The electric heating elements heat up the top volume to 53°C during all hours.

The solar irradiance on the collectors and the daily hot water consumption are the same for both systems. An energy quantity of 1.53 kWh, corresponding to 33 l of hot water heated from 10°C to 50°C or 36 l of hot water heated from 10°C to 47°C is tapped from each system three times each day: 7 am, 12 am and 7 pm.

1.1.2 MEASURED RESULTS

Measured energy quantities for 6 weeks with a draw-off temperature of 50°C are given in Table 2 and measured energy quantities for 7 weeks with a draw-off temperature of 47°C are given in Table 3. The three way valve in the system with the two draw-off pipes ensures that a hot water temperature of 50°C is achieved for all draw-offs in the first test period and that a hot water temperature of 47°C is achieved for all draw-offs in the second test period. However, in periods with very high tank temperatures, the three way valve ensures that the draw-off temperature is as low as possible. In the standard system with one draw-off pipe, a hot water temperature of 52°C is achieved for all draw-offs in both test periods. The net utilized solar energy is the tapped energy from the solar tank minus the energy supply to the electric heating elements.

For the 6 weeks test period with a hot water draw-off temperature of 50°C the net utilized solar energy for the solar heating system with two draw-off levels is 8 kWh higher than the net utilized solar energy for the standard system. For the 7 weeks test period with a hot water draw-off temperature of 47°C the net utilized solar energy for the solar heating system with two draw-off levels is 18 kWh higher than the net utilized solar energy for the standard system.

The thermal performance of the systems is small, since the tests were carried out in the winter. The thermal advantage of two draw-off levels is higher for a draw-off temperature of 47°C than for a draw-off temperature of 50°C.

Yusuke Goto/ Dynax Corporation, Hiroaki Izumi/ Dynax Corporation,

Mamoru Aizawa/ Dynax Corporation, Hiroo Yugami/ Graduate School of Engineering, Tohoku University

It is important to utilize the solar energy efficiently1-2). However, conventional solar thermal collectors have some drawbacks of high cost, heavy weight and low temperature output. These drawbacks prevent from spreading of the market of solar thermal collectors, especially in Japan.

To find a solution to the problems, we propose an absorber / aerogel one-piece structure for high temperature solar thermal collectors. TiOxNy and silica aerogel were employed as a selectively solar-absorbing coating of absorber and as an optically transparent thermal insulator, respectively. Selectively solar-absorbing coatings must show lower reflectance and lower emittance. We have developed selectively solar-absorbing coatings of which optical property was a = 98.5, £ = 3 %. Wet coating of newly designed TiOxNy precursor enabled fabricating selectively solar-absorbing coatings, easily. Shrinkage of silica aerogel had to be minimized to produce an absorber / aerogel one-piece structure. Selectively solar-absorbing coatings and silica aerogel were prepared via sol-gel processing.

Experimental

Michael, Kohl, Volker Kubler, Markus Heck
Fraunhofer-Institut fur Solare Energiesysteme
Heidenhofstr. 2, D-79110 Freiburg
Tel.: 49 (0) 7 61-40 166-82, Fax: 49 (0) 7 61-40 166-81
E-Mail: michael. koehl@ise. fraunhofer. de

Flat-collectors are usually not tight. Therefore they can exchange the air with the environment. Moisture can accumulated in the collector, especially when the thermal insulation material could act as a storage for moisture (like mineral wool, for example). The moisture can increase the corrosivity of the micro-climate in the collector. An optimised ventilation rate which is measure of the air exchange between collector and environment helps to keep the collector dry.

1. Introduction

Sometimes solar collectors exhibit condensed water at the glazing (see figure 1). Usually even identical collectors at the same place in the same field show different degrees of condensation, caused by a unfavourable micro-climate. The micro-climate in ventilated flat — plate collectors is dominated on one hand by the water adsorption behaviour of the thermal insulation material and other components inside collector, e. g. wooden frames of back­planes. The other important property is the ventilation rate, which describes the air-exchange between the collector and the ambient.

