Energy and Exergy Efficiency of a Packed-bed Heat Storage Unit for Greenhouse Heating

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Biosystems Engineering (2003) 86 (2), 231–245 doi:10.1016/S1537-5110(03)00134-X SE}Structures and Environment

Energy and Exergy Efficiency of a Packed-bed Heat Storage Unit for Greenhouse Heating . ztu. rk; A. Ba-sc-etin-celik H.H. O Department of Agricultural Machinery, Faculty of Agriculture, University of C - ukurova, Adana 01330, Turkey; e-mail of corresponding author: [email protected] (Received 23 December 2002; accepted in revised form 7 July 2003; published online 4 September 2003)

In this research, solar energy was stored daily using the volcanic material with the sensible heat technique for heating the tunnel greenhouse of 120 m2. The external heat collection unit consisted of 27 m2 of south-facing solar air heaters mounted at a 558 tilt angle. The dimensions of the packed-bed heat storage unit were 6 m by 2 m by 0
6 m deep. The packed-bed heat storage unit was built under the soil at the centre of the tunnel greenhouse. The heat storage unit volume per square metre of ground surface of the tunnel greenhouse was 0
06 m3, while the storage volume per square metre of the heat collection unit was about 0
27 m3. The heat storage unit was filled with 6480 kg of volcanic material equivalent to 54 kg of heat storage material per square metre of the greenhouse ground surface. Energy and exergy analyses were applied in order to evaluate the system efficiency. During the charging periods, the average daily rates of thermal energy and exergy stored in the heat storage unit were 1242 and 36
33 W, respectively. It was found that the net energy and exergy efficiencies in the charging periods were 39
7 and 2
03%, respectively. During the discharging periods, the average daily rates of the thermal energy and exergy recovered from the heat storage unit were 601
3 and 20
9 W, respectively. The results showed that 18
9% of the total heating requirement of the tunnel greenhouse was obtained from the heat storage unit. # 2003 Silsoe Research Institute. All rights reserved Published by Elsevier Ltd

1. Introduction The primary objective of greenhouses is to produce agricultural products, outside the cultivation season. They have considerable importance in the market of agricultural production. In order to achieve optimum indoor conditions, it is necessary to heat the greenhouses, particularly during the cold seasons. However, even in the southern areas, heating cost exceeds 30% of the overall operational cost of a greenhouse (Santamouris, 1993). Therefore, greenhouse sector has a great potential for energy conservation. Heating applications in greenhouses have an important effect on the yield as well as on the quality and the cultivation time of the products. Optimisation of air temperature in greenhouses is of particular importance in relation to plant growth and development. In order to keep the greenhouse temperature considerably higher than outside temperature, heating is frequently needed during cold seasons. The fuel consumption then becomes an 1537-5110/$30.00

important economic factor. This problem can be overcome by using low-cost heating technologies instead of heating by fossil fuels. Owing to the relatively high cost and uncertain availability of fossil fuels, considerable attention has been given to new and renewable energy sources as alternative means of heating greenhouses. Moreover, developing efficient and economical heat storage systems and related devices is as important as developing new energy sources from the point of view of energy conservation. Solar energy which is an abundant, clean and safe source, is an attractive substitute for conventional fuels for passive and active heating of greenhouses. During the day, excess solar heat is collected for short- or longterm storage and it is recovered at night in order to satisfy the heating needs of the greenhouses. Efficient and economical heat storage is the main factor in the utilisation of solar energy for agricultural purposes. Solar thermal energy can be stored as sensible heat, latent heat, heat of reaction, or a combination of these. 231

# 2003 Silsoe Research Institute. All rights reserved Published by Elsevier Ltd

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. ZTU¨RK; A. BAS,C¸ETINC¸ELIK H.H. O

Notation Ac Ag As cp It l1 l2 m1 m2 Qd Qg Ql Qs Qt t Td Ti Tg Tm To Tos Ts

surface area of the greenhouse cover, m2 ground area of the greenhouse, m2 surface area of the heat storage unit, m2 specific heat of the heat transfer fluid, J kg1 K1 solar radiation, W m2 thickness of the polyethylene, m thickness of glass wool, m flow rate of the charging fluid, m3 s1 flow rate of the discharging fluid, m3 s1 rate of heat recovery from the heat storage unit, W heat requirement of the greenhouse, W m2 rate of heat loss from the heat storage unit, W rate of heat stored into the heat storage unit, W rate of heat transfer into the heat storage unit, W time, s outside air temperature, K inlet temperature of the heat transfer fluid, K air temperature in the greenhouse, K temperature of the heat storage material, K outlet temperature of the heat transfer fluid, K reference outside temperature, K soil temperature inside the greenhouse, K

In most storage systems, it is stored as sensible heat in . ztu. rk et al., 1999). materials such as water and rocks (O In air collection systems, rock beds are normally used to store heat, while water tanks store heat in liquid systems. In latent heat storage systems, the latent heat arising from phase change of a material is used for thermal energy storage. The phase change materials (PCM) can store large amounts of heat (latent heat of fusion) in the change of the phase from solid to liquid. Latent heat storage systems using PCM, in general, provide much higher energy storage density than systems using sensible heat storage. For solar heating applications in the greenhouses, the use of the packedbed heat storage units for thermal energy storage has become an attractive design option in terms of constructive cost and storage efficiency (Willits & Peet, 1987). The efficiency of the seasonal storage as well as that of the daily storage depends on the system configurations, the climate conditions, and various setpoints for environmental control of the greenhouse. Traditional energy analysis, based on the first law of thermodynamics, is concerned only with the efficiency of energy processes (Fang et al., 1995). Exergy analysis, derived from both the first and second laws of thermodynamics, as compared to energy analysis takes

u vw W1 Xd Xl Xs Xt g Zc(net) Zc(total) le l1 l2 r t Cc(net) Cc(total)

overall heat loss coefficient, W m2 K1 wind velocity, m s1 power of the electric motor, W rate of exergy recovery from the heat storage unit, W rate of exergy loss from the heat storage unit, W rate of exergy stored into the heat storage unit, W rate of exergy transfer into the heat storage unit, W factor for the thermal radiation net energy efficiency for charging, % net energy efficiency for charging period, % equivalent heat conduction coefficient, W m1 K1 heat conduction coefficient of polyethylene, W m1 K1 heat conduction coefficient of glass wool, W m1 K1 density of the heat transfer fluid, kg m3 transmittance for solar radiation net exergy efficiency for charging period, % total exergy efficiency for charging period, %

into account the quality of the energy transferred. The main purpose of the exergy analysis is to determine the reasons for the thermodynamic faults of the thermal and chemical processes. Exergy analysis can provide greater insight to cost effective design and management of complex processes (Larson & Cortez, 1995). Exergy analysis is recognised by many engineers to be a powerful tool for the evaluation of the thermodynamic systems in general, and of thermal energy storage systems in particular. While many technically and economically successful thermal storage systems are in operation, no broadly valid basis for comparing the achieved performance of one storage system with that of another operating under different conditions has found general acceptance (Rosen et al., 1988). From a first law perspective, the efficiency of a thermal energy storage system can be assessed in terms of how much thermal energy the system can store. Thus, one system is considered to be more effective than another if, for the same energy input in the hot fluid stream entering the system and the same amount of storage material, it is capable of storing more energy. This approach produces workable designs, but not necessarily those with the highest possible thermodynamic efficiencies. It has been shown in recent years that the design of thermodyna-

