Cooling potential of ventilated PV façade and solar air heaters combined with a desiccant cooling machine

Renewable Energy 31 (2006) 1265–1278 www.elsevier.com/locate/renene

Technical note

Cooling potential of ventilated PV fac¸ade and solar air heaters combined with a desiccant cooling machine Li Meia,*, David Infieldb, Ursula Eickerc, Dennis Lovedaya, Volker Fuxc a

Department of Civil and Building Engineering, Loughborough University, Loughborough, LE11 3TU, UK b CREST, Loughborough University, Loughborough, LE11 3TU, UK c Hochschule fur Technik Schellinstr 24, 70174 Stuttgart, Germany Received 29 November 2004; accepted 12 June 2005 Available online 11 August 2005

Abstract Thermal energy collected from a PV-solar air heating system is being used to provide cooling for the Mataro Library, near Barcelona. The system is designed to utilise surplus heat available from the ventilated PV facade and PV shed elements during the summer season to provide building cooling. A desiccant cooling machine was installed on the library roof with an additional solar air collector and connected to the existing ventilated PV fac¸ade and PV sheds. The desiccant cooling cycle is a novel open heat driven system that can be used to condition the air supplied to the building interior. Cooling power is supplied to the room space within the building by evaporative cooling of the fresh air supply, and the solar heat from the PV-solar air heating system provides the necessary regeneration air temperature for the desiccant machine. This paper describes the system and gives the main technical details. The cooling performance of the solar powered desiccant cooling system is evaluated by the detailed modelling of the complete cooling process. It is shown that air temperature level of the PVsolar air heating system of 70 8C or more can be efficiently used to regenerate the sorption wheel in the desiccant cooling machine. A solar fraction of 75% can be achieved by such an innovative system and the average COP of the cooling machine over the summer season is approximate 0.518. q 2005 Elsevier Ltd. All rights reserved. Keywords: Solar energy; Ventilated PV fac¸ade/shed; Desiccant cooling; Energy balance calculation

* Corresponding author. Tel.: C44 1509223778; fax: 44 1509610031. E-mail address: [email protected] (L. Mei).

0960-1481/$ - see front matter q 2005 Elsevier Ltd. All rights reserved. doi:10.1016/j.renene.2005.06.013

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1. Introduction A building integrated ventilated photovoltaic (PV) system is advantageous both from electrical and thermal point of view. The air circulation behind the PV panel lowers the temperature of PV module and thus improves their electrical performance. Furthermore, a controlled air flow in the cavity between the PV panel and the surface behind leads to solar air heating with the potential to be applied to building for winter preheating or summer cooling. Over a number of years, the authors have investigated the thermal performance and the application of the ventilated PV fac¸ade integrated to building in winter season [1–3]. Directly using of the warmed air from the PV fac¸ade for building heating is environmentally sound. However, the thermal energy generated in the ventilated PV facade is usually not used in summer when it is most abundant. It is also conceivable that the PV fac¸ade structure might even add to the building’s cooling load. An attractive solution to this problem is to use the heated air from the fac¸ade (which in summer is unwanted and normally vented to the outside) to drive an adsorption cooling system for buildings cooling in summer. The authors have recently completed an EU project which concerned the ventilated PV-solar air heating system combining with a desiccant cooling machine constructed and installed in the Mataro Library, near Barcelona. The aim of this project is to design and implement an integrated solar heating and cooling system using building integrated PV solar air heating components coupled to a desiccant cooling system. The main scientific objective of the project is an integrated energy analysis of such systems taking into account the thermal properties of the ventilated PV fac¸ade and its impact on the heating and cooling loads of the building as well as the generation of solar heated air and operation of the desiccant cooler. This has been done by dynamic modelling and validation through an experimental evaluation of the system. Several major innovations have been achieved in the project: † The project proposed the first building integrated solar heating and cooling system in Europe, using ventilated PV and solar air collectors for the heat production and desiccant wheel technology for the solar driven cooling system † An integrated energy analysis for the first time deliver validated results of the summer performance of PV/solar air ventilated facades, consisting both the passive cooling load caused the integrated components and their thermal energy production usable for the active cooling system. † Optimised control algorithms for solar powered desiccant systems have been developed and tested. The control algorithm development and the integration into standard building management system can enhance system performance and significantly contribute to a cost reduction of such innovative cooling concepts.

