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Investigation of a solar assisted heat pump wheat drying system with underground thermal energy storage tank Hatem Hasan Ismaeela, Recep Yumrutaşb, a b

T

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Bardarash Technical Institute, Duhok Polytechnic University, Duhok, Iraq Department of Mechanical Engineering, Gaziantep University, Gaziantep 27310, Turkey

A R T I C LE I N FO

A B S T R A C T

Keywords: Solar energy Solar energy storage Solar drying Heat pump drying Wheat drying System modeling

Storage of solar energy in underground Thermal Energy Storage (TES) tank during sunny days and extraction of the energy in the TES tank and its surrounding ground by a heat pump through the year for drying systems is an attractive subject for eﬀective use of solar energy and ground as heat sources. It is possible to store solar energy in the TES tank and ground to use it for heating systems. In this study, modeling of a wheat drying system with a heat pump and underground TES tank charged by solar energy is developed using energy analysis of the drying system to investigate its long-term performance parameters. The drying system has four components, which are wheat dryer, underground TES tank, solar collectors, and a heat pump. The analytical model consists of a solution of unsteady heat transfer problem around the TES tank and energy expressions for the other components. A MATLAB code is developed to examine the performance parameters hour by hour along the years. These parameters are water temperature in the tank, Coeﬃcient of Performances for the heat pump (COP) and system (COPs), Speciﬁc Moisture Evaporation Rate (SMER), and energy fractions for the system. When wheat mass ﬂow rate and Carnot eﬃciency are selected as 50 kg h−1 and 40% respectively, collector area, TES tank volume, COP, COPs and SMER are 70 m2 and 200 m3, 4.43, 4.3 and 6.05 respectively when the system attains periodic operation time from the 5th year onwards for 10 years of operation.

1. Introduction

conventional drying systems consume intensive energy in the industrial applications. For that reason, alternative and clean energy should be used because of increase in the unit price of fossil fuels, environmental eﬀects, and human health protection. Storage of solar energy is crucial in heating or drying systems when eﬃcient and cost-eﬀective solar energy utilization can be achieved. The great amount of energy demand for drying process can be decreased by the charging of solar energy into Thermal Energy Storage (TES) tank, and the stored energy being used via heat pump for drying of wheat in the drying unit. Using both of these renewable energy sources, which are solar and ground energy by Ground Coupled Heat Pump (GCHP), the performance of the drying systems can be increased. There are many publications in the literature on utilization of solar energy with TES tank by heat pump for heating and drying applications. Esen et al. (2007) compared a techno-economic analysis of GCHP with air-coupled heat pump (ACHP) system for space cooling in Elazig city, Turkey. They found the average COPsys of the GCHP system as 3.85 and 4.26 for horizontal ground heat exchanger at 1 and 2 m depths, respectively, and they obtained the ACHP system COPsys as 3.17. The results indicate that GCHP systems are economically preferable to ACHP systems. Wang and Qi (2008) analyzed the

The growing global population makes it diﬃcult to access aﬀordable, healthy, and normal wheat throughout the year. There are many reasons for this situation. The most important reason is not able to preserve wheat successfully at suitable conditions for many years. Therefore, it is necessary to store wheat by preserving moisture content within acceptable levels. Since, increase in moisture leads to storage problems, including insect and fungal infestation, germination, and respiration (Mrema et al., 2011). Stored temperature is another major factor that which aﬀects wheat storage. From a biological perspective, wheat grains at suitable temperature are active and respire during storage, and in order to eliminate these problems, wheat drying is required (Mrema et al., 2011). Decreasing moisture levels within stored crops is a requirement and is attainable via drying process for agricultural products and foods (Leon et al., 2002). Therefore, drying systems should be designed for drying fresh and cooked wheat in order to store them and have prolonged production. It is known that there are many types of drying applications, but, it is necessary to select a drying system that consumes low energy because

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Corresponding author. E-mail address: [email protected] (R. Yumrutaş).

https://doi.org/10.1016/j.solener.2020.02.022 Received 12 October 2019; Received in revised form 23 January 2020; Accepted 5 February 2020 0038-092X/ © 2020 International Solar Energy Society. Published by Elsevier Ltd. All rights reserved.

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Nomenclature A Ai Ao Af Adesup Acond Ac Acol Afr c Cpa cw C pwheat Di Do Dh f G h hg ha

h p4 h p5 hi ho

IT jH ka kw kf k KL l ṁ a ṁ w ṁ w4 ṁ w5

ṁ p ṁ

Ṁ Mp MC4 MC5 NuD Nua P Pr Pra PrL q Q Q̇H ̇ Qevap ̇ Qu qu

qH r R ri ro ReD

tank surface area (m2) inside heat transfer area for condenser (m2) outside heat transfer area for condenser (m2) ﬁn area (m2) desuperheated heat transfer area for condenser (m2) condensation heat transfer area for condenser (m2) free ﬂow area (m2) collector area (m2) front area (m2) speciﬁc heat of earth (kJ kg−1 K−1) speciﬁc heat of air (kJ kg−1 K−1) speciﬁc heat of water (kJ kg−1 K−1) speciﬁc heat of wheat (kJ kg−1 K−1) inside tube diameter for condenser (m) outside tube diameter for condenser (m) hydraulic diameter (m) friction factor mass velocity for air across condenser (kg m−2 s−1) speciﬁc enthalpy of the moist air (kj kg−1) speciﬁc enthalpy of saturated water vapor (kj kg−1) speciﬁc enthalpy of dry air at the mixture temperature (kj kg−1) speciﬁc enthalpy of dry wheat pre drying (kj kg−1) speciﬁc enthalpy of dry wheat after drying (kj kg−1) heat transfer coeﬃcient of refrigerant side for condenser (W m2 K−1) heat transfer coeﬃcients of air side for condenser (W m2 K−1) hourly solar radiation on tilted surface (W m−2) Colburn factor [-] thermal conductivity for air (W m−1 K−1) thermal conductivity of tube wall for condenser (W m−1 K−1) ﬁn conductivity (W m−1 K−1) thermal conductivity of earth (W m−1 K−1) thermal conductivity of saturated liquid of refrigerant (W m−1 K−1) length of ﬁn (m) mass ﬂow rate of dry air (kg s−1) amount of evaporated water from moist wheat (kg s−1) mass ﬂow rate of water in the product pre-drying (kg s−1) mass ﬂow rate of water in the product after drying (kg s−1) mass ﬂow rate of dry wheat (kg s−1) equivalent mass velocity for two phase of refrigerant (kg m−2) mass ﬂow rate of the refrigerant (kg s−1) moisture content of wheat on a dry basis (kg water kg−1 solid) moisture content pre drying (%) moisture content after drying (%) Nusselt number for refrigerant in superheated region [-] Nusselt number for air [-] dimensionless [-] Prandtl number for refrigerant in desuperheated region [-] Prandtl number for air [-] Prandtl number for saturated liquid of refrigerant [-] dimensionless heat transfer to the tank [-] heat transfer to the tank (W) condenser heating load (W) latent heat of evaporation water from moist wheat (W) solar energy rate (W) dimensionless useful solar energy charge rate to the tank [-]

Rea Ree rh t T Ta To Tmo T2 T3 T4 T5 Tw Tc T∞ u (UA)he Uo V Vȧ Vmax Vfr w Ẇc Ẇf W4 W5 xv

dimensionless condenser heat load [-] radial distance from the tank center (m) tank radius (m) inside tube radius for condenser tubes (m) outside tube radius for condenser tubes (m) Reynolds number for refrigerant in desuperheated region [-] Reynolds number for air [-] equivalent Reynolds number for two-phase ﬂow of refrigerant [-] hydraulic radius for condenser (m) time (s) earth temperature (K) ambient air temperature (K) outside design air temperature (K) dry bulb temperature of moist air (K) air temperature pre drying (K) air temperature after drying (K) wheat temperature pre drying (K) wheat temperature after drying (K) water temperature in the tank (K) average temperature of air pre and after condenser (K) deep ground temperature (K) dimensionless parameter for design condition [-] product of heat transfer coeﬃcient and area for heat pump condenser (W K−1) overall heat transfer coeﬃcient based on outside area of the condenser (Wm−2 K−1) volume of the tank (m3) volume ﬂow rate for air (m3 kg−1) maximum velocity of air in free ﬂow area (m s−1) front velocity of air in frontal area of condenser (m s−1) dimensionless compressor work [-] compressor input power (W) fan input power (W) weight of wheat pre-drying (kg h−1) weight of wheat after drying (kg h−1) vapor quality for refrigerant [-]

