Category Archives: EuroSun2008-7

Experimental validation

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The experimental installation presented in section 2 is used to validate the above presented model. Different inlet conditions were considered with temperature varying form 25°C to 38°C and humidity ratio from 10 to 15 g/kg and for different regeneration temperature (55°C, 60°C 65°C, 70°C and 80°C).

The figure 2 [9] plots the temperature error versus the humidity ratio error for the most significant points. It shows the maximum predicted error 0.4 g/kg and root mean squared error (RMSE) is 0.22 g/kg. For the temperature and the maximum committed error in the outlet temperature is 1 °C for RMSE of 0.65 °C. It must be noted that the uncertainty in the measurement of the mean values is 0.7°C for the temperature and 0.3 g/kg. It shows that the committed errors are in the domain of the uncertainty of the measurements.

In the following section the experiments are performed on the desiccant using experimental measurements and numerical results in order to perform the sensitivity analsysis.

Description of the novel generator design

In this heat and mass transfer prototype, the desiccant solution and the air stream is brought into contact with each other in a cross flow configuration. The regenerator consists of a stack of polycarbonate (PC) twin wall plates. Each plate has an internal heating water circuit. The plates are covered with a wick fibre to facilitate intimate contact between the liquid desiccant and air. This is done to increase the exposure time over the plates and thereby enhance the desired mass transfer and heat exchange. The general design is shown in the left part of Fig. 2.

The advancement of the distribution system consists of:

• a separation between the liquid desiccant distributor and the contact area between air and desiccant. The LiCl solution is distributed over the textile separately in an advance stage i. e. before coming in contact with the air stream. A separator is used to split the generator into two chambers. In the upper small chamber LiCl throttled through perforated plexiglass tubes, completely diffused through the textile fibres and trickles down to come in contact with the horizontal air stream in the lower large chamber. This separator will prevent LiCl particles to drift into the air stream, resulting in a zero carryover regenerator.

• the employment of promising fibres, a new natural man made fibre produced from wood pulp. This fiber can absorb 50% more moisture than cotton did, it has a lower sorption index than cotton, which means higher transport speed of the liquid.

• the employment of perforated plastic tubes with different tested throttling-points diameters.

The LD distributor in this design uses a plurality of parallel plexiglass tubes to horizontally distribute the LD over the wick. The parallel tubes extend outwardly from openings in the lower edge of one of the sides of a LD feed box and are closed from the free end. The tubes penetrate the PC plates horizontally and deliver the desiccant solution over the coated plates at a number of equally spaced-apart locations (discharge-holes). The distributor is shown on the right part of Fig.

2.

The discharge holes are preferably formed on the plexiglass tubes by using a CNC machine to make the fine drip points with high precision. The size and number of the discharge holes are selected to provide the desired liquid flow. The throttling-point density determines the flow into the wick material covering the PC twin-wall plates. Likewise, the distance between the discharge — holes is selected to accommodate the desired LD flow rate with the maximum even distribution.

image618 image619
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The regenerator consists of an upper and a lower hot water-feeder box. Each water feeder box consists of four equal chambers separated by baffles. Hot water will pass through the internal passages of the PC twin-wall plate in a 3-loop serpentine path. It will enter the first chamber from a primary opening in the upper wall of the chamber, and it collides with a liquid spreading surface that faces the internal PC passages. The spreading surface will also disrupt the downward flow of water and will direct it equally to the channels.

Fig. 2. Exploded view of generator design along with the illustration of flow directions of heating water, LD and air (upper left), an isometric enlargement of the LD distribution system and a top-view of how the glass tubes penetrate the polycarbonate twin-plates (upper right).

Operation day in the cooling mode

In Erro! A origem da referenda nao foi encontrada. an example of the operation during a normal summer day is shown. The diagram corresponds to June 13th, 2007. Cooling operation starts at 7:30 and ends at 14:30. In the morning till 11:30 the chiller is operated with heat from the heating net as the temperature in the solar storage was not sufficient to operate the machine. The shaded area between 11:30 and 14:30 shows the time when the system was operated with heat from the solar system.