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mically efficient heat transfer equipment must be based on the second law of thermodynamics in addition to the first law (Krane, 1987). Exergy analysis for the heat storage systems has generally used in order to optimise system efficiency and to model system parameters. Therefore, this study is devoted to an analysis and operation of a large-scale sensible heat storage system used to heat greenhouses. In this paper, energy and exergy analyses were applied for evaluating the efficiency of a specific heat storage system. A number of investigations which aimed at improving the greenhouse thermal performance and environment have been done by some researchers in the design, modelling and testing of energy conserving and solar adopted greenhouses. Arinze et al. (1984) developed a dynamic mathematical model for predicting temperatures and moisture levels in a solar assisted, energy conserving, circular type greenhouse. Willits and Peet (1987) investigated factors affecting the performance of rock storage as solar energy collection/storage systems for the greenhouses. They took the data from two similar rock-storage systems of slightly different design attached to two different greenhouses. Kurata and Takakura (1991) investigated underground storage of solar energy for greenhouse heating. Possibilities for seasonal storage of solar energy in the soil under a greenhouse were investigated and compared with those of daily storage. They assumed a system composed of collectors, a greenhouse, pipes connected to the collectors and buried under the greenhouse for water circulation, and another set of underground pipes for the greenhouse air circulation. A numerical experiment showed that under the conditions tested, electric energy consumed in water and air circulation in seasonal storage is greater than energy saved in greenhouse heating. Comprehensive studies have been carried out concerning the application and modelling of a latent heat storage system and the thermal properties of PCM by many researchers. Short-term latent heat storage systems utilising salt hydrates have been applied to horticulture. The most commonly used PCM for domestic applications are Glauber’s salt (Na2SO4.10H2O) and calcium chloride hexahydrate (CaCl2.6H2O). Calcium chloride hexahydrate, which has a latent heat of fusion of 193 kJ kg1 at the phase change temperature of 278C, has been used to heat 6 m by 6 m fibreglasscoated greenhouse by Kern and Aldrich (1979). They design and built separate internal/external collection and storage systems. The external collection system consisted of two 8
55 m2 of south-facing air collectors mounted at a 588 tilt angle. The internal collection system consisted of a 0
15 m diameter clear tube mounted in each ridge of the greenhouse 0
15 m below

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the top. While collection/storage efficiencies of the internal system were in the range of 6–8%, the efficiencies of the external system ranged from 38 to 43%. Jaffrin and Cadier (1982) used a system utilising a heat storage with 13
5 t of CaCl2.6H2O, placed in 9000 flat bags of 1 l capacity each, weighing 1
5 kg, to heat a 500 m2 multi-span single glazed greenhouse. In terms of greenhouse heating needs, an 80% savings in gas and 75% in terms of total energy (gas+electricity) were achieved compared with the control. Huang et al. (1986) used the commercial cylindrical storage rods which were filled with CaCl2.6H2O in the greenhouse solar system. Their results showed that the designed latent heat storage system had significantly higher compact storage capacity than water or rock storage. Ting et al. (1990) incorporated a PCM energy storage unit into a CO2 enriched, 24
09 m2 polyethylene (PE) greenhouse. The PCM was packed in tubes formed by sealing two flexible sheets, and the tube-sheets were installed in a container located inside of the greenhouse. The melting point of the mixture was approximately 238C which is suitable for the dual purpose of the greenhouse cooling and heating. Ku. rklu. et al. (1995) developed a mathematical model for the prediction of the thermal performance of a PCM store containing 1 m long and 38 mm diameter polypropylene tube. Most of the greenhouse heating demand can be supplied with the latent heat storage systems. Puri and Zuritz (1985) pointed out that the PCM can supply 37% of annual heating demand for the greenhouse without insulation curtains and this fraction can be as high as 80% for an insulated greenhouse. However, their preliminary financial analysis indicated that installation cost of PCM is significantly higher than the energy savings. The system described by Balducci (1985) (2800 kg of chloride material) satisfied 22% of annual heating needs of a glass greenhouse with a 200 m2 ground surface. Baille and Boulard (1987) used CaCl2.6H2O in a greenhouse covered with double skin polycarbonate. For the spring tomato crop, their system supplied 41% of the heating load and energy saving was about 30%. In Turkey, agricultural crops which are mainly vegetables and ornamentals were produced in total 43 139 ha protected cultivation area in 2001. This area consisted of 6015 ha glasshouses and 37 124 ha are plastic houses which include low and high plastic tunnels (SIS, 2001). A small proportion of the greenhouse owners use the auxiliary heating systems only during the coldest winter nights. On the other hand, Turkey has also great solar energy potential due to its location in the Mediterranean Region. The sunshine period of Turkey is 2624 h yr1 with a maximum of 365 h month1 in July and minimum of 103 h month1

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in December. The main solar radiation intensity is about 3
67 kWh m2 day1. Cumulative total of this is about 1
340 MWh m2 yr1. The amount of the solar radiation received over all of Turkey, in other words, the gross . ltanır, 1994). solar energy potential is 3517 EJ yr1 (U In this research, solar energy was stored using a volcanic material as heat storage material with the sensible heat technique for heating a plastic tunnel greenhouse of 120 m2. The performance of the packedbed heat storage unit was evaluated by energy and exergy analyses and, energy and exergy efficiencies of the system were compared during the charging periods. The objectives of the present study are: (1) to determine the annual fraction of the greenhouse heating demand that can be supplied by the packed-bed heat storage unit under ambient conditions of C - ukurova; (2) to investigate the thermal performance of the sensible heat storage unit under the operating conditions; (3) to determine the efficiency of daily storage in the packedbed heat storage unit; (4) to compare energy efficiency of the packed-bed heat storage unit with its exergy efficiency; and (5) to determine the feasibility of using a heat storage material as a possible alternative to other storage techniques for storage of solar thermal energy for C - ukurova climate.

collection unit, (2) heat storage unit, (3) tunnel greenhouse, (4) heat transfer unit, and (5) data acquisition unit. The tunnel greenhouse heating system utilising the daily heat storage located in the C - ukurova region of Turkey. The constructional features of the system units in detail were described in the following sections.