Greek letters

α ηc ηo

ηf ηfc ϕ ϕa ϕc ϕw ϕ2 γ ρa ρ ρw ρL ρG τ Δp ΔTln σ μa f 539

thermal diﬀusivity of earth (m2 s−1) Carnot eﬃciency [-] overall surface eﬀectiveness (air side only) for the condenser [-] ﬁn eﬃciency [-] ﬂat plate solar collector eﬃciency [-] dimensionless temperature for earth [-] dimensionless ambient air temperature [-] dimensionless average temperature for air pre and after condenser [-] dimensionless water temperature in the tank [-] dimensionless air supply temperature pre-dryer [-] dimensionless parameter [-] air density in air side of condenser (kg m−3) density of refrigerant in desuperheated region (kg m−3) water density (kg m−3) density of saturated liquid for refrigerant (kg m−3) density of saturated vapor for refrigerant (kg m−3) dimensionless time [-] pressure drop of air across condenser (Pa) log mean temperature diﬀerence (K) free ﬂow area/frontal area for condenser (Ac Afr−1) air viscosity (Pa. s) friction factor

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fcf ν1 ν2 vm ω ω2

ω3 δ

parameter estimated from a ﬁgure [-] speciﬁc volume for air pre-condenser (m3 kg−1) speciﬁc volume for air after condenser (m3 kg−1) average speciﬁc volume for air (m3 kg−1) humidity ratio for moist air (kg water kg−1 dry air) humidity ratio for moist air pre drying process (kg water kg−1 dry air) humidity ratio for moist air after drying process (kg water kg−1 dry air) ﬁn thickness (m)

Acronyms COP COPs GSHP HPD SAHP SMER TES

coeﬃcient of performance of heat pump coeﬃcient of performance of whole system Ground Source Heat Pump heat pump drying Solar Assisted Heat Pump speciﬁc moisture evaporation rate (kg kW−1 h−1) thermal energy storage

exchanger (GHE) in horizontal and vertical conﬁgurations for solarassisted ground source heat pump using artiﬁcial neural network and adaptive neuro-fuzzy inference systems. They calculated the COP and COPs as 2.88 and 3.55 for horizontal GHE and 2.34 and 2.91 for the vertical GHE type, respectively.. In literature, there are many types of drying systems used for drying of food or agricultural products. A comparative review of Heat Pump Dryers (HPD) with other commonly used dryers such as hot air, vacuum, and freeze drying gives a clear indication of the validity, preference, and high eﬃciency of HPD used among all types (Mujumdar and Jangam, 2011). HPD can be classiﬁed into four main types; air source, chemical, ground source, and hybrid source heat pumps (Wongsuwan et al., 2001; Daghigh et al. 2010). Drying systems that are more eﬃcient in drying than conventional ones should be used to save energy and decrease pollution. SAHP drying systems are more eﬃcient than the conventional ones. When solar energy and heat pump are used together for food drying, the performance of the system will be higher than the conventional one. However, previous studies indicate that there are inadequate studies related to SAHP drying systems and ground source HPD systems (Kivevele & Huan, 2014). Mohanraj and Chandrasekar (2009) investigated chili drying via an experimental study on indirect forced convection solar dryer integrated with diﬀerent sensible heat storage materials. Performance of a forced convection mixed mode solar dryer with TES was experimentally analyzed by (Baniasadi et al., 2017). During the experiments, they found that fresh apricot slices can be dried at diﬀerent working conditions. Another experimental study was conducted by (Şevik et al., 2013) using SAHP drying system with ﬂat plate collectors for drying of mushroom. System COP was calculated as values in a range from 2.1 to 3.1 using experiment results. They also found that SMER values ranged between 0.26 and 0.92 kg kW−1h−1. Another experimental analysis using heat pump assisted solar air collector was carried out by (Şevik, 2014) where four diﬀerent vegetables were dried under diﬀerent climatic conditions. He obtained COP of the drying system between 1.96 and 2.17, and SMER values between 0.03 kg kW−1h−1and 0.46 kg kW−1h−1 for the vegetables. Rahman et al. (2013) studied economic optimization of evaporator and air collector area of a solar-assisted heat pump drying system. They revealed that this drying system has suﬃcient amount of savings during the life cycle with a minimum payback period of about 4 years. Şevik (2013) designed and manufactured a new type of solarheat pump dryer by using double-pass solar air collectors, a heat pump and a photovoltaic unit. He indicated that thermal eﬃciency of doublepass collector ranged between 60% and 78% based on experimental results. He observed that this system can be conveniently run through skipping the heat pump under normal ambient air conditions. Aktas et al. (2017) compared experimental results of a heat pump dryer and an infrared assisted-heat pump dryer to determine the energy and exergy eﬃciencies of the dryers. They indicated several variations, including energy eﬃciency (5.3–50%), COP of the system (2.11–2.96), and exergy eﬃciency (31.6% − 66.8%). Qiu et al. (2016) performed an experimental study to obtain performance and operation mode analyses of a heat recovery and thermal storage solar-assisted heat pump drying system. They have shown that coeﬃcient of performance of the drying

performance of a TES tank in a solar-assisted ground heat pump system for residential buildings. They indicated that the performance of such storage relies heavily on the intensity of solar radiation and the matching between the volume of water tank and solar collector’s area. They observed that the range of 20–40 L/m2 is a reasonable ratio between the volume of the tank and collector area. Yumrutas et al. (2003) used analytical and computational models to predict yearly periodic performance of solar-assisted GCHP for indoor heating with a cylindrical tank. They indicated that the soil type has little eﬀect on yearly energy ratios but has bigger impacts on the storing temperature and yearly heat pump COP. Esen, et al. (2005) performed an experimental study of the eﬀects of various refrigerants on the thermal performance of a two-phase thermosyphon solar collector under various environmental and load conditions. They indicated that the experimental results showed a good agreement with the results in the literature. Yumrutas and Kaska (2004) performed an experimental study through exploiting Solar Assisted Heat Pump (SAHP) room heating system with daily storage of solar energy. They indicated that yearly COP ranged between 2.5 and 3. Inallı (1998) investigated a theoretical study of a SAHP system with spherical TES tank. He solved an unsteady heat transfer problem around the TES tank by using a ﬁnite diﬀerence method and CFFT technique, taking into account the yearly periodic functioning. He indicated that the heat load, collector area, and tank volume were the extreme signiﬁcant inﬂuencing factors on the SAHP system. Chung et al. (1998) adopted TRNSYS for forecasting thermal performance and economic feasibility for a central solar heating system with seasonal storage. They found that the total heat requirement can be supplied by 39% solar energy using 184 m2 as collector area and 600 m3 of storage volume. Wang and Qi (2008) examined the functioning of a SAHP system with TES tank for a house heating unit. It was found that the eﬃciency of the TES based on total solar radiation and absorbed solar energy by collectors may contribute to 40% and 70%, respectively. Wang et al. (2010) experimentally investigated a SAHP system with seasonal TES tank used in housing units in Harbin. They concluded that solar collectors contributed 49.7% of total heating output in the heating system, and the average COP and COPs were 4.39 and 6.55, respectively. It is seen from the results that COP’s are reasonably high. Yumrutas and Unsal (2012) developed an analytical model for a SAHP system with TES tank for space heating. They found that this type of heating system can reach periodic situation in 5–7 years. They also found that temperature of water in the TES tank has a proportional relation with collector area and tank volume. Kuyumcu et al. (2016) investigated the utilization of waste heat rejected from a chiller used for ice rink cooling application and storing the gained energy in a spherical TES tank, which was used to heat a swimming pool. They indicated that ﬁve to seven years are necessary for the swimming pool heating system to attain periodic operation. The results also reveal that the performance of heating system reaches optimum at an ice rink surface area of 475 m2 with swimming pool 625 m2. Tutumlu et al. (2017) investigated thermal performance of an ice rink cooling a system with underground TES tank. They determined that ﬁve to seven years are suﬃcient for the cooling system to attain periodic operation. Esen et al. (2017) established a ground heat 540

