Solar cooling

A. Mahesh and S. C. Kaushik Citation: J. Renewable Sustainable Energy , 022701 (2012); doi: 10.1063/1.3691610 View online: http://dx.doi.org/10.1063/1.3691610 View Table of Contents: http://jrse.aip. http://jrse.aip.org/resource/1/JRSE org/resource/1/JRSEBH/v4/i2 BH/v4/i2 Published by the American the American Institute Institute of Physics.

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JOURNAL OF RENEWABLE AND SUSTAINABLE ENERGY 4, 022701 (2012)

Solar adsorption cooling system: An overview A. Mahesha) and S. C. Kaushik Kaushik Centre for Energy Studies, Indian Institute of Technology Delhi, Hauz Khas, New Delhi 110016, India (Received 9 December 2011; accepted 23 January 2012; published online 16 March 2012)

In thi thiss pap paper, er, a rev review iew of the performa performance nce and developme development nt of var variou iouss sol solar ar powered adsorption refrigeration systems is presented here. The review covers the current curre nt state of theor theoretica eticall study study,, choic choicee of adsorb adsorbent-a ent-adsorba dsorbate, te, experi experimenta mentall performance perfor mance study, and furth further er scope of research in adsorp adsorption tion cooling system. Finally, detailed summary and suggestions are outlined for proper utilization of C 2012 201 2 Ame Americ rican an thermally therm ally opera operated ted adsorp adsorption tion refr refrigera igeration tion techn technologi ologies. es. V http://dx.doi.org/10.1063/1.3691610]] Institute of Physics . [http://dx.doi.org/10.1063/1.3691610

I. INTRO INTRODUCTIO DUCTION N As a result of the projected world energy shortage, the use of solar energy energ y for environ environmental mental 1 control cont rol has rec receive eived d a lot of atte attenti ntion on in engi engineer neering ing and scie science nce lite literatu rature. re. Energy shortages and inconsistent power availability cripple society’s progress like preservation field. Refrigeration is particularly an attractive application for s olar energy because of the near coincidence of peak cooling loads with the available solar power. 2 Almost all modern facilities use energy to generate the power to fulfill for lighting, cooling, refrigeration, and other building services. The resulting impact is to increase the fuel cost and the level of CO2 production dramatically. In addition, energy shortages and inconsistent power availability cripple progress in the food preservation field. Therefore, energy efficient conservation is a vital role for future generation and it can be achieved by using available waste heat or solar energy from a vapor adsorption (VA) system for various cooling applications. Among solar refrigeration alternatives, the solar vapor adsorption system appears to be one of the most promising technologies because it is environmentally friendly with low cost and low maintenance requirements. Hence, the prese present nt artic article le looks at the reexa reexamine mine of exist existing ing literature literature based on performancee tre anc trends nds of ads adsorp orptio tion n coo coolin ling g syst systems ems cou couple pled d wit with h sol solar ar the therma rmall syst systems ems,, whi which ch are referred to as present and past status of theoretical studies, choice of adsorbent–adsorbate pairs, experimental studies, and further scope of research is discussed. It is believed that this literature gives giv es cle clear ar cut ide ideaa to app apprai raise se the fea feasib sibili ility ty for app applyi lying ng sol solar ar ene energy rgy in the therma rmally lly dri driven ven adsorption cooling system. II. DESCRIPTION OF VAPOR VAPOR ADSORPTION SYSTEM Adsorption is a surface phenomenon which can be classified into two groups: (1) Physical adsorption (Physorption) and (2) Chemical adsorption (Chemisorption). Physorption is a reversible proce process ss and mainl mainly y caused by disper dispersionsion-repuls repulsion ion (Vander Walls Walls)) force and elect electrostat rostatic ic forces for ces bet betwee ween n ads adsorb orbate ate (ga (gas) s) mol molecu ecules les and ato atoms ms whi which ch com compos posee the ads adsorb orbent ent (po (porou rouss substa sub stance nce)) sur surfac face. e. Mos Mostt of the VA pro process cess applicabl applicablee to the therma rmall or coo coolin ling g syst systems ems mai mainly nly utilizes physisorption. The basic adsorption cycle relies on the adsorption of a refrigerant vapor (adsorbate) into an adsorbent bed at low pressure and subsequent desorption at a high pressure by heating the adsorbe adso rbent nt bed bed.. In the simplest simplest cas case, e, an ads adsorp orptio tion n ref refrig rigera erator tor can be con consid sidere ered d as a two

a)

Author to whom corr Author correspo esponden ndence ce shou should ld be addr addresse essed. d. Elec Electron tronic ic mai mail: l: mahe maheshiit shiit10@ 10@gmai gmail.co l.com. m. Tele Telephon phone: e: þ91 9899669886 9899669886..

1941-7012/2012/4(2)/022701/20/$30.00

4, 022701-1

C 2012 American Institute of Physics V

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A. Mahesh and S. C. Kaushik

J. Renewable Sustainable Energy 4, 022701 (2012)

vessels connected to each other, one of which is filled with adsorbent and adsorbate as shown in Figure 1. At the beginning stage 1, the system is at a low pressure and temperature. The adsorbent bed is ideally always saturated with vapor at this point. The adsorbent bed is heated and refrigerant start to desorb from the adsorbent bed at stage 2, which raises the system pressure. Desorbed refrigerant condenses in the second vessel and rejects heat as a result of condensation. Finally, the hot adsorbent bed (stage 3) is cooled back to ambient temperature at stage 4 causing the refrigerant to readsorb on the bed. The basic of adsorption refrigeration cycle consists of four thermodynamic steps which can be clearly represented by a clapeyron diagram (ln P Vs À1/T) which is represented in Figure 2. The cycle begins at point Ta2 where the maximum amount of refrigerant is absorbed. The adsorbent is at low temperature and at a low pressure at point T a2. Along the cycle T a2-Tg1, the adsorbent is heated and desorbs refrigerant vapor isosterically (i.e., at constant total adsorbed mass on the adsorbent). This step is isosteric, because there will be refrigerant flow until the pressure inside the adsorbent bed becomes equal or greater than the pressure of the condenser and the amount of desorbed refrigerant is small relative to the total amount adsorbed.

 

ði:eÞ

Dmads

mads

% 0:

Continued heating form T g1ÀTg2 desorbs more refrigerant, forcing it to the condenser until state Tg2 is attained, at which desorption ceases. The second step is isobaric desorption. Then, the hot adsorbent is cooled isosterically causing adsorption and depressurization (Tg2ÀTa1). When the pressure drops below Pevp, refrigerant in the evaporator starts to boil and then it flows to the adsorbent bed, producing a cooling effect. Cooling of the adsorbent continues until the bed is saturated with refrigerant, hence completing the cycle. This process (T a1ÀTa2) is also isobaric adsorption. Coefficient of performance (COP) of the basic adsorption cycle can be expressed as,

COP ¼

Qrefri : Qin þ Qide

Adsorption pairs are one of the most important criteria of any refrigeration system, because working conditions and environmental considerations mainly depend on them. The ideal refrigerant has the following characteristics:3

FIG. 1. Basic adsorption cycles.

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Solar adsorption cooling system


FIG. 2. P-T-x diagrams for basic adsorption cycle.

•

•

•

•

•

•

•

•

•

•

•

•

•

Low specific heat Low viscosity Good thermal stability High thermal conductivity Chemically stable in the working temperature range The molecular dimension should be small enough to allow energy adsorption Non-toxic, non flammable, and non-corrosive. On the other hand, the ideal adsorbent has the following characteristics: High adsorption and desorption capacity Good thermal conductivity Low specific heat capacity Chemically compatible with choosing refrigerant Reversibility of adsorption process for many cycles.