Figure 1: Collector field with condensation effects

A high ventilation rate reduces the difference between the micro-climate and the ambient climate and keeps the collector dry, if it is rain-tight. But the thermal losses increase with the ventilation rate.

Simulation of the ventilation by computational fluid dynamic (CFD) showed that the air entering the collector through a relatively small ventilation hole in a lower corner of the collector is moving in several buoyancy driven loops towards the upper exit hole (figure 2a). There is much time for humidity exchange with humidity adsorbing components during this relatively long residence time.

Simulation of the condensation and re-evaporation of the humidity at the cold glazing showed that the drying by sunshine in the morning starts near the ventilation holes because of the dry air entering at the bottom and the warm air leaving at the top (figure 2b).

Figure 2: Computational fluid dynamics for ventilated flat-plate collectors:

a) Flow lines of air through a diagonally ventilated collector resulting from. The colours indicate the velocity of the test volumes (Blue means low velocity).

b) Effect of the ventilation on the drying of a glazing covered with moisture. The colours indicate the

load of humidity (blue-green means dried areas).________________________________________________________________________

Conventional solar systems are designed to maximise annual solar gains. The basic principle to achieve this is proper orientation and slope of collector field. In the central Europe conditions, maximum annual irradiation is received with a surface with south orientation and slope between 35° and 45°. In the case of facade collectors with slope 90°, the reduction in annual irradiation sum is around 70 %. Figure 2 shows annual profile of daily solar irradiation for roof (45°) and facade (90°) collector based on test reference year for Prague (TRY_Prague). Comparison shows a large difference between summer peak and cold season values for roof collector and relatively uniform profile for facade collector which corresponds closely to hot water demand profile (approx. constant with decrease in summer season). This feature allows the design of solar systems with a high solar fraction (above 50 %) without extremely increased periods of collector stagnation in summer as it appears in roof mounted systems with the same solar fraction.

Solar collector performance generally depends on optical and thermal losses which determine the efficiency of solar energy conversion in the collector. A detailed mathematical model KOLEKTOR was used for an investigation of solar collector thermal performance based on the knowledge of thermal processes in the individual parts of collector. Model consists of absorber outer energy balance (heat transfer through glazing, air gap, frame and absorber surface) and absorber inner energy balance (heat transfer within the absorber fins with solar radiation and piping). Absorber outer energy balance determines the temperature dependent overall U-value of the collector. Absorber inner energy balance obtains the collector performance factors dependent on the absorber material and geometry (F, FR). In the model, the temperature distribution in the collector is solved in an iterative loop from the input parameters. Input parameters are solar collector properties (dimension, physical properties of individual parts), climate data and operation parameters (input temperature, mass flow). Useful gain, efficiency of solar energy conversion and temperature of heat transfer fluid leaving the collector are the outputs from the model. The model was created in Excell sheet processor with use of Visual Basic programming.

The mathematical model was experimentally validated in the research of solar collectors with different covers (single, multiple, transparent insulations) and absorbers (non­selective, selective) [1]. For a specified set of operation conditions, a collector standard efficiency n can be determined in dependence on reduced temperature difference (Tm — Ta)/G. Model is feasible for sensitivity analysis in solar collector research. It has been used for an estimation of facade collector performance compared to roof collector. Further model description can be found in [1,2].

Facade integrated collector, when compared with collector located on the flat roof
(collector slope optimal 45°), shows considerably reduced heat transfer coefficients, especially for:

■ natural convection in the gap between the absorber and glazing

■ wind-related convection

■ back and edge frame heat loss coefficient

Due to vertical orientation of the air gap between absorber and glazing, heat transfer due natural convection is reduced in comparison with the gap under 45°slope to approximately 80 %. In Figure 3, Nusselt number in dependence on the slope of air layer according to different authors and the correlation obtained with statistical methods from the published experimental results is shown [1]. The correlation was used in the solar collector model. Since the air gap is a critical part in the single glazed selective collector, this reduction is reflected in overall collector heat loss coefficient. Standard solar collector efficiency curves determined for different slope angles (20°, 45°, 70°, 90°) are compared in Figure 4. Curves were calculated with use of mathematical model KOLEKTOR for standard weather conditions (ambient temperature Ta = 20 °C, incident solar radiation G = 800W/m2, wind velocity w =4 m/s). From the comparison, the slope impact on collector performance has been shown significant, especially for higher temperatures.