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fractions for solar drying without using heat pump, and SAHP dryer were 3.38, 0.38 kg/kW h, 66.7%, and 44.6% respectively. This study was not performed on a monthly, seasonal, yearly, or continuous basis. Conventional high energy consuming drying systems have been used for decades in industrial applications. However, although renewable energy sources are partially used, they have been used suﬃciently. Because, it is seen from the literature review that there has not been any theoretical or experimental study related with SAHP drying system with underground TES tank up to now, and also there is no yearly and continually operated drying systems. Underground TES application can be used to store solar energy acquired from solar collectors. The utilization of the stored solar energy in the TES tank and its surrounding ground by the heat pump in drying applications should be considered a priority due to ﬂuctuations of fossil fuels’ prices, environmental impacts, and anticipated depletion of conventional fossil fuels. Application of the SAHP drying system by storing solar energy in the underground TES tank and its surrounding ground during the whole year in the form of thermal energy will maximize system performance beyond other types of systems. Due to the multiple advantages of the system under question, in this study, we propose modeling of a wheat drying system operating with ground coupled heat pump and an underground TES tank charged by solar energy in order to ﬁnd performance parameters for the drying system. The drying system comprises of a drying unit, ﬂat plate solar collectors, an underground TES tank, and a ground coupled heat pump. All parts of the drying system are thermally analyzed, and they are combined and linked with each other to obtain a mathematical model for the new drying system. A mathematical model has been developed and incorporated into MATLAB

system ranged from 3.21 to 3.49, and that payback periods for drying radish, pepper, and mushroom in the life span of the system were 6 years, 4 years, and 2 years, respectively. Moreover, during the operation of air-source heat pump in low temperatures, evaporator’s surface temperature may drop to under 0 °C and lead to frost formation and accumulation, which might decrease performance (Song et al., 2018). It also negatively aﬀects air source heat pump. Naemsai et al. (2019) inspected the performance of SAHP drying with heat restoring to minimize thermal energy required to dry chili peppers. The study has shown that the SAHP drying with restoration of heat led to a better drying performance than conventional dying approaches. The dryer system may oﬀer COP, drying time, eﬃciency, and speciﬁc energy consumption of roughly 3.17, 24 h, 33.2%, and 2.21 kWh kg−1, respectively. Wang et al. (2019) experimentally investigated SAHP drying system for mango in which a new drying system with a secondary heat restoring has been developed, with the ability of operating autonomously in solar and heat pump drying method, or operating in a combined mode. A heat pump evaporator and a heat exchanger have been embraced to restore the heat lost from the exhaust moist air from the drying chamber to maximize energy eﬃciency. The eﬃciency of solarassisted drying mode proved 6% greater than the heat pump drying mode, and it saved 3.5kWh. COP’s were 3.69 and 3.48 for SAHP mode and heat pump drying mode, respectively. The typical eﬃciency of a heat exchanger for restoring lost heat may reach 41.7% throughout the entire operation. Yahya et al. (2016) studied the comparison of solar dryer with and without using a heat pump for cassava chips drying. They indicated that the average COP of the heat pump, SMER, solar

Fig. 1. Schematic of the drying system with underground TES tank. 541

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H. Hasan Ismaeel and R. Yumrutaş

as a storage medium in the tank and is buried under the ground. It stores solar energy by circulating water between the collectors and the tank during all seasons, and it gives the stored energy to the drying unit during the whole year by using the heat pump. Bigger volume is necessary for long term storage of solar energy, especially to get beneﬁt from it in winter. Therefore, large scale TES tank is used to collect more solar heat needed for wheat drying process. 3. Modeling of the drying system Energy expressions for each component of the system should be deﬁned to ﬁnd a mathematical modeling of the drying system. The models for the drying system components are related to the drying unit, ground coupled heat pump, underground TES tank, and solar collectors. Mathematical modeling is obtained by combining these expressions. In this section, the expressions will be given in detail.

Fig. 2. Schematic of the dryer unit.

3.1. Energy demand of the dryer unit

program, and then it was executed by using diﬀerent input data to ﬁnd performance parameters for the drying system hour by hour during 24 h along many years. These parameters include temperature of water in the TES tank, Coeﬃcient of performance of the heat pump (COP) and system (COPs), Speciﬁc Moisture Evaporation Rate (SMER), and energy fractions. Results of this study will give an idea to applicants of the SAHP drying system with underground TES tank.

A dryer is employed to evaporate moisture from cooked wheat, which is shown in Fig. 2. There are four main interactions in the drying unit; inﬂux of drying air from condenser to drying chamber, exit of moist air after evaporation of moisture from the unit, input of cooked wheat from top of the dryer, and output of dried wheat with moisture lowered to the desired level. Mass and energy balances may be formulated for the dryer unit by considering it as a control volume. Mass balances for the dryer unit may be expressed for product or wheat, dry air, and water as: ṁ p4 = ṁ p5 = ṁ p for the product and ṁ a2 = ṁ a3 = ṁ a for the air

2. Description of the drying system The wheat drying system under study is an environment-friendly drying system since it uses a great amount of its energy from solar energy charged in the TES tank. The scheme of the drying system is shown in Fig. 1. The system consists of four main parts, which are wheat dryer unit, heat pump, ﬂat plate solar collectors, and underground TES tank. The system is considered to be installed in Gaziantep (37.1 oN), Turkey. The most important part of the drying system is the dryer unit since drying of the wheat or evaporating of water in the wheat is performed in it. It is considered to be cylindrical in shape. Wet product enters the dryer unit from the top and moves slowly towards the bottom. At the same time, fresh air comes to a heating channel and is heated by the condenser. The hot air is pumped by a fan from the bottom of the dryer unit to top of the unit by contacting the product. During this process, water in the product is evaporated and exits with the air. The leaving hot air from the condenser passes over moist wheat in the drying chamber to evaporate undesired water available in the wheat, and transfers it to the environment. Heat pump is another essential part of the drying system. It works as a heat transformer from the TES tank to the dryer unit. It is known that heat pumps are energy eﬃcient devices, and Ground Source Heat Pumps (GSHP) are much more eﬃcient than air source heat pumps. Therefore, GSHP is used in the drying system. Utilizing this type of heat pump in the drying system can supply much more energy to drying air than conventional drying air systems since the GSHP coupled with the TES tank supplies higher energy for less temperature ﬂuctuations. Heat is absorbed from the water in the tank by a refrigerant ﬂowing through the evaporator of the heat pump cycle. The energy extracted and transferred by the compressor is rejected from the condenser of the heat pump to the drying air. The heat pump operates in order to correspond heating requirement of the wheat and evaporate the water present in the wheat throughout the year. Flat plate solar collectors are used in the drying system. They charge solar energy to the tank throughout the entire year. They work when exposed to solar energy. Otherwise, water in the cycle between collectors and the TES tank will not be circulated. The TES tank is the heart of the drying system. Spherical tank is considered to store solar energy as a sensible heat in the water present

ω2 ṁ a + ṁ w4 = ω3 ṁ a + ṁ w5 forthewater

(1)

or Eq. (1) can be rearranged as follows:

ṁ w = ṁ w4 − ṁ w5 = ṁ a (ω3 − ω2)

(2)

The dryer unit is assumed to be insulated. Under this assumption, an energy balance may be expressed for the drying chamber as follows:

ṁ a h2 + ṁ p h p4 + ṁ w4 h w4 = ṁ a h3 + ṁ p h p5 + ṁ w5 hw5

(3)

where ṁ w , ṁ w4 , ṁ w5 , and ω are mass ﬂow rate for evaporated water from the product, mass ﬂow rate of water in the product at state 4, mass ﬂow rate of water in the product at state 5, and humidity ratio respectively. The enthalpy of the moist air is equal to the total individual partial enthalpies of the dry air and water vapor. Thus, the moist air enthalpy may be formulated (ASHRAE, 1997) as:

h = ha + ωhg

(4)

where hg refers to the speciﬁc enthalpy of saturated water vapor, and ha refers to the speciﬁc enthalpy of dry air at the mixture temperature. The speciﬁc enthalpy of the dry air is:

ha = Cp a Tmo

(5)

The enthalpy of the moist air may be given as the following (AlvesFilho, 2015).

h = 1.006Tmo + (2501 + 1.86Tmo ) ω

(6)

where Tmo is dry-bulb temperature of the moist air. The speciﬁc enthalpy of the product or wheat in the energy balance may be written as below:

(h p5 − h p4 ) = C pwheat (T5 − T4 )

(7)

Therefore Eq. (3) can be written as the following expression:

̇ ṁ a (h2 − h3) = ṁ p C pwheat (T5 − T4 ) + Qevap Heat transfer caused by phase change is: 542

(8)

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̇ Qevap = ṁ w hfg

where Af is ﬁn area, and η f is ﬁn eﬃciency. To evaluate eﬃciency for a rectangular ﬁn, we can use the following equations:

(9)

where hfg refers to latent heat of vaporization of water at the average temperature, Tav = (T4 + T5)/2 , ṁ w is mass ﬂow rate for evaporated water from moist wheat, ṁ p is mass ﬂow rate for dry wheat, and T4 and T5 are wheat temperatures before and after drying process. Wheat kernel is supposed to be spherical with a typical diameter of 3.66 mm, a density of 1215 kg/m3, and a speciﬁc heat oﬀered by (Kazarian and Hall, 1962).

