- Email: [email protected]
Experimental and theoretical investigation of combined solar heat pump system for residential heating
Energy Conversion & Management 40 (1999) 1377±1396
Experimental and theoretical investigation of combined solar heat pump system for residential heating K. Kaygusuz a,*, T. Ayhan b b
a Chemistry Department, Karadeniz Technical University, 61080, Trabzon, Turkey Mechanical Engineering Department, Karadeniz Technical University, 61080, Trabzon, Turkey
Received 1 June 1998; accepted 28 December 1998
Abstract In order to investigate the performance of the combined solar-heat pump system with energy storage in encapsulated phase change material (PCM) packings for residential heating in Trabzon, Turkey, an experimental set-up was constructed. Also, a comparative study of the performance of this system has been undertaken theoretically. Simulations have been made with BASIC computer program of three basic combined con®gurations, as well as conventional solar and conventional heat pump systems in Trabzon by using meteorological data. The experimental results were obtained from December to May during the heating season for four heating systems. These systems are a conventional solar system, a series heat pump system, a parallel heat pump system and a dual source heat pump system. The experimentally obtained results are used to calculate the collector eciency, heat pump coecient of performance (COP), seasonal heating performance, the fraction of annual load met by free energy, storage and collector eciencies and total energy consumption of the systems during the heating season. # 1999 Elsevier Science Ltd. All rights reserved. Keywords: Heat pump; Residential heating; Solar heating system; Energy storage
1. Introduction Solar energy systems and heat pumps are two promising means of reducing the consumption of fossil energy resources (coal, petroleum, etc.) and, hopefully, the cost of delivered energy for * Corresponding author. E-mail address: [email protected] (K. Kaygusuz) 0196-8904/99/$ - see front matter # 1999 Elsevier Science Ltd. All rights reserved. PII: S 0 1 9 6 - 8 9 0 4 ( 9 9 ) 0 0 0 2 6 - 6
1378
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
Nomenclature Ac Ae Cpa Cpw I Ucol (ta )e FR Zcol Zdis Zchar Zst mw mac Qcon PCM COP Wcomp Wpump Wcf Tw Tf,in Tf,out Tiwex Towex Tiws Tows Tref1 Tref2 Tacon1 Tacon2 Trcon1 Trcon2 Trev1 Trev2 Tind Ta Tb A U DD
solar collector area (m2) area of collector absorber (m2) speci®c heat of air (kJ/kg K) speci®c heat of water (kJ/kg K) incident solar radiation (W/m2) collector thermal loss coecient (W/m2 K) eective absorptance of cover-absorber assembly heat removal factor collector eciency discharge eciency of storage tank charge eciency of storage tank overall storage eciency mass ¯ow rate of water in system (kg/h) mass ¯ow rate of air in condenser (kg/h) heat extracted by condenser (kJ/h) phase change material heat pump coecient of performance work input to compressor (kJ/h) work input to water circulating pump (kJ/h) work input to compressor fan (kJ/h) water inlet temperature (K) collector water inlet temperature (K) collector water outlet temperature (K) inlet water temperature of water-to-air heat exchanger (K) outlet water temperature of water-to-air heat exchanger (K) inlet water temperature of energy storage tank (K) outlet water temperature of energy storage tank (K) inlet refrigerant temperature of water source evaporator (K) outlet refrigerant temperature of water source evaporator (K) inlet air temperature of condenser (K) outlet air temperature of condenser (K) inlet refrigerant temperature of condenser (K) outlet refrigerant temperature of condenser (K) inlet refrigerant temperature of air source evaporator (K) outlet refrigerant temperature of air source evaporator (K) indoor air temperature (K) ambient air temperature (K) base temperature (K) ¯oor area of building (m2) overall thermal loss coecient [kW/(m2 K day)] number of degree-days per month
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
n h F QL qsolar qair Whp E Cmin (TS)ict
1379
number of days in month heat transfer coecient fraction of annual load met by free energy (%) annual heating load of building (kW) energy supplied by solar system energy supplied by air source heat pump from ambient air energy supplied by series heat pump system eectiveness of heat exchanger thermal capacity of heat exchanger (kJ/h K) ideal control (switch-over) temperature (K)
residential heating. An intelligent extension is to try to combine the two to further reduce the cost of delivered energy. In general, it is widely believed that combined systems will save energy, but what is not often known is the magnitude of the possible energy saving and the value of those savings relative to the additional expense [1]. Solar heat pump systems are classi®ed 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 hot water. In the series system, solar energy is supplied to the evaporator of the heat pump, thereby raising its temperature and increasing the coecient of performance (COP). In the dual source con®guration, 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 [2±6]. The objective of this paper is to analyse several viable types of combined solar heat pump systems and to compare them to conventional solar and conventional heat pump (air-to-air) systems experimentally and theoretically. Our aim is to determine the thermal performance of the dierent solar assisted heat pump system con®gurations and to provide some insight into many of the non-intuitive reasons behind the performance results. We used an energy storage tank with three combined heat pump systems for storing solar energy in the experimental study. This energy storage can reduce the time or rate mismatch between energy supply and energy demand, thereby playing a vital role in energy conservation. In processes which are wasteful of energy, energy storage will result in a saving of premium fuels. Energy storage can also improve system performance and increase reliability. On the other hand, conventional vapour compression heating only heat pumps are not at present an economic alternative to gas-®red heating systems due to their high capital cost and the degradation in their steady state eciency caused by on/o cycling supplementary electric resistance heating and frosting and defrosting losses [7±10]. The combination of a heat pump and a solar energy system would appear to reduce many of the disadvantages that each has when operating separately or alone. During winter conditions, the energy that could be collected by the solar collectors, but which would be too low in temperature to be useful for direct heating, may be used as a heat source for the heat pump. Since the solar collection±storage system can supply energy at temperatures higher than the ambient air temperature, the capacity and heat pump COP would increase over that if the heat
1380
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
pump works alone, the peak auxiliary load requirement would be reduced and the combined heating system (solar and heat pump) will be operated more economically. On the other hand, the operation of the solar system at temperatures near and below the inside temperature of the building will decrease the collector losses and allow more energy to be collected. The lower collection temperature may then allow the use of collectors with one or no cover, which would reduce the ®rst cost from that of a conventional two-cover solar system. Finally, for those areas where warm temperatures occur during cloudy periods, the combined system may compensate for the reduced performance of the conventional solar system under cloudy conditions and the low capacity of the heat pump in cold winter weather conditions. In the present work, an experimental set-up was constructed to determine the performance of the series, parallel and dual source heat pump, solar collectors and energy storage tank ®lled by PCM used for residential heating. The eects of various system parameters on the response of the indoor air temperature of the building, the temperature variation of the PCM in the energy storage tank and the temperatures of the heat transfer ¯uid (water) in the solar collectors and energy storage tank for series, parallel and dual source heat pump systems were investigated. Also, the collector and storage eciencies, heat pump COP, seasonal performance factor (SPF), the building heating load, heat supplied fraction of the heating load by the systems and energy consumption of the combined solar heat pump system during the heating season (December to May) in 1992 were calculated. On the other hand, the theoretical analyses have been made using a BASIC computer simulation program containing a solar assisted heat pump system with latent heat energy storage tank. The original strategy, namely an hour-byhour simulation of the dynamic behaviour of the system, involved driving it by hourly meteorological data. 2. Description of the experimental set-up The combined solar heat pump system linked to a latent heat energy storage tank presented here is at Trabzon, Turkey, latitude 418N, longitude 408E, placed on the Black Sea side. Table 1 gives the climatic data of the heating season for Trabzon. A schematic overview of the system is given in Fig. 1. The solar collectors used in this system were constructed by modifying ¯at-plate waterTable 1 Climatic conditions of Trabzon over heating season in 1992 Month
Nov.
Dec.
Jan.
Feb.
Mar.
Apr.
