J. NuBbicker, Institute for Thermodynamics and Thermal Engineering (ITW), University of Stuttgart

D. Mangold, Solar — und Warmetechnik Stuttgart (SWT, ein Forschungsinstitut der Steinbeis-Stiftung)

W. Heidemann, Institute for Thermodynamics and Thermal Engineering (ITW), University of Stuttgart

H. Mtiller-Steinhagen, Institute for Thermodynamics and Thermal Engineering (ITW), University of Stuttgart, Solar — und Warmetechnik Stuttgart, DLR Stuttgart Institute for Technical Thermodynamics

The solar assisted district heating system with seasonal duct heat store in Neckar­sulm-Amorbach is being realised since 1997. In 2003 about 200 accommodation units, a school with gymnasium and a shopping centre were supplied with heat by the district heating system. So far 5,263 m2 of solar thermal collectors are installed; the volume of the duct heat store is presently 63,360 m3. In the duct heat store heat delivered from solar collectors is stored from summer to winter. The duct heat store was extended twice; the operation of the first and second extension started in 1999 respectively in 2002. The maximum temperature in the duct heat store is expected to be about 85 °C. In 2002 and 2003 a solar fraction based on the total heat demand (space heating and domestic hot water) of 39 % was reached. The planned solar fraction of 50 % is expected to be reached within the next years. This paper presents an overview about the present status of the system as well as operational experiences and characteristic data for the system.

The partial differential equations are converted to algebraic equa­tions by means of finite-volume techniques using rectangular meshes on a staggered arrangement. Diffusive terms at the boundaries of the control volumes are evaluated by means of second-order central dif-

formulation EPF [5]. The resulting algebraic equation system was solved using a pressure — based SIMPLE-like algorithm [8], and the iterative convergence procedure was truncated when relative increments of the computed Nusselt number (Eq. 1) were below the con­vergence prescribed value (10~6). The meshes are intensified near the solid parts using a

tanh-like function with a concentration factor of 1 [4], in order to solve the boundary layer correctly. This aspect is indicated in Fig. 2 with solid triangles.

1.5 Verification.- The numerical solutions presented here have been calculated adopting a global h-refinement criterion. That is, all the numerical parameters (numerical scheme, numerical boundary conditions, etc.) are fixed, and the mesh is refined to yield a set of numerical solutions.

This set of numerical solutions has been post- processed by means of a tool based on the Richard­son extrapolation theory and on the concept of the Grid Convergence Index (GCI) [1][6]. The problems are solved on different meshes related by a mesh ratio r = 2. The tool processes a set of three con­secutive solutions in the h-refinement. The most relevant parameters arisen from the verification pro­cess are the GCI, the observed order of accuracy of the numerical solution (p), and the percentage of nodes of the post-processing grid where has been applied the post-processing procedure, which are called Richardson nodes Rn [1].

For all the post-processing results presented in this work (see for example Table 4), the observed order of accuracy (p), approaches the theorical values of the differential scheme used (between 1 and 3); the GCI values decrease as n increases; and the percentage of Richardson nodes obtained was high.

These results indicate that the estimator GCI is reliable, and that the solution is free of

SHAPE * MERGEFORMAT

According to results in Table 7 the percent­age differences dif% decrease with the de­crease in A’, and the maximum percentage difference was 8.3% at. A good

behaviour of Eq. 8 to calculate Kp parame­ter was found in A’ < 1.0 range. For A’ > 1.0, the percentage differences |<й/|% were higher due to the influence of the overall as­pect ratio A = 20. For the purpose of the design in transparent insulation technology, is an appropriate range.

Comparison with the Scozia results [7]

In Table 8 our numerical results of the Nus — selt number are compared with the Scozia nu­merical results. It was observed that the per­centage differences dif% increases with the decrease in A’, while the maximum percent­age difference was 04% at.

The dynamic simulations were carried out using Matlab/Simulink. The implementation of the thermodynamic model of a plant with 7 stages is depicted in the figure below. The model of a stage with all heat flows is collated to a single block, including thermal inertia. For an easier usage the evaporation area, quality of insulation, spacing between evaporator and condenser and the water volumes can be defined in a menu mask. Supplied heat energy, i. e. from a solar collector and ambient data are put in on the left. A scaling factor defines the quantity of brine entering the system.