MP ⎞ Cp wheat = 1398.3 + 4090.2 ⎛ ⎝ 1 + MP ⎠ ⎜

⎟

(

W4 − W5 W5

(10)

). W refers to pre-drying weight, and 4

W5 refers to post-drying weight. Another equation is used to ﬁnd the weight of the material after the drying process as follows (Mrema et al., 2011): (MC4 − MC5) ⎤ W5 = W4 − ⎡W4 ⎢ 100 − MC5 ⎥ ⎦ ⎣

2ho kf δ

(19)

( ) ∙ (Re − 1000) ∙Pr = 1 + 12.7 ( ) ∙ (Pr − 1) f 8

D

f 0.5 8

2 3

(20)

f = (0.79lnReD − 1.64)−2

(21)

10 4

< ReD < 5 × 106 . Equation (20) is valid for 0.5 < Pr < 2000 and There are many correlations in literature that can be used to ﬁnd the condensation heat transfer coeﬃcient. The correlation obtained by (Akers et al. 1958) is used in the present study to ﬁnd the condensation heat transfer coeﬃcient in the tubes. They assumed complete condensation and used modiﬁed Reynolds number Rem . Local convection coeﬃcient was given by (Akers et al. 1958):

3.2. Power requirement of the fan Fresh air is ﬁrstly taken from atmosphere by a fan and is heated by the condenser of the heat pump, then sent to the dryer unit to evaporate water from the moist wheat. The fan is necessary for this operation. Pressure loss should be ﬁrstly calculated to ﬁnd the energy requirement of the fan. For that reason, pressure loss of air ﬂowing over an aircooled condenser is calculated. In the condenser, there may be heat transfer with three phases, which are desuperheating, condensing, and subcooling (ASHRAE, 2000). Total heat transfer area is the sum of heat transfer areas of the condenser:

hi (x ) Di 1 = C Remn Pr L 3 KL

(22)

The equivalent Reynolds number for two-phase ﬂow Ree is determined by using the following equations given in (Akers et al. 1958):

(12)

Q̇H Uo ΔTln

m=

where f is the friction factor and is calculated for smooth tubes by using the following relation (Icropera and De Witt, 2006):

where MC4 andMC5 are moisture contents of the product at states 4 and 5, respectively.

Ao =

(18)

NuD

(11)

Q̇H = Uo A oΔTln

tanhml ml

where l , kf , δ are length, thermal conductivity, and thickness of the ﬁn respectively. Equation (15) may be applied to air and refrigerant sides to ﬁnd Uo value for the condenser (Cavallaro and Bullard, 1994). Accordingly, the outside area for the condenser is calculated at minimum temperature of Gaziantep city. The correlation obtained by (Icropera and De Witt, 2006) is used for single-phase in the superheated and subcooled regions.

where, Mp refers to moisture content of material on a dry basis (kg water/kg solid), and MP =

ηf =

1

1

⎛ ρ 2⎞ ρ 2⎞ ṁ D ⎛ Rem = Ree ⎜1 + ⎜⎛ L ⎟⎞ ⎟ = e i ⎜1 + ⎜⎛ L ⎟⎞ ⎟ μL ⎝ ρG ⎠ ⎠ ⎝ ρG ⎠ ⎠ ⎝ ⎝

(13)

(23)

1

⎡ ρ 2⎤ ṁ e = ṁ ⎢ (1 − x v ) + x v ⎜⎛ L ⎟⎞ ⎥ ρ ⎢ ⎝ G⎠ ⎥ ⎣ ⎦

where Uo is the overall heat transfer coeﬃcient based on outside area of the condenser. To ﬁnd the heat transfer area of the air cooled condenser, it is necessary to ﬁnd the overall heat transfer coeﬃcients for the refrigerant and air sides. The heat exchanger geometry is given in Table 1. There are two surface areas for the air-cooled condenser: desuperheated and condensation area.

Ao = Adesup + Acond

ṁ =

( ) o

( )+ 1 ( ) hη

Diln 2k w

Ai Ao

+

o o

C = 0.0265, and n = 0.8 for Ree > 50, 000

C = 5.03, and n =

(15)

⎜

⎟

(16)

Af ηo = ⎛1 − (1 − ηf ) ⎞ A o⎠ ⎝ ⎜

1 for Ree < 50, 000 3

Table 1 Input design parameters for compact heat exchanger used as a condenser (Kays and London, 1984).

where Di , Do , ri , ro , hi , ho , k w and ηo are the inner and outer diameters, inner and outer radii of tubes, heat transfer coeﬃcients of air and refrigerant, thermal conductivity of tubes wall, and overall surface effectiveness of air side respectively.

Af ⎞ Ai D = i ⎛1 − Ao Do ⎝ Ao ⎠

(25)

4

The overall heat transfer coeﬃcient, Uo, is a function of air side, tube conductance, and refrigerant side resistances (Kays and London, 1984):

1 1 = A Uo hi A i

Ṁ πDi2

where Ṁis the mass ﬂow rate of the working ﬂuid. The empirical parameters C and n are:

(14)

ro ri

(24)

The equivalent mass velocity is given as:

⎟

(17) 543

Parameter

Value

Tube outside diameter, Do Fin pitch ﬂow passage hydraulic diameter, Dh Fin thickness, t Free ﬂow area/frontal area, σ Heat transfer area/total volume, α Fin area/total area, Af/A

17.17 (mm) 305 (1/m) 3.48 (mm) 0.4064 (mm) 0.481 554 m2/m3 0.95

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H. Hasan Ismaeel and R. Yumrutaş

Q̇H (t ) = (UA)he [Th (t ) − Tc ]

The heat transfer coeﬃcient for air can be evaluated as in the following:

(42)

Q̇H = ṁ a C pa ΔT

(26)

ṁ a = Vȧ ρa

(27)

where Th (t) is condenser temperature, (UA) he is the product of overall heat transfer coeﬃcient and condenser surface area, and Tc represents the average temperature of air entering and leaving the condenser. The heat absorbed by the drying air is given as;

ΔT = T2 − T1 (t )

(28)

Q̇H (t ) = ṁ a Cp a [T2 − Ta (t )]

Vmax Ac ρa = Vfr Afr ρa

(29)

The heated air is used to dry the moist wheat or to evaporate the water present in the wheat. After heating the air, it is ﬂowed through the dryer unit. Therefore, the energy requirement of the dryer unit can be expressed as:

σ=

Ac Afr

Vmax =

(30)

Vfr

Q̇H (t ) = ṁ p C pwheat (T5 − T4 ) + ṁ w hfg

(31)

σ

ṁ a = ρa Vmax Ac

Rea =

(32)

Vmax ρa Dh G Dh = μa μa

(33)

Pra =

Th (t) ⎞ COP = ηc ⎛ ⎝ Th (t) − Tw (t) ⎠ ⎜

h D Nua = o h ka

(34)

(35)

Nua = 0.117

1 Rea0.65 Pra 3

u (ϕc + 1)+ϕ2 − ϕa (t ) ⎞ COP = ηc ⎛⎜ ⎟ uϕ ⎝ a +ϕ2 − ϕa (t ) − uϕw (t ) ⎠

(36)

where Rea , Pra , Nua are Reynolds, Prandtl, and Nusselt numbers of the air respectively, Dh is ﬂow passage hydraulic diameter, and G is mass velocity of air. By neglecting entrance and exit losses, the pressure drop across the condenser for the air side is computed (Kays and London, 1984),

ΔP = G 2

v1 ⎡ v Ao vm ⎤ (1 + σ 2) ⎛ 2 − 1⎞ + fcf 2⎢ v A c v1 ⎥ 1 ⎝ ⎠ ⎦ ⎣ ⎜

u= (37)

a

St =

(38)

jH 2 Pr 3

Nua ho = Pra Rea G Cp a

(UA) he T − Ta (t) = 2 ṁ a Cpa Th (t) − Tc

Wc (t) = (39)

(47)

Q̇H (t) COP

(48)

When Eqs. (46) and (43) are inserted into Eq. (48), the dimensionless work for the compressor, w, willbederivedas below:

(40)

w(τ ) =

where v1 and v2 are speciﬁc volumes of air before and after drying, respectively, vm is average speciﬁc volume for air, jH and St are Calburn factor, and Stanton number respectively. jH and fcf are estimated from a ﬁgure adopted by (Kays and London, 1984). fcf is used to ﬁnd the air side pressure drop for the compact tube-ﬁn heat exchanger. The hourly power input to the fan for the air cooled condenser can be calculated by:

Ẇ fan (t ) = ΔP (t ) Vȧ (t )

(46)

Using the deﬁnition of COP for the heat pump, the work requirement of the compressor for the heat pump is known as:

2

St =

(45)

where u in Eq. (46) is a parameter obtained when equating heat transfer Eq. (42) with (43), ϕc , ϕa , ϕ2 and ϕw are dimensionless average air temperatures before and after condenser, ambient air temperature, air temperature pre-dryer, and water temperature in the TES tank respectively.