2. The description of the system The Mataro´ library building, near Barcelona, has a total surface area of 3200 m2, distributed over three floor levels [4]. Both the south facade and the inclined sheds on

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PV shed 1 PV shed 2

ms, Ts

Air collector field

mc, Tc 3000~9000m3

PV shed 3 PV shed 4

mf, Tf PV Facade

Fig. 1. Solar air system.

the roof are equipped with ventilated PV modules (225 m2 on the south facade and a total of 300 m2 on the roof). The original south PV facade is 6.5 m high, to which the 1.7 m high air collectors have been added to boost the temperature levels of the ventilated PV system. The ventilation channels of the four rows of the PV sheds on the roof (308 inclination) are connected in parallel to the PV fac¸ade channel to supply a merged air volume to an additional solar air collector field. This additional air collector field has the heating area of 105 m2 and is separated in three parallel strings with 35 m length and 95 mm channel height. The further heated air from the additional air collectors is drawn to the desiccant machine to regenerate the sorption wheel. The regeneration air can be controlled between the maximum of 9000 m3/h and minimum of 3000 m3/h volume flow rate to keep a necessary regeneration air temperature. The solar air system is shown in the Fig. 1 The desiccant cooling machine installed on the roof of the Mataro library is an open heat driven cycle which comprises a sorption dehumidify wheel, a heat recovery wheel and the humidifiers. Such system will collect outdoor air as the supply air and will first dehumidify it with a silica-gel desiccant. By this dehumidification, the supply air warms up and pre-cooling with the exhaust air by a heat exchanger is necessary. Thus, the dehumidified supply air is cooled in a heat recovery wheel and further in an evaporative humidifier. After the humidification of the supply air the maximum cooling effect is limited by the maximum water content which is allowed to guarantee thermal comfort in the building. The cooled supply air after humidification is passed to the indoor space to satisfy the human environmental requirements. It is clear that the latent and sensible loads of the room space are more efficiently handled in the desiccant cooling process than in other cooling methods since components can be designed to handle these loads separately. In the exhausting air stream, the silica-gel desiccant is regenerated by the solar heat air from the PV-solar air collectors for releasing water moistures. The desiccant wheel used in the system comprises a matrix containing parallel channels which enables a large heat and mass transfer area to be maintained. The matrix is coated with the silica-gel desiccants and has the 2100 mm of diameter, 360 mm of depth. The wheel rotates slowly at the speed of 20 revolutions per hour. While, the heat recovery wheel having the same diameter as the desiccant wheel rotates at 600 revolutions per hour for the effective heat exchange. The volume flow rate in the supply and exhaust air streams

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PV facade

PV shed Air Collectors

7 8

1

Ambient air

Sorption wheel

2 Recovery wheel

6

5

3

4

Humidifier

Fig. 2. The simple scheme of the system.

are controlled in 3000, 6000 and 12,000 m3/h for the different control strategies. The simple scheme for this desiccant cooling system is shown in the Fig. 2. For such a solar powered desiccant cooling system, the component models were established and the simulation study of its performance was carried out. The thermal building simulation tool of TRNSYS was used for simulate the desiccant cooling system integrated with the Mataro Library building. When simulating the system operation process including the control strategies, the energy balances for the building cooling were yield.

3. Solar application Since the desiccant cooling system was driven by solar air heat, the total useful heat energy and the outlet air temperature of the ventilated PV fac¸ade and PVsheds as well as the additionally installed solar air heater should be determined. As the air channels of the PV fac¸ade and the PV sheds are parallel connected to supply the preheated air to the additional air collectors, the temperature of the merged air is an important factor in the solar cooling procedure. In general, a simplified Duffie’s solar air collector model [5] for the both PV fac¸ade and PV sheds developed by the authors in the previously completed project [1] was utilised to estimate the outlet air temperature of the PV fac¸ade and PV sheds. This model is based on the standard test method for solar collectors. The instantaneous thermal efficiencies of the PV fac¸ade and PV shed were given by: hpf Z

