The results of the evaluations of the roof MaReCo are shown in figure 6. These values are thus the angular dependence of the reflector exclusively, since the influence from the Teflon and the glazing are deducted, as earlier described. The measurements show good agreement for both methods for low angles of incidence. For high angle of incidence, when the irradiance is low, the cosine dependence shows better agreement with outdoor measurements than diode compensated data.

An equal relation between the indoor and outdoor measurements of the solar hybrid window was found for the roof MaReCo. This result gives another indication of that the
method is working satisfactory and that the simulator can be used for characterisation of concentrating collectors. However, at higher incidence angles, the agreement is not as good as for lower angles, suggesting that the method still needs further improvements.

Discussion

The results of incidence angular dependence from outdoor and indoor measurements show that the agreement is good at lower incidence angles. At higher incidence angles, the correspondence is less accurate. There are possibilities to improve the method further. By adding more photodiodes to the array, the accuracy of the measured total irradiance would be better. Improvements of the reflectors of the solar simulator, to obtain a more even light field, are also possible. An explanation for the lower correspondence at higher incidence angles could be that the active test area is smaller and that the power involved is lower.

One of the most critical aspects in the design of thermosyphon solar heater is the design of the absorber. It must be a very good heat exchanger, the flow must be equally distributed along all the tubes and the pressure drop of the thermal fluid due to flow resistance in the absorber must be minimised in order to enhance the thermosyphon effect. With the simple model, design and optimisation studies of the absorber cannot be carried out. On the other hand, the use of high level codes, CFD codes, for the modelling of the complete absorber are not still suitable due to the huge amount of computational resources needed to run a single test. In this section, a code for the modelling of the fluid flow in heat exchangers is explained. This code is being developed in the “Centre Tecnologic de Transferencia de Calor” of the Polytechnical University of Catalonia. It was mainly focused to heat exchangers for automotive applications. Currently it is being extended to be used in order applications such as the absorbers of the solar collector.

The heat exchanger is divided in n different CV in which the conservation laws of mass, momentum and energy are solved. Heat transfer coefficients from the walls of the heat exchanger to the thermal fluid, and pressure drop due two fluid flow resistance due to ex­pansions, contractions, elbows, wyes, friction in the walls.. are evaluated from correlations available in the literature, [10, 8, 13].

A unidimensional adaptation of the SIMPLE (Patankar [16]) methods has been imple­mented to solve the velocity-pressure coupling of the resulting algebraic equations.

This code is currently being used by the authors in order to investigate the fluid flow in different absorber configurations and to obtain the fluid flow parameters of the absorber which are required for the simple model, i. e. the pressure drop through the absorber in terms of the mass flow rate of the thermal fluid.

CFD model

This method represent the highest level of numerical simulation and it provides the values of all the relevant variables such as temperature, pressure and velocity, throughout the do­main of interest by means of the application of any discretization method. In this technique a continue variable in space and time is approached using a set of discreet values over all the studied domain. The method based on finite volumes is commonly employed in CFD and is the one used in this work.

The heat and mass transfer governing equations are continuity, momentum and energy. They constitute a set of non-linear, PDE’s and strongly coupled equations. They can be written in vectorial form, as follows:

V-v = 0 (4)

(5)

pcv (v-V)T = kV2T + V • qT (6)

where the Boussinesq equation of state, dp = —fipdT, has been applied. Here, T is the temperature, t the time, v the velocity, p is the dynamic pressure, Tre/ is the reference temper­ature, g is the gravity and p, f. t, A, and C), are respectively: density, viscosity, conductivity, thermal expansion coefficient and specific heat.

The last term in equation 6 is the energy generation per unit volume due to the imbalance between the absorbed and the emitted radiation. Radiation effects are taken into account via the radiative transfer equation (RTE) which is used to determine the intensity radiation field, I. The RTE considers the variation of the energy carried by a beam in direction due to its interaction with the medium where it is travelling on [5]. Then, the radiative heat
transfer equation is an integro-differential equation that properly describes such variation. For a given wavelength range in a purely absorbing (non scattering) medium, it reduces to

s-VI{r, s) = — k{I{r, s)-IB{r)) (7)

which is a conservation optical balance equation. IB(f) stands for the thermal emission of a black body at a temperature of T(r) for the given wavelength range, and & is the absorption coefficient. Once the intensity radiation field is known, the energy generation or loss due to radiation is calculated as:

рос p

(8)

Jo J 4?r

In general terms, the principal problem in the solution of Navier-Stokes equations is the handling of the coupling between momentum and continuity equations and the fact that it does not exist a specific equation for pressure. Then, the complexity for evaluating RTE equation adds an extra difficulty to the problem.

