## VALUE OF m FOR SIMPLE SOLAR COLLECTOR MODEL

Consider steady state operation of a solar chimney power plant. The solar collector consists of a transparent deck over a floor that receives solar energy. The collector floor

transfers energy to the air flowing over it at the same rate at which it receives energy from the sun, namely at a G Acoll, where a is the effective absorption coefficient of the collector. The air in the collector loses heat through the collector deck at the same rate that the collector deck loses heat to the environment, namely at pATAcoll, where p is an adjusted heat transfer coefficient that allows for radiation and convection losses and the fact that the temperature difference between deck and environment increases from 0 at the outer edge of the collector to AT at the chimney entrance. The real situation is more complex, but this simple model employed by Schlaich (1995) may be used to derive an approximate expression for the collector temperature rise and the exponent m of the analysis above. Find the air temperature rise by equating energy entering and leaving the collector:

Qfloor — Qdeck — Qcfe

a G Acoll PAT Acoll V pcollcpAT

 (11)

AT —___ a G Acoll____

V pcollcp + P Acoll

If the ambient temperature at ground height is T0, the collector exit (chimney inlet) density is pcoll and assuming parallel temperature profiles inside and outside the chimney a chimney of height Hc will generate a hydrostatic pressure potential, pp:

AT

pp — pcoll g Hc

1 0

pcoll g Hc____ a G Acoll (12)

T0 V pcollcp + в Acoll

 This equation is of the form: pp —
 C1
 C2 V + C3
 (13)

It can be shown that for a relationship of the form p = A Vm, m depends only on the local value of the function and the local value of its gradient, and is given by:

 m —-

dpp pp

“dV V (14)

For a function of another form, for example Eq. (13), an equivalent m can be calculated at any point, since m depends only on the coordinates of the point and its local gradient:

— dpp V

dV pp

‘ C1C2

(2 V + C3 )2

V(c2 V + C3)

C2V

(2 V + C3;

Back substitute for C2 and C3 and multiply the denominator and numerator by AT:

V p collc p AT

m = _

(V p coiicp AT + p Aeon AT) = _ V pcollCpAT a G A coll

 Qfloor Q
 deck
 Q
 ncfe
 floor
 (16)

= _ a G A coll _ в A coll AT a G Acoll

Here ncfe is the net rate at which heat is absorbed by the air between the inlet and exit of the collector, expressed as fraction of the rate of heat transfer from the floor to the air. We shall call it the collector floor-to-exit efficiency since it is a measure of how efficiently the collector transfers heat from its floor to the air leaving the collector. Schlaich (1995) writes the standard collector efficiency for his collector model as:

PAT (17)

ncoll =«_ g

The collector transfer efficiency can be written similarly by dividing out aGAcoll in Eq. (16):

 ncfe 1

PAT

aG

= n coll (18)

a

It is remarkable that in the case of the simplified solar chimney collector model, the exponent m turns out to be simply the negative of the collector floor-to-exit efficiency. The immediate implications are the following:

• m must have a value between 0 and -1

• for n = 2, the optimum pt/pp is between 2/3 and 1

• the optimal pt/pp ratio is 2/3 only if the collector efficiency equals zero.

## System Optimization

Two models have been designed: one system includes evacuated tube solar collector and the other includes double-glazed flat plate solar collector. Each model is simulated with different insulation thickness and size of overhang and wing walls. The simulation has been performed to select the optimum collector area and slope, generator temperature so that the maximum possible solar heat is gained, and to select the suitable insulation thickness and overhang and wing walls so that minimum possible cooling load is obtained. As seen from figure 2, Nicosia needs air conditioning for 9 months from March to November. Figure 3 shows cooling load of the building in Assab and from the figure it can be seen that air conditioning is needed all year round.

 Figure 3. Cooling load (kJ) (Assab)
 80 82 84 86 88 90 92 94 hot water inlet temperature (C) Figure 4. Effect of hot water temperature on solar fraction.

For Assab the maximum solar fraction is obtained at the optimum generator hot water inlet temperature of 86°C for the system with evacuated tube solar collector. The corresponding figure for double-glazed flat plate solar collector is 85°C [10]. The cooling water temperature is assumed low enough so that the system works at low generator temperature. Figure 4 shows the generator hot water temperature for Nicosia and the result shows that the optimum temperatures are 86°C and 83°C respectively. The solar fraction is defined as the percentage of the solar energy supplied to the absorption chiller over the total energy supplied (from solar energy and auxiliary heater) to the chiller. When the hot water temperature increases the additional energy is delivered from the heater, which reduces the solar fraction.