Large temperature variations can be observed in the driving circuit. While the cyclic temperature peaks observed in the return flow from the chiller in all circuits (T_HT_out, T_MT_out and T_NT_out) are characteristic of the periodic working adsorption machine, the variations in the driving inlet temperature (T_HT_in) are due to the dynamic behaviour of the plate heat exchanger within this circuit but also variations in the temperature of the heating net. These variations were not expected. A much more stable inlet temperature is observed during the solar operation period. Further it can be seen, that the boreholes effectively attenuate the temperature peaks giving smooth and constant feed temperatures (T_MT_in) to the machine.

4.1. Heating operation

For the heating operation the adsorption chiller is operated as a heat pump. For this operation the evaporator is connected to the borehole system and the heat rejection to the heating coil in the air handling unit.

Sensitivity results and analysis

4.1 Results

Experiments have been conducted on the desiccant wheel with the parameters (e. g. outside temperature, outside humidity ratio, regeneration temperature and regeneration humidity ratio) varying in the range defined in the section 2. The complete DOE of 4 parameters operating between 2 levels needs 24=16 experiments. When a combination of the studied parameters was

difficult to achieve experimentally (3 experiments) the results of the model was used to complete the DOE. Table 1 below shows the results

Table 1. Dehumidification rate of the desiccant wheel for different operating conditions

Ti

w1

T8

W8

w1-w2

T1

w1

T8

w8

w1-w2

[°C]

[g/kg]

[°C]

[g/kg]

[g/kg]

[°C]

[g/kg]

[°C]

[g/kg]

[g/kg]

25

11

55

10

4.8

35

11

55

10

2.9

25

11

55

15

3.6

35

11

55

15

1.7*

25

11

75

10

6.8

35

11

75

10

5.2

25

11

75

15

5.8

35

11

75

15

4.1

25

14.5

55

10

5.4*

35

14.5

55

10

4.1

25

14.5

55

15

4.3

35

14.5

55

15

2.95

25

14.5

75

10

8

35

14.5

75

10

6.7

25

14.5

75

15

7

35

14.5

75

15

5.6*

*

Calculated by

the model

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Solving 16 equations for the effects identified in the equation (1). The dehumidification of the desiccant wheel is then written in function of the operating parameters

4.2 Analysis

Form the equation identified we notice a constant dehumidifying effect of 4,987 g/kg which will be increased or decreased depending on the operating conditions. In order to compare the effect of each parameter with the constant effect, they are divided by this later and a graphical comparison can thus be established. The figure below shows the effect of each parameter.

As commonly known, figure 3 shows tha the regeneration temperature has the most significant impact on the dehumidification performance of the desiccant wheel. In the same time outside conditions has an important impact on too.

• Increasing the regeneration temperature from the mean value (65°C) to its upper limit will increase the dehumidification 23%

• Increasing the outside temperature to its higher limit will decrease the dehumidification of 17%.

• Increasing the outside humidity ratio will increase the performance of the desiccant wheel of 13%

• Increasing the regeneration humidity ratio will decrease the performance of the desiccant wheel of 8%.

The reason behind these observations is the vapor pressure difference of the air and that of the surface of the silica gel. This vapor pressure difference is the driving force of the adsorption phenomena. Reminding that if the desiccant is cold the vapor pressure at its surface. For the moist air the temperature does not have a significant impact on the vapor pressure while humidity ratio does.

The desiccant coming for the regeneration sector is hot and dry, it is first cooled down by the outside air in the process sector and then the adsorption occurs.

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Fig. 3 Effect of the operating conditions and their combinations

So if we increase the temperature of the outside air its vapor pressure will not increase significantly in reversal it will not cool sufficiently the hot desiccant form arriving from the regeneration sector yielding a high vapor pressure at the surface of the desiccant and thus the adsorption will be less efficient. If we increase the humidity ratio of the outside air its vapor pressure will increase significantly yielding a higher vapor pressure difference thus increasing the adsorption. When increasing the regeneration temperature, the regeneration air vapor pressure does not increase significantly while the desiccant is heated and its vapor pressure increases leading to an efficient drying of the desiccant. This desiccant leaving the regeneration sector is dry and thus having fewer vapor particles, yielding a low vapor pressure at its surface and thus high adsorption. When increasing the regeneration humidity ratio we will increase the vapor pressure of regeneration air stream, reducing the transfer from the desiccant to the regeneration air, yielding less drying of the desiccant. When the sorbent will leave the regeneration with more water vapor at its surface its vapor pressure is then higher, reducing thus the adsorption.

