Category Archives: Particle Image Velocimetry (PIV)




Two small low flow SDHW systems were tested side-by-side in a laboratory test facility for solar heating systems. With the exception of the solar tank the systems were identical.

Fig. 1 shows schematic illustrations of the two tanks. Both tanks are standard mantle tanks suitable for low flow systems. In both tanks the auxiliary energy is supplied by electric heating elements.

One tank is equipped with a PEX pipe for hot water draw-off from the very top of the tank. The other tank is equipped with two PEX pipes for hot water draw-off from the very top of the tank and from the middle of the tank.

A three way valve, type RAVI from Danfoss A/S, ensures that the temperature of the tapped water during draw-offs is equal to the required hot water temperature. If this is not possible due to high tank temperatures, the valve ensures that the difference between the tapped water temperature and the required hot water temperature is as low as possible. That is, the three way valve ensures that the right part of the hot water is tapped from the top and from the middle of the tank.

The most important data of the tested systems are given in Table 1. Fig. 2 shows the collectors of the tested systems.


CFD Simulations

As mentioned before, more attention has to be paid to flow phenomena at bifurcations. Therefore simulations were carried out using CFD (Computational Fluid Dynamics). As can be seen from Fig. 10, only the first bifurcations at the absorber inlet were taken into consideration. A constant outlet pressure was assumed. A mass flow of 0.01 kg/s (7.2 l/ (m2h); laminar flow) and 0.1 kg/s (72.1 l/(m2h); turbulent flow), respectively, was chosen as an inlet condition. Fig. 10, which shows the results of a high-flow simulation, confirms the effects observed in the flow experiments: the fluid tends to prefer the inner channels of the fractal structure. A cross section of the first left branch also reveals the Dean vortices pre­sumed to be the cause for the described effect (indicated by vectors in the detail picture).

72.1 l/(m2h) (turbulent)

3.24e 01 I 2 92e 01 2.59Є-01 2.27Є-01 194e01 1.62Є-01

і 1.30e01 9.72e02 6.4ве 02 I 3 24e 02 H O. OOe+OO


An algorithm which is capable of generating fractal hydraulic structures on a given area with fluid in — and outlet was developed and served as a basis for the programme FracTherm. With this programme it is also possible to carry out hydraulic as well as thermal simulations and visualise the results. A total collector efficiency factor F can be determined. The simulated F of a fractal absorber was high compared to measured val­ues of conventional fin absorbers. DXF files can be exported from FracTherm, which al­lows computer aided manufacturing. Flow experiments with ink indicated secondary flows at bifurcations. Thermography pictures showed that a uniform heat transfer can be ob­tained. CFD simulations confirmed the secondary flow effects observed in the flow experi­ments.

In order to be able to compare fractal structures with conventional ones, it is intended to use FracTherm also for simulations with serial and parallel hydraulic structures. Moreover, further CFD simulations as well as experiments are planned to investigate flow phenom­ena and determine the pressure loss coefficients at bifurcations. Variations of the FracTherm parameters will show their influence on the total energy efficiency. In a further step these variations can be carried out automatically and thus lead to an optimisation us­ing "evolution strategy" [7]. The simulation environment ColSim can be used to carry out dynamic system simulations which will also take thermal capacity effects into considera-

tion. Finally, it is intended to produce and measure prototype absorbers. Thus a comparis­on of fractal absorbers and other heat exchangers with conventional ones will be possible.


The author likes to thank the German Federal Environmental Foundation (DBU) for fund­ing this research work in its scholarship programme.


[1] Duffie, J. A. and Beckman, W. A.: Solar Engineering of Thermal Processes. 2nd edition. New York: John Wiley & Sons, 1991

[2] Eisenmann, W.: Untersuchungen zu Leistungsfahigkeit und Materialaufwand von Sonnenkollektoren mit serpentinen — und harfenartiger Rohrverlegung. Fortschr.-Ber. VDI Reihe 6 Nr. 490. DQsseldorf: VDI Verlag, 2003

[3] Frey R., Frei U., Brunold S.: Bestimmung des Kollektorwirkungsgradfaktors F’ an flQssigkeitsfCihrenden Solarabsorbern. Solarenergie PrQf — und Forschungsstelle SPF-ITR, Oberseestr. 10, CH-8640 Rapperswil, 1995

[4] Hermann M., Koschikowski J. und Rommel M.: Corrosion-free solar collectors for thermally driven seawater desalination. Solar Energy 72(5), pp. 415-426, 2002

