Category Archives: EuroSun2008-7

Humidity perception

The Armenian climate in general and the Yerevan climate in particular, are decidedly continental and are very dry, with average relative humidity of less then 20% most of the time in summers. Dry air usually helps to withstand the summer heat, when 40°C may be a common maximum temperature during 2-4 months.

The solar driven DEC machine naturally yields cold air (in our case about 17°C), that has humidity close to the upper comfort limit. When people enter into an environment that has relative humidity of more than 50%, despite of lower temperature, they feel discomfort. This fact suggests to target even at a lower temperature to compensate the “humidity” contrast that is taken negatively by the public in the auditorium. This also suggests customization of the “comfort standards” for regions that have dry climate.

Study of solar cooling systems using absorption heat pumps

T. Mateus* and A. C. Oliveira

New Energy Tec. Unit, Faculty of Engineering, University of Porto, Rua Dr. Roberto Frias, 4200-465 Porto,


* Corresponding Author, tiagomateus@engenheiros. pt

The use of solar energy for cooling is an attractive concept, because the need of cooling happens at the same time of solar radiation availability. Three computational models were developed in TRNSYS for three different buildings: office, hotel and single-family house. The models allow the simulation of solar cooling systems with flat plate collectors or vacuum tubes, with a gas boiler or electrical compression chiller backup. Lisbon, Rome and Berlin were the locations studied. The TRNSYS models are able to run for a whole year (365 days), according to control rules (self-deciding whether to operate in heating or cooling modes), with the possibility of combining cooling, heating and DHW applications. These user friendly models allowed to perform an exhaustive study about the economic viability of solar cooling systems.

Keywords: solar thermal system, solar cooling, heat pump, absorption chiller

1. Introduction

Solar thermal cooling systems are still in their infancy regarding practical applications. A survey done by the European Solar Thermal Industry Federation [1], showed that in 2006 there were only about 100 solar cooling systems installed in Europe. Of those, about 67% where based on absorption cooling technology, and more than half used flat-plate solar collectors. The number of solar cooling systems will increase significantly in the near future, due to the arrival of new players to the market. There are today several absorption chillers driven by hot water available on the market, starting at a cooling capacity of about 4 kW. This makes it possible to install solar absorption cooling systems for several building sizes, from single-family residential to large commercial buildings.

However, for these applications to be economically interesting, namely to decrease their payback period, it would be important to extend the system operation period as much as possible throughout the year. Solar thermal collectors can also be used for water or indoor space heating, thus making it possible to use an integrated system for building cooling and heating.

This study aims to evaluate the potential of integrated solar absorption cooling and heating systems for building applications. The TRNSYS software tool [2], was used as a basis for assessment. Different building types were considered: residential, office and hotel. Three different locations and climates were considered: Berlin (Germany), Lisbon (Portugal), and Rome (Italy). The different local costs for energy (gas, electricity and water) were taken into account — see Table 1. Both energy and economic results are presented for all cases. Savings in CO2 emissions were also assessed for all cases.

to 10th October, Lisbon — Portugal *

Hotel and Office Lisbon Rome Berlin

Sinfle-family house Lisbon Rome Berlin

Electricity price (€/kWh) Gas price (€/kWh)

Water price (€/m3)

0.086 0.139 0.107 0.028 0.032 0.048 1 0.8 3.6

0.15 0.235 0.195 0.05 0.066 0.066 1 0.8 3.6

Cost for yearly maintenance (solar / conventional)

0.5 / 1.0 %

0.5 / 1.0 %

yearly inflation rate (Electricity / Gas)

4.5 / 4.5 %

4.5 / 8.0 %

Product VAT (solar / conventional)

0.0 / 0.0 %

12 / 21 %

Financial support

20% 30% 20%

20% 30% 20%

Table 1. Economic inputs considered in the simulations.

Comparison of the control strategies

To compare both control strategies TRNSYS-simulations of two different places with different irradiation conditions have been carried out.

Firstly, a cloudy summer day in Berlin using a 34 m2 flat plate collector field inclined 30° in direction south separated from the chiller by a counter flow heat exchanger has been simulated. Results are presented in Figures 4 and 5.


Fig. 4. External temperatures for a cloudy day in Berlin.