The details for the selection of adsorbent-adsorbate and its performance are discussed in Sec. IV. The present and past status of theoretical, experimental studies, economic analysis, and further scope of research is discussed in detailed in coming sections. III. THEORETICAL STUDIES A. Adsorption mechanisms In order to understand the adsorption phenomenon thoroughly, adsorption equilibrium should be introduced. Adsorption equilibrium is the state in which adsorption and desorption rates are the same. Investigation of adsorption equilibrium is crucial in determining the performance limitations of the adsorption refrigeration cycles. In practice, the maximum capacity of adsorbent cannot be fully utilized because of the mass transfer effect involved in actual fluid-solid interaction process. Three models are generally used to describe refrigerant trans f er between gaseous and solid phases, i.e., the equilibrium model,4 the solid diffusion model,5 and the linear driven force model (LDF).5 The equilibrium model considers no mass transfer limitations between solid and gas phases, i.e., they are in thermodynamic equilibrium. The adsorption mass transfer rate is

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only related to gaseous pressure and adsorbent temperature through the following adsorption equilibrium equation proposed by Dubinin and Radushkevitch (D-R) at the end of 1947.

dx dxÃ ¼ ; dt dt

(1)

where the adsorption capacity x concentration of refrigerant in bed at temperature T and pressure P, xÃ is the equilibrium adsorption capacity of adsorber which can be given by a D-R equation

xÃ ¼ x0 expðÀk 2 2 Þ;

(2)

where x0 is the saturated adsorption capacity, k is the characteristic parameter for a given adsorbent/adsorbate pair, and [ is the adsorption potential. Adsorption on non-porous solid is different than adsorption on porous solids; surface adsorption dominates the D-R equation where as micropore adsorption dominates the Polonyi model. This is the reason that promoted the development of the D-R Eq. (1), but the D-R equation has several limitations. 6 Mostly applicable for microporous solid with practically homogenous or uniform microstructures, Derived from Henry’s law at low pressure, Valid for degree of micropore filling higher than 0.15. •

•

•

Later, Dubinin and Astakhov (D-A) modified the D-R equation. The D-A equation can be expressed as Ã

x ¼ x0

n

     T ad Àk À1 T sat

;

(3)

where Tsat is the saturation temperature, T ad is the adsorbent bed temperature, and k and n are the adsorption parameters depending on the material of the adsorbent-adsorbate pairs. In fact, due to diffusion through intra-particle macropores and micropores, which is a slow process, the two phases of gas and solid could not reach equilibrium instantly, but only after a certain time, depending on the adsorbent-adsorbate pair used. For this reason, mass transfer kinetics be incorporated into the model to verify whether the equilibrium assumption is justified.7 The LDF model can be used to describe mass transfer between moving and stationary phases in adsorption processes, which is derived from a solid diffusion model based on the assumption made by Gluekauf. 8 In the LDF model expressed by the difference of the amount adsorbed (q), and the maximum amount adsorbed (q0) in equilibrium at certain pressure condition.

dq ¼ K m ðq0 À qÞ; dt K m ¼

15 Ds

R p2

Ds ¼ Dso

(4)

À 

Ea ; RT

where Rp is the radius of the particle, D s is the surface diffusivity of particle, and D so is the pre-exponential constant.

B. Thermodynamic model Among the different models, thermodynamic model is the simplest and in which details of heat transfer are not considered. According to the thermodynamic point of view, vapor adsorption refrigeration cycle consist of four steps which are represented below along with basic adsorption refrigeration cycles in Figure 3.

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(1) (2) (3) (4)



Isosteric heating (constant concentration) Isobaric desorption (heat of adsorption þ sensible heating) Isosteric cooling (constant concentration) Isobaric adsorption (heat of adsorption þ sensible cooling).

1. Isosteric heating

The temperature of adsorbent bed is increased from T a2 to Tg1 (shown in Figure 2) by heating the adsorbent bed. The heat transferred to the adsorbent bed to increase the temperature of the bed from Ta2 to Tg1 is given by the following equation: T g1

Qih ¼

ð

½mðC pad bent þ xC pad bate Þ þ mad mat Â C p ad mat dt :

(5)

T a2

2. Isobaric desorption

After the isosteric heating of the adsorbent bed, the heating process is continued and desorption process is started and the pressure of the cycle remains constant. Tg z

Qide ¼

ð

T g2

½mðC pad bent þ xmin C pad bate Þ þ mad mat Â C pad m at dt þ

T g2

ð

mD H :dx:

(6)

T g1

3. Isosteric cooling

T a1

Qic ¼

ð

½mðC pad bent þ xmin C pad bate Þ þ mad mat Â C pad mat dt :

(7)

T g2

When the adsorption bed reaches its maximum bed temperature T g2, desorption ceases and the adsorbent is allowed to cool down. During this process, decrease in temperature Tg2 to Ta1 induces decrease in pressure from P con to Pevap. 4. Isobaric adsorption

The adsorbent bed is connected to the evaporator, and adsorption evaporation occurs while the adsorbent is cooled from Ta1 to Ta2. During this cooling period, heat is withdrawn to decrease the temperature of the adsorbent. Ta 2

Qiad ¼

ð

Ta 2

½mðC pad bent þ xC pad bate Þ þ mad mat Â C p ad mat dt þ

T a1

ð

mD H :dx:

(8)

Ta1

The useful refrigeration effect which is the energy that must be supplied to the evaporator, Q refri, is calculated as the latent heat of evaporation of the cycled adsorbate, minus the sensible heat of the adsorbate that is entering the evaporator at condensation temperature. Tc on

Qrefri ¼ mð xmax À xmin Þ½ LðT e Þ À

ð

C pl ðT Þdt ;

(9)

T evp

where Cpl is the specific heat of the adsorbate in the liquid phase and L is the latent heat of evaporation of adsorbate. On the basis of the above described equations, the COP for cooling operation can be calculated as the ratio of the useful refrigeration effect produced and heat input to the adsorbent bed

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FIG. 3. Schematic representation of the basic adsorption refrigeration system.

COP ¼

Qrefri : Qith þ Qide

(10)

In literature, theoretical frameworks of adsorption cooling systems with their mathematical explanations have been widely discussed. The relevant studies are chronologically summarized in this article in concise style. In 1988, Douss et al. were developed a numerical model for simple heat and mass recovery cycles with zeolite NaX –water pair.9 According to the numerical prediction, the simple cycle and heat-mass recovery cycle reached the COP 1.38 and 1.56, respectively. The authors stated that condensation and evaporation pressures had the largest effect on the COP values among adsorption bed temperatures and other auxiliary component’s design parameters such as heat transfer coefficients. In 1995, Zheng et al. investigated the effects of operating conditions of a heat recovery adsorption refrigeration system. 10 In this theoretical study, activated carbon–ammonia pair was used. Conditions like maximum bed temperature and ambient temperature were analyzed. The optimized results showed that increasing maximum bed temperature from 180 to 250  C increased the COP by 20%. Moreover, as the ambient temperature increased, COP decreased. One more theoretical analysis was conducted by Zheng et al. (1995) for activated carbon–ammonia pair of two-bed refrigeration system. 11 In this study, the effects of several parameters related to cycle time and heat transfer coefficients on the cycle performance were investigated. Increasing the adsorbent bed heat capacity, decreased the COP and cooling capacity of the cycle whereas increasing the mass of heat transfer fluid within the adsorbent bed increases COP slightly, but decreases cooling capacity considerably. Another model of simple and heat recovery cycle was developed by Teng et al. in 1997.12 In this model, the effects of the several parameters in Dubinin–Astakhov (D-A) equation on cycle performance were investigated. Besides the effects of maximum bed temperature, evaporation temperature and heat capacity were analyzed. As expected, COP values of the activated carbon–methanol pair increased by increasing the maximum bed temperature varied between 90 and 140  C, and the evaporation temperature varied between À20 and 15  C. In this analysis, maximum COP value was reported approximately 1.5. Sun et al. in 1997 developed an adsorption heat pump model based on only heat transfer equations and excluding mass transfer. 13 Herein, zeolite–ammonia pair was used with constant heat of adsorption value for analyses. The model considered two-dimensional heat transfer within the adsorbent bed: axial for heat transfer fluid only (excluding axial heat transfer for adsorbent) and radial heat transfer. This two-dimensional model was solved with a number of assumptions with ordinary differential equations. Accordingly, the authors reported a maximum COP of 0.86 for the following condition: maximum bed temperature at 260  C, condensation temperature of 40  C, evaporation temperature at 5  C, and ambient temperature at 20  C.