⎟

jH = St Pra3

⎟

Then, Eqs. (42) and (43) can be combined and solved for Th . The result of Th can be inserted in Eq. (45) to eventually give:

Cp a μa ka

(44)

Coeﬃcient of Performance of heat pump (COP) and system (COPs), work consumption of the compressor, and Speciﬁc Moisture Evaporation Rate (SMER) are very important for the heat pump drying system as they may serve as indicators for system performance. Therefore, expressions related with these parameters should be obtained. COP of the heat pump can be expressed using the approach given in (Yumrutas and Unsal, 2000; Tarnawski, 1989) as a function of source (Tw) and sink (Th) temperatures. Its derivation is presented in (Yumrutas and Unsal, 2012) and given as:

where Vfr is front velocity, and should be selected carefully, Vȧ is volumetric ﬂow rate for air, T2 is the air temperature before drying, Ac and Afr are free ﬂow area and frontal area for the condenser, respectively. Vmax is the maximum velocity in free ﬂow area.

G=

(43)

[ϕ2 − ϕa (τ )][uϕc +ϕ2 − ϕa (τ ) − uϕw (τ )] ηc [u(ϕc + 1)+ϕ2 − ϕa (τ )]

(49)

Coeﬃcient of performance for the system can be expressed as;

COPs =

Q̇H (t) Wc (t) + Wf (t)

(50)

For a drying system, the most suitable eﬃciency parameter is the SMER (ASHRAE, 1999) which is deﬁned as the energy needed to take away 1 kg of water (Alves-Filho, 2015). So, SMER is recorded as the heat pump dryer eﬃciency.

(41)

3.3. Energy demand of the heat pump

SMER = Heat pumps are devices that extract heat from a source at low temperature, and transfer it to a hotter sink (ASHRAE, 1997). In the drying system, ground source heat pump gains heat via evaporator from the tank and gives it via condenser to the drying air. The given heat by heat pump condenser to the drying air is represented as a function of condenser temperature;

ṁ w ṁ w = Ein Ẇc (t) + Ẇf

(51)

where Ein , Wċ (t ), andẆ f are the total energy, hourly compressor work, and hourly fan power input to the ground heat pump drying system after neglecting pumps input power in this study. By combining Eqs. (50) and (51), SMER can also be deﬁned as a function of COP of the system and heating requirement of the drier: 544

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H. Hasan Ismaeel and R. Yumrutaş

SMER =

Qu̇ (t ) = ηfc (t ) Acol IT (t )

ṁ w COPs Q̇H (t)

(52)

where ηfc is the ﬂat plate solar collector eﬃciency, Acol is the collector area, and IT(t) is the hourly radiation on the solar collector per unit area. The collector eﬃciency variation regarding dimensionless factor, [Tw (t) − Ta (t)]/IT (t) was obtained from the experimental study by (Yumrutas and Kaşka, 2004), and it is also used in the present study;

3.4. Solution of TES tank problem Storage tank is the heart of the drying system since it is used to store solar energy. This energy is given to drying air via heat pump in all seasons. At the same time, ground around the tank is used as a storage medium. Therefore, unsteady heat transfer problem around the tank should be solved to perform energy analysis for the interaction between both of solar collectors and heat pump with TES tank. The TES tank considered in this study is spherical and located deep in earth. The tank is ﬁlled with fully mixed water. As a result, its temperature will change with time, Tw (t) . Water temperature is initially assumed to be equal to the deep underground temperature, T∞. It is assumed that the earth around the tank has a homogeneous structure and constant thermal properties. Fig. 3 demonstrates schematic representation and energy balance of the spherical TES tank. Unsteady heat transfer problem around the tank is given by a differential equation with the following initial and boundary conditions in the spherical coordinate system:

∂ 2T 2 ∂T 1 ∂T + = ∂r 2 r ∂r α ∂t

(52)

T (R, t ) = Tw (t )

(53)

T (∞, t ) = T∞

(54)

T (r , 0) = T∞

(55)

ηfc (t ) = 0.72 − 6.4 ⎡ ⎢ ⎣

dTw ∂T − kA (R, t ) dt ∂r

IT (t) = R I(t)

p

q (τn ) + ⎡ Δτ + ⎣

1 π Δτ

A computer code in MATLAB has been developed for numerical calculations. In this section, main input data for ﬁnding performance parameters for the drying system are outlined. The input data are for dryer unit, heat pump, TES tank, ground around the tank, and solar collectors. The data and information for the main components of the drying system are given in the following subsections. 4.1. Drying unit Energy requirement of the dryer unit is necessary throughout all seasons on hourly basis. Mass ﬂow rate of the cooked wheat is selected basically as 50 kg h−1, which can be changed to ﬁnd its eﬀect on performance of the drying system. Air temperatures at the inlet state to the drier unit are recommended between 55 and 65 °C (Sundaram et al., 2016; KDC, 2019). In the present study, the inlet air temperature to dryer unit, T2 is assumed to be 60 °C. Outside design air temperature for Gaziantep, To, is equal to −9°C, respectively. For design conditions, “u” value in Eq. (47) is 1.7. Ambient air temperature, Ta(t), is heated up by the heat pump condenser to reach design temperature (60 °C), and is blown to the dryer unit by a fan to evaporate water existing in the wheat. The cooked wheat temperature, T4, and drying wheat temperature, T5, are taken as 25 °C and 50 °C, respectively. Moisture contents of moist wheat, MC4, and drying wheat, MC5, have been considered as 30% and 0% respectively, according to (Sundaram et al., 2016). Hourly humidity ratio has been evaluated by using EES (Engineering Equation Solver) program depending on hourly ambient air temperature for Gaziantep, at atmospheric pressure, while relative humidity of 30% is taken in this study based on (Sundaram et al., 2016; Magnussen and Strommen, 1981). Pressure drop through the drying

n − 2 ϕ w (τi + 1) − φw (τi ) π Δτ (n − i)

(57)

Equation (57) ﬁnds the hourly dimensionless water temperature in the tank, while, q(τ ) in Eq. (57) refers to net dimensionless input heat rate to the tank (Yumrutas and Ünsal, 2012). The input heat rate, q(τ ) , represents the diﬀerence between dimensionless useful solar energy and extracted energy. It was expressed by (Yumrutas and Ünsal, 2000; Yumrutas and Ünsal, 2012) as:

q (τ ) = qu (τ ) − qH (τ ) +

w (τ ) γ

(61)

4. Input data for the drying system

(56)

⎤ ϕw (τn − 1) − ∑i = 1 ⎦ p 1 1 + Δτ + π Δτ

(60)

where I(t) is hourly solar radiation on the horizontal surface, and R is the ratio of total radiation on the tilted surface to that on the horizontal surface. R can be estimated by considering the beam, diﬀuse, and ground-reﬂected components of radiation. Calculation procedure for the R is given by (Duﬃe and Backman, 1991).

where Vw , ρw andc w are water volume, density, and speciﬁc heat of the water in the tank. R, A and k are radius, surface area of the tank, and earth’s thermal conductivity. The problem of unsteady heat transfer by conduction is raised in Eqs. (52)–(56) through transferring them into dimensionless forms as presented in (Yumrutas and Ünsal, 2012). Then, dimensionless problem formulation will be obtained. All steps for the solution of the problem are given in (Yumrutas and Ünsal, 2012); therefore, only the solution is presented in this section.