[5] Martin, H.: WarmeQbertrager. Stuttgart; New York: Thieme, 1988

[6] Nachtigall W. und BlQchel K. G.: Das groBe Buch der Bionik. Stuttgart; MQnchen: Deutsche Verlags-Anstalt, 2000

[7] Rechenberg, I.: Evolutionsstrategie ’94 (Werkstatt Bionik und Evolutionstechnik; 1). Stuttgart: frommann-holzboog, 1994

[8] Wittwer, C.: Colsim — Simulation von Regelungssystemen in aktiven solarthermischen Anlagen. Dissertation, Fakultat fQr Architektur, Universitat Karlsruhe (TH), 1999

Summary and Prospect

The supervising of large-scale solar thermal systems in the frame of the program Solarthermie-2000 produced plenty of reliable data which enable an evaluation and optimisation of a system. The technology of preheating domestic hot water in buildings with a high hot water consumption has passed the demonstration phase. The technology is now subsidized by the market incentive program like small scale solar thermal systems as well.

The new program Solarthermie2000plus extended the task for large-scale solar thermal systems. The solar fraction of the total thermal energy needed for a building is supposed to be over 10 %. This means a contribution of the solar plant to the heating. The focus of the new program is on combined systems for domestic hot water and space heating generation, local network heating, especially in combination with other renewable energies like biomass and new applications for large-scale solar thermal systems.


/1/ www. fh-offenburg. de/mv/st2000

/2/ S. Himmelsbach, E. Bollin, U.-M. Klingenberger; “Solare Dusch — und

Beckenwassererwarmung in der albtherme Waldbronn”; Proceedings of the 13. Symposium Thermische Solarenergie in Staffelstein, OTTI Regensburg, 2003 /3/ F.-A. Peuser, K.-H. Remmers, M. Schnauss; “Solar Thermal Systems”; Solarpraxis Berlin, 2002

Facade collector design

Considered facade solar collector is a standard selective liquid flat-plate collector integrated into a building envelope. Layout of investigated type of facade collector is shown in Figure 1. The collector consists of standard spectrally selective absorber (а/є = 0.9/0.09), an air gap and single safety glazing. Collector back and edge insulation is common to building insulation layer. Facade collectors are usually available in wooden frames as large-scale installation panels. Collector panel is directly mounted on insulation envelope of building facade, there is no thermal separation between absorber and insulation envelope in the form of ventilation gap. Solar collector is thermally coupled to the building wall. The integration brings several essential advantages in comparison with solar collectors mounted separately from building envelope (in the front of the envelope or on the flat roof). Additionally to the basic function of solar collector, facade collector serves also as protection shield against atmospheric effects (weather protection) and improves the thermal properties of the building with respect to passive solar gains. Furthermore, collector integration into building

facade is aesthetically more attractive solution when compared to collector plants placed on the flat roofs, which create foreign-like bodies on buildings.

Conventional heating system

In figure 7 the temperatures and volumetric flow rates of the heat distribution net are shown. In 2003 both supply and return temperatures increased in comparison to 2002. At the end of 2002 some new buildings were connected to the district heating system by an approx. 1000 m long pipe. Because of high circulation flow rates and a low heat demand this part of the net causes net return temperatures. The short-term increase of the volumetric flow rate in December 2002 was caused by a temporary supply of an adjacent district heating system.

It is also evident that net supply temperatures in summer 2002 / 2003 and in 2003 even in spring are higher than required. The reason therefore are the hydraulic conditions in the 3- pipe heat distribution net, see [3]. Because the collector volumetric flow rate is higher than the volumetric flow rate in the heat distribution net not enough cold water is available to cool down the supply temperature by admixing cold return flow to the hot supply flow.

Heat balances

In figure 8 a schematic heat balance for the district heating system in Neckarsulm-Amor — bach is depicted for 2003. About 70 % of the total solar heat delivery (2,121 MWh) was used to charge the duct heat store. 1,109 MWh heat were delivered by an auxilary gas boiler. The heat demand of the buildings amounts to 1,305 MWh and the heat losses in the heat distribution and solar net to 586 MWh. 153 MWh were discharged from the duct heat store and 548 MWh of thermal solar heat were directly used for heat supply in the heat distribution net. In 2003 a solar fraction of the total heat demand of 39 % was reached.

The heat losses in the heat distribution and solar net are high because the net is almost completely installed but less buildings are connected than expected. The discharge of the duct heat store is about 10 % of the charging heat amount because duct heat stores need a 5-8 years heating-up period to reach a quasi-steady-state behaviour. A significant

Figure 8: Schematic heat balance for Neckarsulm-Amorbach in 2003

In table 2 some characteristic data for the heat distribution system in Neckarsulm-Amor- bach are listed. In 2003 the heat demand decreased eventhrough more buildings were connected to the district heating system. The solar fraction increased from 18 % in 1999 to 39 % in 2002 and 2003. The planned solar fraction of ~50 % is expected to be reached in the next years due to the heating-up period of the duct heat store.


