Category Archives: Sonar-Collecttors

Process simulation software APROS

The time dependent processes are studied on a numerical basis, using an unsteady process simulation software. Such software tools solve the one-dimensional un­steady conservation equations for mass, energy and momentum [2]:

• TOC o "1-5" h z mass balance: dA^+dApv = Q (1)

dt d z

• momentum balance: dApv + dApv +dAp = Sm (2)

dt d z d z

• energy balance: dAph + dApvh = se (3)

dt d z e

On the basis of a detailed software evaluation [2,3,4], reviewing more than 15 simula­tion tools, the unsteady simulation software APROS from the finish supplier VTT was selected. The Advanced PROcess Simulation environment APROS is one of the very few software tools, that provides a 6-equation model for calculating the conservation equations for water and steam separately. This is of interest for the calculation of so called zero flow conditions appearing during start up and shut down when the flow velocities decrease to zero. Most of the other tools provide 5-equation models only, neglecting the momentum differences between the liquid and gaseous phases. Fur­ther advantages are the various interfaces providing several alternatives for connect­ing external software to the simulation programme like the IAPWS-IF97 for the thermodynamic properties of water and steam [6]. A long list of reference users working with APROS especially in Europe is another advantage. The number of available power plant components is very large and nearly every plant type from fluidized bed boilers to even coal gasification combined cycles including automation and electrical systems can be modelled.

RESULTS AND DISCUSSION

The models developed were initially verified by simple species and energy balances. The results were accurate within 1%.

(25)

(26)

For the constant thickness model, the results were compared to results available in the literature [10] for simulations using CaCl2-water as the liquid-desiccant solution, flowing in a co-current configuration. The authors defined average Nusselt and Sherwood numbers as follows:

It is not clear if the Cai* adopted by the authors was calculated at the wall temperature or the desiccant inlet temperature. The former was used in the present work. Table 1 presents the variables and properties used that were common to all simulations.

Table 1: properties and variables used in simulations.

CaCl2

Air

k

W/m2

0.525

0.02635

CP

J/kg K

2330

1.028 x 103

r

kg/m3

1394

1.172

kg/m s

1.19 x 10-2

1.83 x 10-5

D

m2/s

2.5 x 10"5

4.2 x 10-10

m

kg/s

7.0 x 10"3

1.23795 x 10-2

Ti

oC

25

35

Ci

kg/kg

0.6

0.02

hfa

J/kg K

2.448 x 106

Tw

oC

1 0

H

8 cx 2

Present work

Rahamah et al [10]

% Difference

m

m

Nu

Sh

Nu

Sh

Nu

Sh

4.0 x 10-1

2.6 x 10-3

2.42

2.08

2.32

1.92

4.1

7.7

4.0 x 10-1

3.0 x 10-3

2.79

2.42

2.68

2.25

3.9

7.0

4.0 x 10-1

3.5 x 10-3

3.20

2.80

3.08

2.67

3.8

4.6

4.0 x 10-1

4.0 x 10-3

3.56

3.15

3.44

2.95

3.4

6.3

4.0 x 10-1

3.3 x 10-3

3.04

2.65

2.93

2.50

3.6

5.7

5.0 x 10-1

3.3 x 10-3

2.56

2.20

2.45

2.07

4.3

5.9

6.0 x 10-1

3.3 x 10-3

2.18

1.88

2.08

1.75

4.6

6.9

7.0 x 10-1

3.3 x 10-3

1.98

1.63

1.84

1.51

7.1

7.4

8.0 x 10-1

3.3 x 10-3

1.73

1.40

1.65

1.32

4.6

5.7

Table 2: comparison between the results for Nu and Sh obtained in the present work and

Rahamah et al[10].

The results agree well, in special if one considers that the values used for the properties were not presented by Rahamah et al. This is particularly important for the air mass flow rate. The authors used a Reynolds number for the air flow of 1350, but did not inform the value used for air viscosity. Since the air flow rate has a significant impact on Nu and Sh, it would be difficult to obtain closer results, even if using the same numerical method and computer algorithms. For example, in the case of the 0.70 high channel, the one which presented the largest difference between the works, a 10% reduction in air mass flow rate causes a 9.6% reduction in the Nusselt number and a 9.8% reduction in the Sherwood number.

To compare the three models, another set of simulations was performed. All conditions were kept constant, with the exception of the desiccant mass flow. A co-current flow configuration was used, at this time with LiCl as the desiccant material. Table 3 presents the properties and variables used for the simulations.