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Wang et al. in 2000 (Ref. 14) predicted the performance of solar powered ice maker coupled with water heater through energy analysis with an annual year. In this investigation, activated carbon–methanol pair was preferred as working pair. The results showed that the maximum COP of the refrigeration cycle was 0.51 for the period from October to December according to the simulated operating conditions. The simulation results were higher than the experimental results in which the COP was 0.386 in two days in December 1998. A solar adsorptive ice maker model is presented and experimentally validated with an innovative approach of collector condenser technology. 15 The model first was used to identify convection heat transfer coefficients on the upper plate outside of the collector, (h av) and outside the condenser, (har ) and the equivalent heat transfer coefficient inside the condenser (i.e., equivalent heat transfer coefficient means the global heat transfer coefficient including convection, gas conduction, and radiation between the rear adsorbent bed and the inner condenser plate). Then, to study the sensitivity of the daily ice production (DIP) vis-a-vis critical physical parameters of the unit and to estimate the limit of the collector–condenser technology. As from the predicted results, the authors reported the DIP increases with increasing h ar and decreases slightly beyond optimum with increasing h av. According to this model, DIP reached to 11.5 kg/m2 of collector with corresponding COP of 19% reached with best condition of physical and meteorological parameters. In an investigation made by Alam et al., (2000), two dimensional heat transfer equations were established for both the fluid and adsorbent sides.16 This study analyzed a set of nondimensional parameters, which presented the different physical design and operating parameters of the system. A parametric study was also conducted to show the effect of different nondimensional parameters on the system performance and their results were simil ar to the results obtained by Hajji and Khalloufi (1995) (Ref. 17) as well as Zheng et al. (1995).18 In a novel approach, a uniform pressure model describing the heat and mass transfer in the adsorbent bed for a flat plate solar ice maker was presented by Li and Wang in 2003. 19 The model predicts the dynamic performance of an intermittent adsorption cycle using activated carbon–methanol as pairs. The numerical results show that the proposed model predicted the dynamic response of the solar solid adsorption system with good accuracy. The errors between the calculated and measured results are less than 4%. The author’s suggested that this model is a good tool for optimizing the adsorption system driven by solar energy. Solar powered adsorption (SPA) refrigerator using an evacuated tube for ther mal insulation was numerically made by Li et al. in 2003 with zeolite–water pair as an adsorber. 20 Generally, because of the limit to the mass of adsorbent in an evacuated tube, a number of evacuated tubes are needed to meet the given cooling capacity. They are usually assembled in parallel, thus radiation heat transfer will occur between two adjacent tubes. As shown in Figure 4, the small area d F1 on tube 1 has a range of radiation (arc BFC) on tube 2. The angle factor is calculated as follows:

X dF1;dF2 ¼

Cosb Â Cosc dF2 : p Â r 2

FIG. 4. Radiation between two adjacent tubes.

(11)

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The authors numerically computed relationships between the diameter of evacuated tube collector (ETC) and performance with respect to distance among two adjacent tube centers, when the diameter is equal to 70 mm, both the COP and the cooling capacity reach their maximum values 0.25 and 4377 kJ mÀ2, respectively. Although whose diameters are less than 70 mm can show higher values, their COP and cooling capacity may be lower due to the smaller mass quantity of the adsorbent. The investigator finally suggested that structure parameters, such as the diameter of the evacuated tubes, the mass quantity of adsorbent, and the distance between two adjacent tube centers, have important effects on the maximum average temperature and the performance of the system. A theoretical study of the two-stage adsorption cooling cycle was developed by Alam et al. in 2004.21 Silica gel–water pair was selected as in the theoretical study, and effects of cycle time, maximum bed and condensation temperature, and mass of adsorbent on the system performance were investigated. According to the results, COP values increased, achieved a peak value and then decreased as the maximum bed temperature increased. Cooling capacity increased with decreasing cycle time and increasing the mass of adsorbent, while with longer cycle times, better COP values could be attained. Finally, as the condensation temperature decreased, both COP and cooling capacity increased. Wang et al. in 2005 developed an analytical model for a two-bed adsorption cooling cycle with silica gel–water pair. 22 Effects of heat and mass recovery and operating temperatures on the system performance were investigated. The highest COP was obtained when the maximum bed temperature was 65–85  C range. The maximum predicted COP of the cycle was 0.65 when the condensation temperature was 20  C. Another theoretical model for solar-powered regenerative adsorption refrigeration was developed by Lambert in 2007.23 In this study, activated carbon–ammonia pair with CaCl 2 additions were used. As a result of CaCl2 addition, the adsorption capacity of the pair increased by 35% on average. According to predicted results, a maximum COP of 1.60 was reported. Additionally, different types of solar collectors (single or double-glazed flat plate with or without coating, evacuated tube, and parabolic concentrator) were compared under certain conditions and the highest solar thermal efficiency was obtained from an ETC while parabolic concentrator had second highest solar thermal efficiency. A two-dimensional numerical heat and mass transfer model that accounted for heat transfer limitations was developed by Liu and Leong in 2008 for a transient simple adsorption cooling cycle. 24 The investigators claimed that the condensation pressure is not constant as time progresses for the isosteric process due to heat transfer limitations. Contradictory to the authors’ stated that the simulation results showed that the pressure in the condenser does not change significantly for the simple adsorption cycle and COP increases as the mass flow rate within the condenser increases. Fadar et al. in 2009 examined numerically of a continuous adsorption refrigeration system consisting of two adsorbent beds and powered by parabolic trough solar collector (PTC). 25 Activated carbon as adsorbents and ammonia as refrigerant were used in the study. A predictive model accounting for heat balance of the solar collector components and instantaneous heat and mass transfer in adsorbent bed was presented. The validity of the theoretical model was tested by comparison with experimental data of the temperature evolution within the adsorber during isosteric heating phase. A good agreement is obtained. The system performance is assessed in terms of specific cooling power (SCP), refrigeration cycle (COPcycle), and solar COP, which were evaluated by a cycle simulation computer program. Under the climatic conditions of daily solar radiation being about 14MJ per 0.8 m2 (17.5 MJ/m2) and operating conditions of evaporating temperature, Tev ¼ 0  C, condensing temperature, T con ¼ 30  C and a heat source temperature of 100  C, the results indicate that the system could achieve an SCP of the order of 104 W/kg, a refrigeration cycle COP of 0.43, and it could produce a daily useful cooling of 2515 kJ per 0.8 m2 of collector area, while its gross solar COP could reach 0.18. A dynamic heat and mass transfer model was established by the Wu et al. (2009) based on the LDF model.7 The model was solved using the finite difference method, and the performance of the adsorber and condenser/evaporator of the module was analyzed. The calculated results were validated with experimental data and good agreement was observed. A solar powered two-bed adsorption cooling cycle with heat and mass recovery was modeled for silica gel–water