Fig. 3 Comparison of Nusselt number for Fig. 4 Solar collector efficiency curves for inclined air layer based on experiments by different slope angles

different authors

Calculation of wind-related forced convection heat transfer coefficients for solar collectors is not a distinct problem. There is a large number of models, which gives completely different transfer coefficient values in dependence on wind velocity. Some of them result from very detailed wind tunnel measurements, the other from measurements in real turbulent wind, but only for specified conditions and collector-building configuration. In solar engineering, McAdams’s [3] simple linear model

hw = 5.6 + 3.8 w

is regarded as reliable for usual heat transfer coefficient calculation. Sparrow et al. [4, 5] carried out a number of experiments to investigate the local heat transfer coefficients on the heated plates (e. g. solar collectors) under airflow at different conditions (angles of inclination of the plate relative to oncoming airstream, different velocities, framing surfaces). It was realised that average convection heat transfer coefficients are practically independent of the incident angle of airstream. From the airflow patterns on the plate, a considerable difference between the local heat transfer coefficients in the center of plate and on the edges was found. Higher velocity on the edge leads to higher local heat transfer coefficient, while the coefficients near the center are lower due the airflow stagnation. Consequently, the average convection heat transfer coefficients can be
substantially reduced when thermally active surface (solar collector) is framed by another thermally inactive surface (facade surface). Sparrow gives an equation to obtain the rate of reduction of the average heat transfer coefficient [5]

h/h* = (Lc/Lf)1/2 where h and h* respectively denote the coefficients in the presence and in the absence of the frame. The hydrodynamic dimensions Lc (collector) and Lf (framing surface) are determined as characteristic lengths from

L = 2L1La/(L1 + L>)

In the case of flat roof located solar collector and facade collector integrated in the building envelope, different values of wind-related heat transfer coefficients will be achieved. While at the glazing surface of roof collector, the average heat transfer coefficient corresponding to wind velocity is present, the average heat transfer coefficient at the surface of facade collector is lower due to framing effect (see Figure 5). For an investigated common block of flats case, the wind-related coefficient can be reduced to 60-80 % of the value for roof located collector.

Back and edge heat loss coefficient is reduced to minimum in dependence on thermal resistance of adjacent facade construction and "outer" collector frame temperature at value 20 °C (room temperature).

Synergetic impact of these individual heat transfer reductions is shown in the Figure 6 through the standard efficiency curves comparison for the roof and the facade collector with adjacent construction thermal resistance R =1,3 and 6 m2K/W. Facade integration brings qualitative improvement in solar energy conversion efficiency and better thermal performance especially for increased collector-ambient temperature differences.

Henderson D., Odeh N., Muneer T., Grassie T.

School of Engineering, Napier University, Edinburgh, EH10 5DT, UK

A photovoltaic-driven fan-duct system is being optimised as part of a solar air heater. The fan is used to draw warm air through roof slates into a duct system. The optimisation of this flow system is based on maximising the daily volume of air delivered for a given irradiance profile. This partially necessitates the accurate prediction of the speed and head-flow characteristic of the fan. This paper is concerned with analysing the performance of this solar electrical system composed of a PV module, a permanent-magnet brushless DC motor and an axial flow fan. The fan’s operating speed was predicted as a function of irradiance and module temperature. The resulting model predicts the speed of the fan to within 10 % of the measured values. The dependence of the head-flow characteristic of the fan on its speed is simplified by an easily manipulated relationship. The performance of two fans with different ratings (9.5 W with 69 l/s capacity and 20.3 W with 110 l/s capacity) was studied. The model predicts that the smaller fan delivers more volume of air than the 20.3 W fan. It is concluded that the start-up point of a motor/fan is an important factor in maximising the volume of air delivered.