2.1. Materials

2.1.1. Heat collection unit Solar air heaters were used to recover solar energy. The external heat collection consisted of 27 m2 of southfacing solar air heaters mounted at a 558 tilt angle (Fig. 2). The solar air heaters that have packed airflow passages were used in the heat collection unit. Each of the solar air heaters has an absorber plate area of 1
5 m2. The Raschig ring type of packing was used to increase heat transfer from the plates to the heat transfer fluid underneath the absorber plates of the air heaters. The characteristic diameter of the Raschig rings, made of the polyvinyl chloride tube, was 0
05 m. The aluminiumbased absorber of the air heaters is covered with a 4 mm glass sheet and the underside is insulated with glass wool. The dimensions of the air heaters are 1
9 m by 0
9 m and 18 of them form the external heat collection system. The absorber surface of the installed air heaters was 0
225 m2 per square metre of the tunnel greenhouse ground surface. The heat collection unit was supplied with air from the environment by a centrifugal fan that had a volumetric flow rate of 600 m3 h1. The heat collection unit was installed outside the greenhouse and mounted on a galvanised steel structure.

In the following work, heating the tunnel greenhouse of 120 m2 floor area utilising solar energy and daily heat storage within the packed-bed heat storage unit using the sensible heat storage technique is attempted. The schematic arrangement of the system is given in Fig. 1. The system consists of mainly five units: (1) heat

2.1.2. Heat storage unit The packed-bed heat storage unit, which consisted of a 6 m by 2 m by 0
6 m deep, was used as the heat storage unit. The packed-bed heat storage unit was installed under the soil at the centre of the tunnel greenhouse. The heat storage unit volume per square metre of the

2. Materials and methods

FAN Fan A

Valve 1

Fig. 1. The arrangement of the heat storage and greenhouse heating system

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Outlet 1900 1470

Inlet

1450

1960

Fan

5840

940

Fig. 2. The external heat collection unit; all dimensions in mm

tunnel greenhouse ground surface was 0
06 m3, while the storage volume per square metre of the heat collection unit was about 0
27 m3. The whole surface of the heat storage unit was insulated with 0
2 mm of PE film and 0
05 m of glasswool, respectively. The heat storage unit was attached to the tunnel greenhouse by means of PE pipes. Solar energy collected by the solar air heaters was transferred to the heat storage unit by circulating air through PE pipes. Perforated PE pipe as the heat exchanger was installed into the heat storage unit. The total length and diameter of the perforated PE pipe were 44 and 0
1 m, respectively. 2.1.3. Heat storage material In general, materials that have a large change in internal energy per unit volume minimise the space needed to store energy. However, the following properties of the heat storage material for the sensible heat storage systems must be taken into account in the selection of storage material: heat capacity, density, heat storage temperature, storage material cost, containment and heat exchange costs, thermal conductivity, vapour pressure, toxicity, and corrosiveness. The values of

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specific heat and density of water, granite, limestone, brick and concrete for sensible heat storage are 4190 J kg1 K1 and 1000 kg m3, 820 J kg1 K1 and 2640 kg m3, 900 J kg1 K1 and 2500 kg m3, 1 1 3 840 J kg K and 1698 kg m , and 1130 J kg1 K1 3 . ztu. rk & Ba-sc-etin-celik, and 2240 kg m , respectively (O 2002). The cost per unit storage capacity must be low if the system economics are to be reasonable. This requires not only that the cost of the storage media be low, but also that containment and heat exchange costs be reasonable. The difficulties of the vapour pressure of water and the limitations of other liquids can be avoided by storing thermal energy as sensible heat in solids. However, larger amounts of solids are needed than water since the heat storage density of solids is usually less than water. The cost of the storage media per unit energy stored, although not as low as for water, is still acceptable for rocks. Direct contact between the solid storage media and a heat transfer fluid is vital to minimise the cost of heat exchange in a solid storage medium. Due to the high storage capacity per unit volume and inexpensive containment of solid heat storage media, volcanic material was used as the heat storage material in the packed-bed heat storage unit. The heat storage material required for greenhouse heating was calculated on the basis of heat storage capacity per unit volume of selected volcanic material and the total heat requirement of the tunnel greenhouse. The heat storage unit was filled with 6480 kg of volcanic material equivalent to 54 kg of heat storage material per square metre of the greenhouse ground surface. Bulk density and porosity of the heat storage material were 900 kg m3 and 41
22, respectively. 2.1.4. Tunnel greenhouse The experiments were carried out in a semi-cylindrical tunnel greenhouse that was aligned north–south. The tunnel greenhouse which consisted of galvanised steel tube has continuous side openings operated by a rolling mechanism. The openings in the sidewalls which were created by rolling up or down the plastic film was used for ventilation. The tunnel greenhouse was covered with double skin PE material (thickness of 0
35 mm) that contains ultraviolet and infrared stabilisers. The dimensions of the tunnel greenhouse were: width 6 m; length 20 m and height 3 m. The warm air from the heat storage unit was distributed by perforated PE ducts lying on the ground surface inside the tunnel greenhouse. 2.1.5. Heat transfer unit In this experiment, heat transfer with the forced convection between the heat collection unit, heat storage unit and tunnel greenhouse was carried out two centrifugal fans. Two differential thermostats were used

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to control the ventilating and extraction process by controlling the centrifugal fans which can be operated independently. When the difference between the outside air temperature and the storage system temperature exceeds the preset value, one of the differential thermostats activated one of the fans (ventilating fan). During the sunny days, heat collected by the heat collection unit was used to charge the heat storage unit. During the ventilating operation, the air from outside was circulated through the heat collection unit by the ventilating fan, and then, was passed through the heat storage unit. When the air temperature of the tunnel greenhouse fell below a preset value, the second fan (extractor fan) was activated during the extraction operation. The extractor fan drove the greenhouse air through the packed-bed heat storage unit during the night, and then returned it to interior of the plastic tunnel via the perforated PE ducts. The operation of the electric motor activating the extractor fan was controlled by time-clock between 22:00 h and 02:00 h. The directions of the heat flow between the heat collection unit, heat storage unit and tunnel greenhouse were controlled by two valves. 2.1.6. Data acquisition unit All air temperatures at the inlets and outlets of the heat collection and the heat storage units, and inside the tunnel greenhouse were measured with the thermistors. The range of the thermistors was 20 to +808C and the accuracy was 0
28C over 0–708C. The sensors consisted of a stainless-steel clad thermistor probe with a 5 m cable. The temperatures of the heat storage materials and the circulating air as the heat transfer fluid were measured with 2 kO thermistors distributed uniformly throughout the heat storage unit. Twelve thermistors were placed in the packed-bed heat storage unit to measure the temperature of the heat storage material. The thermistors that were used to measure the temperature of the heat storage material were installed in two rows at 0
6 m depth in the packed-bed heat storage unit. The distance between the two rows was 0
7 m. The thermistors were located at the distance of 1 m in the packed-bed heat storage unit. These sensors were positioned in such a way that the surface temperature of the heat storage material was measured in the heat storage unit. Eight thermistors were located to measure air temperature at the inlet and outlet of the heat storage unit. The average temperatures of the heat storage material and heat transfer fluid were determined by averaging the measurements of the sensors. Air temperature inside the tunnel greenhouse was measured with 2 kO hermetically sealed thermistors with an accuracy of 0
18C over 0–808C; the accuracy was 0
138C at –208C. The sensors was mounted in an open-cylindrical probe made from a material which has

