Comparative Study of Different Solar Cooling Systems for Buildings

Comparative study of different solar cooling systems for buildingsin subtropical city

K.F. Fong *, T.T. Chow, C.K. Lee, Z. Lin, L.S. Chan

Division of Building Science and Technology, College of Science and Engineering, City University of Hong Kong, Tat Chee Avenue,

Kowloon Tong, Kowloon, Hong Kong, China

Received 11 June 2009; received in revised form 28 October 2009; accepted 7 November 2009

Available online 8 December 2009

Communicated by: Associated Editor Ruzhu Wang

Abstract

In recent years, more and more attention has been paid on the application potential of solar cooling for buildings. Due to the fact thatthe efficiency of solar collectors is generally low at the time being, the effectiveness of solar cooling would be closely related to the avail-ability of solar irradiation, climatic conditions and geographical location of a place. In this paper, five types of solar cooling systems wereinvolved in a comparative study for subtropical city, which is commonly featured with long hot and humid summer. The solar coolingsystems included the solar electric compression refrigeration, solar mechanical compression refrigeration, solar absorption refrigeration,solar adsorption refrigeration and solar solid desiccant cooling. Component-based simulation models of these systems were developed,and their performances were evaluated throughout a year. The key performance indicators are solar fraction, coefficient of performance,solar thermal gain, and primary energy consumption. In addition, different installation strategies and types of solar collectors were com-pared for each kind of solar cooling system. Through this comparative study, it was found that solar electric compression refrigerationand solar absorption refrigeration had the highest energy saving potential in the subtropical Hong Kong. The former is to make use ofthe solar electric gain, while the latter is to adopt the solar thermal gain. These two solar cooling systems would have even better per-formances through the continual advancement of the solar collectors. It will provide a promising application potential of solar coolingfor buildings in the subtropical region.Ó 2009 Elsevier Ltd. All rights reserved.

Keywords: Solar cooling; Photovoltaic; Solar thermal technology; Solar collectors; Air-conditioning; Plant simulation

1. Introduction

In the field of heating, ventilating and air-conditioning(HVAC) engineering, electricity has been used to energizethe air-conditioning and refrigeration systems for a cen-tury. It is undoubted that HVAC is one of the indispens-able provisions in modern life, no matter in working orliving environments. Normally one-third to half of theannual total electricity consumption is used for air-condi-tioning and refrigeration in the metropolis worldwide.80% of electricity is still generated by burning the fossil

fuels, leading to the non-stopping emission of global warm-ing gases. In a typical subtropical city, like Hong Kong, thesummer is hot and humid, and the winter is generally tem-perate. Therefore air-conditioning is essential in a subtrop-ical city throughout a year, particularly for the commercialand institutional buildings. In the subtropical Hong Kong,air-conditioning and refrigeration are the largest area ofelectricity consumption, accounting for 33% in 2006 (HongKong Energy End-use Data, 2008).

Many facts show that the climate change is gettingworse. The changes of global surface temperature, globalaverage sea level, snow and ice over 1850–2000 arereported (IPCC, 2007). To have wider application ofrenewable energy is a crucial sustainable strategy to allevi-

0038-092X/$ - see front matter Ó 2009 Elsevier Ltd. All rights reserved.

doi:10.1016/j.solener.2009.11.002

* Corresponding author. Tel.: +852 27888724; fax: +852 27889716.E-mail address: [email protected] (K.F. Fong).

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Available online at www.sciencedirect.com

Solar Energy 84 (2010) 227–244

ate the climate change. Global warming has caused theincreasing trend of the environmental temperatures. Thisgradual change would affect the future air-conditioningrequirements in Hong Kong (Lam et al., 2008). Thedemand of air-conditioning would be increasing due tothe effect of climate change and global warming. If we stillrely on the conventional electric-driven air-conditioningbut the electricity is generated from the fossil fuels, thegreenhouse gas emission would continuously worsen globalwarming, in turn the demand of air-conditioning would befurther increasing. If this dead loop cannot be solved, even‘global heating’ would come one day. Therefore it is urgentto minimize the consumption of fossil fuels and promotewider use of solar energy, particularly in refrigerationand air-conditioning.

Solar cooling is envisioned as the sustainable means toprovide air-conditioning and refrigeration, since solarenergy is considered to be the primary energy source. Tech-nically, solar cooling is a feasible way to replace the electricrefrigeration machines for buildings. In recent years, more

reviews have been made about the feasibility of widerapplication of solar cooling technologies (Fan et al.,2007; Hwang et al., 2008; Kim and Infante Ferreira,2008). A number of demonstration projects have beenlaunched to gain more experience in the design and opera-tion of solar refrigeration and air-conditioning (Eicker,2003; Henning, 2004; Zhai et al., 2008). Air-conditioningis already part of our daily lives, solar cooling certainlyhas a positive and long-term impact. For the continualpopulation and economic growth, wider use of solar energyin air-conditioning would secure the increasing energydemand.

Solar cooling can be broadly categorized into solar elec-tric refrigeration, solar thermal refrigeration and solarthermal air-conditioning. In the first category, the solarelectric compression refrigeration uses photovoltaic (PV)panels to power conventional refrigeration machine. Inthe second category, the refrigeration effect can be pro-duced through solar thermal gain, the solar mechanicalcompression refrigeration, solar absorption refrigeration

Nomenclature

COP coefficient of performanceCOPab COP of solar absorption refrigerationCOPad COP of solar adsorption refrigerationCOPdesi COP of solar solid desiccant coolingCOPelect COP of solar electric compression refrigera-

tionCOPmech COP of solar mechanical compression refrig-

erationEp primary energy consumption (kWh)Ep,desi primary energy consumption of solar solid des-

iccant cooling (kWh)Ep,elect primary energy consumption of solar electric

compression refrigeration (kWh)Elp,therma

primary energy consumption of solar thermalrefrigeration (kWh)

Gelect solar electric gain (kWh)Gsolar solar thermal gain (kWh)Gsolar,eq solar thermal gain equivalent (kWh)ho specific enthalpy of outdoor air (kW/kg)hs specific enthalpy of supply air (kW/kg)m total number of time steps in a month for hot

water pump in operationm0 total number of time steps in a month for chiller

in operation_ma mass flow rate of supply air (kg/s)Qaux heat output of auxiliary heater (kWh)Qboil heat input to boiler of heat engine (kW)Qde heat input for desorption (kW)Qe refrigeration effect (kW)Qgen heat input to generator (kW)Qregen heat input for regeneration (kW)

Qsolar solar thermal gain from the solar collectors(kW)

SF solar fractionSFelect solar fraction of solar electric compression

refrigerationSFthermal

solar fraction of solar thermal coolingW power input to DC motor (kW)WAHU electrical energy consumption of AHU (kWh)Waux auxiliary electrical power from the city power

grid (kW)Wfans electrical energy consumption of supply air fan

and exhaust air fan (kWh)Wparasitic

electrical energy consumption of parasiticequipment, including pumps and cooling tower(kWh)

W para,desi electrical energy consumption of parasiticequipment, including pumps, desiccant wheel,heat wheel and evaporative coolers (kWh)

Wsolar solar electric gain from photovoltaic panels(kW)

y total number of time steps in a year for hotwater pump in operation

y0 total number of time steps in a year for chiller inoperation

Greek symbols

ge energy efficiency for electrical energy convertedinto primary energy

gg energy efficiency for gas energy converted intoprimary energy

228 K.F. Fong et al. / Solar Energy 84 (2010) 227–244

and solar adsorption refrigeration are the three commonoptions. In the third category, the conditioned air can bedirectly provided through the solar thermal gain by meansof desiccant cooling. Both solid and liquid sorbents areavailable, like silica gel and lithium chloride, respectively.However there is a risk of carry-over of the liquid sorbentinto the indoor space by the conditioned air, so it is moresuitable to apply the solar solid desiccant cooling. As aresult, the following five types of solar cooling systems wereincluded in this study:

� solar electric compression refrigeration;� solar mechanical compression refrigeration;� solar absorption refrigeration;� solar adsorption refrigeration; and� solar solid desiccant cooing.

The first one uses the solar electric approach, while therest apply the solar thermal technology.

