performance of solar-assisted heat-pump systems

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ELSEVIER

Applied Energy 51 (1995) 93-109 © 1995 Elsevier Science Limited

Printed in Great Britain. All rights reserved 0306-2619/95/$9.50

0 3 0 6 - 2 6 1 9 ( 9 4 ) 0 0 0 4 2 - 5

Performance of Solar-Assisted Heat-Pump Systems

Kamil Kaygusuz

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

ABSTRA CT

Simulations have been made with SOLSIM (Howell, J. R. et al., Solar- Thermal Energy Systems. McGraw-Hill, New York, 1982) for solar and air-source heat-pump systems with energy storage in encapsulated phase- change material ( PCM) packings. An experimental study was performed for a laboratory building in Trabzon, Turkey for space heating. The results indicate that the dual-source system is technically the most convenient solar heat-pump configuration, but before installing a solar-assisted heat- pump system, a detailed economic analysis is needed

A C

Ae Cmi. C.a COW COP E

f F r . I

Is,ref L mac maex mw

NOTATION

Solar collector area (m 2) Area of collector absorber (m 2) Thermal capacity (kW/K) Specific heat of air (kJ/(kg K)) Specific heat of water (kJ/(kg K)) Heat-pump coefficient of performance Effectiveness of heat exchanger Fraction of load met by heat-pump system Fraction of annual load met by free energy Heat-removal factor Incident solar isolation (W/m 2) Reference solar isolation (W/m 2) Rate of heat loss from the storage unit (W/m 2 °C) Mass flow rate of air in condenser (kg/h) Mass flow rate of air in heat exchanger (kg/h) Mass flow rate of water in system (kg/h)

93

94 K. Kaygusuz

nehar n¢ol ndis nst qair qa~x qsolar Qeon QI. Ta Tacon 1 Tacon2 Tcxl Tox2 Tf, in Tf, out Tind Tiws Tows (Ts)iso Tst Tw T3 T4 U Wcf Wcomp Wevf W~xf Whp Wpump

(ra)o

Charge efficiency of the storage unit Collector efficiency Discharge efficiency of the storage unit Storage efficiency Energy supplied from ambient air by air-sourced heat pump Energy supplied by auxiliary heating system Energy supplied by insolation Heat extracted by condenser (k J/h) Annual heating load (kWh) Average daily ambient air temperature (K) Inlet air temperature of condenser (K) Outlet air temperature of condenser (K) Inlet water temperature of water-to-air heat exchanger (K) Outlet water temperature of water-to-air heat exchanger (K) Collector fluid (water) inlet temperature (K) Collector fluid (water) outlet temperature (K) Indoor air temperature (K) Inlet water temperature of store (K) Outlet water temperature of store (K) Ideal switch-over (control) temperature (K) Storage temperature (K) Water inlet temperature (K) Inlet refrigerant temperature of air-sourced condenser (K) Outlet refrigerant temperature of air-sourced condenser (K) Collector thermal loss coefficient (W/(m 2 K)) Work input to the compressor fan (k J/h) Work input to the compressor (k J/h) Work input to the evaporator fan (k J/h) Work input to the heat-exchanger fan (k J/h) Energy supplied by series heat-pump system (k J/h) Work input to the water-circulating pump (k J/h)

Effective absorptance of cover-absorber assembly

INTRODUCTION

Solar-energy systems and heat pumps are two promising means of reducing the consumptions of fossil-fuel resources (coal, petroleum etc.), and, hope- fully, the cost of delivered energy for residential heating. An intelligent extension of each is to try to combine the two in order to further reduce the cost of delivered energy. In general, it is widely believed that combined

Solar-assisted heat-pump systems 95

systems will save energy, but what is not often known is the magnitude of the possible energy saving and the value of that saving relative to the additional expense.