a low affinity for water and which fits inside a cylindrical louvred radiation screen, made of anodised aluminium, which protects the sensor against solar radiation and rain. Six sensors were used to measure air temperature inside the tunnel greenhouse. These sensors were placed at height of 1
5 m at the three different locations of the tunnel greenhouse. Two sensors were placed in the inlet, at the middle and in the end of the tunnel greenhouse. The flow rates of the heat transfer fluid were measured at the inlet and outlet of the heat storage unit. A data-logger was used for taking and storing readings from the sensors; it could accept digital inputs in the form of voltages, resistances, counts and frequencies. The recorded data were stored in memory for output to a printer or to a computer for storage on disk. Data can be retrieved from the logger and the current readings of the sensors can be examined without interrupting the logging process. Readings can be taken at regular intervals, which can be different for each channel. To optimise use of the logger memory, timed readings taken on a channel over specified periods can be recorded as single values, representing the average, maximum or minimum reading for the period. Recording of the temperatures was made at 1 s intervals and averaged over 30 min in the experiment. 2.2. Methods 2.2.1. Energy analysis for the charging and discharging periods The rate of heat transfer Q’ t in W from the heat collection unit into the heat storage unit was calculated during the charging period by using the following equation: ’ 1 rcp ½Ti ðtÞ  To ðtÞ Q’ t ðtÞ ¼ m ð1Þ ’ 1 is the flow rate of the charging fluid in m3 s1; where: m r is the density of the heat transfer fluid in kg m3; cp is the specific heat of the heat transfer fluid in J kg1 K1; Ti and To are the inlet and outlet temperatures of the heat transfer fluid in K; and t is the time in s. The rate of heat stored in the heat storage unit Q’ s in W was determined with respect to the heat transfer rate into the heat storage unit and heat losses from the heat storage unit for the charging period: Q’ ðtÞ ¼ Q’ ðtÞ  Q’ ðtÞ ð2Þ s

t

l

where: Q’ l is rate of heat loss from the heat storage unit in W. The rate of heat loss from the whole surface area of the heat storage unit was calculated from Q’ l ðtÞ ¼ As le ½Tm ðtÞ  Ts ðtÞ ð3Þ where: As is the surface area of the heat storage unit in m2; le is the equivalent heat conduction coefficient in

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W m1 K1; Tm is the temperature of the heat storage material in K; and Ts is the soil temperature inside the greenhouse in K. It was assumed that heat was transferred from the heat storage unit to the surroundings by the conduction. The equivalent heat conduction coefficient was calculated on the basis of the heat conduction coefficients and the thickness of the insulating materials as follows: l1 þ l2 le ¼ ð4Þ l1 =l1 þ l2 =l2 where: l1 is the thickness of the PE in m; l2 is thickness of glass wool in m; l1 is the heat conduction coefficient of PE in W m1 K1; and l2 is the heat conduction coefficient of glass wool in W m1 K1. Energy efficiency for the charging period was defined as the ratio of the heat stored in the heat storage unit to the heat transfer from the heat collection unit. Then, the total energy efficiency during the charging period Zc(total) in % can be formulated as follows: Q’ ðtÞ ð5Þ ZcðtotalÞ ðtÞ ¼ s 100 Q’ ðtÞ t

When the power consumed by the electric motor which was used to activate the charging fan for the charging period is considered, the net energy efficiency during the charging period Zc(net) in % can be written as follows: Q’ s ðtÞ

100 ð6Þ ZcðnetÞ ðtÞ ¼ ’1 Q’ t ðtÞ þ W ’ 1 is the power of the electric motor in W. where: W The rate of heat recovered from the heat storage unit during the discharging period Q’ d was calculated from ’ 2 rcp ½To ðtÞ  Ti ðtÞ Q’ ðtÞ ¼ m ð7Þ d

where: m2 is the flow rate of the discharging fluid during the discharging period in m3 s1. 2.2.2. Exergy analysis for the charging and discharging periods The rate of thermal exergy transfer from the heat collection unit into the heat storage unit X’ t in W was calculated during the charging period from the following . ztu. rk, 1997): equation (O Ti ðtÞ ’ 1 rcp ln ð8Þ X’ t ðtÞ ¼ Q’ t ðtÞ  Tos :m To ðtÞ where: Tos is the reference outside temperature in K. It is important to mention that the rate of thermal exergy transfer is usually calculated on the basis of the rate of thermal energy transfer. The symbol X refers to the rate of exergy transfer in the literature of thermodynamics. The rate of the thermal exergy stored in the heat storage unit X’ s in W was determined in relation to the exergy transfer rate into the heat storage unit and the

exergy losses from the heat storage unit for the charging period: ’ t ðtÞ  X ’ l ðtÞ ’ s ðtÞ ¼ X ð9Þ X where: X’ l is the rate of thermal exergy loss from the heat storage unit in W. During the charging period, the rate of the thermal exergy losses associated with the heat losses from the heat storage unit to the surroundings was evaluated as follows:   ’ l ðtÞ ¼ Q’ l ðtÞ 1  Ts ðtÞ X ð10Þ Tm ðtÞ Similarly energy efficiency, the total and net exergy efficiencies during the charging period were determined by the following equations, respectively. X’ s ðtÞ

100 ð11Þ CcðtotalÞ ðtÞ ¼ X’ t ðtÞ CcðnetÞ ðtÞ ¼

’ s ðtÞ X

100 ’ t ðtÞ þ W1 X

ð12Þ

where: Cc(net) is the net exergy efficiency during the charging period in %; and Cc(total) is the total exergy efficiency during the charging period in %. The rate of thermal exergy recovered from the heat ’ d in W was storage unit for the discharging period X calculated on the basis of the rate of heat recovered from the heat storage unit Q’ d as follows: To ðtÞ ’ d ðtÞ ¼ Q’ d ðtÞ  Tos m ’ 2 rcp ln X ð13Þ Ti ðtÞ 2.2.3. Heat requirement of the tunnel greenhouse The heat requirement of the tunnel greenhouse ground surface Q’ g in W m2 was calculated from the following equation (Bailey, 1988). Ac Q’ g ðtÞ ¼ u½Tg  Td ðtÞ  I’t tg ð14Þ Ag where: Ac is the surface area of the greenhouse cover in m2; Ag is the ground area of the greenhouse in m2; u is the overall heat loss coefficient in W m2 K1; Tg is the air temperature in the greenhouse in K; Td is the outside air temperature in K; I’t is the total solar radiation in W m2; t is the transmittance of the greenhouse cover for solar radiation; and g is the conversion factor of global radiation energy to thermal energy within the greenhouse. g is the portion of the solar radiation entering the greenhouse which is used to increase the internal temperature. The overall heat loss coefficient u represents the total energy loss in W per m2 of external area of the greenhouse for a difference of 1 K between the inside and outside temperatures. The value of the overall heat loss coefficient depends on especially external climatic

