With reference to Figure 1, in winter functioning if the internal air temperature is lower than 19°C, the control system activates the pump (b) and determines the inlet temperature by means of the Equation (4): if the water temperature within the tank is lower than that required, the three way valve system is activated (3) and the inlet flow rate is completely supplied by the auxiliary system at the required temperature. On the contrary, if the temperature within the tank is higher, the three way valve system (1) is activated and the water flow rate from the tank is mixed with a fraction of the recirculating flow rate returning from the radiant ceilings in such a way as to obtain the required temperature.

During summer, the generator inlet temperature is regulated based on the thermal power required by the radiant ceiling and of the temperature in the condenser of the water flow rate provided by the evaporative tower, hypothesised as 8°C higher than the wet bulb temperature of the external air. The dependence of these two temperatures on the power supplied by the absorption machine considered is shown in Figure 2, which highlights the operative limits of the chiller, for example the minimum functioning temperature of the generator is equal to 76.7°C. Moreover, in the hypothesis that the condenser temperature is 29°C, the minimum distributable power at the minimum functioning temperature of the generator results as being equal to 21.4 kW. For condenser temperatures of 30°C and 31°C, the minimum distributable powers at the minimum functioning temperature of the generator become 12.9 kW and 2.4 kW respectively. From these considerations it is possible to deduce that, on the basis of the temperature supplied to the condenser, acting upon the inlet temperature of the generator it is not always possible to regulate the absorption chiller in such a way as to provide the required refrigerating power. It is in fact possible to observe that powers greater than 12.9 kW can be distributed at the minimum functioning temperature of the generator only for condenser temperatures lower than 30°C. The summer control system is therefore capable of determining the required generator temperature based upon the water temperature in the condenser and upon the thermal power required by the building [7].

In order to guarantee the complete coverage of the refrigerant load required by the building, it is necessary to use an electrical heat pump which intervenes in situations in which the chiller is not capable of supplying the required refrigerating power. The auxiliary chiller intervenes each time that within the tank there is a lower temperature compared to that determined by the control system for the generator. The auxiliary heating system used for the winter period could increase the water temperature of the generator, yet is it not used as this solution does not permit the attainment of primary energy saving. The control strategy used for the summer period regulates the generator inlet temperature according to the following logic:

♦ If the water temperature in the tank is greater than that required by the generator, the three way valve system (1) is activated which mixes the load extracted from the tank with that exiting the generator, and the absorption machine distributes precisely the required load. If the inlet temperature in the tank is lower than that required by the generator, the remaining quota is supplied by the traditional auxiliary chiller.

♦ If the conditions are such that it is not possible to guarantee the effectiveness of the absorption chiller (temperature in the tank of less than 76.7 °C or required powers that are incompatible with the condenser inlet temperature), the refrigerator power is totally supplied by the auxiliary chiller.

The refrigerator flow rate supplies the radiant ceiling system only when the internal air temperatures exceeds 27°C, with the control system that activates pumps (c) and (d) of figure 1.

The refrigerant power requested by the environment is estimated with the relation successive based on the difference in temperature between the internal air and the reference value, set at 26°C:

Qcool = K2 ‘ (tIA — 26) (6)

A preliminary simulation campaign permitted the determination of the value of constant K2 which resulted as being equal to 47.35 kW°C-1:

The fifth step consists in analysing the performance data coming from the last step so that the user is able to quantify the economic gain given by the studied solar cooling system. For this purpose and in the framework of the EU Specific Support Action ROCOCO [5], a analysis method has been developed by the Spanish solar cooling engineering specialist, Aiguasol, in partnership with the other participants of this project to analyse technical results on the economical point of view. Numerous criteria such as investment, operation maintenance and energy prices based on existing experiences and cost reduction trends are used to make a very precise and interesting analysis. Several graphs such as the one presented in Figure 3 should permit to present specific ratios such as economical savings in primary energy in comparison with a reference system (electricity, gas for example). This powerful tool is still under development but could be adapted to the method with the agreement of Aiguasol.