Solar heat-pump systems can be classified according to the source of heat that supplies the evaporator of the heat pump as either parallel, series or dual. In parallel systems, the heat pump receives energy from the atmosphere, and the collected solar energy is supplied directly for either space heating or for heating water. In the series system, solar energy is supplied to the evaporator of the heat pump, thereby raising its temperature and increasing the coefficient of performance (COP). In the dual-source configuration, the evaporator is designed so that it can receive energy from either the atmosphere or from the solar-energy store. The performance and economics of these three systems have been analysed theoretically in the literature) -5

In the present study, the performance of a dual-source heat-pump system for residential heating is investigated. The theoretical results, obtained using a modified computer package, SOLSIM, 6 are compared with experimental results which were obtained over the heating season of 1992, 7 (from November to May). Also, the heat-pump COP, storage and collector efficiencies, the number of operating days, percentage of the heating load met by the systems, total energy savings and the total heating loads have been calculated.

THE EXPERIMENTAL SET-UP

The solar-assisted heat-pump system linked to a phase-change energy storage tank described here is at Trabzon. A schematic overview of the heating system is shown in Fig. 1. A detailed description of the experimental set-up was given in a previous study. 7

Ideal switch-over (control) temperature

A series solar heat-pump system may consume the least energy if the water-to-air heat exchanger is used when the water temperature is low. s-l~ The general practice is to have the switch-over from using direct solar heating to using the heat pump, or vice versa, to take place at some fixed entry-water temperature, which is called the switch-over (control) temperature. In previous studies, Manton and Mitchell s used 303 K, McGraw et al. 9 used 302.4 K and Bond ~° proposed various temperatures depending on the building load as the control temperature. For example, Bond proposed 303 K with a 12.5 kW building load. Bessler and Hwang ~1

96 K. Kaygusuz

/So ;

(,, ...... ) ~ 1 ~ l . ' f / E~,po,,,o,/J J / / ~ ~ ] " - - ] ] / "J (watcr ~uurcc) .-xJ

Air ~ ~ l o t i l g c lilllk Al l OUt Condenser

c o l l e c t o r s

Fig. 1. Schematic overview of the system.

used a different control strategy in their simulation study, in which direct solar heating was used when the entry-water temperature exceeded 313.7 K. The heat pump was used when it was below 294.8 K and both direct solar heating and the heat pump were used at intermediate temperatures. But these temperatures were determined with the primary purpose of maintaining the supply air temperature instead of minimizing the rate of energy consumption.

Comparing the system's power consumptions, when the heat exchanger is used for direct solar heating and when the heat pump is used, as shown in Fig. 2(a), it is seen that, when the entry-water temperature is below the desired indoor temperature (295.1 K), the system always consumes less power when using the heat pump than when using the heat exchanger regardless of the building heating load. This is because the water-to- air heat exchanger cannot provide any heat from the water at this low temperature, and all heating of the building has to be provided with the auxiliary heater or with the heat pump. However, when the entry-water temperature is sufficiently high (333 K), it is always more energy conserving to use the heat exchanger. But, in the solar-assisted series heat-pump system, the entry water to the water-sourced evaporator of the heat pump is at a maximum of 313 K at 1200 kg/h water circulating rate through the system under the heating season conditions. Because the melting temperature of the PCM is around 303 K, its (latent) energy storage temperature is nearly constant at 303-310 K. So, at this maximum entry temperature, sometimes it is more energy conserving to use the

Solar -ass i s t ed hea t -pump sy s t ems 97

1o

s

F i g . 2 .

11:1

i 303

/ , ,~.p .

i ]9,3 +

I I I I I0 I0 lO 40

I '~al ing Load ( k W I d l y )

I l I0 20 lO

Healing "Load ( k W l d a y )

(b)

(a) Power consumption; (b) ideal switch-over temperature versus heating load.

water-to-air heat exchanger and sometimes more energy conserving to use the heat pump.