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conditions (principally on the wind speed but also on rain and snow), it is always given in relation to the wind speed. A number of relationships have been developed to predict the overall heat loss coefficient of different types of greenhouses. The relationship between the overall heat loss coefficient and the wind speed for the single transparent covered greenhouse was reported by Damrath (1982) as follows: u ¼ 3
97 þ 0
16vw

ð15Þ

where: v is the wind velocity in m s1. In this experimental study, the following relationship that . ztu. rk (1997) for the obtained by Ba-sc-etin-celik and O tunnel greenhouse (double skin PE) in which the experiment was carried out was used to calculate the overall heat loss coefficient: u ¼ 3
55 þ 0
11vw

ð16Þ

In Eqn (14), the value of t for single foil ranges from 0
6 to 0
7, for double foil greenhouses from 0
5 to 0
6. During the day, a part or all of the energy may be supplied by solar radiation. During the night, the heat storage system will provide most of the energy. The value of g depends on the proportion of the floor, which is covered by plants, and generally lies in the range of 0
3–0
7 (Bailey, 1988). In this study, 0
6 and 0
5 as the values for t and g were used to calculate the heat requirement of the tunnel greenhouse.

3. Results and discussion There are only a few publications relevant to exergetic efficiency of the large-scale heat storage applications. Previous studies on this subject were restricted to theoretical analysis of the system efficiency. On the other hand, there are a lot of factors affecting the efficiency of the packed-bed heat storage systems as solar energy collection/storage systems for the greenhouses. The efficiency of the heat storage system depends on thermal and physical properties of the heat storage material (specific heat, thermal conductivity, density), heat storage temperature, geometry of heat exchanger, and system configuration. Furthermore, an exergetic efficiency of the heat storage systems is a new approach to determine the net efficiency of the system. Therefore, the results of this experimental study could not be compared with those in previous studies in details. The results of an energy and exergy analyses of the sensible heat storage system for the greenhouse heating were evaluated for the following charging periods: (1) first charging period (from 13 to 18 January, 1998), (2) second charging period (from 4 to 9 March,

1998), and (3) third charging period (from 1 to 7 April, 1998).

3.1. The results of the energy and exergy analyses during the charging periods 3.1.1. The rate of the heat and thermal exergy stored in the heat storage unit The rates of heat and thermal exergy stored in the heat storage unit during the charging periods were calculated by using Eqns (2) and (9), respectively. The results of an energy and exergy analyses during the charging periods are shown in Fig. 3. The rates of heat and thermal exergy stored in the heat storage unit increased with increasing inlet temperature of the heat transfer fluid during the charging periods. The changes of the rate of heat and thermal exergy stored in the heat storage unit during the charging periods are shown as a function of time in Fig. 3. The values of heat and thermal exergy stored in the heat storage unit during the charging periods are given in Table 1. Figure 3 shows that the average hourly rate of thermal energy and thermal exergy changed with time during the charging periods. During the first charging period [Fig. 3(a)], the rate of the heat stored in the heat storage unit ranged from 519 W to 1
47 kW, whereas the rate of the thermal exergy stored in the heat storage unit was in the range of 1
223
1 W. The rate of thermal exergy which was only 1
2 W at 10:00 h reached to the maximum value (23
1 W) at 13:00 h. In this charging period, average daily rates of the heat and thermal exergy stored in the heat storage unit were 1
15 kW and 14
3 W, respectively. During the second charging period [Fig. 3(b)], the rate of heat and thermal exergy stored in the heat storage unit were in the range of 0
7341
47 kW and 14
646
6 W, respectively. During the second charging period, average daily rate of the heat and thermal exergy stored in the heat storage unit were 1
16 kW and 33
7 W, respectively. During the third charging period [Fig. 3(c)], similar to the first and second charging periods, the rate of the heat stored in the heat storage unit ranged from 78
4 W to 2
02 kW, while the rate of the thermal exergy stored in the heat storage unit was in the range of 689
9 W. In this charging period, the average daily rates of the heat and thermal exergy stored into the heat storage unit were 1
40 kW and 61 W, respectively. The rate of thermal exergy was significantly different from the rate of heat stored into the heat storage unit during the charging periods. This difference is due to the fact that the quality of the energy was taken into account to calculate the thermal exergy transfer [Eqn (9)]. However, in the calculation of thermal energy

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25

15 1000 10 500

5

0

0 10:00

11:00

(a)

12:00 13:00 Time, h

14:00

15:00

50 40

1500

30 1000 20 500

10

0

Rate of exergy, W

Rate of heat, W

2000

0 10:00

11:00

(b)

Rate of heat, W

Rate of exergy, W

20

1500

12:00 13:00 Time, h

14:00

15:00

2500

100

2000

80

1500

60

1000

40

500

20

0

Rate of exergy, W

Rate of heat, W

2000

0 10:00

11:00

(c)

12:00 13:00 Time, h

14:00

15:00

Fig. 3. The rates of heat and thermal exergy stored in the heat storage unit during the charging periods: (a) first charging period; (b) second charging period; (c) third charging period: , rate of heat , rate of exergy

transfer [Eqn (2)], the quantity of the energy was taken into account. In other words, the useless part of the thermal energy stored in the heat storage unit was ignored in the calculation of the thermal exergy transfer. Differences between the charging periods in terms of the heat and exergy stored in the heat storage unit could be explained as follows. (1) During the third charging period, the average daily rate of the heat and thermal exergy stored in the heat storage unit were higher than that of the first and second charging periods. Since the rate of exergy depends on the temperature of the heat transfer fluid and surrounding, the rate of exergy increased with the increasing difference between the inlet and outlet temperatures of the heat transfer fluid during the charging periods. (2) The rate of exergy stored in the heat storage unit during the first charging period was lower than that of the second and third charging periods due to the fact that the rate of exergy is higher at high temperature than that of low temperature. As can be expected, the temperature of the heat transfer fluid was the highest during the third charging period. (3) Moreover, the percentage increase in the thermal exergy at high temperatures was higher compared with the thermal energy. In this study, it was also found that the percentage increase in the thermal exergy was higher than that of the thermal energy during the charging periods. As shown in Table 1, the average daily rates of heat stored in the heat storage unit during the second and third charging periods were 1
16 and 1
40 kW, while the average daily rates of the thermal exergy stored in the heat storage unit during the second and third charging periods were 33
7 and 61 W, respectively. In this case, the thermal energy increased 16
7% from the second charging period to the third charging period, whereas the thermal exergy increased 44
8% for the same charging periods. These results may be simply explained in terms of heat transfer considerations. The possible amount of thermal energy stored in the heat storage unit decreased with the decreasing difference between the inlet and outlet temperatures of the

Table 1 The values of heat and thermal exergy stored into the heat storage unit during the charging periods Charging periods

1 2 3

Heat, W

Exergy, W

Min.

Max.

Average

Min.

Max.