A further extension of the duct heat store is planned when the heat demand in the residential area increases. However instead of extending the duct heat store, the installation of a heat pump will be taken into consideration to increase the usable temperature level of the duct heat store.



J. NuBbicker, D. Mangold, W. Heidemann, H. MQller-Steinhagen: Erfahrungen aus Betrieb und Ausbau der solar unterstQtzten Nahwarmeversorgung mit Erdsonden-Warmespeicher in Neckarsulm-Amorbach. Proc. of 12. Symposium Thermische Solarenergie, Staffelstein, Germany 24.-26.04.2002, pp. 471-475


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, Stuttgart, 2004


J. NuBbicker, D. Mangold, W. Heidemann, H. MQller-Steinhagen: Solar assisted district heating system with duct heat store in Neckarsulm-Amorbach (Germany), Proc. of ISES, Goteborg (Sweden), June 14-19, 2003

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.

Seasonal hot water heat store

The volume of the heat store amounts to 12 000 m3 (height: 20 m, diameter:

32 m). Figure 5 shows the hot water heat store during construction. The charging and discharging behaviour for 2002 is shown in figure 6. Charging of the store mainly occurs from May to August, discharging in the autumn months. Approximately 20 % of the heat delivered by the solar collectors is used directly for preheating the district heating net. In figure 7 the variation of temperature in the store is depicted for the years 1997 to 2003. The seasonal variation of temperature can be seen easily. The highest temperatures reached in the store in the end of summer amount to 80 °C on the top and to about 60 °C at the bottom. This also means that the heat capacity of the store is not sufficiently used. The building development and the solar collector fields of the second extension are realised unexpectedly slow and different to the planning. The lowest temperatures reached in the store are about 40 to 45 °C at the bottom depending on the year, since the return temperatures of the district heating net, which are the lowest temperatures in the whole system, are higher than expected. The difference between the lowest temperature in the store and the start temperature in October 1996 (11 °C) represents the unusable heat content of the store which was needed to put the store into regular operation. Furthermore in figure 7 an increase of the temperature besides the store can be seen from November 2000 to April 2001. This is due to a defect of the drainage pump and the resulting wetting of the thermal insulation. The heat store is surrounded by a drainage system to protect the thermal insulation of the heat store from flooding by ground water. The water is collected in a drainage duct and delivered to a nearby lake by the drainage pump. After repairing of the pump the temperature around the store continuously decreased during the years 2001 and

2002. The thermal insulation seems to dry up.

The heat losses of the store are between 322 and 360 MWh/a, corresponding to a moderate efficiency of 60 % for the seasonal heat storage. The calculated value amounts to 220 MWh/a. This difference is due to the operation of the heat store at higher temperatures than expected caused by the high net return temperature. It generates an offset of the store temperature of at least 10 K at the bottom and increases the heat losses significantly since the lower third of the heat store is not thermally insulated. It is also assumed that the thermal insulation of the heat store is partly wet due to insufficient drainage of ground water. Approximately calculations to take into consideration the wetness of the thermal insulation yield additional heat losses of 200 MWh/a.

The contribution of the connecting pipes between heating plant and heat store is also not negligible. The pipes have an overall length of 55 m. These pipes were used for charging and discharging operations for 6 600 h in 2002. Assuming a temperature drop of 0.5 K between in — and outlet of the pipes the resulting heat losses amount to 50 MWh/a.


О Discharging

Direct usage of solar heat

Energy content of the heat store

1 2 3 4 5 6 7 8

Figure 6: Charging and discharging of the heat store in 2002

Micro-V Covering Materials with High Light Transmittance for Solar Collectors

P. J. Sonneveld and G. L.A. M. Swinkels

Wageningen UR, A&F, P. O. Box 17, 6700 AA Wageningen, The Netherlands, tel. +31.317 476 438, fax. +31.317 4 75 347, E-mail: piet. sonneveld@wur. nl


Maximal light transmittance of the covering material is important for solar collectors maximising yield of the system. Furthermore a second sheet of covering material can be applied with low light loss to increase thermal insulation. Therefore research is aimed at improving light transmission. Ray tracing method has been applied to design the optimal geometry of the material. Light transmission, thermal insulation, structural performance and yield aspects of solar collectors are combined in this research with glass as basic covering material