The method employed to couple pressure-velocity field is one of the SIMPLE(Semi-Implicit Method for Pressure-Linked Equations)[16] family. This resolves in a segregated way mo­mentum equation (5), with a guessed pressure. Then, velocities and pressure are corrected fulfilling continuity equation (4). So it is necessary to derive a pressure-correction equation using equations (4) and (5) [16].

If participant medium to radiation is considered, RTE equation has to be solved using some technique like Discrete Ordinates Method (DOM) [5]. The object of the DOM is to solve equation (7) for a chosen set of N ordinates in such manner that any integral magnitude in relation to the intensity radiation field may be changed by a weighted summation of the integrand at selected points (ordinates) of the integration domain.

Once momentum equations (5) are solved the procedure to resolve energy equation (6) is as follows. Firstly, a temperature field is supposed. Then, the RTE is solved using DOM, assuming that the temperature is not modified by radiation field. Thus, the energy generation or loss per unit volume is calculated using equation (8). The next step is to solve equation (6), given as an additional source term, that varies on each iteration due to it depends on the temperature field. With the new thermal field, the RTE is solved again until convergence is accomplished.

Due to the elliptic nature of the governing equations with respect to the spatial coor­dinates, boundary conditions have to be specified at all external boundaries to take into account the interactions with the environment.

In order to assure that the numerical solution are an appropriate approximation of the mathematical model, the authors use a post processing tool based on the generalised Richardson extrapolation for h-refinement studies [2]. From this method, estimations of the numerical uncertainty of the solutions are calculated.

Screening testing is thereafter conducted with the purpose of qualitatively assessing the importance of the different degradation mechanisms and degradation factors identified in the initial risk analysis of potential life-limiting processes.

When selecting the most suitable test methods for screening testing, it is important to se­lect those with test conditions representing the most critical combination of degradation factors.

Using artificially aged samples from the screening testing, changes in the key functional properties or the selected degradation indicators are analysed with respect to associated material changes. This is made in order to identify the predominant degradation mecha­nisms of the materials in the component. When the predominant degradation mechanisms have been identified also the predominant degradation factors and the critical service con­ditions determining the service life will be known.

Screening testing and analysis of material change associated with deterioration in per­formance during ageing should therefore be performed in parallel. Suitable techniques for analysis of material changes due to ageing may vary considerably.

On the static solar materials of Task 27, a number of accelerated screening have been performed including simulation of possible degradation in performance under the influence of high temperature, high humidity/condensation, UV, and corrosion loads; either single or combined loads; see Table 5.

In Figure 3 the results from a series of screening tests on pure aluminium, used as refer­ence reflector material, are shown as an example of result from the Task 27 study. Degra­dation in optical performance is observed mainly, as expected, in the corrosion tests. In Figure 4 the result from the testing of a number of antireflective glazing materials at 80 °C and 95 %RH is given. The cause of degradation in optical performance is in this case not understood and the degradation therefore needs to further analysed. To identify degrada­tion mechanisms for the tested materials various analytical techniques are presently em­ployed.

Figure 3 Results from screening tests on pure Aluminium in the IEA Task 27 study

Figure 4 Result of the most significant screening test on antireflective glazing glazing ma­terials involving exposure at 80 C and 95 %RH in the IEA Task 27 study

Martin Meier, Werner Korner, Andreas Beck, Jochen Fricke

Bavarian Center for Applied EnergyResearch (ZAE Bayern) Thermal Insulation and Heat Transfer Division Wurzburg, Germany

Modern solar glazing units normally make use of low-e coatings, i. e. coatings with emissivities less than 0.05 in the IR spectral range. These coatings drastically reduce radiative heat transfer between the inner and the outer pane. Visible transmittance is hardly affected by these coatings. Unfortunately, due to their IR low-e characteristics these low-e coatings decrease the solar transmittance of the glazing, which can be a hindrance in solar applications. This paper demonstrates how the solar transmittance of solar glazing can be considerably improved by using micro-structured low-e-coatings, while the IR-emissivity of the system is not significantly increased.

SolarTransmittance of Low Emissivity Coatings

Low-e coatings mainly consist of one or two thin metal layers (usually silver) that determine the low-e behavior, and several additional layers for anti-oxidant, anti-reflex and adhesion-improving purposes.

As the main application of these coatings is in architectural glass, the coatings are designed in such a way that visible transmittance is hardly affected. Unfortunately, due to the decrease in transmission above 700 nm, these low-e coatings reduce the solar transmittance ofthe glazing, which can be a hindrance in solar applications.