The installation strategy and types of solar collectorswould directly affect the system performance since theamount of solar irradiation acquired would be different.The flat plate collectors and evacuated tubes are the com-mon types of solar collectors used in the solar thermal sys-tems. The parabolic concentrators are also getting morepopular in building applications. It is worthy to understandtheir performances in response to different system designand operating requirements based on the solar irradiation,climatic and loading conditions in the subtropical region.In the previous research works (Hwang et al., 2008; Kimand Infante Ferreira, 2008), simulation and experimentalstudy on a certain type of solar cooling system has beenpinpointed, but not across a variety of solar cooling sys-tems with different installation strategies and types of solarcollectors.

The rest of this paper is organized as follows: Section 2describes development of the simulation models of the fivetypes of solar cooling systems serving a typical local office.Section 3 addresses the framework of the comparativestudy of different scenarios, including different installationstrategies and types of solar collectors associated with thesolar cooling systems. Section 4 presents the definitionsof the performance indicators, including solar fraction,coefficient of performance, solar thermal gain, and primaryenergy consumption. Section 5 discusses the results of thepreliminary study for the installation strategies and typesof solar collectors. Section 6 discusses results of the fullcomparative study based on the performance indicators.The feasible choice of solar cooling systems would be iden-tified and recommended. Section 7 is the conclusion.

2. Simulation of the five solar cooling systems

2.1. General design information

In this study, each type of the solar cooling systems wasdesigned to serve a common typical office in the subtropical

Hong Kong located at 22.32°N and 114.17°E. The officewas single storey with floor area of 200 m2. The daily occu-pancy schedule was from 08:00 to 18:00, totally 10 h. Thewall-fenestration ratio was 0.5. The internal and externalshading factors of fenestration were 0.8 and 0.2, respec-tively. The internal heat gains included 24 persons seatedat work, 230 W/person of personal computer, and 17 W/m2 lighting with 30% convective part. The provisions andschedules of the internal heat gains were based on the localdesign practice (Code of Practice for Energy Efficiency ofAir Conditioning Installations, 2007). The fresh air flowrate was 10 litre/s/person. The design indoor conditionswere 25.5 °C and 60% RH, and the year-round weatherdata followed the Typical Meteorological Year of HongKong (Chan et al., 2006). The design entering hot water,condenser water and chilled water temperatures were90 °C, 30 °C and 13 °C, respectively. The chiller was sizedbased on the design entering condenser and chilled watertemperatures at 30 °C and 13 °C, respectively. The con-denser and chilled water flow rates were designed basedon the 5 °C change between entering and leaving tempera-tures. The component-based simulation program TRNSYS(2006) and its component library TESS(2006) were used tobuild and run the system simulations. The simulation timestep was 6 min. Table 1 summarizes the common parame-ters used in modeling different solar cooling systems.

2.2. Solar electric compression refrigeration

2.2.1. Model building

Generally the solar electric compression refrigeration,the air side system and the corresponding control provi-sions were built by using TRNSYS and TESS, as shownin Fig. 1. The major equipment included the PV panels,direct current (DC) motor, power regulator, vapour com-pression chiller, chilled water pump, cooling tower, con-denser water pump, air handling unit (AHU). A detailedchiller model was developed separately. The modelingmethod of DC motor was based on Hughes (1993), thatof condenser and evaporator based on Lee (2008), and thatof the compressor followed the parameter estimation tech-nique adopted by Jin and Spitler(2002). The total and spacecooling load were 29 kW and 20 kW, respectively. The gen-eral modeling parameters of the solar electric compressionrefrigeration is shown in Table 1.

2.2.2. Part-load control of chiller

The PV panels were designed to furnish the necessaryelectric power to drive the chiller. When the supply fromthe PV panels was insufficient due to the weather condition,a power regulator was used to draw auxiliary electricalpower from the city power grid. The power regulator wascapable of tracking the maximum power point of the PVpanels at any circumstances, so that the auxiliary electricalpower requirement could be minimized. To prevent fre-quent on/off of the chiller in low-load condition, a part-load controller for the chiller was used to modulate the

K.F. Fong et al. / Solar Energy 84 (2010) 227–244 229

set point of the power regulator between 1.5 and 7.5 kW infive steps. The set point started at 1.5 kW (the minimumrequired power input for the chiller) when the chilled water

return temperature of the chiller exceeded 12 °C, andstepped up if the water temperature dropped less than1 °C after one simulation time step. On the contrary, when

Table 1

General modeling parameters for the solar cooling systems.

Solar electric

compression

refrigeration

Solar mechanical

compression

refrigeration

Solar

absorption

refrigeration

Solar

adsorption

refrigeration

Solar solid

desiccant

cooling

Air handling unit

Supply air mass flow rate (kg sÿ1) 2.15 2.15 2.15 2.15 3.36

Supply fan power (kW) 2.15 2.15 2.15 2.15 3.36

Supply fan efficiency 70% 70% 70% 70% 70%

Face area of cooling coil (m2) 0.9 0.9 0.9 0.9 N/A

Chiller

Chilled water mass flow rate (kg sÿ1) 1.4 1.4 1.4 1.4 N/A

Condenser water mass flow rate (kg sÿ1) 1.9 1.9 1.6 1.8 N/A

Specific heat capacity of liquid water (kJ kgÿ1 Kÿ1) 4.19 4.19 4.19 4.19 N/A

Specific heat capacity of water vapour (kJ kgÿ1 Kÿ1) N/A N/A 1.866 1.866 2.028

Overall heat transfer value of evaporator coil (kW Kÿ1) 2.1 2.1 4.3 6.0 N/A

Overall heat transfer value of condenser coil (kW Kÿ1) 3.1 3.1 4.8 6.5 N/A

Condenser/cooling water system

Cooling tower air volume flow rate (m3 sÿ1) 1.17 11.11 2.22 2.78 N/A

Cooling tower fan power (kW) 0.36 3.42 0.683 0.855 N/A

Cooling tower fan efficiency 65% 65% 65% 65% N/A

Condenser water pump power (kW) 0.131 3.23 0.093 0.115 N/A

Condenser water pump efficiency 60% 60% 60% 60% N/A

Chilled water system

Chilled water pump power (kW) 0.084 0.084 0.084 0.084 N/A

Chilled water pump efficiency 60% 60% 60% 60% N/A

Hot/regenerative water system

Hot water mass flow rate (kg sÿ1) N/A 2.8 2.0 2.8 1.92

Hot water pump power (kW) N/A 0.233 0.2 0.233 0.192

Hot water pump efficiency N/A 60% 60% 60% 60%

Regenerative water mass flow rate (kg sÿ1) N/A 15 2.0 2.8 1.92

Regenerative water pump power (kW) N/A 1.55 0.247 0.341 0.141

Regenerative water pump efficiency N/A 60% 60% 60% 60%

General

Area of solar collectors (m2) 100 100 100 100 100

Minimum driving temperature (°C) N/A 82 67 64 60

Capacity of auxiliary heater (kW) N/A 323 42 69 89

Remark: N/A means ‘not applicable’.

DC

Motor

Condenser

Compressor

Evaporator

PV Panels

Expansion

Device

Cooling

Tower

-

Indoor Space

Supply

Air

Fresh

Air

Return Air T

Vapour

Compression

Chiller

Chilled Water Pump

Condenser Water Pump

AHU

3-way

Valve

PR

City Power

Grid

Power

Regulator

Cooling Coil

Controller

Part-load

controller

Thermostat with

Cooling Tower

Controller

T

T

C

C

Fig. 1. Schematic diagram of HVAC system designed with solar electric compression refrigeration.


the chilled water return temperature was below 12 °C, theset point would be stepped down if the water temperaturedropped more than 1 °C after one simulation time step.By employing the part-load control, the chiller capacitycould be adjusted through the input control of powerregulator.

2.2.3. Control and operation of chiller and associated

equipment

The chilled water entering the cooling coil of AHU wascontrolled by a three-way valve based on space tempera-ture. The cooling coil controller generated a linear signalfrom 0 to 1 corresponding to no to full flow to the coolingcoil when the space temperature varied from 24.5 to26.5 °C. To avoid instability when simulating the coolingcoil, the valve would open at least 30% when the space tem-perature was above 24.5 °C. The operation of the chillerwas determined by a controller that provided an on/off sig-nal between 12 and 9 °C with hysteresis based on thechilled water return temperature at the chiller. When thechiller was on, the condenser water pump would operate.The operation of the cooling tower would be controlledby a thermostat with hysteresis between 20 and 15 °C basedon the condenser water temperature leaving the chiller.Basically the chilled water pump and AHU were in opera-tion according to the daily occupancy schedule.