ϕw (τn ) =

Tw (t ) − Ta (t ) ⎤ ⎥ IT (t ) ⎦

where Tw(t) is water temperature in the TES tank, and Ta(t) is the ambient air temperature. IT (t) in Eq. (60) is given by;

Some of the energy is charged in the tank in the form of sensible heat, and the remaining will be transferred to the surrounding earth as conduction heat loss. The energy balance equation can be expressed as:

Q = ρw c w Vw

(59)

(58)

where w(τ ), qH (τ) and qu (τ ) are dimensionless work input to heat pump, heat demand of the drying unit, and rate of solar energy charge to the tank, respectively while, γ is a dimensionless parameter (4π Rk/ṁ a Cpa ) . 3.5. Useful solar energy collection rate to the tank The solar energy charge rate, Q̇u (t) , is calculated by using the formula given by (Duﬃe and Beckman, 1991);

Fig. 3. Schematic of the underground TES tank problem. 545

Solar Energy 199 (2020) 538–551

H. Hasan Ismaeel and R. Yumrutaş

hourly heating requirement for drying unit is has been calculated through Eqs. (3–9). Depending on the mass ﬂow rate of the wheat, the hourly heating requirement of the drying unit is constant. Secondly, since ambient air temperature changes hourly, mass or volume ﬂow rate of the air will change. The power input required by the fan is calculated by using Eqs. (12) through (41). Thirdly, hourly useful solar ̇ (t), is computed by using Eqs. (59)–(61). Collector eﬃciency energy, Qu given by Eq. (60) was obtained by correlating results from the experimental study by (Yumrutas, and Kaska, 2004). It was computed via hourly solar radiation rates on tilted surfaces, water temperature in the tank, and outdoor air temperature. Solar energy gain is later transformed into dimensionless form using dimensionless expressions referred to in (Yumrutaş and Ünsal, 2012). Next, new hourly water temperature in the tank has been estimated through Eq. (57) by utilizing dimensionless net energy charge rate to TES tank expressed in Eq. (58). Finally, hourly, monthly, and annual values of water temperature in TES tank, COP, COPs, SMER, and energy fractions have been computed. These computations were done during each hour until annual water temperature distribution in the tank reached annually periodic operating conditions for 10 years of operation.

unit is neglected since it is very low in relation to compressor work. EES program is also used to calculate hourly power input to the blower or fan. 4.2. Heat pump Heat is absorbed by the evaporator of the heat pump from the underground TES tank, and some amount of energy is added to the working ﬂuid by the compressor. Heat is rejected to drying air to evaporate water in the wheat. Expressions for energy requirement and heat pump COP are derived in Section 3.3. A practical Carnot Eﬃciency (CE) value is set basically set 0.4 in the present study. Zogou and Stamatelos (1998) expressed that CE values range between 0.30 and 0.50 for small electric heat pumps, and consequently three CE values of 0.30, 0.40 and 0.50 were taken in this study. In the present study, CE value was set at 0.4, and tank volume was set at 200 m3 for all calculations unless stated otherwise. 4.3. The TES tank and its surrounding structure Water in the TES tank is assumed to be fully mixed. Speciﬁc heat and density of the water are 4.18 kJ/kgoC, and 1000 kg/m3, respectively. Temperature of the water at the beginning of operation is taken as equal to deep earth temperature of 15 °C. Solar energy is charged to the water if there is useful solar energy during the year. Some of the stored energy is absorbed by the heat pump, some is lost to the earth surrounding the tank, and the remaining part is stored as a sensible heat in the TES tank. In this study, limestone is basically considered as the earth surrounding the tank since it is the prevalent geo-structure of Gaziantep. For examining the impact of the earth type on system performance, three types of earth soils are considered: limestone, granite, and coarse graveled. Thermophysical properties of the earths are given in (Ozisik, 1985) and illustrated in Table 2. Tank volume is selected as 200 m3 in all calculations unless another volume is used.

6. Results and discussion It is necessary to ﬁnd performance parameters for the drying system in order to know whether the system within which the parameters operate are eﬃcient or not. The performance parameters are: water temperature in the TES tank, coeﬃcient of performance for the heat pump (COP) and system (COPs), energy ratio, and speciﬁc moisture evaporation ratio (SMER). They have been computed by executing a computer program in MATLAB hour by hour during 24 h along many years, and using drying system parameters such as Carnot Eﬃciency (CE), type of ground, collector area, mass ﬂow rate, and TES tank volume. In the calculations, limestone is used as a geological structure, and CE, collector area, volume of TES tank are set as 0.4, 70 m2, and 200 m3, respectively, unless other values of these input parameters are used. In order to ﬁnd suitable values of the drying system such as tank volume, collector area, mass ﬂow rate, results of numerical computations are presented in ﬁgures and discussed in this section.

4.4. Solar collectors In this study, ﬂat plate solar collectors have been utilized to collect useful solar energy in the tank. They are assumed to be directed toward the south, and their tilt angle is taken as latitude angle (37.1°N) of Gaziantep, Turkey since the highest solar energy is collected when the slope angle is equal to the latitude angle for yearly operation (Elminir et al., 2006). Solar energy charge to the underground TES tank given by Eq. (58) is calculated by using Eqs. (59) and (60). Hourly solar radiation values and fresh air temperatures are used in the numerical calculations. The values of the hourly solar radiation on horizontal surface and ambient air temperature were taken from Gaziantep Meteorological Station. Eﬀect of solar energy collection rate on performance of the drying system is investigated. During the calculations of the model, collector area as an input data was set as 70 m2 if another collector area was not used.

6.1. TES tank volume Storage tank volume is very important to determine size of the drying system. That is, storage volume should be minimized to oﬀer reasonable performance parameters. Fig. 5 is depicted to determine the storage volume for the selected cooked wheat mass ﬂow rate of 50 kg h−1. It is observed from the ﬁgure that lower TES tank volume gives higher water temperature during summer and lower water temperature during winter. That is, lower tank volume gives higher water temperature amplitude. When the tank volume is selected as 100 m2, water temperature decreases below or close to zero degrees Celsius during December, January and February. There is an important criterion that water temperature should be above 0 °C, and also, it should be higher than ambient air temperature for reasonable performance parameters. Once the tank volume is 200 m3, water temperature approaches 0 °C, but higher than it, and then, performance parameters can be acceptable values. For that reason, tank volume is selected as 200 m3

5. Computation procedure for the drying system In order to examine performance parameters of wheat drying system, a computation procedure is developed by using interrelated analytical models and expressions for each system component. The drying system consists of four main components; drying unit, heat pump, solar collectors and TES tank. The current drying system is named as solar-assisted heat pump wheat drying system with underground TES tank. A computer program in MATLAB based on expressions given in the model is developed for the numerical computation. The computation procedure for the program is illustrated as ﬂowcharts in Fig. 4, and explained clearly in this section. The computation procedure passes through several stages. Firstly, wheat mass ﬂow rate and moisture contents are selected, and then the

Table 2 Characteristics of the ground structures (Ozisik, 1985).

546

Ground type

Conductivity (W/ m K)

Diﬀusivity (m2/s)

Speciﬁc heat (J/kg K)

Heat capacity (kJ/m3 K)

Coarse

0.519

1.39 × 10−7

1842

3772

Limestone

1.3

5.75 × 10−7

900

2250

Granite

3.0

14.00 × 10−7

820

2164.8

Solar Energy 199 (2020) 538–551

H. Hasan Ismaeel and R. Yumrutaş

40

Tw (°C)

30 20 10 0 July

Aug

Sep

Oct

Nov

Dec

Jan

Feb

Mar

Apr

May

Jun

-10

Months Fig. 5. Eﬀect of tank volume on temperature of water in the tank during ﬁrst year. 40

Tw (°C)

30

20

10

0 July Aug

Sep

Oct

Nov

Dec

Jan

Feb

Mar

Apr May

Jun

Months Fig. 6. Eﬀect of collector area on annual temperature of water in the TES tank during ﬁrst year of operation.

and employed to examine the impact of other parameters upon the drying system performance. 6.2. Collector area In order to deﬁne collector area for selected load or mass ﬂow rate of cooked wheat, Fig. 6 illustrates three diﬀerent collector areas, which are 65, 70, and 75 m2. It is known that higher collector area gives higher water temperature. This indicates that there is a proportional relationship between collector area and water temperature in the TES tank. An important point is that the selection of collector area should provide reasonable temperature for the water in the TES tank. It is known that minimum collector area should be selected for reasonable performance parameters, and three selected areas were suitable. However, when the collector area is taken as 65 m2, or below it, water temperature approaches 0 °C, and consequently gives lower performance parameters of COP, COPs and SMER. Collector area should be selected as a minimum value for reasonable performance parameters. Therefore, the suitable collector area is 70 m2. 6.3. Mass ﬂow rate of cooked wheat Mass ﬂow rate of cooked wheat aﬀects performance parameters such as water temperature in TES tank, SMER, COP and COPs. Heating requirement of the wheat dryer is calculated when wheat mass ﬂow rate is selected. Fig. 7 shows variation of yearly water temperature in the TES tank for three diﬀerent cooked wheat mass ﬂow rate of 45, 50 and 55 kg h−1. When the mass ﬂow rate is selected as 55 kg h−1, water temperature approaches to 0 °C during January, and then performance parameters will take unreasonable value. It seen from this ﬁgure that mass ﬂow rate of cooked wheat can be selected as 50 kg h−1 since it is

Fig. 4. Flowchart of the MATLAB program..