May
Average outdoor temperature (8C) Minimum outdoor temperature (8C) Maximum outdoor temperature (8C) Average relative humidity (%) Average wind velocity (m/s) Average solar radiation (MJ/m2 day) Degree days
12.6 6.8 19.0 76.4 2.4 5.78 125
8.3 2.1 20.1 68.6 2.4 4.38 283
4.1 ÿ1.3 11.4 66.7 2.3 5.43 297
3.7 12.0 18.4 67.2 2.1 7.17 305
8.9 1.0 26.2 65.2 2.1 11.74 313
11.6 6.0 25.0 71.9 2.3 14.32 246
13.9 7.6 22.1 81.4 1.9 14.46 158
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
1381
Fig. 1. Schematic overview of the heating system.
cooled collectors. Each collector was made of pipe and ®n type; the absorber unit consists of nine copper tubes (1.8 m length and 0.022 m outside diameter) ®tted longitudinally at 0.1 m pitch across an aluminium sheet (1.8 m length and 0.9 m width). The aluminium sheet has 0.55 mm thickness and 1.62 m2 eective absorber area. The plate is painted with blackboard paint (a=0.90 and e=0.80) and placed inside a metal box made of 0.8 mm thick aluminium sheet and insulated at the bottom and all sides with glass wool of thickness 6 cm [k=0.038 W/ (m 8C)]. The top of the metal box is glazed with a 3.5 mm thick glass cover having (ta )e=0.80. The loss coecient of the collector is 8.20 (W/m2 8C), and the constant heat removal factor FR=0.9 for each collector. The 18 collectors were connected in a parallel series combination. The mass ¯ow rate of the heat transfer ¯uid through the collectors was 43 kg/h m2. The solar collectors were installed at an angle of 488 from the horizontal and faced due south. The compressor used for the combined solar heat pump system was a hermetic type. It was driven by a 1490 watt electrical motor. The heat pump can use ambient air and the solar energy storage tank as a heat source. So, it has two evaporators, air-cooled and water-cooled. The heat pump condenser is an aircooled ®nned tube type heat exchanger. Recently, the use of phase change materials (PCMs) for thermal energy storage in solarassisted heating systems has received considerable attention. The motivation for using phase change energy storage (PCES) is the reduction in storage volume which can be achieved compared to sensible heat storage. Therefore, we used commercial calcium chloride hexahydrate (CaCl2.6H2O) as a PCM for solar energy storage because it is cheaper (10$/100 kg), and it has a better thermal stability than the other PCMs [11±15]. It is commercial grade material produced by the Dow Chemical Company. The melting temperature (288C) of this material is almost an optimum temperature for residential heating applications in some regions
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
1382
of Turkey. Also, the large amounts of heat that can be stored in the phase change of CaCl2.6H2O can reduce the storage volume to 1/8 to 1/15 of that required for rock or gravel storage and to 1/5 to 1/10 of that required for water storage [16±20]. The storage tank was made of sheet iron, and has a diameter of 1.3 m and a length of 3.2 m. It consists of a vessel packed in the horizontal direction with cylindrical tubes. The energy storage material is inside the tubes (tubes are made of PVC plastic), and the heat transfer ¯uid (HTF) ¯ows parallel to it. The storage tank contains cylindrical PVC containers ®lled with PCM. The heat storage tank was linked to the heat pump by means of an evaporator for use as a heat source. The heat storage tank was insulated entirely with 60 mm thick glass wool. The heat loss from the tank to its surroundings is 0.25 (W/m2 8C). In the experiments, the inlet and outlet pressures of the compressor and evaporators were measured with manometers. Also, the inlet and outlet temperatures of the refrigerant (R-22) at the condenser, compressor and evaporators were measured with iron-constantan thermocouples. In addition, the inlet and outlet temperatures of the HTF (water) at the storage tank, collectors, water source evaporator and water-to-air heat exchanger were measured with the same kind of thermocouples. Also, the temperatures of the PCM in the energy store were measured. In addition, the ambient and indoor air temperatures were measured with normal thermometers. A Kipp±Zonen solarimeter, mounted on the vertical plane of the solar collectors, was used to measure the global solar radiation on the tilted surface, and some results are given in Table 2. The ¯ow rate of the circulated water through the system was measured by means of two ¯ow meters. The power consumption of the systems was measured with a wattmeter. A laboratory building of 75 m2 ¯oor area was used as the space to be heated. The structural properties of the building are given in Table 3.
Table 2 Hourly average global solar radiation on a horizontal surface (W/m2) Hour
Jan.
Feb.
Mar.
Apr.
May
June
July
Aug.
Sept.
Oct.
Nov.
Dec.
4±5 5±6 6±7 7±8 8±9 9±10 10±11 11±12 12±13 13±14 14±15 15±16 16±17 17±18 Total
0.0 0.0 0.0 94 223 384 433 419 356 230 98 0.0 0.0 0.0 2442
0.0 0.0 66 244 419 523 615 631 614 530 370 174 0.0 0.0 4187
0.0 0.0 87 230 475 656 733 742 719 607 447 166 73 0.0 4936
0.0 133 321 474 565 663 740 747 733 621 530 370 174 27 6097
63 210 381 545 678 768 817 827 795 719 614 433 258 70 7178
81 253 477 644 777 860 902 916 853 777 637 519 344 120 8160
75 217 392 588 692 779 827 837 805 728 624 443 267 77 7351
82 193 335 481 617 691 726 719 677 610 481 368 187 68 6235
0.0 140 314 453 586 670 698 684 642 537 384 209 56 0.0 5372
0.0 31 167 300 454 544 593 572 523 412 244 77 0.0 0.0 3907
0.0 0.0 70 216 342 426 475 461 398 279 140 27 0.0 0.0 2836
0.0 0.0 0.0 77 216 356 391 405 349 244 147 49 0.0 0.0 2236
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
1383
Table 3 Construction properties of the laboratory building Properties
Value
Window area (single glass, U=4.8 W/m2 8C day) Wall area (single brick, U=1.6 W/m2 8C day) Floor area (concrete, U=2.5 W/m2 8C day) Ceiling area (concrete+¯at metal, U=2.0 W/m2 8C day) Eective UA [kWh/(8C day)] Comfort temperature (8C) Average degree-days in the heating season Average total heating load in the heating season (kW) Dimensions of the building (3.5 m 60 m 12.0 m)
75 m2 60 m2 75 m2 75 m2 0.800 20.0 1727 20772 242 m3
3. System description and operation modes We will de®ne a base solar heating system and a base heat pump heating system as standards of comparison for the combined solar heat pump systems to follow. The base solar system is the standard liquid heating system shown in Fig. 2. It consists of conventional ¯atplate water-cooled solar collectors, an energy storage tank, a liquid-to-air heat exchanger coil in the laboratory building supply duct, a water circulating pump, auxiliary space heaters (electrical resistance heaters) and other control equipment. We have designed our experimental set-up to form dual sources for the heat pump. Because of this, the system can be used in three forms, series, parallel and dual source systems. This process can be accomplished by changing the heat sources for the heat pump. More detailed description of the systems is given below. 3.1. Solar assisted series heat pump system with storage As shown in Fig. 3, this series system consists of conventional ¯at-plate water-cooled solar collectors, an energy storage tank ®lled by PCM as a heat storage material, a heat pump with water-to-refrigerant heat exchanger, an air-cooled condenser, a water circulating pump, a liquid-to-air heat exchanger for direct solar heating, auxiliary heater and other conventional equipment. In this system, the hot water which comes from the collectors ®rst goes to the
Fig. 2. Base solar system with storage.
1384
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
Fig. 3. Solar assisted series heat pump system with storage.
energy storage tank where it releases some energy to the PCM in the storage, and after this, it is used as a heat source by the water source evaporator. After that, it is sent to the solar collectors by the water circulating pump. However, at night and on cloudy days, the less-hot water that comes from the evaporator of the heat pump was sent to the energy storage tank instead of the solar collectors. The cold water extracts energy from the PCM in the storage, and it ¯ows to the evaporator for use as a heat source. 3.2. Solar assisted parallel heat pump system with storage As shown in Fig. 4, this parallel system consists of conventional ¯at-plate water-cooled solar collectors, water-to-air heat exchanger, a heat pump with air-to-refrigerant heat exchanger, an air-cooled condenser, a latent heat energy storage tank, a water circulating pump, an auxiliary electrical heater and other control equipment. The solar-assisted parallel heat pump system is combining two main components, a conventional solar heating system and a conventional airto-air heat pump system. In this system, the heat pump uses ambient air as an energy source, while the water-to-air heat exchanger uses solar energy as a heat source, and they each individually give their energies to the heating load of the building. Solar energy is used to meet as much of the heating requirement as possible. Thus, the total available energy of the system
Fig. 4. Solar assisted parallel heat pump system with storage.
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
1385
is the sum of the extracted energies from two dierent systems (solar system and heat pump system). 3.3. Dual source system The dual source system contains two systems: the series system and the parallel system described above. In the dual source system, the heat pump has two evaporators: one is the water source evaporator which is placed in the storage tank loop and the other is the air source evaporator which is placed outside of the room. This allows the heat pump to use either the collected solar energy or the ambient air as a heat source, depending on which results in a higher COP. There are three heating modes for the dual source system. In the ®rst case, the dual source system is operated in the direct solar heating mode. Here, the storage tank is hotter than the predetermined control value (below 303 K) and able to heat the laboratory building directly. The heat pump is o, and any excess solar energy is being stored in the tank. In the second case, when the energy storage tank temperature is below the control value, but higher than the minimum using value (280 K) and ambient air temperature, the hot water in the storage tank is pumped to the evaporator and used as a heat source for the heat pump. Depending on the magnitude of the building heating load, the excess solar energy is collected in the energy storage tank. In the third case, when the storage temperature is either at its minimum temperature or less than the ambient air temperature, the ambient air is used as a heat source for the evaporator, and auxiliary heat is supplied when needed.
