Category Archives: EuroSun2008-10

Performance Parameters

The COP of the heat pump was determined by:

COP = Q^ (1)

W

comp

where Wcomp is the true power input to the compressor measured directly by the wattmeter, and Qc was determined by:

Qc = mr (h2- h) (2)

Where mr is the mass flow rate of the refrigerant, and h2 and h3 are the enthalpies at the inlet and outlet of the condenser, on the refrigerant side. Enthalpy is a function of both pressure and temperature, which are measured at each point throughout the cycle.

The natural convection flow rate, mNC can be determined by performing an energy balance across the condenser. Assuming the heat lost to the environment is negligible, as the condenser was well insulated, the energy balance is expressed as:

Qc = mr (h2 — h3) = mNCCp (T8 — T7 ) (3)

image183 Подпись: (4)

where Cp is the specific heat of the water. Rearranging Eq. 3, the natural convection flow rate is then expressed as:

2. Results and Discussion

An analysis was conducted on the operation of an ISAHP with a varying temperature input. The mains water temperature, and therefore initial storage tank temperature was approximately 20°C. Figure 3 shows the power delivered to the glycol through the heaters, as well as; heat transferred to the refrigerant through the evaporator; heat rejected to the natural convection loop through the condenser; and power consumed by the compressor.

image184

Fig. 3. Variation of heat transfer rates, and compressor input power throughout the test

As shown in Figure 3, the heat transferred from the glycol through the evaporator exceeds the heat delivered to the glycol through the auxiliary heaters. This is due to a large amount of glycol in the auxiliary heaters which are initially at a higher temperature. The heat pump unit extracts the extra energy until the heat transferred through the evaporator matches the heat input to the glycol through the heaters. Finally, the heat transfer through the evaporator is observed to lag the heater input, again due to the volume of glycol in the auxiliary heaters.

A temperature probe installed within the storage tank, with thermocouples spaced 0.15 m apart, was used to determine the level of stratification of the tank. Figure 4 shows the temperatures measured in the tank throughout the test, with T11 being the temperature measured at the top of the tank, and T20 measured at the bottom. During the test the tank was observed to stratify well, and throughout the test the temperature at the bottom of the tank, T20, remains constant at approximately 21°C. Therefore, over the course of the test, the temperature of the water entering the natural convection heat exchanger/condenser was effectively constant at 21°C.

The natural convection flow rate was also calculated over the length of the test. Figure 5 shows the natural convection flow rate of water through the condenser, plotted with the temperature of glycol delivered to the evaporator on the same graph. Consistent with the previous constant temperature tests, it was observed that higher glycol temperatures delivered to the evaporator increased the flow rate, and an increase in tank temperature decrease the flow rate. This resulted in the natural convection flow rate remaining constant for the first half of the test, and then decreasing as the glycol temperature declined, and tank temperature continued to increase.

In order to compare the dynamic operation of the heat pump with the steady-state model, the measured compressor power and heat pump COP were plotted with the simulated results on the same figure. To assist in validating the computer model, the simulation was run under the same conditions as the experiment. The model predicted the power consumed by the compressor, the evaporator and condenser heat transfer rates, and the COP of the system. Figure 6 shows the heat

transfer rates through the condenser and evaporator and Figure 7 shows the COP and compressor power curves throughout the test. Both simulated and experimental values are displayed.

image185

image186

Fig. 5. Natural convection flow rate of water through the condenser

 

image187

COP values measured throughout the test ranged from 2.4 to 3.2, and the measured compressor power ranged from 484 to 635 W. The amount of heat transferred through the condenser to the natural convection loop ranged from 1300 to 2000 kW.

Similar to the previous tests [13] run at constant glycol temperatures, the simulation program was found to overestimate the COP and heat transferred through both heat exchangers, but predicted the compressor power consumption reasonably well. Both heat transfer rates, and the COP, were over-predicted by the simulation by approximately 12%. These results were previously attributed to an over estimation of the heat exchanger effectiveness values. Neglecting this offset, the simulation, based on a steady-state governing equations, predicts the trend of the heat pump system well.