2.3. Solar mechanical compression refrigeration


Development of the solar mechanical compressionrefrigeration was based on the schematic diagram shownin Fig. 2. The essential equipment consisted of the solarcollectors, hot water pump, hot water tank, auxiliary hea-ter, regenerative water pump, heat engine using Rankinecycle, vapour compression chiller, chilled water pump,cooling tower, condenser water pump and AHU.

A number of validated models were involved in the solarmechanical compression refrigeration. Modeling approachof the Rankine cycle followed that proposed by Putten andColonna (2007). The performance of a multistage turbineat different operating conditions was approximated by Sto-dola’s ‘law of the ellipse’ according to Dixon (1998). Mod-eling methods of condenser, evaporator and compressorwere same as those for the solar electric compression refrig-eration in Section 2.2.1. The general modeling parametersfor the solar mechanical compression refrigeration areshown in Table 1, while the specific parameters are pre-sented in Table 2.

2.3.2. Control and operation of Rankine-cycle-driven chiller

The general provisions for the year-round control andoperation of the chiller and associated equipment weresame as those in Section 2.2.3 for the solar electric com-pression refrigeration. However, a specific considerationwas required for modeling the Rankine-cycle-driven chillerin order to ensure its year-round operation. Chiller opera-tion at an insufficient driving temperature for the Rankine

Condenser

Compressor

Evaporator

Chilled water supply and

return for the AHU serving the

indoor space (same as Fig. 1)

Cooler

Turbine

Heat Engine

(Rankine Cycle)

Vapour

Compression

Chiller

Feed

Pump Solar

Collectors

Expansion Device Boiler

Hot Water Pump

Regen.

Water

Pump

Hot

Water

Tank

T

T

C

Hot Water Pump

Controller

Cooling

Tower

T Condenser Water Pump

Town Gas

Supply

Aux.

Heater

Thermostat with Cooling

Tower Controller

Fig. 2. Schematic diagram of solar mechanical compression refrigeration.

Table 2

Modeling parameters for the Rankine cycle of solar mechanical compres-

sion refrigeration.

Parameter Value

Cooler water mass flow rate (kg sÿ1) 14

Specific heat capacity of circulating fluid (kJ kgÿ1 Kÿ1) 4.19

Density of circulating fluid (kg mÿ3) 1000

Volume of circulating fluid in boiler coil (m3) 0.08

Volume of circulating fluid in cooler coil (m3) 0.08

Volume of circulating fluid in feed pump and connecting pipeline

(m3)

0.004

Overall heat transfer value of boiler coil (kW Kÿ1) 50

Overall heat transfer value of cooler coil (kW Kÿ1) 50

Polytropic efficiency of turbine 80%

Mechanical efficiency of turbine 90%

Ratio of specific heat for circulating fluid 1.3


cycle would cause heat loss from the hot water side to thecooling water side with no cooling capacity produced. Toavoid this, the chiller would start only when the regenera-tive water temperature entering the boiler exceeded a presetminimum driving temperature, which was required by theRankine-cycle-driven chiller according to the design enter-ing cooling and chilled water conditions. This minimumdriving temperature was 82 °C in this case.

2.3.3. Control and operation of hot water pump and auxiliary

heater

Ahot water tankwas used to store the heat transferred bythe solar collectors. To ensure fulfillment of the cooling loaddemand even when the solar irradiation was insufficient,auxiliary heating was provided inside the storage tank as abackup heat source. The set point for the auxiliary heaterwas 90 °C. The temperature inside the hot water tank mightbe higher than that at the outlet of the solar collectors in caseof weak solar irradiation, then the solar collectors would dis-sipate heat to the surrounding. To prevent this, the hot waterpump, which was used to circulate hot water between thesolar collectors and the hot water tank, was controlled bythe hot water pump controller using a hysteresis loop con-trol. Whenever the temperature at the solar collector outletexceeded the water temperature inside the hot water tankby 5 °C, the hot water pump would be switched on untilthe temperature difference dropped to zero.

2.4. Solar absorption refrigeration


The entire solar absorption refrigeration was developedas shown in Fig. 3. The major equipment covered the solar

collectors, hot water pump, hot water tank, auxiliary hea-ter, regenerative water pump, absorption chiller, chilledwater pump, cooling tower, condenser water pump andAHU.

The model of absorption refrigeration cycle was con-structed according to the validated single-effect absorptionchiller of Kim and Infante Ferreira (2008). In addition, theenthalpy of lithium bromide (LiBr) solution was given byFlorides et al. (2003) and the saturated vapour pressureof LiBr solution by Patek and Klomfar(2006). The variousproperties of refrigerant (water) were determined accordingto Florides et al. (2003) and Zhang et al. (2003). The gen-eral parameters for solar absorption refrigeration areshown in Table 1, while the specific parameters are listedin Table 3.

2.4.2. Control and operation of absorption chiller and

auxiliary heater

The general provisions for the year-round control andoperation of hot water pump and auxiliary heater weresimilar to those in Section 2.3.3 for solar mechanical com-pression refrigeration, expect that the auxiliary heater

Solution

Pump

Expansion

Device

Heat

Exchanger

Throttling

Valve

Solar

Collectors

Hot Water Pump

Regenerative

Water Pump

T

T

C

Hot Water Pump

Controller

Generator

Cooling

Tower

T

Condenser Water Pump

Condenser

Evaporator Absorber

Absorption

Chiller

Town Gas

Supply

Hot

Water

Tank

Aux.

Heater

Chilled water supply and

return for the AHU serving the

indoor space (same as Fig. 1)

T

C Part-load

Controller

Fig. 3. Schematic diagram of solar absorption refrigeration.

Table 3

Modeling parameters for solar absorption refrigeration.

Parameter Value

Absorber water mass flow rate (kg sÿ1) 2.0

Overall heat transfer value of generator (kW Kÿ1) 4.5

Overall heat transfer value of absorber (kW Kÿ1) 4.5

Overall heat transfer value of solution-to-solution heat exchanger

(kW Kÿ1)

1.0

Degree of superheat at evaporator outlet (°C) 5.0

Solution volume flow rate at absorber outlet (L sÿ1) 0.1


would be activated through an on/off signal between 12and 9 °C with hysteresis based on the chilled water returntemperature at the chiller. The operation of chiller at aninsufficient driving temperature would cause heat loss fromthe hot water side to the condenser water side with no cool-ing capacity produced. To avoid this, the chiller wouldstart only when the auxiliary heater controller was onand the regenerative water temperature entering the chillerexceeded a preset minimum, which was 5 °C lower than theminimum driving temperature 67 °C of the absorption chil-ler. When the chiller was on, the condenser water pumpand the regenerative water pump would be in operationas well.

2.4.3. Part-load control of absorption chiller

To reduce the utilization of auxiliary heating, the hotwater temperature used to drive the chiller should be main-tained as low as possible. This was achieved by adopting apart-load control for the chiller which modulated the setpoint of the auxiliary heater between the minimum drivingtemperature 67 °C and the ceiling temperature 90 °C in 5steps. The set point started at 67 °C when the chilled waterreturn temperature at the chiller exceeded 12 °C, andstepped up if the water temperature dropped less than1 °C after one simulation time step. Conversely, when thechilled water return temperature was below 12 °C, the setpoint would step down if the water temperature droppedmore than 1 °C after one simulation time step. By employ-ing this part-load control, the chiller capacity could be reg-ulated through the control of driving temperature.

2.5. Solar adsorption refrigeration


The system simulation model of solar adsorption refrig-eration was built according to Fig. 4. The system compo-

nents were similar to those of the solar absorptionrefrigeration, except the adsorption chiller was usedinstead.

