Category Archives: Sonar-Collecttors

Temperatures inside the absorber

The tests were disturbed by the clouds passing in the sky and by the drift of the heliostats during the solar tracking. The second effect was sensibly reduced after the change of the tracking algorithm in the control software of the heliostats field. For this reason, the last days of tests were less affected by instantaneous defocus to correct the aiming point, and the system behaved more stably. It yields more quasy-steady states to compare the results. Thus, only the results of the square metallic cups tests are presented, in reason of clarity. Moreover the position of the inner thermocouples is better fixed, because they were installed by Emitec during he manufacturing and they could not move, what leads to a higher data reliability of these tests.

Fig.3. Test results for the metallic square cups under a radiation intensity of 370 ± 20 kW/m2

Effect of cell size and mass flow rate on the axial temperature distributions.

For a roughly constant value of the solar radiation flux, the axial temperature distribution around the central channel is shown in fig.3 for different mass flow rates per cup. It is assumed that the mass flow rates per cup are all in the same range and then, the mass flow rate per cup is estimated by dividing the overall mass flow rate per the number of cups (35 ones for the square cups tests, since once cup of the Solair receiver was removed and closed). In this chart, the wall temperatures are plotted versus the parameter x/L, being x the distance from the irradiated entrance along the channel and L the total channel length. Inside each cup, a similar temperature profile is maintained for the three flow rates, but the mean temperature increases as the mass flow rate decreases, as it was expected.

The influence of the cell size on the axial distribution is also obtained from data depicted in the figure: for a smaller cell or channel diameter, the maximum wall temperature is reached earlier, and then the temperatures decrease with a stronger slope. On the other hand, for absorbers with a larger cell size, the maximum temperature is reached deeper in
the channel. With only 4 data we cannot assure, as it seems in the chart, that the maximum for the 500 cpsi absorber is located at a x/L value of 0.43, which corresponds to a distance of 30 mm from the entrance. It should be somewhere between 10 and 30 mm, but the positions chosen for the thermocouples do not allow to determine the location of the maximum temperature. Similarly, for the absorber with 600 cpsi, the maximum temperature should be near the thermocouple located at 10 mm on the channel axis.

These temperature profiles give information about the penetrability of the radiation into the absorber channels: for a higher cell density (600 cpsi), the cell size is smaller and the rays penetrate less than for an absorber with lower cell density (500 cpsi). This behaviour is related to the volumetric effect, since the radiation is absorbed in the first millimetres from the entrance and not only on the aperture surface. Absorbers with wider channels have more “volumetricity” since the volume irradiated is bigger, but the specific surface to absorb and transfer the heat are both smaller, resulting in lower temperature values.

Fig.4. Effect of the solar radiation intensity for an overall mass flow rate of 0.3090 ± 0.0003 kg/s, that corresponds to an estimated mass fow rate per cup of 0.0088 ± 0.0001 kg/s.

The behaviour for different solar radiation fluxes under the same mass flow rate is as expected: higher radiative fluxes lead to higher absorber temperatures, but the temperature distribution is similar in all cases. In fig. 4 only the results for the cup with 600 cpsi are presented, the profiles of the other absorber are analogue to those shown in the previous chart.

Air temperature distribution at the absorber outlet:

To obtain the temperature distribution for the air exiting the absorber, the data from the thermocouples located in several positions at a cross section 2 cm behind the absorber have been interpolated in Matlab with the function called Griddata, which fits a surface to the data in nonuniformly-spaced vectors. With the ‘v4’ method, griddata produces smooth surfaces. Once the data are interpolated for the defined grid, the level contours of air outlet temperature are plotted.

Fig.5. Level contours of the air outlet temperature for 3 mass flow rates through the 600 cpsi absorber in the first row and the 500 cpsi absorber in the second row

The temperatures extrapolated near to the edges of the square absorbers must not to be considered because there are not experimental points near the edges to function as boundary conditions. For this reason the calculated temperatures are too low with respect to the real temperatures in the square corners.