Table 3: properties and

variables used in simulations for comparison between the three models.

LiCl

Air

k

(W/m2K)

0.558

0.0275

cp

(J/kg K)

3140

990

r

(kg/m3)

1394

1.11

P

(kg/m s)

1.86 x 10-3

1.9 x 10-5

D

(m2/s)

2.65 x 10-5

1.2 x 10-9

m

(kg/s)

0.01264

Ti

(oC)

25

30

Ci

(kg/kg)

0.6

0.015

hfg

(J/kg K)

2.430 x 106

Tw

(oC)

25

H

(m)

0.5

p cX 2

(m)

5.0 x 10-3

Table 4 and Fig. 5 present the results obtained with the simulations. It can be seen that the outlet air temperatures for the three models are relatively close, but the results for outlet water mass fraction in air differ significantly when the desiccant mass flow is reduced. The constant and variable thickness models give results that disagree by only 1.2%, when the air mass flow rate equals the desiccant mass flow rate.

Fig. 5: variation of outlet water mass fraction in air with desiccant solution mass flow. (conditions in Table 3).

For higher desiccant flow rates the variable and constant thickness model converge to the same results. However, for a desiccant flow rate 400 times lower than the air flow rate, the difference between the two models is 19%.In terms of the E Ca, the constant thickness model predicts a value 4 times smaller than the variable thickness model.

The variable thickness model was further tested against experimental data available in the literature. Keeling et al [11] made measurements with one single channel using LiCl as the desiccant and a counter-flow configuration.

Variable Thickness Model

Constant Thickness Model

Simplified Model

md

Tao ( C)

Cao

Tao

Cao

Tao

Cao

(kg/s)

(x10-3kgw/kgair)

(oC)

(x10-3kgw/kgair)

(oC)

(x10-3kgw/kgair)

ma

25.87

5.83

25.86

5.90

26.13

6.24

ma/2.5

25.86

5.93

25.86

6.13

26.13

6.37

ma/5

25.86

6.07

25.86

6.46

26.13

6.57

ma/10

25.86

6.32

25.85

7.08

26.13

6.96

ma/20

25.86

6.76

25.85

8.18

26.13

7.66

ma/30

25.86

7.15

25.85

9.08

26.13

8.25

ma/40

25.86

7.50

25.85

9.80

26.13

8.75

ma/50

25.86

7.81

25.85

10.4

26.13

9.18

ma/60

25.86

8.09

25.85

10.9

26.13

9.56

ma/70

25.86

8.34

25.85

11.2

26.13

9.88

ma/80

25.86

8.57

25.85

11.6

26.13

10.2

ma/90

25.86

8.78

25.85

11.8

26.13

10.4

ma/100

25.86

8.98

25.85

12.0

26.13

10.7

ma/150

25.86

9.73

25.85

12.8

26.13

11.5

ma/200

25.86

10.3

25.86

13.3

26.13

12.1

ma/250

25.87

10.7

25.86

13.6

26.13

12.5

ma/300

25.87

11.0

25.86

13.8

26.13

12.8

ma/350

25.87

11.2

25.86

14.0

26.13

13.0

ma/400

25.87

11.4

25.86

14.1

26.13

13.2

Table 4: comparison between the ^results obtained with the three different models.

The model was adapted for the water cooling channel employed in the experiment, that used a cross-flow configuration. In this configuration, for every 50 mm along the channel, there was 10 mm with no water flow. Therefore, the boundary condition for the regions with no water flow was changed to adiabatic, instead of isothermal. The regions with water flow were considered isothermal at a temperature equals to the average between the measured inlet and outlet water cooling temperatures. In the experiments, the water flow was kept high enough that the maximum increase in the water cooling temperature was

0. 4 oC. The channel was 0.46 m high and 5.5 mm deep, and the desiccant solution had an inlet water mass fraction of 59.8% for all experiments. The results are presented in Table

5. The outlet air temperature results agree very well, with less than 1.5% difference between the model and the experimental results. For the outlet water mass fraction in air the results agree very well for the lower desiccant mass flow rates. The difference increases for higher desiccant flow rates, reaching 13% for the case with the highest flow rate. It is not clear, however, what could cause such discrepancies. The desiccant mass flow range is still in the laminar smooth region, therefore, it is unlike that a wavy flow could have caused the difference. The values of uncertainty for the experiments published by the authors are bellow the difference obtained. Therefore, it remains to be investigated the causes for the discrepancies for the higher desiccant flow rates. For lower desiccant flow rates, the results agree very well. Although the use of the constant thickness model was not tested against the data from Keeling et al, the results would not likely be better, since the constant thickness model predicts a lower level of dehumidification than the variable thickness model.

