Category Archives: Sonar-Collecttors

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.


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.


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


Sorbent thickness

0.25 mm

Heat exchanger plates material


Plates thickness

0.5 mm

Plates distance

3.5 mm

Table 1 — Structural heat exchanger characteristics

Return air temperature


Return air relative humidity


Return air humidity ratio

10.5 g/kg

Ambient air temperature


Ambient air humidity ratio

20 g/kg

Heat recovery efficiency (reg.)


Adsorption Phase

1 t*

Regeneration phase

0.8 t

Pre-cooling phase

0.2 t


Regeneration temperature range


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).


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















6 8 10 12 14 16 18 20 22

humidity ratio [g/kg]

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




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.


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


[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,










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


— 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


• 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.


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


[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

Structural Components of Skylights

Figure 3: Structural Components of Skylights

The structural components of the skylights represent a constructional solution which will ensure that all transparent surfaces (even with multiple layers) will properly join the internal surfaces of the building interior while also establishing a unified and coherent and integral architectural, building constructional and structural solution. The structural complexity of skylights is well represented by this skylight detail, where the horizontal steel I beam behaves not only as main load-bearing component but also utilizes the drainage system and even hosts the "hidden” artificial lights of the interior space bellow. The illumination characteristics of the skylights are determined by the structure of their transparent — and reflective (light-guiding)

surfaces and by the obstructions. The resulting illumination characteristics of the overall system will be determined by the material characteristics of the surfaces and by the geometry of the skylight. The daylighting characteristics of the interior spaces are the joint result of the illumination interactions of large — and small-; commensurable and incommensurable surfaces of the hemisphere, of the skylight and of the interior space of the building.

Absorption Process

ATequ ~ f (Tsol,^sol) Equ. 1

Figure 1. Two connected vessels containing liquid refrigerant and solution composed of the absorbent and the refrigerant.

Absorption occurs when one material, the refrigerant, is absorbed into another, the absorbent, to form a ‘solution’. The vapour pressure of the refrigerant for this solution (Ps0i) is lower than that for the liquid refrigerant (Pref). If two vessels are connected together as in Figure 1, one containing liquid refrigerant and the other the absorbent, refrigerant will be transported from the left hand vessel to the right due to this difference in vapour pressure. This results in evaporation and cooling in the refrigerant and absorption and heating in the solution, leading to a lower absorbent concentration. This transfer will continue until equilibrium is achieved when the vapour pressures in the two vessels are equal. However, the refrigerant temperature (Tref) will be lower than the solution temperature (Tsol). The difference between these two temperatures (ATequ) will be dependent on both the temperature and the concentration of the solution, as shown in Equ. 1, where ^sol is the mass fraction of the absorbent in the refrigerant.

If heat is applied to the refrigerant vessel at this lower temperature, the temperature will rise and with it the vapour pressure. This will result in transport of refrigerant vapour to the solution vessel where it is absorbed in the solution, releasing heat. If this heat is removed from the solution vessel, at a higher temperature than the refrigerant vessel, the process can continue. This is essentially a heat pumping process that can be used for cooling or heating. However, the solution gets weaker in terms of absorbent, and the temperature required to give a certain vapour pressure will decrease. Thus the temperature difference, the temperature lift, between the two vessels will also decrease, reducing the usefulness of the heat pump. In order to maintain the temperature lift, the solution needs to be regenerated by desorption, which in principle is the reverse process, with heat applied at a higher temperature to the solution vessel and removed at a lower temperature from the refrigerant vessel. Two vessels connected as in Figure 1 can be used for intermittent cooling, but in order to be able to simultaneously provide cooling while regenerating the solution, two pairs of vessels are required. For a single effect absorption chiller, these are connected to form a continuous cycle.