Solar heat gain from the collector varies with slope of the collector and latitude. It is proportional to the angle between the collector surface and the incident ray from the sun. For Assab the maximum is obtained at a collector slope of 13° for both types of collectors

[10]

 Figure 5. Effect of collector slope on solar fraction
 -SLFR EVC — SLFR DGC — SYEFF EVC SYEFFDGC

. For Nicosia a maximum heat is gained when the evacuated tube collector slope is 25° and the double-glazed flat plate collector slope is 24° (Figure 5).

The effect of both collector area and storage tank volume on solar fraction (SLFR) and system efficiency (SYEFF) for evacuated tube and double-glazed collectors are shown in
figures 6 and 7 respectively. System efficiency is defined as the percentage of the incident solar energy converted into cooling effect. For the case of Assab it is given in [10]. As seen from the figures the solar fraction always increases with solar collector size and storage tank volume. In the case of collector area the increase is higher at the smaller collector area, but decreases its increment as the area increases. For the case of storage tank the variation is high at high volume.

 Figure 6. Effect of evacuated tube solar collector (ESC) area on solar fraction and system efficiency (Nicosia).
 collector area Figure 7. Effect of double-glazed flat plate solar collector (DGC) area on solar fraction and system efficiency (Nicosia).
 -1 SLFR -1 SYEFF -2SLFR 2 SYEFF -3SLFR 3SYEFF 4 SLFR 4 SYEFF 5 SLFR 5 SYEFF
 -1 SLFR -1 SYEFF -1.5 SLFR -1.5 SYEFF -2SLFR 2 SYEFF 3 SLFR 3 SYEFF

Figures 6-10 show the effect of collector area and storage tank volume on solar fraction and system efficiency for the insulated building.

The effect of solar collector area on solar fraction and system efficiency for the 2.0m3 hot water storage tank is shown for Nicosia (Figure 8) and for Assab (Figure 9). From the figure it can be seen that the solar fraction of Assab is smaller than that of Nicosia, but the system efficiency is higher. In all cases solar fraction increases with collector area. The maximum system efficiency of the evacuated tube solar collector for both locations is obtained when the collector area is 20m2. For Nicosia the maximum system efficiency is obtained when the double-glazed collector is 20 m2 while for Assab when it is 30 m2.

The effect of thermal storage tank to the solar fraction and system efficiency is very small comparing to that of the solar collector (Figure 10).

20 10 0

3

storage tank volume (m3)

Figure 10. Effect of storage tank on solar fraction and system efficiency (Nicosia).

 Figure 11. Effect of insulation thickness on cooling load.

Figure 11 shows the effect of insulation thickness on cooling load for both locations. The highest cooling load is obtained when the building is not insulated and the lowest is obtained when the insulation thickness is 0.20 m. The maximum cooling load for Assab is 265 MJ while for Nicosia it is 78.6MJ. The reduction of cooling load in the first 0.05m insulation thickness is 34% for Assab and 25% for Nicosia. For the second 0.05m insulation thickness the reduction drops to 6.2% for Assab and 2% for Nicosia. As the thickness increases the effect of reduction decreases and is insignificant for insulation thickness greater than 0.2m.

 overhang and wing wall (m) Figure 12. Effect of overhang and wing wall on cooling load.

Overhang and wing wall reduce cooling load since the direct solar radiation is prevented from entering the building. As the size increases the reduction also increases (Figure 12). But the effect is very small when compared with the effect of insulation. In the first 0.5 m overhang and wing wall addition the reduction is 4% for Assab and 5% for Nicosia. When another 0.5 m is added the reduction is 2.8% and 2.6% respectively.

 Figure 13. Effect of insulation thickness on solar fraction and system efficiency load (Nicosia).

Figure 13 shows that solar fraction increases with insulation. This is due to the fact that energy demand decreases with insulation. But the system efficiency more or less remains the same.

Conclusion

A TRNSYS simulation model of solar-assisted air conditioning system has been developed to simulate long-term performance of buildings in Assab, Eritrea and Nicosia, Cyprus. Two collectors types: evacuated tube collector and double­glazed flat plate collectors are used for comparison purposes. The model was used to perform a parametric study of the system to investigate the effect of collector area, storage tank size and generator temperature on solar fraction and over all system efficiency. The model was also used to analyze the possibility of reducing cooling by means of addition of insulation, overhang and wing walls to the buildings.