In this study, for most of the days, the entry-water temperature varied between 298 and 308 K and the building load was around 20 kW/day (17 200 kcal/day). Therefore, when the heat pump is used, more energy conservation can be achieved. The ideal control temperature was determined by using the following equation: the calculated temperatures were plotted against the building heating loads in Fig. 2(b). The ideal switch-over temperature is given by: 1°

(Ts)is ° = Tind + Qc - Wcomp (1) ECmin

EXPERIMENTAL RESULTS

The heat pump COP is defined as:

COP = (Qco.)/(Weomp) (2)

The COP of the solar-assisted heat-pump system with storage can be calculated by using following equation: 7

COP = m a c f p a ( T a c ° n 2 - Tac°nl) (3)

Wcomp + Wpump + Wcf

The COP of the solar-assisted parallel heat-pump system with storage is defined as: 7

C O P = macCpa(Tac°n2 - Tae°n0 + maexCpa(Tex2 - Tex') (4)

Wcomp + Wpump + Wet + Wex f + Wev f

98 K. Kaygusuz

The instantaneous collector efficiency is given by: 6

UAe(Tr, i . - Ta)] (5) n¢°l = FR ('r°t)err - A f t

Also, the net collector efficiency is:

mwCpw(rr, i . - Tf, ou,) ncol = (6) Aj

Storage efficiency of phase-change material

The thermal performance of the energy storage can be evaluated by its discharge efficiency, charge efficiency, and overall storage efficiency. The definitions of these efficiencies are as follows:

Actual heat output Discharge efficiency =

Ideal heat output

mwCpw(Tiws - Tows) dt ndi s = (7)

mwCpw(Tst - Tiws) dt

Actual heat received Charge efficiency =

Ideal heat input

mwCpw(T iw s - Tows) dt - L dt ncha, = (8)

mwCpw(Tiws - /'st ) d t

The overall storage efficiency is the product of the discharge efficiency and the charge efficiency:

nst = ndisnchar (9)

METHOD OF THEORETICAL ANALYSIS

The complexity of the thermal analysis of solar-assisted heat-pump systems makes the use of computer simulations the only feasible method for determining the system dynamics and performance. These simulations were performed with the simulation program SOLSIM. 6 This was modified to include the heat-pump system performance behaviour as indicated by the experimental set-up. So, this computer program contains some sub- routines, which model the behaviours of individual pieces of hardware (i.e. collectors, storage tanks, heat pumps, building heating load), and an executive routine which links these component models and solves the resulting system of equations. The simulation calculations are performed


TABLE 1 SOLSIM required inputs.

Specific heat capacity of working fluid (J/(kg °C)) Length of simulation (min) Calculation interval (rain) Collector aperture area (m 2) Value of (a~')erf Value of a Value of b Value of F R Collector mass flow rate (kg/min) Value of Is,ra used in b (W/m 2) Constant heating load (J/h) Load mass flow rate (kg/min) Thermal capacity of storage (J/K) Initial temperature (K) Minimum temperature to load (K) Maximum allowed storage temperature (K) Tank overall loss coefficient (W/m 2 °C) Tank L/D ratio Stratified latent heat storage (number of tank levels) Treated as sinusoidal around ambient average (K) With a positive maximum swing (K) Is a simple sine curve to be used? Maximum insolation (W/m 2) Day length (sunrise-sunset) (min) Time of sunrise (h)

4183 900

15 30 0-80 0 1.0 0.9

21.6 800

1-77 x 107 21.6

2.55 x 108 293 293 312

0.250 2.46 3

281 12 Yes

800 600

7-0

with a 30 min c o m p u t a t i o n a l t ime step to a l low cons ide ra t ion o f the t rans ien t effects and sho r t - t e rm in terac t ions o f co m p o n en t s .

T h e h e a t - p u m p mode l used in these s imula t ions is quasi s teady-s ta te in na ture . The hea t p u m p has two hea t sources fo r the evapora to r : these are the wa te r and air sources. T h e actual p e r f o r m a n c e d a t a o b t a in ed f r o m exper imen ta l results are used to genera te t h i rd -o rde r po lynomia l s re la t ing the heat p u m p ' s COP to the source t empera tu re . F o r the dua l - source heat p u m p , two different sets o f po lynomia l s are used, one set re la t ing to the wa te r source and the o the r fo r the air source.