Fig. 5. Power consumption versus heating load for the heat exchanger and heat pump for entering water temperature at 295 K.
1386
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
3.4. Ideal control (switch-over) 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 [21±24]. The general practice is to have the control (switch-over) change from using direct solar heating to using the heat pump, or vice versa, take place at some ®xed entering water temperature which is called the switch-over temperature. In previous studies, Manton and Mitchell [21] used 303 K, McGraw et al. [22] used 302.4 K and Bond [23] proposed various temperatures depending on the building load at the switch-over temperature. For example, Bond proposed 303 K with a 12.5 kW building heating load. Bessoler and Hwang [24] used a dierent control strategy in their simulation study in which direct solar heating was used when the entering water temperature was above 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. However, these temperatures were determined with the primary purpose of maintaining the supply air temperature instead of minimizing energy consumption. Comparing the system's power consumption when the heat exchanger is used for direct solar heating and when the heat pump is used, as shown in Fig. 5, it is seen that when the entering water temperature is below the desired indoor air temperature (293 K), the system always consumes less power when using the heat pump than when using the heat exchanger, regardless of the building heating load, because the water-to-air heat exchanger cannot provide any heat by using water at this low temperature, and all heating of the building has to be provided with
Fig. 6. Ideal switch-over (control) temperature versus building heating load.
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
1387
the auxiliary heater or by using the heat pump. However, when the entering water temperature is suciently high (333 K), it is always more energy conserving to use the heat exchanger. In the case of the solar assisted series heat pump system, the entering water to the water source evaporator is a maximum temperature of 313 K at a 1200 kg/h water circulating rate through the system under the heating season conditions. The melting temperature of the CaCl2.6H2O is around 303 K, and its latent heat energy storage temperature is nearly constant at temperatures between 303±310 K. So, at this maximum entering temperature, sometimes it is more energy conserving to use the water-to-air heat exchanger and sometimes more energy conserving to use the heat pump. In this study, for most days, the entering water temperature varies between 298 to 308 K, and the building load is around 20 kW/day. So, if we use the heat pump as much as possible, more energy conservation can occur for residential heating. In the present study, the ideal control (switch-over) temperature was determined by using the following equation, and then the calculated temperature values were plotted against building heating load as shown in Fig. 6. The ideal control temperature is given by [23] as TS ict Tind
QL ÿ Wcomp : ECmin
1
4. Calculation of the experimental results The heat pump COP is de®ned as COP Qcon =Wcomp :
2
The COP of the solar assisted series heat pump with storage can be calculated as [25] COP
mac Cpa Tacon2 ÿ Tacon1 : Wcomp Wpump Wcf
3
The COP of the solar assisted parallel heat pump system with storage can be calculated as COP
mac Cpa Tacon2 ÿ Tacon1 maex Cpa Tex2 ÿ Tex1 : Wcomp Wpump Wcf Wexf Wevf
The instantaneous collector eciency is given by [1] Ucol Ae Tfi ÿ Ta : Zcol FR taeff ÿ Ac I
4
5
Also, the net collector eciency can be calculated as Zcol
mw Cpw Tf,in ÿ Tf,out : Ac I
6
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
1388
The thermal performance of the energy storage can be evaluated by its discharge eciency, charge eciency and overall storage eciency. The de®nitions of these eciencies are [26] Discharge efficiency
Zdis
Actual heat output Ideal heat output
mw Cpw Tiws ÿ Tows dt mw Cpw Tst ÿ Tiws dt
Charge efficiency
Zchar
7
Actual heat received Ideal heat input
mw Cpw Tiws ÿ Tows dt ÿ L dt mw Cpw Tiws ÿ Tst dt
8
The overall storage eciency is the product of the discharge and charge eciencies: Zst Zdis Zchar :
9
The monthly heating load can be calculated by using the following equation [27] QL UAh :DD: The monthly average degree-days (DD) can be approximated by ln cosh 1:698h 3=2 h DD sm n 0:2042 2 3:396
10
11
where n is the number of days in the month and h is de®ned as h
Tb ÿ Ta : sm :Vn
12
The standard deviation of the monthly average ambient temperature is sm which is generally not available from weather summaries. It can be approximated by the following with essentially no loss in accuracy: sm 1:45 ÿ 0:0290Ta 0:0664syr
13
where syr is the standard deviation of the monthly average ambient temperature from the annual average ambient temperature. So, the annual heating loads of the laboratory building were calculated by using the above equations, and the derived results are given in Table 3. 5. Computer simulation of the system The complexity of the thermal analyses of solar assisted heat pump systems makes use of computer simulations the only method for suitably determining the system dynamics and
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
1389
performance. These simulations were performed with the BASIC computer simulation program. This computer program contains some subroutines which model individual pieces of hardware (e.g. collectors, energy store, air and water source heat pump, building heating load), and an executive routine which links these component models and solves the resulting system of equations. The heat pump model used in these simulations is quasi-steady state in nature. The heat pump has two heat sources for the evaporator, i.e. air and water. The actual performance data obtained from the experimental results are used to generate third-order polynomials relating heat pump COP to the source temperature. For the dual source heat pump, two dierent sets of polynomials are used, one set relating to the water source and the other for the air source. For the water source heat pump: COPH 5:46 5:33 10ÿ2 Tw ÿ 5:53 10ÿ4 Tw 2 1:20 10ÿ6 Tw 3
14
Qcon 21:42 ÿ 5:62 10ÿ2 Tw ÿ 7:47 10ÿ4 Tw 2 2:63 10ÿ6 Tw 3 :
15
For the air source heat pump: COPH ÿ27:86 0:121 Ta 1:601 10ÿ4 Ta 2 ÿ 7:035 10ÿ7 Ta 3
16
Table 4 BASIC computer simulation required inputs Data Value of a Value of b Value of FR Value of (ta )e Length of simulation (min) Calculation interval (min) Collector aperture area (m2) Collector mass ¯ow rate (kg/min) Value of Is,ref used in b (W/m2) Constant heating load (J/h) Load mass ¯ow rate (kg/min) Thermal capacity of storage (J/K) Initial temperature (K) Minimum temperature to heating load (K) Maximum allowed storage temperature (K) Storage tank overall loss coecient [W/(m2 8C)] Energy storage tank length/diameter ratio Treated as sinusoidal around ambient average (K) Day length (sun rise±sun set) (min) Maximum solar radiation (W/m2) Time of sunrise (h) Speci®c heat capacity of working ¯uid [J/(kg 8C)]
Value 0.0 1.0 0.9 0.80 900 15 30 21.6 800 1.45 105 21.6 2.53 108 293 293 312 0.250 2.46 281 600 900 07:00 4183
1390
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
Qcon 18:45 ÿ 0:101 Ta 6:508 10ÿ5 Ta 2 5:044 10ÿ7 Ta 3 :
17
The solar system modelled is a conventional liquid medium system. The collector parameters include FR, b, (ta )e and Is,ref. The solar radiation was chosen as a repeating sinusoidally varying function, and in this case, the time of sunrise, day length and peak radiation were speci®ed. The heating load was speci®ed as constant, given at each time. The storage capacity is speci®ed, and the storage was chosen as being strati®ed according to the experimental results. Heat losses from storage are taken into account by specifying the overall thermal loss coecient. The ambient temperature is modelled as a daily sinusoidal variation around an average ambient temperature. The BASIC computer simulation required inputs are given in Table 4. The building used in the computer simulation is the laboratory, and its structure properties are given in Table 3. The building has 75 m2 ¯oor area and is not well insulated. 6. Results and discussion Fig. 5 shows the power consumption versus heating load for the heat exchanger and heat pump at the entering water temperature of 295 K. The ®gure also shows that the power consumption increases with building heating load.
Fig. 7. Fraction of annual load meet by free energy as a function of collector area.