In both locations the highest system efficiency was obtained from the system, which used an evacuated solar collector. For Assab it is 1.85 times greater than that of the system, which used double-glazed collector and for Nicosia the figure is 1.72.

The lowest cooling load is obtained for both locations when the insulation thickness is 0.20 m and the size of the overhung is 1.5 m. Further reduction could be obtained without economic advantage. Reduction of cooling load is not enough when optimum condition is needed. In addition to the cooling load reduction analysis of the overall cost of the system is important. But for economic comparison the cost of equipment, building materials and fuel oil used in the systems are not the same in all locations.

## Study on the Passive Cooling Methods by the Evaporation and Solar Reflection on the Rooftop in a Temperate Climate Region

Seonghwan YOON, Yasunobu ASHIE and Toshio ABE

Building Research Institute, 1 Tachihara, Tsukuba, Ibaragi, 305-0802 JAPAN

Introduction: There have been a number of studies on the passive cooling of the outer surface of buildings for the purpose of conserving energy without resorting to fossil fuels, and simultaneously, improving the indoor thermal comfort. In addition to this, the heat island phenomenon has been regarded as a problem (Ichinose et al., 1999). Thus, it is necessary to strive to reduce the thermal load of buildings on the urban environment. It is particularly important to examine thermal properties of finishing materials on the rooftop surface where the incoming direct solar radiation is great. For this reason, the purpose of this study is to examine the cooling effects by the evaporation and the reflection of solar radiation on the rooftop surface by conducting outdoor experiment and numerical calculation. In the outdoor experiment, various test pieces were installed on an existing rooftop to examine the change in the surface temperature on a clear, summer day. Next, numerical calculation was carried out assuming different states of insulation to clarify the influence of rooftop surface cooling on the thermal load both indoors and outdoors in summertime.

## EXPERIMENTS IN THE DAYLIGHT CHAMBER

Preliminary experiments are done in the daylight chamber of the university of Delft in order to make an estimation of the size of the illuminance in underground spaces. The artificial sky simulates a CIE-overcast sky. The illuminance Ehorf in the daylight chamber is 6800 lux. Experiments are done with scale models of scale 1:20 because the height below ground level in the daylight chamber is only 0,8 m in reality (fig. 1). The dimensions of the standard underground space was taken to be 4 x 4 x 4 m3 with a tube of 1 x 1 x 3 m3 on scale.

Figure 2 shows this standard space and the different variants of which the level of illuminance has been tested. The dimensions of the cylindric model (H) and the half sphere

(I) are chosen in a way that the surface of the walls is the same as in the standard model.

All scale models of the underground space are made of white foam, so that internal reflections are taken into account. The inside material of the tube was a material of high specularity in order to send as much daylight as possible to the underground space.

The effects on the illuminance in relation to the dimensions and the form of the space and the tube are investigated. The illuminance is measured at different points on the floor and the wall. On the floor the distance between the grid points was 0,5 m on scale (fig.3). On the wall there were two rows of measurement points: one row on 1,5 m and one on 3 m height, with distance of 0,6 m between the points (fig. 4).

Table 1 and 2 show the results of the measurements. On the floor almost all the models show illuminances between 150 lux and 450 lux and on the wall illuminances between 80 lux and 180 lux. Raising the height of the space or the tube reduced the illuminance in the space, but raising the height of the tube has the most impact. From the symmetries of the spaces it can be concluded that the measurement uncertainty is below 10 lux.

The measured illuminances are too low for offices and livings, but these buildings are yet forbidden below ground level by the building norms in the Netherlands. But exposition spaces can be built underground. Illuminances of 150 lux on walls are very well suited to expose paintings, and illuminances of 300 lux — 450 lux on the floor are suited to show 3D art objects like statues [2,3].

fig. 1. Measurements below ground level in the daylight chamber.

 A. Standard
 Ci^£aceJieic]hU5m
 E. Space length 6m
 F. Space length 8m (1 tube)
 D. Space height 6m
 Bi^£acejieic]ht_3m

SHAPE * MERGEFORMAT

fig.2. Different shapes of the scale models used in the daylight chamber.

0,5 m

fig.3. The gridpoints on the floor of the underground space, distance 0,5 m (scale 1:20).