3. Future Work

Further testing is to be carried out, investigating wider ranges of temperatures and operating conditions. New heat exchanger relationships will be derived to better predict the effectiveness values, which will bring the model in to better agreement with the actual results. With the model refined, full year simulations in TRNSYS will be carried out to determine seasonal solar fraction values for the ISAHP system.

4. Conclusion

The experimental results matched well with the simulated results for the compressor power input, but the simulation over predicted the performance on the system. The compressor power input ranged from 484 — 635 W, and the COP of the system ranged from 2.4 — 3.2 over the duration of the test. The computer model predicted the dynamic operation of the system well, except for a 12% overestimation of performance due to the model’s effectiveness relationships. New seasonal solar fraction values and life cycle cost numbers will be calculated once the full year simulations are completed in TRNSYS.

3. Acknowledgments

Support for this work was provided by the Solar Buildings Research Network of Canada, the Ontario Graduate Scholarship Program, and the Natural Science and Engineering Research Council of Canada.

References

[1] NRCan, (2006). Energy Use Data Handbook, Natural Resources Canada.

[2] C. Aguilar, D. J. White, D. L. Ryan, (2005). Domestic Water Heating and Water Heater Energy Consumption in Canada, Canadian Building Energy End-Use Data and Analysis Centre

[3] G. A. Freeman, (1997). Indirect Solar-Assisted Heat Pumps for Application in the Canadian Environment, Masters Thesis, Department of Mechanical Engineering, Queen’s University.

[4] G. A. Freeman, S. J. Harrison, (1997). Solar Assisted Heat Pump Hot Water Heaters for the Canadian Environment, Proceedings of 1997 SESCI Conference, Vancouver, BC.

[5] K. Chaturvedi, J. Y. Shen, (1982). Analysis of Two-Phase Flow Solar Collectors with Application to Heat Pumps, Journal of Solar Energy Engineering, Vol. 104, 358

[6] Morrison, G. L, (1994). Simulation of Packaged Solar Heat-Pump Water Heaters, Solar Energy, Vol. 53, 249

[7] P. Sporn, E. R. Ambrose, (1955). The Heat Pump and Solar Energy, Proceedings of the World Symposium on Applied Solar Energy, Phoenix, Ariz.

[8] Chaturvedi, K., Abazeri, M, (1987). Transient Simulation of a Capacity-Modulated Direct-Expansion, Solar-Assisted Heat Pump, Journal of Solar Energy, Vol. 39, 441

[9] B. J. Huang, J. P. Chyng, (1998). Integral-Type Solar-Assisted Heat Pump Water Heater, Journal of Renewable Energy, Vol. 16, 731

[10] B. J. Huang, C. P. Lee, (2003). Long-Term Performance of Solar-Assisted Heat Pump Water Heater, Journal of Renewable Energy, Vol. 29, 633

[11] J. M. Purdy, S. J. Harrison, P.H. Oosthuizen, (1998). Compact Heat Exchanger Evaluation for Natural Convection Applications, Proceedings of the 11th IHTC Heat Transfer 1998, Korea, Vol. 6, 305

[12] University of Wisconsin, Solar Energy Laboratory, (2006). TRNSYS: A Transient Simulation Program, Madison

[13] A. Bridgeman, S. J. Harrison, (2008). Preliminary Experimental Evaluation of Indirect Solar Assisted Heat Pump Systems, Proceedings for The 3rd Annual Canadian Solar Buildings Conference, Fredericton, NB

[14] A. H. Fanney, S. A. Klein, (1988). Thermal Performance Comparisons for Solar Hot Water Systems Subjected to Various Collector Array Flow Rates, Proceedings of Intersol 85, Montreal, QC.

Insulation of the storage tank

The typical thickness of a PUR storage insulation is in the range of 10 to 12 cm [5]. Within this study the thickness has been varied between 7.5 cm and 17.5 cm (base case 15 cm). The results of the optimisation process for these variations are shown in Figure 3 (d) and Table 1.