The main difference between the distribution of the two absorbers is caused by the temperature at the absorber centre: in the 600 cpsi absorber the central temperature is in the same temperature range as the thermocouples next to it, while the central temperature of the 500 cpsi absorber is always lower than the temperatures at each side of the center. The colder central region in the 500 cpsi absorber suggests a non-uniform air mass flow distribution through this cup. The air flow distribution through the 600 cpsi absorber seems to be more uniform. From this, it is possible to conclude that the air flow distribution for both cups is different.

On the other hand, the warmer region that is observed at the left upper part of the 500 cpsi absorber may be due to the fact that the solar radiation was focused at the centre of the Solair, which is located in that position regarding to the 500 cpsi absorber cup (position 22).

Other important aspect is the great difference between the absorber wall temperature and the air outlet temperature. In volumetric solar receivers the air is supposed to reach the same temperature of the absorber matrix before exiting the channel. However, in these tests between the wall temperature at 20 mm from the absorber exit and the air temperature measured 20 mm after the exit, there is always a difference about 150°C for the 600 cpsi absorber and about 70°C for the 500 cpsi absorber. This means that the air does not reach the same temperature as the absorber in the first centimeters of the absorber channel, as some simulations suggest. Air temperature is lower than the absorber temperature all along the channel. Thus, air is always cooling the absorber. Nevertheless, in the first part of the
absorber channel, radiative heating up dominates over convective air cooling. From the point where radiation has no more effect over the channel walls, the absorber temperature begins to drop, more rapidly when that point is nearer to the entrance.

The extrapolation of the temperatures inside the absorber, considering only the last falling slope, gives quite a good estimation for the air temperatures.

A way of comparing the relative "heat transfer efficiency” from the absorber monolith to the air for each absorber module is to calculate the relative difference from the last absorber wall temperature (measured in the central axis at 20 mm from the exit) to the air outlet temperature in the same axis. The results show a better relative efficiency for the 500 cpsi absorber, but the absolute values are lower than for the other absorber. The variation in mass flow rates does not affect to the relative temperature decrease, while the variation in radiative fluxes has a slight effect: for higher radiation, worse heat transfer.


— The temperature distribution along the channel for a determined channel geometry (or absorber cell density) is the same for different mass flow rates or solar radiation fluxes. It achieves a maximum value inside the channel and then the temperature decreases due to the heat transfer to the cooling air.

— For a smaller cell or channel diameter, the maximum wall temperature is reached earlier, and then the temperatures decrease with a stronger slope. On the other hand, for absorber with a larger cell size, the maximum temperature is reached deeper in the channel.

— — The great difference between the absorber wall temperature and the air outlet temperature indicates that the studied absorbers are not sufficiently efficient to transfer all their energy to the air. The relative heat transfer efficiency is better for lower cell densities.

To sum up, the studied absorbers do not behave as it is expected for an ideal volumetric solar receiver, where the maximum matrix temperature would be at the end of the channel and the air would reach the same temperature of the absorber matrix before exiting the channel. These experimental results can be used to validate theoretical (numerical or analytical) models that reproduce the heat transfer in duct volumetric solar receivers.


The authors acknowledge the collaboration with the German enterprise Emitec, and the efforts of Mr. Arndt-Udo Rolle in the manufacturing of the samples are specially appreciated. Results discussions with Ma Jesus Marcos are also very appreciated.