Intensive research is aimed at developing the special geometry of the sheet material with maximal light transmission. With ray-tracing software light transmission of structured materials was optimised. These investigations were started to increase light transmittance of double insulated greenhouses, the approach is promising for solar collectors too. A beam of light incident to a flat transparent sheet will be partly transmitted, according to Fresnell law, and partly reflected (Fig. 1a above). At lower angles of incidence, reflection increases. For a flat sheet the reflected part of the light does not enter the system and therefore is lost for the solar collector. For a zigzag surface, the primary reflected light hits the other part of the sheet surface with favourable angle if incidence and after transmission will (partly) enter the system after all (Fig. 1a below). This is especially effective at low angles of incidence of the primary light. By this effect the transmittance for direct and diffuse light of a zigzag-shaped single sheet of PC increases with about 5% compared to a flat single sheet. A&F developed the idea and the optimal shape, thickness and grid of such a zigzag sheet. Calculations and measurements of single and double sheets with different pigment additives and with and without coatings were performed. For the calculations of the light transmission a ray tracing computer program was developed. In Fig. 1b a prototype sample of the zigzag-structure material is shown. In Figure 2 the calculated data of direct perpendicular light transmission is presented for a single PC — sheet.

In Figure 3 the calculated and measured data of diffuse light transmission are presented for a single PC-sheet and a double-wall sheet with and without additives. The measurements were carried out at A&F, applying an integrating sphere allowing a sample size of 50 x 50 cm.

The diagram shows a good agreement between calculations and measurements. A considerable improvement of the light transmittance can be realised. The diffuse light transmittance of the sheets increases with 4% (single layer) to 6 % (double layer).

It can also be observed that the local optimum for the inclination of the zigzag-shape is 45- 60o. When regarding other criteria like insulation value, material consumed and material strength an inclination of 48o is ideal. The transmittance for diffuse light of a double zigzag sheet with an inclination of 48o is 78.8%.

Fig. 1a:The principle of the transmittance and reflection of light beams hitting a flat sheet (above) and a zigzag-sheet (below).

The developed zigzag-sheet is manufactured of polycarbonate (lexan) by GE-Plastics and applied in new types of greenhouses (sonneveld et al. 2003). The principle can also be applied for solar application but a difficulty for this application is the rather high stagnation temperatures with sometimes occur in solar collectors. Therefore the typical covering material for solar collectors should be glass. However the ZigZag shape is difficult to be made of a glass sheet. Therefore an adaptation to very small surface V-structure (Fig.4) is studied.

0 10 23304)50 60 70 80 90

angle [deg

Fig.3 Overview of the diffuse light transmission of a 1mm thick Zigzag-designs as a function of the zigzag angle (zero is planar material) for different absorption coefficients o the materials

This surface V-structure of the material will result in a decrease of light reflections in the same way as the ZigZag structure. This will result in extra light transmission over the complete solar spectrum with increasing angles of the microstructure.



Light transmittance is depicted in Fig. 5 as function of the angle of the Micro-V structure. It shows that zigzag angle near 45 degrees decreases transmission considerably and that the optimum angle is at about 50 degrees. This increase of about 5 % is comparable with other antireflection methods (Furbo 2003)

The collector efficiency is the most important factor. It can be calculated as a function of the radiation normalized temperature difference and the heat loss coefficient U.

Both transmission and absorption are dependent on the angle of incidence. The efficiency decreases due to heat losses, which depend on the heat transfer coefficient U of the cover for well-insulated collectors. For single layer covering U value is 3.5 W/m2K and for double covering approximately 2.3 W/m2K. The collector efficiency can be determined with:

V = Vo[r(<P, в),а(<р, в)]- U ■ T*

with р0[т(ф, в),а(ф,0) the reference efficiency dependent on the product of the transmission т of the collector covering and the absorption coefficient a of the absorber dependent on the angle of incidence ф and the tilt angle Ф of the north-south oriented solar collector, T* a radiation normalized temperature difference according to

and G the global radiation (in the Netherlands maximum at 800 W/m2 in this

case 400 W/m2 is chosen).

With the transmission values given in Table 1 and a typical absorption coefficient of 0,96 for the collector the efficiency of the collector can be calculated. The results are presented in Fig. 6 and 7 for respectively direct perpendicular radiation and diffuse radiation.

For single Micro-V covering material an increase of 4-5 % can be observed compared with single glass. For a double sheet covering this increase in efficiency is 8 %. The total result for double layer Micro-V is extra yield at lower and higher temperature differences. At higher temperature differences the efficiency can be doubles compared with a single glass covering.