This clearly shows the limitations of the desiccant cooling technique regarding outside conditions. It demonstrates that high outside temperature reduces significantly the performance of the desiccant wheel. Regarding the outside humidity ratio even if the dehumidification increase with increasing outside humidity ratio, we noticed that for outside temperature beyond 30°C the maximum dehumidification rate is 6 g/kg. Taking into account the maximum humidity inside the building (e. g. 11.8 g/kg) and the humidification across the supply humidifier we conclude that the maximum outside humidity under which a desiccant system will operate efficiently is 14.5 g/kg.

3. Conclusion

In this paper a sensitivity analysis based on the design of experiments was conducted on a desiccant wheel using mainly experimental results. A numerical model was validated experimentally and provided the missing combination of the experiments. The impact of the outside and regeneration conditions on the dehumidification rate of the desiccant wheel was studied. As widely known the regeneration temperature has the most significant impact on the dehumidification rate, but the impact of the outside temperature, outside humidity ratio and the

regeneration humidity ratio are very important as well. These results showed that desiccant cooling is an interesting option for moderately hot and moderately humid climates

Nomenclature

Cpa

heat capacity of air [J. Kg-1.K-1]

NUT

number of transfer unit [-]

cpv

heat capacity of water vapour

RMSE

Root mean squared error

[J. Kg’.K-1]

[unity of the variable]

cpm

heat capacity of the regenerator matrix

t

time [s]

[J. Kg-‘.K-1]

T

temperature [K]

C

heat capacity (J. Kg"1.K"1)

T

a

air temperature [K]

Fi

potential characteristic [-]

T

eq

equilibrium temperature [K]

ha

enthalpy of moist air [J. Kg-1]

Tm

matrix temperature [K]

H

enthalpy of the desiccant [J. Kg-1]

u

fluid velocity (m. s-1)

Jt

lumped heat transfer coefficient [s-1]

Wa

humidity ratio of moist air. [Kg/Kg]

Jm

lumped mass transfer coefficient [s-1]

Wd

water content of desiccant [Kg/Kg]

Le

Lewis number [-]

z

coordinate in the flow direction [m]

ma

air mass flow rate [Kg. s-1]

T

ratio of matrix mass over air mass [-

Md

mass of the desiccant matrix [Kg]

n

efficiency [-]

Mm

mass of the aluminium matrix [Kg]

p

density [Kg. m-3]

N

angular speed of the wheel [rd. s-1]

Yi

parameter [-]

Acknowledgments

The authors would like to thank Mr Michel Burlot for his valuable technical support on the experimental

installation. This work was supported by ADEME (French Agency of the Environment and Energy

Management) and the regional council of Poitou Charentes.

References

[1] P. Bourdoukan., E. Wurtz., P. Joubert., M. Sperandio,. (2008). Potential of solar heat pipe vacuum collectors in the desiccant cooling process: modelling and experimental results. Solar Energy. 0.1016/j. solener.2008.06.003

[2] J. Goupy, (2001). Introduction aux plans d’experiences, Dunod.

[3] P. J. Banks, (1972). Coupled equilibrium heat and single adsorbate transfer in field flow through a porous medium — 1. Caracteristics potentials and specific capacity ratios. Chemical Engineering Science, 27, 1143-1155.

[4] I. L. Maclaine-cross, P. J. Banks, (1972). Coupled heat and mass transfer in regenerators — prediction using analogy with heat transfer. Int. J. Heat Mass Transfer, 15, 1225-1242.

[5] J. J. Jurinak. (1982). Open cycle solid dessicant cooling — component models and system simulation, Ph. D, Winconsin, Madison.

[6] P. Stabat, (2003). Modelisation de composants de systemes de climatisation mettant en oeuvre l’adsorption et l’evaporation d’eau, Ph. D, Ecole des Mines, Paris.