For solar architecture, transparent insulations (Tl) and solar collectors, the solar and not the visible transmittance is the important factor. An ideal “solar low-e coating” for these applications would be extremely wavelength selective: a coating that reflects in the spectral range of 300 К heat radiation but is perfectly transparent in the range of solar radiation (see Figure 1).

Figure 1:

Reflectance of an ideal solar low-e coating. The coating is transparent in the spectral range ofsolar radiation (below a wavelength of2.5 pm), but reflects in the spectral range of heat radiation at a temperature of 300 K (above a wavelength of2.5 pm).

Considerations show that the solar transmittance can be increased by about 18 to 20 percent using this ideal solar low-e coating compared to a conventional coating. This value is calculated from the difference in transmittance of an uncoated pane of float glass and of a low-e-coated pane, each averaged over the solar spectrum [1] (see Figure 2).

A silicon photovoltaic module (30x10cm) was used to charge 12Volts battery needed to drive the N-S tracking axis DC motor, control components, and a DC closed loop water pump (in case of forced convection). The PV module was fixed on the tracking frame to improve PV cell efficiency by minimizing solar incident angle. Having this type of power supply the system can operate independently and remotely. The selection of PV and battery sizes was based on the power required by the DC electric motor. The required motor torque was estimated experimentally and found equal to (100 N. m).

Testing platform

The testing platform of the parabolic collector is shown in Fig. 7. The platform allows the flexibility of different outdoor tests as well as different tracking axis inclinations. The system working fluid (water in this study) inside a closed loop pipe delivers heat to the storage tank via a tube coil heat exchanger. The storage tank height is adjustable to allow the study of natural circulation effect of the working fluid.

A DC fluid pump is also connected to have forced circulation of the fluid.

Conclusion

Parabolic-trough solar water heating is a well proven solar energy technology which is being used on a commercial scale to produce heat for industrial and residential applications. In this work, a single tracking axis parabolic trough collector was designed and constructed for moderate heat load processes. The solar tracking collector was designed to be a self powered system so it can operate remotely and independently under moderate radiation levels. Simplicity in manufacturing and operation was considered in the design of this collector such as using stainless steel as a reflecting surface and a closed loop control system for the single solar tracking axis.

References

[1] Agenda 21, Preliminary Report Prepared by Several Organizations, 2000, Amman.

[2] May K., Barker G., Hancock E., Walker A., Dominick J., and Westly B., Performance of a large parabolic trough solar water heating system at Phoenix federal correctional institution, J. of Solar Energy Engineering, 122, 2000, pp.165-169.

[3] Kruger D., Hennecke K., Schwarzbozl P., Lokurlu A., and Richarts R, Parabolic trough collectors for cooling and heat supply of a hotel in Turkey, World Renewable Energy Congress, 2002, VII, Cologne — Germany.

[4] Grass C., Schoelkopf W., Staudacher L., and Hacker Z., Comparison of the optics of non-tracking and novel types of tracking solar thermal collectors for process heat applications up to 300 °C, Solar Energy, 76, 2004, p207-216.

[5] Almanza R., Lentz A., and Jimenez G., Receiver behavior in direct steam generation with parabolic troughs, Solar Energy, 61, 1997,pp. 275-278

[6] Odeh S., Morrison G. & Behnia M., Modeling of parabolic trough direct steam generation solar collectors, Solar Energy, 62, 1998, pp. 395-406.

[7] Odeh S., Unified model of solar thermal electric generation systems, Renewable Energy, 28, 2003, pp.755-767.

[8] Odeh S., Evaluation of solar thermal electric generation system in Jordan, Training Work Shop on Renewable Energy Technologies for a Clean Environment and Sustainable Development, 2002, Amman-Jordan.

[9] Odeh S., Modeling of solar industrial process heat system with direct steam generation parabolic collector, sent to Int. Journal of Heat and Mass Transfer, 2004.

[10] Abdallah S. and Nijmeh S., Design construction and operation of One Axis Sun Tracking Sytem With PLC Control, Jordan Journal of Applied Science, 2002, 4, 2, pp.45-53.

[11] Abdallah S. and Nijmeh S., Two-axis sun tracking with PLC control, Accepted in Energy Conversion and Management, 2004.

[12] J. Duffie, W. Beckman, Solar Engineering of Thermal Processes, 2nd ed. J. Wily and Sons, 1991.