3 m

1,5 m

0,6 m

fig.4. The gridpoints on the wall of the underground space, distance 0,6 m (scale 1:20).

table 1. Illuminances on the floor of the underground space for the different variants.

 lux 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 A 260 250 225 190 252 243 219 189 229 223 200 174 197 190 172 152 B 370 358 305 235 352 337 295 228 300 291 255 204 243 242 215 182 C 200 195 183 162 197 193 181 162 190 184 173 155 179 174 161 144 D 161 157 150 141 160 156 150 141 154 152 146 139 148 144 140 128 E 222 217 195 158 219 213 193 157 203 198 178 149 180 177 159 135 F 204 198 178 141 201 196 176 140 188 183 166 133 169 166 149 123 G 221 219 243 265 219 218 240 261 210 208 232 248 193 192 205 220 H 249 241 218 188 242 233 214 178 227 219 197 169 201 197 179 162 I 436 396 290 226 426 404 309 245 350 328 260 218 279 267 230 192 J 381 370 342 293 383 369 342 296 366 357 326 293 343 335 306 279 K 196 183 155 130 181 170 149 123 151 146 127 108 134 131 117 95 L 251 245 221 189 249 240 219 186 229 211 223 176 202 198 180 157 M 259 250 226 189 255 246 223 189 234 227 208 178 212 207 187 159 N 260 251 230 195 256 247 225 194 235 228 208 183 210 206 196 168 O 265 258 229 203 259 253 229 199 239 228 212 189 213 205 197 165

table 2. Illuminances on the wall of the underground space for the different variants.

 lux 1 2 3 4 5 6 A 169 167 154 132 128 116 B 185 173 151 C 150 139 123 130 129 122 D 123 123 116 107 108 104 E 137 131 116 98 97 85 F 118 112 95 86 84 75 G 160 160 169 129 132 141 H 161 123 I 163 J 196 181 162 141 135 128 K 147 141 132 132 130 121 L 165 160 144 128 125 110 M 167 165 155 130 126 116 N 168 167 154 132 125 114 O 172 170 151 143 136 128

## The Skylight-Dimensioning Method

 Figure 8: Notation system of the Skylight-Dimensioning Method

This existing method can be applied to determine the minimal necessary area of transparent surfaces of those skylights that are effected by the hole hemisphere (that "can see” the whole hemisphere).

Ff represents the necessary transparent area of the skylight, it can be calculated with the following equation:

 {2}

eaverage a • b m2

100 Tg — Ts -Td ■ ka ■ ^

Nomination of the initial coefficients and parameters:

 e average a, b,h r Pc, Pw, Pf
 p, q,m Pa
 a

: the average illumination coefficient [%],

: the main dimensions — width, depth, height — of the interior space [m],

: the distance between the reference plane and the floor level [m],

: the average reflection coefficients of the surfaces — ceiling, walls, floor — of the interior space [%],

: the main dimensions — width, depth, height — of the light-well [m],

: the average reflection coefficient of the internal surface of the light-well [%],

: skylight transmittance modification coefficients: rg — transmittance, rs — obstruction and rd — dirt coefficients [-].

: the efficiency of the light-well [-],

: the efficiency of the internal space [-].

Calculation and determination of values of the initial coefficients and parameters:

eaverage : the required illumination coefficient, which can be determined using

standards and tables or preliminary recommendations, or it also can be calculated using the En interior (function) specific nominal illumination:

eaverage = EE^ 100%]. {3}

Ekm

Tg, rs, rd : the Tg — transmittance, rs — obstruction and rd — dirt modification values of the

transparent surfaces of the skylight can be determined using existing standards and tables according to the type of the applied — glazing, — structure or depending on the actual environment, for instance:

 Tg As *d simple, clear glazing 0.9 structural glazing 1 countryside, regular cleaning 0.8 double, clear glazing 0.8 metal structure 0.9 countryside, accidental cln. 0.55 simple wire-gazing 0.77 timber structure 0.8 residential, regular cleaning 0.7 double wire-gazing 0.63 RC structure 0.8 residential, accidental clean. 0.4 plastics 0.8 industrial, regular cleaning 0.55 industrial, accidental cleaning 0.25 clean state at handover 1
 Table 2: Approximate Transmittance, Obstruction and Dirt modification values

 a

: the efficiency of light-well, that can be determined using the following equation, where i is the light-well index and pa is the average reflection value of the internal surface of the light-well:

i = m-(p + q)