1. Introduction

Roof slates are generally warmer than the ambient by a few degrees [1] and so mechanical ventilation systems can be designed to both ventilate the house underneath and harness this extra heat and supply it to the house by means of a small fan. This supplied heat can contribute to the heating demand thus reducing the need for auxiliary heating as well as reducing heating bills. A mechanical ventilation system has the disadvantage of requiring an auxiliary electric source to operate the fan in order to deliver warm air from the slates through the duct and into the house. This drawback of the system can be overcome by using a photovoltaic (PV) module to power the fan.

The use of a PV module has another advantage. The PV-driven system acts as a fast — response sensor to solar radiation and so the fan will deliver air only when the slates are also receiving radiation [2]. However, due to the comparatively high cost of PV modules, for such a system to be viable, its performance must be optimised. Optimum performance may be considered so that which delivers the greatest proportion of the incident energy to the house. The energy yield will be maximised when the daily volume of air delivered is maximised. A model, relating system flow rate to irradiation and PV module temperature has therefore been developed to determine the optimum configuration of system components to meet the above criterion.

For a given length and properties of delivery duct, the flow rate of air in the system depends on the head-flow (H-Q) characteristic of the fan. The head developed across the fan is a function of the flow rate and the speed of the motor of the fan. Furthermore, the speed of the motor is a function of irradiance and the PV module temperature. The current work is concerned with the estimation of the motor’s speed and the H-Q characteristic of the fan as a function of irradiance and module temperature.

In figure 11 the temperatures of the 1st extension are shown for 2003. In the summer months (June to September) the net return temperature varies between 55 and 60 °C due to the indirect preparation of domestic hot water by storage charging systems. In the summer months normally there is no heating system required. In 2003 the installation of the 2nd extension results in a disconnection of the heat store from 7/21/03 to 8/06/03. In the winter months with high heating demand the return flow temperature is slightly below 50 °C. Since the net return temperature is the lowest temperature in the whole system, the seasonal heat store can not be discharged to temperatures below 50 °C. This results in higher heat losses to the environment compared to the design net return temperatures of below 40 °C. The integration of a heat pump into the system would lead to an improved discharging of the store increasing the efficiency of all components due to lower heat losses to the environment.

— net supply temperature —- net return temperature

— preheating temperature —- ambient temperature

Figure 11: Ambient temperature and temperatures in the district heating net in 2003

Summary and outlook

Between 1997 and 2003 solar fractions from 21 to 30 % were reached. The specific solar heat gains vary from 330 to 400 kWh/(m2 a) resp. 176 to 241 kWh/(m2 a) (gross resp. net). One major reason for not reaching higher solar fractions are the net return temperatures of around 50 °C which are more than 10 K higher than expected value of less than 40 °C (yearly average weighted by volume flow). In addition, the heat losses of the heat store are higher than expected, mainly due to wet thermal insulation, and the connecting pipes with a length of 55 m. Apart from this no major problems occurred during the last seven operational years.

References

[1] M. Benner, M. Bodmann, D. Mangold, J. NuBbicker, S. Raab, Th. Schmidt, H. Seiwald: Solar unterstQtzte Nahwarmeversorgung mit und ohne Langzeit — Warmespeicher (Nov. 98 bis Jan. 03). Forschungsbericht zum BMWi-Vorhaben 0329606 S, ISBN 3-9805274-2-5

[2] TRNSYS Version 15.0 — User Manual. Solar Energy Laboratory, University of Wisconsin, Madison und Transsolar, Stuttgart.

[3] DF — Dynamic Fitting Version 2.7, InSitu Scientific Software, c/o W. Spirkl, Germering, Germany.