The optimisation with 7.5 cm of insulation leads to a system with a storage device capacity of 0.77 m3 and a solar collector area of 12 m2. Thicker storage insulation leads to an improvement of the thermal behaviour of the system. Therefore with the same dimensioning and investment cost but thicker insulation a dot in Figure 3 would shift towards higher primary energy savings and simultaneously towards smaller additional costs (due to a reduction of running costs). Thus, the optimisation curves in Figure 3 (d) shift down and flatten with an enlargement in insulation. A solely parallel movement of the curves to smaller additional costs would mean that the point of intersection of the tangent would move left, whereas a flattening of the curves would lead to a point of intersection with higher primary energy savings. Figure 3 (d) shows that the optimised

system configurations shift towards smaller primary energy savings and Table 1 indicates that the resulting dimensions are getting smaller. This means that the reduction of additional cost dominates the increasing primary energy savings. The bigger the thickness of insulation gets the smaller the distance between the resulting curves become. With an insulation thickness of 17.5 cm nearly no difference is apparent to the base case. A cross check taking into account increasing investment costs with thicker insulation showed that the results summarised in Table 1 change insignificantly.

Considerations on mobility

The mobility of a solar thermal plant depends partly on tank size. The tank sizing for Case 2 (600m3) starts to be over the limit of going from factory built to site built, at least as a single tank. Up to a certain size, however, this can be arranged with 2-3 smaller tanks. Based on a telephone interview with a storage tank manufacturer this solution may also be cheaper up to 500-700 m3.

The cost of moving the panel array is unknown. Nobody has reportedly done this or documented the costs of doing it. The collector manufacturers for large arrays do not give out the information of their installation costs nor do they deliver collectors only; they sell only whole solar loop installations.

Some assumptions can be made based on a few documented projects. One german project reports [13] panel array installation costs to 60 EUR/m2 (although crane and transport not separated). Assuming the same 60 EUR/m2 for re-installation, and that dismantling and transport would be half of that, we come

to a rough estimation of 90 EUR/m2 which is about 20-25% of the initial investment. Making an estimation from the feasibility graphs for Case 1 and Case 2 this means roughly 3-6 years of prolonged payback time. It seems that with the presented feasible payback periods and current energy prices there is no room for added costs from moving the solar plant within its lifetime.

3. Conclusion

A feasibility study to find out general boundary conditions for combining solar thermal plants with small scale district heating networks under north European conditions was done. Even though the scenario was such that we connect solar to an existing plant, the conclusions made based on the results should be valid also for new plants with reasonable accuracy.

Cost breakdown:

The most costly part of the system is the collector array. The fixed costs play a smaller and smaller part in the total cost with increased system size, resulting in lower specific costs and thus improved economic viability, assuming the saved fuels are the same. The tank is the second largest investment. Other costs are relatively small compared to these, and only marginally affect the economic viability.

Pellets as fuel:

In the considered range 0.5-2 MWth, if the load is mainly space heating (Case 1) the results would imply that with todays pellet prices it will be difficult to find a feasible payback period during the plants estimated lifetime in the studied climate. Even with annual price increase rates of 5-10 % the investment seems difficult to justify economically without subsidies. For Case 2 we can find feasible payback periods shorter than the estimated lifetime of the system, but based on the studied plant portfolio, this kind of load profiles seem to be exceptions.

Oil as fuel:

If the replaced fuel is oil, then the feasible payback periods look different from the case of pellets. For Case 1 they are between 10-20 years and for Case 2 between 7-15 years depending on country and scenario.

Plant mobility:

The rough estimation made about dismantling and moving a plant implies that with current fuel prices the plant payback period is prolonged by 3-6 years. This cannot be justified regarding the feasible payback periods of the whole plant in general.

Acknowledgements

We are grateful to VAPO Oy for collaboration and financing.