In Figure 4 we see the inlet temperatures of hot and cooling water and the outlet temperature of the generated chilled water temperature for the conventional and the new control. As described before the conventional control cuts driving temperature peaks in order to ensure a minimum chilled water temperature. Cooling water temperature is constant at 27°C. The new control can use the whole solar offer. It does not cut any peaks. Cooling water temperature is adapted depending on driving temperature but limited to a minimum value of 23°C. At lower values electric energy consumption of the cooling tower fan would increase too much as it rises with the third power of the fan speed. The chilled water temperature can not be kept at the required level of 15°C for the whole working period but it is better approached as in case of the conventional control strategy.

The advantage of the new control strategy in respect of meeting the required cooling load even at cloudy days with moderate irradiation is better seen in Figure 5. Assuming a cooling load of 10 kW we see that the new control strategy can provide the full load nearly an hour earlier than the conventional. Small drops in driving temperature can be compensated completely, bigger ones better than in the case of the conventional control.

Total cooling energy supplied could be increased by 13% to 72 kWh compared to 64 kWh with the conventional control strategy (see Table 1). Total water consumption increases by 12% but specific water consumption drops slightly. Specific fan power consumption is about 11% higher. 8 kWh more cooling energy are generated by an additional fan power consumption of only 1.3 Wh. This corresponds to a COP of more than 6,000. That means that the small additional fan power consumption is well justified.


Fig. 5. Cooling capacity provided by conventional and new control strategy.

Table 1. Comparison of cooling tower power and water consumption for Berlin, 25th August 2003.

Conventional control

New control

Difference [%]

Total cooling energy supplied [kWh]




Total fan power consumption [kWh]




Specific fan power consumption [kWh/kWh]




Total water consumption [kg]




Specific water consumption [kg/kWh]




Secondly, a sunny day in Iran has been simulated. In this case higher driving temperatures up to 115°C have been available. In Figure 6, the temperatures of the external water circuits are presented.


Fig. 6. External temperatures for a sunny day in Iran.

Again, after start-up and before shut-down when irradiation is low the new control strategy has the advantage to meet the required chilled water temperature or to meet it better than the conventional. Shortly after start-up a cooling capacity of around 9 kW can be supplied, already. The full cooling capacity is reached 40 minutes earlier than in the case of the conventional control strategy. In the evening the full cooling capacity can be supplied for 20 minutes longer.

Figure 7 shows the comparison of cooling tower fan power consumption for both control strategies.


Fig. 7. Comparison of cooling tower fan power consumption.

The advantage of the new control strategy regarding parasitic electricity consumption is here clearly to be seen. At peak irradiation at 2 pm the conventional control needs nearly five times more fan power (270 W compared to 56 W). Table 2 gives an overview about fan power and water consumption for the complete day. Total fan power consumption can be reduced by 29% even though 3% more cooling energy is supplied.

Table 2. Comparison of cooling tower power and water consumption for Iran, 22th July 2004.

Conventional control

New control

Difference [%]

Total cooling energy supplied [kWh]




Total fan power consumption [kWh]




Specific fan power consumption [kWh/kWh]




Total water consumption [kg]




Specific water consumption [kg/kWh]




In this simulation both control targets — energy and water saving and meeting of the required chilled water temperatures — are combined. If the main target of the control strategy is to minimise electricity consumption and not to meet exactly the chilled water temperature a lot more energy can be saved. In the morning it is certainly often advantageous to provide the full cooling capacity earlier. But if we assume that in the evening after 5 pm full cooling load is not required anymore we can limit cooling water temperature to 27°C as in the case of the conventional control strategy. Thus, specific fan power consumption can be reduced by 50%!

2. Conclusion

The new control strategy produces the expected result: stable chilled water outlet temperature by adjusting the cooling water temperature. It provides the opportunity to operate a solar cooling system with low electricity consumption when irradiation is high enough to cover the cooling load. Simulations showed that at a sunny day in Iran up to 50 % of the fan power consumption can be saved compared to the conventional control strategy. On the other hand, a high cooling capacity can be provided even if the irradiation is not yet sufficient as e. g. in morning hours. At a cloudy day in Berlin full cooling capacity can be supplied nearly an hour earlier compared to the conventional strategy. This could also be a possibility to dispose of a short-time storage tank.