F o r the wa te r - source hea t pump:

COP = 5.46 + 5.33 x 10-2Tw - 5-53 x 10-4Tw 2 + 1.20 x 10-6Tw 3 (10)

Qcon = 21-42 - 5-62 x 10-2Tw - 7.47 x 10~Tw z + 2.63 x 10-6Tw 3 (11)

F o r the a i r - source hea t pump:

COP= - 2 7 . 8 6 + 0"121Ta + 1"601 X 1 0 ~ T 2 - 7"035 X 10-7T 3 (12)

Qco, = 18.45 - 0 .101Z a + 6-508 x 10-ST f - 5.044 x 10-7Ta 3 (13)

100 K. Kaygusuz

TABLE 2 Construction properties of the laboratory building.

Window area (single glass, U = 4.8 W/m 2 °C) Wall area (single brick, U -- 1.6 W/m 2 °C) Floor area (concrete, U -- 2-5 W/m 2 °C) Ceiling area (concrete + fiat metal, U = 2.0 W/m: °C) Effective UA Comfort temperature Average degree days for two seasons

75 m z 60 m 2 75 m 2 75 m 2

0.800 kWh/°C 22.0°C

2340 Average per season total heating load over two heating seasons (kWh) 20 214 Dimensions of the building 3.5 m × 6.0 m

× 12.0 m

The solar system modelled is a conventional l iquid-medium system. The collector parameters include FR, b, (ra)~tr and Is,r~ t. Insolation was chosen as a repeating, sinusoidally varying function, and, in this case, the time of sunrise, day length and peak insolation were specified. The heating load was specified as constant, given at each time. The storage capacity was specified and the storage was chosen as stratified according to the experimental results. Heat losses from the store are accounted for by specifying the overall thermal loss coefficient. The ambient temperature is modelled as a daily sinusoidal variation around an average ambient temperature. The SOLSIM required inputs are given in Table 1.

The building used in the simulation is the laboratory, whose structural properties are given in Table 2. The building has 75 m z floor area and was not well insulated.

R E S U L T S A N D D I S C U S S I O N

The function, F, is shown in Fig. 3, for Trabzon, as a function of collector area for the conventional heat pump, conventional single-cover solar system, and series, parallel and dual source heat-pump systems. For a building with neither a conventional solar-energy system nor a heat- pump system, F equals zero. For a building (in our simulations, the laboratory building with a 75 m 2 floor area) with only a conventional (air-to-air) heat pump, the fraction of the heating requirement supplied by free energy (non-purchased) is qair divided by the total heating requirement of the building and equals 50%. Since the air-to-air heat pump does not contribute to the heating load, the value of F depends on the C O P and the relative size of the space to be heated and building heating loads during the heating season. In the case of a conventional

Solar-assisted heat-pump systems l 01

A

v

U .

Fig. 3.

100

Dual Source (1-cover)

' °

. / ~ , ~onventional ~ Pump ~

60 / f / . ~ / / "

/ , / "

/ . / ~ . o n v e n t l o n a t - - Solar ( 1.cover ) / ' / . /

I I I I I I 0 I0 20 30 40 50 SO

Col lec tor Area (m 2)

Fraction of annual load met by free energy as a function of collector area.

solar system, F depends on the collector area and the storage mass; as the collector area and storage mass increase, F increases.

The F curves for the conventional solar systems are in agreement with results predicted by the f -char t method ~2 to within a few percent as shown in Fig. 3 for Trabzon over the heating seasons of 1992. The collector size necessary for the solar-energy system to consume less auxiliary energy than the conventional heat-pump system is between 20 and 30 m 2 for a climate like Trabzon.

tO ~,...~. Series System (I-cever)

I v Duet Source ( I -cover) ~'~'~"

,10 Conventional SoLar _~ sad ParaUeL System

I I I I I 0 10 20 30 40 SO

Collector Area (m 2 )

Fig. 4. Collector efficiency of combined heating systems.

102 K. Kaygusuz

'100 +/. .=.

._~ 8 L"

or,

,.i¢

..r+

WHP I

6~

t.O

20

0 P o r o t e t

3 . 0 C O P

Fig. 5.