References

[1] Streicher W. et al., Solarunterstutzte Warmenetze, Endbericht zum gleichnamigen Projekt in der Forschungsausschreibung „Haus der Zukunft“ im Auftrag des BMVIT, Technischen Universitat Graz (2002).

[2] Hahne E. et al., Solar unterstutzte Nahwarmeversorgung mit und ohne Langzeit-Warmespeicher „Forschungsbericht zum BMBF-Vorhaben“, Universitat Stuttgart (1998).

1st International Congress on Heating, Cooling, and Buildings — 7th to 10th October, Lisbon — Portugal /

[3] Hahne E. et al., Solar unterstutzte Nahwarmeversorgung mit und ohne Langzeit-Warmespeicher „Forschungsbericht zum BMBF-Vorhaben“, Universitat Stuttgart (2003).

[4] Muller-Steinhagen H. et al., Solar unterstutzte Nahwarmeversorgung und Langzeit-Warmespeicher „Forschungsbericht zum BMBF-Vorhaben“, Universitat Stuttgart (2005).

[5] Schmidt T., Mangold D., Status der Solaren Nahwarme in Deutschland, Universitat Stuttgart (2003 Conference presentation at „Solaren Kombianlagen fur Mehrfamilienhauser im europaischen Vergleich“).

[6] Holter C., Solarenergie & Biomasse, eine Erfolgsstory, S. O.L. I.D. Gmbh (2005 Conference presentation at „Mitteleuropaische Biomassekonferenz“).

[7] Calminder B. et al., Medelstora solvarmeanlaggningar — En utvardering av medelstora solvarmeanlaggningar utforda under perioden 1993-2000, K-Konsult Energi Stockholm AB (2002).

[8] Kovacs et al., Solenergi i industriell processvarme — En forstudie av svenska mojligheter, SP-Rapport 2003:16 (2003).

[9] Lundgren J., Hermansson R., Solar Assisted Small-Scale Biomass District Heating System in the Northern Part of Sweden, International Journal of Green Energy (2004).

[10] Dalenback J-O., Solar thermal market development in Europe, (Conference presentation at Northsun 2005).

[11] Dalenback J-O., Large-scale solar heating systems — a challenge for Europe, (Conference presentation at Gleisdorf Solar 2004).

[12] Isaksson et al., Report on technical investigations of large solar thermal systems, NEGST report WP2.D5 (2007).

[13] Reuss M., ZAE Bayern, Conference presentation at “Fachforum Solarstadt Munchen 2006”.

Heat exchange with n+1 and n-1 elements

image222 image223 image224

We consider that only conduction plays an important role in the heat transfer between neighbor elements. Energies transferred to the inferior element and from the superior element within a time step (p) are:

3.1 Energy input from the electrical heater

The auxiliary electrical heater has a 2kW power. Experiments showed that during heating water in the tank is mixed up in the entire volume just above the heater. We consider in this model that all elements are brought at the same temperature after one time step of electrical heating. Thus, the set-up temperature is reached by all elements above the heater in the same time, after one time step or more. If during a time step of heating the hot water temperature might rise above the set-up temperature, the electrical heater will stop even if the time step is not over.

Evaluation of virtual case studies

Out of the large number of virtual case studies, a handy number of standard system configurations, which work best under different conditions, are identified. Based on these, the industry partners will provide consistent package solutions. These will enable planers and independent craftsmen to install reliable systems. The economical and ecological rating of the virtual case studies will also allow identifying the most promising markets, where systems are yet at the edge of economical breakeven point or beyond. Last but not least the results of the virtual case studies will be made available online with an easy to handle web-based tool, which can query it under different aspects.

3.3. Training on package solutions

Special training courses for solar thermal installers on standard system configurations and package solutions will be prepared and 15 pilot courses will be evaluated. Target group are (solar thermal) installers, because the goal of the packaged solutions is to avoid the need of engineering.