Fig. 7 Efficiency of diffuse perpendicular radiation for different types of transparent

covering materials,…………….. single layer glass, …………….. double layer glass,————- single layer

Micro-V and ———— double layer Micro-V

With Dutch climate data (Reference Year van weather station De Bilt, The Netherlands) the yearly yield is calculated with a simulation program for Solar Boilers VABI for a ZEN Solar system with 2.75 m2 collector surface area and a boiler of 90 dm3. The tap water demand is 110 dm3 per day with an input water temperature of 15 oC and an end temperature of 65 oC. This corresponds with an energy load of 8391 MJ per year. The optical efficiency of the absorber is 0,867. In Table 2 the yield of the systems with different cover materials are summarized. Changing from single glass to Micro-V glass will result in 4 % extra yield. An extra yield of 9 % is possible with a double Micro-V covering.

Table 2 Yield with the different covering materials

Transparent cover

Single normal glass

Single glass with Micro-V




Double glass with Micro-V






Optical efficiency





Thermal loss factor





Yield per year [MJ] Yield compared with





reference system [%]






A new covering material for solar collectors is in development with 5 % enhanced transmission over the whole solar spectrum. For high temperature applications of solar collectors the efficiency can increase up to 50 % with a double covering due to the higher insulation value of the double sheet material with a good transmission. An yield increase of 4 % is seen by changing the standard glass by micro-V. With double Micro-V glass the extra yield is 9 % as a result of the higher insulation value.


Sonneveld, P. J., G. L.A. M. Swinkels and D. Waaijenberg, 2002, Greenhouse design for the future, which combines high insulation roof material with high light transmittance, Paper no. 02SE013, International Conference on Agricultural Engineering (AgEng), Budapest, Hungary, 30 June — 4 Juli 2002, pp. 8 Sonneveld, P. J, Adriaanse F., 2002, New Energy Saving Greenhouse Roof with a High Light Transmittance — Zigzag greenhouse glazing, Paper no. 02SE004, International Conference on Agricultural Engineering (AgEng), Budapest, Hungary, 30 June — 4 Juli

2002, pp. 102

Furbo S., Shah L. J., Thermal advantages for solar heating system with a glass cover with antireflection surfaces, Solar Energy, 75, pp. 513-523

Keywords: transparent material, light transmission, thermal insulation

Assumptions for the simulations and calculations

The following cases were defined and simulated to elaborate possible saving potentials when adding solar heating systems:

• The simulations were only carried out for Solar Domestic Hot Water Systems (SDHW).

• All simulations and comparisons were made for two boiler types. A condensing natural gas burner and an oil boiler, both with the same settings for calculating the boiler efficiency as described for Fig 8.

The reference system was assumed to be designed similar to the systems of the two monitored boilers described in the beginning [2]. This leads to the following key figures: Nominal power: 20 kW

Hot water tank: 160 litres, completely kept on set temperature

Insulation of the tank: 40 mm, thermal conductivity: 0.04 W/mK

Set temperature: 50°C with a hysteresis of 5K

Hot water tap temperature: 45°C

For the Solar domestic hot water systems, the following system design was simulated: Collector area: 2 m2 or 4 m2

Collector efficiency: Лс = 0.8; a1 = 3.12 W/ m2K; a2 = 0.012 W/ m2K2

Incident angle modifier: ke = 1 — xana(&/2); a = 3.3; for 0 = 50°: k% = 0.92

Hot water tank: 200 litres, auxiliary volume: 61 litres

Insulation of the tank: 40 mm, thermal conductivity: 0.04 W/mK

Hot water tap temperature: 45°C

The space heating system is always supplied directly by the boiler.

In Table 1 the monthly main data for the reference system with low hot water demand and a natural gas burner is shown. The space heating demand was defined to be 90 kWh/m2a, which in Denmark is a typical existing house. The average size of the one family houses is around 180 m2 which results in a yearly space heating demand of 16,200 kWh/a. Also it is assumed that there is no space heating demand during summer time, which in the Danish climate is the case from May till September.

Two domestic hot water demands (related to Fig 6 and Fig 7) were simulated, a high demand with 3,000 kWh/a and a low demand with 1,500 kWh/a. The monthly distribution of the consumption was simulated according to Fig 7. Additional variations were simulated with varying daily values between 80% and 120% of the average daily consumption during the week. The daily profile was created with different hourly domestic hot water consumptions as a percentage of the daily consumption, beginning at 5 a. m. and ending at 10 p. m. and with values between 0.8 % and 13.1 % resulting in 100 % for the whole day.


switched off because of no space heating demand during this period. Instead of the natural gas burner an electric heating element with an efficiency of 100 % supplies the missing energy to prepare the hot water.