ANALYSIS. Constant Thickness Model

To simplify the mathematical analysis, the following assumptions were made for the constant thickness film model: the wall is considered isothermal, i. e., the cooling water flow is not considered, the flow is laminar and fully developed for the liquid-desiccant and the air, the desiccant flow is smooth (not wavy), the physical properties are constant, there is thermodynamic equilibrium at the desiccant/air interface, air is an ideal gas, no shear forces are exerted by the air on the desiccant, the body force in the air is negligible, diffusion in the direction of the flow is negligible, species thermo-diffusion and diffusion — thermo effects are negligible, the rate of water vapour absorption is small, i. e., the velocity in the transverse direction is negligible, the solubility of air in the desiccant solution is negligible and the film thickness is constant.

for the liquid-desiccant film,

and for the air,

Fig. 2: schematic diagram of the dehumidifier channel with notation.

d u








— + Pdg = 0

■ = ая

d 2T


dy2 5 C

"d dx Dd dy2




The boundary conditions are:

at x=0,

Td=Tdi, Ta=T

at y=0,




at y=Sd,

Td=Ta, ud=ua

= 0, ud=0,




‘ dy

dp d2 ua

dx = A. y 2


Ua STa = «. SlT; dx a dy2


u C. — d ^

a dx D dy2





= 0,


Considering the assumptions above and the geometry presented in Fig. 2, the governing momentum, energy and species equations are:


at y=Sc


dT dC du

a = 0 , ^ = 0 , ^ = 0 .

dy dy dy

The velocity profile for the liquid-desiccant is calculated integrating (1) twice across the desiccant film:





and considering the continuity across the film and a channel 1 m wide:

md = £’ Pduddy (12)

Solving (12) allows the development of the expressions for film thickness and velocity profile, according to the desiccant mass flow, md:




Pdg j




3md ( 2у _ т! л

2 Pd v^d ^d J

The velocity profile for the air stream can be calculated also through the momentum and continuity equations for the air stream:


ma = 2‘PaUady




3 Pa


fo — sc )2





2Pa {Sd ~SC )3 _

2 (y2~5l )+(A — y)

dp 1 dx pa


The energy and species balances at the interface (y=8d) give the additional matching equations:

_ k k T. П h C

d dy a dy Pa “ ff dy













18.0153 pw

28.9645p -10.949pw

The equilibrium water mass fraction in the air at the interface is given by the water vapour pressure of the water in the desiccant solution at the interface, pw(Td, Cd), and Dalton’s law applied for a water vapour/air mixture:

Equations (2),(3),(5),(6),(18) and (19) were approximated using the finite-difference method. Central-differencing was used for the diffusion terms and backward or forward — scheme for the convection terms, depending on whether the air/desiccant flow configuration was co-current or counter-current. The equations were solved simultaneously using the software package Microsoft EXCEL[6]. In the “y” direction, 7 nodes were used for the liquid-desiccant film and 40 for the air (Figure 3). The interface between the last node in the desiccant side and the first node in the air side was represented by an additional node. This interface node has no volume, but provides, through equations (18), (19) and (20), the coupling between the air and liquid streams and between the energy and species equations. The number of nodes in the flow direction varied according to the height of the channel.

representation for one step in flow direction.




The variable thickness model uses the same equations, assumptions and number of grid nodes of the constant thickness model. The difference is that the thickness of the desiccant film is allowed to vary and is recalculated using equation (13), for every step of the simulation. Once the film thickness is recalculated, it is possible to obtain a new velocity profile for the desiccant using (14), i. e., the profile is still the profile for the fully developed laminar flow, but recalculated to account for the change in

Fig. 4: schematic of grid with thickness change in the desiccant film.

Variable Thickness Model

mass flow and film thickness. This procedure is similar to the one adopted by Jernqvist and Kockum [7] for developing falling film flow. To reduce the computational time, and because the changes in film thickness are small, the change in the air flow velocity profile was neglected.


A solar cooling system consisting of solar thermal collectors, a heat storage (optional), an absorption chiller, a cold storage (optional) and the application system (e. g. room cooling) is a very complex arrangement. The performance of some components (e. g. collector and cooling tower) depends on ambient conditions (e. g. air temperature, air humidity). Because of the high number of interrelations between the different components and between some components and the ambient conditions it is not possible to calculate the performance of the whole system and to compare different types of certain components (e. g. collectors) manually. Simulation programs have to be used.