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
1391
Fig. 6 shows the ideal switch-over (control) temperature versus building heating load; it also shows that the ideal switch-over temperature increases with building load. Its minimum value is equal to the indoor temperature, and its maximum value is determined by the characteristic of the heat exchanger, as given by its ECmin value. The function F is shown in Fig. 7 for Trabzon as a function of collector area for the conventional air source heat pump, conventional single-cover solar system, series heat pump, parallel heat pump and dual source heat pump systems. For a building with neither a conventional solar heating system nor a heat pump system, the F value equals zero. For a building (in our simulations the laboratory building with 75 m2 ¯oor area) with only a conventional heat pump (air-to-air), 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 COP and the relative size of the space to be heated and building heating load during the heating season. In the case of the conventional solar system, F depends on collector area and storage mass. As collector area and storage mass increase, F increases. The F curves for the conventional solar system are in agreement with results predicted by the f-Chart method [28] to within a few percent as shown in Fig. 7 for Trabzon during the heating season. 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 m2 for a climate like Trabzon. The simulation results show that the seasonal collector performance for the solar assisted parallel heat pump and conventional solar systems of the same collector area are equal. In the case of the solar assisted series and dual source heat pump systems, the collector performance of the same collector area are not equal. Clearly, the improved collection eciency is a direct
Fig. 8. Collector eciencies of combined heating system.
1392
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
result of the solar assisted heat pump capability, which maintains lower average storage temperatures and, hence, lower collector temperature in the series system. Fig. 8 shows the collector eciencies for 1-cover solar assisted series and dual source heat pump systems and for 1-cover conventional solar and parallel heat pump systems for the entire heating season. The seasonal energy balance requires that the sum of all energy supplied equal the heating load of the building, or qsolar qair Whp qaux ÿ QL :
18
The relative contributions from each of these four heat sources is shown in the bar graphs of Fig. 9. The combined height of the qsolar and qair bars in Fig. 9 represent the percentage of the total heating requirement supplied by free energy and is, therefore, equal to the F value. As shown in Fig. 9, adding the solar energy capability to the conventional solar system to create the series system increases qsolar modestly by Dqsolar. The balance of the heating load must be supplied with purchased energy (Whp and qaux), since qair is equal to zero for both systems. The net increase in F is simply (qsolar/QL). The heat pump seasonal COP (ratio of rejected heat to work input) varies between the systems. The use of a solar source for the heat pump raises the seasonal COP over that of the conventional heat pump and parallel heat pump systems. As shown in Fig. 9, the seasonal heat pump heating COP for the parallel, dual source and series heat pump systems are 3.0, 3.5 and
Fig. 9. Heating contributions from all possible sources.
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
1393
4.0, respectively. As expected, the COP for the series and dual source heat pumps are substantially higher, since they utilize the solar energy storage tank as a heat source. 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 of either ambient air or energy storage. This is because the solar assisted series heat pump system operates only down to a source temperature of 108C, while the dual source heat pump system utilizes colder ambient air as a source when the energy storage tank has reached the solidi®cation temperature of the PCM (<108C). 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. Fig. 10 shows the monthly average daily global, extraterrestrial, diuse and beam radiation. We measured the global and diuse radiation, and then we calculated the extraterrestrial and beam radiation by using equations given in the literature [27] from experimental results. The maximum values are observed at the month of June. Fig. 11 shows the temperature variation of calcium chloride hexahydrate with time of day in the storage tank during the experiment. As shown in Fig. 11, there is strati®cation in the energy storage tank, and the temperature variation of the PCM with time of day is acceptable from the view point of melting behaviour. Fig. 12 also shows the temperature variation of the heat transfer ¯uid (water) with time of day during the charging and discharging periods in the experimental study.
Fig. 10. Monthly average daily extraterrestrial, global, diuse and beam radiations.
1394
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
Fig. 11. Temperature variation of calcium chloride hexahydrate with time of day in the storage tank.
Fig. 12. Temperature variation with time of day during charging and discharging period.
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
1395
7. Conclusions We have investigated the performance of combined solar heat pump heating systems both experimentally and theoretically at the Karadeniz Technical University. In the experimental study, over the heating season of 1992 (from December to May), the COP, net energy savings, collector and storage eciencies of the system are measured and calculated. Some experimental results are given in Table 5. The average seasonal COP values of the series, parallel and dual source heat pump systems for two heating seasons are 4.0, 3.0 and 3.5, respectively. The average seasonal collector eciencies of the solar, parallel and series systems are also 50, 50 and 60%, respectively. The average seasonal storage eciencies of the same systems are 55, 53 and 60%, respectively. The percent of heat load met (F ) by the series, parallel, dual and solar systems are also 0.60, 0.75, 0.80 and 0.25, respectively. The seasonal performance factors (SPF) of the series, parallel and dual source systems are 3.30, 3.70 and 4.20, respectively. The dual source system saved net energy of 12,056 kW per heating season, while the parallel system saved 10,120 kW and the series system saved 9390 kW of energy. It is concluded that the dual source heat pump system takes advantage of the best features of the series and parallel systems. The solar-only system is not convenient alone for our region because the F value is much less than for the other systems, since our region has more cloudy days. Acknowledgements The authors wish to acknowledge support for this research by the Karadeniz Technical University Research Fund under research grant No. 89.112.003 and TUB|Ç TAK under the research grant No. MISAG-11.
Table 5 Experimental performance of the heat pump systems over heating season of 1992. Note: the stated values of QL, COP, ncol, nst and ¦ are averages taken over a month Series system
Parallel system
Month
QL (kW)
Ta (8C)
Na
COP
ncol
nst
f
Na
COP
ncol
nst
f
Nov. Dec. Jan. Feb. Mar. Apr. May
1973 3175 4337 4383 3028 2216 1665
12.6 8.3 4.1 3.7 8.9 11.6 13.6
20 11 15 18 25 23 25
4.60 4.53 4.50 4.45 4.53 4.51 4.70
0.62 0.60 0.57 0.63 0.64 0.56 0.60
0.58 0.62 0.61 0.63 0.58 0.60 0.57
0.66 0.40 0.50 0.64 0.82 0.78 0.82
30 31 27 20 31 30 31
3.00 3.02 2.80 2.79 3.04 3.16 3.21
0.52 0.53 0.54 0.49 0.50 0.48 0.51
0.57 0.56 0.55 0.53 0.53 0.55 0.54
0.98 0.84 0.51 0.37 0.86 0.98 0.99
a
Number of working days of the heat pump systems per month.
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
1396
References [1] [2] [3] [4] [5] [6] [7] [8] [9] [10] [11] [12] [13] [14] [15] [16] [17] [18] [19] [20] [21] [22] [23] [24] [25] [26] [27] [28]
Howell JR, Bannerot RB, Vliet GC. Solar thermal energy systems. New York: McGraw-Hill, 1982. Freeman TL, Mitchell JW, Audit TE. Solar Energy 1979;22:125±35. Karman VD, Freeman TL, Mitchell JW. Proc. ISES, 1976. 15±20 August, Canada. Anderson JV, Mitchell JW, Beckman WA. Solar Energy 1980;25:155±63. Mitchell JW, Freeman TL, Beckman WA. Solar Age 1978;3:24±8. Chandrashekar M, Sullivan HF, Hollands KT. Solar Energy 1982;28:217±26. Tassou SA, Marquant CJ, Wilson DR. Applied Energy 1986;24:127±38. Calm JM. The heat pump. ASRAE J 1984;40±45 (August). York T. Heat pumps: Developing the dual fuel option. EPRI J 1990;22±27 (December). Kaygusuz K. Applied Energy 1993;45:257±67. Abhat A. Solar Energy 1983;30:313±32. Carlsson B, Stymne H, Wettermark G. Solar Energy 1979;23:343±50. Parisini FC. Solar Energy 1988;41:193±7. GuÈltekin N, Ayhan T, Kaygusuz K. Energy Convers Mgmt 1991;32:311±8. Brandstetter A. Solar Energy 1988;41:183±91. Telkes M. ASHRAE J 1974;16:38±46. Huang BK, Toksoy M, Cengel YA. Solar Energy 1986;37:279±92. Kaygusuz K. Energy Convers Mgmt 1995;36:315±23. Ghoneim AA, Klein SA. Solar Energy 1989;42:441±7. Dietz D. Trans ASME J Solar Energy Engineering 1984;106:106±11. Manton BE, Mitchell JW. Trans ASME J Solar Energy Engng 1982; (August). McGraw BA, Bedinger A, Reid R. Proc ISES Annual Meeting, 1981. p. 562±566. Bond TY. ASHRAE Transactions 1984;90:617±27. Bessler WF, Hwang BC. 1979 ASME/AIChEng 18th Heat Transfer Conf, August. Kaygusuz K, GuÈltekin N, Ayhan T. Energy Convers Mgmt 1993;34:335±46. Kaygusuz K. Applied Energy 1995;51:93±109. Due JA, Beckman WA. Solar engineering of thermal processes. New York: John Wiley, 1991. Beckman WA, Klein SA, Due JA. Solar heating design by the f-chart method. New York: John Wiley, 1977.