3.4. Dissemination, communication and training

Tailored dissemination, communication and training plans were elaborated to reach the different key actors. They include besides the presentation of results at relevant conferences and trade fairs addressing a wider audience (i) the dissemination of both the elaborated brochure and the online tool to query the virtual study cases towards professional groups (HVAC planners, architects, engineers, building industry), through their interest groups and associations (e. g. ESTIF, ECTP, chambers), where possible on the occasion of annual meetings or in synergy with related national and international projects, (ii) the provision of information and advice to (national) authorities on the potential of Solar Combi+ with the aim to include it in support programmes and (iii) the approach of local authorities in promising regions promoting pilot installations. Finally, information through public media in the most promising regions should give an important push to market entry. On the website all public deliverables will be available for download and most attention is given to the integration of the webpage in the existing information network on solar heating in general and combined solar heating and cooling in special.

4. Market analysis

Analysis of Results

In spite of an early settlement of the monitoring system, the STS has undergone a period of either reduced load and/or deficient operation, which has prevented the collection of a representative set of monitoring data enabling a thorough analysis to the system.

To the present, monitoring operations have suited particularly the detection of system faults, rather than evaluation of system behavior and performance. Nevertheless, and beside a short fault examples list, the data acquired allows a preliminary analysis of system performance and trends.

2.1 Data system results

The actual monitoring period started in June 2007, when the building was becoming occupied and starting to be close to project conditions. The deployment of the STS started three months before, with the solar field working at limited capacity considering the low occupancy of the building (only a single 4 collectors group uncovered on each orientation).

2.1.1. System Deficiencies

After the deployment of the system, a number of fault situations were detected; either related to installation problems or inadequate load conditions.

The system proved to be hydraulically unbalanced in its East-facing collectors.

There were leaks observed in the pumping area.

After the leakages were fixed and pressure reset they reappeared (after 1 month) as well as in the East-facing collectors.

Design of Hot Water Heating System for Low-rise Apartment House

S. Kaneko1* , M. Udagawa1 and T. Kusunoki1

1 Kogakuin University, 1-24-2 Nishi-Shinjuku, Shinjuku-ku, Tokyo, 163-8677, JAPAN
* Corresponding Author, dm07014@ns. kogakuin. ac. jp

Abstract

Total performance of solar hot water heating(DHW) system for an apartment house of 10 housing units was examined using the detailed simulation with EESLISM [4]. The simulation was carried out to examine the appropriate collector area and the storage tank volume for the central type of DHW supply system. The result for an apartment house showed that the collector area of 30-50m2 is appropriate to expect the solar contribution of 37.1-91.8% while the solar contribution is strongly depending on DHW supply rate. The appropriate storage tank volume is 1.0-1.5m3 in the studied cases. The economical efficiency of DHW system showed that the equipment cost should be suppressed below 2 million JPY if expecting the pay back period of 10 years.

Keywords: Collector area, Storage tank volume, DHW heating load, Equipment cost

1. Introduction

In the past studies [1-3], the solar hot water heating(DHW) system for an apartment house of 10 housing units was simulated in order to examine the difference of tilt angles and azimuths of the collector. In this study, in order to find the relationship of collector area and storage tank volume of the DHW system, the simulation study was carried out. The suitable combination of the collector area and the storage tank volume with considering initial cost are examined by simulating the yearly performance of the solar DHW system for the apartment house.

Rise in energy prices

The simulations of the base case were calculated with a comparably small rise in energy prices of

1.3 Подпись: 0Подпись:Подпись:image254%/a for natural gas and 0.3 %/a over 20 years for electricity [6]. However the current development of prices in Germany amounts to 9.7 %/a (non inflation-adjusted) over the last seven years [7] which indicates a higher annual rate of growth.