For adsorption refrigeration, an economical adsorptionpair is silica gel and water, where silica gel is the adsorbentand water the adsorbate (or refrigerant). For the configura-tion of adsorption cycle in the model, the two-chamberdesign was used. In a working cycle, one of the chamberswas used for adsorption while another for desorption.The role of these two chambers was interchanged inanother cycle, so as to maintain a pseudo-continuous pro-duction of refrigeration effect. Model development of thesolar adsorption refrigeration cycle was based on the vali-dated model of Cho and Kim(1992). The general modelingparameters for the solar adsorption refrigeration are shownin Table 1, while the specific parameters are outlined inTable 4.

2.5.2. Control and operation of solar adsorption refrigeration

The details of control and operation for the chiller andthe associated equipment were same as those in Section2.2.3 for the solar electric compression refrigeration. Theprovisions of control and operation of the hot water pump

Chilled water supply and return

for the AHU serving the indoor

space (same as Fig. 1)

Condenser

Evaporator

Adsorption

Compart-

ment 1

Adsorption

Compart-

ment 2

Solar

Collectors

Hot Water Pump

Regen.

Water

Pump

T

T

C

Hot Water Pump

Controller

Cooling

Tower

T Condenser Water Pump

Adsorption

ChillerPart-load

Controller

Town Gas

Supply

T

Hot

Water

Tank

Aux.

Heater

C

Fig. 4. Schematic diagram of solar adsorption refrigeration.

Table 4

Modeling parameters for solar adsorption refrigeration.

Parameter Value

Adsorption chamber water mass flow rate (kg sÿ1) 3.0

Specific heat capacity of metal (kJ kgÿ1 Kÿ1) 0.448

Specific heat capacity of silica gel (kJ kgÿ1 Kÿ1) 0.92

Specific heat of adsorption (kJ kgÿ1) 2,800

Mass of metal in adsorption/desorption chamber (kg) 120

Mass of silica gel in adsorption/desorption chamber (kg) 80

Mass of metal in condenser coil (kg) 150

Mass of metal in evaporator coil (kg) 150

Overall heat transfer value of adsorption chamber (kW Kÿ1) 16

Overall heat transfer value of desorption chamber (kW Kÿ1) 16

Adsorption/desorption period (s) 360


and auxiliary heater were same as those in Section 2.4.2 forthe solar absorption refrigeration. The part-load control ofthe adsorption chiller was same as that in Section 2.4.3 forthe solar absorption refrigeration, except the minimumdriving temperature was 64 °C for the solar adsorptionrefrigeration.

2.6. Solar solid desiccant cooling


In the solar solid desiccant cooling, the system simula-tion model was constructed based on Fig. 5. It encom-passed the solar collectors, hot water pump, hot watertank, auxiliary heater, regenerative water pump, regenera-tive heating coil, desiccant wheel, heat wheel, evaporativecoolers, supply air fan and exhaust air fan.

Themodel of desiccant wheel was developed based on thevalidated results of Zhang et al. (2003), the equilibrium rela-tive humidity was given by Pesaran andMills (1987a,b). Thegeneral parameters for modeling the solar solid desiccantcooling are shown in Table 1, while the specific parametersare presented in Table 5.

2.6.2. Additional design considerations of solar desiccant

cooling

Unlike the other four types of solar cooling systems con-taining chillers, the solar desiccant cooling would directlytreat the fresh air for air-conditioning purpose. Althoughthe indoor space was identical, the total cooling load wasincreased to 109 kW because ventilation load was muchhigher for the process air. The desiccant wheel was sizedaccording to the design regeneration temperature of85 °C, with 5 °C temperature difference between the enter-ing regenerative water and the leaving exhaust air of theregenerative heating coil. The desiccant wheel speed wasset to a value which would yield the lowest supply air tem-

perature at the design conditions. The flow rate of theregenerative water was selected so that the temperaturedrop after passing through the regenerative heating coilwould be 10 °C. The flow rate of the hot water pumpwas same as that of the regenerative water pump. Thecapacity of the auxiliary heater was 89 kW based on theworst situation that all the regenerative heat needed bythe desiccant cooling system was provided by the auxiliaryheater.

Evapor-

ative

Cooler 2

Exhaust

Air Fan

Outdoor Air

Regenerative

Heating Coil

Heat

Wheel

Supply

Air FanEvapor-

ative

Cooler 1

Solar

Collectors

Hot Water Pump

Regenerative

Water PumpT

T

C

Controller for Hot

Water Pump

Indoor Space

Exhaust Air

T

T

C

Outdoor

Air

Auxiliary Heater

Controller

Town Gas

Supply Aux.

Heater

Desiccant

Wheel

T

Hot

Water

Tank

Fig. 5. Schematic diagram of solar solid desiccant cooling.

Table 5

Modeling parameters for the solar solid desiccant cooling.

Parameter Value

Process air stream

Process air evaporative cooler saturation efficiency 90%

Process air evaporative cooler power consumption (kW) 0.1

Exhaust air stream

Exhaust air mass flow rate (kg sÿ1) 3.36

Exhaust air fan power (kW) 3.36

Exhaust air fan efficiency 70%

Exhaust air evaporative cooler saturation efficiency 90%

Exhaust air evaporative cooler power consumption (kW) 0.1

Overall heat transfer value of regenerative heating coil (kW Kÿ1) 9.72

Heat wheel

Temperature effectiveness of heat wheel 80%

Heat wheel power consumption (kW) 0.2

Desiccant wheel

Specific heat capacity of air (kJ kgÿ1Kÿ1) 1.005

Specific heat capacity of metal (kJ kgÿ1Kÿ1) 0.9

Specific heat capacity of silica gel (kJ kgÿ1Kÿ1) 0.921

Specific latent heat of vaporization of water (kJ kgÿ1) 2.400

Thermal conductivity of air (W mÿ1 Kÿ1) 0.0321

Density of air (kg mÿ3) 1.2

Diameter of desiccant wheel (m) 2

Width of desiccant wheel (m) 0.2

Effective area ratio of desiccant wheel 0.744

Fraction of wheel area for regeneration 0.5

Desiccant wheel speed (rph) 12

Desiccant wheel power consumption (kW) 0.2


2.6.3. Control and operation of solar desiccant cooling

The auxiliary heater was controlled through a desiccantcooling controller that provided an on/off signal between26.5 and 24.5 °C with hysteresis based on the space temper-ature. Moreover, the auxiliary heater only started when theambient temperature was above 20 °C. The operation ofthe desiccant cooling system at insufficient regenerationtemperature would cause heat loss from the hot water sideto the exhaust air with no cooling capacity produced. Toavoid this, the desiccant cooling system would run onlywhen the following three conditions were satisfied:

� The auxiliary heater controller was on.� The ambient temperature was above 20 °C.� The regenerative water temperature entering the regen-erative heating coil exceeded a preset minimum, whichwas 5 °C lower than the minimum driving temperature60 °C.

When the desiccant cooling system was on, the regener-ative water pump, desiccant wheel, heat wheel and theevaporative coolers would operate. The supply and exhaustair fans would be in operation according to the daily occu-pancy schedule.

3. Framework of comparative study

3.1. Preliminary study and full study

In order to have a thorough comparative study of differ-ent solar cooling systems, the types of installation strategiesand solar collectors were also included. Generally, there aretwo feasible installation strategies for the solar collectors.The first one is to make use of any unobstructed openspace, practically on the roof, to mount the solar collectorsat an inclination angle same as the local latitude. The sec-ond strategy is to apply the building-integrated approachby mounting the solar collectors vertically on the buildingfacades. For the high-rise buildings commonly found inHong Kong, the former would be limited by the usableroof area, while the latter can make use of the availablearea of the external facades for solar collector installation.About the common choices of solar thermal collectors,

there are flat plate collectors, evacuated tubes and para-bolic concentrators. By accounting for these installationstrategies and options of solar collectors, a variety of com-binations can be formed for the five types of solar coolingsystems as shown in Table 6.

From Table 6, there are altogether 26 combinations ofdifferent installation strategies and solar collectors for thefive types of solar cooling systems. It is an extensive simu-lation study about the solar cooling application in the con-text of year-round operation. Since certain installationstrategy and solar collector would have relatively inferioreffect to the system performances, a preliminary studywas carried out in order to remove the unnecessary combi-nations in the later full study. As a result, the followingcomparison works would be conducted first in the preli-minary study:

� Effect of installation strategies: The performances ofsolar electric compression refrigeration and solarabsorption refrigeration with flat plate collectors basedon the two installation strategies would be compared,so as to evaluate the more feasible option in view ofsolar electric gain and solar thermal gain in the subtrop-ical region.