R. Consul, I. Rodriguez, K. Claramunt and A. Oliva

Centre TecnOlogic de Transferencia de Calor (CTTC)
Lab. de Termotecnia i Energetica
Universitat Politecnica de Catalunya (UPC)
labtie@labtie. mmt. upc. es, www. cttc. upc. edu

This work presents the numerical studies performed by means of multidimensional Complutational Fluid Dynamics (CFD) simulations to analyse the thermal behaviour of solar water storage tanks operated by rigid distribution manifolds. Advantatges and limitations of the use of these kinds of diffusers to improve thermal stratification in storage tanks are discussed. A test procedure to characterise the level of mix­ing produced by natural convection when fluid entering the tank is colder than the surrounding fluid is proposed.

Introduction

The importance of developing improved storage devices is a key aspect of the thermal optimisation of many energy systems that are characterised by the delay between energy production and consumption. One of the most illustrative cases are thermal solar energy systems, where a good performance of the storage devices means a considerable increase in the overall efficiency of the installation.

Within the wide range of storage equipment, sensible heat storage tanks of liquid water are routinely used in standard thermal solar systems (i. e. solar domestic hot water and heating). The good thermal properties of water and its relatively low cost make the water storage tank the most traditional and extended storing equipment.

Even though the configuration of storage tanks is apparently simple, the extreme weak­ness of the level of temperature stratification (one of the most desirable properties in these equipments) make optimised designs very difficult. The store performance is dominated by mixing and diffusion processes over the interface between hot and cold fluid layers, but one of the major causes of loss of stratification in storage tanks is produced by the inlet mass flow rates (from collector and load loops) due to the enhancement of mixing and diffusion phenomena.

In order to improve the level of temperature stratification inside the storage devices, dif­ferent constructive possibilities have been explored in the literature such as the study of the position of the inlet/outlet ports, different kinds of materials, diffusers [1,2, 3, 4, 5], etc. Some of these improvements are being implemented in commercial products. Special interest has motivated the design and optimization of distribution manifolds [5]. These kinds of diffusers are used to reduce the mixing due to natural convection that occurs when the fluid entering the storage tank from the collector loop is colder than the surrounding fluid. Their principle of operation is based on the reduction of the momentum of water entering the tank, allowing buoyancy forces to direct the fluid to the location at which the resident store fluid temperature most closely matches the inlet fluid temperature.

In this work, and by means of multi-dimensional transient numerical simulations of heat transfer and fluid flow, the thermal performance of such internal elements will be virtually analysed. Special attention will be given to the advantages and limitations of available com­mercial designs, discussing possible improvements and/or alternatives to be considered.

Numerical modeling of thermal storage tanks is often used technique to investigate thermal and mass processes. Although two or three-dimensional analysis are possible to describe flow and temperature distribution in storage tank, they are not applicable to simulation calculations of the long-term performance due to insuperable difficulties of calculation algorithms. Hence, one-dimensional modeling is a possible alternative, because of its simplicity and sufficient accuracy of computational procedures.

Simple one-dimensional numerical model has been developed for predicting the transient behavior of the vertical temperature distribution in the tank. The model describes temperature changing in different layers of the tank by means of momentary energy balance for defined quantity of water.

The stratified accumulator considered in this study is divided into m sectors (corresponding to the number of thermo sensors) with the equal volume, as depicted schematically in Fig. 1. The sectors are numbered from the bottom to the top of tank. Different sectors contain different parts of heater area (serpentine). This means that in sectors work different heat sources (heat exchange area).

Hot water is consumed from the upper sector of the tank. Consumed water quantity is compensated by injection cold water at the bottom sector. This water is assumed to mix

with the water in the sector. Some quantity of mm:,

water from bottom (1) sector enters the next upper sector (2) and mixes with the water in the sector. This process occurs in all next sectors of the tank. The temperature change in sectors by discharging process (water consumption) can be written by:

(1)

where i, n are sector and time step number; V is sector volume and T — temperature of water.

AV is quantity (volume) of consumed water for time step n. For the bottom sector (1) the temperature Ti. i,n. i is the net supply water temperature Tnet. Discharging process is simulated by consecutive passing across the sectors from the bottom to the top.

In the same time, another process is taking place — the thermal charging process (hot water accumulation). The heat from solar collectors is transferred to the water in the tank by serpentine elements. This causes temperature rise in tank. Temperature rise depends on outlet temperature from solar collectors and flow rate of the working fluid. The charging process can be considered as independent (superposition principle). Hence, a second passing across the sectors for the same time step is needed to determine the temperature rising. Energy balance in sectors gives the temperature change:

TOC o "1-5" h z K F.

T = t + ^ ‘jul(t — t )Vr, (2)

i, av i, ac

where T, n’ is the temperature in /-sector of the accumulator at n-time step, after the discharging process has passed; Tf — average fluid temperature in /-serpentine element; At — time step interval for charging process; K, serp and F, serp — heat transfer coefficient and
heat exchange area of serpentine element for /-sector; p and cp — density and specific heat capacitance of water in tank.