Therefore the rise in energy prices has been varied between

1.3 Подпись: 12%/a and 7.5 %/a. Figure 3 (e) shows that as expected the lowest values of the objective function get smaller with a growing rise in energy prices which means that the heat generation costs per kWh become smaller. More surprisingly, the dimensions of each optimal system in terms of the underlying objective function and therefore also the primary energy savings stay constant, showing that the dimensioning does not depend on the rise in energy prices within the examined range. Figure 5 shows an extrapolation of the simulation results calculated with different rises in energy prices. The trend line derived from the calculated points indicates that the analysed solar heating system would be economically rewarding with a rise in energy prices of 9 %/a without any subsidies.

4.3 Subsumption of simulation results

Reducing the price of the solar collectors by 30 % improves the cost/benefit ratio by 21 %. The resulting optimal collector area increases by 3.5 m2, whereby the storage device capacity keeps unaltered. Using a high efficiency flat-plate collector instead of the initially defined model reduces the cost/benefit ratio by 8 %, whereas a low efficiency flat-plate collector increases the cost/benefit ratio by 13 % without having stronger impact on the dimensioning of the system. A 40 % reduction in the storage cost improves the cost/benefit ratio by 24 % again without changing the optimal dimensioning. The difference in the cost/benefit ratio between a system with 7.5 cm of storage insulation and a system with 17.5 cm of storage insulation amounts to 21 %, whereas the optimal system with the thickest isolation consists of a storage 120 litres larger than the thin isolated tank, connected to solar collectors that are 3.2 m2 smaller than the pendant with the thin isolated tank. Considering a rise in energy prices of 7.5 %/a instead of 1.3 %/a leads to a reduction of additional costs of 250 € /a with equal dimensioning parameters.

Comparison of the Thermal Performance of Different Working. Fluids in a Closed Two-phase Solar Water Heating Thermosyphon

A. Ordaz-Flores1, O. Garcfa-Valladares2*, V. H. Gomez2

1 Posgrado en Ingenieria (Energia), Universidad National Autonoma de Mexico, Privada Xochicalco s/n,

Temixco, Mor. 62580, Mexico

2 Centro de Investigation en Energia, Universidad National Autonoma de Mexico, Privada Xochicalco s/n,

Temixco, Mor. 62580, Mexico

* Corresponding author, ogv@cie. unam. mx
Abstract

A closed two-phase thermosyphon solar system was designed and built to produce hot water for sanitary purposes. The aim of this work is to compare the thermal performance of a two — phase closed thermosyphon using different phase change working fluids (acetone, R134a and R410A). The choice of using a closed two-phase thermosyphon, instead of a conven­tional solar water heating thermosyphons obeys to the some advantages as the lower freez­ing point of the two-phase system compared to water, and elimination of fueling, scaling and corrosion. Disadvantages of these systems are the higher cost because of the working fluid used and the additional coil heat exchanger; moreover, refrigerants reach high pressures. A witness conventional solar water heating system has being installed to compare its perform­ance versus that of the two-phase closed system. The two-phase system consists of a flat plate solar collector coupled to a thermotank by a continuous copper tubing in which the working fluid circulates. The working fluid evaporates in the collector and condensates in the thermotank transferring its latent heat to the water through a coil heat exchanger. The conventional thermosyphon system has the same characteristics (materials and dimensions), with the exception that it lacks the coil presented in the two-phase system. Data were col­lected from the two kind of solar water heating systems, operating simultaneously, and com­parisons of performance were made. Results show that the performance of the two-phase systems is strongly dependent on the load of the working fluid: an optimum point should be found. R134a and R410A show better performance than acetone. The two-phase closed sys­tem shows hardly any difference in performance (when working with both R134a and R410A) compared to the conventional solar water heating thermosyphon.

Keywords: acetone, test, R134a, R410A, phase change.

1. Introduction

The increasing interest of preserving the non-renewable resources has led to focus on sustainable growing, based mainly on using renewable energy. The use of renewable sources helps to save economical expenses, as well as to prevent the inherent environmental impact of conventional sources. Renewable energy sources are the Sun, biomass, hydrogen, wind, etc. The Sun leads to thermosolar and photovoltaic technologies, mainly.