The model had to be validated with measurement data from the unit in connection with a flat plate collector. The comparison of the simulation and the measured data can be seen in diagram figure 7. The accordance of simulated and measured data is very good. Differences in temperatures are mainly due to the uneven temperature distribution in the stages. So it occurs that a sensor is nearer to the cold seawater input and so measures a lower temperature.

With the help of a validated complete system, it is possible to predict the performance of the unit in other climatic regions. As an initial prognosis, the simulation was performed using weather data from Sidi Barrani on the Egyptian Mediterranean coast. The average daily production over a period of one year of a 1 m2 model with a 4 m2 flat plate collector is shown in Figure 8. The daily production is seen to be between 20 and 60 kg/d, depending on the time of year. The average yearly production achieved by the unit reached 47 kg/d.

In order to reduce the energy requirements even further, an additional system to recover the heat from the condensate and the discharged salt water, especially designed for solar thermal desalination units, is now being investigated with the financial support of RWE Aqua (Thames Water). Initial attempts have already shown very promising results.

Alexander Thur, Simon Furbo, Louise Jivan Shah

Department of Civil Engineering, BYG DTU, Technical University of Denmark
Brovej — Building 118, DK-2800 Kgs. Lyngby, Denmark
e-mail: alt@bvg. dtu. dk. sf@byg. dtu. dk. lis@byg. dtu. dk
phone: +45 /4525 — 1887, fax: +45 / 4525 — 1755

Introduction

The energy savings for a solar heating system depend on a number of quantities related to the energy system before and without a solar heating system and after installation of the solar heating system. Often these quantities are not known. It is therefore difficult to determine the energy savings with a good accuracy. In many European countries most solar heating systems are installed in houses with an oil fired boiler or a natural gas burner. For such energy systems the savings when adding a solar heating system are determined by the following quantities:

Before solar heating:

The efficiency of the boiler/burner The electricity consumption of the boiler/burner Service costs of the boiler/burner After installation of a solar heating system:

The net utilized solar energy of the solar heating system

Energy savings due to turning off the boiler/burner during summer periods

The efficiency of the boiler/burner

The electricity consumption of the boiler/burner

The electricity consumption of the solar heating system

Service costs of the boiler/burner

Service costs of the solar heating system.

Some of the most important quantities in connection with energy savings for solar heating systems are the energy savings achieved by turning off the boiler/burner during summer periods (when efficiency of the boiler/burner typically is rather low) after installation of the solar heating system. The paper will present measurements from practice of the efficiency of an oil fired boiler and of two natural gas burners. Based on these measurements it is possible to determine the energy savings by turning off the boiler/burner during summer. Further, based on theoretical calculations, estimates of the energy savings for solar heating systems will be presented. Similar results are reported from a field test in Sweden [1].

For these types of 2D concentrating collectors, the incidence angle dependence of the optical efficiency is composed by the angular dependence of both the transversal and the longitudinal plane, as illustrated in figure 5.

A propsed biaxial model, based on separate measurements of the effect on the optical efficiency of the reflector and the glazing, is shown in equation (1).

K (0L,0T) = RT (вт )fL (0t)

Here, fL (0) gives the influence of the glazing and is obtained by measuring the angular optical efficiency in the longitudinal plane at a given Or, when the reflector gives a constant contribution. RT (вт) gives the influence of the reflector and is obtained by measuring the angular optical efficiency of the collector in the transversal plane when 6L = 0. If glazing is used, the function K(0L= 0, вт) gets contribution from both the glazing and the reflector according to equation (1). Then R. T(eT) will be obtained from equation (2) according to equation (3).

K(0L=O, 0т)= Rt (вт)* fL (вт). (2)

This means that the ratio between the measured dependency in the 0T and 0L directions gives the influence of the reflector only.

The following complete expression for the angular dependency is obtained from equation (4).

In the longitudinal plane, the angular dependence consists only of the dependence of the glazing, f(0L). This means Km in the longitudinal plane can be described as flat plate collector by equation (5), where b0 is an incidence angle modifier coefficient, characteristic for the glazing (Duffie and Beckman, 1991).

J. Jaramillo, E. Mas de les Valls, C. Oliet and J. Cadafalch

Centre Tecnoldgic de Transferencia de Calor (CTTC)

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

Introduction

In the solar domestic water heating sector, the thermosyphon systems have achieved a good acceptance due to its relatively good efficiency, low installation and maintainance cost and due to the absence of mobile parts. As a result, these systems have become a really interesting device in order to exploit solar energy.