Different working pairs have been suggested in the literature (Macriss et al., 1988; Macriss and Zawacki, 1989), but only two are commonly available commercially: LiBr as the
absorbent and water as the refrigerant for comfort cooling, where the evaporation of water cannot go below 0°C; and refrigerators using water as the absorbent with NH3 as the refrigerant. Cycles using water/NH3 can also be used for comfort cooling, but they are not common. A number of different cycles have been developed and tested, and are treated in various studies (Herold et al., 1996; Srikhirin et al., 2001). Nearly all studies have worked on the cycles themselves, and very few have looked at the possibility of energy storage, although it is possible by storing the relatively concentrated solution between the generator and the absorber. Although this requires several extra vessels, it could be used instead of external storage devices (Berlitz et al., 1998). However, the potential for this type of storage is limited by the practical concentration variation achievable in a machine, where the heat exchanger in the absorber and generator are critical. In addition, crystallisation has to be avoided so that the solution can be pumped between vessels.

Adsorption Process

Adsorption, the binding of a sorbate onto the surface of a sorbent, can also be used in a similar way to absorption as in Figure 1. The major difference here being that adsorption is a surface phenomenon and can only be used with solid adsorbents, and thus a complete heat pump cycle cannot be built up in the same way. Instead the desorption/condensation phase, also called charging phase, and the evaporation/adsorption phase, also called discharging phase, must be separated in time. Again there are a number of different working pairs that have been studied (Dieng and Wang, 2001; Wongsuwan et al., 2001; Henning and Wiemken, 2003), and similar to absorption, those with water as the sorbate are limited to comfort cooling applications. Adsorption can be used in open or closed cycles. Common to all is the fact that the sorbate must be transported into the structure of the substance and also that the heat has to be transported to/from the solid. This creates practical problems for the design of heat exchangers and the matrix for the solid.

Thermal Storage

Adsorption has also been studied for thermal energy storage, especially for solar heating and cooling applications in recent years (Mittelbach et al., 2000). Due to their high energy density compared to sensible heat storage in water, the potential for long-term heat storage has been studied. This work has focussed mainly on water together with zeolite, silica gel or modified silica gels. An energy density of 134 kWh/m3 silica gel material has been achieved in a practical system (Nunez et al., 2003) whereas 160 kWh/m3 has been achieved for a small (1 kg) sample of zeolite and theoretically 233 kWh/m3 for impregnated aluminosilicates (Janchen et al., 2004). However, the storage density is dependent on the pressure in the system and thus the desorption temperature. Zeolites require, in general, higher desorption temperatures than silica gels. A comparison of storage capabilities for different materials for sorption systems (Mugnier and Goetz, 2001) showed that for refrigeration at -20°C, a solid-gas chemical reaction with ammonia gave the highest energy densities, whereas for comfort cooling the highest energy densities were achieved by water with NaOH for absorption, and for CaCl2, MgCl2 and Na2S for chemical reactions. These chemical reactions are the binding of water to hydrates of the salt.

Optimization of a small-scale solar-driven ejector refrigeration system

Wimolsiri Pridasawas, M. Sc., Department of Energy Technology, Royal Institute of Technology, Sweden

Per Lundqvist, Ph. D., Assoc. Prof., Department of Energy Technology, Royal Institute of Technology, Sweden


The TRNSYS-EES simulation tool is used to simulate the characteristic of the solar- driven ejector refrigeration system with a flat-plate, double-glazed solar collector. Butane is used as a refrigerant in the cooling subsystem and water is used as a heating medium in a solar-collector subsystem. The performance of the system is shown in terms of coefficient of performance (COP) for the refrigeration subsystem and system thermal ratio (STR) for the whole system. The simulation results show the performance of the system, the annual electricity usage by the pumps and the auxiliary heater at different solar collector area, storage tank volume and water flow rate. The system performance depends on the solar radiation and the operating temperatures in the refrigeration subsystem. The STR is high when the solar radiation is high. The maximum STR that can be obtained is about 0.25 at a COP of

0. 55. The optimum solar collector area for the average cooling load 4 kW is about 50 m2. The system operates only during daytime, thus the volume of the well-mixed storage tank does not significantly affect the performance and the electricity usage of the system.


Refrigerators and air-conditioning systems are mostly driven by electricity and account for about 15% of the world’s electricity consumption (Lucas, 1998). Solar energy can be converted to both thermal and electrical energy, both of which can be used to drive refrigeration systems. The demand for cooling is generally high when the solar radiation is high. The performance of an electricity-driven refrigeration system is quite high but it requires photovoltaic panels, which are expensive and have low efficiencies. These systems, however, can be built in small sizes, making them suitable for applications such as vaccine transportation or cooling boxes. An air-conditioning system is used to control temperature and humidity for human thermal comfort. The demand for this application is high in a densely populated area such as big cities. The solar thermal-driven refrigeration systems are more suitable for air-conditioning applications due to the lower installation cost, furthermore it can provide high cooling capacity.