Table 5: comparison between the results obtained with the variable thickness model and Keeling et ^ al[11]. ___________________________________________

ma

md

Tai

Tw

Cai

Tao

Cao

x 10-3 kg/s

x 10-3 kg/s

oC

oC

x 10-3 kgw/kgair

oC

x 10-3 kgw/kgair

1

12.64

0.621

24.5

24.3

14.2

Exp.

25.4

5.3

Model

25.1

6.1

2

12.64

0.327

24.8

24.3

14.3

Exp.

25.6

5.7

Model

25.3

6.4

3

12.64

0.187

24.6

24.3

14.3

Exp.

25.5

6.4

Model

25.3

6.8

4

12.64

0.115

24.4

24.3

14.3

Exp.

25.4

7.1

Model

25.3

7.4

5

12.64

0.085

24.4

24.3

14.3

Exp.

25.4

7.7

Model

25.3

7.8

6

12.64

0.071

24.1

24.3

14.3

Exp.

25.3

7.9

Model

25.3

8.1

7

12.64

0.058

23.9

24.2

14.2

Exp.

25.3

8.4

Model

25.2

8.4

CONCLUSIONS

Numerical models were developed for simultaneous heat and mass transfer in parallel — plate dehumidifiers. The model with variable film thickness was compared to experimental results available in the literature. The results for temperature agree well for all cases tested. The results for water mass fraction agree within 6% for desiccant flow rates more than 34 times lower than the air mass flow, considering desiccant flow on both sides of channel, i. e., 2 x md. For higher desiccant flow rates the discrepancies between the experimental and numerical results increase. Further investigation should be conducted to better explain the reasons for such discrepancies.

The constant thickness and simplified model under-predict the dehumidification, in special for low desiccant flow rates. The numerical model can be adapted for non-isothermal conditions, with the introduction of the cooling water flow equations.

NOMENCLATURE

C water mass fraction in solution (kg/kg)

Cp specific heat (J/kg K)

D diffusivity (m2/s)

dh hydraulic diameter (m)

g gravity acceleration (m/s2)

hh heat transfer coefficient (W/m2K)

hm mass transfer coefficient (m/s)

hfg latent heat of vaporization (J/kg)

ka heat conduction coefficient (W/mK)

m mass flow rate (kg/s)

Nu Nusselt number

p pressure (Pa)

pw water vapour pressure (Pa)

T temperature (oC)

u velocity (m/s)

x coordinate in the direction across the channel (m)

y coordinate in the direction along the channel (m)

Greek characters

a thermal diffusivity (m2/s)

g kinematic viscosity (m2/s)

dd film thickness (m)

dc channel half depth (m)

p. dynamic viscosity (kg/ms)

Subscripts

a

d

o

w

air-water

water-salt solution inlet

outlet

wall

SOLAR AUTONOMOUS DESICCANT COOLING TECHNOLOGY

Within the last years a lot of research in the field of solar supported desiccant and evaporative cooling (DEC) systems was done and many system variations were developed /1-10/. The solar input is typically "collected” by solar water collectors, the energy is then stored or directly used in the process for heating up the return air flow for regenerating the desiccant wheel.

The use of solar air collectors as energy source was analysed within simulation studies. The idea of keeping the system as simple as possible was supported by the promising results of the simulations. Thus the pilot plant in Freiburg uses directly solar heated air as exclusive energy source for the cooling case. Figure 1 shows the flow pattern and the different process steps in the temperature-humidity-diagram.

The general aim is to cool and to dry the ambient air. Therefore ambient air is
dehumidified by adsorbing the water vapour to a silicagel bed or some other hygroscopic
material between point 1 and 2. Due to the heat of adsorption the air is warmed up. For

removing the heat of adsorption the heat recovery between point 2 and 3 is needed. The fresh air stream coming from the dehumidifer wheel is the warm side of the heat recovery wheel, the return air from the building, which is cooled down by humidification (point 6 to 7) after leaving the room, is the cold side. After cooling down by heat recovery the air is humidified again (point 4 to 5) until reaching the set point of humidity and temperature.

The regeneration of the dehumidification wheel is done by air heated up in the solar air collectors (point 1 to 9). For regeneration air with high temperature and low relative humidity is needed. Therefore the regeneration temperature should reach 45 to 90°C /11/ depending on the cooling load.