In the case of “SDHW_1500_2m2”the total yearly energy saving per m2 collector area increased about 10 % compared to the results before presented in Fig 11.

The second group of simulations was done with an oil boiler as an auxiliary source. In Fig 13 the efficiency of the oil boiler used in different energy systems is shown. Compared to the natural gas burner the boiler efficiency in general is lower. In the summer period the difference is quite large. The lowest efficiency in August is less than 45 % in the reference case with low hot water demand (Ref_1500). Therefore the fuel reduction potential in this case is much higher. In Fig 14, the fuel reduction of the SDHW-systems

compared to the reference systems is shown. In Fig 15 the fuel reduction is shown for the case that in the summer period from May till September the oil boiler is switched off and an electric heating element with 100% efficiency is supplying the additional energy. In the case of “SDHW_1500_2m2” the total yearly energy saving per m2 collector area increased about 60 % compared to the results before presented in Fig 14.


In Fig 16 a summary on yearly values of all calculated systems is shown. In the case of the small hot water demand of 1,500 kWh per year and a SDHW-system optimized for high solar gains (2 m2 collector area) the calculations result in the highest specific fuel reduction per m2 collector area: 960 kWh/(m2a).

This is twice as much as the pure solar gain of about 480 kWh/(m2a). The solar fraction of this SDHW-system is about 39 % for the whole year. In the summer period in this case the solar fraction is between 60 and 80 %. The solar fraction here is calculated by hot water demand minus boiler supply (= net utilised solar gain) and then divided by the hot water demand. This means that all the heat losses are covered by the SDHW-system.

The SDHW-system with 4 m2 collector area and 1,500 kWh hot water load leads to a solar fraction of 57 % for the whole year and between 90 and 98 % in the summer period. When increasing the hot water load to 3,000 kWh per year, the solar fraction is 47 % for the whole year and between 72 and 92 % in the summer months.

The results show very clearly that the lower the system efficiency of the reference system is, the higher the additional reduction potential of the fuel consumption by improving the system efficiency is when a SDHW-system is added to the system.


Yearly energy saving potentials with SDHW-plants

Gas boiler Gas boiler + Electricity Oil boiler Oil boiler + Electricity








1500 / 4mz / Fuel red. 1500 / 4mz / Solar Gain 3000 / 4mz / Fuel red. 3000 / 4m3 / Solar Gain 1500 / 2mz / Fuel red. 1500 / 2mz / Solar Gain

1500 / 4mz / Fuel red. 1500 / 4mz / Solar Gain 3000 / 4mz / Fuel red. 3000 / 4m3 / Solar Gain 1500 / 2mz / Fuel red. 1500 / 2mz / Solar Gain

1500 / 4mz / Fuel red. 1500 / 4mz / Solar Gain 3000 / 4mz / Fuel red. 3000 / 4m3 / Solar Gain 1500 / 2mz / Fuel red. 1500 / 2mz / Solar Gain

1500 / 4mz / Fuel red. 1500 / 4mz / Solar Gain 3000 / 4mz / Fuel red. 3000 / 4m3 / Solar Gain 1500 / 2mz / Fuel red. 1500 / 2mz / Solar Gain

Fig 16 Overview of all simulation results on a yearly basis

A short estimation on the basis of primary energy consumption showed that because of the high solar fraction in the summer period only a very small amount of electricity is necessary. Even in the case of the SDHW-system with the smallest solar fraction of 39 % per year the reduction of primary energy consumption still is approx. 880 kWh/m2 (assuming a primary factor 2.5 for electricity and 1.15 for oil [3]). Compared to 960 kWh/m2 mentioned before this is a very little influence. If the solar fraction of the SDHW-system increases the difference between the reduction potential of final to primary energy decreases to increasingly smaller values.

In principle a solar thermal system has a very little demand for maintenance. Mainly there are 3 parts where some work is necessary. The first point is to check the glycol and to change the liquid maybe all 5 to 10 years. The second point is the pump which can
perhaps break and then has to be replaced. The third point is the collector sensor and the controller which can be destroyed by lightning in the case of bad luck. On the other hand if the boiler can be switched off 40% of the year, the lifetime of the boiler will probably increase. In general the advantages and disadvantages will be more or less of the same magnitude.