For the final adaptation of Type 307 the values fNC and fDEI are determined using new performance data files. These have been prepared from laboratory test measurements of an EAW Wegracal SE 15 chiller carried out by ILK-Dresden. After normalising these data two different files have been prepared (PD1 and PD2). This difference is explained by Fig. 1 where fNC is shown for increasing hot water inlet temperatures tGi while chilled and cooling water inlet temperatures are constant. From the PD1 curve it is seen that the TRNSYS look-up approach interpolates stepwise between the neighbouring supporting points of the data file, i. e. the ‘tines’ of the solid curve. Since the supporting points in the performance file are not allowed to change, i. e. once the first values have been fixed the same values have to be used for all combinations, there is a slight difference to the measured values. Furthermore it can be seen that extrapolation to lower as well as to upper values is not possible. In these cases the TRNSYS simulation gets stuck at the lower or upper end (fNC = 0.89 or 1.19, respectively). This behaviour is important to know for users working with Type 107 based models, since TRNSYS neither interrupts the simulation nor writes a warning to the list file. This could result in a wrong plant dimensioning in the worst case and has to be avoided. For example, feeding the model with tGi = 70°C, tACi = 27°C, and tEi = 18,5°C will return a value of fNC « 0.9 from the look-up approach with PD1. In reality the fNC is considerably lower (cf. Fig. 1 for PD2 and additional independent data points). Thus, the solar cooling system would have been simulated too efficient for this operating condition.

Since the data reading and interpolation procedure in Type 307 is the same as in Type 107 (stepwise interpolation, no extrapolation) a trick has been applied to avoid the above mentioned problems. From a linear fit of the measured values a global slope is achieved for a certain group of independent variables (e. g. in Fig. 1 tGi = var., tACi, tEi = const.). The global slope is a result of the fit over the whole range 80°C < tGi < 95°C at fixed tACi/tEi conditions («fixed temperature lift, see section 2.4).

Obviously it is steady enough to allow extrapolation. This is confirmed by two additional independent points at lower driving temperatures which have not been used in the fit for the global slope. The new interpolation limits (e. g. in Fig. 1 fNCmin = 0.54 and fNC, max = 1.30) have to account for physical and/or safety limits. Up to now, for the EAW Wegracal SE 15 chiller they have been set to the known temperature limits of operating conditions where measurements are available

from ILK test data and from the plant at Ebner Solartechnik. Now the new PD2 performance file includes only the fNC and fDEI values at these borders (i. e. only two points per tGi/tACi/tEi combination). During simulation the interpolation — which is still done stepwise, but just between these end points — merges into a global interpolation.

Within the Heat Pump Programme (HPP) of the International Energy Agency (IEA) Annex 34 “Thermally Driven Heat Pumps for Heating and Cooling” has been started last year. As outlined in the letter of intentions of the Annex, “the main goal of this annex is to reduce the environmental impact of heating and cooling by the use of thermally driven heat pumps”.

Taking as a base the results of the finished Annex 24 and cooperating with the IEA Solar Heating and Cooling Program Task 38 “Solar Air Conditioning and Refrigeration” the economical quantification, the environmental implications and energy performance of integrated thermally driven heat pumps in cooling and heating systems in a range of climates, countries and applications are going to be studied. The applications are mainly focused on the use of the heat pumps in domestic and small commercial buildings, although some industrial applications will be taken into account.

• Larger DHWS-systems with cooling

• Small Heat Pumps for domestic heating and cooling

• Small Abs/Ads Chillers on heat form

• Small Abs/Ads Chillers on solar heat

• Industrial processes using waste heat

As can be seen in the figure 1 [2] or in the obtained results of the ECOHEATCOOL Project [3,4], heating needs in Europe are usually more important than cooling needs. Therfore, reducing the primary energy amount with a more efficient machine during the months when heating is needed, is a good scenario for the development of new applications for thermally driven heat pumps that would have potential for saving primary energy (fossil fuel) and reducing its adverse impact of classical fuels on global warming and environmental pollution. (i. e.: as heat source renewable energies as solar energy, waste heat, process heat or district heating could be used [5])

In [6] a comparison of the energy demand for heating and cooling for single house sited in the middle of Spain, a place where cooling is needed during the summer season, is exposed. Although this example is placed in one of the hottest zones of Europe, the heating demand is considerably higher than the cooling one — and the example therefore emphasises the importance of savings in heating for nearly all climates in Europe.