Beside dehumidification and cooling, the DEC-plant is also used for humidification and/or heating of supply air or for pure ventilation of the rooms only. The operation modes internally numbered from one to five are:

• Modus 1 — Heating: The desiccant wheel is used as heat recovery wheel. Both wheels are used to warm up the inlet air by cooling down the outlet air. For meeting the room conditions the air is additionally heated by the backup heater and/ or solar gains.

• Modus 2 — Heat Recovery: The desiccant wheel is used as heat recovery wheel. Both wheels are used to warm up the inlet air by cooling down the outlet air. No additional heating is needed.

• Modus 3 — Free Ventilation: For realising the demanded air change rate the ventilators are delivering the inlet and outlet air. No other components of the plant are active.

• Modus 4 — Adiabatic Cooling: The inlet air is cooled down within the heat recovery wheel by humidified outlet air and/or by humidifying the inlet air.

• Modus 5 — Desiccant and Evaporative Cooling: The air is treated as described above.

Theory. Definition of good contrast

As mentioned in the introduction, large contrasts within the field of view are very unpleasant to the eye. This sensation is known as discomfort glare and has been examined from various perspectives since the early 1920s. Discomfort glare is usually expressed by a difference in luminance value within the field of view, and more precise by the ratio of the smallest and largest luminance value within the field of view. Although different researchers have formulated various equations with which to predict the discomfort glare, the agreement between the various equations and the subject response found by other researchers is still not so good that these equations are used in this paper. In this paper the simpler value of the maximum luminance ratio between the wall and the window is employed. This value, however, is also not so clearly defined as one would hope. In literature the following values are found, see table 1.

Source

Maximum

Specification

NEN-3087 [4]

1: 10 to 1:30

task and surrounding area

Moore [5]

1 : 20

task and surrounding area

Velds [1]

1 : 40

window and surrounding area

Table 1: luminance ratios found in the literature.

In this paper the value of 1:20 is taken as the maximum allowed value for the luminance ratio between the window and the surrounding wall. As this is not the maximum value found in literature, if a solution meets the demand of 1:20, it will always satisfy the higher demands found in the literature. As mentioned in the literature, the other parameters that influence the occurrence of discomfort glare from windows, besides the luminance contrast ratio is the lighting quality inside, the information given by the dagylight, the view through the window, the prior experience of the subjects and the subjects state of mind [1].

The integration of solar technologies

Solar collectors, Solarwall and PV modules are integrated in only one part of the building as reported in figure 13 and 14. This solution has been proposed since it could be applied in any similar condition without changing the local architectural typologies.

Solar systems are installed close to the technical rooms in order to reduce distances and thus energy losses between solar surfaces and other parts of the systems, like boilers, inverters, batteries, and so on.

Solar surfaces are also easy to clean and maintain, since no particular machineries or stairs are needed. Attention is also been paid in the design of the support system of the PV modules to prevent vandalism or thefts.

Fig. 14 — View of the house

5. Conclusion

The house design highlighted the potential of the whole building approach and of the use of energy efficient and solar technologies in improving the energy performance of a Mediterranean house. This has been recognized by the Regional Government of Lazio providing funding support for the construction of the house, which will be started in 2004. The house will monitored for a period of 10 years. Detailed cost analysis is still underway. This example is already stimulating interest among local contractor and technicians.

Solar thermal systems

Of the 20 projects 14 have active solar thermal systems. Ten systems are for domestic water heating (dhw), four are combined systems also contributing space heating. On average, the houses with solar dhw heating have 5 m[18] of collector coupled to a 500 liter storage tank.

The four houses with combined dhw and space heating systems have an average of 12.5 m2 of collector coupled with 2200 liter of storage. Two of those buildings have high efficiency vacuum collectors, which explains the smaller collector area per storage tank volume.

Conclusions

In planning a high performance house for the first time it can be useful to observe which design characteristics occur frequently in successful projects. Of particular importance are the key parameters of compactness, wall and window U-values, window to facade ratio and air tightness. The resulting very low remaining space heating requirement makes the energy needed for domestic water heating also an important end-use. Three quarters of the houses reported here have a solar energy dhw or combined dhw and space heating system to further reduce the consumption of non-renewable energy.