� Effect of solar collectors: The performances of the flatplate collectors, evacuated tubes and parabolic concen-trators applied to the absorption refrigeration wouldbe compared, so as to find out the more feasible typesof solar collectors for the subtropical region.

The models of solar collectors were based on those inthe TESS library for TRNSYS. The coefficients of the effi-ciency curve for the evacuated tubes were based on(Hochschul Rapperswil of Switzerland, 1997). To modelthe building-integrated approach, the component modelof building-integrated photovoltaic/thermal system (TESSType 567, 2006) from TESS was used to interact to theindoor space of the building system under study. In thebuilding-integrated strategy, the solar collectors were verti-cally installed to the east, south and west. The north sidewas not used since this study was about the subtropicalregion in the northern hemisphere. In addition, three direc-tions were used instead of one because the three sides of

Table 6

Possible combinations of installation strategies and solar collectors for the solar cooling systems.

Installation

strategy

Photovoltaic

panels

Flat plate

collectors

Evacuated

tubes

Parabolic

concentrators

Solar electric compression refrigeration Roof-mountedp

Building-integratedp

Solar mechanical compression

refrigeration

Roof-mountedp p p

Building-integratedp p p

Solar absorption refrigeration Roof-mountedp p p


Solar adsorption refrigeration Roof-mountedp p p


Solar solid desiccant cooling Roof-mountedp p p



external facades were all needed in order to sufficientlyaccommodate the design mounting area of solar collectors.

3.2. Solar cooling systems against conventional electric-

driven vapour compression refrigeration

In order to promotewider application of the solar coolingsystems, it is necessary to benchmark their performancesagainst the conventional vapour compression refrigerationsolely energized by electricity. There are two kinds of chillerplants commonly applied in the subtropical region:

� The air-cooled packaged vapour compression refrigera-tion; and

� The water-cooled vapour compression refrigeration.

Based on the small and medium applications, the coeffi-cient of performance of air-cooled chiller was 2.4 and thatof water-cooled chiller was 3.0 in this study. The generalmodeling parameters of the chillers and the associatedequipment are same as those in Table 1, so as to have acommon ground for comparison.

4. Performance indicators

A number of performance indicators are needed in orderto have a full understanding of different aspects of the solarcooling systems. These performance indicators include thesolar fraction, coefficient of performance, solar thermalgain and primary energy consumption. A year-round per-spective of these performance indicators was used, so thatthe system performances could take into account thechanging loading and climatic conditions in this compara-tive study. The definitions, particularly the first two, woulddepend on the types of solar cooling systems, therefore theyare clearly described in this section.

4.1. Solar fraction, SF

Solar fraction is commonly used to evaluate the perfor-mance of a solar cooling system. The fraction shows theportion of solar energy contribution as compared to thetotal energy required to drive the refrigeration part of solarcooling system. This total amount of energy input is com-monly supplemented by the auxiliary provision. In thisstudy, the solar fractions are defined below.

For solar electric compression refrigeration,

SF elect ¼W solar

W solar þ W aux

ð1Þ

where SFelect = solar fraction of solar electric compressionrefrigeration; Wsolar = solar electric gain from PV panels(kW); Waux = auxiliary electrical power from the citypower grid (kW).

For the solar thermal cooling system including solarmechanical compression refrigeration, absorption refriger-ation, adsorption refrigeration and desiccant cooling,

SF thermal ¼Qsolar

Qsolar þ Qaux

ð2Þ

where SFthermal = solar fraction of solar thermal coolingsystem; Qsolar = solar thermal gain from the solar collectors(kW); Qaux = heat output of auxiliary heater (kW).

The solar fraction would be averaged in monthly oryearly basis, it is based on the SF acquired at each time stepwhen solar thermal gain is produced due to the hot waterpump in operation. The monthly averaged SF and yearlyaveraged SF are formulated below:

Monthly averaged SF ¼Xm

i¼1

SF i

mð3Þ

where m = total number of time steps in a month for hotwater pump in operation.

Yearly averaged SF ¼Xy

i¼1

SF i

yð4Þ

where y = total number of time steps in a year for hotwater pump in operation.

For the dynamic simulation of the year-round operationin this study, the simulation time step was 6 min as men-tioned in Section 2.1. However the number of time stepsinvolved finding the averaged solar fraction depended onwhether the related equipment of the solar cooling systemwas really in operation, since the indoor conditions couldstill be maintained at certain climatic and loading condi-tions in a year. The number of time steps involved to deter-mine the monthly/yearly averaged solar fraction is basedon the total number of time steps for hot water pump inoperation only. As discussed in Section 2.3.3, a hot waterpump controller was included to activate the hot waterpump whenever the temperature at the solar collector out-let exceeded the water temperature inside the hot watertank. In this situation, solar thermal gain could be actuallyacquired from the solar collectors. Therefore the solar frac-tion was counted and averaged when the hot water pumpwas in operation.

4.2. Coefficient of performance, COP

Coefficient of performance is an useful performanceindicator to evaluate the key air-conditioning equipmentof the solar cooling systems. In the solar electric refrigera-tion, solar mechanical refrigeration, solar absorptionrefrigeration and solar adsorption refrigeration, the keyair-conditioning equipment are the respective chillers. Inthe solar solid desiccant cooling, it is the desiccant wheel.Therefore the COP of different solar cooling systems aredefined below.


COP elect ¼Qe

Wð5Þ


where COPelect = COP of solar electric compression refrig-eration; Qe = refrigeration effect (kW);W = power input toDC motor (kW).

For solar mechanical compression refrigeration.

COPmech ¼Qe

Qboil

ð6Þ

where COPmech = COP of solar mechanical compressionrefrigeration; Qboil = heat input to boiler of Rackine cycle(kW).

For solar absorption refrigeration,

COP ab ¼Qe

Qgen

ð7Þ

where COPab = COP of solar absorption refrigeration;Qgen = heat input to generator (kW).

For solar adsorption refrigeration,

COP ad ¼Qe

Qde

ð8Þ

where COPad = COP of solar adsorption refrigeration;Qde = heat input for desorption (kW).

For solar solid desiccant cooling,

COP desi ¼_maðho ÿ hsÞ

Qregen

ð9Þ

where COPdesi = COP of solar solid desiccant cooling; _ma=mass flow rate of supply air (kg/s); ho = specific enthalpy ofoutdoor air (kW/kg); hs = specific enthalpy of supply air(kW/kg); Qregen = heat input for regeneration (kW).

The COP would be averaged in monthly or yearly basis,it is based on the COP acquired at the time step when thekey air-conditioning equipment (chiller or desiccant wheel)is in operation. The monthly averaged COP and yearlyaveraged COP are formulated below:

Monthly averaged COP ¼Xm0

i¼1

COP i

m0 ð10Þ

where m0 = total number of time steps in a month for chil-ler or desiccant wheel in operation.

Yearly averaged COP ¼Xy0

i¼1

COP i

y 0ð11Þ

where y0 = total number of time steps in a year for chilleror desiccant wheel in operation.

For the solar electric compression refrigeration, solarmechanical compression refrigeration, solar absorptionrefrigeration and solar adsorption refrigeration, the totalnumber of time steps involved were based on the chillerin operation. As discussed in Section 2.2.3, the operationof the chiller was determined by a controller that providedan on/off signal based on the chilled water return tempera-ture. For the solar desiccant cooling, the total number oftime steps was counted according to the desiccant wheelin operation. As mentioned in Section 2.6.3, the desiccantcooling system would be activated only when the auxiliary

heater controller was on; the ambient temperature wasabove 20 °C and the regenerative water temperature enter-ing the regenerative heating coil exceeded a presetminimum.

4.3. Solar thermal gain, Gsolar

Solar thermal gain is the useful energy acquired throughthe solar collectors to drive a solar cooling system. It is notjust related to the efficiency of the solar collectors, but alsothe nature of energy demand of the solar cooling system. Inthe four solar thermal cooling systems under study, it isabout the gain of thermal energy from the flat plate collec-tors, evacuated tubes or parabolic concentrators. Theamount of gain is related to both the type of solar collec-tors and system operation due to different climatic andloading conditions. Therefore it is also one of the perfor-mance indicators to evaluate the effectiveness of a solarcooling system with certain type of solar collector.