Obtained results will help in the identification of the applications that are suitable of being optimized from the economical and environmental point of view, as well as the most promising countries, markets, barriers — and solutions to eliminate the latter — in order to help launching the technology in all possible climates.

2. Structure of the project

More than fourteen research and development groups and producers from nowadays 6 different countries (Austria, Canada, Germany, Italy, Netherlands and USA) are joining efforts to develop this project. It is divided in 5 different tasks as is shown in the figure 2.

Fraunhofer ISE (Germany) has the responsibility of leading the Annex, helped by the leaders of each one of the Tasks.

Respectively, the responsibles are:

• ECN (Netherlands)

• Arsenal Research (Austria)

• CNR Messina (Italy)

• Eurac Research (Italy)

• TU Berlin (Germany)

R. T. Montilla1* and A. Abanades2

1 SONNEGEX S. L., 31, Sta. Cruz de Marcenado St., 28015 — Madrid — Spam
2 Dpto. Termotecnia, ETSII, Universidad Politecnica de Madrid, Madrid — Spain
Corresponding Author, rmontilla@icai. es

Abstract

This report draws the conclusions of an economic analysis carried out to study the viability of domestic solar cooling systems in southern Spain. Due to the reduced range of small-powered absorption chillers, a 32kW machine has been selected and analysed in three different housing typologies: apartments, terraced houses, and detached houses. The results of this analysis show, as expected, that solar cooling is not already economically acceptable for domestic use, with a payback period longer than 25 years, mainly because of the high cost of the solar field. By considering public subsidies to the installation of this component, and its complementary use to pre-heat sanitary hot water and central heating water, the payback period decreases to less than five years in the most optimistic scenario.

Keywords: solar, absorption, cooling, Spain, viability

1. Introduction

Parallel to the improvement of life conditions, air conditioning devices have become almost a basic need in houses of developed countries. The extra energy demand that cooling entails has not meant a problem while electricity has been cheap. However, linked to the evolution of fuel price, the cost of electricity has started a raising period with no upper limit. This situation forces the industry to optimize the efficiency of all the energy consumer processes.

Electric vapour compression chillers are the most commonly used in domestic applications, representing almost 95% of all the domestic cooling systems. Despite the fact that these machines have a high EER, mainly caused by a high research and development grade, their electricity demand is already significant. Regarding energetic efficiency criteria, the use of solar absorption chillers, which could obtain all the needed thermal energy from a solar collector field, could represent an important decrease of this electricity demand.

With the objective of reducing energy consumption and its consequent impact on the environment, installation of solar cooling systems has been considered and analysed from an economic point of view in southern Spain.

A. Millioud1* and J. Dreyer2

1 NEP Europe GmbH, Badenerstrasse 18, 8004 Zurich, Switzerland
2 NEP Solar Pty Ltd, Unit 21, 14 Jubilee Avenue, Warriewood, NSW 2102, Australia
* Corresponding Author, antoine. millioud@newenergvpartners. com

Abstract

A small aperture medium temperature solar parabolic collector system has been developed. Its primary application is to drive high temperature and high efficiency thermal refrigeration systems such as double stage chillers; a combination which is shown to achieve the best economic performance in solar cooling. The collector system makes use of composite reflector carriers. The carrier consists of an efficient sandwich structure of fibre-reinforced polymeric skins and a specialised core material, onto which an aluminium reflector is bonded. Multiple design iterations have lead to a structure which is form stable, accurate and cost effective. Finite element analysis and photogrammetric testing on the prototype were conducted to confirm the optical precision of the solution. A 50m2 aperture pilot solar field was installed and is currently undergoing thermal testing.