Maximum Fluid Power Condition in Solar Chimney. Power Plants — An Analytical Approach

TW von Backstrom and TP Fluri

Department of Mechanical Engineering, University of Stellenbosch
Private Bag X1, Matieland 7602, South Africa, Tel: +27 (0)21 808 4267
Fax: +27 (0)21 808 4958, E-mail: twvb@sun. ac. za

Abstract — Main features of a solar chimney power plant are a circular greenhouse type collector and a tall chimney at its centre. Air flowing radially inwards under the collector roof heats up and enters the chimney after passing through a turbo­generator. The objective of the study was to investigate analytically the validity and applicability of the assumption that, for maximum fluid power, the optimum ratio of turbine pressure drop to pressure potential (available system pressure difference) is 2/3. An initial power law model assumes that pressure potential is proportional to volume flow to the power m, where m is typically a negative number between 0 and -1, and that the system pressure drop is proportional to the power n, where typically n = 2. The analysis shows that the optimum turbine pressure drop as fraction of the pressure potential is (n-m)/(n+1), which is equal to 2/3 only when m = 0, implying a constant pressure potential, independent of flow rate. Consideration of a basic collector model proposed by Schlaich led to the conclusion that the value of m is equal to the negative of the collector floor-to-exit heat transfer efficiency. A more comprehensive optimization scheme, incorporating the basic collector model of Schlaich in the analysis, shows that the power law approach is sound and conservative. It is shown that the constant pressure potential assumption (m = 0) may lead to appreciable under estimation of the performance of a solar chimney power plant, when compared to the analyses presented in the paper. More important is that both these analyses predict that maximum fluid power is available at much lower flow rate and much higher turbine pressure drop than predicted by the constant pressure potential assumption. Thus, the constant pressure potential assumption may lead to overestimating the size of the flow passages in the plant, and designing a turbine with inadequate stall margin and excessive runaway speed margin. The derived equations may be useful in the initial estimation of plant performance, in plant performance analysis and in control algorithm design. The analyses may also serve to set up test cases for more comprehensive plant models.

INTRODUCTION

In order to design a flow system containing a turbine for maximum power production and to run it at maximum power, engineers need to find the optimal pressure drop across the turbine as a fraction of the total available system pressure difference. The design flow rate through the system determines the size and cost of the plant flow passages as well as the size, design and cost of the turbine. In the design phase some iterative algorithm may suffice to find the optimum, but a simple analytical method would be more convenient in a control algorithm. It could also serve to set up test cases for more comprehensive methods.

Many solar chimney investigators have made the assumption that the optimum ratio of pt/pp is 2/3, (Haaf et al., 1983; Lautenschlager et al., 1984; Mullett, 1987; Schlaich, 1995). In more detailed calculations Schlaich (1995) apparently used an optimum value of pt/pp = 0.82 as evident from the values of pt and pp reported in tables. Hedderwick (2001)
presented graphs showing values around 0.7. Von Backstrom and Gannon (2000) used the 2/3 assumption only for optimization at constant available pressure difference, but Gannon and Von Backstrom (2000) employed an optimization procedure under conditions of constant solar irradiation. Schlaich et al. (2003) reported a pt/pp value of about 0.80, while Bernardes et al (2003) reported a value of as high as 0.97. The wide variation in values warrants further investigation.

The question is the existence or not of a relevant optimum pt/pp in solar chimney power plants, and how to determine it. Even under conditions of constant solar irradiation the pressure potential of a solar chimney plant is not fixed but is a function of the air temperature rise in the collector, which varies with flow rate.

The first objective of this paper applies to any general process where the pressure potential is not constant and the system pressure drop is not necessarily proportional to the flow rate to the power 2.0. The objective is to derive simple, generally applicable equations for the determination of the volume flow for maximum fluid power (MFP) and the associated ratio of turbine total pressure drop to pressure potential. The second objective applies to solar chimney power plants. It is to derive equations for finding the optimum flow rate and pt/pp conditions as dependent on the relevant design and operating conditions of the plant, using a simple solar collector model.

Parametric study results

Supply air volume flow rate

200 m3/h

Reqen. air volume flow rate

250 m3/h

Pre-cooling air volume flow rate

250 m3/h

Sorbent Material

Zeolite

Sorbent thickness

0.25 mm

Heat exchanger plates material

Aluminium

Plates thickness

0.5 mm

Plates distance

3.5 mm

Table 1 — Structural heat exchanger characteristics

Return air temperature

26°C

Return air relative humidity

50%

Return air humidity ratio

10.5 g/kg

Ambient air temperature

35°C

Ambient air humidity ratio

20 g/kg

Heat recovery efficiency (reg.)