Purchased ~ A oir "~ Free energy [-"--I Q'sotar J' energy

Series Dual SolQP ~.0 3.S

Heating contributions from all possible sources.

The simulation results show that the seasonal performance of the collector for the solar-assisted parallel heat pump and conventional solar systems of the same collector area are equal. In the case of solar-assisted series and dual-source heat-pump systems, the collector performances of the same collector area are not equal. Clearly, the improved collection efficiency is the direct result of the solar-assisted heat-pump capability, which maintains lower average storage temperatures and hence lower collector temperatures in the series system. Figure 4 shows the collector efficiencies for single-cover solar-assisted series and dual-source systems and for single-cover conventional solar and parallel heat-pump systems over the heating season.

The seasonal energy balance requires that the sum of all the energies supplied equals the heating load of the building, or:

qsolar + qair + Whp q- qaux -- QL (14)

The relative contributions from each of these four heat sources is shown in the bar graphs of Fig. 5 for the parallel, series and dual-source heat- pump systems and for the conventional solar and heat-pump systems. The combined height of the qsolar and qair bars in Fig. 5 represents the percentage of the total heating requirement supplied by free energy and is therefore equal to the F value.

As shown in Fig. 5, adding the solar-energy capability to the conventional solar system to create the series system increases qsolar modestly. The balance of the heating load must be supplied by purchased energy (Whp and q,,.x) because q,ir is equal to zero for both systems. The net increase in F is qsolJqload.


~q t~

o z=

Z

N o

Fig. 6.

400

• S e r l e s system

300 / \ / \

/ ~ / Parallel

/ \ 200 / / ~ - . - ~ \ A ~ " Dual s°uree

/ .Jr"- ~ ~ \ system

273 283 293 303 313 323

STORAGE TANK TEMPERATURE (K)

~ral~l, sefiesanddualsoumestoragetem~ratum distribution.

10o

The heat-pump's seasonal COP varies between the systems. The use of a solar source for the heat pump raises the seasonal COP over that of the conventional and parallel heat-pump systems. As shown in Fig. 5, the seasonal heat-pump heating COPs for the parallel, dual source and series heat-pump systems are 3.0, 3.5 and 4.0 respectively. As expected, the COP for the series and dual source heat pumps are substantially higher since they utilize the stored solar energy.

The series heat-pump COP is higher than that of the dual-source system even though the latter has the apparent advantage of utilizing the more favourable energy source, i.e. either ambient air or energy storage. This is because the solar-assisted series heat-pump system operates only down to a source temperature of 10°C, while the dual-source heat-pump system utilizes colder ambient air as a source when the energy storage tank has reached the solidification temperature of the PCM (<10°C). As a result, the dual-source heat pump supplies more heat to the house than the series heat pump but does so at a lower COP.

Frequency distributions of temperatures and operating conditions during the heating season afford a greater understanding of these overall performance results. Figure 6 is plot of frequency of occurrence of the storage temperature during the heating season of 1992 for the parallel, series and dual-source systems with 30 m 2 single-cover collectors in Trabzon. The average storage temperature in a properly sized solar-heating system will be near the minimum usable temperature in the middle of the winter, when most of the heating load occurs. In the conventional solar and parallel systems, this minimum usable storage temperature is some-

104 K. Kaygusuz

w.

o

~s

0-1

o

Z 0

,< t~

o

0

o Z

O"

120

100

80

60

40

20

0 273 278 283 288 293 2 9 8 3 0 3

TEMPERATURE OF SOURCE ( K )

Fig. 7. Heat-pumpfrequency of operation distribution.

what above room temperature. However, in the series and the dual-source systems, the average storage temperature is the minimum usable heat- pump source temperature (~7°C). During the winter (January-March), the storage temperature very rarely rises enough to permit direct heating by the solar system. As a result, the heat pump must operate virtually whenever winter-time space heating is required.

Figure 7 shows the frequency of operation of the heat pump as a function of source temperature in the series, parallel and dual-source systems. The heat pump operates most often in the parallel system, because the heat pump then uses ambient air as a heat source. When used as the auxiliary to a solar system, i.e. in the parallel system, the hours of operation are reduced but there is also a tendency towards a lower average source temperature which accounts for the lower COP.