Experimental and theoretical investigation of combined solar heat pump system for residential heating K. Kaygusuz a,*, T. Ayhan b b
a Chemistry Department, Karadeniz Technical University, 61080, Trabzon, Turkey Mechanical Engineering Department, Karadeniz Technical University, 61080, Trabzon, Turkey
Received 1 June 1998; accepted 28 December 1998
Abstract In order to investigate the performance of the combined solar-heat pump system with energy storage in encapsulated phase change material (PCM) packings for residential heating in Trabzon, Turkey, an experimental set-up was constructed. Also, a comparative study of the performance of this system has been undertaken theoretically. Simulations have been made with BASIC computer program of three basic combined con®gurations, as well as conventional solar and conventional heat pump systems in Trabzon by using meteorological data. The experimental results were obtained from December to May during the heating season for four heating systems. These systems are a conventional solar system, a series heat pump system, a parallel heat pump system and a dual source heat pump system. The experimentally obtained results are used to calculate the collector eciency, heat pump coecient of performance (COP), seasonal heating performance, the fraction of annual load met by free energy, storage and collector eciencies and total energy consumption of the systems during the heating season. # 1999 Elsevier Science Ltd. All rights reserved. Keywords: Heat pump; Residential heating; Solar heating system; Energy storage
1. Introduction Solar energy systems and heat pumps are two promising means of reducing the consumption of fossil energy resources (coal, petroleum, etc.) and, hopefully, the cost of delivered energy for * Corresponding author. E-mail address: [email protected] (K. Kaygusuz) 0196-8904/99/$ - see front matter # 1999 Elsevier Science Ltd. All rights reserved. PII: S 0 1 9 6 - 8 9 0 4 ( 9 9 ) 0 0 0 2 6 - 6
1378
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
Nomenclature Ac Ae Cpa Cpw I Ucol (ta )e FR Zcol Zdis Zchar Zst mw mac Qcon PCM COP Wcomp Wpump Wcf Tw Tf,in Tf,out Tiwex Towex Tiws Tows Tref1 Tref2 Tacon1 Tacon2 Trcon1 Trcon2 Trev1 Trev2 Tind Ta Tb A U DD
solar collector area (m2) area of collector absorber (m2) speci®c heat of air (kJ/kg K) speci®c heat of water (kJ/kg K) incident solar radiation (W/m2) collector thermal loss coecient (W/m2 K) eective absorptance of cover-absorber assembly heat removal factor collector eciency discharge eciency of storage tank charge eciency of storage tank overall storage eciency mass ¯ow rate of water in system (kg/h) mass ¯ow rate of air in condenser (kg/h) heat extracted by condenser (kJ/h) phase change material heat pump coecient of performance work input to compressor (kJ/h) work input to water circulating pump (kJ/h) work input to compressor fan (kJ/h) water inlet temperature (K) collector water inlet temperature (K) collector water outlet temperature (K) inlet water temperature of water-to-air heat exchanger (K) outlet water temperature of water-to-air heat exchanger (K) inlet water temperature of energy storage tank (K) outlet water temperature of energy storage tank (K) inlet refrigerant temperature of water source evaporator (K) outlet refrigerant temperature of water source evaporator (K) inlet air temperature of condenser (K) outlet air temperature of condenser (K) inlet refrigerant temperature of condenser (K) outlet refrigerant temperature of condenser (K) inlet refrigerant temperature of air source evaporator (K) outlet refrigerant temperature of air source evaporator (K) indoor air temperature (K) ambient air temperature (K) base temperature (K) ¯oor area of building (m2) overall thermal loss coecient [kW/(m2 K day)] number of degree-days per month
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
n h F QL qsolar qair Whp E Cmin (TS)ict
1379
number of days in month heat transfer coecient fraction of annual load met by free energy (%) annual heating load of building (kW) energy supplied by solar system energy supplied by air source heat pump from ambient air energy supplied by series heat pump system eectiveness of heat exchanger thermal capacity of heat exchanger (kJ/h K) ideal control (switch-over) temperature (K)
residential heating. An intelligent extension is to try to combine the two to further reduce the cost of delivered energy. In general, it is widely believed that combined systems will save energy, but what is not often known is the magnitude of the possible energy saving and the value of those savings relative to the additional expense [1]. Solar heat pump systems are classi®ed 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 hot water. In the series system, solar energy is supplied to the evaporator of the heat pump, thereby raising its temperature and increasing the coecient of performance (COP). In the dual source con®guration, 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 [2±6]. The objective of this paper is to analyse several viable types of combined solar heat pump systems and to compare them to conventional solar and conventional heat pump (air-to-air) systems experimentally and theoretically. Our aim is to determine the thermal performance of the dierent solar assisted heat pump system con®gurations and to provide some insight into many of the non-intuitive reasons behind the performance results. We used an energy storage tank with three combined heat pump systems for storing solar energy in the experimental study. This energy storage can reduce the time or rate mismatch between energy supply and energy demand, thereby playing a vital role in energy conservation. In processes which are wasteful of energy, energy storage will result in a saving of premium fuels. Energy storage can also improve system performance and increase reliability. On the other hand, conventional vapour compression heating only heat pumps are not at present an economic alternative to gas-®red heating systems due to their high capital cost and the degradation in their steady state eciency caused by on/o cycling supplementary electric resistance heating and frosting and defrosting losses [7±10]. The combination of a heat pump and a solar energy system would appear to reduce many of the disadvantages that each has when operating separately or alone. During winter conditions, the energy that could be collected by the solar collectors, but which would be too low in temperature to be useful for direct heating, may be used as a heat source for the heat pump. Since the solar collection±storage system can supply energy at temperatures higher than the ambient air temperature, the capacity and heat pump COP would increase over that if the heat
1380
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
pump works alone, the peak auxiliary load requirement would be reduced and the combined heating system (solar and heat pump) will be operated more economically. On the other hand, the operation of the solar system at temperatures near and below the inside temperature of the building will decrease the collector losses and allow more energy to be collected. The lower collection temperature may then allow the use of collectors with one or no cover, which would reduce the ®rst cost from that of a conventional two-cover solar system. Finally, for those areas where warm temperatures occur during cloudy periods, the combined system may compensate for the reduced performance of the conventional solar system under cloudy conditions and the low capacity of the heat pump in cold winter weather conditions. In the present work, an experimental set-up was constructed to determine the performance of the series, parallel and dual source heat pump, solar collectors and energy storage tank ®lled by PCM used for residential heating. The eects of various system parameters on the response of the indoor air temperature of the building, the temperature variation of the PCM in the energy storage tank and the temperatures of the heat transfer ¯uid (water) in the solar collectors and energy storage tank for series, parallel and dual source heat pump systems were investigated. Also, the collector and storage eciencies, heat pump COP, seasonal performance factor (SPF), the building heating load, heat supplied fraction of the heating load by the systems and energy consumption of the combined solar heat pump system during the heating season (December to May) in 1992 were calculated. On the other hand, the theoretical analyses have been made using a BASIC computer simulation program containing a solar assisted heat pump system with latent heat energy storage tank. The original strategy, namely an hour-byhour simulation of the dynamic behaviour of the system, involved driving it by hourly meteorological data. 2. Description of the experimental set-up The combined solar heat pump system linked to a latent heat energy storage tank presented here is at Trabzon, Turkey, latitude 418N, longitude 408E, placed on the Black Sea side. Table 1 gives the climatic data of the heating season for Trabzon. A schematic overview of the system is given in Fig. 1. The solar collectors used in this system were constructed by modifying ¯at-plate waterTable 1 Climatic conditions of Trabzon over heating season in 1992 Month
Nov.
Dec.
Jan.
Feb.
Mar.
Apr.