Using the capabilities of the TRNSYS /1/ package the performance of the small capacity absorption chiller in combination with different kinds of collectors and re-cooling options was evaluated /2/. For flat plate and vacuum tube collectors a comparison of different models (manufacturers) was done using published test data. For each type a reference collector model with a good performance and an acceptable cost-performance ratio was chosen. The properties of these models were used for further calculations. For parabolic trough collectors the STEC-Library was used (DLR model of the IST trough) /3/.

At first a system with a fixed cold water inlet temperature (12 °C; 18 °C) to the absorption chiller was simulated with climate data of Huelva/Spain.

For all combinations of the three collector types and two re-cooling options (wet cooling tower and dry cooler) the delivered cooling energy from April to August was calculated for five collector areas. The results are shown in Figure 8 and Figure 9.

The results show the great influence of the re-cooling version. If a dry cooler is used in­stead of a wet cooling tower the collector area has to be increased to obtain the same quantity of cooling during the summer.

Comparing the figures 8 and 9 the influence of the cold water temperature can be evalu­ated. If the room cooling system is capable to meet the cooling load with a higher cold wa­ter temperature (greater heat exchanger area needed) the cooling energy delivered by the absorption chiller increases (or a smaller collector area is sufficient to receive the same amount of cooling).

The simulations also provided other information such as the maximum cooling perform­ance of the absorption chiller during the operation period, distribution of the cooling energy over the month (and days), water demand for the wet cooling tower, delivered cooling en­ergy per collector area. By comparing the results for different collector areas an optimal size can be chosen.

At a second step a more realistic system with changing cold water temperatures and in­cluding the thermal behaviour of the building (cooling load) was simulated. To compare the different arrangements

— the excess temperature: ATe = TRoom — 26 °C

— the duration (tet) of the periods with inside room temperature above 26 °C and

— the integral: |ДТЕ dTET [Kh]

were calculated. Figure 10 shows the results for one configuration with different collector areas.

Simulations using sea water for re-cooling the absorption chiller were also done. The re­sults for Huelva showed that there is no higher performance of the absorption chiller than using a wet cooling tower. Sea water cooling has some other restrictions too (only near the sea, high cost, need to use suitable materials).

40 50 60 70 80 90 100

Gross Collector Area [m2]

Figure 8: Delivered cooling energy for different configurations in Huelva and a fixed cold water inlet temperature to the absorption chiller of 12°C

Figure 9: Delivered cooling energy for different configurations in Huelva and a fixed cold water inlet temperature to the absorption chiller of 18°C

All the results are valid for the specified equipment (collectors, coolers) only. For evalua­tion of a certain project simulations have to be done again with the characteristics of the equipment to be used and suitable climate data.

The experiences gained during the simulations and supervising solar cooling installations showed that great care should be taken regarding system control, storage integration and design of the room cooling system. As mentioned before the cold water temperature is important for the achievable cooling output.

—І і і p і і T ‘ I T

40 50 60 70 80 90

Gross Collector Area [m2]

Figure 10: Integral of the excess room temperature in Huelva for different configurations


A H2O/LiBr absorption chiller with a nominal cooling capacity of 15 kW was developed. A special heat exchanger design permits the use of low temperature heat such as solar or waste heat to drive the absorption chiller.

After prototype testing a field test at three sites was carried out in the summer of 2003. The field test showed good results. The absorption chiller is working reliable and flexible over a wide range of external conditions.

Because of the high number of cooling applications in the capacity range below 50 kW the new chiller has the potential to increase the usage of solar thermal energy for cooling pur­poses.

The waste heat of cogeneration systems can also be used to drive the absorption chiller creating new possibilities for trigeneration systems (power, heat and cold).

The performance of solar thermal cooling systems can be predicted using the capabilities of system simulation programs.



coefficient of performance


excess temperature


global solar insolation


solar collector efficiency


ambient temperature


duration of periods with temperature


temperature of the collector fluid

above 26 °C

T Room

room temperature


/1/ Solar Energy Laboratory, University of Wisconsin (2000). TRNSYS 15, A Transient System Simulation Program, Madison, WI, USA.