Thermosyphon systems typically consists of a collector and a storage tank which are mounted together outdoors. The outlet of the collector is connected to the top of the storage tank, and the inlet of the collector is connected to the bottom of the storage tank. When the water in the collector is heated up, density of water in the collector decreases with respect to the density of the water in the tank. When this difference of the density is high enough to overcome the resistance to flow forces (friction, expansions, contractions, other singular­ities…), circulation starts. During the night, water in the collector is cooled dawn and the density of the water in the collector increases. To avoid inverse circulation of hot water in the tank to the collector due to this fact, a check valve is normally installed in the pipe connecting the outlet of the collector to the tank.

Numerical studies on the designs of thermosyphon solar heaters are very often based on simple mathematical models as in [9, 14, 6]. In these models, thermal and fluid flow behaviour of the different parts of the thermosyphon cycle are evaluated by means of ex­perimental data (as the steady state efficiency curve of the collector) or unidimensional or zero-dimensional energy and momentum balances. All parts are linked together forming the cycle, and are solved by means of an iterative procedure.

According to the huge computational resources available nowadays, it is possible to go a step further in the modelling of thermosyphon solar heaters by making use of CFD codes. These codes are based on the numerical resolution of the multi-dimensional governing equa­tions (energy, momentum, mass, radiative heat transfer). A large number of control volumes would be required to discretize a whole thermosyphon system, therefore, CFD codes are not still suitable to do it. However, this kind of high level simulation is essential to get deeper insight into the physical phenomena of the flow and heat transfer in characteristic parts of the system and in order to calculate specific data that may be required by more simplified models such as heat transfer coefficients and friction factors. An example of CFD studies developed on solar collectors using CFD techniques can be found in [4], a flat plate trans­parently insulated cover is modelled solving the energy conservation low and the radiative
transfer equation. Another example of a detailed simulation of the components of a ther­mosyphon solar system can be found in [3]. There, the storage tank is studied by means of three-dimensional and transient CFD simulations of the temperature, pressure and velocity field (resolution of mass, momentum and energy equations). From this simulations conclu­sions on the influence of the inlet mass flow rate on the degree of thermal stratification are drawn.

The authors are currently using different simulation levels for the analysis of thermosyphon systems, they are divided in three groups, i) Simplified models; ii) Intermediate models; iii) CFD models. This papers describes the three levels of modelling focusing on the CFD mod­els. Some illustrative examples of results that can be achieved with these models are also

Bo Carlsson, SP Swedish National Testing and Research Institute, Sweden Stefan Brunold, Institut fur Solartechnik SPF Hochschule Rapperswil, Switzerland Andreas Gombert, Fraunhofer Institut fur Solare Energiesysteme, Germany Markus Heck, Fraunhofer Institut fur Solare Energiesysteme, Germany

To achieve successful commercialisation of new advanced windows and solar fa­cade components for buildings, the durability of these need to be demonstrated prior to installation by use of reliable and well-accepted test methods.

In Task 27 of the International Energy Agency Solar Heating and Cooling Pro­gramme, a general methodology for durability test procedures and service lifetime prediction (SLP) methods therefore has been developed that should be adaptable to the wide variety of advanced optical materials and components used in energy efficient solar thermal and buildings applications. The general durability assess­ment methodology is now adopted to some static solar materials to allow prediction of service lifetime.

Introduction

The IEA Solar Heating and Cooling Programme, Task 27 on the Performance of Solar Fa­cade Components started at the beginning of year 2000 with the objectives of developing and applying appropriate methods for assessment of durability, reliability and environ­mental impact of advanced components for solar building facades [1].

For the work on durability there are two main objectives. The first is to develop a general framework for durability test procedures and service lifetime prediction (SLP) methods that are applicable to a wide variety of advanced optical materials and components used in energy efficient solar thermal and buildings applications. The second is to apply the ap­propriate durability test tools to specific materials/components to allow prediction of service lifetime and to generate proposals for international standards.

As the result of this work, a general methodology has been developed [2], which is now adopted to some static solar materials. The work is performed in three case studies on anti-reflective glazing materials, reflectors and solar facade absorbers. Anti-reflective ma­terials that are studied include sol-gel coated and etched AR glasses. Reflectors that are studied include aluminium alloy based mirrors; some protected by clear coats, and glass mirror reflectors. Solar Fagade Absorbers that are studied include coloured sputtered se­lective solar absorber coatings, absorber coatings made with sol-gel technology and thick­ness insensitive spectrally selective paints.