Average

519 734 78
4

1470 1470 2020

1150 1160 1400

1
2 14
6 6
0

23
1 46
6 89
9

14
3 33
7 61
0

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4

30

3

20

2

10

1

Exergy efficiency, %

50

0 0 10:00 10:30 11:00 11:30 12:00 12:30 13:00 13:30 14:00 14:30 15:00 15:30

(a) 50

5

40

4

30

3

20

2

10

1

Exergy efficiency, %

Time, h

0 0 10:00 10:30 11:00 11:30 12:00 12:30 13:00 13:30 14:00 14:30 15:00 15:30

(b)

Time, h 60

6

50

5

40

4

30

3

20

2

10

1

Exergy efficiency, %

3.1.2. Energy and exergy efficiency during the charging periods The net energy and exergy efficiencies of the heat storage system during the charging periods were calculated by using Eqns (6) and (13), respectively. The changes of the average hourly energy and exergy efficiencies during the charging periods are shown as a function of time in Fig. 4. The values of the net energy and exergy efficiencies can be seen in Table 2. The net energy and exergy efficiencies increased with the increasing the inlet temperature of the heat transfer fluid. The net energy efficiency ranged from 22
6 to 45
3%, while the net exergy efficiency was in the range of 0
071
3% during the first charging period [Fig. 4(a)]. While the net exergy efficiency at 10:00 h was only 0
07%, it reached to the maximum value (1
3%) at 13:00 h. During the first charging period, it was found that the average daily net energy and exergy efficiencies were 38
8 and 0
8%, respectively. During the second charging period [Fig. 4(b)], the net energy efficiency was in the range of 28
645
3%. On the other hand, the net exergy efficiency varied from 0
8 to 2
6% for the same charging period. The average daily net energy and exergy efficiencies were 40
0 and 1
9%, respectively. During the third charging period [Fig. 4(c)], the average daily net energy and exergy efficiencies were higher compared with first and second charging periods. The maximum values of the net energy and exergy efficiencies shown in Fig. 4(c) were 52
9% and 4
9%, respectively (Table 2). When the net energy efficiency is compared with net exergy efficiency during the charging periods, the following results can be drawn. (1) The net energy efficiency was always higher than the net exergy efficiency. This is expected because the total energy content of the hot air used as heat transfer fluid is taken into account in order to calculate the net energy efficiency. In other words, to calculate the net energy efficiency, the quantity of the energy transferred is taken into account, and the quality of the energy transferred is neglected. (2) For all charging periods, similar results were obtained in terms of the net energy efficiency of the heat storage system. The average daily net energy efficiency of the heat storage system remained nearly

Energy efficiency, %

heat transfer fluid during the charging periods. The percentage decrease in the possible amount of thermal energy stored in the heat storage unit was higher compared with the thermal exergy during the charging periods. The differences between the charging periods in terms of the heat and exergy stored in the heat storage unit are attributed to the inlet and outlet temperatures of the heat transfer fluid and the outside temperature.

Energy efficiency, %

. ZTU¨RK; A. BAS,C¸ETINC¸ELIK H.H. O

Energy efficiency, %

240

0 0 10:00 10:30 11:00 11:30 12:00 12:30 13:00 13:30 14:00 14:30 15:00 15:30

(c)

Time, h

Fig. 4. The change of the net energy and exergy efficiencies for the charging periods: (a) first charging period; (b) second charging period; (c) third charging period: &, energy efficiency, *, exergy efficiency

constant (approximately 40%) during the charging periods. The net energy efficiency of the heat storage system did not change much during the charging periods. However, the net exergy efficiency of the heat storage system changed depending on the charging periods. While the average daily net exergy efficiency was only 0
8% during the first charging period, it increased more than twofold (1
9%) during the second charging period. Then, the net exergy efficiency reached to the maximum value (3
4%) during the third charging

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Table 2 The values of energy and exergy efficiencies for the charging periods Exergy efficiency, %

Min.

Max.

Average

Min.

Max.

Average

22
6 28
6 4
1

45
3 45
3 52
9

38
8 40
0 40
4

0
07 0
80 0
34

1
3 2
6 4
9

0
8 1
9 3
4

1000

20

15 600 10 400 5

200 0

3.2.1. The rate of heat and exergy recovered from the heat storage unit The rates of heat and thermal exergy recovered from the heat storage unit during the discharging periods were calculated by using Eqns (7) and (14), respectively. The changes of the average hourly rates of heat and thermal exergy recovered from the heat storage unit during the discharging periods are shown as a function of time in Fig. 5. The amounts of heat and thermal exergy recovered from the heat storage unit during the discharging periods were also given in Table 3. The amount of heat recovered from the heat storage unit was in the range of 716873 W during the first

(a) 1000

30

800

25

Rate of heat, W

When the air temperatures inside the plastic tunnel greenhouse decreased below a set-value, the cold air inside the tunnel greenhouse was circulated through the heat storage unit during the night. In other words, the heat stored in the heat storage unit during the charging periods was recovered during the nights (22:00–02:00 h). The results of energy and exergy analyses of the sensible heat storage system for the greenhouse heating were evaluated for the following discharging periods: (1) the first discharging period (from 13 to 18 January, 1998), (2) the second discharging period (from 4 to 9 March, 1998); and (3) the third discharging period (from 1 to 7 April, 1998).

0 22:30 32:00 23:30 00:00 00:30 01:00 01:30 02:00 Time, h

20 600 15 400 10 200

5 0

0 (b)

22:30 32:00 23:30 00:00 00:30 01:00 01:30 02:00 Time, h 500

35 30

400 Rate of heat, W

3.2. The results of the energy and exergy analysis during the discharging periods

Rate of exergy, W

800

25 300

20

200

15 10

100

5

0

0 22:30

(c)

Rate of exergy, W

period. The above results indicate that the net exergy efficiency of the heat storage system was always higher than the net energy efficiency at higher temperatures. (3) It was found that the average net exergy efficiency was only 2% during charging periods. This result indicates that the heat storage system investigated in this study is inefficient in terms of the exergy efficiency. Krane (1987) also concluded that is sensible heat energy storage systems are inherently inefficient devices in terms of the exergy efficiency.

Rate of heat, W

1 2 3

Energy efficiency, %

Rate of exergy, W

Charging periods

23:30

00:30 Time, h

01:30

Fig. 5. The rates of heat and thermal exergy recovered from the heat storage unit for the discharging periods: (a) first charging period; (b) second charging period; (c) third charging period: , rate of heat; , rate of exergy

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Table 3 The values of heat and thermal exergy recovered from the heat storage unit during the discharging periods Discharging periods

1 2 3

Heat, W

Exergy, W

Min.

Max.

Average

Min.

Max.