Qu;pf Tin;pf KTa Z 0:27K6:37 GT Apf GT

(1)

and Tout;pf Z

Qu;pf C Tin;pf Cp m_ f

(2)

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where Tin,pf(8C), Ta(8C) and GT (W/m2) obtained from the data logging system of the Mataro Library are the inlet air temperature of the PV fac¸ade, the ambient air temperature and the solar irradiance, respectively. The parameters of 0.27and 6.37 (W/Km2) were estimated using the measured data, such as the air mass flow rate, the inlet and outlet air temperatures of the PV channel, from Mataro Library data acquisition system during the period 01/08/2000–06/08/2000, for 40 min periods around midday. Consequently, the outlet air temperature of the PV fac¸ade, Tout,pf (8C)can be calculated by the inlet air temperature, Tin,pf (8C), mass flow rate m_ f (kg/s) and the total useful thermal energy Qu, ¸ ade. The same model pf (W). In the Eqs. (1) and (2), the subscripts only stand for the PV fac was used for the PV sheds and not repeated here. Then, it is clear to notice that the different collecting areas collect the different useful thermal energy and cause the different outlet air temperature for both the PV fac¸ade and PV sheds. The performance characteristic of the additional solar air collectors (105 m2) was given by the manufacture’s test data which can be expressed as: hc

Qu;c T KTa Z 0:86K6:5 in;c GT A GT

(3)

where, Tin,c (8C), the air collector inlet air temperature is the merged air temperature of the PV fac¸ade and PV sheds. The regenerating air temperature to the desiccant cooling machine is thus obtained by: Tout;c Z

Qu;c C Tin;c Cp m_ c

(4)

In order to maintain the regeneration air temperature at a significant level (above 70 8C), the ventilation air volume through the PV fac¸ade and PV sheds and the additional solar air collectors should be controlled between 3000 and 9000 m3/h. In the simulation model, a simple PI controller was designed for this purpose and implemented in the simulation program. Both measured and simulated results show the total temperature increase from ambient to the regeneration side of 35–40 K at high irradiance levels. In the solar air system, the air volume through the PV fac¸ade and the PV sheds are connected to the air collectors in parallel. The merged air temperature to the air collector inlet thus approximates to the following relationship: Tin;c Z

m_ s m_ f Tout;ps C T m_ s C m_ f m_ s C m_ f out;pf

(5)

where, the outlet air temperatures of the PV fac¸ade and PV sheds, Tout,pf and Tout,ps are inverse to the air mass flow rates of the PV fac¸ade and PV sheds, m_ f and m_ s . Thus, the solar air collector inlet temperature, Tin;c , is not varied with the different ration of the m_ s =m_ f . Furthermore, the regeneration solar energy delivered to the desiccant cooling machine is not varied with the different ratio of the m_ s =m_ f .

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4. Desiccant cooling The desiccant cooling machine including the sorption wheel, heat recovery wheel and the humidifiers was modelled to simulate the summe_rtime psychrometric desiccant processes and utilise hourly meteorological data to predict energy consumption For the sorption wheel, the dimensionless effectiveness was used to characterise the performance of the dehumidifier. In this study, the effectiveness of the dehumidifier was assumed to be a constant. The humidity ratio and temperature of the air leaving sorption wheel in the supply stream were determined by: X2 Z X1 Khdw ðX1 KXre Þ

(6)

and T2 Z

h2 K2500X2 1:004 C 1:875X2

(7)

where, X (kg/kg), h (kJ/kg) and T (8C) are the humidity ratio, specific enthalpy and temperature of air, respectively. The subscripts of 1,2 and so on in the following equations indicate the positions shown in the Fig. 2. Xre (kg/kg) is the humidity ratio of the regeneration air. h2 is determined by an approximate expression for the enthalpy of the mixture air and explained in the Eq. (11). In the heat recovery wheel, only the sensible heat exchange is in the process and the humidity ratio is always constant. The temperature of the supply air stream leaving the heat recovery wheel was: T3 Z T2 Khhw ðT2 KT6 Þ

(8)

For the humidifier, the enthalpy is approximately a constant in the evaporative cooling process. The humidity ratio and temperature of the air leaving the humidifier can be determined by: X4 Z X3 Khhd ðXs KX5 Þ