Keywords: Solar cooling, parabolic trough collector, polymer carrier, absorption chiller, composite materials, polymeric materials, reflector, concentrating solar collector

1. Introduction

Air-conditioning loads in cities around the Sun Belt regions of the earth are increasing leading to rising consumption of electricity with associated green house gas emissions and increased pressure at peak hours on congested municipal electricity distribution networks. The deployment of solar heat driven absorption chillers is an effective way to address these issues. In particular, double stage absorption chillers driven by parabolic trough solar collectors custom-designed for solar cooling applications, is a financially viable approach to solar cooling and in general to the substitution of grid electricity with solar energy. Solar cooling projects are thought to be most viable where they are integrated in large building air-conditioning systems and where they take the role of peaking chillers. Ideally such a system is implemented as part of an upgrade or new build. Solar thermal systems can easily be integrated into existing building thermal energy systems and can accommodate thermal storages. The integration solution has to be optimised such that the collected solar energy is fully utilised, if necessary through cascading (from high temperature requirements down to lower temperature requirements). The concentrating solar collection system has to be designed such that it can be installed on roofs of varying sizes and geometries. This flexibility is required at minimal additional costs. Furthermore concentration factors, absorber design and choice of materials must be selected in function of the desired output temperature of 150-200°C.

The results of experimental testing at varying supply inlet temperature and inlet relative humidity for different regeneration temperatures at constant regeneration humidity are shown in the following graphs.

Figure 4. Moisture removal in supply air stream for Figure 5. Moisture removal in supply air stream for varying supply inlet humidity and temperature at 50°C varying supply inlet humidity and temperature at 80°C regeneration temperature. regeneration temperature.

Figure 4 shows the moisture removal from the supply air stream at varying relative humidity and temperature for a regeneration temperature of 50 °С. It is obvious that the moisture removal increases with both increasing relative humidity and temperature of the supply air. The maximum moisture removal for supply inlet conditions of 40°C/95% RH was measured at 17 g/kg d. a.. In this point the driving temperature difference is only 10K between supply and regeneration air, however the difference in absolute humidity between supply and regeneration air is 40 g/kg d. a..

The same parameters as in Figure 4 are shown in Figure 5 for a regeneration temperature of 80°C. It can be observed that the moisture removal at 80°C regeneration temperature is generally greater than at 50°C regeneration temperature. A larger moisture removal capacity is clearly noticeable in Figure 5 at high supply inlet humidity and supply inlet temperatures of 30 and 40°C. However, at supply inlet temperatures of 10 and 20°C the difference in moisture removal between 50 and 80°C regeneration temperature is relatively small. In fact, the difference in moisture removal between 50 and 80°C regeneration temperature for supply inlet temperatures between 10 and 30°C and supply inlet relative humidity between 20 and 50% is smaller than 1 g/kg d. a. This shows that a regeneration temperature of 50 °C can almost achieve the same moisture removal at lower inlet humidity supply air.

Dae-Young Lee

Energy Mechanics Research Center, Korea Institute of Science and Technology, Seoul 136-791, KOREA

*E-mail : ldv@,kist. re. kr

Abstract

The cooling capacity and the energy efficiency of the desiccant cooling system can be improved by incorporating a regenerative evaporative cooler instead of a direct evaporative cooler. The regenerative evaporative cooler is to cool a stream of air using the evaporative cooling effect without an increase in the humidity ratio. In this study, a prototype of the desiccant cooling system incorporating a regenerative evaporative cooler was designed, fabricated and tested for the performance evaluation. To this purpose, two important components, i. e., the regenerative evaporative cooler and the desiccant rotor were developed and assembled into the system. The regenerative evaporative cooler was built by compiling multiple pairs of aluminium plates and fins and the desiccant rotor was fabricated using polymeric desiccant. The exterior dimension of the completed prototype is 700(W) x 800(D) x 1,900 mm(H). The prototype was tested at the ARI condition (indoor: 27oC, 50%RH, outdoor: 35oC, 40%RH) for performance evaluation. With the regeneration air temperature of 60oC and the ventilation ratio of 0.3, the cooling capacity was measured as 4.4 kW and COP as 0.76.