This project is being supported by the German Federal Ministry for the Environment, Nature Conservation and Nuclear Safety (Bundesministerium fur Umwelt, Naturschutz und Reaktorsicherheit), FKZ 0329607F. The authors gratefully acknowledge this support and carry the full responsibility for the content of this paper.

K. Schwarzer, C. Faber, T. Hartz, F. Spate Solar-I nstitut Julich / Fachhochschule Aachen Heinrich-MuBmann-Str. 5, D-52428 Julich Tel.: 0246/ 99 35-52, Fax: 02461 / 99 35-70 hartz@sij. fh-aachen. de, www. sij. fh-aachen. de

C. Petersdorff, J. Backes

ECOFYS GmbH, Eupener StraBe 59, D-50933 Koln Tel.: 02217/ 510 907-0, Fax.: 0221 / 510 907-49

c. petersdorff@ecofys. de, www. ecofys. de

1 Background Information

The State of North Rhine Westfalia has, in its project entitled "50 Solar Settlements in NRW” brought together leading innovations made in solar technology, so that the benefits of this research is made available to all. Within the framework of the AG Solar department at the Ministry for Science and Research in North Rhine Westfalia, the evaluation and scientific support for the first "solar settlement” ‘Gelsenkirchen Bismarck’ was established. The knowledge and expertise gained through this project will influence current and future solar settlement projects.

T. J.Sarvoththama Jothi,

Lecturer, School Of Mechanical Engineering, SASTRA Deemed University, Tirumalaisamudram, Thanjavur 613 402. India. Tel: 04362 264101 (office)

Fax: 04362 264120. Email: tisiothi@mech. sastra. edu. tjsjothi@hotmail. com

ABSTRACT

Enormous amount of energy is wasted in the form of heat for the purpose of cooking all around the world. Broad ranges of technologies are required around the world to incorporate the energy required for cooking. We have efficiently designed and developed a device named Line Concentrated Solar Cooker for the purpose of cooking and heating water or even pasteurization of drinking water. It is distinct from other type of cooker that is using the same old technologies. More over this device can be constructed by means of an inexpensive, commonly available material, thus providing a low-cost option suitable for household use in the developing world. This device was mainly designed from the input taken from the houses of four members each at various places. Its design and performance were evaluated at the laboratory including the efficiency tests. A model of such device was developed which gave the maximum efficiency of around 27 %. This Line Concentrated Solar Cooker has been mainly designed to prevent tracking mechanism, which is the main draw back for other concentrated type solar cooker. In order to prevent tracking mechanism, the design has been made in such a manner that the maximum sunrays are impinging on the reflecting surface of the Line Concentrated Solar Cooker all the time. Hence, minimum of at least 35 percent of the area of the Line Concentrated Solar Cooker is exposed to the sunlight at 8:00 AM and maximum of 100 percentage by noon and gradually decreases by evening as the sun sets. This model gave us a good results leading to excellent heating effect from morning to evening. Hence the heating effect gradually increased from morning to maximum at noon.

INTRODUCTION

We all know the fact that half of the world’s population is burning wood and dried dung in order to cook their food. It is learned that it causes more illnesses by breathing smoke through out the day, and there are more environmental impacts like deforestation, pollution, health hazardous etc., many of these billions of people live near the equator, where sunshine is abundant and free. One of the countries where most of the sunshine available abundantly is our own country INDIA, where many types of equipment (devices) related to solar studies can be developed and used with maximum efficiency. There are many devices related to solar that exists like solar parabolic cooker, solar box type cooker, solar water heater and many more. But what we tried to develop is Line Concentrated Line Concentrated Solar Cooker. The Line Concentrated Solar Cooker is safe, easy to make, inexpensive yet effective solar cooker in capturing the sun’s energy for cooking and pasteurizing water. It looks like a large, deep funnel, and incorporates the best features of the parabolic cooker and the box cooker. As the name indicates it looks like a cone-shaped reflecting funnel, sunlight would be concentrated along the axis of the funnel at the bottom, where the cooking pot or jar would be placed. Safety comes from the fact that the sunlight is concentrated along a line deep inside
the funnel. If you place your arm up the bottom of the funnel, warmth can be felt with our burning the hand as in the case of point concentrated solar cooker. This is much safer than concentrating sunlight to a point above a parabolic collector. The heat loss in Line Concentrated Solar Cooker is very less when compared to that of the box type solar cooker in which heat is gained only from one plane and loss occurs through the other planes. In Line Concentrated type solar cooker, the whole area of the funnel is used for capturing the heat radiation.