[7] W. M. Kays, A. L London, (1984). Compact Heat Exchangers, McGraw-Hill, New York.

[8] SPARK. (2003). Simulation Problem Analysis and Research Kernel. LBNL, Berkeley, California.

[9] P. Bourdoukan., E. Wurtz., P. Joubert., M. Sperandio, (2008). Critical efficiencies of components in desiccant cooling cycles and a comparison between the conventional mode and the recirculation mode."

In proceedings of ECOS conference, Krakow, Poland, 1, 435-443

Experimental setup

Two different sets of experiments were carried out to prove the advancements on the distribution system.

Подпись: Fig. 3. Experimental setup to determine the optimum size and number of discharge points for the distributor (upper), and to check the LiCl diffusion through the textile (lower).

1. The objective of the first test is to determine the optimal size and number of discharge points distributed along the plastic tube. The optimized tube will serve the LD flow rate required by the application and ensure fairly equal amounts of the LD throttled-out through each discharge bore. Five perforated plastic tubes are tested. Each tube is perforated by a CNC machine in equally-spaced positions. H2O/LiCl (30% salt concentration by weight) is used as the liquid desiccant with three volume flow rate values. In order to investigate the effect of the distance between the LiCl feed-box and the discharge-point location, different positions along the tube are selected. Transparent PVC hoses are used to enclose the discharge-bores, without affecting the throttled LiCl. The free ends of the PVC hoses are connected to plastic bottles in order to collect the LiCl throttled out of the intended discharge-bore. Fig. 3 shows the experimental setup.

2. The second test is a comparison between different types of textiles (100% cotton, 100% viscose, 100% polyester, 100% polyamide and wood pulp based textiles) with different thicknesses and different weaving. Each type of the mentioned textiles has been tested to measure the absorption capacity and the diffusion speed by simply pouring a quantity of LiCl onto the upper surface of a taut piece of textile, and measuring the needed time for the LiCl droplets to be completely absorbed by the textile fibers. Each type has been tested in both, dry and wet state. Then PC plates were coated with the textiles which show the best absorption and diffusion speed, and exposed to a LiCl solution throttled through the perforated plexiglass tube. A violet fluorescent light and dyes have been used to support the visual inspection of LiCl distribution through the textile fibers. Fig. 3.

Absorption chiller

AAWC building is a newly reconstructed five-storey building with approximately 3300m2 total area. Space heating is based on fan-coils system and 300kWgas boiler. The estimated cooling load of the building is about 150kW. In summer time the ambient temperature in Yerevan often reaches 400C. For cooling needs single effect hot water lithium bromide absorption chiller YORK Model YIA-HW-1A1-50-C-S-C has been installed in the basement of the building near the gas boiler. The chiller is shown on Figure 3.

The absorption chiller has nominal cooling capacity of 420kW and is connected to the VXT 70 cooling tower with 400kW capacity placed on the roof top of the building 25m above the boiler house. The cooling tower can be seen on the roof of the building on the right side of the PV array (Fig.1).

In cooling season the gas boiler provides hot water to the absorption chiller installed inside the same boiler house. Chilled water enters the fan-coils system of the building. The gas boiler provides hot water for chiller’s generator loop at 90°C/70°C supply/return temperature. For similar applications the hot water could be supplied by appropriate solar water heating system with gas boiler back-up. The water temperature in absorber/condenser-cooling tower loop is 290C/ 330C, and chilled water supply/return temperature is 80C/120C.

The power need of the chiller is about 24kW including chiller’s solution, refrigerant, purge pumps and cooling tower fan and water pump. A conventional compression chiller with

Подпись: Figure 3 Lithium bromide absorption chiller at the AAWC
similar cooling capacity would require almost 10 times more electric power and, consequently, too large size cabling from building to transformer sub-station. .

Absorption chiller is an attractive energy efficiency solution, particularly when gas or solar heat source is available.

The absorption chiller has been tested with building’s fan-coils system and commissioned in June2008. The building is still under renovation and at the moment there is no need for absorption chiller which is kept in standby mode.