2 • p ■ q

The value of ka can be determined using the light-well efficiency table bellow:

 kr
 kr
 {5}
 a ■ b (a + b)-(h — r)

: the interior coefficient can be calculated using the following equation:

 П

: the efficiency of the interior can be determined with the help of the following table (the kr interior coefficient values and the actual pc — ceiling, pw — wall and pf — floor average reflection coefficients shall be also substituted to the table):

 pf 0,3 0,1 Pc 0,8 0,5 0,8 0,5 0,3 Pw 0,8 0,5 0,3 0,5 0,3 0,3 0,5 0,3 0,3 kr 0,6 0,63 0,39 0,33 0,39 0,33 0,34 0,37 0,33 0,32 0,8 0,76 0,53 0,45 0,51 0,45 0,45 0,50 0,45 0,44 1 0,88 0,62 0,54 0,60 0,54 0,53 0,58 0,53 0,50 1,25 0,95 0,71 0,63 0,68 0,62 0,62 0,66 0,60 0,60 1,5 1,02 0,78 0,70 0,76 0,69 0,68 0,72 0,68 0,66 2 1,1 0,89 0,81 0,85 0,78 0,77 0,80 0,77 0,74 2,5 1,14 0,96 0,88 0,91 0,85 0,83 0,85 0,82 0,80 3 1,17 1,01 0,94 0,95 0,89 0,87 0,88 0,86 0,84 4 1,21 1,07 1,01 1,00 0,95 0,92 0,93 0,90 0,89 5 1,23 1,12 1,06 1,03 0,98 0,95 0,96 0,93 0,92

## Comparison against coal technology

In a recent U. S. regional power plan discussion document (North West Council, 2002) the cost of a 400 MWe pulverised coal plant was found to be \$1468/kWe in the North West USA. This plant is used as a coal cost baseline for comparison costings against two CLFR/cavern scenarios, one with 54% capacity factor and one with 68%. In Table 1, the coal plant is given an 80% capacity factor, within the normal range for capacity factors in the USA. David and Herzog (2003), for example, use 75% in a study of carbon sequestration. The coal plant IRR was held to 14%, assumed as a reasonable payback for solar plants in NREL (2003), by adjusting the wholesale price for electricity. premium charged for peaking sales, because as the capacity factor is reduced, there is a greater opportunity to indulge in ‘peak lopping’, giving a higher return per kWhe. The IRR for such trading can only be determined using a complex grid pricing model not available to the authors.

 Annual Output MW(th) 14,038,462 3,620,800 4,526,000 Thermal to Electrical efficiency 39.0% 31.5% 31.5% Online Status 0.98 0.98 0.98 Total Annual Equivalent MWH Output 2,522,880 1,117,741 1,397,176 Annual Gross Plant Revenue US\$ 111,315,773 45,291,605 56,614,506

 Coal cost 0.71 MMBTU 42, 522 644 — — Reflector Array Cleaning — 3, 587, ,143 4, 483 929 Operations and Maintenance 15, 646 080 2, 022, ,019 3, ,159 ,404 Debt Payment 28, ,146, 027 20, 522, 963 24, 359 ,041 Annual Gross Costs of Service US\$ 58, ,168, 724 26, ,132, ,125 32, О о го 373

Annual Net Plant Revenue US\$ 53,147,049 19,159,481 24,612,133

 Net Present Revenue per MWh \$43.05 \$43.05 \$43.05 Net present cost per MWh -\$28.97 -\$24.43 -\$23.84 Net present profit per MWh \$14.07 \$18.61 \$19.21 CPI 2.5% 2.5% 2.5% Debt cost 7.2% 7.2% 7.2% Debt ratio 50% 50% 50% 25YR IRR 14.00% 13.87% 14.76%

 2nd Year Example Revenue Sheet400 MWe Coal CLFR/Cavern CLFR/Cavern Capacity Factor 0.80 0.54 0.68 Electricity Sale \$/MWH 45.23 45.23 45.23 Environmental Support \$/MWH 0.00 0.00 Total Revenue \$ per MWh 45.23 45.23 45.23 Collector Area m2 0 3,188,571 3,985,714 Array related cost \$/kWe 0 1435 1744 Storage Cost \$/kWe 0 68 92 Power block and BOP cost \$/kWe 1468 281 281 Total Cost \$ per kWe 1468 1784 2117
 Table 1. Costs and IRR of coal and CLFR systems