However in the solar electric refrigeration, the PV panelsare used and the solar electric gain Gelect is producedinstead of solar thermal gain. Since the solar electric gainand solar thermal gain are different in nature, a term calledsolar thermal gain equivalent Gsolar,eq is defined for com-parison purpose:

Gsolar;eq ¼ Gelect=ge ð12Þwhere ge is the energy efficiency used to convert the electri-cal energy into the primary energy, and it is taken to be33% in this study.

4.4. Primary energy consumption, Ep

When calculating the primary energy consumption, it isin a system approach. All the parasitic energy consump-tions from the associated equipment were considered uponthose of the main equipment. Since two different types ofenergy sources, electricity and gas, were involved, theirenergy efficiencies were assumed below:

� energy efficiency for electrical energy converted into pri-mary energy, ge = 33%;

� energy efficiency for gas energy converted into primaryenergy, gg = 90%.

The former was used to determine the primary energyinput of the electric-driven equipment, while the latter forthe auxiliary heater operated by burning gas. Ep is there-fore determined in different ways for the correspondingsolar cooling systems as follows.


Ep;elect ¼ ðW aux þ W parasitic þ W AHU Þ=ge ð13Þwhere Ep,elect = primary energy consumption of solar elec-tric compression refrigeration (kWh); Wparasitic = electricalenergy consumption of parasitic equipment, includingpumps and cooling tower (kWh); WAHU = electrical energy


consumption of AHU (kWh).For solar mechanical com-pression refrigeration, absorption refrigeration and adsorp-tion refrigeration,

Ep;thermal ¼ Qaux=gg þ ðW parasitic þ W AHU Þ=ge ð14Þ

where Ep,thermal = primary energy consumption of solarthermal refrigeration (kWh).

For solar solid desiccant cooling,

Ep;desi ¼ Qaux=gg þ ðW para;desi þ W fanÞ=ge ð15Þ

where Ep,desi = primary energy consumption of solar soliddesiccant cooling (kWh); Wpara,desi = electrical energy con-sumption of parasitic equipment, including pumps, desic-cant wheel, heat wheel and evaporative coolers (kWh);Wfan = electrical energy consumption of supply air fanand exhaust air fan (kWh).

5. Preliminary study

5.1. Effect of installation strategies of solar collectors

The results of the two feasible installation strategies forthe solar electric compression refrigeration and solarabsorption refrigeration are summarized in Table 7. Theperformances were evaluated on a year-round basis,including the four performance indicators and the averagedspace cooling load. Due to the fact that the solar collectorsintegrated to the building facades can reduce the transmis-sion heat gains, it is found that the year-round averagedspace cooling load was reduced by 12.7% and 9.2% forthe building-integrated solar electric compression refrigera-tion and solar absorption refrigeration, as compared to theroof-mounted strategy. However, both cases could not out-weigh the significant drop of the corresponding solar frac-tions at 50.2% and 82.3%, as well as the decrease of therespective solar thermal gains at 65.8% and 83%. Finally

the year-round totals of primary energy consumptions wereincreased by 13.9% and 35.7% for the building-integratedsolar electric refrigeration and solar absorption refrigera-tion respectively. The COP was about the same or slightlyincreased, it was not due to the improvement of the build-ing-integrated approach, but just because of the auxiliaryprovisions operated whenever in need. Therefore, COP

was not effective to reflect the performance of solar coolingsystems in this case.

The poorer result of the building-integrated solarabsorption system was further due to the performancesof the solar collectors installed at the three different direc-tions. Based on the typical design, the hot water pumpwould only operate when the water leaving temperatureof the solar collectors was higher than that of the hot waterstorage tank. As the size of solar collector facing to eachdirection was reduced to one-third, the hot water tempera-ture that could be achieved would be lower, causing signif-icant drops of solar fraction and solar thermal gain.

As a result, the building-integrated strategy was notinvolved in the ongoing comparative study for differentsolar cooling systems. Such strategy is only recommendedwhen the roof or the available open space is not enoughto accommodate the required number of solar collectors.

5.2. Effect of solar collectors

Table 8 summarizes the year-round performances ofthe solar absorption refrigeration system with differenttypes of solar collectors, including the flat plate collectors,evacuated tubes and parabolic concentrators. The four per-formance indicators were involved to evaluate the effective-ness of solar collectors.

The type of flat plate collectors was used as the baselineof comparison. It can be observed that both the evacuatedtubes and parabolic concentrators had higher solar fraction

Table 7

Year-round performances of the two feasible installation strategies for the solar electric compression refrigeration and solar absorption refrigeration.

Solar cooling system Installation

strategy

Year-round

averaged space

load (kW)

Year-round

averaged SF

Year-round

averaged COP

Year-round

total of

Gsolar (kWh)

Year-round

total of Ep (kWh)

Solar electric compression refrigeration Roof-mounted 10.99 0.687 4.599 31,678 44,589

Building-integrated 9.59 (;12.7%) 0.342 (;50.2%) 4.658 ("1.3%) 10,844 (;65.8%) 50,781 ("13.9%)

Solar absorption refrigeration Roof-mounted 11.11 0.497 0.769 37,234 72,797

Building-integrated 10.09 (;9.2%) 0.088 (;82.3%) 0.763 (;0.8%) 6,326 (;83.0%) 98,787 ("35.7%)

Table 8

Year-round performances of different types of solar collectors for the solar absorption refrigeration.

Type of solar

collectors

Year-round averaged

SF

Year-round averaged

COP

Year-round total of Gsolar

(kWh)

Year-round total of Ep (kWh)

Flat plate 0.497 0.769 37,234 72,797

Evacuated tubes 0.818 ("64.6%) 0.763 (;0.8%) 67,383 ("81.0%) 49,425 (;32.1%)

Parabolic

concentrators

0.596 ("19.9%) 0.777 ("1.0%) 47,929 ("28.7%) 67,450 (;7.3% vs. flat plate; ;36.5% vs.

evacuated tubes)


and solar thermal gain, and lower primary energyconsumption. However, since the parabolic concentratorsonly responded to the direct solar irradiation, their perfor-mances were lower than that of the evacuated tubes despitea higher collector temperature could be achieved. In theyear-round total of primary energy consumption, the par-abolic concentrators were 36.5% worse than the evacuatedtubes, and just 7.3% better than the flat plate collectors. Infact, the direct solar irradiation is only 42.4% of the total inthe subtropical Hong Kong (Fong et al., 2007). Moreover,within the same collector installation area, the totalcollector tubes of parabolic concentrators would befewer due to the substantial coverage of the parabolicmirrors.

The year-round averaged COP varies only within a verynarrow range, from 0.763 to 0.777, for these three types ofsolar collectors. The presence of the auxiliary heater tendedto minimize the variation of effectiveness of different typesof solar collectors to the absorption chiller.

As a result, the option of parabolic concentrators wasexcluded from the following comparative study, since itcould not have a distinctive performance as compared tothe evacuated tubes in the subtropical Hong Kong. How-ever, the flat plate collectors were still involved becauseof their popularity and simple configuration.

6. Comparative study of solar cooling systems

From the preliminary performance study, the options ofbuilding-integrated strategy and parabolic concentratorswere removed, therefore the variety of combinations forthe full comparative study could be trimmed down asshown in Table 9.

6.1. Overall year-round performances of different solar

cooling systems

The performance results of different solar cooling sys-tems are summarized in Table 10. The four performanceindicators were involved in the analysis. The solar electriccompression refrigeration was used as the baseline of thiscomparative study, except its COP due to different nature.In terms of the year-round total of primary energy con-sumption, the order of the solar cooling systems from thebest to the worst is found as follows:

� solar electric compression refrigeration� solar absorption refrigeration� solar adsorption refrigeration� solar solid desiccant cooling� solar mechanical compression refrigeration

In fact, the solar electric compression refrigeration andsolar absorption with evacuated tubes had comparable per-formances. Although the primary energy consumption ofthe solar absorption refrigeration was 10.8% higher, itssolar faction was 19.1% better. It is also observed that thisorder is independent to the types of solar collectorsadopted. For the same kind of solar thermal cooling sys-tem, it is clear that the evacuated tubes had better perfor-mances than the flat plate collectors. In this study, thePV efficiency is based on the typical information of thePV component of TRNSYS (2006). The simulation resultshowed the yearly averaged efficiency was around 10%according to the climatic conditions of Hong Kong.Although the solar electric compression refrigeration hasthe best overall performances, particularly primary energy

Table 9

Comparative study of different solar cooling systems.