The current paper has special interest in Solar Domestic Water Heating Systems (SDWHS). SDWHS permit to diminish the consumption of liquid gas and electricity, helping to reduce the quantity of pollutants expelled to the atmosphere. In 2004, Kalogirou [1] studied the environmental impact of energy utilisation and the potential benefits to swap conventional for solar assisted sys-

tems. He estimated that, for the case of solar water heating (one of the two most widely used re­newable energy) the savings would reach up to 80%. Hence, the importance of solar water heating.

For instance, in Mexico, the use of flat plate solar collectors to heat 500 L of daily water would yield savings of 433 kg/year of LP gas [2].

The most common currently available solar equipments to heat water are the thermosyphons in which the water is heated in a flat plate solar collector and stored in a thermotank. Active systems use a pump to circulate the water, while in passive systems the water circulates by the thermosy­phon effect. The water presented in the flat plate solar collector is heated by the Sun energy, so its density diminishes; the lower density of the water in the collector, compared to that of the thermo­tank makes the water to circulate: that is the thermosyphon effect. In direct systems, the water is heated in the collector; in indirect systems, some fluid is heated in the collector, and it transfers the energy to the water by means of a heat exchanger; in a closed system, the working fluid is sealed from the atmosphere, in an open system, the heat transfer fluid is in contact with the atmosphere. If the fluid changes its phase in the collector, the system is called a two-phase or a phase-change sys­tem.

The system studied in this paper is a passive, indirect, closed, two-phase system. This kind of sys­tem prevents problems like freezing, corrosion, scaling and fouling [3], which are presented in the conventional systems, increasing the life of the system.

In 1979, Soin et al. [3] described an experimental set up to evaluate the performance of a solar col­lector with a phase change working fluid. They used acetone and petroleum ether as working flu­ids, because of their high boiling and condensation heat transfer coefficient. They demonstrated that the collector efficiency increases linearly with liquid level.

In 1981, Schreyer [4] used a refrigerant, trichlorofluoromethane, to evaluate the energy recovery in a solar collector coupled to a heat exchanger, and the latter to a storage tank. The primary loop was passive and the secondary needed a recirculation pump. His system recovered up to 83% energy at low collector temperature difference.

Evaluation of R134a (among others) as replacing working fluids of ozone depletion promoting chlorofluorocarbons was made by Calm and Didion [5]. They concluded that there is no perfect fluid to prevent every environmental impact. R134a has a high latent heat of vaporization, does not contributes to ozone depletion but, yet low, does have impact on global warming.

Ong and Haider-E-Alahi [6] studied the performance of a heat pipe filled up with R134a, and found that the heat flux transferred increased with high refrigerant flow rates, high fill ratios and greater temperature difference between bath and condenser.

More recently, Hussein [7] studied a two-phase closed thermosyphon with the heat exchanger (condenser) in the solar collector; however, he did not mention the working fluid used. He carried out both experimental and numerical tests and set some dimensionless variables to determine ade­quate storage dimensions for the tank to improve the solar energy gain.

In 2005, Esen and Esen [8] studied a thermosyphon heat-pipe solar collector, to evaluate its ther­mal performance using three different working fluids, R134a, R407C and R410A. They found that the latter offered the highest solar energy collection.

In this work, refrigerants R134a and R410A were chosen due to their availability, low cost and small impact to environment. Acetone is also cheap and available, but it avoids the high pressures reached with the former ones; on the other hand, acetone is flammable.

2. Experiment

A water heating two-phase closed thermosyphon, using either R134a, R410A and acetone as work­ing fluids, and a conventional natural thermosyphon are compared simultaneously. Both systems have the same geometry, except for the coil presented in the two-phase system. The construction materials for the whole system are the same. Each collector has an absorption area of 1.62 m2 and the volume capacity of each thermotank is 160 L. The two-phase system consists of a flat plate solar collector coupled to a thermotank by a copper tubing circuit in which the working fluid circu­lates. A scheme of the systems is shown in Fig. 1.