Overall heat transfer coefficient K/ser for serpentine element includes convective coefficients hf and hfree, corresponding to the forced circulation in serpentine pipe and free convection from external surface and conductive transfer parameters for serpentine wall:

free

where 5s and As are serpentine wall thickness and conductivity coefficient of the serpentine material.

Convective coefficients hf and hfree depend on fluid temperatures, which are unknown values in the beginning of the calculations. Known parameters for calculation start are the inlet fluid temperature for serpentine (outlet collector temperature) and water temperature in accumulator (temperature distribution in accumulator). Initial temperature distribution in accumulator must be adopted in the beginning of the calculations (initial conditions). This predestines the calculation consequence — from the top to the bottom of accumulator because the inlet of serpentine is in top region of accumulator. Calculation begins for the top accumulator sector with the known fluid temperature in entrance of serpentine element. Iteration procedure for transfer coefficient K/:Serp is adopted.

Inlet fluid temperature T/:/n in /-sector is known — the outlet temperature from the previous (upper) serpentine element (sector /+1). It stays constant in iteration process. For the top sector inlet fluid temperature in serpentine element is determined by solar collector’s performance. The outlet temperature T/:OUt depends on transferred heat energy in sector and is determined in last iteration step.

Mathematical model for solar collectors is well-established matter and detailed information can be fond in solar energy publications. Outlet temperature of working fluid for solar collectors is defined by next equaton [1]:

F

= — [9,(*-■«)e -UL(TsolM — Ta) (4)

m ■ cp

where Fr is heat removal factor, m — mass flow of working fluid, (ra)e — effective transmittance absorbing coefficient for optical part of solar collectors, qs — solar radiation flux for tilted surface [W/m2], UL — overall collector heat loss coefficient [W/m2 K], Ta — ambient temperature. Inlet temperature for solar collectors Tcol, in is the outlet temperature from the bottom serpentine element. Because of dependence between inlet temperature of working fluid for serpentine and inlet temperature for solar collectors (outlet temperature for serpentine), new iteration calculations are needed.

Special simulation algorithm binds the collector and accumulator models in a working unit. It takes into account the heat losses in pipes and accumulator. Computer program is created to manage the theoretical calculations.

Y. Tripanagnostopoulos1, M. Souliotis1, R. Battisti2 and A. Corrado2

1Physics Department, University of Patras, Patra 26500, Greece
Tel/Fax: +30 2610 997472, e-mail:1 yiantrip@physics. upatras. gr,

2Dept of Mechanics and Aeronautics, Univ. of Rome “La Sapienza”, Rome 00184, Italy
Tel:+39 06 44585271, Fax:+39 06 4881759, e-mail: 2riccardo. battisti@uniroma1.it

Hybrid PV/T systems with air heat extraction are an alternative and cost effective solution to building integrated PV systems, because of their easier construction and operation. These systems are usually consisted of PV modules with air channel at their rear surface, where ambient air is circulating in the channel for PV cooling and the extracted heat can be used for building thermal needs. To increase the system thermal efficiency, an additional glazing is necessary, but it has as result the decrease of the PV module electrical output from the additional optical losses of the solar radiation. An extensive study on air cooled PV/T solar systems has been conducted at the University of Patras, where hybrid PVT/AIR prototypes have been experimentally studied in their standard form and also with a low cost modification. The methodology of Life Cycle Assessment (LCA) has been used to do an energetic and environmental assessment of the heat recovery system by the University of Rome “La Sapienza”, implementing a specific software for LCA, SimaPro 5.1. In this paper we provide electrical and thermal energy output results for PV and PVT/AIR systems, analyzing them with respect to their performance improvements and environmental impact, considering their construction and operation requirements.

PVT/AIR system concept

The temperature of PV modules increases due to the absorbed solar radiation that is not converted into electricity causing a decrease in their efficiency. This undesirable effect can be partially avoided by applying a heat recovery unit with a fluid circulation. In hybrid Photovoltaic/Thermal (PV/T) solar systems the reduction of PV module temperature can be combined with a useful fluid heating. Hybrid PV/T systems can simultaneously provide electrical and thermal energy, thus achieving a higher energy conversion rate of the absorbed solar radiation. These systems consist of PV modules coupled to heat extraction devices, in which air or water of lower temperature than that of PV modules is heated whilst at the same time the PV module temperature is reduced.

In PV/T system application electricity is of priority and therefore the operation of the PV modules at low temperatures keeps cell electrical efficiency at a sufficient level. This demand limits the effective operation range of PV/T system thermal unit in low temperatures and the extracted heat can be mainly used for low temperature thermal needs (space heating and natural ventilation of buildings, air or water preheating, etc).