/2/ Heinrich, C. (2004). Modelling and Simulation of solar thermal driven single effect absorption chillers of small capacity for climatisation, Diploma Thesis (German), Dresden, Germany.

/3/ Schwarzbozl, P.; Eiden, U. e.; Pitz-Paal, R.; Jones, S. (2002). A TRNSYS Model Library for Solar Ther­mal Electric Components (STEC) — Reference Manual — Release 2.2, Cologne, Germany.

Figure 13: Radiance model of Amon Carter Museum’s south gallery on November 28 at 3:40 PM; isolux contours (top). North galleries adjacent to atrium (2nd floor)

Two of the windowless north galleries of the second floor receive direct sunlight that passes through the high southwest windows of the atrium (see yellow areas in Figure 4- right). Illuminance levels measured over the painting in one of the two galleries on January 19, 2003 (Figure 15) reached an extremely high value of 11,500 lux, which is about 57 times higher than the recommended IES standards for museums. The total illuminance-hour per year over these display areas over exceeds the total exposure limits for moderately light susceptible displayed materials. The fish-eye photo (Figure 16) shows that the painting receives direct sun for one and a half hours in the early afternoon from November to January. The high illuminance levels over the display areas are due to the high visible transmittance of the atrium’s glass (Tvis=59%). The glass of the atrium does not have any shading device to intercept the incoming direct sun to the adjacent galleries. To protect paintings from over exposure to light, curators have been rotating oil paintings frequently (see dates on Figures 15 and 16).

Figure 15: Oil painting (1848), Amon Carter Museum’s gallery adjacent to atrium, January 19, 2003, 4:30 PM.

Figure 16: Fish eye photo taken from painting’s viewpoint with sun path diagram, at gallery adjacent to atrium (March 5, 2004).

Figure 17: Direct glare on visitor’s eye at Amon Carter’s north gallery adjacent to atrium (March 5, 2004, at 4:50 PM).

Another problem observed in these galleries is the glare that visitors experience while moving through the gallery and to the atrium. Figure 17 illustrates how a visitor has to protect her eyes from direct sun while trying to see the painting and read the sign next to it. This visitor is receiving about 7,770 lux over her eyes when not covered. The extremely high variations of light levels within the field of view of the visitors to these galleries makes them uncomfortable to adjust their eyes to the low light levels over the paintings and to the bright atrium’s glass areas that are within her field of view.

Table 1: Summary of Lighting Conditions at the three museums.

Modern Art Museum

Kimbell Art Museum

Amon Carter Museum

Sunlight penetration on galleries

■ West-facing galleries

■ No (entrance of daylight is well controlled)

■ East-facing lobby

■ South-facing gallery (2nd floor)

■ Interior galleries adjacent to atrium (2nd floor)

Daylighting systems (orientation)

■ Toplighting (north-, south-facing clerestories)

■ Sidelighting (west­facing)

■ Windowless galleries

■ Toplighting (narrow- strip skylights with reflectors)

■ Sidelight high strip windows (east-, west­facing)

■ Toplighting (high windows in atrium)

■ Sidelighting (east-, south-facing)

■ Windowless galleries

Range of illuminance levels on display areas

■ 50 lux to 4,000 lux (under direct sunlight)

■ 50 lux and under

■ 50 lux (windowless galleries)

■ Up to 11,000 lux (under direct sun)

Display objects

■ Oil paintings

■ Sculptures (wood)

■ Paper, prints

■ Photographs

■ Oil paintings

■ Watercolors

■ Oil paintings

■ Daguerrotypes

■ Photographs

■ Sculptures (metals)


■ Blue tinted, Tvis=59% with white interior screens.