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pair by Luo et al. in 2010.26 The COP values were predicted using energy balances with a maximum bed temperature range of 55-90  C. The results predicted that a COP of the system higher than 0.25 can be achieved under fair solar radiation and with an evacuated tube collector. Results also stated that the evacuated tube collector always has higher thermal efficiencies than the flat plate collector for the investigated cases. A silica gel–water adsorption chiller integrated with a closed wet cooling tower was developed and numerically studied. 27 A transient one-dimensional model was proposed and validated using experimental data. The predicted results agreed well with the experimental measurement. Based on this reliable mathematical model, the following conclusions were achieved: the cooling capacity and COP of the chiller are 10.76 kW and 0.51, respectively, the cycle mass of heat and mass recovery cycle is increased by 34.9%. A forward-looking study on the operating characteristic of silica gel–water adsorption chiller driven by solar energy was introduced by Zhang et al. (2011) with the help of a lumped parameter model. 28 MATLAB – SIMULINK, as a high-performance computing and programming tool, was used to simulate the operating characteristics of the chiller. The effects of the hot water tank capacity, the cycle time, and the initial hot water temperature on the performance of the chiller were analyzed when the chiller was driven by a stable heat source and solar energy, respectively. The simulation results indicated that when the chiller was driven by solar energy, the open circulation of the hot water with a short cycle time and the closed circulation of hot water with a longer cycle time were better. The average refrigerating capacity and the average COP in a day are about 6.04 kW and 0.458, respectively. The adsorption refrigeration system mainly depends on the heat and mass transfer in the inside of the system. In this view, Hassan et al. (2011) represented a mathematical model using activated carbon methanol (ACM) pair and it stand for heat and mass transfer in the adsorption bed, the condenser, and the evaporator. 29 The simulation technique takes into account the variations of ambient temperature and solar radiation along the day. Furthermore, the local pressure and local thermal conductivity variations in space and time inside the tubular reactor are investigated as well. It found that, the solar coefficient of performance and the specific cooling power of the system were 0.211 and 2.326, respectively. In the same manner, Zhao et al. (2011) was validated a dynamic mathematical model under the non-uniform pressure a ssumption and the introduction of a transient boundary condition of the diffusion equation.30 The experimental data and numerical results were compared in terms of temperature development inside the carbon bed, with the transient and two popular simplified boundary conditions of vapor density, respectively. The comparison showed that the transient boundary condition improves accuracy of the model, and more importantly, it is capable of reflecting the dynamic shift of dominant driving forces of the adsorption process inside the generator, i.e., shifting to temperature driving gradually from diffusion driving. Various theoretical results have been discussed in the above section and it also gives a tool to optimize of adsorption systems driven by solar heat or other low grade heat source temperature. IV. CHOICE OF ADSORBENT–ADSORBATE PAIRS Adsorption refrigeration is achieved using a combination of adsorbate and adsorbent. Significant improvements of such system performances may be realized through using of different adsorbent and adsorbate materials. The selection of any pair of adsorbent-adsorbate combination depends on certain desirable characteristics of their constituents, including the affinity for each other. These characteristics range from their thermodynamic and chemical properties to their physical properties. A number of studies have been carried out, both experimentally and theoretically for selection of adsorbent-adsorbate material; but the cost for selection of adsorbent–adsorbate materials still makes them non-competitive for commercialization. Therefore, some investigations are focused on cost reduction and on increasing the efficiency of the machines. This section summarizes the most widely used working pairs are zeolite–water, activated carbon–methanol, silica gel–water, and activated carbon–ammonia. In addition, the performance of adsorbent-adsorbate pair is reviewed based on heat source temperature.

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A. Zeolite–water The development of sorption refrigeration systems powered by zeolite–water emerged in 1978 following the pioneering work of Tchernev.31 Then, in 1982, significant efforts were made by Tchernev to develop a solar-powered natural zeolite-water system. Natural zeolite, with a heat of adsorption of 2.8 MJ/kg, was found to be more suitable for cooling systems than the synthetic zeolite having a heat of adsorption of 4.2 MJ/kg. This is due to the fact that with lower heat of adsorption, the adsorbent bed can be cooled more effectively, which in turn results in better adsorption of refrigerant. In the same manner, Tchernev (1982) fabricated and tested a 100 dm3 zeolite–water solar powered refrigerator. The system contained 50 kg of zeolite per square meter. For a solar energy input of 6 kWh, the refrigerator produced 900 Wh of cooling per square meter of collector area with a COP of 0.15. At the end of the 1980 s, in France, Grenier et al. (1988) (Ref. 32) designed a solar adsorption air conditioning system using zeolite-water working pair under mass of 22 to 15 kg/m 2. This system was designed to refrigerate a 12 m 3 room for food preservation with a collector area of 13.10 m2. In this study, the collector temperature varied from 93 to 103  C correspondingly the COP of the system varies from 0.10 to 0.11, respectively. The temperature difference of 20  C between the outside ambient and the cold room was varied. The adsorbent bed is filled with 140kg of 13 Â Zeolite grains and 185 kg of water was charged into the system.33 The cooling power ranging from 3.0 to 4.2 kW had a COP of 0.21 under the suitable heat source temperature of 220–250  C, which is fair enough to make the comfort level. In an innovative approach, a truck air-conditioning driven by engine waste heat was presented by Dong Wu (2011). 35 The working pair of 13X Zeolite–water was used with the mass of 45 kg and 10.50 kg, respectively. Under the typical condition of the heat source temperature of 325  C, the value of COP and SCP namely 0.26 and 95.0 W/kg, respectively. A thermally powered prototype adsorption cooling system using natural zeolite–water was investigated by Ismail Solmus in 2011. 34 The mass and volume of working pair were 1.87 kg and 2.5 liter. The mean COP and SCP values of the experimental prototype were 0.25 and 6.4 W/kg, respectively for the heat source temperature of 150  C. Form the above summarization, the present investigation found out that zeolite have significantly larger quantities and high temperature would be required if water is the adsorbate since only a modest amount of adsorbate is desorbed in going from room temperature to flat plate solar collector temperature. Table I shows some of the achieved performance of solar adsorption refrigeration system. As a result, the zeolite–water pair is not suitable for flat plate solar collector adsorption refrigeration system. However, zeolites have another unique property in that their adsorption isotherms have extremely nonlinear pressure dependence, as reported by Tchernev (1982), which is of important criteria in solar refrigeration applications.

B. Activated carbon–methanol Activated carbon–methanol is one of the most widely used working pair in adsorption system because of their large cyclic adsorption capacity, low desorption temperature, low adsorption temperature, and high latent heat of methanol. In this way, Table II shows the some appraisal results of activated carbon-methanol refrigeration system. In this concern, last twenty four years back Pons and Guilleminot (1986) studied a prototype solar powered ice maker with an activated carbon/methanol pair.36 The solar collector area of 6 m 2 contains on the whole, 130 kg of activated carbon produced the net COP was 0.12 for the maximum heat source temperature of 95  C, which reported this machine one of the most efficient solar ice maker in 1986. Li and Sumathy 37 was loaded 17 kg of activated carbon inside the collector to form the adsorption bed, which is a little less than the specific loading reported by Pons and Guilleminot in 1986. At last, the system achieved the COP of about 0.1 to 0.12. In a novel approach to increase the overall efficiency of the adsorption icemaker assisted with prototype hybrid system was designed by Wang et al.14 The activated carbon and methanol was laden with 22 kg and 25 kg, respectively, for the heat source temperature of 98  C to