(9)

and T4 Z

h4 K2500X4 1:004 C 1:875X4

(10)

where, Xs (kg/kg) is the humidity ratio at saturation state. The modelling of condition along the entire process of desiccant cooling was carried out for multiple psychrometric calculation. Within the calculations, the values of specific enthalpy of the mixed humid air, h2 and h4, which are the energy content of humid air, were calculated by an approximate expression of the sum of the enthalpy of dry air and the enthalpy of water vapour related to the reference temperature 0 8C: h2 Z 1:005T2 C X2 ð2500 C 1:875T2 Þ h4 Z 1:005T4 C X4 ð2500 C 1:875T4 Þ

(11)

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The psychrometric equations from (6)–(11) were numerically solved to obtain the steady state psychrometric values for each position in the desiccant cooling process. In addition, the energy balance calculations based on the air temperature, humidity ratio and specific enthalpy at the inlet of the desiccant system and the room space as well as the solar air collector outlet were represented as: † The cooling energy generated from the desiccant system: Q_dsc Z m_ d ðh1 Kh4 ðWÞ

(12)

† The sensible cooling load removed from the room: Q_supply Z m_ d Cp ðTroom KT4 ÞðWÞ

(13)

† The regeneration energy from solar heating system: Q_reg Z m_ c Cp ðTreg KT8 ÞðWÞ

(14)

† The auxiliary cooling energy supplied to the room: Q_aux Z m_ d Cp ðTroom K25:5 8CÞðWÞ

(15)

† COP of the desiccant system: Q_dsc Q_reg Q_supply † Solar fraction Z Q_load COP Z COP

(16) (17)

Where, m_ d (kg/s) is the supply air mass flow rate of the desiccant cooling machine.

5. Temperature and humidity assessment The cooling process of the solar powered desiccant cooling system can be illustrated by a typical steady state psychrometric process from the model simulation results. For example, on 15th of June 2002 at 2:00pm, the warm moist air at 29.2 8C and 0.011 kg/kg moisture content (relative humidity is about 38.9%) was vent through the sorption wheel so that it came off at 41.3 8C and 0.0088 kg/kg moisture content (relative humidity is about 17.7%). The supply air stream then past through the heat recovery wheel and was sensible cooled to 24.6 8C. The cooled air then past through an evaporative humidifier with an adiabatic efficiency of approximate 85%. In this case, air can be supplied to the room space at 16.4 8C and 0.0108 kg/kg moisture content (relative humidity is about 93%). On the return air side, air from the room space at 25.5 8C and 0.0102 kg/kg moisture content (relative humidity is about 48%) was drawn through the evaporative cooler so that it entered the thermal wheel at approximate 18.6 8C and 0.011 kg/kg moisture content (relative humidity is about 91%). When the retune air stream past through the heat recovery wheel, it was sensibly heated to approximate 35 8C and exhausted to the outside.

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Table 1 Comparison of simulated and measured data for cooling process