The purpose of the CFD FLUENT analysis was to determine the enclosure geometry most conducive to natural convection heat and mass transfer, given the limitations of a distillation enclosure. As the available geometric shapes are somewhat infinite, the processing time to examine each and every possibility was unrealistic. Rather than trying various geometric shapes randomly, a more logical approach was taken. The process is summed up as follows —

1. The basic geometric shapes that could conceivably meet the limitations and demands of a distillation enclosure were drawn and meshed. These included a rectangle, oval, circular cylinder, and trapezoid.

2. Each of these basic geometries was assigned reasonable boundary conditions representing the realistic operating conditions of a low temperature solar distillation enclosure.

3. The meshed models of these basic geometries were then solved for their steady state behavior in 2D.

4. The heat transfer between the hot-sink and cold-sink, due to natural convection, was determined in each case and compared with the heat transfer in the other geometries.

5. Various parameters such as aspect ratio, tilt angle, fillets, and other minute geometric considerations were varied in order to find the maximum amount of heat transfer possible for each basic geometry.

6. The basic geometries were then compared with respect to their maximum heat transfer to determine the optimum basic geometry.

7. Once the optimum geometry type was determined it was remodeled in 3D and its final volumetric geometry was determined.

The above iterative optimization process resulted in the geometry shown in Figure 2. It can be best described as an “ovalized rectangle” as it benefits from the "best” of both of these two geometries. A rectangle in general had high values of heat transfer when compared to the other basic geometries, this agrees well with other studies done [4]. In the case of a simple rectangle however, with a horizontal top and bottom, the heat transfer rate was slightly reduced due to the severe deflection of flow away from the vertical walls. The dome shaped top cover allows the convective current to move more smoothly over the curved surface, which Fl9ure 2 — Ovalized rectangle enhances the heat transfer from the surface. These results are in agreement with a similar study done which also examined the natural convection in dome shaped enclosures [5].

Further examining and modifying led two additional improvements. Inserting a partition between the two heat transfer surfaces yields further improvement of ~ 5-10% in heat transfer. This is due to forcing the air to enter the hot-sink from its coldest point and not allowing it to "cut across” the center. The same is true for the hot air on top. This way large temperature gradients for driving heat transfer are kept. An additional benefit of the partition is that it prevents ‘cross mixing’ between the saline solution and the condensing (and dripping) distillate. The tilt angle at which the enclosure is at with respect to gravity also has an influence on the natural convection. Certain tilt angles were shown to improve the overall heat transfer due to natural convection.

The partitioned ovalized rectangle was determined to be the optimal geometry for a distillation enclosure. It was remodeled in 3D (Figure 3) and solved iteratively, so as to determine the exact volumetric parameters that would provide the best heat transfer by natural convection through the enclosure.

In this examination three factors were tested — separation distance between heat transfer surfaces, enclosure tilt angle (0), and aspect ratio (H/L). The following graphs (Figure 4 — 6) illustrate the results.

Figure 5 — Effect of tilt angle Figure 6 — Effect of aspect on heat flux ratio on heat flux

The results described previously show that by optimizing the geometry in 3D we can improve the heat transfer by natural convection. As for the separation distance and aspect ratio there is a significant influence on heat transfer while the influence of the tilt angle seems to be less. When applying the results we can propose an enclosure geometry that is optimal for heat and mass transfer, as shown in Figure 7.