Roof-mounted PV panels Roof-mounted flat plate collectors Roof-mounted evacuated tubes

Solar electric compression refrigerationp

Solar mechanical compression refrigerationp p

Solar absorption refrigerationp p

Solar adsorption refrigerationp p

Solar solid desiccant coolingp p

Table 10

Year-round performances of different solar cooling systems.

Solar cooling system Type of solar

collector

Year-round

averaged SF

Year-round

averaged COP

Year-round total of Gsolar

(kWh)

Year-round total of Ep

(kWh)

Solar electric compression

refrigeration

PV 0.687 4.599 31,678, Gsolar,eq 44,589

Solar mechanical compression

refrigeration

Flat plate 0.046 (;93.9%) 0.102 26,911 (;15.0%) 629,897 ("1313%)

Evacuated tubes 0.112 (;83.7%) 0.102 65,513 ("107%) 592,292 ("1228%)

Solar absorption refrigeration Flat plate 0.497 (;27.7%) 0.769 37,234 ("17.5%) 72,797 ("63.3%)

Evacuated tubes 0.818 ("19.1%) 0.763 67,383 ("113%) 49,425 ("10.8%)

Solar adsorption refrigeration Flat plate 0.313 (;54.4%) 0.435 40,421 ("27.6%) 128,715 ("189%)

Evacuated tubes 0.577 (;16.0%) 0.437 72,885 ("130%) 93,692 ("110%)

Solar solid desiccant cooling Flat plate 0.336 (;51.1%) 1.066 40,435 ("27.6%) 149,356 ("235%)

Evacuated tubes 0.552 (;19.7%) 1.059 71,832 ("127%) 128,052 ("187%)


consumption, they were acquired based on a typical PVpanel, not due to a certain product with exceptionally highefficiency.

The solar mechanical compression refrigeration hadexceptionally high year-round primary energy consump-tion, showing that the auxiliary heater, instead of solar col-lectors, was the essential contributor for driving theRankine cycle of the heat engine. The narrow range ofoperative driving temperature of the Rankine-cycle-drivenchiller maintained a high-temperature demand and thepart-load control became useless. The solar collectorscould not have effective contribution due to its relativelylow-temperature thermal gain.

The solar adsorption refrigeration was the third in viewof the primary energy consumption. Although its primaryenergy consumption was about double of that of solar elec-tric compression refrigeration or solar absorption refriger-ation, continual research is being carried out for moreeffective adsorbent–adsorbate pairs (Fan et al., 2007; Khat-tab, 2004; Wang et al., 2005), instead of the typical pair ofsilica gel and water used in this model. On the other hand,there are emerging configurations for enhancing the effec-tiveness of adsorption–desorption process (Wang et al.,2005, 2008; Voyiatzis et al.,2007). Therefore the applicationpotential of solar adsorption refrigeration still exists.

The solar solid desiccant cooling had high year-roundprimary energy consumption, but it was different fromthe other solar cooling systems because it had to handlethe extra ventilation load due to the inherent nature of fullfresh air design. Therefore the total cooling capacity of des-iccant cooling system would be much higher than those ofthe other four solar cooling systems, hence higher primaryenergy consumption. In addition, the parasitic energy con-sumption was comparatively high, particularly the supplyair fan and exhaust air fan. The feature of solar desiccantcooling in effect is more than only to satisfy the requiredcooling load, it is also capable to supply the fresh airamount far above the minimum requirement of a func-tional area. This unique merit can guarantee a good indoorair quality and ventilation effectiveness. Therefore theapplication potential of solar solid desiccant cooling is stillvalid.

The overall performances of the average room and sup-ply air conditions are summarized in Table 11. As the solar

mechanical compression refrigeration had exceptionallyhigh year-round primary energy consumption, it wasexcluded in this part of evaluation. From Table 11, it isfound that the average room air conditions of various solarcooling systems are close to the design conditions at25.5 °C and 60% RH, showing that they could provide sat-isfactory indoor environment for the occupants. The per-formances of solar electric compression refrigeration,solar absorption refrigeration and solar adsorption refrig-eration are very close (24.4–24.9 °C and 59–61%RH)regardless the type of solar collectors. It is because theyprovided air-conditioning in a similar way, by using thechilled water supply to cooling coil for cooling and dehu-midification purposes. Solar desiccant cooling had slightdifference of room temperature and relative humidity fromthe other three types because the supply air was handled inanother way, by using the desiccant wheel, heat wheel andevaporative cooler.

6.2. Monthly performances of different solar cooling systems

Apart from the year-round performances, it is also wor-thy to understand the monthly variations of the perfor-mance indicators of different solar cooling systems. Thechanges would be according to the climatic conditions,loading conditions and features of solar cooling systems.The annual profile of each performance indicator is there-fore presented and discussed in this section.

6.2.1. Monthly averaged solar fraction

Fig. 6 shows the annual profiles of solar fraction. It isfound that the solar fraction was commonly low fromMay to September, which is the typical summer periodand the cooling load is around the peak. On the otherhand, the solar fraction was usually high from Decemberto February, which is the typical winter season. Generally,the solar fraction was less than one, indicating that just thesolar electric or thermal gain acquired from the solar col-lectors was not sufficient to fully drive the respective solarcooling system (except the solar desiccant cooling), so theauxiliary provision was inevitably activated to supplementthe deficit. As a whole, the solar absorption refrigerationhad the highest solar fraction; the solar electric compres-sion refrigeration the second; the solar adsorption refriger-

Table 11

Yearly average room and supply air conditions of various solar cooling systems.

Solar cooling system Type of solar

collector

Average room

temperature (°C)

Average room

relative humidity (%)

Average supply air

temperature (°C)

Average supply air

humidity ratio

(kg/kg dry air)

Solar electric compression refrigeration PV 24.9 61 19.4 11.7 � 10ÿ3

Solar absorption refrigeration Flat plate 24.5 59 19.7 11.1 � 10ÿ3

Evacuated tubes 24.4 59 19.6 11.0 � 10ÿ3

Solar adsorption refrigeration Flat plate 24.5 59 19.7 11.2 � 10ÿ3


Solar solid desiccant cooling Flat plate 23.6 67 20.5 12.2 � 10ÿ3



ation the third; the solar desiccant cooling the fourth; andthe solar mechanical compression refrigeration the last.

For the solar solid desiccant cooling, the solar fractionwas unity from December to February. This shows thatthe solar thermal gain from solar collectors was enoughto drive the system in such period, and free air coolingmight also be involved. On the other hand, the extremelylow solar fraction of solar mechanical compression refrig-eration shows its heavy reliance on auxiliary heating andvery weak support from the solar collectors.

6.2.2. Monthly averaged COP

Fig. 7 shows the monthly averaged COP of differentsolar cooling systems. The solar electric compression refrig-eration is excluded from this analysis because the nature ofits COP is different from that of the four solar thermalcooling systems. For the COP in the summer period

(May to September), the order from the best to the worstis found as follows:

� solar solid desiccant cooling� solar absorption refrigeration� solar adsorption refrigeration� solar mechanical compression refrigeration

The solar solid desiccant cooling could have COP

greater than unity in such period. From Eq. (9), its COP

definition is different to the three solar thermal refrigera-tion systems. The numerator is the enthalpy change ofthe process air, instead of refrigeration effect. For the threesolar thermal refrigeration systems, their patterns are sim-ilar and nearly flat throughout a year. This is because thesupport from auxiliary heating would not stop wheneverthe required driving temperature is not sufficient. On theother hand, suitable control and operation of auxiliary

Fig. 6. Annual profiles of solar fraction of different solar cooling systems. (Abbreviation: AB, absorption refrigeration; AD, adsorption refrigeration; DE,

desiccant cooling; EL, electric compression refrigeration; ME, mechanical compression refrigeration; et, evacuated tubes; fp, flat plate collectors; pv,

photovoltaic panels).