Focusing on the fluid refrigerant behaviour, the solar collector is the evaporator of the system and the copper coil immersed in the thermotank is the condenser. The incoming solar radiation makes the temperature of the refrigerant in the collector to grow higher to reach the saturation liquid state. From this point, the working fluid starts to evaporate to reach the saturated vapour state and even the superheated vapour zone. As the refrigerant has a higher temperature than the water, the former donates its phase change latent heat to the latter and leaves the thermotank as sub-cooled liquid to come back to the solar collector to repeat the cycle.

image159

Fig. 1. Two-phase closed thermosyphon and conventional thermosyphon.

Refrigerant R134a is one of the replacing working fluids of chlorofluorocarbons since it does not contribute to ozone depletion. R134a evaporates at -26.1 °C at atmospheric pressure [9] with an enthalpy of vaporisation of 216.98 kJ/kg; its freezing point at this pressure is -101 °С.

Acetone (also known as propanone) is a colourless liquid, used mainly as solvent, for cleaning, or as a drying agent; is flammable, and should not be inhaled. At atmospheric pressure, it evaporates at 56.05 °С [10] with an enthalpy of vaporisation of 501.03 kJ/kg and a freezing point of -94.7 °С.

R410A is a mixture of refrigerants R32 and R125 (50% of the volume of each one), it is used in air conditioning as substitute of R22; it is not toxic and does not contribute to ozone depletion; its boiling point at atmospheric pressure is -52.7 °С; its enthalpy of vaporisation is 275.93 kJ/kg. The freezing point of R410A is not determined yet, but the freezing points of its components are -103°C for R125 and -136°C for R32, at atmospheric pressure [9].

The combination of boiling point temperature (the lower, the better) and heat of vaporization (the higher, the better) will show which of the fluids is more suitable for these operating conditions; other parameters as viscosity and pressure must also be considered.

The main disadvantage of R134a and R410A is that they reach high pressures; for instance, the pressure of these fluids at 50°C is 13.18 bar for R134a and 30.71 bar for R410A; their main advan­tage of the refrigerants is their low boiling points; that means that the heat transfer will start soon after the beginning of the test. Acetone does not have problems of pressure: at 50°C, it only reaches 0.81 bar; and its enthalpy of vaporisation is higher related to the refrigerants, but it lacks of a low boiling point at atmospheric pressure: 56.5°C.

The two-phase system was loaded up to 91% when operating with R134a, up to 83% when operat­ing with acetone, and up to 62% when operating with R410A. The systems were loaded differently because of the characteristic of the fluids and the difficulty to load refrigerants. On the other hand, acetone is very easy to load and permits to have better control.

Energy input from the solar loop

The solar heat exchanger is a cupper coil immerged at the bottom of the tank. One or more elements contain this exchanger as the user can decide. We consider that the heat provided by the

4

solar loop is transferred in a time step to one or more elements depending of their temperature. For example, if the last three elements have a lower temperature than the hot water produced by the solar exchanger than only those three elements will be heated. No influence is considered for the fourth element, during that time step.

The solar panel temperature is computed depending on the operation of the pump: if the pump is not running than

Tpanel(i + 1) = Tpanel(i) + •

panel

• Esun(i + [ • P-K1 • (Tpanel(i) — T0(i + 1))-K2 • (Tpanel(i) — T0(i + 1 ))2

image225 Подпись: (9)

And if the pump is running than

Where mpanel stands for the mass of water contained by the panels in kg, msolar stands for the water quantity (kg) flowing through the solar loop in a time step p, S panel is the active surface of the panels, Tho and Thi stand for the outlet and inlet temperatures of the exchanger, and T0 is the ambient temperature.

• (Tho(i + 1) — T0(i + 1))2 .• Spanel • P

Подпись: Esolar (i + 1) Подпись: Esun(i + 1) •P -K1 • (Tho(i + 1) - T0(i + V) -K2

Using the panel temperature, the control system can decide on the pump operation for the next time step, and finally we can compute the energy input of the solar loop if the pump is on.

(10)