2. Conclusion

The 10kW grid connected BIPV station and 420kW lithium bromide absorption chiller are installed, tested, and ready for operation and will be monitored during operation.

Their performance monitoring data, particularly related to the power consumption and hot water supplied to absorption chiller, can be used when considering combinations of solar PV and/or water heating system with similar absorption chiller. The performance of the liquid absorption chiller can be compared in the same climatic conditions with the performance of the solar driven desiccant cooling system operating in Yerevan since February 2002 [2].

The combination of the grid connected PV power station and lithium bromide absorption chiller on the same place provides also the opportunity for determination of produced PV power’s

contribution to chiller’s power consumption in different cooling conditions. The results of the study can be useful for countries with similar climate conditions.

References

[1] http://www. renewableenergyarmenia. am

[2] EU INCO-COPERNICUS Programme. Design and Installation of Solar Driven Desiccant Cooling Demonstration System, Contract #ICOP-DEMO-4034/98. INETI (PT), AUA (AM), FhG-ISE (D), CONTACT-A (AM), INTERSOLARCENTER (RU), Final Report, 2002

Experimental results and discussion

Two different kinds of experiments have been performed with regard to the regeneration conditions. In the first series of experiments constant regeneration humidity has been used. This allows the comparison of the FAM wheel data with other desiccant wheels. However, it prevents predictions for the wheel performance in a desiccant-evaporative system where ambient air is heated up for regeneration purposes. Obviously, depending on the ambient air conditions, the regeneration air conditions will vary, too. Therefore a second series of experiments has been conducted with the
regeneration humidity matching the ambient (supply) humidity. Constant parameter in both series of experiments included desiccant wheel speed of 20 1/hr, supply/regeneration air flow of 400 m3/hr, a counter flow arrangement and a resulting air velocity at the wheel face of 3.25 m/s. The wheel face area has been split equally between supply and regeneration stream.

Cold storage configuration

The basic scheme in this case is shown in the following figure:

image254

Fig.4. Configuration with cold storage

In this case, to have cold storage allows having an intermediate case between the two previous, so that they can be stored excess production and face bigger loads of instantaneous production of the chiller.

The cold storage has been determined to have an autonomy of one hour. And so that, from the flow of the evaporator (that coincides with the one of cold load) it’s determined the volume necessary.

3. Execution of the simulations

As it was mentioned before, for each one of those three schemes they have been tested different control strategies: three for the solar camp (C1, C2 y C3) and another three for the chiller (C4, C5 У C6).

With regard to the solar part:

C1: Constant flow control on the solar part and constant flow on the absorption.

In the case of the solar pump, it has been used the output temperature of the collectors as variable to control, as well as the reference variable is the one at the bottom of the tank. It was used hysteresis, and the values adopted were: upper for the start 6 °C, and for the stop 1 °C. To start up the absorption the storage temperature and the room temperature of the zone to be conditioned are taken into account.

C2: Control with variable flow on the solar part starting by differential temperature and constant flow on the absorption.

The system is similar to the one on the previous case, being the start up in function of a differential temperature but using flow regulation for the solar pump. In this way can be compared the improvement using constant flow on the solar flow.

C3: Control with variable flow on the solar part starting by critical radiation and constant flow on the absorption

In this case the start up of the solar pump is made in function of the critical radiation and the flow is regulated to maximize the output temperature of the collectors.

Regarding to the absorption chiller: For regulation on the following schemes, it was chosen for the solar part the C3 strategy and some changes have been applied to the control of the chiller in order to compare their effect using a constant system for the solar control. For the control of the chiller, once there were the necessary conditions for its start up, was established the control of the corresponding variable, in order to obtain a power produced by the evaporator, equal to the one demanded by the system. Thus facing the real demand in the system taking into account the higher delay produced by the effect of the exterior conditions on the load.

C4: Control with variable flow on the solar part starting by critical radiation and control by temperature on the generator for the absorption

In this case there is a three way valve on the suction of the generator pump, that allows controlling the temperature on the input, and with it the output power for the evaporator.

C5: Control with variable flow on the solar part starting by critical radiation and control with variable flow on the generator using a diversion valve.