The operating time of the solar pump is typically in the range of about 1.500 to 2.000 hours per year. If we calculate a pump with 50 W power, this leads to an electricity consumption of about 75 to 100 kWh per year. This is about 4%-10% of the fuel reduction. The summer months May till September have 3.672 hours. If in the reference case the boiler consumes 20 W in average (for standby and running two to three times a day for really supplying energy; also see Fig 3) this also leads to 73 kWh only in the summer period. This means, if there is additional electricity consumption, it is definitely much lower than 4%-10%.


In this paper the realistic behaviour and efficiency of heating systems were analysed, based on long term monitoring projects. Based on the measurements a boiler model was evaluated. Comparisons of measured and calculated fuel consumptions showed a good degree of similarity. With the boiler model, various simulations of solar heating systems were done for different hot water demands and collector sizes. The result shows that the potential of fuel reduction can be much higher than the solar gain of the solar thermal system. For some conditions the fuel reduction can be up to the double of the solar gain due to a strong increase of the system efficiency. As the monitored boilers were not older than 3 years, it can be assumed that the saving potential with older boilers could be even higher than calculated in this paper.


[1] T. Larsson, Enkatundersokning om energibesparing och drift med solfangare, Orebro University, Department of technology, Sweden

[2] S. Furbo, L. J.Shah, C. H.Christiansen, K. V.Frederiksen, Kedeleffektiviteter for oliefyr og naturgaskedler I enfamiliehuse, Sagsrapport BYG. DTU R-072, 2004, (www. byg. dtu. dk/ =>Publications =>Scientific Reports: byg-r072.pdf)

[3] H. Krapmeier, et. al., CEPHEUS-Austria Final Report, 2004, Energieinstitut Vorarlberg, Austria (www. cepheus. at)

[4] F. Kristiansen, Lavenergffikkehuse — IEA-Task13 — Malinger og beregninger, BYG. DTU, Denmark; Rapport R-025, 2000, ISSN 1396-4011

[5] K. Ellehauge, Malinger pa solvarmeanlffig til kombineret brugsvands — og rumopvarmning, Danmarks Tekniske Hojskole, Denmark; Meddelelse nr. 255, 1993

[6] K. Ellehauge, Monitoring of Danish Combisystems, Ellehauge&Kildemoes, Denmark; aLtENER Solar Combisystems, 2003, www. elle-kilde. dk/altener- combi/dwload/Monitoring_of_Danish_combisystems-17-02-04.pdf

[7] L. L.Overgaard, et. al., Erfaringer fra malinger pa kombinerede solvarme — og biobrffindselsanlffig, Teknologisk Institut, Denmark; SEC-R-7, ISSN-nr 1600-3780, 2000

[8] S. Knudson, Consumers influence on the thermal performance of small SDHW-Systems — Theoretical Investigations, BYG. DTU, Denmark; Solar Energy, Volume 73 , Issue 1, July 2002, pages 33-42; www. elsevier. com/locate/solener

[9] W. Streicher, et. al., Simulation programm SHWwin, TU-Graz, 1999, www. iwt. tugraz. at/de/main. html => Solarthermie => downloads => SHWwin

Thermodynamics of Regenerative Distillation

Desalination of seawater can be accomplished in principle by many different processes. The same fundamental thermodynamic laws govern every such process. By means of a thermodynamic analysis for the so-called "differential process”, as it proceeds in series of infinitesimal steps with complete equilibrium being maintained at all times, the reversible, isothermal work for any steady state process regardless of mechanism may be found to be approximately —

Wmin = 2.98 kw-hr/1000 gallons or 2.83 kJ/kg

This figure (or a similar one depending on operating temperature and data chosen) is the one usually quoted as the minimum theoretical work. Practically, it is a very unrealistic figure, not only because it assumes complete reversibility of all operations, but also because it would involve pumping an infinite amount of feed water and the pumping work would then be infinite [2].

Distillation is a vaporization process driven by heat. Essentially, it is a heat pumping process driven by a heat engine. Here the work requirement depends not only on the quantity of heat to be pumped but also on the temperature difference over which it is pumped. When heat is used to vaporize water from a salt solution, the vapor evolved contains the full amount of heat of evaporation and the only net result is a slight degradation in temperature due to the Boiling Point Elevation (BPE) of a salt solution. This heat can all be reutilized merely by restoring it to its former temperature — i. e. by heat pumping. By using the first law of thermodynamics and the Carnot relationships, the minimum heat (Qk) necessary to drive an ideal distillation process over a given temperature range AK can be shown to be —

This parameter represents the amount of "effects” that the device is able to perform. "Effects” stand for the amount of kilograms of distillate obtained from a heat input equal to the latent heat of evaporation of 1 kilogram of seawater. A GOR of 1 thus means that there was no gained effect and the unit mass of distilled water corresponds with the amount of energy needed to evaporate that unit mass. Obviously, GOR above 1 indicates a system that is regenerating its heat for multiple effects. gOr serves as a factor by which distillation devices of different design may be compared with respect to their heat efficiencies.