[1] P. Kohlenbach (2005). “Solar cooling with absorption chillers: Control strategies and transient chiller performance“, Dissertation, Technische Universitat Berlin, Germany.

[2] C. Schweigler, A. Costa, M. Hogenauer-Lego, M. Harm, F. Ziegler (2001). “Absorptionskaltwassersatz zur solaren Klimatisierung mit 10 kW Kalteleistung“, Tagungsbericht der Deutschen Kalte-Klima — Tagung 2001 Ulm, Deutscher Kalte — und Klimatechnischer Verein, Stuttgart, Germany.

[3] A. Kuhn, F. Ziegler (2005). “Operational results of a 10 kW absorption chiller and adaptation of the characteristic equation”, Proceedings of the 1st International Conference Solar Air Conditioning, 6/7th October 2005, Bad Staffelstein, Germany.

[4] V. Claufi, A. Kuhn, C. Schweigler (2007). “Field testing of a compact 10 kW water/LiBr absorption chiller“, Proceedings of the 2nd International Conference Solar Air Conditioning, 18/19th October 2007, Tarragona, Spain.

[5] E. Wiegand, P. Kohlenbach, A. Kuhn, S. Petersen, F. Ziegler (2005). “Entwicklung eines offenen Nasskuhlturmes kleiner Leistungsklasse“, KI Luft — und Kaltetechnik 10/2005, S. 413-415.

[6] J. Albers (2002). “TRNSYS Type 107.Part Load Simulation of single staged absorption chillers in quasi steady states”, Contribution to a design tool for solar assisted air conditioning systems developed in IEA TASK25 Subtask B Final Report, Forschungsbericht IEMB Nr. 2-67/2002.

[7] S. A. Klein et al.(2006). “TRNSYS Manual for TRNSYS 16”, Solar Energy Laboratory, University of Wisconsin-Madison, USA.

[8] V. Claufi, A. Kuhn, F. Ziegler (2007). “A new control strategy for solar driven absorption chillers”, Proceedings of the 2nd International Conference Solar Air Conditioning, 18/19th October 2007, Tarragona, Spain.

TRNSYS simulation

The TRNSYS® interface interacts with the user as a graphic programming tool. It permits to build a virtual facility and easily change from different types of configurations. Using an already validated simulation of solar cooling facilities benchmarked with the UC3M’s experimental solar cooling facility, the model of the wet cooling tower has been substituted by the model of a Ground Heat Exchanger (GHE). This model is the Type 557a from the TESS libraries for TRNSYS 16 and simulates a U-tube GHE. For more information about the TRNSYS simulation program and TESS libraries please refer to [6]. The simulation has been conducted in a trial an error way in order to size the GHE. Different numbers of boreholes connected in parallel have been simulated at different depths until the heat rejected to the ground equals the heat generated in the absorber and condenser.

During this simulation the soil thermal properties are going to be estimated because lack of information about the soil in the Madrid region. Normally, to estimate the thermal properties of the soil

a Thermal Response Test has to be performed first in order to be accurate. In the Thermal Response Test a probe is introduced into the ground and a defined heat load is circulated. The temperature difference is recorded and the properties of the soil can be calculated easily. This technique allows the sizing of the GHE to be accurate [7].

Typical market dimensions have been used in order to simulate the GHE. The borehole radius is of 11 cm and the single U-tube in each borehole has an outer radius of 2 cm and an inner radius of 1.6 cm. Different number of boreholes corresponding to 4, 6 and 8 connected in parallel has been simulated at different depths. A summary of the input values for the GHE model is presented in Table 3.

Table 3. Values for the GHE model.

Storage thermal conductivity

2.6 W/m2K

Fill thermal conductivity (Clay)

1.3 W/m2K

Pipe thermal conductivity (Copper)

52 W/m2K

Annual amplitude of air temperature (Madrid)

22 °C

Annual average air temperature (Madrid)


1.3. Simulation Results

Simulation of July 9 is conducted in order to compare the behaviour of the simulated facility incorporating the GHE with the current experimental facility. Figure 5 shows the simulation results for July 9.

Simulation vs. experimental results


Figure 5. Simulation vs. Experimental results for July 9.

The simulated heat rejection demanded by the absorption chiller reached 135.339 kWh for the day.