3.40Є+02

3.24e+02

Figure 7 — Contours of static temperature for the proposed distillation enclosure

Experimental Device

Initially, the enclosure with optimized geometry was constructed and insulated. To prevent corrosion the material primarily used in the enclosure was stainless steel with the few features that demanded more flexibility made from plexi-glass. Once the enclosure "shell” was complete the evaporator and condenser sections were assembled into it.

The evaporator is made of 15 individual evaporation sheets, each fed hot brine from thin, holed, feed pipes each connected to the incoming main pipe. The best material for the evaporation sheets, after experimentation, was found to be a tight-weave cotton cloth. Each individual cloth was "tunneled” at the top to improve the uniform spread of the flow as it leaves the feed pipes onto the cloths. The bottoms of the cloths are V-shaped so as to centralize the down flowing brine into one "drip line” at the bottom (refer to Figure 8).

The condenser, on the opposing side, is a tube bank made up of seven, parallel, 9 mm copper tube sections. Each copper tubing section is fed from the seawater main entering at the bottom of the enclosure and feeds into the central pipe supplying the solar collectors at the top of the enclosure. The enclosure volume is divided into half by a partition made of plexi-glass which prevents the humid air flows from prematurely mixing and which also keeps the distillate from being lost or contaminated. The bottom of the enclosure is divided into two by a dividing wall that completely separates the two pools of fluid — brine and distillate — with each pool draining out of its respective pipe. Figure 8 shows the open evaporator / condenser.

Bottom of evaporation cloths

Results

The salt content of the brine feed water (Ms) was made to be approximately 23,000 ppm by adding table salt. The salt content of the distillate (Md) produced by the device was found to be ~ 16.7 ppm. These results were measured on the feed and distillate samples by titration, and were corroborated by an electric conductance measurement.

The device was tested while operating at various feed water flow rates, enclosure tilt angles, and temperatures. Following each experiment the optimal efficiency (GORopt — equation (3)) and the ideal efficiency (GORi — equation (1)) were calculated to serve as indicators to the device’s actual performance. Results for the actual efficiency (GORa) and
the ideal efficiency (GOR), as a function of the optimal (GORopt) (both calculated at the same operating temperatures), are shown in Figure 9.

Stated differently, Figure 9 represents a comparison between three efficiencies for each

Figure 9 — Actual efficiencies compared to the optimal and ideal calculated efficiencies

experiment. The uppermost line, GOR, is higher than the other two as it corresponds to an efficiency representing the ideal situation — an ideal heat engine and heat pump. The lower line, GORopt represents the efficiency that is theoretically possible in an optimal but realistic regeneration device, and the scattered data GORa represents the actual measured efficiencies. A definite correlation can be seen between the actual measured GORa in the laboratory device and the GOR attainable in an optimal device operating under the same temperature limitations. The device’s efficiency was also measured as a function of the tilt angles starting at 0° degrees through 25° degrees (at 5° degree intervals). This test sequence was conducted twice, the results are shown in Figure 10.

Discussion

The dissolved salt content in the distillate of 16.7 ppm shows the device’s ability to fulfill its basic requirement — desalinate water. When considering the actual thermal efficiency of the laboratory device one can see that it corresponds relatively well with the optimal efficiencies possible at the same temperature values (see Figure 9). In many cases however, the actual efficiency exceeds that of the optimal efficiency — an "impossibility”. By examining this contradiction statistically we see that more than likely the inherent error in thermocouple measurement is responsible for the "overshooting” of GORa. However, even when taking the worst possible error in measurement and subtracting it from the least-squares average, the data representing GORa still corresponds well to the optimum curve (GOR0pt). On one hand the experimental data does seem to have an inherent amount of error due to measurements, but on the other hand even when taking this error into account the device is still operating at very near the best possible efficiency.