Hybrid PV/T systems with air heat extraction (for simplicity PVT/AIR) are more extensively studied, mainly as an alternative and cost effective solution to building integrated PV systems, because of their easier construction and operation. In typical BIPV applications the increase of PV module temperature results to the increase of undesirable heat transfer to the building, mainly during summer. Air cooled hybrid PV/T systems are usually consisted of PV modules with air channel at their rear surface and usually ambient air is circulating in the channel for achieving both PV cooling and thermal energy output, which
can be used for building thermal needs. In PVT/AIR systems the thermal unit for the heat extraction, the necessary pump and the external pipes for air circulation constitute the complete system that extracts the heat from PV module and brings it to the final use. To increase the system operating temperature, an additional glazing is necessary, but it has as result the decrease of the PV module electrical output from the additional optical losses of the solar radiation.

Theoretical and experimental studies are referred to hybrid PV/T systems, with most of them including work on air heat extraction from the PV modules. Among the recent works we can notice the papers of Brinkworth et al (1997), on design and performance of building integrated hybrid PVT/AIR systems and of Hegazy (2000), who compares four PV/T air collectors. We also could refer the work of Eicker et al (2000), with the monitoring results from a BIPV PV/T system that operates during winter for space heating and during summer for active cooling and of Bazilian et al (2001) for the practical use of several PV/T systems with air heat extraction in the built environment. The building integrated PVs is going to be a sector of a wider PV module application and Lee et al (2001), Ito and Miura (2003) and Chow et al (2003), give interesting results on air cooled BIPV modules.

University of Patras has been involved in the research of PV/T systems with work on water and air cooled photovoltaics towards the increase of electrical and thermal output of BIPV PVT/AIR systems. The work aims to air heat extraction improvements with modifications in PVT/AIR systems (Tripanagnostopoulos et al, 2000, 2001a). In addition, improved PV/T systems with dual (air or water) heat extraction operation (Tripanagnostopoulos et al, 2001b) and modeling results confirming the improvements of a modified air cooled PV/T model (Tripanagnostopoulos et al, 2002a) are recently presented.

The electrical output of PV/T systems is of priority, as the cost of PV modules is some times higher than the thermal unit. The different performance of the two subsystems regarding temperature affects system cost and optimised modifications for both electrical and thermal efficient operation must be considered. The consideration of the environmental impact of PV modules by using Life Cycle Assessment (LCA) methodology has been presented for typical photovoltaic systems by several authors. LCA has been extensively used at University of Rome "La Sapienza”, starting with the PhD Thesis of Frankl (1996) on LCA for photovoltaic systems and followed by the study on the simplified Life-Cycle analysis in buildings (Frankl et al,1998), on the overview and future outlook of LCA for photovoltaics (Frankl, 2002) and also on the comparison of PV/T systems with standard PV and thermal systems confirming the environmental advantage of PV/T systems (Frankl et al, 2000).

In the present paper we give results for system energy performance and environmental aspects by the LCA method for standard PV and air-cooled PV/T systems. The work is based on the combined evaluation, for PVT/AIR systems, of both the "energy assessment”, that is experimentally investigated at the University of Patras, and the "environmental assessment”, in terms of LCA results, performed at the University of Rome with the aid of a specific software (SimaPro 5.1). These results are referred to typical PV modules and to glazed and unglazed PV/T solar systems for horizontal and tilted building roof installation, including also modified systems. The use of a booster diffuse reflector between the parallel rows for the horizontal installations is also suggested to increase the solar input to the PV and PVT/AIR systems and the corresponding results are presented. The calculated energy performance and the LCA results can be considered useful as guidelines for the application of the studied standard PV and PV/T systems as well as the modified ones.

Matthias Gebauer / Karsten Lambert
Solarverein Trier e. V. / FH Trier
Am Knieberg 29, D-54293 Trier
Tel.: (0049-651) 9960245, Fax: (0049-651) 65295
E-Mail: gebauer@fh-trier. de
Internet: www. solarverein-trier. de

1. Introduction

A thermal solar installation is thereby more susceptible to errors than conventional heating systems, because

• the system becomes actually more complicated

• new techniques and components often become "tried out" and

• a control by the user is difficult

It is therefore necessary to develop suitable assistance programs which help the user to analyse the disposition of the installation, to identify errors and to eliminate them.