Other possibility to regulate the power of the generator is the variation of the flow on the chiller. In this case all the water of the solar part is driven, and a three-way valve is used to let pass more or less water through the generator in function of the demand.

C6: Control with variable flow on the solar part starting by critical radiation and control with variable flow on the generator by means of a pump.

In this case, the flow regulation on the generator has been performed by means of pump with variable flow.

4. Results

Three tables are shown next with the comparative of the different data obtained:

As can be seen, two of the configurations are better than the rest: on the configuration of the solar part the one with cold storage, and with regard to the performance of the performance of the absorption chiller, the one with infinite storage. The fact of not having to fulfill the condition of demand of the load, allows a better use of the available energy. On the other hand, the cold storage allows having a bigger impact over the temperature of the solar installation, what increase the solar collection, but in return, penalizes the chiller COP.

Table 1. Results for the simulations with direct use

C1

C2

C3

C4

C5

C6

Energy produced on the collectors [kWh]

7.400

8.650

8.640

8.590

8.650

9.100

Performance of the collectors [p. u.]

0,35

0,32

0,41

0,40

0,41

0,43

Energy given to the generator [kWh]

5.770

6.710

6.670

6.700

6.650

7.660

Energy produced on the evaporator [kWh]

2.575

3.050

3.040

3.180

3.030

3.880

Energy given to the load [kWh]

2.575

3.050

3.040

3.180

3.030

3.880

Medium COP of the chiller [p. u.]

0,45

0,45

0,46

0,47

0,46

0,51

Load Fraction [p. u.]

0,50

0,59

0,59

0,61

0,58

0,75

Solar COP [p. u.]

0,16

0,14

0,19

0,19

0,19

0,22

Table 2. Results for the simulations with total use

C1

C2

C3

C4

C5

C6

Energy produced on the collectors [kWh]

9.100

8.650

8.800

8.800

8.740

8.740

Performance of the collectors [p. u.]

0,43

0,41

0,41

0,41

0,41

0,41

Energy given to the generator [kWh]

8.100

7.850

7.950

7.650

7,650

7.950

Energy produced on the evaporator [kWh]

3.550

4.600

4.600

4.450

4.520

4.770

Energy given to the load [kWh]

3.550

4.600

4.600

4.450

4.520

4.770

Medium COP of the chiller [p. u.]

0,44

0,59

0,58

0,58

0,59

0,60

Load Fraction [p. u.]

0,69

0,89

0,89

0,86

0,87

0,92

Solar COP [p. u.]

0,19

0,24

0,24

0,24

0,24

0,25

Table 3. Results for the simulations with finite storage

C1

C2

C3

C4

C5

C6

Energy produced on the collectors [kWh]

8.850

8.850

9.050

9.800

9.000

9.500

Performance of the collectors [p. u.]

0,42

0,42

0,43

0,46

0,42

0,45

Energy given to the generator [kWh]

7.210

7.300

7.350

8.270

7.280

8.250

5.

Подпись: 3.150 3.250 3.250 3.270 3.250 3.710 3.015 3.150 3.160 3.180 3.150 3.590 0,44 0,45 0,44 0,40 0,45 0,45 0,59 0,61 0,61 0,61 0,61 0,69 0,18 0,19 0,19 0,18 0,19 0,20

Conclusions

Different control actions on the solar part have repercussion on the chiller performance and vice versa, different control systems for the chiller, change the working conditions of the solar installation.

From the previous results, we can deduce that the configuration with total use is the one that better performance shows, as it has more hours of use than the rest of configurations.

As well, it can be seen how the most adequate control for the solar installation is the one with variable flow, combined whit start based on critical radiation. The control for the chiller that better results offers is the one based on variable flow by means of the pump. The cold storage, reduces the performance of the installation, but from the point of view of design, allows the use of a smaller absorption chiller.

References

[1] National Renewable Energy Laboratory. User’s Manual for TMY2s (Typical Meteorological Years), NREL/SP-463-7668, and TMY2s, Typical Meteorological Years Derived from the 1961-1990 National Solar Radiation Data Base, June 1995, CDROM. Golden: NREL, 1995

[2] TRNSYS 16 Documentation.. A transient Simulation Program. Solar Energy Laboratory, University of Wisconsin, Madison, 2006.