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91  C, corresponding values of COP increase from 0.38 to 0.43. The daily ice production was about 10 kg when the insolation was about 22 MJ/m2, which is higher than Wang et al. (1998) earlier study. 38 A forward-looking way, the temperature gradient within the adsorbent bed and cooling performance were analyzed. 39 These authors deduce that in order to improve the performance of this system, the heat transfer properties of the adsorber needed to be enhanced. This could be achieved by increasing the number of fins or using consolidated adsorbent. Accordingly, this ice-maker had a COP reached for 0.42 to 0.45 and produced between 5 to 6 kg of ice per m 2 of collector when charged with 22 kg o f activated carbon. This study is much improvement than Pons and Guillminot (1986) analysis. 36 An adsorption ice-maker was tested by way of 40 kg of AC and 7.0 l of methanol40 at Burkina Faso by Buchter et al. (2003) for the heat source temperature from 29–100  C. The outcome of this prototype was compared with similar system of previously studied by Pons and Grenier in Orsay (1987) (Ref. 82) and Boubakri et al. (1992) in Morocco.41,42 The machine (Burkina Faso) presented a cooling performance of about 28% and 35% higher than that of the machine tested in Orsay and Morocco, respectively. The limitation of the system was mainly due to the ambient temperature, an increase in the ambient temperature beyond 23  C reduced the system performance significantly. Wang et al. (2003) tested two different types of adorbers in which two types of bed are filled with compressively solidified activated carbon and another one is filled with granular activated carbon (AC).43 The authors finally reported that the performance of solidified AC adsorber is much better than that of granular AC adsorber since the heat transfer coefficient is much better than granular AC. Accordingly, COP and SCP of the granular AC were 0.125 and 16 W/kg correspondingly for the granular. Thus, granular AC demise the COP and SCP were 0.104 and 13.1 W/kg, respectively, for the heat source temperature of 110  C. Ferreira Leite presented 44 characterization and pre-dimensioning of an adsorption chiller in which each adsorber consist of with 25 kg of activated carbon and 72 kg of methanol, respectively. The effective COP was obtained to be 0.5, which is relatively low, in relation to that of conventional vapor compression systems. C. Silica gel–water Silica gel is a partially dehydrated form of polymeric colloidal silicic acid. The adsorptive property of silica gel arises from its high porosity and its pores are sub-microscopic scale approximately 20$200. The silica-gel-water is an alternative method, allowing cycles to be driven by relatively lower driving heat-source temperatures. Table II illustrated the previous researchers achieved performance results of siligca gel–water adsorption refrigeration system. Hildbrand et al. (2004) focused a high performance new solar powered adsorption refrigerator.45 This final result shows COP varies between 0.10 and 0.25 with a mean value of 0.16. These values are higher than those obtained by earlier solar powered refrigerators (0.10–0.12). The adsorber consists of 12 parallel tubes (72.5 mm in diameter) that contain 78.8 kg mass of silica gel. Based on the above research, Liu et al. (2005) developed an adsorption46 chiller could be regenerated by hot water between 75  C and 90  C. The whole chiller contained 52.8 kg of silica gel divided between two adsorbent beds, which operated out of phase and thus, produced constant cooling. Experiments with the first prototype showed that a cooling power of 3.56 kW and a COP of 0.26 could be obtained when the mass and heat recovery processes were employed under the following operating conditions: an evaporation temperature of 7.0  C, a heat sink temperature of 28  C, and a heat sour ce temperature of 85  C. Novel silica gel–water adsorption chillers47 with two single bed systems were built and tested in Shanghai Jiao Tong University (SJTU). Each adsorber contains 52 kg of silica gel. As the results, the refrigerating capacity and the COP of the chiller are, respectively, 8.69 kW and 0.388 for the heat source temperature of 82.5  C. There is an improvement at least 12% for the COP compared with the former chiller (2005). In the same way, contemporary research on silica gel water adsorption chiller was built in SJTU in 2009. The performance of the chiller

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TABLE I. Some achieved performance of solar adsorption refrigeration system. Authors

Mass of adsorbent-adsorbate pair

COP

50 kg/m2

0.15

Tchernev 31 Grenier et al. Lu et al.

32

33

Ismail Solmus

2

Dong Wu35

—

15-22 kg/m

0.10

93 C to 103  C

2

0.21

220  C–250  C

2

0.25

150  C

0.26

325  C

140kg/m 34

Heat source temperature

1.87kg/m

45-10.5 kg/m2



was tested with 104 kg mass of silica gel with driven heat source temperature of 82  C and compared with former adsorption chiller. The results show that COP increases to 0.43 while the cooling power is about 11.0 kW. Compared with that of the former chiller, the present system COP is increases by 20%. In an innovative approach, a compact adsorption chiller integrated with a closed wet cooling tower was 27 designed by Chen et al. (2010). The system performance was investigated against the heat source temperature of 85  C with 65 kg of silica gel and 169 kg of water for corresponding COP and SCP were 0.51 and 10.76 Kw, respectively. One more adsorption chiller without vacuum valve was manufactured and 48 experimentally studied by Chen et al. in 2010. Fin-tube heat exchangers are adopted as adsorbers. Each adsorber contains 27 adsorbent units. Each adsorber contains 48.5 kg of silica gel. The cooling power and COP were 9.60 kW and 0.49, respectively, when the heat source te mperature of 82.0  C. The present investigation compared to Chen et al. (2010) existing system 27 based on mass of adsorbent, present system showed better efficiency. Luo et al. (2010) designed and constructed26 a solar powered adsorption chiller with a mass of silica gel 112 kg. The chiller was produced a cooling power of about 66-90 W/m2 and daily solar cooling COP is about 0.1–0.13 for a heat source temperature of 65-85  C. Farid et al. (2011) deal with the performance investigation of a silica gel/water based on two-stage, four-bed adsorption chiller with different mass allocation 49 whereas heat source temperature of 90  C. The COP improvement ratios for the mass allocation 3:2 are 4.618%, 3.253%, 2.808%, 0.518%, and 1.752% (decreasing). In addition, the COP improvement ratios for the mass allocation 2:1 are 7.07%, 5.13%, 4.04%, 0.71%, and 4.94% (decreasing). But the COP decreasing ratios for the mass allocation ratio 2:3 are 6.31%, 4.87%, 3.09%, 0.42%, and 7.53% (increasing). A low-grade waste heat driven silica gel-water adsorption chiller was successfully designed50 and tested by Grisel in (2010). The average cooling power was 3.6 kW, COP was 0.62, and the power density was 17 kW/m3 for the system as whole. D. Activated carbon–ammonia Explorations in the use of charcoal–ammonia are apparently more recent mainly during the 1990 s. Jones (1993) used a novel carbon-moulding technique and incorporated a thermal waveregeneration concept; a small unit consuming 0.51 kg of charcoal produced 293 W cooling with TABLE II. Adsorbent–adsorbate pair used in the solar adsorption refrigeration system. Adsorbent-adsorbate pair