The solar plants were then evaluated on this selling price and it was found that their IRR is comparable to coal; slightly higher than coal for the 68% capacity factor plant and slightly lower for the 54% plant. The optimal capacity factor depends upon the

In Fig. 2, the capacity factor of new pulverized coal plant is now varied to produce a range of electricity wholesale prices which meet the desired IRR of 14%. This is compared to the 68% CF CLFR/cavern storage solar plant which is also held to an IRR of 14%. The graph shows that the coal fired plant is more costly up to about a CF of 82%, and even at a CF of 90% is only \$5 per MWhe less expensive than the 68% CF solar plant. This suggests that minimal measures such as low priced carbon trading

 Wholesale Price of Electricity for 14% IRR
 Coal Plant Capacity Factor Fig. 2. Cost of electricity in the second project year required to produce a 14% IRR in high CF Coal and Solar scenarios. The Coal CF is allowed to vary while the CLFR storage plant is held at a 68% CF, close to the higher range of solar CFs possible using daily storage in mid-latitudes such as NSW and California.

would be sufficient to provide solar competitiveness against the cheapest baselaod coal fired plant.

## EXPERIMENTAL STUDY OF A LIQUID DESICCANT. SYSTEM FOR SOLAR COOLING AND. DEHUMIDIFICATION

K. Gommed and G. Grossman

Faculty of Mechanical Engineering Technion, Israel Institute of Technology, Haifa 32000,

ISRAEL

Growing demand for air conditioning in recent years, even in European countries with no air conditioning tradition, has caused a significant increase in demand for primary energy resources. Solar cooling has been recognized as one of the environmentally-friendly techniques toward alleviating the problem. A promising method is through the use of a liquid desiccant system, where humidity is absorbed directly from the process air by direct contact with the desiccant. The desiccant is then regenerated, again in direct contact with an external air stream, by solar heat at relatively low temperatures. The liquid desiccant system has many potential advantages over other solar air conditioning systems and can provide a promising alternative to absorption or to solid desiccant systems.

Earlier work included theoretical simulations and preliminary experiments on the key components of the liquid desiccant system. The objective of the present study has been to construct a prototype system based on the knowledge gained, to monitor its performance, identify problems and carry out preliminary design optimization. A 12 kWt system was installed at the Energy Engineering Center at the Technion, in the Mediterranean city of Haifa. The system comprises a dehumidifier and a regenerator with their associated components operating together to dehumidify the fresh air supply to a group of offices on the top floor of the building. LiCl-water is employed as the working fluid. The system is coupled to a solar collector field and employs two methods of storage — hot water and desiccant solution in the regenerated state. The performance of the system was monitored for five months during the summer of 2003 under varying operating conditions. The paper describes the operation of the experimental system and presents the measured data and the calculated performance parameters.

Background

The growing demand for air conditioning, particularly in hot and humid climates such as in Mediterranean countries, has caused a significant increase in demand for energy resources. Electric utilities have their peak loads in hot summer days, and are often barely capable of meeting the demand, with brown-out situations. With suitable technology, solar cooling can help alleviate, if not eliminate the problem. It is a good application for solar energy due to the fact that the greatest demand for air conditioning occurs during times of highest insolation (Grossman and Johannsen, 1981; Grossman, 2002).

Conventional closed-cycle absorption chillers require heat source temperatures that are significantly higher than the temperatures of corresponding heat sinks. Thus, they have to be operated with high-grade heat extracted from natural gas, steam, concentrating solar collectors and the like. A promising alternative is the use of an open absorption system, where humidity is absorbed directly from the air to be treated by direct contact with the absorbent. The absorbent is then regenerated, again in direct contact with an external air stream, at relatively low temperatures of the heat source. The entire operation takes place at atmospheric pressure, thus eliminating vacuum vessels and the like.

Earlier work has been conducted on liquid desiccant systems for cooling and dehumidification, using solar energy for regeneration. In several cases, direct regeneration of the solution in the sun has been considered, using a special type of collector. Wood and co-workers at Arizona State University (Ameel et al., 1995; Nelson and Wood, 1989a, 1989b, 1989c) have constructed and tested a full-scale liquid desiccant system, employing LiCl as well as a mixture of LiCl and CaCl2 as liquid sorbents. Kessling (1997) studied a LiCl-water system operating at a large concentration difference between the strong and weak desiccant, to facilitate cold storage by means of a regenerated solution. Kababaev et al.(1976, 1977, 1981) report on the operation of a large scale air-conditioning system employing LiCl-water, where both direct regeneration in open collectors and cold storage in the form of regenerated solution have been attempted. Noteworthy are also the liquid desiccant system analyses of Collier (1979), Haim et al. (1992), and Gandihdasan and Al — Farayedhi (1995).

