capacity with time. As shown in Fig. 8, the COP at 65°/,, humidity decreased from an

1 0 0

~" 60 W

n -

Ae" 0 0

0

q

40--

20--

0

Fig. I0.

nont rosting z)ne

I I 0 5

(XJTDOOR DRY-BULB TE~F~RATURE (°C)

Ambient conditions for defrosting.

10

238 Kamil Kaygusuz

150

E • £: i 20 v

J ~ 9o o z_ I..- ffl

~ 60 W 121

30

Fig. 11.

0 5O

0 O ~ 4 ° C

Jr - 3 ~ 3 ° C

oo o°o o %Oo

oo+ " " ~ o + -t- +

°"~e~. + +

, I , I , I , I , 60 70 80 90 100

OUTDOOR-AIR RELATIVE HUMIDITY (°/o)

Outdoor relative humidity and defrosting interval.

initial value of 2.50 to 2.25, i.e. a 10% reduction within 180 min of operation at an ambient temperature of -2.0°C. At the higher humidity of 75%, the COP decreased from 2-40 to 2.12, i.e. an 11% reduction within 180 min of operation while the COP degraded by 15% of the steady-state value within 180 min of the heat-pump operation for the 85% relative humidity test conducted at -2-0°C. The heating capacity observed during the 65, 75 and 85% relative-humidity tests decreased by 9, 16 and 28%, respectively (Fig. 9).

Figure 10 shows the relationships between outdoor-air temperature and outdoor relative-humidity, when defrosting occurs, for this case in which a residential heat-pump is controlled by the defrost system. There is no defrosting unless the outdoor-air temperature is below 4°C and the outdoor relative-humidity exceeds 70%. This shows that the defrosting is limited to the frost-formation range of the heat pump and there is no unnecessary defrosting. Because defrosting is greatly influenced by the relative humidity when the outdoor temperature is below 4°C, we have studied the relationship between the relative humidity and the defrosting interval. The results are shown in Fig. 11. Although the data points are scattered widely, the defrosting interval tends to increase rapidly as the outdoor-air's relative humidity is decreased.

Figure 12 shows the defrosting of the outdoor heat-exchanger pro- duced transients in refrigerant (R-22) circuit temperatures and pressures and compressor power-consumptions. The results of a typical defrosting

3000

~250

0

rr

W o~2o

oo

Q. 8 &.

1ooo

5o0

Suction pressure

X

X

X

Discharge

pressure

--0---0--0-- Compressor

power

B,0

~5

0 (~

15.0

I

I I

Refrigerant

temperature

I I

I I

I I

I I

30

Fig.

12.

12.0

~/3

0

9.0

~-~2

o I n

°.o

3.0

O

I I -10

0.0

60

90

12

0 15

0 15

0 21

0 24

0 27

0 O

PERA

TING

TIM

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UR

ING

DEF

ROST

(s)

Def

rost

cyc

le o

bser

ved

foll

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g a

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. 75

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elat

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hum

idit

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st.

i rb

~b

La~

240 Kamil Kaygusuz

1000 5

800

600 <,

z 400

2OO

0

370

l n s o l a t _ i o n

330 ~ i

6..0 8.0 10.00 12.00 14.00 16.00 18.00 TIME OF DAY (h)

Fig. 13.

4

3

g

~2

1

O

Temperature, COP and insolation variations with time of day.

test, following the -2-0°C ambient-air temperature and 75% relative- humidity frosting tests are presented.

Figure 13 shows the variation of the COP, daily total solar-insolation and temperature of ~ , T4, Tind and T a with time of day.

CONCLUSIONS

The results presented show the effects of frosting and defrosting on the performance of air-to-air heat pumps and provide information on the selection of control parameters for defrost initiation. The conclusions that can be drawn from the results are as follows:

(1) Frosting and defrosting losses were insignificant for all experimental tests conducted with outdoor-air relative-humidities of 65% or less at an ambient-air temperature of 4.0°C.

(2) Seasonal performance analyses show that the frosting reductions in the heating season's COP to be approximately 3% for Trabzon, Turkey.

(3) The use of auxiliary heat, during the defrosting periods, caused significant cumulative reductions in the heat-pump's COP. Defrost power-consumption yielded nominally 15% degradations of COP for laboratory tests conducted at -2.0°C and at relative humidities greater that 65%.

A ir-to-air heat pump under frosting/deJ?osting conditions 241

(4) For non-frosting dehumidifying conditions at an ambient temperature of IO°C, the COP improved 13% and the heating capacity improved 25% by increasing the ou tdoor relative-humidity from 65 to 85%.

(5) In order to optimize the performance of the heat pump when operating under conditions of frost formation, the supplementary heat necessary to maintain a constant heat-output to the building and the number of defrost cycles should be minimized.

R E F E R E N C E S

1. Tassou, S. A., Marquand, C. J. & Wilson, D. R., Energy and economic comparisons of domestic heat pumps and conventional heating systems in the British climate. Appl. Energy, 24 (1986) 127-38.

2. Calm, J. M., The heat pump, ASHRAE J., Aug. (1984) 40-5. 3. York, T., . Heat pumps: Developing the dual-fuel option EPRI J., Dec.

(1990) 22-27. 4. Matsuda, T., Miyamoto, S. & Minoshima, Y., A new air-source heat pump

system, ASHRAE J., Aug. (1978) 324. 5. Goldschmidt, W, V. & Hart, H. G., Heat-pump system performance: Ex-

perimental and theoretical results. ASHRAE Trans., 88 (1982) 479 89. 6. Tassou, S. A., Marquand, C. J. & Wilson, D. R., Comparison of the

performance of capacity controlled and conventional on/off controlled heat- pumps. Appl. Energy, 14 (1983) 241-56.

7. Reay, D. A. & Macmichael, D. B. A., Heat Pumps: Design and Application. Pergamon Press, Oxford, 1979.

8. Merril, P., Heat pumps 'on'-'ofF capacity control and defrost performance tests using demand and time-temperature defrost controls. ASHRAE Trans., 87 (1981) 381-93.

9. Kuwahara, E., Kawamura, T. & Yamazaki, M., Shortening the defrost time on a heat pump air conditioner. ASHRAE Trans., 92 (1986) 20-9.

10. Miller, W. A., Frosting experiments for a heat pump having a one-row spine-fin outdoor coil. A SHRAE Trans., 90 (1984) 1009-25.

11. Imaiida, T., Kojima, S., Aoi, F., Isaka, Y. & Ohta, M., Development of demand defrost control system for residential heat pumps. ASHRAE Trans., 91 (1985) 152740.

12. Bittle, B. B. & Goldschmidt, V. A., Effect of cycling and frost formation on heat-pump performance: Implications of cyclic test data. ASHRAE Trans., 89 (1983) 743-54.

13. Votsis, P. P., Tassou, S. A., Wilson, D. R. & Marquand, C. J., Investigation of the performance of a heat pump under frosting and defrosting conditions. Heat Recovery Systems & CHP, 9 (1989) 399406.

14. Duffle, J. A. & Beckman, W. A., Solar Engineering of Thermal Processes. John Wiley, New York, 1991.

15. Kaygusuz, K., Energy and economic comparisons of air-to-air heat pumps and conventional heating systems for the Turkish climate. Appl. Energy, 45 (1993) 257-67.

performance of an air-to-air heat pump under frosting and defrosting conditions

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