[3] SACE: Solar Air Conditioning in Europe. Final Report, EU Project NNE5-2001-25, 2003

[4] ESESA (1996), Manual de instalacion: Grupos refrigerantes por absorcion de agua caliente WFC-10.

[5] M. Kovarik, F. Lesse, Optimal control of flow in low temperature solar heat collectors, Solar Energy 18 (1976) 431-435

[6] C. B. Winn, D. E. Hull III, Optimal controllers of the second kind, Solar Energy, 23 (1979) 529-234.

[7] P. Dorato, Optimal temperature control of solar energy systems, Solar Energy, 30 n 2 (1983) 147-153

[8] D. Zambrano, E. F. Camacho, Application of MPC with multiple objective of a solar refrigeration plant., proceeding the 2002 IEEE International Conference on Control Applications, 2002.

[9] M. A. Corchero, M. G. Ortega, R. R. Rubio. .Aplicacion del control robusto H® a una planta solar. XXV Jornadas de Automatica de Ciudad Real. 2004

[10] J. C. Blinn. (1979) Simulation of solar absorption air conditioning. Ms. D. Thesis University, of Wisconsin-Madison.

[11] R. Lazzarin, Steady and transient behaviour of LiBr absorption chillers of low capacity, Revue Internationale du Froid, 3, n 4, (1980), 213-218.

[12] J. A. Duffie, W. A. Beckman, (1991). Solar engineering of thermal processes 2nd edition, John Wiley & Sons, New York.

Control Strategies for Solar Thermal Cooling System in Office Building in Almeria, Spain

J. Bote Garcia, A. Gal, D. Tavan*

Energy Efficiency & Comfort Group, R&D Division, Acciona Infraestructuras
Calle Valportillo II, 8, 28108 Alcobendas, Spain

* Corresponding Author, dtavan@acciona. es

Abstract

This article gives an insight on working solar-driven cooling systems that were integrate in an office building used by researchers at the Plataforma Solar de Almeria. Implemented systems combined both passive and active solar cooling techniques in order to minimize the energy consumption of HVAC systems while maintaining an adequate thermal comfort inside the building. The control strategy is briefly described whose goal is to maximize the use of renewable sources over conventional ones whenever they are available. This is the key to achieving a reduction by 80 to 90% of overall energy consumption with regard to typical energy needs for office buildings.

Keywords: Solar driven cooling, Absorption Heat-pump, Night cooling, Radiative cooling, Supervision control and data acquisition

1. Introduction

With the ARFRISOL (Bioclimatic Architecture and Solar Cooling) project, the Spanish Ministry of Innovation and Science is promoting energy efficiency in office buildings. As part of the project objectives, five office buildings are to be built or rehabilitated in different climatic zone of Spain in order to experiment various energy saving and renewable energy techniques and demonstrate that the energy consumption of office buildings can be cut by 80-90%, while the remaining conventional energy consumption can be supplied by active solar systems (e. g. solar thermal collectors for heating and cooling and photovoltaic panels for electricity). Acciona Infraestructuras is responsible for the construction of one of these high energy performance buildings situated in the Solar Platform of Almeria (PSA) in the south of Spain. The purpose of this paper is to explain the solar cooling techniques used in this building and the control strategies that have been designed to achieve high energy performance.

Technical specifications of Ao Sol’s solar chiller

The novel chiller unit is developed for the year-round thermal energy supply of residential buildings under Mediterranean conditions. Envisaged are detached houses with a heated area from 150 to 250 m2 or more. This building segment was chosen for the high relevance of air­conditioning in buildings with high living comfort.