May
Average outdoor temperature (8C) Minimum outdoor temperature (8C) Maximum outdoor temperature (8C) Average relative humidity (%) Average wind velocity (m/s) Average solar radiation (MJ/m2 day) Degree days
12.6 6.8 19.0 76.4 2.4 5.78 125
8.3 2.1 20.1 68.6 2.4 4.38 283
4.1 ÿ1.3 11.4 66.7 2.3 5.43 297
3.7 12.0 18.4 67.2 2.1 7.17 305
8.9 1.0 26.2 65.2 2.1 11.74 313
11.6 6.0 25.0 71.9 2.3 14.32 246
13.9 7.6 22.1 81.4 1.9 14.46 158
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
1381
Fig. 1. Schematic overview of the heating system.
cooled collectors. Each collector was made of pipe and ®n type; the absorber unit consists of nine copper tubes (1.8 m length and 0.022 m outside diameter) ®tted longitudinally at 0.1 m pitch across an aluminium sheet (1.8 m length and 0.9 m width). The aluminium sheet has 0.55 mm thickness and 1.62 m2 eective absorber area. The plate is painted with blackboard paint (a=0.90 and e=0.80) and placed inside a metal box made of 0.8 mm thick aluminium sheet and insulated at the bottom and all sides with glass wool of thickness 6 cm [k=0.038 W/ (m 8C)]. The top of the metal box is glazed with a 3.5 mm thick glass cover having (ta )e=0.80. The loss coecient of the collector is 8.20 (W/m2 8C), and the constant heat removal factor FR=0.9 for each collector. The 18 collectors were connected in a parallel series combination. The mass ¯ow rate of the heat transfer ¯uid through the collectors was 43 kg/h m2. The solar collectors were installed at an angle of 488 from the horizontal and faced due south. The compressor used for the combined solar heat pump system was a hermetic type. It was driven by a 1490 watt electrical motor. The heat pump can use ambient air and the solar energy storage tank as a heat source. So, it has two evaporators, air-cooled and water-cooled. The heat pump condenser is an aircooled ®nned tube type heat exchanger. Recently, the use of phase change materials (PCMs) for thermal energy storage in solarassisted heating systems has received considerable attention. The motivation for using phase change energy storage (PCES) is the reduction in storage volume which can be achieved compared to sensible heat storage. Therefore, we used commercial calcium chloride hexahydrate (CaCl2.6H2O) as a PCM for solar energy storage because it is cheaper (10$/100 kg), and it has a better thermal stability than the other PCMs [11±15]. It is commercial grade material produced by the Dow Chemical Company. The melting temperature (288C) of this material is almost an optimum temperature for residential heating applications in some regions
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
1382
of Turkey. Also, the large amounts of heat that can be stored in the phase change of CaCl2.6H2O can reduce the storage volume to 1/8 to 1/15 of that required for rock or gravel storage and to 1/5 to 1/10 of that required for water storage [16±20]. The storage tank was made of sheet iron, and has a diameter of 1.3 m and a length of 3.2 m. It consists of a vessel packed in the horizontal direction with cylindrical tubes. The energy storage material is inside the tubes (tubes are made of PVC plastic), and the heat transfer ¯uid (HTF) ¯ows parallel to it. The storage tank contains cylindrical PVC containers ®lled with PCM. The heat storage tank was linked to the heat pump by means of an evaporator for use as a heat source. The heat storage tank was insulated entirely with 60 mm thick glass wool. The heat loss from the tank to its surroundings is 0.25 (W/m2 8C). In the experiments, the inlet and outlet pressures of the compressor and evaporators were measured with manometers. Also, the inlet and outlet temperatures of the refrigerant (R-22) at the condenser, compressor and evaporators were measured with iron-constantan thermocouples. In addition, the inlet and outlet temperatures of the HTF (water) at the storage tank, collectors, water source evaporator and water-to-air heat exchanger were measured with the same kind of thermocouples. Also, the temperatures of the PCM in the energy store were measured. In addition, the ambient and indoor air temperatures were measured with normal thermometers. A Kipp±Zonen solarimeter, mounted on the vertical plane of the solar collectors, was used to measure the global solar radiation on the tilted surface, and some results are given in Table 2. The ¯ow rate of the circulated water through the system was measured by means of two ¯ow meters. The power consumption of the systems was measured with a wattmeter. A laboratory building of 75 m2 ¯oor area was used as the space to be heated. The structural properties of the building are given in Table 3.
Table 2 Hourly average global solar radiation on a horizontal surface (W/m2) Hour
Jan.
Feb.
Mar.
Apr.
May
June
July
Aug.
Sept.
Oct.
Nov.
Dec.
4±5 5±6 6±7 7±8 8±9 9±10 10±11 11±12 12±13 13±14 14±15 15±16 16±17 17±18 Total
0.0 0.0 0.0 94 223 384 433 419 356 230 98 0.0 0.0 0.0 2442
0.0 0.0 66 244 419 523 615 631 614 530 370 174 0.0 0.0 4187
0.0 0.0 87 230 475 656 733 742 719 607 447 166 73 0.0 4936
0.0 133 321 474 565 663 740 747 733 621 530 370 174 27 6097
63 210 381 545 678 768 817 827 795 719 614 433 258 70 7178
81 253 477 644 777 860 902 916 853 777 637 519 344 120 8160
75 217 392 588 692 779 827 837 805 728 624 443 267 77 7351
82 193 335 481 617 691 726 719 677 610 481 368 187 68 6235
0.0 140 314 453 586 670 698 684 642 537 384 209 56 0.0 5372
0.0 31 167 300 454 544 593 572 523 412 244 77 0.0 0.0 3907
0.0 0.0 70 216 342 426 475 461 398 279 140 27 0.0 0.0 2836
0.0 0.0 0.0 77 216 356 391 405 349 244 147 49 0.0 0.0 2236
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
1383
Table 3 Construction properties of the laboratory building Properties
Value
Window area (single glass, U=4.8 W/m2 8C day) Wall area (single brick, U=1.6 W/m2 8C day) Floor area (concrete, U=2.5 W/m2 8C day) Ceiling area (concrete+¯at metal, U=2.0 W/m2 8C day) Eective UA [kWh/(8C day)] Comfort temperature (8C) Average degree-days in the heating season Average total heating load in the heating season (kW) Dimensions of the building (3.5 m 60 m 12.0 m)
75 m2 60 m2 75 m2 75 m2 0.800 20.0 1727 20772 242 m3
3. System description and operation modes We will de®ne a base solar heating system and a base heat pump heating system as standards of comparison for the combined solar heat pump systems to follow. The base solar system is the standard liquid heating system shown in Fig. 2. It consists of conventional ¯atplate water-cooled solar collectors, an energy storage tank, a liquid-to-air heat exchanger coil in the laboratory building supply duct, a water circulating pump, auxiliary space heaters (electrical resistance heaters) and other control equipment. We have designed our experimental set-up to form dual sources for the heat pump. Because of this, the system can be used in three forms, series, parallel and dual source systems. This process can be accomplished by changing the heat sources for the heat pump. More detailed description of the systems is given below. 3.1. Solar assisted series heat pump system with storage As shown in Fig. 3, this series system consists of conventional ¯at-plate water-cooled solar collectors, an energy storage tank ®lled by PCM as a heat storage material, a heat pump with water-to-refrigerant heat exchanger, an air-cooled condenser, a water circulating pump, a liquid-to-air heat exchanger for direct solar heating, auxiliary heater and other conventional equipment. In this system, the hot water which comes from the collectors ®rst goes to the
Fig. 2. Base solar system with storage.
1384
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
Fig. 3. Solar assisted series heat pump system with storage.
energy storage tank where it releases some energy to the PCM in the storage, and after this, it is used as a heat source by the water source evaporator. After that, it is sent to the solar collectors by the water circulating pump. However, at night and on cloudy days, the less-hot water that comes from the evaporator of the heat pump was sent to the energy storage tank instead of the solar collectors. The cold water extracts energy from the PCM in the storage, and it ¯ows to the evaporator for use as a heat source. 3.2. Solar assisted parallel heat pump system with storage As shown in Fig. 4, this parallel system consists of conventional ¯at-plate water-cooled solar collectors, water-to-air heat exchanger, a heat pump with air-to-refrigerant heat exchanger, an air-cooled condenser, a latent heat energy storage tank, a water circulating pump, an auxiliary electrical heater and other control equipment. The solar-assisted parallel heat pump system is combining two main components, a conventional solar heating system and a conventional airto-air heat pump system. In this system, the heat pump uses ambient air as an energy source, while the water-to-air heat exchanger uses solar energy as a heat source, and they each individually give their energies to the heating load of the building. Solar energy is used to meet as much of the heating requirement as possible. Thus, the total available energy of the system
Fig. 4. Solar assisted parallel heat pump system with storage.
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
1385
is the sum of the extracted energies from two dierent systems (solar system and heat pump system). 3.3. Dual source system The dual source system contains two systems: the series system and the parallel system described above. In the dual source system, the heat pump has two evaporators: one is the water source evaporator which is placed in the storage tank loop and the other is the air source evaporator which is placed outside of the room. This allows the heat pump to use either the collected solar energy or the ambient air as a heat source, depending on which results in a higher COP. There are three heating modes for the dual source system. In the ®rst case, the dual source system is operated in the direct solar heating mode. Here, the storage tank is hotter than the predetermined control value (below 303 K) and able to heat the laboratory building directly. The heat pump is o, and any excess solar energy is being stored in the tank. In the second case, when the energy storage tank temperature is below the control value, but higher than the minimum using value (280 K) and ambient air temperature, the hot water in the storage tank is pumped to the evaporator and used as a heat source for the heat pump. Depending on the magnitude of the building heating load, the excess solar energy is collected in the energy storage tank. In the third case, when the storage temperature is either at its minimum temperature or less than the ambient air temperature, the ambient air is used as a heat source for the evaporator, and auxiliary heat is supplied when needed.