Ulrike Jordan, Simon Furbo

Technical University of Denmark, Dpt. of Civil Engineering
DK-2800 Kgs. Lyngby
Tel.: 0045-4525-1889, sf@byg. dtu. dk

Abstract: Experimental and computational investigations of the flow fields around buffer plates in a small domestic hot water tank are presented. Inlet devices with different buffer plate diameters were placed at the bottom of an experimental glass tank. Temperatures were measured in different tank levels and two-dimensional velocity fields were measured in the centre plane of the tank with an optical method called Particle Image Velocimetry (PIV). The experimental results were used to model the influence of the buffer plate diameter on the stratification in the tank. The model is suitable for a limited range of buffer plate diameters. It was implemented into the simulation tool TRNSYS. Annual system simulations of a typical small solar domestic hot water system confirm earlier findings that the net utilized solar energy of the investigated and typically marketed system in Denmark could be improved by about 3 to 5 percent by employing a suitable buffer plate.

1. Introduction

In Denmark small solar domestic hot water systems typically consist of smaller collector areas and storage tanks than in Central Europe. The volume of the storage tanks is usually about 180 litres and the pipes are connected through the bottom of the tank. A buffer plate is placed above the vertical inlet pipe for cold water. Previous studies about stratification in solar domestic water tanks showed already that the design of buffer plates plays an important role for the system performance, especially if water is drawn off the tank with high flow rates. For example, (Carlsson 1993) carried out experiments with four different cold-water inlet devices (direct inlet, bent, perforated pipe and parallel plates) connected through the tank side of stores with a volume of 2 m3. A horizontally bent tube (facing in tangential direction to the side) turned out to cause the whole water volume in the tank to circulate in the direction of the entering fluid during and after a draw-off. In general, a large cross sectional area of the inlet was found to be of advantage. (Huhn et al. 2002) carried out experiments with inlet pipes entering the tank from the tank bottom and determined characteristic numbers for a tank with an open pipe, a bent tube and two buffer plate inlets. These numbers, however, could not directly be used to predict the system performance reduction or increase due to the buffer plates.

Investigations of the thermal stratification in solar stores with Particle Image (or Tracking) Velocimetry (PIV) were carried out, e. g. by (van Berkel 1997), (Shah, 1999) and (Knudsen et al. 2003). PIV is an optical experimental method to visualize velocity fields in a fluid. Van Berkel investigated the case of a flow stream entering the tank from the side with high velocities. The investigations were focused on the thickness of the thermocline situated between two initial temperature layers in the tank. Shah and Knudsen visualized the flow fields of different mantle tanks to determine thermal stratification inside the mantle as well as inside the tank.

Previous to the presented study, a model has been developed, based on measurements with two differently shaped buffer plates built into marketed steel tanks (Jordan and Furbo 2003-1). Based on the previous study, a further development of the model is described in the following, containing a relation used to quantify the impact of the buffer plate diameter on the system performance of a solar heating system.

2. Experimental Set-up

A scheme of the experimental set-up is shown in figure 1. It consists of a glass tank with a water volume of about 136 litres. Cold water enters the bottom of the tank and warm water is drawn from the top. The inlet — and outlet-pipes are connected to a cooling unit, a heating unit and to a buffer tank.

Temperatures are measured at 14 different levels of the tank. Instantaneous values are captured in time intervals of 10s. The thermocouples (of type TT) are mounted close to a tank corner. Average flow rate values over a time interval of 10s are measured with a magnetic-inductive flow meter.

The Particle Image Velocimetry — (PIV) equipment consists of two lasers, a camera, a processor unit (used to trigger camera and laser, as well as for data processing) and a computer (for further data processing).

PIV is a non-intrusive optical method to measure two — or three-dimensional velocity fields in a fluid. Small tracer particles (with a diameter of 20pm) are added into the fluid and illuminated by a pulsed laser sheet. The scattered images of the particles are recorded with a camera, based on electronic solid-state imagers (charge couples device (CCD) cameras). The time delay between two laser pulses is being adapted to the mean velocity of the flow and the magnification at imaging. It is assumed that the tracer particles move with the local flow velocity between two illuminations.

Measurements were carried out with three different buffer plates shown in figure 2.

Uniform initial tank temperatures at three different levels (of about 28°C/31°C, 42°C and 57°C) and up to four different flow rates were applied as reference conditions for the experiments of each inlet device. The inlet temperatures varied between about 7 and

Fig. 2: Investigated inlet devices. Max. diameters: 28, 52, and 70 mm; height of the buffer plates 12, 18, and 26 mm; widths of inlet gaps: 10,10, and 20 mm respectively.

3. Measurements

The flow field around the buffer plates is influenced significantly by the conditions applied. To investigate the effect of solely the flow rate and the plate diameter on the velocity field, in a first step, ambient temperature was used for both the water inside the tank and the

entering water. The field of view is marked by a rectangle shown in figure 1. It reaches from the centre of the tank to the tank wall on the right side of the tank, in a vertical pla The inlet devices are placed in the bottom centre of the field of view.