Simulated 8C Kg/kg RH Measured 8C RH

Exterior air

After sorption

After heat recovery

After humidifier

Return air

After return humidifier

Air to outside

Regeneration air

29.2 0.011 38.9%

41.3 0.0088 17.7%

24.6 0.0068 –

16.4 0.0108 93%

25.5 0.0102 48%

18.6 0.011 –

35 0.011 –

69.5 –

29.2 37.6%

38.9 19.9%

26.5 –

17.2 94.8%

25.8 50.5%

20.2 –

37.4 –

61.7

On the regeneration air stream, the solar heated air at approximate 69.5 8C was vent through the sorption wheel to regenerate the desiccants. For validation purpose, a group of data measured from the Mataro library data acquisition system on 5th of June 2002 at 2:00pm were recorded in the Table 1 and compared with the simulation results given above. It can be seen that the simulated air temperatures and humilities along the air flow stream were quite close to the measured data. However, at some points, there were the reasonable differences between the measured and simulated values. This perhaps related to the process modelling accuracy or the system measuring errors. Then, in the simulation the desiccant system reached 16.4 8C inlet air temperature, which is the design value. In this study we also focused on the system dynamical simulation so that the solar energy used and the cooling energy consumed for summer time can be analysed. As the cooling machine relies purely on solar thermal heat from the ventilated PV-solar heat system, the regeneration air temperature cannot be assumed to be constant. Thus, the achievable solar fraction should be carefully analysed. So far a dynamical simulation based on the component models described before and the hourly meteor weather data of Barcelona has been carried out. For the computer simulation, the sophisticated control strategies of the system were simplified as the following supply/exhaust air volume control cascade: † Supply air volume control: if the room temperature and the outside temperature are below 24 8C and above 21 8C, the free ventilation operation will start—the supply and exhaust air volume is 3000 m3/h. If the room temperature is above 24 8C, the desiccant cooling starts operating and the supply/exhaust air volume is 6000 m3/h. If the room temperature is higher than 24.5 8C, the supply and exhaust air volume will be gradually increased to the maximum value, 12,000 m3/h. † Auxiliary cooling supply control: if desiccant system cannot supply satisfactory cooling energy to maintain the room temperature at 25.5 8C (or 26 8C), the auxiliary cooling starts operation. The supply and the exhaust air volume control cascade is illustrated in Fig. 3.

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3

12000m /hr

2

6000m /hr 2

3000m /hr

Aux. Cooler

Fig. 3. Supply and exhaust air volume control cascade.

Temperatures (C)

The complete solar powered cooling system was simulated for a volume flow on the fresh air side of 6000–12,000 m3/h (the volume flow rate was depending on the control strategies) and a variable volume flow of 3000–9000 m3/h on the regeneration side. The input irradiance, temperature and humidity values are taken from the Meteonorm database. From simulation, it can be seen that the temperature increase of PV-solar heater can reach from the exterior to the regeneration of 35–40 K at high irradiance levels. Approximate 70 8C of the regeneration air temperature can be obtained by a proper controller to vary the volume flow on the regeneration air side between 3000 and 9000 m3/h. Fig. 4 shows this effect. From Fig. 5, the room inlet air reaches 15–17 8C during cooling operation period in the first ten days of July. On days with low irradiance levels (for example the 6th of July) the regeneration air temperature is so low that no lower cooling effect is achieved. As ambient air temperature levels are also rather low on that day, this should not present a problem. During the free ventilation period (night), the room inlet air temperature is as the same as the ambient air temperature. Depending on the control strategies, the room air temperature can be controlled between 24 and 25.5 8C during the daytime. The maximum room temperature in July is not higher than 25.6 8C. Fig. 6 gives the absolute humidity of the ambient air, (supply air), room air and the air after the sorption wheel for the first 10 days of July. For humidity simulation, there are no control strategies implemented. So, the humidity level in the room only depends on the ambient humidity ratio and the efficiencies of the dehumidifier and the humidifier. However, the absolute humidity for the room inlet air is not allowed to be higher than 80 60 40 20 Tamb

Treg

0 1

25

49

73

97 121 145 169 193 217 241 First 10 days in July

Fig. 4. Regeneration and ambient air temperatures with simulation time.

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Temperatures (C)

1274 40 30 20 10

Tamb

Troom

Tsupply

0 1

25

49

73

97 121 145 169 193 217 241 First 10 days in July

Absolute humidity (kg/kg)

Fig. 5. The room, ambient and supply air temperatures with simulation time.

0.020 0.015 0.010 0.005 Xamb

Xroom

Xsupply

Xdhum

0.000 1

25

49

73 97 121 145 169 193 217 241 First 10 days in July

Fig. 6. Absolute humidity from simulation.

0.01 kg/kg (RH 95%, 15 8C) when the desiccant cooling system operating. The effective dehumidification of the sorption reaches about 0.005 kg/kg despite using the fresh air humidifier.

6. Cooling load of inf/aud room The solar powered desiccant cooling system in the Mataro library was designed to supply the cooling energy to the large reading room—Inf/Aud room. The Inf/Aud room is located in the west side and ground level of the Mataro library. The north wall, the large part of south wall and the ground floor of the Inf/Aud room are adjoining to the soil. The east wall and the ceiling are the interior partition. Thus, in the cooling load calculation for the Inf/Aud room, solar radiation contributes the thermal flux only to the west wall, the west window and the small part of the south wall. As required in the cooling load calculation, the room dimension, the construction material and the interior radiation of the Inf/Aud room are listed in Table 2. Consequently, the shading influence, the adjoin soil temperature and the wind speed were carefully considered and the Barcelona climatic data were used for the cooling load calculation over 12 months period. The average cooling loads of the Inf/Aud room for the August is around 9000 kWh.