A solar-driven ejector refrigeration cycle is quite a reliable and simple system. An interesting advantage that can be noticed is its ‘low temperature heat supply’ that allows it to be integrated with a simple solar collector such as a flat-plate solar collector. Furthermore, this system is easy to install, design and operate. Several research groups have studied the ejector refrigeration cycle in different perspectives but only a few of the solar-driven systems have been presented. Huang (1998), has developed a solar ejector cooling system using R141b as the refrigerant; the overall COP is about 0.22 at a generating temperature of 95°C, an evaporating temperature of 8°C, and solar radiation of 700 W m-2. Several simulation models are found in the literature of Dorantes (1996), Sokolov (1992) and Al-Khalidy (1997). Chlorinated refrigerants such as R142b (Dorantes, 1996), R114 (Sokolov, 1993) or R113 (Al-Khalidy, 1997) were recommended due to a high performance. These refrigerants, however, have negative environmental effects.

Some environmentally benign refrigerants for solar-driven ejector refrigeration systems are introduced in the literature of Pridasawas (2003) including a comparison of the technical feasibility and performance of each refrigerant.

In this paper, a TRNSYS-EES simulation tool was used to model and analyse the performance of a solar-driven ejector refrigeration system using butane as a refrigerant. TRNSYS is a transient systems simulation program with a modular structure (Klein, 2000). It is widely used for analysis of time dependent systems such as solar systems, low energy buildings and HVAC systems. The Engineering Equation Solver program or EES is generally used for solving a set of algebraic equations and initial value differential equations (Klein, 2002). It provides built-in mathematical and thermophysical property functions suitable for cycle simuations. The whole system is simulated by using TRNSYS but the model of the ejector refrigeration sub-system is developed in EES. The weather data from Bangkok, Thailand is chosen to represent the warm climate for this simulation.

The system’s performance mainly depends on the solar radiation and the operating temperature. The performance decreases in inverse proportion to the condensing temperature but it increases when the generating temperature increases. High generating temperature requires high outlet solar collector temperature but the efficiency of the solar collector decreases at the high outlet solar collector temperature. The optimum operating condition for the highest system performance should be considered. The optimum generating temperature, solar collector area and storage tank volume were studied by using TRNSYS-EES tool as mentioned above.

Diffusing projection screen

The dimensions, positioning and coating characteristics of the triangular projection panel are detailed in (Andersen et al., 2001; Andersen, 2004): a diffusing white paint manufactured by LMT allows to obtain an almost lambertian surface (perfectly diffusing), with only a 2.6% difference to the theoretical model.

The removal of screen covers, necessary to perform BRDF measurements, aims at leaving the incident beam path free, while the controlling of its shape is taken care of by the ellipses cut out from the metal sheet.

To minimize the blind zones, these screen covers must present elliptic shapes as well. Their exact geometry was determined following a similar procedure as for the metal sheet:

• First, their theoretical dimensions and positions were deduced by trigonometry on the basis of the intersection of a perfectly parallel beam (reaching the sample at different Qi angles) with the tilted detection surface (accounting for the shift between sample and detection screen base planes.

• Then, using on the results provided by the sample illumination analysis with the actual light source and on the metal sheet ellipses dimensions, adjusted horizontal and verti­cal axes for the screen ellipses were estimated, to which a 2 mm margin was added to avoid edge effects.

• After that, to determine the actual dimensions of the cut out covers, the thickness of the screen had to be taken into account; on the other hand, the covers insertion required a slant between the upper (external) and lower (internal) sides of the screen, chosen unique and equal to 20° to ease the screen manufacturing. To leave the beam’s pas­sage free through a screen of significant thickness, larger upper ellipses are required when the angle between the incoming beam and the screen plane increases (i. e. when I Q — Q0 I increases). The ellipses were thus adjusted accordingly, depending on each one’s incident tilt angle.