The LHP principle also provides great scope for various design embodiments. This refers, first of all, to the design of the evaporator and the condenser, and also to the variants of their joining.

The main form of the evaporator is cylindrical with a diameter from 5 mm to 30 mm and more. The length of its active zone (heating zone) depending on the diameter may range from 20 to 400 mm. The compensation chamber may be located in the same body as the evaporator or separately. If the dimensions of the heat-load source are sufficiently large, an LHP may contain several parallel evaporators joined by a common thermal interface.

The shape, dimensions and design of the condenser — heat exchanger may differ greatly depending on the conditions of heat exchange with the heat receiver. Fig. 4 presents the main types of condensers used in LHPs.

Since the evaporator and the condenser are joined by means of separate pipe-lines of a relatively small diameter, which contain no capillary structure, they can bend quite easily and take practically any configuration if necessary.

Fig. 5 gives as an example the external view of an experimental LHP, whose pipe-lines 6-8 mm in diameter and about 21 m in length are bent in the form of a flat coil. The device is equipped with a cylindrical evaporator 24 mm in diameter with an active zone length of 190 mm and a condenser 300 mm long made by the variant “d”. The maximum capacity demonstrated by the LHP with ammonia as a working fluid when operated in a horizontal position was 1.7 KW at a vapor temperature of 60oC. The total thermal resistance in this case did not exceed 0.04 K/W. Another example characterizing the potentialities of these devices is an ammonia LHP 4.5 m “high” with pipe-lines 6 mm in diameter, which on trials transferred downwards a heat flow of 1 KW at the same vapor operating temperature. The tests and calculations performed show that the indicated technical data are far from being limiting for devices of such a type even with ammonia as a working fluid. With the use of water at an operating temperature of 100-150°C the heat-transfer distance and its value may be considerably increased.

Parabolic trough collectors usually track the sun with one degree of freedom using east-west or north-south axis. Solar tracking by these modes maintain the plane of solar beam always normal to collector aperture. Thus solar beam from different points of the parabolic trough reflecting surface is collected on the focal line receiver as shown in Fig. 1. The selection of tracking axis configuration is based on load profile and site latitude. To find a proper tracking axis configuration for Jordan, long term evaluation of collected solar energy on different tracking modes was conducted using TRNSYS package. Hourly weather radiation file was generated by averaging the measured data for number of years in Amman (latitude 32.02°). The results of TRNSYS simulation are presented in Fig.2. It is shown clearly that for the given latitude (32.02°) N-S tracking mode is more efficient than the E-W one.

Additional improvement of the system annual performance could be obtained by inclining the tracking axis. An optimum inclination (adopted in this paper) was found equal to 30° which produces 6% increase in the collected solar energy as compared with the horizontal N-S axis (see Fig.2).

Parabolic Collector System Design

An inclined single axis tracking parabolic trough collector was designed to generate hot water or steam at a pressure close to ambient. The parabolic collector system consists of four main parts: parabolic trough reflector, solar receiver, power supply, tracking control and

mechanism. A brief description of each component is as follows:

Parabolic Trough Reflector

A parabolic steel frame of aperture width 1.8m, rim angle of 74°, focal line 2m, and focal point length of 0.6m was considered.

The type of reflecting surface used in this study is stainless steel sheet (2mx1m) of a reflectivity close to (0.92). The sheet has enough flexibility to adopt the shape of the parabolic steel frame. This frame was constructed from steel trusses (as shown in Fig.3) to withstand the effect of wind force and any deformation in the shape of parabola that may occur. Two stainless steel sheets were connected to the frame from its edges by sliding them inside 3cm grooves at the edge of the frame. Total frame mass including the reflecting surface was found approximately equal to 70 kg. This frame configuration allows the testing of different reflecting surfaces.

Solar Receiver

A flat type solar receiver located in the focal length of the parabolic trough was used. The receiver consists of a 2cm diameter copper tube fixed on a 6cm wide flat plate. The receiver is insulated and inserted inside a rectangular casing with single glazed opening in the side facing the reflector as shown in Fig. 4. Heat loss from this type of receiver is affected by thermal resistance between flat plate-glass cover (convection and radiation), glass cover — surroundings (convection and radiation), polytheren insulation (conductivity), and metal casing-surroundings (convection and radiation). Solar energy concentrated on the receiver is delivered to a thermal storage tank by circulating distilled water in a closed loop. The rim angle of the parabola was designed to be small enough to concentrate reflected beam inside the solar receiver. The receiver was designed with two supports arrangements: tracking the sun with the parabola, or fixed in the plane normal to the tracking shaft. In the tracking arrangement the shadow of the receiver is projected continuously on the center of the parabola, while in the fixed arrangement the projection of the shadow of the receiver moves with sun position form east to west.