The importance of the minimum allowable tank temperature, Tmin, was also evaluated. The parallel, series and dual-source heat-pump systems were simulated for Trabzon with the 30 m 2 single-cover collectors and a minimum tank temperature of 7°C. The resultant distribution of the parallel and series system source temperature utilization is shown in Fig. 8. For sunny days, the storage source is used more often for the


Fig. 8.

i00

Parallel system with air source

" 7 BO

~ Series system / ~ . / \yith s t o r a g e tank

60 • = PC)

~ 4 0

~ 20

~ I I I I

-5 0 5 to ]5 20 25 30

TEMPERATURE OF SOURCE (C)

Effect of minimum source temperature on parallel and series heat-pump operation.

series system. In contrast, the air source is used more often for cloudy, hot days (Ta > Tmi,) for a parallel system via the conventional air-to-air heat pump.

Figure 9 also shows the comparison of the measured collector efficiency and the efficiency as predicted by SOLSIM. The measured collector efficiency is somewhat below the predicted values early in the solar day,

Fig. 9.

-6 t.J

1.0 1.0

0 . . / \ "J \ ~. C~cul l ted Measured / / ur or , . . . a . + ~ . ~ Meal; ed Collect E/ | i . ~ - ~ l n c a l ~ t i a n , - ~ - ~ " ;, \ . . . . . . . . . . . .

. ~ , / / / \ o o "~. /7 o / j ° < o ' k

~ ¢aleuLated Collecter (ffk; iency 0 ' ~ It4 ,~ 0.4

o , o ,

\ \

0 - 0 f / I I I I I I I I I ~l'~e ~ II } 10 11 12 O 14 IS II IT

Time of Day (In)

Comparison of the measured collector efficiency and the efficiency as predicted by SOLSIM.

106 K. Kaygusuz

&S

a.

0 4 .0

u

l

Fig. 10.

1 march I$$2

• measwed o I~redicted by SOI.$1M

I I I I I I I

lime Qt Day (h)

Comparison of the measured COP and the COP as predicted by SOLSlM.

but the match is good over the period of high insolation. The daily average collector efficiency measured for the system was 0-47, and the predicted value from SOLSIM was 0.42 for this particular day.

Figure 10 shows a comparison of the measured COP and the COP as predicted by SOLSIM for the solar-assisted series heat-pump system with storage. The predicted COP is somewhat below the measured values. The daily average COP measured for the system was 4.2, and the predicted value from SOLSIM was 4-0 over this particular day.

Figure 11 shows the measured variations of COP, daily total solar insolation and temperatures T3, T4, Ti,d and Ta with time of day for the parallel heat-pump system.

,°oo I 1,,o , t . ,qolnt l u . T

86001- _. . - - ' /~- ~ ~ -" - I ~o ~

,'.o , o.o ,,'oo doo TIME OF DAY ( h i

Fig. ! L Temperature, COP and insolation variations with time of day.


3 0 8

306

304

3o2

Tupper

3OO

298 I I I ~ I I

10 12 14 16 18 20 22

TIME OF DAY

Temperature variation of C a C I 2 . 6 H 2 0 with time of day in storage tank. Fig. 12.

Figure 12 shows the variation of temperature of the CaC12 . 6H20. As shown in the figure, there is a three temperature-layer stratification in the storage tank. So this was considered in the SOLSIM simulation program.

CONCLUSIONS

The author analysed the performance of solar-assisted heat pumps with latent-heat energy storage at the Karadeniz Technical University. In the experimental study, over the heating season, the mean value of COP, the number of operating days per month, the percentage of heat load met by the systems, the average-collector and storage efficiencies of the parallel and series heat-pump systems have been deduced from experimental data and tabulated in Table 3. The following conclusions were obtained:

1. The dual-source system saved a net energy of 12 056 kWh while the parallel system saved 10120 kWh and the series system saved 9390 kWh net energy all per heating season. Also, the dual-source system takes advantage of the best features of the series and parallel systems.