System COP

Reference

Silica gel–water

0.16

45

Silica gel–water

0.10 À0.13

85

Activated carbon-methanol

0.1 À0.12

36 and 83


0.1 À0.12

41 and 42


0.1 À0.12

86


0.16

87

0.14 À0.16

71 and 88

0.05

89

Domestic charcoal–methanol Activated carbon-ammonia

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ammonia as the adsorbate.51 A small solar adsorption refrigerator was built 52 and tested under preliminary level by Critoph (1994). The collector contains 17 kg of active carbon and 1.60 kg of ammonia was tested for 150  C heat sources. It produced up to 4 kg of ice per day in a diurnal cycle. In this investigation, COP was varied between 0.33–0.24. Similarly, a low cost rotary regenerative adsorption air conditioning system53 using monolithic carbon–ammonia was built by Tamainot-Telto and Critoph (1997). The total mass of monolithic carbon in the generator is about 40 g and 16 g of ammonia refrigerant, respectively. The typical COP of the system varied from 0.20–0.50. Mass recovery can play an important r ole to better the performance of an adsorption refrigeration pair of activated carbon–ammonia.54 The influence of mass recovery is mainly dominated by the sorption characteristics of the pair. The total volume of the ammonia was about 500 liter and 31.7 kg of activated carbon. The COP of the activated carbon/ammonia adsorption refrigeration cycle might be were varied from 0.09-0.12 with the mass recovery process due to different working conditions. Fadar et al. (2009) proposed a novel solar adsorptive cooling system coupled with a parabolic trough collector. 55 The authors were used 5-30 kg of activated carbon and its performance was tested for the heat source temperature of 70–170  C. From the investigation, the authors reported that, COP increases with increases in adsorbent mass and once the adsorbent mass reaches a critical value (146 kg of AC), the COP decreases. The reason is that the increase in adsorbent mass induces adsorption of high quantity of ammonia initially at adsorption phase and hence, desorption of large amounts of ammonia in subsequent desorption process. This produces more cooling and consequently, results in high COPs. Nevertheless, beyond this optimal value of 14.5 kg, the adsorbent bed is heated, but the heat adsorbed is not sufficient to desorb the required amount ammonia. In order to develop a consolidated solidified adsorbent with better heat transfer 56 and mass transfer performance was developed and studied by Wang et al. (2011). For this investigation, solidified AC and granular AC was used. It stated that granular activated carbon was better than that of solidified because of the reduced mass transfer of ammonia at low saturated pressure. The working pair of activated carbon–ammonia was developed 57 by Miles and Shelton in 1996. The COP obtained varied between 1.19 and 0.42, for ambient temperatures from 20  C to 35  C. Another kind of heat regenerative cycle was suggested by Critoph (1998), in which the refrigerant served as a heat transfer medium.58 A thermodynamic model presented by the author predicted a COP of 0.95 with an evaporator temperature of 0  C and a condensation temperature of 42  C. Wade et al. (1989) simulated a regenerative adsorption cycle with eight beds that could recover 76% of the waste energy from the adsorption process. 59 The previous work was the base for the design of a prototype with four beds that was used to produce cooling at À136  C.60 The energy input necessary to produce 1.0 W of cooling in this prototype was 76.6 W, which was much lower than the 165 W required in the system without heat regeneration. However, the difficulties and practical disadvantages of activated carbon–ammonia systems are the high pressure requirement, resulting in the bulkiness of the refrigerator, and the corrosive nature of the refrigerant, ammonia. The problem of the great bulk of large systems can be avoided by the development of rapid cycling units. V. EXPERIMENTAL STUDIES OF ADSORPTION COOLING SYSTEM The design of solar collectors is a significant factor for solar cooling applications. It depends on the level of temperature required by the process. For example, an efficient waterammonia absorption solar refrigerator requires a temperature of collector around 150  C. This can be performed generally by collectors which are either parabolic or with evacuated tubes. 61 The Table III. demonstrated the selected performance attainment of different collector systems. Adsorptive units require a temperature of the collector around 100  C; these can be easily achieved by flat-plate collectors, which is an advantage. Consequently, selection of solar collector is a vital role in the adsorption refrigeration system. In this context, this section review is based on the dimension and design aspect of the solar cooling system.

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A. Flat plate collector At the end of the 1980 s, Grenier et al. (1988) studied a solar adsorption air conditioning system that had used 20 m2 of solar flat plate collector.32 A single glass cover was used 3 cm ahead of the front face of the collectors. The system has attained maximum heat source temperature of 130  C. This system was designed to refrigerate a 12 m3 room for food preservation and the cold room could store 1000 kg of vegetables with a rotation of 130 kg per day for a temperature difference of 20  C between the outside ambient temperature and the cold room when insolation received by the solar collector was about 22 MJ/m2. The COP in this case was 0.10. In an effort was made to utilize solar heat, Sakoda and Suzuki (1986) achieved a solar COP of about 0.2 with solar flat plate collector dimensions of 500 Â 500 Â 50 mm.62 Attempts were taken to increase the heat transfer in the collector was divided into four blocks by ten heat transfer fins of 50 mmÂ 50 mm square and 0.5 mm thick. This kind of arrangement got maximum bed temperature of 80  C. In addition, present system performance was estimated based on changing various design parameters like effect of selective surface with and without a glass cover and selective surface with and without two glass cover. Under the above said condition, the COP was 0.32 and 0.30, respectively. To enhance the heat transfer inside the adsorbent bed, Li et al. (2002) arranged a number of fins in the two flat plate collectors. 39 Each collector was charged with 22 kg of activated carbon with a surface area of 0.75 m 2. The top surface of the adsorbent bed is coated with black paint to enhance receiving solar flux radiation. Within this type of setup, the adsorbent bed reaches maximum temperature of 110  C. The experimental result showed that for this kind of arrangement, the COP was fluctuated beyond 0.12–0.14 and approximately 5–6 kg of ice/m 2 of area. Several experiments were carried out by Li et al. (2002) with a different collector area of 1.5 and 0.5 m 2 viz. and achieved the COP of 0.14 and 0.147, respectively. In a creative way, Boubakri et al. (2000) was developed a collector–condenser technology by using flat plate collector. 15 The units are mainly composed of a single glassed collector-condenser connected by a flexible tube with an evaporator. The collectorcondenser (1 m2, 90mm thick, and 20 tilted) was made to two identical stainless steel shells. A grid holds 20 kg of activated carbon in the upper shell, which plays the role of solar collector. The rear shell of the collector–condenser plays the role of air cooled condenser with external fins (7.5 m2). In this experiment, the collector reaches the maximum temperature of 110  C. The author’s find out that this technology leads to a 40% reduction of adsorbent rearless. It leads to affect the system performance like (a) a better heating of the collector (b) a poor cooling. This poor cooling effect limits the improvement of daily ice production. Many parameters that affect the performance of solar refrigerator a nd ice mass, such as solar radiation, wind speed, ambient temperature, numbers of glass cover, 63 coating material etc., cited by Li and Wang (2002). Among these parameters, two parameters are most important; they are the number of glazings and the selective coating material. Usually, the increase of glazing number is limited by collector structure and dimension, in common practice, not more than two or three glazings. In this view, the authors found out that the effects of COP and ice mass when single glass cover and double glass covers are used, respectively, 0.116, 7.47 kg and 0.113, 10.51 kg. In addition, the effects on COP and ice mass when black coating material and selective coating material are used and obtained results were 0.116, 7.47 kg and 0.145, 9.37 kg, respectively. Into the accumulation with the effects on COP and ice mass when double glass covers and selective coating material are used simultaneously and got the results of 0.116, 7.47 kg and 0.193, 12.43 kg correspondingly. The authors finally conclude that in the practical design of solar refrigerator should use double glass covers and selective material to improve the characteristics of solar refrigerator. Instead of two valves are necessary to fulfill the adsorption and desorption processes of a basic refrigeration system, a no valve, flat plate solar ice maker was developed by Li et al. (2004) in China which performance was tested for the real solar radiation condition. 64 In this system, there are no any reservoirs, connecting valve, and throttling valve, and the structure of

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TABLE III. Design aspects of solar cooling system with typical achievements. Reference