Fig. 7. Annual profiles of coefficient of performance of different solar cooling systems. (Abbreviation: AB, absorption refrigeration; AD, adsorption

refrigeration; DE, desiccant cooling; EL, electric compression refrigeration; ME, mechanical compression refrigeration; et, evacuated tubes; fp, flat plate

collectors; pv, photovoltaic panels).


heating would lead to the COP independent to the type ofsolar collectors. It is found that both flat plate collectorsand evacuated tubes had similar COP for the same typeof solar thermal systems.

For solar solid desiccant cooling system, the COP waslower than one in the months other than the summer per-iod, in between 0.67 and 0.98. In January and February,the COP was zero. It is because the operation mode of freeair cooling was effective to handle the space cooling load,so the desiccant wheel and related equipment were notneeded to operate.

6.2.3. Monthly total of solar thermal gain

Fig. 8 shows the variation of monthly averaged solarthermal gain of different solar cooling systems. It is foundthat there are two sets of curves clustered together due tothe type of solar collectors. The set of performance curvesof solar cooling systems using evacuated tubes are higherthan those using flat plate collectors. This clearly showsthe evacuated tubes have a more superior solar thermalgain compared to the flat plate collectors. It is noted thatthe performance curve of solar electric compression refrig-eration is within the lower cluster. Although it was trans-formed to Gsolar,eq due to its nature of solar electric gain,its performance was not as low as those systems with flatplate collectors. It can be found that the yearly averagedsolar fraction of solar electric compression refrigeration is0.687, which is relatively high in Table 10, showing thatthe solar electric gain was quite enough to drive the refrig-eration cycle.

Generally the solar thermal gain was high in the periodof July to October, which covers from the mid-summer toearly autumn. It is not fully matched to the solar fractionprofiles, which are mainly affected by the cooling loaddemand in the hot season from May to September. Thevariation of solar thermal gain primarily follows the avail-ability of solar irradiation in the subtropical Hong Kong,

in which the profile is also high from July to October, whilelow in January and February (Fong et al., 2007).

6.2.4. Monthly total of primary energy consumption

Fig. 9 shows the annual profiles of the primary energyconsumption of different solar cooling systems. The perfor-mance of the solar mechanical compression refrigerationwas excluded from the analysis since its year-round totalwas far more higher than the other types of solar coolingsystems as shown in Table 10. It is clear that the solar elec-tric compression refrigeration has the lowest primaryenergy consumption. In fact, the annual profile of solarabsorption refrigeration with evacuated tubes was closeto that of the solar electric compression refrigeration, theslight difference was mainly due to the additional energyrequired for the parasitic equipment for the solar thermalpart, particularly the hot water pump and regenerativepump. Generally these profiles are similar to the annualcooling load variation. The primary energy consumptionis high from June to August, which is coincident to thepeak cooling load demand in the summer period. The pri-mary energy consumption is low from December to Febru-ary, which is the typical winter.

6.3. Performances of solar cooling systems against

conventional electric-driven vapour compression refrigeration

Application potential of solar cooling systems depends ontheir performances against the conventional electric-drivenvapour compression refrigeration, particularly the primaryenergy consumption. Both the air-cooled and water-cooledvapour compression chiller plants were involved in thisstudy, it is found that their year-round totals of primaryenergy consumption were 86,289 kWh and 79,138 kWh,respectively, and their annual profiles are plotted togetherwith those for the solar cooling systems in Fig. 9. From theyear-round totals of different solar cooling systems shown

Fig. 8. Annual profiles of solar thermal gain of different solar cooling systems. (Abbreviation: AB, absorption refrigeration; AD, adsorption refrigeration;

DE, desiccant cooling; EL, electric compression refrigeration; ME, mechanical compression refrigeration; et, evacuated tubes; fp, flat plate collectors; pv,

photovoltaic panels).


in Table 10, only the solar electric compression refrigerationand solar absorption refrigeration (no matter with evacu-ated tubes or flat plate collectors) have lower primary energyconsumption. The same results can also be observed inFig. 9. As compared to the conventional air-cooled andwater-cooled refrigeration, the primary energy savings ofsolar electric compression refrigeration were decreased by48.3% and 43.7%, while those of solar absorption refrigera-tion with evacuated tubes by 42.7% and 37.5%, and those ofsolar absorption refrigeration with flat plate collectors by15.6% and 8.0%. Although the solar adsorption refrigera-tion with evacuated tubes had relatively higher primaryenergy consumption, it was only 8.6% and 18.4% abovethe conventional air-cooled and water-cooled refrigerationsystems, respectively.

For the other types of solar cooling systems, their primaryenergy consumptions were far higher than those of the twoconventional electric-driven refrigeration systems. There-fore they do not have any energy saving potential even thefree solar energy is consumed. If life-cycle cost analysis is car-ried out, both initial and operating costs of such solar cool-ing systems do not have any advantages at all.

7. Conclusion

A comparative study was thoroughly carried out for thefive types of solar cooling systems for a typical office in thesubtropical Hong Kong. The results were worked out withthe emphasis of suitable system control and operation inresponse to the year-round changing climatic and loadingconditions. Based on the best year-round total of primaryenergy consumption, the order of the five types of solarcooling systems is:

� solar electric compression refrigeration,� solar absorption refrigeration,� solar adsorption refrigeration,

� solar solid desiccant cooling, and� solar mechanical compression refrigeration.

In the preliminary study for the options of installationstrategies and types of solar collectors, the alternatives ofbuilding-integrated approach and parabolic concentratorswere found not energy-efficient enough. Although thebuilding-integrated approach could reduce the space cool-ing load by about 10%, but the overall system perfor-mances were worse than the typical roof-mountedapproach. For the choice of solar collectors, the parabolicconcentrators had the primary energy consumption onlyslightly better than the flat plate collectors by 7.3% forthe absorption refrigeration, but worse than the evacuatedtubes by 36.5%.

If the conventional electric-driven air-cooled and water-cooled vapour compression chiller plants were included inthe comparison, the solar electric compression refrigerationand solar absorption refrigeration (with evacuated tubes orflat plate collectors) were the two types of solar cooling sys-tems that could have attractive energy saving potential.The year-round energy savings would be from 15.6% to48.3% compared to the conventional electric-driven air-cooled refrigeration, while 8.0% to 43.7% to the water-cooled refrigeration. In fact, it is possible to have continualimprovement in the solar electric compression refrigerationand solar absorption refrigeration, particularly the technol-ogy advancement of the solar collectors.

The application potential of solar adsorption refrigera-tion with evacuated tubes still exists, since its year-roundtotal of primary energy consumption is just slightly higherthan the conventional electric-driven air-cooled vapourcompression refrigeration. Continual advancement of theadsorption pair and chiller configuration would enhancethe use of its distinctive merit of low driving temperature.

Although the solar solid desiccant cooling has high year-round primary energy consumption, it is different from the

Fig. 9. Annual profiles of primary energy consumption of different solar cooling systems (including those of electric-driven vapour compression

refrigeration). (Abbreviation: AB, absorption refrigeration; ACVCR, conventional air-cooled vapour compression refrigeration; AD, adsorption

refrigeration; DE, desiccant cooling; EL, electric compression refrigeration; ME, mechanical compression refrigeration; WCVCR, conventional water-

cooled vapour compression refrigeration; et, evacuated tubes; fp, flat plate collectors; pv, photovoltaic panels).


other types of solar cooling systems because it has to tacklethe high ventilation load from its inherent full fresh airprovision. Subsequently it has high cooling capacity andlarge energy consumption of supply and exhaust air fans.In fact, the solar desiccant cooling is able to supplyexcessive fresh air for a functional area, resulting a goodindoor air quality and ventilation effectiveness. Thereforeapplication potential of solar solid desiccant cooling is stillvalid.

This paper also describes the framework and useful per-formance indicators for a comparative study of differentsolar cooling systems. The methodology can be appliedto evaluate the potential systems and configurations, andto benchmark the conventional electric-driven vapour com-pression refrigeration.

Acknowledgement

The work described in this paper was fully supported bya grant from City University of Hong Kong (Project No.7002310).

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Comparative Study of Different Solar Cooling Systems for Buildings