Fig. 5. Power consumption versus heating load for the heat exchanger and heat pump for entering water temperature at 295 K.
1386
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
3.4. Ideal control (switch-over) 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 [21±24]. The general practice is to have the control (switch-over) change from using direct solar heating to using the heat pump, or vice versa, take place at some ®xed entering water temperature which is called the switch-over temperature. In previous studies, Manton and Mitchell [21] used 303 K, McGraw et al. [22] used 302.4 K and Bond [23] proposed various temperatures depending on the building load at the switch-over temperature. For example, Bond proposed 303 K with a 12.5 kW building heating load. Bessoler and Hwang [24] used a dierent control strategy in their simulation study in which direct solar heating was used when the entering water temperature was above 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. However, these temperatures were determined with the primary purpose of maintaining the supply air temperature instead of minimizing energy consumption. Comparing the system's power consumption when the heat exchanger is used for direct solar heating and when the heat pump is used, as shown in Fig. 5, it is seen that when the entering water temperature is below the desired indoor air temperature (293 K), the system always consumes less power when using the heat pump than when using the heat exchanger, regardless of the building heating load, because the water-to-air heat exchanger cannot provide any heat by using water at this low temperature, and all heating of the building has to be provided with
Fig. 6. Ideal switch-over (control) temperature versus building heating load.
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
1387
the auxiliary heater or by using the heat pump. However, when the entering water temperature is suciently high (333 K), it is always more energy conserving to use the heat exchanger. In the case of the solar assisted series heat pump system, the entering water to the water source evaporator is a maximum temperature of 313 K at a 1200 kg/h water circulating rate through the system under the heating season conditions. The melting temperature of the CaCl2.6H2O is around 303 K, and its latent heat energy storage temperature is nearly constant at temperatures between 303±310 K. So, at this maximum entering temperature, sometimes it is more energy conserving to use the water-to-air heat exchanger and sometimes more energy conserving to use the heat pump. In this study, for most days, the entering water temperature varies between 298 to 308 K, and the building load is around 20 kW/day. So, if we use the heat pump as much as possible, more energy conservation can occur for residential heating. In the present study, the ideal control (switch-over) temperature was determined by using the following equation, and then the calculated temperature values were plotted against building heating load as shown in Fig. 6. The ideal control temperature is given by [23] as TS ict Tind
QL ÿ Wcomp : ECmin
1
4. Calculation of the experimental results The heat pump COP is de®ned as COP Qcon =Wcomp :
2
The COP of the solar assisted series heat pump with storage can be calculated as [25] COP
mac Cpa Tacon2 ÿ Tacon1 : Wcomp Wpump Wcf
3
The COP of the solar assisted parallel heat pump system with storage can be calculated as COP
mac Cpa Tacon2 ÿ Tacon1 maex Cpa Tex2 ÿ Tex1 : Wcomp Wpump Wcf Wexf Wevf
The instantaneous collector eciency is given by [1] Ucol Ae Tfi ÿ Ta : Zcol FR taeff ÿ Ac I
4
5
Also, the net collector eciency can be calculated as Zcol
mw Cpw Tf,in ÿ Tf,out : Ac I
6
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
1388
The thermal performance of the energy storage can be evaluated by its discharge eciency, charge eciency and overall storage eciency. The de®nitions of these eciencies are [26] Discharge efficiency
Zdis
Actual heat output Ideal heat output
mw Cpw Tiws ÿ Tows dt mw Cpw Tst ÿ Tiws dt
Charge efficiency
Zchar
7
Actual heat received Ideal heat input
mw Cpw Tiws ÿ Tows dt ÿ L dt mw Cpw Tiws ÿ Tst dt
8
The overall storage eciency is the product of the discharge and charge eciencies: Zst Zdis Zchar :
9
The monthly heating load can be calculated by using the following equation [27] QL UAh :DD: The monthly average degree-days (DD) can be approximated by ln cosh 1:698h 3=2 h DD sm n 0:2042 2 3:396
10
11
where n is the number of days in the month and h is de®ned as h
Tb ÿ Ta : sm :Vn
12
The standard deviation of the monthly average ambient temperature is sm which is generally not available from weather summaries. It can be approximated by the following with essentially no loss in accuracy: sm 1:45 ÿ 0:0290Ta 0:0664syr
13
where syr is the standard deviation of the monthly average ambient temperature from the annual average ambient temperature. So, the annual heating loads of the laboratory building were calculated by using the above equations, and the derived results are given in Table 3. 5. Computer simulation of the system The complexity of the thermal analyses of solar assisted heat pump systems makes use of computer simulations the only method for suitably determining the system dynamics and
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
1389
performance. These simulations were performed with the BASIC computer simulation program. This computer program contains some subroutines which model individual pieces of hardware (e.g. collectors, energy store, air and water source heat pump, building heating load), and an executive routine which links these component models and solves the resulting system of equations. The heat pump model used in these simulations is quasi-steady state in nature. The heat pump has two heat sources for the evaporator, i.e. air and water. The actual performance data obtained from the experimental results are used to generate third-order polynomials relating heat pump COP to the source temperature. For the dual source heat pump, two dierent sets of polynomials are used, one set relating to the water source and the other for the air source. For the water source heat pump: COPH 5:46 5:33 10ÿ2 Tw ÿ 5:53 10ÿ4 Tw 2 1:20 10ÿ6 Tw 3
14
Qcon 21:42 ÿ 5:62 10ÿ2 Tw ÿ 7:47 10ÿ4 Tw 2 2:63 10ÿ6 Tw 3 :
15
For the air source heat pump: COPH ÿ27:86 0:121 Ta 1:601 10ÿ4 Ta 2 ÿ 7:035 10ÿ7 Ta 3
16
Table 4 BASIC computer simulation required inputs Data Value of a Value of b Value of FR Value of (ta )e Length of simulation (min) Calculation interval (min) Collector aperture area (m2) Collector mass ¯ow rate (kg/min) Value of Is,ref used in b (W/m2) Constant heating load (J/h) Load mass ¯ow rate (kg/min) Thermal capacity of storage (J/K) Initial temperature (K) Minimum temperature to heating load (K) Maximum allowed storage temperature (K) Storage tank overall loss coecient [W/(m2 8C)] Energy storage tank length/diameter ratio Treated as sinusoidal around ambient average (K) Day length (sun rise±sun set) (min) Maximum solar radiation (W/m2) Time of sunrise (h) Speci®c heat capacity of working ¯uid [J/(kg 8C)]
Value 0.0 1.0 0.9 0.80 900 15 30 21.6 800 1.45 105 21.6 2.53 108 293 293 312 0.250 2.46 281 600 900 07:00 4183
1390
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
Qcon 18:45 ÿ 0:101 Ta 6:508 10ÿ5 Ta 2 5:044 10ÿ7 Ta 3 :
17
The solar system modelled is a conventional liquid medium system. The collector parameters include FR, b, (ta )e and Is,ref. The solar radiation was chosen as a repeating sinusoidally varying function, and in this case, the time of sunrise, day length and peak radiation were speci®ed. The heating load was speci®ed as constant, given at each time. The storage capacity is speci®ed, and the storage was chosen as being strati®ed according to the experimental results. Heat losses from storage are taken into account by specifying the overall thermal loss coecient. The ambient temperature is modelled as a daily sinusoidal variation around an average ambient temperature. The BASIC computer simulation required inputs are given in Table 4. The building used in the computer simulation is the laboratory, and its structure properties are given in Table 3. The building has 75 m2 ¯oor area and is not well insulated. 6. Results and discussion Fig. 5 shows the power consumption versus heating load for the heat exchanger and heat pump at the entering water temperature of 295 K. The ®gure also shows that the power consumption increases with building heating load.
Fig. 7. Fraction of annual load meet by free energy as a function of collector area.