547

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H. Hasan Ismaeel and R. Yumrutaş

summer months and lower values during winter months whenever the tank is surrounded by the coarse graveled earth. On the other hand, water temperatures for granite have lower values in summer and higher values in winter. Limestone gives medium temperature values that are between coarse graveled and granite, which accordingly give medium performance parameters. Temperatures of water in the tank take values about between 5 and 10 °C during winter, which are reasonable values. However, granite and coarse graveled earth types are not widely present as a geological structure around Gaziantep City, Turkey. All ﬁgures are depicted for limestone ground since it is widely present in Gaziantep unless mentioned otherwise. Fig. 10 shows long term temperature variations of water in the tank surrounded by three kinds of earth; coarse graveled earth, limestone, and granite for September. It is seen from Fig. 10 that coarse graveled earth gives higher water temperatures in the tank than limestone and granite. Such a variation may be attributed to three thermophysical properties for the ground, which are thermal conductivity, speciﬁc heat and density. Heat capacity (C = ρCp) is a property, and depends on density, and speciﬁc heat. Thermal diﬀusivity (α = k/ρCp) is another property and depends on thermal conductivity, density and speciﬁc heat. From Table 2, which includes the thermophysical properties, it can be seen that thermal conductivity and diﬀusivity for the coarse graveled earth are below the corresponding properties of the other two types, while speciﬁc heat and heat capacity are higher than those of granite and limestone. It is observed from these properties that lower thermal conductivity and diﬀusivity decrease heat transfer and propagation from the TES tank to the surrounding earth, while higher speciﬁc heat or heat capacity increases heat absorption capability. Variation of water temperature during the ﬁrst few years and after the ﬁfth year is another observation. Water temperature in the tank increases rapidly until year 5 of operation. This indicates that the drying system reaches periodic operation in the ﬁfth year for the given input system parameters, which adds more signiﬁcance to the drying system.

Fig. 7. Eﬀect of cooked wheat mass ﬂow rate on annual temperature of water in the TES tank during ﬁrst year of operation.

Fig. 8. Eﬀect of wheat mass ﬂow rate variation on COP

suitable. The eﬀect of cooked wheat mass ﬂow rate on COP during the ﬁfth year or periodic operation is shown in Fig. 8. As it is shown in the ﬁgure, three diﬀerent wheat ﬂow rates (45, 50 and 55 kg h−1) are selected to determine suitable ﬂow rate under the given input data. The ﬁgure also shows that when the wheat mass ﬂow rate is increased, the COP decreases. However, COP of heat pump range from 3 to 4 for mass ﬂow rate of 55 kg h−1. While, COP values varied from 4 to 5 for the mass ﬂow rates of 45 ad 50 kg h−1 during the periodic operation year. Thus, these values are reasonable for the solar-assisted heat pump drying system. When mass rate of wheat is higher than 55 kg h−1for TES tank volume of 200 m3 and collector area of 70 m2, water temperature will approach to 0 °C, and consequently COP will go under 3 during the ﬁrst year, which makes this value unreasonable. Consequently, W4 or mass rate of cooked wheat is taken as 50 kg h−1 for all calculations since it is suitable for all performance parameters of the drying system.

6.5. Periodic operation of the drying system Yearly water temperature variation in the tank for the ﬁrst, second, third and ﬁfth year of operation is shown in Fig. 11. It is seen from the ﬁgure that water temperature varies rapidly during the ﬁrst few years. After that this variation decreases, and the drying system reaches the annual periodic operation in year 5 of operation onwards. It is evident that the number of periodical operation is equal to 5 years. That is, energy input to the tank is equal to heat loss from the tank, and energy balance takes place starting from the ﬁfth year of operation. Fig. 12 is depicted for yearly variations of COP, COPs and SMER in order to see the eﬀect of fan power on COP and periodic operation year. It is seen from the ﬁgure that COP and COPs are reasonably closer to each other for several years. Its reason is that power consumption of the 50 40

Tw (oC)

6.4. Type of earth surrounding TES tank Each type of earth surrounding the TES tank gives diﬀerent performance parameters because of diﬀerent thermophysical properties. Water temperature variation in the tank is a type of performance parameter, which is heart of the drying system. Hence, every performance parameter depends on the storage temperature. For that reason, Fig. 9 depicts the individual eﬀects of the three selected types of ground on water temperature variation in the TES tank during ﬁfth year of operation by using these input parameters; collector area (70 m2), Carnot eﬃciency factor (40%), storage volume (200 m3), and mass ﬂow rate of cooked wheat (50 kg h−1). The reasons behind selecting these parameters are discussed in this study by indicating ﬁgures. It is observed from Fig. 9 that water temperatures have higher values during

Limestone Coarse graveled Granite

30 20 10 0 July Aug Sep

Oct Nov Dec

Jan

Feb Mar Apr May Jun

Months Fig. 9. Eﬀect of earth type on annual temperature variation of water in the TES tank during ﬁfth year of operation. 548

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H. Hasan Ismaeel and R. Yumrutaş

40

6

30

COP

Tw (°C)

4

Coarse graveled Limestone Granite

20

CE=0.3 CE=0.4 CE=0.5

2

10

0

0 1

2

3

Years

4

5

1

6

2

3

Years

4

5

6

5

6

Fig. 14. Eﬀect of CE on COP.

Fig. 10. Eﬀect of earth type on annual temperature variation of water in the TES tank for September.

8 40 1st Year 2nd Year 3th Year 5th Year

6

SMER

Tw (°C)

30

20

4

CE=0.3 CE=0.4 CE=0.5

2

10

0 July Aug

Sep

Oct

Nov

Dec

Jan

Feb

Mar

Apr May

0

Jun

1

Months

2

3

4

Years

Fig. 11. Annual temperature variation of water in the TES tank for September.

Fig. 15. Eﬀect of CE on SMER.

7

100 80

5

Energy fraction

SMER, COP & COPs

6

4 3

SMER COP COPs

2

0 2

3

Years

4

5

Solar

Tw (°C)

Work

Fan

Stored

Lost

Load

Fig. 16. Variation of energy fractions with years of operation.

fan is very low with respect to compressor power. Yearly fan power required is about 3% of the compressor load. Another point that can be drawn from the ﬁgure is that the performance parameters do not change or stay constant after the ﬁfth year. It means that the drying system reaches periodic operation at the ﬁfth year. This type of heating system requires periodic operation time from 5 to 7 years (Yumrutaş and Ünsal, 2012).

50 CE=0.3 CE=0.4 CE=0.5

30

40

0

6

Fig. 12. Variation of yearly COP, COPs and SMER.

40

1.Year 2.Year 5.Year

20

1 1

60

20 10

6.6. Carnot eﬃciency

0 July

Aug

Sep

Oct

Nov

Dec

Jan

Feb

Mar

Apr

May

Carnot eﬃciency (CE) is deﬁned as the ratio of Carnot COP to the actual COP. Actual COP relies upon real heat pump types and sizes. Zogou and Stamatelos (1998) reported in their study that CE values in the range from 0.30 to 0.50 for small electric heat pumps. Therefore, three CE values of 0.30, 0.40, and 0.50 were selected in this study. Annual variation of water temperature in the TES tank is shown in

Jun

Months Fig. 13. Eﬀect of CE on annual temperature variation of water in the TES tank during ﬁfth year of operation.