Solar flux pattern in the receiver is estimated by evaluating the width of the image reflected by the different reflector segments. Total flux (W/m2) on focal line is evaluated by segmenting the parabolic reflector and evaluating the reflected solar energy of each segment:

n

Total flux = flux(Y) + flux(2) + flux(S) +……………….. + flux(n) = ^ flux(i) (1)

i=l

V у

Where, n is number of reflector segments, i is segment number, Ib solar energy collected from each segment (W), and Lr receiver length (m). Image width (Wmg) depends on segment number or local rim angle (0ri) of the parabola and evaluated by the method given in [12]. To evaluate flux pattern (Watts) in the focal point of the parabola the accumulation of flux from all parabola segments have to be considered. Therefore at any distance from the focal point:

Solar flux pattern is generated at different beam incident angles using Equation (4) and the results are shown in Fig.5. Solar flux reduces as we move away from the receiver center. The deviation of beam incident angle from normal causes significant reduction in solar flux and increase in maximum width of solar image. Based on Fig.5, it can be concluded that receiver width must be around 6cm to intercept all solar image during low solar elevation angle hours.

The selected store consist of a vertical tank made of plexiglass of approximately 200 liters. The tank length is 1.0 m with an internal diameter of 0.5 m. The Plexiglass material is 8 mm thick. The tank has two inlet/oulet ports located at the top and bottom walls. The ports have an inner diameter of 0.04 m. Two plates with a diameter of 0.30 m and located 0.04 m away from the ports are use to reduce the inertia of the entering mass flow rates.

The PCM modules consist of rectangular section rings (0.01×0.02 m) located at different positions of the tank. The PCM used is a commercial hydrate salt (Sodium Acetate) with a latent heat of 240 kJ/kg and with a melting temperature of 56 C.

Three devices have been tested: a) a sensible storage device without PCM modules; b) an hybrid device with a 6% of its volume occupied by PCM modules; c) an hybrid device with a 10% of its volume occupied by PCM modules. These devices will be hereafter referred as designs S, H6 and H10 respectively (see Figure 1).

The stores have been tested considering the assay C in ENV 12977-3. The test has been performed as follows: i) Charging the store with a mass flow rate of 50 l/h and with an inlet temperature of 60zC until the outlet temperature at the bottom of the store exceeds 55 C; ii) Discharging the complete store with a mass flow rate of 100 l/h and with an inlet temperature of 20 C. Ambient temperature is fixed to 20 C.

CFD results

Considering the geometry of the stores and the working conditions explained above, the following hypothesis have been assumed: the heat transfer and fluid flow phenomena is axialsymmetric, convection is negelected in the phase change phenomena involved in the PCM modules, the physical properties are constant in accordance with the Boussinesq ap­proximation (density variations are only relevant in the buoyancy terms of the momentum
equations), the fluid behaviour is Newtonian, the viscous dissipation and the influence of pressure in temperature is negligible, and the radiating medium is non-participant. The ther­mal loss of the tank have been modelled considering heat transfer coefficients of 3 W/m2K.

The computational domain has been discretized using cylindrical coordinates. The size of the control volumes (CVs) has been maintained almost constant throughout the domain, except for the zones near the inlet/outlet ports. The computational grid was composed of nTx CVs. The simulated time was discretized using a constant time increment.

In order to verify the numerical solutions obtained, and to analyse the sensitivity of numer­ical parameters that account for the discretization (mesh spacing and time increment), the store without PCM modules has been simulated considering different grid densities (16×53, 32×106 64×212 and 128×424 CVs) and time-steps (0.25 s, 1 s and 4 s). After this analy­sis, a discretization of 32×106 CVs and a time-step of 1s has been selected to perform the numerical simulations.

In Figure 1 illustrative temperature countours for the charging sequence and for the tested devices are plotted. As can be observed, the thermal behaviour of all tanks are very similiar. Thermal stratification is clearly showed. However, the thermocline thickness is slightly higher in the case of H10 design.

In Figure 2 temperature profiles are given for the discharging test sequence. As can be

seen, after 2 hours the store without PCM is practically discharged while in the hybrid stores PCM modules of the top part of the tank, still have a high temperatures.