Making a trial and error analysis it is found that to dissipate the heat generated in the absorption chiller with 4 boreholes, a depth of 160 meters is necessary. With 6 boreholes, the depth needed for the heat rejection is of 100 meters and with 8 boreholes, a depth of 80 meters is necessary. Estimating a price per borehole of 45-65€ (depending of type of soil) per meter, it is found that the best design should be 6 boreholes of 100 meters deep.

In the Figure it is shown a time delay between the experimental and the simulated heat rejected. This is motivated by the temperature control of the recooling loop. Not circulating water through the condenser when is not needed lowers the thermal inertia of the chiller. Nevertheless, almost the same value of heat rejected is achieved but there is a slight increment in the value of the outlet water temperature from the condenser. Experimental values for July 9, 2008 reached maximum values of 30 °C while the simulation reached 34 °С. Nevertheless, the cooling energy produced does not experience major changes. Simulated cooling energy produced reached 36,547, with simulated weather conditions, a slight difference from experimental.

2. Conclusions

Although the wet cooling towers behave well in lowering the water temperature, the GHE seems to be the way to promote the use of solar cooling facilities in the sector.

The simulation conducted shows that a GHE formed by 6 boreholes of 100 meters deep and each containing single U-tubes, connected in parallel could be sufficient to supply the heat rejection rate to cool down the absorption chiller in this kind of facilities.

The construction of this kind of heat sink is more complicated than the installation of a wet cooling tower, but once installed its maintenance cost is low.

Another good characteristic of coupling a GHE to an absorption chiller is that in winter time the facility could operate to supply low temperature heat using the absorption chiller as a heat pump and the cold produced in the evaporator sent to the GHE supplying the load.

A more extensive simulation should be conducted in order to evaluate the thermal depletion of the soil. This could be counteracted by operating the facility during the whole year.


[1] G. Grossman, A. Johannsen, Prog. Energy Combust. Sci., 7 (1981) 185-228.

[2] G. A. Florides, S. A. Kalogirou, S. A. Tassou, L. C. Wrobel, Energy Conversion and Management, 44 (2003) 2483-2508.

[3] M. Izquierdo, R. Lizarte, J. D. Marcos, G. Gutierrez, Applied Thermal Engineering, 28 (2008) 1074-1081.

[4] M. C. Rodriguez, P. Rodriguez, M. Izquierdo, A. Lecuona, R. Salgado, Applied Thermal Engineering, 28 (2008) 1734-1744.

[5] R. Salgado, P. Rodriguez, M. Venegas, A. Lecuona, M. C. Rodriguez, 5th European Thermal-Sciences Conference (Eurotherm 2008), ISBN 978-90-386-1274-4, 123-124.

[6] TRNSYS 16 User’s Manual, Solar Energy Laboratory, University of Wisconsin-Madison.

[7] B. Sanner, C. Karytsas, D. Mendrinos, L. Rybach, Geothermics, 32 (2003) 579-588.

Absorption chiller TRNSYS type

The absorption chiller model is a static “black box” model (see Figure6), in which experimental COP and chilling power are used to predict the chiller response at different boundary conditions. The chiller main parameters are: specific heat of oil (Cp0il), specific heat of chilled water (Cpch), auxiliary power for pumps and fans operation (Paux) and minimum refrigerant outlet temperature (Tset). The chiller inputs are: oil mass flow rate (moil), oil inlet temperature (Toili), ambient temperature (Tamb), chilled water mass flow rate (mch), chilled water inlet temperature (Tchi). The main outputs are oil outlet temperature (Toil, o) and the chilled water outlet temperature (Tch, o), from

Подпись: Fig. 4. COP variation with ambient temperature and oil inlet temperature for: oil flow rate 3500 l/h, chilled water inlet temperature 0 °C, chilled water flow rate 2500 l/h.

which driving heat rate (Q0ii), chilling power (Qch), rejected heat (Qamb) and chiller coefficient of performance (COP) are easily calculated. For a detailed description see [7].



Nominal capacity

13 kW

Chilled water flow rate min / nom / max

2300 l/h 2500 l/h 2900 l/h

Initial charge of H2O

11.5 kg

Initial charge of NH3

8 kg

Max. generator pressure

35 bar

Min. evaporator temperature

-20 0C

Table. 1. Main technical features of the Robur ACF 60 LB absorption chiller.