0.8

Adsorption Phase

1 t*

Regeneration phase

0.8 t

Pre-cooling phase

0.2 t

Parameters:

Regeneration temperature range

60-95°C

Cycle duration range

150-600 s

Table 2 — Main data used for the simulation campaign

In Figures 4 to 6 the results of a complete set of simulation calculations are represented as a function of the two considered parameters. The COP map in Figure 4 shows values significantly higher than in standard DEC cycles (typically 0.6 — 1). As expected the values are higher for low regeneration temperatures and long cycle duration, where the indirect adiabatic cooling effect is predominant on the desiccant cooling process. On the other hand for these values of the two parameters the system does not provide acceptable supply air conditions — at least in hot-humid climates — since the level of dehumidification is extremely small. Values of COP in the range of 1.2-1.4 are reachable for short cycles (i. e., below 300s) at any regeneration temperature. Figures 5 and 6 present the supply air temperature and humidity ratio maps for different values of cycle duration and regeneration temperature, respectively.

It can be seen that the level of dehumidification reachable with short cycles and regeneration temperatures higher than 80°C lies between 11 and 12.5 g/kg. These values show a dehumidification process significantly more efficient than in standard DEC systems in the same range of regeneration temperature. In the psychometric chart (Figure 7) are presented, as example, the two process paths.

Moreover the isotherm curves in Figure 5 have a different shape than the ones related to the humidity ratio for the same values of the studied parameters. From a deep analysis of the simulations data for each time-step, resulted that with high regeneration temperatures (i. e., 90-95°C) and short cycle durations (below 250s) a considerable amount of energy is still stored in the heat exchanger thermal mass at the beginning of the adsorption phase. The latter is due to the fact that the pre-cooling phase duration is fixed as percentage of the cycle
duration, and it is not optimised towards the process performance according to the regeneration temperature.

60 65 70 75 80 85 90 95

Regeneration temperature [°d

Figure 4 — COP maps as result of the parametric study versus regeneration temperature and cycle duration

60 65 70 75 80 85 90 95

Regeneration temperature [°d

Figure 5 — Supply air temperature maps as result of the parametric study versus regeneration temperature and cycle duration

60 65 70 75 80 85 90 95

Regeneration temperature [°d

Figure 6 — Supply air humidity ratio maps as result of the parametric study versus regeneration temperature and cycle duration

The data presented in Figures 4 to 6 do not take into account the possible further direct evaporative cooling (of the supply air).

If a humidifier would be operated after the heat exchanger supply air channels (see optional humidifier in Figure 3) the air could be further cooled and a full air-conditioning process could be carried out without using any conventional refrigeration machinery. Therefore another simulation run has been carried out including a direct humidifier in the supply air. For each time step the supply air conditions have been calculated implementing a direct evaporative cooling process simulation. The set supply air humidity ratio has been assumed 8.8 g/kg in order to cover the internal latent cooling loads. Figures 8 and 9 show the resulting supply air temperature and humidity ratio maps.

The direct evaporative cooling process causes a significant drop in the temperature values for regeneration temperatures higher than 80°C and cycle durations below 300s. The system manages to reach a minimum temperature of 22°C for the highest regeneration temperature (i. e. 95°C) and the shortest cycle duration (i. e.,150s).

2 CONCLUSIONS

The simulation results show a good ECOS’s performance for heat driven air­conditioning applications in the range of temperatures interesting for the use with solar thermal plants. In particular the process reaches a very efficient dehumidification with simultaneous temperature reduction. At the same time COP values are achievable which are significantly higher when compared to those of standard desiccant and evaporative cooling systems employing rotors. Therefore the ECOS process results very promising in particular for climate

Comparison Standard DEC and ECOS’s cycle paths

80

70

60

U

50

ra

40

Ф

a

30

ф

i-

20

10

6 8 10 12 14 16 18 20 22

humidity ratio [g/kg]

Figure 7 — Comparison standard DEC and ECOS’s cycle paths

24

0

4

zones with high ambient air humidity (e. g. Mediterranean and tropic areas). Nevertheless in these climatic conditions in order to ensure a proper air-conditioning operation the regeneration temperature required could be in a range not optimal for standard flat plate collectors. In these cases, evacuated tube or CPC collectors would be desirable.

Furthermore the system design does not pose limits for the realisation of low capacities (200 m3/h) units. Consequently the ECOS system results a good candidate for "split” air­conditioning applications to be connected with the heat distribution network driven by solar combi systems.

The results of the parametric study shown that a further analysis of the single phases duration is needed. Moreover an optimised choice of the employed sorbent material it would desirable, in order to achieve higher performances.