Keyword: desiccant cooling system, regenerative evaporative cooler, prototype

1. Introduction

The desiccant cooling system has been investigated for years as an alternative option to conventional vapor compression cooling systems[1-4]. In a desiccant cooling system, air is dehumidified passing through a desiccant rotor and then is cooled in a series of a sensible heat exchanger and an evaporative cooler to achieve the desired air condition. This system works without CFC’s or other similar ozone-depleting chemicals but only uses water as the working fluid. It provides cooling by the use of thermal energy instead of electricity contributing to peak demand reduction of the electricity. The solar energy, the waste heat from plants, etc. can be applied as the source of the thermal energy. In such cases, the desiccant cooling technology can greatly contribute to energy saving and CO2 reduction by using the energy unutilized otherwise and thus reducing the consumption of fossil fuels.

In this study, a prototype of the desiccant cooling system incorporating a regenerative evaporative cooler was designed, fabricated and tested for the performance evaluation. The prototype was designed to show 4 kW cooling capacity at the ARI condition (indoor: 27oC, 50%RH, outdoor: 35oC, 40%RH) with the hot water supply of 70oC as the heat source. To this purpose, two important components, i. e., the regenerative evaporative cooler and the desiccant rotor were developed and assembled into the system. The regenerative evaporative cooler was built by

compiling multiple pairs of aluminium plates and fins and the desiccant rotor was fabricated using the polymeric desiccant, SDP.

The considered building was designed following the typical requirements for sustainable bioclimatic constructions to adapt the building to its surrounding environment (e. g., study of on­site climatic conditions, correct building orientation, adapted solar protection, rational use of natural resources, environmental integration). Being located in the Tabernas desert in Almeria, the building will have to deal with rather drastic weather conditions, with very high temperatures observed in summer and relatively low ones at night and in winter, which involves a high energy demand both in summer and winter in order to maintain satisfying indoor thermal conditions. However, thanks to bioclimatic design and integration of the various passive heating and cooling techniques, the total energy demand for this building is expected to be as low as 35 kWh/m2.

Being part of a national R&D program, this building was designed to be a test platform for the integration of various active and passive HVAC techniques and systems. As such it integrates both active and passive solar heating and cooling systems alongside with conventional HVAC installations. The purpose was to make it possible to compare the performance of unconventional systems with regard to conventional ones in a building continually occupied throughout the year.

Given the particular location of the building in one of the hottest and driest region of Europe, efforts were centred on cooling techniques. With regard to thermal comfort in summer, both passive and active solar cooling techniques are combined so that indoor thermal comfort can be ensured while minimizing energy expense, even at high ambient temperatures. As a result, the following systems were integrated:

The choice to develop a direct air-cooled chiller was ambitious. Air has a poor heat transfer coefficient that raises the required driving temperature, increases the air heat exchanger area and requires a scrupulous thermodynamic design. Nevertheless this choice was based on the careful judgment of the issues discussed above, applied to the Southern European climatic conditions. The vision behind Ao Sol’s development endeavour can be synthesized as follows:

The chiller must be cost-effective in the manufacturing in terms of components used and assembling procedures. The chiller must be direct-cooled, without any dry or wet cooling tower. This saves investment cost, room to place the tower, operation costs for a pump (to bring the heat from the chiller to the tower), a strong fan, water consumption and chemical water treatment, and tower maintenance. The chiller must be developed firstly for fan coil use. The use with radiant ceilings is thermodynamically less demanding and can be achieved without any further development. The other way around requires a rather new development. The control system must allow a very reduced part-load operation, thus reducing startup/shutdown losses. The chiller must be placed outside regardless of the climate, in order to save valuable space indoor.

To confirm the economic potential of the novel chiller an analysis was performed using Excel software to compare the novel ammonia chiller with a standard compression chiller. The comparison was set up for a 200 m2 residential building in Lisbon, Portugal. The data concerning heating, cooling and (DHW) demand were calculated by means of commercial software [7]. Ao Sol’s CPC collectors were simulated with internal software [8] for the calculation of the solar input leading to the solar fraction. Both solar energy yield and house energy demand were averaged on a monthly base. Efficiencies were taken into consideration for the chillers, the back-up gas burner and the electric grid.

The price for the ammonia chiller is calculated on the base of a company’s internal cost breakdown for the full production phase; the compression chiller price was scaled up from a market product of 5 kW [9]. Installation costs were assumed equal for both chillers, as well as the cost for hot water storage tank and gas boiler. Finally, the cost for solar collectors was set according to Ao Sol’s selling price for their new CPC MAXI collector.