From figure 10 it is clear that there does seem to be some influence of the tilt angle on the overall device efficiency. As the errors in measurement are on the order of the experienced improvement, however, it is difficult to state a definite conclusion as to the exact contribution of the tilt angle. The tilt angle that seems to be most conducive to natural convection is somewhere in the vicinity of 15° degrees which is less than predicted by Fluent (15° vs. 35°). In any case, further investigation is necessary to determine the precise effect the enclosure tilt angle has on the overall efficiency.

The lag time the device takes to reach steady state conditions (approx. 8 hours) is a detriment. In order for the device to operate at its optimum the fluctuations in temperature need to be prevented. In addition to this, as the device is completely insulated from the environment, it is not dependant on outside temperatures and may take advantage of 24- hour a day operation. For these two reasons it seems that the only way to effectively operate an efficient regenerative solar distillation device is to have it supplied hot water from an intermediate (storage) reservoir.

Conclusion

This research project, in contrast to other research [6] investigating low temperature distillation devices employing regeneration, incorporated a CFD analysis that formed the base for the optimization process. By using such an analysis the optimum geometry for the distillation enclosure was determined. Along with the basic geometry, other features such as aspect ratio, tilt angle, and partitions that also contribute to the overall thermal efficiency, could be tested for their contribution and incorporated into the design. As a CFD analysis was incorporated, not only could the distillation enclosure be optimized, but the behavior of the natural convection flow field could be predicted with relative ease.

The product of the CFD analysis was an optimized low temperature distillation device. The device was built, tested experimentally to verify its functionality, and finally tested for its thermal efficiency. The following points conclude the performance of the device.

— The proposed optimized design is functional and is capable of producing distilled water of high quality (dissolved salt content of distillate ~ 17 ppm).

— The actual operating efficiency of the proposed design compares well with the efficiency of an optimal device, thus the design may be considered optimized.

— The device built is durable and simple to operate — two additional benefits that add to the overall effectiveness of the device.

This research has shown that by incorporating a CFD analysis the originally complex natural convection flow found in distillation enclosures could be investigated. This way a low temperature distillation device incorporating regeneration was optimized and designed.

A three dimensional steady state simulation has been carried out in order to study the riser tubes of a plastic absorber configured by two manifolds and a set of riser tubes without fins.

The physic domain of the studied case is shown in figure 3(a). It consists of a square duct with thin walls (e = 0.4mm), length of 1 m ( ) and different side length (I =

4.6.8.10mm).

The temperature of the water and the air surrounding are assumed constant (Ta = Tin = 25°C), and a mass flow of fn = 501/hrn2 is imposed. In order to simulate solar irradiance and convection losses to the environment, at the top wall (wall facing the cover) a constant heat ^^^Яф2) is imposed and an equivalent convection coefficient ( ) is

Figure 4: Numerical results for riser tube: Heat removal factor (FR) and pressure drop (Ap) along channel vs. side length (I = 4,6.8 and 10 mm).

’ЛіШУШРЛ

2® 29 30 31 32 33 3+ 35 35 36 37 38 38 39 40 41 TiYl

Figure 5: Numerical results riser. Riser tube temperature profiles for the four set of dimen­sions at different sections in у-direction at x-midplane. ,Y* = y/L.

assumed. At the bottom wall, the coefficient imposed is Ub = 0.5W/rn2K and at the outlet a pressure outflow boundary condition is imposed [12].

The computational domain where the governing equations are solved is shown in figure 3(b). The number of control volumes used in each direction is represented by the n parame­ter, where n = 40. Near the solid walls the grid has been intensified by means of a tanh-like function[17], it is indicated by solid triangles in the figure.

As from the CFD simulation local values of all the variables (temperature and velocity field) are know, is possible to evaluate a posteriori any integral quantity of interest. As an example, the values of the heat removal factor Fr and of the pressure drop Ap through the riser in terms of the side of the duct have been computed. They are given in figure 4.

Further information of the numerical results are shown in figure 5. It consist of a compar­ison of the temperature profiles in the central section of the riser tube at different values of y (0.25, 0.50, 0.75 and 1.0 m).