ACKNOWLEDGEMENTS

The work of M. Motta has been supported by a Marie Curie Fellowship of the European Community programme "Improving human potential and the socio-economic knowledge base” under contract number ENK6 — CT — 2002-50515

REFERENCES

[1] EC (1999): Study for the Directorate-General for Energy (DGXVII) of the Commission of the European Communities (1999): Energy Efficiency of Room Air-Conditioners

[2] Henning H. M., (2004): Hans-Martin Henning (Ed.) — Solar-Assisted Air-Conditioning in Buildings, A handbook for planners — (2004) Springer Verlag

[3] Motta M. et al. (2004): M. Motta, H. M. Henning — An original heat driven air-conditioning concept: advanced desiccant and evaporative cooling cycle numerical analysis, Proc. 44° Convegno Internazionale AICARR 2004 — Milano 3-4 Marzo 2004 Vol. II — p. 1149 — 1166

0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1 Specific collector area, m2/m2 -♦-SFC / 0 h —*—SFC / 1 h -■-SFC / 3 h SFC / 6 h — SFC / 12 h ♦ eta / 0 h eta / 1 h ■ eta / 3 h • eta / 6 h —•-eta / 12 h Figure 3: Example of an output graph of the SolarCoolingLight calculation tool. The solar fraction for cooling SFC (left axis) and the net collector efficiency eta (right axis) is shown versus the specific collector area (m2 of collector are per m2 of conditioned room area). The curves correspond to a variation of the heat storage size, expressed in hours of covering the peak cooling load. For this calculation, the meteorological site of Madrid and a load structure of a typical office room was selected and a stationary CPC-collector was chosen. . Easy Solar Cooling tool

An extended version of the SolarCoolingLight calculation tool was created at Fraunhofer ISE, Freiburg. While the identical structure of the combined input file containing meteorological and load data is used, the tool allows more distinction between different cooling technologies and system designs. The desired system configuration may be selected from 11 pre-defined configurations, covering solar assisted air-conditioning systems either with a desiccant cooling system or with a thermally driven chiller and with different types of a backup system. As a further advantage, a reference system may be selected and automatically, a comparison of investment cost, annual cost and other key figures between the solar assisted system and the reference system is generated by running the program. Figure 4 shows two configurations, selected from the program: a desiccant cooling system and a reference system, consisting of an air handling unit with a compression chiller as cold backup.

Although for each hour of the year a complete energy balance including auxiliary energy consumption (e. g., pumps, fans) is carried out in order to determine the annual performance figures, the tool is developed for pre-design studies, as no modifications in the system control can be applied and for the chillers and desiccant cooling systems global performance values can be specified only without considering part-load behaviour.

Nevertheless, the tool is useful for comparative studies, like the study on the energy — economic performance of solar assisted air-conditioning in the SACE-project. Within this study, an energetic and economic assessment of different solar cooling technologies for European sites and for different types of application was performed.

For each investigated system configuration, a reference system was defined and the energetic performance and costs relative to the reference system were compared. Figure 5 shows a result, extracted from this study. For a desiccant system configuration and a reference system as shown in Figure 4, the annual cost (annual payments for investment,

COL

Qc

Qh

BAH

BAH

Й

STH

1

FAN DEC Qh

operation and maintenance) are drawn versus the investment cost (first cost) of the plant. Both, annual cost and investment cost are given as a percentage of the corresponding cost of the reference system. The calculation was repeated for the same application (a lecture room), but located at three different sites: Freiburg (Germany), Madrid (Spain) and Palermo (Italy).

Figure 4: Two of 11 system configurations which may be selected in the EasySolarCooling tool. The system on the left presents the solar assisted configuration of a desiccant cooling system (DEC) with a solar collector array (COL), a heat storage (STH), and a thermal backup (BAH). The system on the right is the conventional reference system with an air handling unit (AHU), a compression chiller (CCH) and a thermal backup (BAH) for air heating in winter.

100% 110% 120% 130% 140% 150%

first cost, relative to reference

FREIBURG_FPC_DEC MADRID_FPC_DEC —A— PALERMO SAC DEC

— 100% annual cost

( = break-even condition)

Figure 5: Simulation results from the SACE economic study/3/: Annual cost versus investment cost of a desiccant cooling system at three European locations (Palermo, Madrid and Freiburg). The costs are given in percent of the costs of the reference system (see Figure 4). The actual first cost correspond to the right hand end-positions of each line; the required first cost to obtain competitiveness in the annual cost compared to the reference system (100% annual cost, relative to reference) are marked by the vertical dashed lines for each site. In the systems at Freiburg and Madrid, a common flat plate water collector (FPC) is assumed, while at Palermo site a solar air collector (SAC) is applied.