We see therefore that in order to maximize the efficiency of a distillation process from a thermodynamic approach (and minimize irreversibilities) the whole process of heat insertion and removal should take place at just a fraction of a degree difference from each other (small AT). This however would require infinitely large surface areas and is thus impractical. Likewise, ideally, one would desire that the salinity of the incoming water remain constant and not increase as part of it is evaporated. This however, would require an infinite flow of water to be processed through the system and is also impractical. As

the thermal energy inserted into the system must be ultimately removed it would make sense to benefit from the removed heat in such a way that it could be reinserted into the system. Finally, in order to operate at a higher thermodynamic efficiency the temperature range of operation must also be large. Following is a basic diagram describing a process that incorporates these issues.

Figure 1- Parameters of a thermally optimized distiller

The described optimized enclosure operating over a large temperature range, with a small temperature drop, while regenerating its latent heat of condensation, will still have a limited ability to produce distillate. The amount of heat needed to "pre-heat" an amount of cold incoming seawater is considerably less than the latent heat released while the same equivalent amount of distillate condenses. Thus it is apparent that there is a practical limit to which a realistic regeneration device is able to reuse its heat. This is a function of the high and low temperature heat reservoirs, and the temperature drop necessary for heat transfer between the evaporator and condenser. Thus, the highest possible GOR any distillation device is able to feasibly operate at, may be expressed as

(3)11 MdT^T^.AT)-

Where Mr is the mass flow ratio between the cold incoming water (Ms) and the distillate (Md) and is defined using the first law of thermodynamics and humid air relationships.

____________________________________________ С pm *(Тк» ~ AT ~ Г„и )_______________________________________

I Cpa ■ {Thli — )- Сr ■ [g7to ■ (Тш )~ ПГрри ‘(Tpau )]+ hfgO^tJbq, 4J)

Ц 07 ha ~ .

Here Cpa and Cpw are the specific heats of the air and the water respectively and represents the specific humidity.

It is important to remember that this relationship represents the maximum efficiency attainable in a realistic regenerative distillation process and gives an indication as to how close a given device is operating with respect to the maximum possible. Thot and Tcoid are defined by the environment, and temperature drop — AT is defined by the size and quality
of the heat transfer surfaces. In effect therefore, one may “pre-define” the theoretical efficiency limitations of any given regenerative distillation device simply by defining these three factors. The upper and lower temperature limits are generally defined by the environment and AT depends on how much investment is made in the size and quality of the heat transfer surfaces. The goal of this research project has been to determine, by using CFD tools, how to promote the most effective natural convection so as to allow the device to approach this optimal efficiency at the given operating parameters.

Intermediate model: flow distribution in a fin-and-tube absorber

In this subsection, the flow distribution through an absorber is analysed. The absorber has 327 riser tubs with a diameter of 6 mm and a length of 1m which are connected to two manifolds with a diameter of фм. The riser tubs are mounted with no separation (no fins) between them. Therefore, the width of the absorber is in the range of 2 m. The risers will be numbered from 1 upwards, where the riser 1 is the one next to the inlet. The absorber has a z-configuration: the water enters the absorber through one ending of the inlet-manifold and is distributed through the different risers; after passing through the risers, water is collected in the outlet-manifold, and exists the absorber from the ending closer to the riser number 327. Investigated is the flow distribution in the risers for different diameters of the manifold, фм, and for two different total mass flow rate through the whole absorber: 25 l/hrn2 and 50 , using water as thermal fluid. Results are shown in figure 2. Given are the mass flow rates through each riser і, йц. These values are normalised by the ideal flow rate m0 which will result from a uniform distribution of the fluid flow through the riser, i. e fn0 = m/n, where n is the number of riser tubs.

iii= 25 1/hm2 m= 50 1/hm2

Figure 2: Numerical results with the intermediate model: flow distribution through a absorber with a z-conflguration, with two manifolds of a diameter of and 327 riser tubs (riser 1 = riser next to the inlet). Given is the mass flow rate through each riser i, , normalised by tha mass flow rate for a uniform distribution of the flow, . Two different total mass

flow rates, m, are analysed: a) m = 25 l/hrn2; b) .

Studies like this are of major importance in order to assure an appropriate flow distribution through the absorbers in order to obtain best thermal performance of the collector, [20].