Type of collector

Collector area

COP

Flat plate

0.4 m 2

0.40

32

Flat plate

20 m

2

0.10

67

Heat pump

2 m2

0.15

2

0.36

62

90 22, 46, and 65 2 36 and 83

Flat plate

2m

Natural gas or liquid petroleum gas

—

Evacuated tube

150 m 2

Flat plate

6m

0.40 2

0.35 0.1À0.12

the system is simple. The adsorbent bed was made of flat plate stainless steel box having a surface area of 1.0 m2 and also 19 kg of adsorbent was charged and selective coating was used in the system. The results showed that 6.0–1.0 kg of ice was produced under the condition of about 17–20 MJ/m2 solar radiation; the COP of this system was about 0.13-0.15. Finally, the authors suggested that no valve solar ice maker is expected to be economical in west of China in the near future. In no valve concept, Liu et al. (2005) designed for simplified construction of the adsorption refrigeration system. 65 The advantage of this system is there is no refrigerant valve. This feature reduced the cost of the chiller and made it more reliable as there were minimum moving parts that could lead to air infiltration. The whole chiller contained 52.8 kg of silica gel divided between two adsorbent beds, thus produced constant cooling power of 3.56 kW and a COP of 0.26 were obtained under the following operating conditions: an evaporation temperature of 7.0  C, a heat sink temperature of 28  C, and a heat source temperature of 85  C. Louajari et al. (2011) presented the performance coefficient of the machine according to with and without fins in the solar flat plate collector. 66 The COP increases from 7.5% corresponding to a tube without fins to 11.1% which corresponds to a tube with fins. Owing motivation for that cycled mass of the refrigerant is also higher than the one of the tubes without fins. It results with fins in the solar collector increases COP of the system. In order to improve the heat transfer effectiveness in the adsorption bed, finned tubes were used instead of a simple flat-bed collector (Dieng and Wang, 2001), which had an outer diameter of 53 mm and a length of 0.2 m.67 The system was tested at various operating conditions, when the condensation temperature and regeneration temperature were 20  C andÀ5.0  C, respectively. In an another way, Anyanwu and Ezakwe 68 also designed a combined flat-plate type collector/generator/adsorber technology whose effective exposed area was 1.5 m 2 with the efficiency of 11.6%À16.4%. The maximum temperature was attained in the collector about 109  C with an area of 1.2 m2. The useful overall COP ranged over 0.007–0.015. Thus, only 2% of the incident solar energy was converted into the refrigeration effect. B. Compound parabolic collector Solar adsorption cooling systems are usually based on the flat plate collector, whereas little attention has been paid to concentrate collectors by Tamainot Telto and Critoph. 53 It consists of two compound parabolic collector (CPC) reflectors, two adsorbers, and a cover. The width, aperture, and height are, respectively, 420 mm and 658 mm. The concentration ratio of the collector was 2.37. Each absorber was made with 2 m long stainless steel tube and is covered in selective surface. The two adsorbers contain about 400 aluminium fins and about 17 kg of granular carbon. The maximum collector temperature reaches and cooling power value of 173  C and 120 W, respectively, under the solar radiation condition of 820 W/m2. Manuel et al. (2007) presented a CPC collector which separated by four tubular receiver contains the sorption bed using 13.8 kg of activated carbon and where only a portion of the

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receiver is exposed to sunlight.69 The total collection area was 0.55 m2. The prototype was tested during the springer and summer of 2005 in Burgos. The experimental results showed that solar COP range from 0.078 to 0.096. The maximum and minimum bed temperatures were 116  C and 38  C, respectively. In order to increase the desorption temperatures and good cooling effect during the adsorption period at night, Headley et al. (1994) constructed a charcoal–methanol adsorption refrigerator powered by CPC concentrating collectors of concentration ratio 3.9 and aperture area 2.0 m 2 was used, but the solar COP achieved was very low,70 of about 0.02, due to the reason the temperatures in excess of 150  C are undesirable since they favour the conversion of methanol to dimethyl ether, a noncondensable gas which inhibit both condensation and adsorption. There was an excessive heating capacity in the system and only 2% of the incident solar radiation was converted to the refrigeration effect. The author reported that system was not economically viable. Khattab (2004) designed an innovative approach to heat the adsorption bed via reflectors. 71 The testing of the module is mainly focused on the sorption bed, therefore, four types of bed techniques and four reflector arrangements to heat the sorption bed were proposed and tested under climatic condition of Cairo (30  latitude). The angles of inclination of the reflectors are varied every month to receive maximum solar energy at noon time. Testing module as a refrigerator realize daily ice production of 6.9 and 9.4 kg/m2 and net solar COP of 0.136 and 0.159 in cold and hot climate, respectively. The study has put in evidence the ability of such a system to achieve a promising performance and to overcome the intermittent of the adsorption refrigeration systems 55 driven by parabolic trough collector developed by Fadar et al. in 2009. Under the climatic conditions of daily solar radiation being about 14 MJ per 0.8 m 2 area of parabolic trough collector could achieve an SCP of the order of 104 W/kg, a refrigeration cycle COP of 0.43, and it could produce a daily useful cooling of 2515 kJ per 0.8 m2 of collector area, while its gross solar COP could reach 0.18. In the same way, the author tried to increase the efficiency of the system as a result of parabolic trough collector 55 coupled with heat pipe (Fadar et al. 2009). In the investigation, it was put in evidence that the system is optimum (COPs ¼ 0.18), when the adsorber external radius and aperture width of collector are of the order of 14.5 and 70 cm, respectively. So, the results show a promising performance in comparison with before published data, which were obtained with intermittent adsorption refrigerators. In addition, it does not need any moving parts or external pumping. C. Evacuated tube collector A different approach to increase the overall efficiency of the adsorption icemaker was studied by various researchers for the past years. In this way, Niemann et al. (1997) designed and constructed an ETC collector coupled with an external parabolic circle concentrator (PCC) to operate a large size adsorption refrigeration system. 72 The envisaged collector field consists of evacuated tube collectors 1.6 m2 area and the concentration ratio of 5.0 was mounted in an east-west direction. The collector fluid temperature varied from 0–170  C. Finally, the author reported that the collector area was needed to produce 500 kg of ice per day amount to 140 m 2 of PCC collectors, whereas for the ETC system, an area of 205 m2 was needed. Wang et al. (2000) presented a new hybrid of solar powered water heater and adsorption ice maker with an ETC solar collector area of 2 m2 which can produce the COP of about 0.15–0.23 and heating efficiency of about 0.35–0.38.14 The specialty of the proposed system was adsorption bed separated and immersed into a water bath which was powered directly by vacuum tube solar collector. Hence, no thermal insulation or enhanced convection is needed for the adsorber and also which guarantees either good heating or cooling of the adsorber. In this arrangement, adsorption bed reaches the temperature of 80–100  C. This hybrid system was provided 50 kg of water at 84–100  C for domestic purpose. In order to reduce the heat loss in the adsorber 73 and to improve the COP, ETC was used by Li et al. (2003). The performance of the cooling system was analyzed based on several design parameters like effect of the diameter of the evacuated tubes, the effect of distance between two adjacent tube centers, and effect of mass quantities of adsorbent. It is found out