■ Skylights

(polycarbonate with UV filter)

■ Dark tinted, Tvis=5%, UV protection

■ Tinted, Tvis=12%

■ Clear glass, Tvis=60%


The most noticeable problem, in the galleries of two of the museums presented in this paper, has been the sunlight penetration over the displayed museum objects. This fact is harming valuable art collections, and also creating visually uncomfortable environments for visitors. Sunlight penetration occurs mainly in galleries with side lighting windows that face the sun — east, south, and west-; orientations that are the most difficult to control. Even though the time of over exposure is relatively limited, the illuminance levels reached at these times are extremely high, 10 to 57 times the maximum recommended IES standards for light susceptible objects. Dark tinted glass, screens, or small overhangs are not enough measures to block the entrance of direct sun over the exhibit areas at the Modern Museum (west-facing galleries) and at the Amon Carter Museum (east-facing galleries).

To control the light levels in these galleries more aggressive changes should be done, like modifying the fagade by adding horizontal or vertical shading devices, external louvers, trellises, or trees to filter sunlight. All these modifications could change the image of the building, which may not be acceptable by designers. There are other less intrusive solutions that include the use of miniaturized sunscreens, interior louvers or baffles; or computer controlled dynamic window systems (Reference 6), where direct sunlight over sidelight windows can be intercepted and filtered reducing interior light levels as low as 25 lux or under.

On the other hand the third museum, the Kimbell Art, presents a well thought and carefully designed toplighting system that can introduce adequate light levels throughout the year to accommodate the lighting requirements of exhibits, and at the same time provide a connection to the exterior environment by rendering the galleries with natural light. This museum is an excellent example that it is possible to successfully illuminate museum galleries with daylighting.

If we already know that sidelight windows in museum galleries create many problems to the display of light susceptible artifacts, why are we still including them as a source of illumination in museum galleries? Is it due to the lack of knowledge about solar geometry? Or it is just that architects underestimate the effects of sunlighting in museums? The answers to these questions may not be clear and simple, meanwhile we might be finding museum galleries with sidelight windows with strong sunlight over display areas, such as in two of the three museums presented in this paper (see summary in Table 1). In the meantime, curators and facility managers at these museums have to create means to protect their valuable art collection (i. e. rotating display objects, use boards to cover them on a daily basis, display least light susceptible objects, cover completely the windows with boards or black cloths, among other solutions) (Reference 7). The main goal of this paper was to make us reflect on the way that museums are still being designed, and on how could we make them better to preserve and display artifacts that are important part of human history.

The air solar system

To assure a complete self-sufficient system, the air solar system should cover the 20% of the remaining demand. With this aim, 12 sq m of Solarwall system (the same surface of the solar panels) has been added.

Figure 11 shows how the system works. A fan collects the preheat air from the Solarwall and it is introduced through a special insulated channel directly to the living room. The system keeps the house dry also when it is not occupied for long periods. In sunny days the air system will delay the switching on of the conventional solar water system.

The calculation indicates that only in December the self-sufficiency could not be reached. Nevertheless, the results depend completely on the weather. In fact, on a sunny day the house could be heated mainly by the air solar system, but on a cloudy day the risk to be insufficient is very high.

Then, it is necessary to integrate the solar system with a boiler, using renewable energy source, which means a wood heater.

The wood heater

When necessary, a wood heater integrates the solar system and covers the small daily energy need, about 5,3 kWh/day. This small requirement can be satisfied burning around 1 kg of pellets o using directly the local olives trees pruning in a stove with a 85% efficiency.

In order to cover the remaining winter heating demand not cover by solar energy of 171 kWh, only 43 kg of pellets (that means 7,74 Euro for the whole period) or 40 kg of olives trees pruning is required.

Form and orientation

One of the first decisions a house builder has to make is the form and its orientation on the building site. While many of the analyzed projects are oriented exactly to the south, some deviate by as much as 45° off south. It appears that such deviation can be easily compensated by other design features in highly insulated houses, as can be seen in Fig. 7.