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Table 2 The thermal information of the Inf/Aud room Wall description

Construction

Surface area

U_value (W/m2K)

West window

30.4 m!3. 2 mZ97.28 m2

4

East wall

Adiabatic two 0.02 m gypsum boards

South wall 1

Floor

Adjoining soil, with internal common block and face brick Not adjoining soil, with internal common block and face brick Concrete with insulation

Ceiling

Adiabatic 0.2 m lightweight concrete

30.4 m!1. 3 mZ39.52 m2 15.5 m!4. 5 mZ69.75 m2 30.4 m!4. 5 mZ136.8 m2 15.5 m!3. 2 mZ49.6 m2 15.5 m!1. 3 mZ20.15 m2 30.4 m!15. 5 mZ417.2 m2 30.4 m!15. 5 mZ417.2 m2 30.4 m!15. 5 m!4.5 mZ 2120 m3

0.334

North wall

Double glazing with 6 mm distance between two 4 mm glasses, 10% aluminium frame without any thermal break 0.25 m concrete insulation with 0.007 m steel cover Concrete wall completely adjoining soil (insulated)

West wall

South wall 2

Room volume

Interior heat radiation

0.530 0.406 0.530 0.334 0.460 0.406

40 persons, 6000 w lights and computers

7. The energy balance One of the important aims of the project was being capable of calculating the energy balance of the system by the models simulation so that the valuations of system performance and cost can be forecasted. Using the system models developed above, the energy balance has been calculated relating to the Barcelona climatic data and the system control strategies. For cooling purpose, the cooling energy generated by the desiccant cooling system was calculated using the enthalpy difference between the ambient air and the room inlet air from April to October. The total amount of 40,639 kWh cooling energy obtained refers to the total enthalpy reduction from the ambient air to the room inlet air. The regeneration energy for driving the desiccant cooling machine, for the period of April to October, from ventilated PV-solar air heating system was calculated as 68, 829 kWh. Table 3 The Cooling and regeneration power for summer season

Q_dsc (kWh) Q_reg (kWh) COP

April

May

June

July

Aug

Sept

Oct

Total

501 3356 0.149

2692 7120 0.378

6562 12,790 0.513

11,831 18,103 0.654

11,395 15,686 0.726

6624 9832 0.674

1034 1942 0.533

40,639 68,829 0.518

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Table 4 The supplied cooling energy to the room

Q_supply (kWh) (at 3000 m3/h) Q_supply (kWh) (at 6000 m3/h) Q_supply (kWh) (at 6000–12, 000 m3/h) Q_supply (kWh) (at 12,000 m3/h) Total Q_supply (kWh)

April

May

June

July

Aug

Sept

Oct

Total

4285

3428

2098

1006

1214

1764

2810

16,605

750

1234

2543

2592

2128

1981

753

11,981

344

1305

2026

3100

2473

1443

0

10,691

0

223

528

1739

1209

604

0

4303

5379

6190

7195

8437

7024

5792

3563

43,580

The coefficient of performance (COP), i.e. the ratio of produced cooling power to required regeneration power, is on average 0.518 for this application. Table 3 shows the monthly data of the regeneration energy, the desiccant cooling energy and the COP calculated. Using the room temperature calculated simultaneously for the aud/inf room, the cooling energy effectively delivered to the room was obtained which depends on the enthalpy difference between the room air and the room inlet air. The total cooling energy supplied to the room from the desiccant cooling system is about 43,580 kWh. Depending on the control strategies, both free ventilation and cooling ventilation were applied to the system and three different ventilation air volumes were used. With respect to the total cooling energy supplied to the aud/inf room, the free ventilation cooling energy (3000 m3/h supply volume flow rate) is about 38%; the desiccant cooling energy supplied at the volume flow rates of 6000 m3/h and between 6000 and 12,000 m3/h are 27 and 24%, respectively; only 9% desiccant cooling energy is supplied to the room at 12,000 m3/h volume flow rate. Table 4 shows the detailed monthly values. With respect to the control strategies, the auxiliary cooling energy starts operating when the room temperature is higher than 25.5 8C. Therefore, the auxiliary cooling energy was calculated from the simulation, which can keep the room temperature at 25.5 8C during the high cooling load periods. If considering the total amount of desiccant cooling supply plus the auxiliary cooling energy as the total cooling demands for the aud/inf room, the monthly solar fraction for desiccant cooling system was calculated and given in Table 5. In average, 93% of the cooling demand for summer season can be covered by Table 5 The solar fraction