As an example, velocity vector maps, measured with a flow rate of 4 l/min, applying th small, medium scale and large buffer plate, respectively, are shown in figure 3. The velocity vector maps show mean values of 100 instantaneous vector maps, measured time intervals of 10 s to 200 s (time between recordings: 0.25 s to 2 s).

As shown in figure 3a, the water entering the store through the small inlet device pass the device with a fairly high vertical velocity component, with an angle of the flow direc to the vertical of about 45°. In contrary, with the medium scale inlet device and the sar flow rate (figure 3 b), the entering water is first directed to the tank bottom, then strean the wall, and from there in vertical direction. For the large inlet device, the water is also first directed to the tank bottom. However, no steady flow direction can be specified frc the vector map of mean values, due to the low velocities of the flow.

In figure 4 the vertical velocity component vz of a measurement with the small inlet dev and a flow rate of about 8 l/min is shown for six different storage levels, between h = 40 mm and 140 mm. The position of the lower outlet height of the opening of the smal inlet device is placed at h = 40 mm (see figure 2). As expected, the maximum value of situated in the curve closest to the inlet gap (at h = 60 mm), with a value of about 0.12 At h = 120 mm, the maximum value dropped by about two thirds.

In order to quantify and compare the upward velocities for different flow rates, the mea square roots of vzm in the horizontal was calculated for 4 measurements with different rates, uniform temperatures, using the small inlet device. vzm is defined as:

with N = 39 velocity vectors in horizontal direction

As shown in figure 5, the peak of vzm(h) moves only slightly upwards to increased stor heights for increasing flow rates. This means that the angle of the entering flow directs the vertical is approximately independent of the flow rate for the given reference conditions.

The strong deviation between the curves for a flow rate of 12 l/min shows that the flow was not stationary and that more images maps needed to be captured in order to get statistically independent results. The difference between the inlet temperature and the temperatures in the storage tank was within the measurement accuracy band of 0.3 K If the temperature of the water entering the tank is smaller than the temperature of the water in the tank, buoyancy effects occur for the flow stream. The cold water is presse downwards and the mixing is reduced. Therefore, the stratification of the tank is impro

the higher the temperature difference between the water in the bottom part of the store and the entering water.

In figure 6 and 7 measurements of the temperature distribution and velocity fields captured during a draw-off are shown. The large buffer plate was applied. Figure 7a) shows storage temperatures at different heights, the inlet temperature and the flow rate over the time. The initial tank temperature was 57°C, the flow rate about 11 l/min, and the inlet temperature about 7°C. The two temperatures measured at the lowest tank heights (h = 20 and 60 mm) are approximately equal throughout the measurement. The slope of these two curves is relatively small, which corresponds to a high degree of mixing of the lower part of the store at the beginning of the measurement. Nevertheless, the slope of the curves increases during the draw-off, for the (later) temperature drop of higher temperature layers. The size of the boundary layer that remains after the draw-off can be regarded as relatively small. The temperature gradient is higher than 30 K / 60 mm.

In Figure 6 a-d PIV velocity vector maps are shown. Each vector map shows mean values of 10 instantaneous vector maps. The corresponding particle images were captured within a time interval of 2.5 seconds. The cold water streams from the centre of the tank towards the tank wall. Next to the wall a vortex is developed, that rises throughout the draw-off.

Measurements of the thermal stratifications were carried out for draw-offs of a volume of about 40 l with three different (uniform) initial tank temperatures and four different flow rates for the three inlet devices. As an example, the distributions of the relative temperatures in the tank with the small buffer plate are shown in figure 8. The relative tank temperature is defined as:

T — T with: T : temperature at a given storage height

Trel = Tm : temperature of entering cold water

Tmax _ Tin Tmax: mean initial tank temperature

Fig. 7: Large inlet device (d = 70 mm).

7a) red temperatures and flow rate over the time. Mean temperature of the

entering water: 7°C, mean flow rate: 11 l/min, initial storage temperature: 57°C, draw­off volume: 401. The bars show (roughly) the time intervals in which PIV image maps were captured.

7b) Calculated temperatures (TRNSYS) over the time. Input values to the simulation Measured flow rate and inlet temperature as shown in a), as well as the initial storage temperature. Simulation time step: 4.5s.

7c) Inlet height of the water entering the store used for TRNSYS simulations over the time. hin = 55mm (At = 10s), 45mm (1min), 110mm (2min) and 150mm (3min).

SHAPE * MERGEFORMAT

SHAPE * MERGEFORMAT