022701-17



that the greater the diameter of ETC, the better performance but when the distance of is larger than the diameter, the cooling capacity decreased. In large level, air conditioning system was developed by Zhai and Wang (2010) to be powered by ETCs with heat source74 temperatures close to 100  C. The system had two adsorbers with 26 kg of carbon in each one and methanol used as refrigerant. The COP and the SCP of this system were significantly influenced by the area of ETC. The presented system with an ETC area of 80 m2 lead to a COP of 0.32 and a cooling power of 8.2 kW while operating with an ETC area of 240 m 2 lead to a COP of 0.37 and cooling power of 16.4 kW. To improve the performance of the system, the authors used a tube and plate heat exchanger between the plates. With this new design, the improved COP obtained was 0.4 and the cooling power was 16.57 kW. The experimental conditions in this case were: a heat source temperature of 100  C, evaporation temperature of 10  C, and a condensing temperature of 24  C. Nkwetta and Smyth in 2011 suggested that concentrated augmented evacuated tube heats pipe array of solar cooling applications. 75 Normally, non-concentrated evacuated tube heat pipe solar collector had been reported to show high fluid temperature with improved thermal performance in the low to medium temperature range ( <60  C) due to low heat losses but suffer higher heat losses in the medium to higher temperature range ( >80  C) which reduces their efficiency compared to concentrated evacuated tube heat pipe solar collectors. The proposed system was attaining temperatures in the range of 70–120  C. VI. ECONOMIC ANALYSIS OF SOLAR ADSORPTION REFRIGERATION SYSTEM When assessing the performance of a refrigeration system or heat pump, its COP value is usually considered, this undoubtedly can account for the economic character well in the system powered by electricity or oil-boiling. But, if it is powered by solar energy, more attention will be paid to the ratio of the cooling capacity to the costs of materials and manufacture. Therefore, another assessing index is used here, and according to it, the optimum structure of the system is determined. The defined parameter is SPA, 73 SPA ¼ Qf =ðC Â AÞ; where Qf is the cooling capacity, C is the total cost per unit area, and A is the area of the adsorber. Obviously, a larger parameter SPA may have a better economic character and performance. VII. PRESENT STATUS OF ADSORPTION SYSTEM AND FURTHER SCOPE OF RESEARCH Recent advances in solar adsorption cooling technologies make it feasible to convert sunlight into cooling power, although net COP is currently in the range of 0.3 À0.6. In this regard, a major leap has been realized in the net COP for solar cooling by introducing (Gordon and Kim Choon 2000) solar fiber optic mini-dish concentrators that offer a significan t increase in collecting efficiency while retaining the possibility of high temperature delivery. 76 This new technology can be useful to overcome the inherent limitations for adsorption cooling applications. In addition, the bulky size is the major hurdle in the solar cooling system. So, compactness of the system is very important in the solar cooling system. In this view, no valve solar ice64 maker was built by Li et al. 2004. For this system, there are no any reservoirs, connecting valves, or throttling valve, and the structure of the system is very simple. Although solid adsorption refrigeration has received much attention from most of the refrigeration companies and laboratories around the world, some of the main challenges such as long adsorption/desorption time and small refrigeration capacity per unit mass of adsorbent have become obstacles for the real mass production of the system. Finally, present investigation suggested that these two obstacles rectified by using activated carbon fiber. Since activated carbon fiber would have the higher adsorption capacity due to its larger surface area, total pore volume, and average pore diameter properties. Saha et al. (2006) compared the adsorption 77 capacity of activated carbon

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with activated carbon fiber (ACF), the ACF adsorption capacity is very high 0.797kg/kg compared with others. This ACF is recommended for short adsorption–desorption cycle time. VIII. CONCLUSIONS In this paper, an extensive review of the technologies related to the better utilization of solar energy for the production of cooling energy is presented. Among the various types of refrigeration systems, emphasis is made on the design concepts, the different adsorbent-adsorbat, and the different aspects of model concepts. The above described literature reveals that the performance of various adsorption systems varies over a wide range. Most of the works are experimental in nature under varying operating conditions. Some of them use sophisticated solar collectors with concentration, 70,78,79 others used the adsorbent itself, contained in a transparent tube, as the solar energy absorbing material,78,80 but the most efficient configuration seems to consist of metallic flat-plate solar collectors, single- or double-glazed, covered with a selective surface, filled with the adsorbent bed and evacuated tube solar collector. 14,32,41,42,52,81 – 83 In this connection, the type of solar collectors is an important factor for the adsorption refrigeration system. It depends on the level of temperature required by the process. An efficient water-ammonia adsorption solar refrigerator requires a temperature of collector around 150  C. This can be done generally by collectors which are either parabolic 61 or with evacuated tubes. Nevertheless, most of the adsorptive units required a temperature of collector around 100  C. This can easily powered by flat-plate collectors, which is an advantage. Another concern in the performance of the adsorption cooling system is mainly depends on the working pairs. A well designed system should have the characteristics of large adsorption capacity, large change of adsorption capacity with temperature variation, and more flat desorption isotherm and refrigerant fluid whose choice depends on the evaporator temperature. It must possess high latent heat of vaporization and small molecular dimensions to secure an easy adsorption. Since the refrigeration system will operate at a temperature largely below 0  C, thus methanol seems to be a good adsorbate because it can evaporate at a temperature largely below 0  C, its enthalpy of vaporization is high, its molecule is small enough to be easily adsorbed into micropores, its working pressure is always lower than the atmospheric one, which means a safety factor in case of leakage. As a result activated–carbon methanol is the most widely used adsorbent reported in the literature due to its extremely high surface area and micro pore volume. The operating temperature is one of the limitations that constrain the operation of an activated carbon–methanol pair cooling system. At temperature more than 150  C, the methanol decomposes into dimethyl. Yuan84 got very low cooling performance, due to the decomposition products may be one of the main reasons for the diminishing of solar cooling performance. As a result, flat plate collector is enough will achieve for nominal cooling. However, no need for expensive solar collector systems like concentrating and evacuated tube collector. Silicagel–water systems require the operation under vacuum condition which poses a hurdle. In addition, the practical difficulties in water are precluding using it at an ice-making temperature. So, it is a more correct choice for air conditioning purposes only. Zeolite–water pair requires a regeneration temperature of above 200  C, activated carbon–ammonia pair also requires more than 150  C for its regeneration. These temperatures are not obtainable by simple flat plate collectors or evacuated tube solar collector systems. So it is achieved only by solar concentrating collectors. It is suggested that this type of solar collectors increases the initial cost of the system. Although activated carbon methanol pair also works at a low regeneration temperature, but it is more suitable for ice production and freezing the application. Comparing with activated carbon–ammonia system, activated carbon methanol system is a vacuum system and that is safer than high–pressure system though not so reliable if a leak exists. Activated carbonammonia adsorption system operates in the high–pressure state, and ammonia is a corrosive and flammable refrigerant. A comprehensive theoretical analysis to assess the performance of adsorption system is important. It also ensures successful operation of such design concept and the important role of the adsorbent–adsorbate materials. In addition, these systems also require good thermal

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insulation as well as provision to increase convection current during the adsorption period. These factors dominate in dictating the performance of the system. However, the point to be observed is that the heat transfer and mass transfer inside the bed were not effective due to the thickness of the bed being very deep. Finally concluded that activated carbon-methanol is an ideal working pair of solar energy because of its high COP and low generation temperature, low freezing point, and no corrosion of copper. In addition, maximum adsorption capacity of the activated carbon methanol pair is very good compared to the other adsorbent material. ACKNOWLEDGMENTS This research was supported by the MNRE, Government of India under PDRF-National Renewable Energy fellowship program at IIT Delhi. Nomenclature M ¼ Mass of dry adsorbent (kg) Cp ¼ Specific heat (kJ/kg) Qih ¼ Heat of isosteric heating process (kJ) Qide ¼ Heat of isobaric desorption (kJ) Qic ¼ Heat of isosteric cooling (kJ) Qiad ¼ Heat of isobaric adsorption process (kJ) Qrefri ¼ Refrigeration effect (kJ) Tg1 ¼ Temperature to start desorption (  C) Tg2 ¼ Desorption temperature (  C) Ta1 ¼ Temperature to start adsorption (  C) Ta2 ¼ Adsorption temperature (  C) Tevp ¼ Evaporation temperature (  C) DH ¼ Heat of adsorption (kJ) X ¼ Adsorption capacity (kg of refrigerant/kg of adsorbent) Xmax ¼ Adsorption capacity at adsorbed state (kg/kg) Xmin ¼ Adsorption capacity at desorbed state (kg/kg) ad bent ¼ Adsorbent ad bate ¼ Adsorbate ad mat ¼ Adsorbent bed 1

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