Total Q_supply (kWh) Q_aux (kWh) Q_supplyC Q_aux (kWh) Solar fraction

April

May

June

July

Aug

Sept

Oct

Total

5379 0 5379

6190 21 6211

7195 131 7326

8437 486 8923

7024 2099 9123

5792 605 6397

3563 0 3563

43,580 3341 46,921

1.000

0.993

0.975

0.939

0.735

0.869

1.000

0.930

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Table 6 Efficiency study

Humidifier

Dehumidifier

Heat recovery

Efficiency

Q_dsc (kWh)

Q_supply (kWh)

Q_aux (kWh)

0.8 0.9 0.95 0.7 0.8 0.9 0.7 0.8 0.9

11,831 11,072 10,609 11,256 11,831 12,334 10,368 11,831 12,888

8439 8673 8726 8395 8439 8481 7979 8439 8711

486 355 300 533 486 448 780 486 278

the solar powered cooling energy. At least, 73.5% of the cooling load can be removed by the desiccant cooling in August.

8. Discussion The results of the studies described above indicate that the amount of the cooling energy supplied to the room space is sensible to the system parameters, such as the efficiencies of the sorption wheel, heat recovery wheel and humidifier. If the efficiencies for the three main components varied between 0.7 and 0.9 (in the above simulation, the efficiencies are fixed as 0.8), the cooling energy supplied to the room also has G5.4% change. This results to the changes of the auxiliary cooling supply amount and the solar fraction value. The cooling energy generated from the desiccant system, the cooling energy supplied to the room and the auxiliary cooling energy are given in Table 6 with the different efficiencies for the simulation of the typical month of July. In this study, the electrical consume of three ventilation fans was considered. The total electric consumption of fans was calculated for the different three ventilation flow volume rates. 13,557 kWh of electric energy could be used for summer season operation for the desiccant cooling. This electrical consume should be added to the auxiliary cooling and the solar fraction of the system should be analysed carefully. It is necessary to indicate that the solar powered desiccant system can be operated in heating mode for winter season. When operating in heating mode, the combination of solar energy and recovered heat from the exhaust air stream appears to be capable of one-third of the total heating energy required during the winter months. It also can be seen that most of the auxiliary heating energy supplied to the room is during the no-solar period of the day-evening and night.

9. Conclusion In summary, building integrated ventilated photovoltaic fac¸ade and PV sheds can significantly contribute to the thermal energy needs of buildings in summer season.

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The typical application of the thermal energy transmitted from PV fac¸ade and PV sheds is to use it to drive a solar powered desiccant cooling system for buildings cooling. In this paper, the establishment of the component models of a desiccant cooling system regenerated by solar heat was described. The simulation study of the system performance was carried out. When simulating the system operation process including the control strategies, the energy balances for the building cooling were yield. The useful solar thermal energy observed by ventilated PV-solar heating system and the effective cooling energy generated from the desiccant cooling system have been calculated. With respect to the regeneration energy from the solar air system, an average COP of 0.518 was obtained. The supplied cooling energy from the desiccant cooling system for summer season were estimated as 43,580 kWh which can achieve a solar fraction of 0.75 at least. The results of the studies described in this paper demonstrate the potential of desiccant cooling systems combined with ventilated PV fac¸ade/sheds. The results also confirmed that the ventilated PV fac¸ade/sheds can produce the same order of magnitude of thermal energy as the photovoltaic electric production for buildings heating and cooling. Therefore, the concept of building integrating ventilated PV fac¸ade/sheds has the good potential in the future.

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