Building compactness is, however, important. The surface area to volume ratio (A/V) for all projects is 0.63. For the single family houses the average A/V is 0.73; for apartment buildings the A/V averages 0.49. There are two buildings with a relatively high A/V of 0.91 and 0.99, circled in Both of these buildings compensate their lack of compactness

by the heat gains of solar wall systems. The one house has the wooden Lucido-System: slotted wood acting as a solar buffer, between an exterior glass facade and the opaque insulated wall. The other house has a translucent solar storage wall of paraffin phase change material in a glass facade.

Equipment description

Basically CSE equipment with the Fresnel lens consist in an axis orientated to the Polar Sun, that is, parallel to the earth rotation axis. This axis is computer controlled to have a single daily rotation (Fig. 1).

The altitude of the sun is hand controlled because the equipment is simpler. Every morning we change this altitude and then the focal spot is on position.

Then, after we focalize the sun in the morning in the focal point the focal spot will be maintained at the same place the whole day because the axis makes 1 rev. per day following exactly the sun track. By computer control we can change the turning speed and register the temperature time heating and cooling curves of the sample.

Because of the angle of the axis and the weight of the Fresnel lens frame and other elements, the axis has a tie with a screw to adjust exactly the angle and to avoid oscillations when the wind is large or the weigh is heavy.

2.3. Fresnel lens

Fresnel lens

Quartz window

Horizontal hand

Horzontal axis hand




The Fresnel lens is ca. 0.8 sq. m and exist in the market at moderate price, ca. 200 €. Is made in polymeric material, 3 mm thick, and, consequently is very light in weight. The lens is covered while no working. A cover with a hole in the center let us to put the focal spot in the right place. To work the cover is put out and the heating of the sample start.





Grooves per inch



Transmission 400-1100 nm





0.125 +/- 40 %

30.0 +/- 5 %



92 %

353 K (80 C)

TABLE I. Fresnel lens characteristics (figures in inches)

controlled axis

Polar oriented axis computer controlled

Fig. 2. General view of equipment

The Fresnel lens is made by Edmund Optics, Ltd. Main characteristics are summarized in Table I. Fig. 2. show a general view of the equipment.

Typical design of the Fresnel lens consist in the substitution of the curve of the conventional lens with a series the concentric groove. This lens is molded in the surface of a thin, light weight plastic sheet with the appropriate curvature. The conical grooves act as individual refracting surfaces like tiny prisms when viewed in cross section bending parallels rays in a very close approximation to a common focal length.

One added advantage of the thinner plastic lens to the light weight is that because it is very thin there is very low absorbance of solar beams, i. e., it is a very high transmission. Fresnel lenses are a compromise between efficiency and image quality that depends of the groove density. High groove density increases image quality but low grooves density
produces higher efficiency that is our main interest. In infinite conjugated systems, as is in our case, the grooved side of the lens should face the longer conjugated.

2.4. Reaction chamber

Our installation has a double wall watere refrigerated reaction chamber that allow us to work in atmospheres of different composition. An inlet and outlet allows to introduce the gas of the required composition. To avoid the introduction of air we work with a slight over pressure of some cm of water.

The cylindrical chamber has a quartz window in one of the circular sides to allow the concentrated beams reach the sample inside the chamber. In the opposite side there is a system to introduce and withdraw the sample and for the thermocouples exit.

2.5. Samples

The samples we used to obtain SHS intermetallic coatings on steel consist in a small steel cylinder 6 mm thick, 14 mm diameter with an internal hole 12 mm in diameter.

ECOS a novel DEC concept

ECOS is a sorptive cooled heat exchanger employed for building air-conditioning. The novel system is the implementation of an original desiccant and evaporative cooling process. The new concept aims to overcome the thermodynamic limits of conventional DEC systems, allowing a higher energy efficiency and performance. The design of the process results in a higher dehumidification potential in comparison with conventional systems. It is particularly intended as a desiccant cooling system for climates with high ambient air humidity (e. g., Mediterranean and tropical). Moreover the novel system avoids the complexity of the rotating parts necessary in standard systems and gives the possibility to apply the DEC concept even at small scale plants.