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compressor work, and fan power are 76.6%, 22.7%, and 0.7% respectively from year 5 onwards for 10 years of operation. Condenser fan power is below 1% of total energy input to the drying system.

Fig. 13. When CE increases, water temperature decreases, and COP of the heat pump and SMER increase. Yearly variation of COP and SMER with three diﬀerent Carnot Eﬃciencies are shown in Figs. 14 and 15, respectively. They increase with the number of operation years and reach periodic operation during ﬁfth year, and they remain constant after periodic conditions. It is seen from these two ﬁgures that COP and SMER change between about 3 and 5.2, and 4.2 and 7 for all years, respectively. These values are reasonable high values. It is important to make yearly energy balance for the drying system. The energy provided to the drying system consists of solar energy, heat pump work, and fan power. The energy supplied to the system should be equal to the total useful energy or load for the dryer unit, partially stored energy in the TES tank and energy lost to the surrounding earth. Energy fractions are the ratio of individual energy components to the overall energy provided to the system. Fig. 16 shows energy fractions during year 5 of operation for the ﬁrst, second, and ﬁfth year of operation. The ﬁgure shows that solar energy increases, compressor work decreases, and fan power stays constant throughout the years. At the same time, both stored and lost energy decrease, and load increases with the years. It is observed that there is no stored energy during the ﬁfth year of operation; that is, the system operates periodically during and after the ﬁfth year. Total energy input to the drying system supplied by solar energy, compressor work, and fan power are 76.63%, 22.68%, and 0.7% respectively during year 5 of operation. The ﬁgure demonstrates that fan power is below 1% of total energy input.

The followings are recommended for further studies related with the SAHP drying systems with underground TES tank.

• Optimization of the SAHP drying system can be studied, and most suitable size of the system can be obtained. • Thermo-economic analysis for the SAHP drying system is important, and it can be studied.

Declaration of Competing Interest The authors declare that they have no known competing ﬁnancial interests or personal relationships that could have appeared to inﬂuence the work reported in this paper. References Aktaş, M., Khanlari, A., Amini, A., Şevik, S., 2017. Performance analysis of heat pump and infrared–heat pump drying of grated carrot using energy-exergy methodology. Energy Convers. Manage. 132, 327–338. ASHRAE, 1997. ASHRAE handbook fundamentals. American society of heating, refrigerating and air conditioning engineers. Atlanta, USA. ASHRAE, 1999. ASHRAE handbook applications. American society of heating, refrigerating and air conditioning engineers. Atlanta, USA. ASHRAE, 2000. ASHRAE handbook systems and equipment. American society of heating, refrigerating, and air- conditioning engineers. Atlanta, USA. Alves-Filho, O., 2015. Heat pump dryers: Theory, design and industrial applications. CRC Press, Taylor & Francis Group. Akers, W.W., Deans, H.A., Crosser, O.K., 1958. Condensing heat transfer within horizontal tubes. Chem. Eng. Progr., pp. 54. Baniasadi, E., Ranjbar, S., Boostanipour, O., 2017. Experimental investigation of the performance of a mixed-mode solar dryer with thermal energy storage. Renew. Energy 112, 143–150. Cavallaro, A. R., Bullard, C. W., 1994. Eﬀects of varying fan speed on a refrigerator/ freezer system. Air Conditioning and Refrigeration Center, College of Engineering. University of Illinois at Urbana Champaign. Chung, M., Park, J., Yoon, H.K., 1998. Simulation of a Central Solar Heating System with Seasonal Storage in Korea. Sol. Energy 64, 163–178. Duﬃe, J.A., Beckman, W.A., 1991. Solar engineering of thermal processes. John Wiley & Sons, New York. Daghigh, R., Ruslan, M.H., Sulaiman, M.Y., Sopian, K., 2010. Review of Solar Assisted Heat Pump Drying Systems for Agricultural and Marine Products. Renew. Sustain. Energy Rev. 14, 2564–2579. Elminir, H.K., Ghitas, A.E., El-Hussainy, F., Hamid, R., Beheary, M.M., Abdel-Moneim, K.M., 2006. Optimum solar ﬂat-plate collector slope: case study for Helwan, Egypt. Energy Convers. Manage 47, 624–637. Esen, H., Esen, M., Ozsolak, O., 2017. Modelling and experimental performance analysis of solar-assisted ground source heat pump system. J. Exp. Theor. Artif. Intell. 29, 1–17. Esen, H., Inalli, M., Esen, M., 2007. A techno-economic comparison of ground-coupled and air-coupled heat pump system for space cooling. Build. Environ. 42, 1955–1965. Esen, M., Esen, H., 2005. Experimental investigation of a two-phase closed thermosyphon solar water heater. Sol. Energy 79 (5), 459–468. Icropera, F.P., De Witt, D.P., 2006. Fundamentals of heat and mass transfer, 6th Edition. John Wiley and Sons Inc., USA. Inallı, M., 1998. Design Parameters for a Solar Heating System with an Underground Cylindrical Tank. Energy 23, 1015–1027. Kuyumcu, M.E., Tutumlu, H., Yumrutaş, R., 2016. Performance of a Swimming Pool Heating System by Utilizing Waste Energy Rejected from an Ice Rink with an Energy Storage Tank. Energy Convers. Manage. 121, 349–357. Kivevele, T., Huan, Z., 2014. A review on opportunities for the development of heat pump drying systems in South Africa. S. Afr. J. Sci. 110, 01–11. Kazarian, E.A., Hall, C.W., 1962. Thermal properties of grain. (Doctoral dissertation, Michigan State University of Agriculture and Applied Science. Dept. of Agricultural Engineering). Kays, W.M., London, A.L., 1984. Compact heat exchangers, 3rd ed. McGraw Hill, Blacklick, Ohio, USA. KCD dryer continuo’s mixed ﬂow, 2019. Operation manual. www.kongskilde.com, kcdmk4_man-gb_117012512. Leon, M.A., Kumar, S., Bhattacharya, S.C., 2002. A comprehensive procedure for performance evaluation of solar food dryers. Renew. Sustain. Energy Rev. 6, 367–393. Mrema, G.C., Gumbe, L.O., Chepete, H.J., Agullo, J.O., 2011. Rural structures in the tropics, Design and development. Food and Agriculture Organization of the United Nations. Available: http://www.fao.org/3/i2433e/i2433e00.pdf. Mohanraj, M., Chandrasekar, P., 2009. Performance of a forced convection solar dryer integrated with gravel as heat storage material for chili drying. Journal of Engineering Science and Technology, School of Engineering, Taylor’s University

7. Conclusions and recommendations In this study, mathematical modeling of a wheat drying system operating with a ground coupled heat pump and underground Thermal Energy Storage (TES) tank charged by solar energy is developed to investigate long-term performance parameters for the drying system. These parameters are hourly temperature of water in the TES tank, coeﬃcient of performance of the heat pump (COP) and system (COPs), speciﬁc moisture evaporation rate (SMER) and energy fractions. These parameters are obtained by using the solution of heat transfer problem around the TES tank and energy expressions for other components of the SAHP drying system. The performance parameters are computed by a computer program prepared in MATLAB based on the analytical model for the solar assisted heat pump (SAHP) drying system. The drying system is operated hour by hour for 24 h in a day and during the whole year. There are many important input system parameters for the program, such as collector area, type of earth, TES tank volume, Carnot Eﬃciency (CE), mass ﬂow rate of product, design conditions, etc. The program is executed by using the input parameters in order to ﬁnd the eﬀect of these parameters on the performance parameters for the drying system and to decide suitable system parameters. The most important results are outlined below:

• It is observed that all performance parameters, such as water tem• •

•

perature in the tank, heat pump COP and system COPs, SMER, and energy fraction for the load increase gradually from the beginning of drying operation until the ﬁfth year of operation. The system works as a periodic operation from ﬁfth year onwards. Earth type or thermophysical characteristics of the geological structure around the underground TES tank have a great eﬀect on the drying system performance. Thus, it is recommended that a suitable region should be selected for installing such a drying system. When wheat mass ﬂow rate and CE are selected as 50 kg h−1 and 40%, more suitable collector area and TES tank volume are obtained as 70 m2 and 200 m3, respectively. For these values, the performance parameters of the COP, COPs and SMER are obtained as 4.43, 4.3 and 6.05 respectively, when the drying system reaches periodic operation from ﬁfth year onwards. Total energy input to the drying system supplied by solar energy, 550

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