2. The solar system is not suitable on its own for heating the laboratory building for given technical and climatic conditions, because the region has many cloudy days.

3. For the most energy conserving operation of a series solar heat-pump system, it is necessary to have a control temperature, which varies with the building's heating-load. Therefore a value of 303 K was chosen (with 20 kWh/day building load) for the control temperature.

108 K. Kaygusuz

TABLE 3 The experimental performance of the heat-pump systems over the heating season of 1992.

Months Q~ T a Series system Parallel system (kWh) (°C)

N b COP n~ot n~, f N b COP n~o t n~, f

November 1991 1972 12.6 20 4.60 0.62 0.58 0.66 30 3-00 0.52 0-57 0-98 December1991 3172 8.3 11 4.53 0.60 0-62 0.40 31 3.02 0.53 0-56 0.84 January 1992 4336 4.1 15 4.50 0.57 0.61 0.50 27 2.80 0.54 0.55 0.51 February 1992 4383 3.7 18 4.45 0.63 0.63 0-64 20 2.79 0.49 0.53 0.37 March 1992 3027 8.9 25 4.53 0.64 0.58 0.82 31 3-04 0.50 0.53 0-86 April1992 2216 11-6 23 4.51 0.56 0-60 0.78 30 3.16 0.48 0,55 0-98 May 1992 1666 13.6 25 4.70 0.60 0.57 0.82 31 3.21 0-51 0,54 0-99

a Average heating load (kWh). b Number of working days of the heat-pump system per month. Note: the stated values of COP, n~oj, ns t and fare averages taken over a month.

4. F rom the experimental and theoretical investigations, it was con- cluded that heat storage is an important component in moderate climatic conditions such as those encountered in Trabzon and for this purpose, C a C I 2 . 6 H 2 0 can be used as the PCM in the energy storage tank. The PCM stores energy as latent heat at a nearly constant transition temperature (during charge and discharge), so it is preferable as an energy storage material to water or rock storage for the solar-assisted heat-pump systems in the region,

R E F E R E N C E S

1. Freeman, T. L., Mitchell, J. W. & Audit, T. E., Performance of combined solar heat-pump systems. Solar Energy, 22 (1979) 125-35.

2. Karman, V. D., Freeman, T. L. & Mitchell, J. W., Simulation study of solar heat pump systems. Proc. ISES, Canada 1976.

3. Anderson, J. V., Mitchell, J. W. & Beckman, W. A., A design method for parallel solar heat pump systems. Solar Energy, 25 (1980) 155-63.

4. Mitchell, J. W., Freeman, T. L. & Beckman, W. A., Heat pumps do they make economic and performance sense with solar? Solar Age, 3 (1978) 24-8.

5. Chandrashekar, M., Le, N. T., Sullivan, H. F. & Hollands, K. G. T., A comparative study of solar assisted heat-pump systems for Canadian locations. Solar Energy, 28 (1982) 217-26.

6. Howell, J. R., Bannerot, R. B. & Vliet, G. C., Solar-Thermal Energy Systems. McGraw-Hill, New York, 1982.

7. Kaygusuz, K., Gt~ltekin, N. & Ayhan, T., Solar assisted heat-pump and energy storage for domestic heating in Turkey. Energy Conversion Managmt, 34 (1993) 335~16.


8. Manton, B. E. & Mitchell, J. W., Performance of solar heat-pumps and solar heat-pump systems. ASME J. Solar Energy Engng (1982).

9. McGraw, B. A., Bedinger, A. F. & Reid, R. L., Experimental evaluation of a series solar assisted heat-pump system. Proc. 1981 American Section, 1SES Annual Meeting, 1981, pp. 562-6.

10. Bond, T. Y., Toward an efficient operation of a series solar heat-pump system. ASHRAE Trans., 90 (1984) 617-27.

11. Bessler, W. F. & Hwang, B. C., Performance of solar assisted heat pump heating systems for residential use. Paper presented at ASME/AIChEng, 18th National Heat Transfer Conference, San Diego, 1979.

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

performance of solar-assisted heat-pump systems

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