The objective of this project has been to design, build, test and evaluate a fully — instrumented liquid desiccant cooling system, to provide a demonstration and supply operating data under varying conditions. The system is capable of using as its source of power low-grade solar heat from low-cost flat plate collectors and has the potential to provide both cooling and dehumidification in variable ratios, as required by the load. The significance of this work lies in the potential to provide solar powered cooling, dehumidification and air conditioning for residential and commercial applications.

As part of the design phase of the liquid desiccant system, a complete system simulation was conducted, in order to predict trends and attempt a preliminary optimization. The lack of reliable data on heat and mass transfer coefficients in the absorption and desorption processes had been a serious impediment in earlier simulation studies to obtaining a good prediction of the system’s performance. Particularly critical are the performances of the dehumidifier (absorber) and regenerator (desorber), forming the two key components of the liquid desiccant system. Such data has now been obtained through the experimental work described by Gommed, Grossman and Ziegler, (2002). This made it possible to conduct an extensive parametric study of the overall system behavior(Gommed and Grossman, 2002). The computer simulation yielded the temperature and humidity of the air at the system outlets as well as heat duties of the various system components as functions of the specified conditions at inlets and other operating conditions.

## 60 55 50 45 40 35 30 25 20 15 10 О и & Н I — Q Month Figure 3. Average global radiation and ambient temperature of Bangkok (Meteonorm, 2001) RESULTS

The location of Bangkok is at latitude 13.44 degree above equator thus the cooling load as well as the ambient temperature are quite stable for the year round. The average global radiation (kWh/m2) and the average ambient temperature are showed in figure 3.

The load generally depends on the solar radiation as shown in figure 4. The average load capacity is 4 kW. The high cooling load can be observed during the dry season, from March to June.

 Generating temperature (°C) Figure 5. Efficiency of the ejector refrigeration subsystem

The performance of the ejector refrigeration subsystem strongly depends on the operating temperatures and the ejector design (Pridasawas, 2003). The COP of the refrigeration subsystem increases proportionally to the generation and evaporation temperature. A high generating temperature, however, requires a high temperature output from the solar collector, thus resulting in high heat losses and more expensive solar collectors. The performance of the ejector increases when the evaporator temperature increases, thus refrigeration subsystem should be provided at the highest evaporation temperature possible in the given application. In this case, the evaporation temperature is fixed at 10°C and the generating temperature is studied. The changing in cooling load does not significantly affect the COP, if the operating temperatures are kept constant. Generally, the generating temperature is high when the cooling load is high due to the high temperature heat gain from the solar collector. On the other hand, the solar collector efficiency is in inverse proportion to the temperature output of the solar collector. Figure 5 shows that the highest STR is obtained at the generating temperature between 90 and 110 °C. The average COP of the system of the simulation condition described above is about 0.3.

The simulation result for the whole year at 40 m2 solar collector area, 2 m3 storage tank volume and 400 kg/hr water flow rate is shown in figure 6. The highest performance is obtained from March to August. During the dry season (March to June), the ejector

subsystem operates at high cooling capacity and high performance due to the higher outlet temperature of the solar collector. The performance of the refrigeration subsystem does not only depend on the generating temperature, other operating conditions e. g. condensing temperature and evaporation temperature, the geometry of the ejector also influence the performance (Lundqvist, 1987). In this simulation other parameters are set constant at normal operating conditions, thus, their influence cannot be observed. Further information can be found in the literature of Lundqvist (1987), and Pridasawas (2003).

 0 800 Jan Feb Mar Apr May Jun July Aug Sep Oct Now Dec ЗДічМИоп Tim# ««TWO, iv Figure 6. A year round COPejc
 Figure 7. A year round solar collector efficiency
 Figure 8. A year round system thermal ratio

The annual energy usage by pumps and auxiliary heater decreases remarkably when the solar collector area increases from 20 m2 to 50 m2. The optimum solar collector area should be traded off with the solar collector installation cost. Figure 9 shows the annual energy (heat and electricity) usage by the pumps and auxiliary heater at different solar collector area for the storage tank 2 m3 and the water flow rate 450 kg/hr. Ninety five percent of the energy is consumed by the auxiliary heater, 3% by the circulation pump in the solar collector subsystem and 2% by the pump in the refrigeration subsystem. It shows clearly that the auxiliary heater is needed for the reliability of the system.