Average

716 591 189

873 794 383

810 690 304

5
7 20
7 23
3

17
0 24
4 31
1

13
1 22
3 27
3

discharging period as shown in Fig. 5(a). However, it was found that the thermal exergy recovered from the heat storage unit ranged from 5
7 to 17 W for the same discharging period. During the first discharging period the average daily heat and thermal exergy recovered from the heat storage unit were 810 and 13
1 W, respectively (Table 3). During the second discharging period [Fig. 5(b)], the rate of heat recovered from the heat storage unit ranged from 591 to 794 W, whereas the thermal exergy recovered from the heat storage unit was in the range of 20
724
4 W. During the second discharging period, the average daily heat and thermal exergy recovered from the heat storage unit were 690 and 22
3 W, respectively. During the third discharging period (Table 3), the average daily heat and thermal exergy recovered from the heat storage unit were 304 and 27
3 W, respectively. The differences between the discharging periods in terms of the heat and exergy recovered from the heat storage unit could be explained as follows. (1) The maximum heat was recovered from the heat storage unit during the first discharging period. After the first discharging period, the rate of heat recovered from the heat storage unit decreased due to the decreasing the amount of the heat into the heat storage unit. While the average daily heat recovered from the heat storage unit was 809
7 W during the first discharging period, it decreased to 304 W during the third discharging period. This indicates that the heat recovered from the heat storage unit decreased by 62
5% from the first discharging period to the third. (2) Contrary to the amount of heat recovered from the heat storage unit, the thermal exergy recovered from the heat storage unit increased from the first discharging period to the third. While the average daily thermal exergy recovered from the heat storage unit was 13
1 W during the first discharging period, it increased to 27
3 W during the third discharging period. It is clear that the thermal exergy recovered from the heat storage unit increased more than twofold from the first discharging period to the third. This is because the outlet temperature of the heat transfers fluid during the third discharging period was higher than that of the first and second discharging periods.

3.2.2. Heat requirement of the tunnel greenhouse The overall heat loss coefficient and the total heat requirement of the tunnel greenhouse during the discharging periods were calculated from Eqns (16) and (14), respectively. The changes of the total heat requirement of the tunnel greenhouse and the amount of heat recovered from the heat storage unit during the discharging periods are shown as a function of time in Fig. 6. The total heat requirement of the tunnel greenhouse and the rate of heat recovered from the heat storage unit during the discharging periods were given in Table 4. During the first discharging period [Fig. 6(a)], the total heat requirement of the tunnel greenhouse ranged from 7
53 to 8
60 kW, while the rate of heat recovered from the heat storage unit to the tunnel greenhouse was in the range of 716873 W. While the average daily heat requirement of the tunnel greenhouse was 8
25 kW, the amount of heat recovered from the heat storage unit was only 810 W during this discharging period. In this case, 9
8% of the total heat requirement of the tunnel greenhouse was obtained from the heat storage unit. During the second discharging period [Fig. 6(b)], the rate of heat recovered from the heat storage unit to the tunnel greenhouse varied from 591 to 794 W. In this discharging period, 11
6% of the total heat requirement of the tunnel greenhouse was obtained from the heat storage unit. During the third discharging period [Fig. 6(c)], the total heat requirement of the tunnel greenhouse ranged from 788 W to 1
17 kW, while the amount of heat recovered from the heat storage unit was in the range of 189383 W. In this discharging period, the average daily heat requirement was 828 W, but the amount of heat recovered from the heat storage unit was only 305 W. In other words, 35
4% of the total heat requirement of the tunnel greenhouse was obtained from the heat storage unit. A comparison of the total heat requirement of the tunnel greenhouse with the heat recovered from the heat storage unit during the discharging periods, shows the difference between the total heat requirement and the heat recovered from the heat storage unit is important. The total heat requirement of the tunnel greenhouse was

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10

12

8 Rate of heat, kW

10 6

8

4

6 4

2

2

0

6

12

5

10

4

8

3

6

2

4

1

2

Supplied by the storage unit,%

Rate of heat, kW

14

0

0 22:30 23:00 23:30 00:00 00:30 01:00 01:30 02:00 Time, h

1.2

40

1.0 0.8

30

0.6

20

0.4 10

0.2 0

Supplied by the storage unit, %

50

1.4

Rate of heat, kW

4. Conclusions

7

(b)

the third discharging period was lower compared with the first and second discharging periods, because the air temperature outside the tunnel greenhouse during the third discharging period was higher compared with the first and second discharging periods. It is worth mentioning that the total heat requirement decreased by 90%, while the heat recovered from the heat storage unit decreased by 62
4% from the first discharging period to the third.

0 22:30 23:00 23:30 00:00 00:30 01:00 01:30 02:00 Time, h

(a)

0 23:00

(c)

Supplied by the storage unit,%

14

243

23:30 00:00 00:30 01:00 01:30 02:00 Time, h

Fig. 6. The changes of the total heat requirement of the tunnel greenhouse and the rate of heat recovered from the heat storage unit during the discharging periods: (a) first charging period; (b) second charging period; (c) third charging period: , recovered heat; , heat requirement; m, supplied by the heat storage unit

The packed-bed heat storage unit for greenhouse heating was analysed and energy and exergy efficiencies of the system were obtained. The effects of various factors on the net energy and exergy efficiencies were examined. The following conclusions are general in nature and are expected to be valid independent of location and weather. ð1Þ Greenhouse heating with solar energy The use of solar energy for greenhouse heating has gained an increasing acceptance during the last years. The main problem is related with the selection and sizing or the more appropriate, passive or active technology for the specific application. Active solar systems applied to greenhouses can supply a significant part of the heating requirements. However, there are some problems related to the cost of the heat collection unit, of the occupied land, the backup system, and the heat storage methods. A significant investment cost is necessary with active solar systems, especially metallic glazed collectors are used as the heat collection unit. An active solar system for greenhouse heating can be coupled with a heat pump in order to increase its efficiency. In addition, the storage volume and capacity of the heat storage material should be chosen carefully. Research and development on active solar systems for greenhouse heating has given a certain number of technical solutions aiming to overcome the above problems. ð2Þ Heat storage material for greenhouse heating

always higher than the heat recovered from the heat storage unit. The ratio of the total heat requirement to the heat recovered from the heat storage unit was the lowest during the first discharging period. This means that the total heat requirement of the tunnel greenhouse was the highest during the first discharging period. The total heat requirement of the tunnel greenhouse during

For intermediate temperature sensible heat storage, the commercially available liquids are generally expensive. Research directed toward improving the lifetime of sensible heat storage fluids could be useful. Solid materials are economically more attractive for hightemperature storage than the fluids and their volume requirements are nearly comparable. However, research

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Table 4 The values of the total heat requirement of the tunnel greenhouse and the rate of heat recovered from the heat storage unit for the discharging periods Discharging periods

1 2 3

Heat requirement of the tunnel greenhouse, W

Heat recovered from the heat storage unit, W

Heat supplied by the heat storage unit, %

Min.

Max.

Av.

Min.

Max.

Av.

Min.

Max.

Av.