The present development is based on previous work accomplished at the University of Lisbon, Instituto Superior Tecnico, by Prof. Mendes [5], who developed several water-cooled ammonia/water absorption chillers. The thermodynamic cycle of the present solar absorption chiller has been consequently re-calculated and fine-tuned for direct air-cooled operation. The absorption unit is designed to work using solar thermal energy from CPC collectors or waste heat from a cogeneration engine fuelled by biomass. For the use linked to solar collectors a back-up gas burner is foreseen. The device is designed to produce 8 kW of chilled water in a range of temperatures between 5 °C and 18 °C. The chilled water is normally used in buildings either for air-conditioning via fan coil elements or directly for space cooling via radiant ceilings. The chiller developed can be operated for both concepts. Moreover, using ammonia as the refrigerant, ice production or even deep refrigeration could be achieved via changes in the control strategy. The solar chiller is driven by hot water at temperatures at around 95 °C, stored in a solar buffer. The device directly dissipates the own waste heat produced at around 40 °C without any need for an external wet, hybrid or dry cooling tower.

The detailed technical distinctive features of the AO SOL solar chiller are summarized below:

The chiller is air-cooled. Overall sizes of the chiller and operation cost are minimized. No water is needed; this ensures market compatibility even in the extremely hot and dry continental regions of the peninsula’s interior. The absence of a cooling tower decreases considerably the overall dimensions compared to concurrent products.

Plate heat exchangers are used wherever possible in the thermodynamic cycle and take advantage of the high efficiency typical of this technology (up to more than 95 %!) coupled with very compact dimensions and low specific weight (kg/kW).

The remote monitoring system acting via mobile phone technology ensures complete safety. A timely alarm transmission with the possibility of intervening remotely on the chiller control adds to the on-site safety features and it minimizes the service needs.

2.2. Control strategy

The concept envisaged for the chiller is completely self-sufficient. Automatic procedures for start­up, load changes, shutdown, and safety issues have been implemented. The control acts on two 3­way valves, which stabilize the water temperatures in the hot and chilled water loops. Further, it triggers the adjustment of fan and solution pump speed, and the actuation of the refrigerant throttle valve. The chiller is controlled through the adjustment of temperatures in the hot and chilled water loops. The adjustment is activated by means of PID controllers. The machine reaches stable operation within 15 minutes. The automatic shutdown procedure is immediate; the machine reacts quickly and without swings. A new start without any manual intervention in between is secured.

The chiller is the core component of a harmonised chain of components representing a turn-key solution: an absorption cooling machine driven by hot water and producing chilled water for fan coil or radiant ceiling use, and Ao Sol’s unique solar CPC MAXI collector, which will deliver on the order of 25% more energy than a very good flat plate collector of today [6].

The ideally combined system will provide hot water all year around, heating in winter and cooling for half of the year. The control unit developed combines and actuates all components of the solar

heating, cooling and DHW system, from the solar collector, to solar buffer and backup, up to the chiller and the space to be conditioned. Figure 3 shows the flow scheme.

3. Experimental results

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The prototype has been taken successfully into operation. The novel machine showed a satisfactory behaviour in the start-up phase and was from the very beginning stable and predictable in its reactions. The generator capacity ranged roughly between 7 and 14 kW, whereas the evaporator has been regularly producing cooling between some 3 and a peak power of 7.8 kW. The COP, defined as the ratio between evaporator and generator power, showed a maximum of 0.53. The results of a test rig at nominal conditions are shown in Fig. 4 and 5. The unit was run with 96 °C hot water and produced chilled water at 12.5 °C. The cooling air left the machine at around 36 °C. Under these conditions more than 7 kW of chilled water can be already steadily produced with a COP above 0.5.

Fluctuations due to oscillation in air temperature can be clearly noticed in the evolvement of thermal capacities and relative COP. Nevertheless, these fluctuations did not show any trend and did not compromise the reliability of the operations.

Fig. 4. Experimental working conditions for hot water, cooling air and chilled water.

The experimental tests already gave an impression of the capacity of the prototype. So, it can be said that most components seem to have the right size or even to be oversized, as e. g. the air heat exchangers and the refiner. The air fan — run in part load — provided the necessary cooling for condenser and absorber, even if the laboratory room is very narrow and a certain air backflow could not be excluded

However the prototype has shown some limitations, which are now fully identified and can be easily corrected in the next prototype, to be built ant tested until the end of 2008. It is the expectation of the authors that the nominal values will be easily reached then, namely the design

COP of 0.6

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Fig. 5. Generator and chilled water capacity and reached COP.