• Finally, as the above adjustment was only necessary for the ellipses half farthest from the Qj = ©0 direction, their vertical axes (and thus the blind zones) were reduced by re-centering them to open a passage for the actual beam only, still accounting for the screen thickness and a constant 20° slant.

The elliptic covers are held in place by small and strong permanent magnets inserted in the screen central piece. To achieve their removal and repositioning, a “permanent electro­magnet” (PEM) is used, i. e. a permanent magnet that can be deactivated by powering the surrounding coil. This PEM is mounted on a small wagon running on two rails parallel to the main axis of the screen thanks to an indented belt forming a closed loop. An additional on-board mechanism allows it to move up and down from approximately 3 cm, in order to extract and replace the covers. To ensure a reliable lifting, a mechanical “extractor” was added, using four screw-like pins that get inserted in four slots carved in each cover, shown on Figure 5(a); centering pins were added as well on protruding fingers to ensure a reliable positioning. An extra shift was implemented for the wagon movements to allow the extraction system to have a secure grip on the covers.

The limitations in the rails length made it impossible for this extractor to reach the tip cover. Its handling thus required an additional PEM device, together with some extra commands.

(a) Screen covers (b) Wagon and steering rails

(c) Obstructing cover (d) Extraction (e) Removal (f) Illumination

Figure 5: Motorized screen with removable covers for incident beam path.

The wagon is driven by a stepping motor, controlled by a specific ISEL micro-controller with a RS-232 interface. A typical cycle of extraction, removal and replacement of a cover is sequenced as follows: [12]

• Wagon positioned out of the beam path and kept in place as long as needed to com­plete the image acquisition and processing phase;

• Wagon moved back above the open hole, PEM lowered, deactivated then lifted up empty, the cover being back in place.

Once the wagon movements were adequately calibrated to position it right above each cover, this new design was tested successfully with hundreds of random extractions at different screen inclinations.

The definitive screen panel is shown on Figure 5(b), where the wagon is in position to re­move the tip cover and where all other covers are missing. Figures 5(c) to 5(f) illustrate the sequence of events taking place when the projection screen obstructs the incident beam path.

Numerical optimisation. Domestic hot water applications

The facade design shown in Figure 1 was numerically investigated. The results obtained for two climatic conditions, Barcelona and Geneve, are shown in Figure 4.

Figure 4: Monthly performance of the facade for domestic hot water production: solar fraction and solar efficiency. Consumption profile 1, (with noon draw), tank stratification considered as given by 5 nodes.

It is observed that solar fraction is less variable in Barcelona climate, while it suffers depreciations in Geneve climate in the winter months. Solar Efficiency presents for both climates a similar performance. For April month, for instance, 32% of total domestic hot water load may be satisfied by the solar facade if it is located in Geneve. For a Barcelona location, facade provides 40% of total load. Regarding Solar Efficiency, the variation is lower, 46% for Geneve and 49% for Barcelona. Annual values are shown in Table 4, where OTL stands for the mean annual outlet temperature from water tank.

Table 4: Annual perfomance for both climatic conditions: Barcelona and Geneve. Consump­tion profile 1, stratification represented by 5 nodes________________
















The influence of the consumption profile has been analyzed for a design addressed to Barcelona climate. Monthly solar fraction and solar efficiency are shown in Figure 5 for the two profiles described in Tables 1 and 2. Consumption profile 1 (with noon draw), presents better results all around the year. The differences are more noticeable in the central months (June, July and August). In the rest of the year, differences are negligible.

The influence of stratification within the tank has been analyzed conside­ring two possible le­vels. Stratification represented by 5 and 10 nodes. Di­fferences obtained are negligible, al­though it is pos­sible to get larger useful energy with

a higher level of Figure 5: Monthly performance of the facade for domestic hot water stratification in wa- production: solar fraction and solar efficiency. Two consumption pro — ter tank Data are files (1: morning, noon and evening, 2: morning and evening). Data corresponding to Barcelona

Table 5: Annual perfomance for both stratification models (10 and 5 nodes) and two con­sumption profiles. Data corresponding to Barcelona climate



] OTL[°C]



10 nodes,

consumption 1





5 nodes,

consumption 1





10 nodes,

consumption 2





5 nodes,

consumption 2