Operation parameters, such as gas [10] and power [11] prices and their respective escalations have been set for the dynamic economic calculation. The prices are averages for Portugal and the price escalation formulated by official sites. Table 1 gives the overview.

The economics of the two systems have been compared by means of a differential present net value (Table 2). The solar fraction was set at 67 % by choosing the appropriate amount of solar collectors according to the chillers COP and collector efficiency. This value for the solar coverage was chosen in order to achieve energy savings compared to the reference system connected to the grid [12]. The total investment includes all parameters listed in Table 1. The gas consumption relates to heating and DHW supply by the gas boiler, as well as driving heat for the ammonia chiller.

Several items summed up to the parasitic power consumption, which turned out to be not negligible for the absorption chiller: solar, hot water, chilled water and solution pump, gas burner and chiller control, and cooling fan were taken into account for the absorption chiller. Gas burner control was added to the compressor consumption for the compression chiller. The parasitic consumption has been weighted with the operation hours of each item.

Investment and cumulated yearly operation cost were then calculated for each chiller and summed up in a net present value for a period of 20 years, which represents the technical lifetime of a compression chiller. It must be noted that an absorption chiller has a rather longer lifetime of 25 years, which was not taken into account here. The difference of the cumulated investment and yearly operation cost for both systems tells which system provides the best economic performance over its operational lifetime. As it can be seen, the ammonia chiller causes during its lifetime roughly half of the expenses caused by the compression chiller. The dynamic payback time of the ammonia chiller, i. e. the time span before the higher investment is recovered, occurs during the 8th year (Figure 6).

4. Conclusion

Ao Sol’s chiller development is based on a strong product displaying superiorly in terms of manufacturing effort, maintenance-free reliability, low overall operating cost and compact overall

size. The combination with Ao Sol’s own CPC-MAXI collectors proves to be economically appealing.

The experimental tests already gave an impression of the capacity of the prototype reaching a cooling capacity of 7.8 kW and a COP of 0.53. So, it can be said that most components seem to have the right size or even to be oversized, as e. g. the air heat exchangers and the rectifier. The solution pump worked well and performed the full flow rate requested. The cooling fan — run in part load — provided the necessary cooling for condenser and absorber, even if the laboratory room is very narrow and a certain air backflow could not be excluded. As referred the prototype has shown some limitations, which are now fully identified and can be easily corrected in the next prototype, to be built ant tested until the end of 2008. It is the expectation of the authors that the nominal values will be easily reached then, namely the design COP of 0.6.

The comparison for all-year supply of heating, cooling and domestic hot water of the solar-assisted absorption chiller with a compression chiller system shows the economic viability of the novel technology. With an overall solar fraction of ca. 70 % the novel chiller requires investment and operation expenses roughly half of them of a compression-based system after 20 years of lifetime. The break-even point between the two systems is reached during the 8th year (dynamic payback time).

References

[1] European Commission (2008), EUROPA, http://europa. eu.

[2] ESTIF (2007), ESTTP (European Solar Thermal Technology Platform), Intersolar, June 2007.

[3] Eurostat (2008) — http://epp. eurostat. ec. europa. eu.

[4] The Potential for Solar Assisted Cooling in Southern European Countries. Commission of the European Communities. Contract number: RENA-CT94-0017 (1995).

[5] Mendes, L. F., Collares-Pereira, M., (1999), A Solar Assisted and Air Cooled Absorption Machine to Provide Small Power Heating and Cooling, International Sorption Heat Pump Conference, Munich, pp 129-136

[6] comunication “Collector Testing” AO SOL

[7] W. Feist, (2007). Passive House Planning Package.

[8] Solpro (2008), v.16.5, Ao Sol R&D.

[9] Daikin EWAQ-ACV3P.

[10] ERSE (2007). http://www. erse. pt.

[11] Eurostat (2007).

[12] H. M. Henning, (2005). Solar Assisted Air-Conditioning in Buildings, Springer, Wien, New York.