The increase in size of the storage tank and the water flow rate (in the solar collector subsystem) does not significantly affect the energy usage.

The increase of the water flow rate every 100 kg/hr decreases the supplementary energy usage about 2% while the increase of the storage tank volumes every 1 m3 decreases the energy usage only about 0.5%. The system operates during daytime while the high radiation is proportional to the cooling load, thus the size of the storage tank does not significantly affect the system. If the system also operates during nighttime, the size of the storage tank becomes an important parameter. The model and the degree of stratification of the storage tank should also be taken in consideration.

CONCLUSION

The system thermal ratio (STR) is influenced by both the COPejc and the solar collector efficiency. The COPejc increases in proportion to the generating temperature while the efficiency of the solar collector decreases in inverse proportion to the outlet temperature. The optimum generating temperature for achieving the highest STR is about 100°C. The highest STR that can be obtained is about 0.25 at the COP of 0.55. The optimum solar collector area is about 50 m2 for the average cooling load at 4 kW. The size of the well — mixed vertical storage tank does not significantly affect the annual electricity usage of the system, which operates during daytime. The water flow rate in the solar collector subsystem slightly influences the electricity usage, however, it should be high enough to provide the heat for the generator and to maintain the reliability of the system.

NOMENCLATURE

TOC o "1-5" h z Asc = solar collector area (m2)

c = velocity of refrigerant in the refrigeration subsystem. (m/s)

COPejC = coefficient of performance of the ejector refrigeration system (-)

FR(za)e= Bliss coefficient (-)

(FrUl) = Bliss heat loss coefficient (kW/m2 K)

h = enthalpy (kJ/kg)

hg, exp = enthalpy of the driving fluid from the generator after expanded

through the nozzle. (kJ/kg)

I = Incident solar radiation (kW/m2)

m = mass flow rate (kg/s)

Qe = cooling capacity at an evaporator (kW)

Qg = energy input to a generator (kW)

Qs = incident solar radiation over the solar collector (kW)

Qu = useful energy output from a solar collector (kW) 03

SER = system electrical efficiency STR = system thermal ratio

T = absolute temperature (Kelvin)

Wpump = necessary work required by pump (kW)

r/sc = solar collector efficiency

• = mass ratio

Subscripts

1- 9 = the points in the refrigeration subsystem as shown on figure 2.

c = condenser

e = evaporator

el = electricity

g = generator

ht = auxiliary heater

p = pump

## The Modern Art Museum

The Modern Art Museum of Fort Worth was designed by Japanese architect Tadao Ando in 2000-2001. This two-story museum opened its new building at the end of 2002. The museum’s permanent collection consists of more than 2,600 works, including paintings, sculpture, site-specific installations, drawings, prints, photographs, and video and digital imagery. The building is set on eleven acres, including a large reflecting pool and areas of outdoor sculptures. The building has 53,000 square feet of gallery space (Reference 1).

The Modern Art Museum introduces natural light into the galleries through different fenestration systems that include skylights, clerestories with interior and exterior louvers, and side light windows with and without interior screens (Figure 2).

## A Mediterranean house designed with the whole. building approach

Patricia Ferro

Via Portoferraio, 22 — 00182 Roma — criferro@tiscali. it

1. Introduction

Yearly energy consumption (heating, hot water and electricity) for Italian residential buildings averages from 150 in south and central Italy to 200 kWh/m2 in north Italy1. This paper describes a house designed with the whole building approach whose estimated energy consumption (heating, hot water and electricity) is estimated less than 70 kWh/m2 year.

The house will be built in central Italy (Casaprota, Rieti Province, 450 meters a. s.l., 42° 24’ N — 12° 52’ E ), a typical Italian Mediterranean hilly landscape.

The paper focuses on some of the design results, in particular those regarding daylighting, space heating and electricity demand, whose production will be 100% from renewable energy (solar and biomass).

To achieve these results, special attention has been given during the design to the siting, the house lay out, the local renewable energy resources available, the local architecture and the construction traditions.

Lazio Regional Government awarded the design of this house within the "Zero Emission residential buildings" competition. The award includes a funding support for the house construction.