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
1391
Fig. 6 shows the ideal switch-over (control) temperature versus building heating load; it also shows that the ideal switch-over temperature increases with building load. Its minimum value is equal to the indoor temperature, and its maximum value is determined by the characteristic of the heat exchanger, as given by its ECmin value. The function F is shown in Fig. 7 for Trabzon as a function of collector area for the conventional air source heat pump, conventional single-cover solar system, series heat pump, parallel heat pump and dual source heat pump systems. For a building with neither a conventional solar heating system nor a heat pump system, the F value equals zero. For a building (in our simulations the laboratory building with 75 m2 ¯oor area) with only a conventional heat pump (air-to-air), 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 COP and the relative size of the space to be heated and building heating load during the heating season. In the case of the conventional solar system, F depends on collector area and storage mass. As collector area and storage mass increase, F increases. The F curves for the conventional solar system are in agreement with results predicted by the f-Chart method [28] to within a few percent as shown in Fig. 7 for Trabzon during the heating season. 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 m2 for a climate like Trabzon. The simulation results show that the seasonal collector performance for the solar assisted parallel heat pump and conventional solar systems of the same collector area are equal. In the case of the solar assisted series and dual source heat pump systems, the collector performance of the same collector area are not equal. Clearly, the improved collection eciency is a direct
Fig. 8. Collector eciencies of combined heating system.
1392
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
result of the solar assisted heat pump capability, which maintains lower average storage temperatures and, hence, lower collector temperature in the series system. Fig. 8 shows the collector eciencies for 1-cover solar assisted series and dual source heat pump systems and for 1-cover conventional solar and parallel heat pump systems for the entire heating season. The seasonal energy balance requires that the sum of all energy supplied equal the heating load of the building, or qsolar qair Whp qaux ÿ QL :
18
The relative contributions from each of these four heat sources is shown in the bar graphs of Fig. 9. The combined height of the qsolar and qair bars in Fig. 9 represent the percentage of the total heating requirement supplied by free energy and is, therefore, equal to the F value. As shown in Fig. 9, adding the solar energy capability to the conventional solar system to create the series system increases qsolar modestly by Dqsolar. The balance of the heating load must be supplied with purchased energy (Whp and qaux), since qair is equal to zero for both systems. The net increase in F is simply (qsolar/QL). The heat pump seasonal COP (ratio of rejected heat to work input) varies between the systems. The use of a solar source for the heat pump raises the seasonal COP over that of the conventional heat pump and parallel heat pump systems. As shown in Fig. 9, the seasonal heat pump heating COP for the parallel, dual source and series heat pump systems are 3.0, 3.5 and
Fig. 9. Heating contributions from all possible sources.
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
1393
4.0, respectively. As expected, the COP for the series and dual source heat pumps are substantially higher, since they utilize the solar energy storage tank as a heat source. 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 of either ambient air or energy storage. This is because the solar assisted series heat pump system operates only down to a source temperature of 108C, while the dual source heat pump system utilizes colder ambient air as a source when the energy storage tank has reached the solidi®cation temperature of the PCM (<108C). 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. Fig. 10 shows the monthly average daily global, extraterrestrial, diuse and beam radiation. We measured the global and diuse radiation, and then we calculated the extraterrestrial and beam radiation by using equations given in the literature [27] from experimental results. The maximum values are observed at the month of June. Fig. 11 shows the temperature variation of calcium chloride hexahydrate with time of day in the storage tank during the experiment. As shown in Fig. 11, there is strati®cation in the energy storage tank, and the temperature variation of the PCM with time of day is acceptable from the view point of melting behaviour. Fig. 12 also shows the temperature variation of the heat transfer ¯uid (water) with time of day during the charging and discharging periods in the experimental study.
Fig. 10. Monthly average daily extraterrestrial, global, diuse and beam radiations.
1394
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
Fig. 11. Temperature variation of calcium chloride hexahydrate with time of day in the storage tank.
Fig. 12. Temperature variation with time of day during charging and discharging period.
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
1395
7. Conclusions We have investigated the performance of combined solar heat pump heating systems both experimentally and theoretically at the Karadeniz Technical University. In the experimental study, over the heating season of 1992 (from December to May), the COP, net energy savings, collector and storage eciencies of the system are measured and calculated. Some experimental results are given in Table 5. The average seasonal COP values of the series, parallel and dual source heat pump systems for two heating seasons are 4.0, 3.0 and 3.5, respectively. The average seasonal collector eciencies of the solar, parallel and series systems are also 50, 50 and 60%, respectively. The average seasonal storage eciencies of the same systems are 55, 53 and 60%, respectively. The percent of heat load met (F ) by the series, parallel, dual and solar systems are also 0.60, 0.75, 0.80 and 0.25, respectively. The seasonal performance factors (SPF) of the series, parallel and dual source systems are 3.30, 3.70 and 4.20, respectively. The dual source system saved net energy of 12,056 kW per heating season, while the parallel system saved 10,120 kW and the series system saved 9390 kW of energy. It is concluded that the dual source heat pump system takes advantage of the best features of the series and parallel systems. The solar-only system is not convenient alone for our region because the F value is much less than for the other systems, since our region has more cloudy days. Acknowledgements The authors wish to acknowledge support for this research by the Karadeniz Technical University Research Fund under research grant No. 89.112.003 and TUB|Ç TAK under the research grant No. MISAG-11.
Table 5 Experimental performance of the heat pump systems over heating season of 1992. Note: the stated values of QL, COP, ncol, nst and ¦ are averages taken over a month Series system
Parallel system
Month
QL (kW)
Ta (8C)
Na
COP
ncol
nst
f
Na
COP
ncol
nst
f
Nov. Dec. Jan. Feb. Mar. Apr. May
1973 3175 4337 4383 3028 2216 1665
12.6 8.3 4.1 3.7 8.9 11.6 13.6
20 11 15 18 25 23 25
4.60 4.53 4.50 4.45 4.53 4.51 4.70
0.62 0.60 0.57 0.63 0.64 0.56 0.60
0.58 0.62 0.61 0.63 0.58 0.60 0.57
0.66 0.40 0.50 0.64 0.82 0.78 0.82
30 31 27 20 31 30 31
3.00 3.02 2.80 2.79 3.04 3.16 3.21
0.52 0.53 0.54 0.49 0.50 0.48 0.51
0.57 0.56 0.55 0.53 0.53 0.55 0.54
0.98 0.84 0.51 0.37 0.86 0.98 0.99
a
Number of working days of the heat pump systems per month.
K. Kaygusuz, T. Ayhan / Energy Conversion & Management 40 (1999) 1377±1396
1396
References [1] [2] [3] [4] [5] [6] [7] [8] [9] [10] [11] [12] [13] [14] [15] [16] [17] [18] [19] [20] [21] [22] [23] [24] [25] [26] [27] [28]
Howell JR, Bannerot RB, Vliet GC. Solar thermal energy systems. New York: McGraw-Hill, 1982. Freeman TL, Mitchell JW, Audit TE. Solar Energy 1979;22:125±35. Karman VD, Freeman TL, Mitchell JW. Proc. ISES, 1976. 15±20 August, Canada. Anderson JV, Mitchell JW, Beckman WA. Solar Energy 1980;25:155±63. Mitchell JW, Freeman TL, Beckman WA. Solar Age 1978;3:24±8. Chandrashekar M, Sullivan HF, Hollands KT. Solar Energy 1982;28:217±26. Tassou SA, Marquant CJ, Wilson DR. Applied Energy 1986;24:127±38. Calm JM. The heat pump. ASRAE J 1984;40±45 (August). York T. Heat pumps: Developing the dual fuel option. EPRI J 1990;22±27 (December). Kaygusuz K. Applied Energy 1993;45:257±67. Abhat A. Solar Energy 1983;30:313±32. Carlsson B, Stymne H, Wettermark G. Solar Energy 1979;23:343±50. Parisini FC. Solar Energy 1988;41:193±7. GuÈltekin N, Ayhan T, Kaygusuz K. Energy Convers Mgmt 1991;32:311±8. Brandstetter A. Solar Energy 1988;41:183±91. Telkes M. ASHRAE J 1974;16:38±46. Huang BK, Toksoy M, Cengel YA. Solar Energy 1986;37:279±92. Kaygusuz K. Energy Convers Mgmt 1995;36:315±23. Ghoneim AA, Klein SA. Solar Energy 1989;42:441±7. Dietz D. Trans ASME J Solar Energy Engineering 1984;106:106±11. Manton BE, Mitchell JW. Trans ASME J Solar Energy Engng 1982; (August). McGraw BA, Bedinger A, Reid R. Proc ISES Annual Meeting, 1981. p. 562±566. Bond TY. ASHRAE Transactions 1984;90:617±27. Bessler WF, Hwang BC. 1979 ASME/AIChEng 18th Heat Transfer Conf, August. Kaygusuz K, GuÈltekin N, Ayhan T. Energy Convers Mgmt 1993;34:335±46. Kaygusuz K. Applied Energy 1995;51:93±109. Due JA, Beckman WA. Solar engineering of thermal processes. New York: John Wiley, 1991. Beckman WA, Klein SA, Due JA. Solar heating design by the f-chart method. New York: John Wiley, 1977.