Figure 3- Simplified scheme of the ECOS system

The process is based on simultaneous sorptive dehumidification and indirect evaporative cooling of the supply air stream. Moreover the indirect evaporative cooling is obtained through a continuous humidification process, ensuring a high heat exchange potential.

The system implementing the process is based on a counter-flow air-to-air heat exchanger technology. The heat exchanger is divided in sorptive (black line in Figure 3) and cooling (grey line) channels, which are physically separated but in thermal contact. The sorptive material is fixed on the heat exchanger sorptive channels. In the sorptive channels the supply air is dehumidified. In the cooling channels a continuous humidification of the cooling stream takes place. The latter, used for indirect evaporative cooling of the supply air stream, is for this purpose always kept in close-to — saturated conditions during the process.

The complete system consists of two sorptive heat exchangers, operated periodically. The periodic operation of two heat exchangers enables a quasi-continuous air-conditioning process. While one component is used in air-conditioning operation mode the other one is regenerated and pre-cooled before the next air-conditioning operation mode. In the regeneration the water vapour load of the sorbent material is released to the environment by means of a hot air stream (60-95°C). The air is heated through a heat exchanger connected to the solar plant. The subsequent pre-cooling phase is intended to lower the temperature of the heat exchanger after the regeneration, taking up the heat stored in the heat exchanger thermal mass. Moreover, a complete air-conditioning system will include at least a humidifier at the entrance of the cooling channel which brings the air at almost saturated conditions before entering the channel. Another optional humidifier can be installed on the supply air side in order to exploit the potential for direct evaporative cooling of the process.

Design tools

To assess the feasibility of a solar assisted air conditioning system, to find an appropriate system design or to investigate the advantages of selected technical solutions, computer based design tools have to be applied. Different types of design tools exist to meet the different requirements in the planning phase. Not in any case, a detailed simulation of a solar assisted air-conditioning system is required from the very beginning. On the other hand, there is no tool which meets all requirements at different levels of the planning phase. The range of existing tools may be sorted into different categories: easy-to-use pre­feasibility tools, allowing a fast calculation of basic key figures; simulation and sizing tools which also provide fast results but allow more distinction between different system configurations; and open simulation tools for advanced studies of a certain system configuration.

In this chapter, a non-exhaustive overview of a few computer tools is provided.

Best practice integration of solar heating in residential buildings

Fig. 3 Example of row houses in the Netherlands with 3 m2 solar collector per house. The rectangular collector are an esthetical improvement over the square version.(Project Boerenstreek, Soest, NL, [1])

There are a large number of examples to be found around the world of projects where special attention was given to the integration of solar heating products into buildings. Beside the technical and economical reasons there is clearly an increasing trend to consider the esthetical reasons for building integration of solar hot water systems. In this paper we give some examples of the different forms of and approaches to building integration in one family houses and in apartment buildings. The use of a small (3-4 m2) solar water heaters in a single family row house is rather common. However there are many projects with large solar roofs, where thermal solar energy is used for both domestic hot water and for space heating. This is more common in apartments with central heating system.

Fig. 4 Roof completely filled with solar collectors, Gneis-Moos, Austria [1]

Fig. 5 Student houses in Chemnitz (Germany), example of facade integration [1]

In the Northern European countries the use of solar heating systems in the facade is also rather common and useful because of the lower sun angle in the heating season. Depending on the latitude a different angle is needed. This gives the designer new ideas and possibilities for the integration of solar heating systems in the facade. For example, some systems are integrated as large awnings above the windows.

Fig. 6 Building integration in Shanghai, China [1]

Integration of solar heating systems in a

single family house is rather easy. Only a few square meters of the roof are needed for a domestic hot water system. Also in the case of space heating the roof will be large enough for the solar collectors. In the case of an apartment building the integration approach needs to be different. Depending on the amount of apartments and the floors in the building, the amount of square meters for solar hot water systems will exceed the available surface of the roof. In this case creative solutions are needed to integrate the collectors in the facade.