The figure reveals the following information, explained for the system located at Freiburg: the investment cost under current market prices are expected to be approx. 140% of the investment cost of the reference system, referring to 115% of the reference systems annual cost. If these annual cost are not allowed to exceed the annual cost of the reference system
(100%), the relative first cost of the solar assisted cooling system should not rise above approx. 113%. The difference in first cost between this value and the initial value of 140% has to be overcome by funding measures and by a reduction of component and installation cost, but is also subject to modifications in the running cost of the system (e. g. an increase in electricity cost), to make the system economically competitive to conventional system solutions. A further decrease in investment cost down to identical investment cost of the reference system (first cost of 100%) finally would lead to annual cost of only 91% compared to the reference system annual cost and hence, economic benefits could be expected beside primary energy savings.

The EasySolarCooling software has been designed for internal use only and is employed within pre-feasibility studies on solar assisted air conditioning. It will be continuously developed according to needs in project work.

Standards for building integration

Solar heating systems are a rather new technology and most building standards have little or no mention of them. In some countries the first guidelines for building integration are developed [6]. The solar heating systems has of course to fulfil all the regulations concerning water tightness, fire protection, strength, water quality etc.. A special item is the regulations for the aesthetics of buildings that exist, especially in northern Europe. They form a barrier for the use of solar heating systems. Proper integration into the building design is the only solution to overcome this esthetical problems. This not only an item in Europe. In China some cities are proposing to ban solar heating systems in new high rise buildings.

Integration in the building process

In the market for solar heating systems for existing houses, the system is sold directly to the house owner. Mostly by the manufacturer or the installer. This sales process gives very little possibility to integrate the solar heating system in the building. For new buildings a solar heater is often not part of the original design by the architect. Many architects are therefor not in favour of solar collectors, because it changes their design. This can be avoided by including the solar heating system from the beginning of the design and building process. It also gives maximum possibility for building integration. The challenge is to get the parties involved interested in solar heating, but it requires another type of sales and marketing.

Analysis and conclusions

The status of building integration can be summarised in the following statements:

• Building integration is not the standard yet, most installed heating systems are not building integrated.

• Building integration has several aspects: Integration in the aesthetics of the building, the building technology site with issues like water tightness, integration in the design and building process, integration in the building standards and guidelines.

• The reasons for integration are cost saving, better aesthetics, better approval by the users, policy makers and architects.

Building integration of solar heating systems has many advantages, but it is little practised.

There are enough good examples available, but this has not yet led to wide spread

application. For further growth of the market for solar heating systems, better building

integration will be essential. This can be explained with several examples:

• In China large cities are considering a ban on solar heating systems, because they look poor.

• In northern Europe the aesthetics for new buildings is an important item. An aesthetically good design can only be reached with building integration.

• In Greece the market for solar water heaters is saturated, further growth is possible for apartment buildings, but therefor better building integration will be essential.

• Building integration can save costs by avoiding building material and saving on installation work

Conclusion

• For further growth of the market for solar heating, it will be essential to better integrate the solar heating system into the building and into the building process.

Acknowledgments

The work in the paper is based on a study funded by UNDESA, UNF and UNFIP.

References

[1] Ree B v. d., Bosselaar L., Li H., Reijenga T, Westlake A., Zhu J., Yang J, He Z. Global status report on the integration of solar heating into residential buildings and implications for the china market

[2] Sun in Action II — A Solar Thermal Strategy for Europe (2 volumes), European Solar Thermal Industry, www. estif. org

[3] Weiss, W., Faninger, G.: Solar heating worldwide — markets and contribution to the energy supply, IEA Solar Heating and Cooling Programme, 2004, www. iea-shc. org

[4] Weiss, W. (Ed.): Solar Heating Systems for Houses, A Design Handbook for Solar Combisystems,

James & James, London, 2003

[5] ‘Soltherm Europe — European Market Report’ — B. van der Ree (Ed.), Ecofys, February 2003, downloadable from www. soltherm. org

[6] Dutch Pre-Standard NVN 7250:2003: Solar energy systems — Integration in roofs and facades — Constructional aspects (in Dutch, draft English version available). July 2003, NEN, Delft, the Netherlands www. nen. nl .

[7] Peuser, A., Remmers, K. H., Schnauss, M.: Solar Thermal Systems — Successful Planing and Construction, James & James, London 2002, ISBN 1-902916 -39-5