7530 5940 788

8600 6040 1170

8250 5950 828

716 591 189

873 794 383

810 690 304

9
8 9
9 28
9

10
1 13
2 43
0

9
8 11
6 35
4

and development are needed to find heat transfer fluids that can be used in direct contact with the solids over long time periods. Development and low-cost containment and direct heat technology could significantly reduce the cost of the greenhouse heating systems with solar energy in the intermediate temperature range. Direct contact between the solid storage media and a heat transfer fluid is vital to minimise the cost of heat exchange in a sensible heat storage system. If the same heat exchanger is used for charging and discharging processes, considerable attention should be given to design the heat exchanger. In this case, technical limitations should be taken into account. Latent heat storage using phase change materials (PCM) in general provides much higher energy storage density than systems using sensible heat storage. High energy storage densities over a narrow temperature range make PCM attractive for greenhouse heating. There are problems with repeatable cycling, heat transfer rate and containment that need to be solved before latent heat storage systems can become commercially and economically viable for intermediate temperatures. Future studies should focus on: latent heat storage for greenhouse heating, and modelling the efficiency of the heat storage systems. Such experimental studies will be very useful to optimise the management of the heat storage systems. The economics and thermal efficiency of each new system must be carefully evaluated in a total system context.

ð3Þ Evaluation of the system efficiency The results of this study show that the difference between the results of energy and exergy analyses is significant. Since exergy is a measure of the quality of energy, exergy efficiency is more significant than energy efficiency, and that exergy analysis should be considered in the evaluation and comparison of the thermal energy storage systems. Exergy analysis clearly takes into account the loss of availability of heat in storage operations and, hence, it more correctly reflects the

thermodynamic and economic value of the storage operation. The optimisation of the design and exploitation of the thermal energy storage systems can be made by means of the exergy analysis. When optimising the thermodynamic efficiency of a thermal energy storage system, both design and operational parameters must be considered. The real purpose of a thermal energy storage system is not to store energy, but the store exergy. According to the results of the experiments, sensible heat energy storage systems are inherently inefficient devices in terms of the exergy efficiency. Exergy loss and auxiliary energy consumptions for the charging and discharging processes could be reduced to improve the exergy efficiency of the thermal energy storage systems for a particular application. The unit should be capable of receiving energy at the maximum rate without excessive driving forces. Analysis of thermal performance of systems costs of solar equipment and costs of auxiliary energy systems can be used to determine the optimum size of the system components for a particular application. Exergy analysis is essential to cost effective design and management of the thermal energy storage systems. Therefore, exergy analysis must be used to design thermal energy storage systems with the highest possible thermodynamic efficiencies. In conclusion, charging and discharging process of a thermal energy storage system must be analysed in order to optimise system efficiency.

References Arinze E A; Schoenau G J; Besant R W (1984). A dynamic thermal performance simulation model of an energy conserving greenhouse with thermal storage. ASAE Paper No. 84-2702 Bailey B J (1988) Principles of environmental controls. In: Energy Conservation and Renewable Energies for Greenhouse Heating (Zabeltitz C V, ed.), pp 17–41. FAO-Reur Technical Series, Vol. 3. Food and Agriculture Organization of the United Nations, Rome, Italy Baille A; Boulard T (1987). Phase change material for heat storage in greenhouse (France). In: Greenhouse Heating with Solar Energy (Zabeltitz C V, ed.), pp 139–142. FAO-

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Reur Technical Series, Vol. 1. Food and Agriculture Organization of the United Nations, Rome, Italy Balducci M (1985). Solar agricultural greenhouse. Flag Brochure, EEC, DG 17 Damrath J (1982). Solarenergienutzung im Gew.achshaus eine energetische Darstellung des doppeltbedachten Gew.achshauses mit solarunterstu. tzter Heinzung. [Energy conservation and use in solar greenhouses covered double layer.] Heft 14, Institut fu. r Technik in Gartenbau und Landwirtschaft Universit.at Hannover Fang Z; Larson D L; Fleischmen G (1995). Exergy analysis of a milk processing system. Transactions of the ASAE, 38(6), 1825–1832 Huang B K; Toksoy M; Cengel Y A (1986). Transient response of latent heat storage in greenhouse solar system. Solar Energy, 37(4), 279–292 Jaffrin A; Cadier P (1982). Latent heat storage applied to greenhouse. Solar Energy, 28(4), 313–321 Kern M; Aldrich A (1979). Phase change energy storage in a greenhouse solar heating system. ASAE Paper No. 79-4028 Krane R J (1987). A second law analysis of the optimum design and operation of thermal energy storage systems. International Journal of Heat Mass Transfer, 30(1), 43–57 Kurata K; Takakura T (1991). Underground storage of solar energy for greenhouse heating. I. Analysis of seasonal storage system by scale and numerical models. Transactions of ASAE, 34(2), 563–569 K.urkl.u A; Wheldon A; Hadley P (1995). Mathematical modelling of the thermal performance of a phase-change material (PCM) store: cooling cycle. Applied Thermal Engineering, 16(7), 613–623 Larson D L; Cortez L A B (1995). Exergy analysis: essential to effective energy management. Transactions of ASAE, 38(4), 1173–1178 . zt.urk H H (1997). Seraısıtma i-cin gu. ne-s enerjisinin faz O de&gi-stiren materyalde (PCM) depolanması u. zerine bir ara-stırma. [The research on storage of solar energy in phase

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change material (PCM) for greenhouse heating.] PhD Thesis, Department of Agricultural Machinery Institute of Natural and Applied Sciences, University of C - ukurova, Turkey . zt.urk H H; Ba-sc-etin-celik A (2002). Isı Depolama Tekni&gi. O [Thermal energy storage techniques.] The Union of Turkish Chambers of Agriculture, Publication Number 230, Ankara . zt.urk H H; Ba-sc-etin-celik A; Paksoy H O . ; Demirel Y (1999). O The research on storage of solar energy in phase change material (PCM) for greenhouse heating. Proceedings of ICAME ’99 ‘7th International Congress on Agricultural Mechanisation and Energy’, 26–27 May 1999, Adana, Turkey, pp 326–331 Puri V M; Zuritz C M (1985). Feasibility of subsurface latent heat storage for plant root zone and greenhouse heating. ASAE Paper No. 85-4045 Rosen M A; Hooper F C; Barbaris L N (1988). Exergy analysis for the evaluation of the performance of closed thermal energy storage systems. Transactions of the ASME, Journal of Solar Energy Engineering, 110, 255–261 Santamouris M I (1993). Active solar agricultural greenhouses. The state of art. Solar Energy, 14, 19–32 SIS (2001). Agricultural structure: production, price, value. Publication Number 2758, State Institute of Statistics, Prime Ministry Republic of Turkey, Ankara Ting K C; Giacomelli G A; Wu S W (1990). PCM energy storage for a CO2 enriched greenhouse. ASAE Paper No. 90-4040 . (1994). Potential of new and renewable energy . ltanır M O U sources in long-term utilisation for Turkish rural areas. Proceedings of AGENG ’94 ‘International Conference on Agricultural Engineering’, 29th August–1st September 1994, Milano, Italy, pp 822–828 Willits D H; Peet M M (1987). Factors affecting the performance of rockstorages as solar energy collection/ storage systems for greenhouses. Transactions of ASAE, 30(1), 221–232