Fig. 5. Chilling power (fraction of nominal value)
variation with ambient temperature and oil inlet
temperature for: oil flow rate 3500 l/h, chilled water
inlet temperature 0 °C, chilled water flow rate 2500l/h

Подпись: mbii mh Toil,i Tch,i Tamb Подпись: Toil,o Tch,o chiller
black box

Fig. 6. Absorption chiller black box model.



Подпись: air Fig8. Capsule charge phase

Integration of the solar plant in the existing system

The principle of the integration of the solar plant into the heating distribution system of the building is show in Figure 1. In winter the produced heat of the solar system is provided to the thermally activated ceilings with a maximum capacity of 200 kW on a low temperature level of 50/30°C. If the solar system delivers e. g. in autumn and spring more than 200 kW, the heating energy is switched to a higher temperature level and connected to the main heating distribution system of the building for the rest of the day. In summer the heat is directly provided to the main heating distribution system, which supplies the adsorption chillers with the required heating energy.

— Waste heat recovered (400 -500 kW)

— 3 Gas boilers 5.5 MW





Heating distribution Chilled distribution

Figure 1: Hydraulic scheme of the solar adsorption cooling plant (FESTO AG)

The installation of the solar plant is supported by the German government within the Solarthermie 2000plus program and the monitoring system was installed by the University of Applied Sciences Offenburg [1].

Energy analysis

The results obtained for the different simulations are summarised in the tables 1 and 2. They show the required solar collector surface to achieve the objective of 700 MWh/year of chilled water and

different energy performance parameters for the different thermal chiller and solar thermal technologies.

As expected, both tables show that solar collectors having higher efficiency parameters give a higher heating and cooling production per area unit. Also it could be observed that because of the lower operation of the solar adsorption systems, these have higher solar gain coefficients than the absorption ones. Alternatively, the specific chilled energy coefficients depends not only of the collector technology but also of the thermal chiller technology, being the best options the ab/adsorption systems with ETC-CPC collector with a maximum performance of 606 kWh/m2y for the combination with the BDH-65 chiller. The adsorption systems specific chilled energy coefficients only exceed the ones of the absorption systems when FPC collectors are used.

Comparing the Broad solar cooling systems themselves, the BDH-65 systems give a better performance than the BDH-50 ones, especially for ETC and FPC collectors. That fact leads to lower solar collector surface requirement to achieve the 700 MWh of chilled energy. The reason of that behaviour is the better annual average performance of the thermal collectors in BDH-65 systems. Looking at only to the adsorption results of table 2, the MYCOM ADR-60 SYSTEMS present better performance than the ADR-80. In that case the solar system has more or less the same performance and the explanation of this result is that the ADR-60 systems operate at slight higher temperatures in the generators-receivers, obtaining then slight higher values of their COP.

T able 1. Collector surface and energy performance parameters for different solar collector technologies and

absorption chillers.







Collector Area m2




Spec. Solar gain


Annual chilled energy


Spec chilled energy




Broad BDH-65 663 kWc
























Broad BDH-50 512 kWc
























Table 2. Collector surface and energy performance parameters for different solar collector technologies and

adsorption chillers.













Spec. Solar gain


Annual chilled energy


Spec chilled energy
























































Comparing this analysis to the previous study [3], it could be observed that the MYCOM ADR-80 results are very similar with differences lower than 3.5 %. In the previous study we used some empirical correlations to simulate that chiller. This entire means that the method suggested by Ktihn and al. [3] is valid to model this adsorption chiller. As regards the absorption chiller model, in [3] we used the Thermax LT 21 S with 739 kW with an average COP of 0.64. In the case we are dealing with now we selected the Broad chillers because they present a better COP (0.76). As a consequence, the values of the specific chiller capacity calculated now are between a 4 and 15 % higher. Obviously, the collectors’ surface requirements are also reduced in the same amount.

It should be remarked that, due to the special chilled water demand profile, the solar cooling system should be in operation the whole year. As a result the cooling water temperature is almost 2/3 of the year close to 22 °C. One of the most important consequences of working at lower cooling tower water temperatures is that the temperatures needed in the generator of the chiller could be lowered maintaining the capacity and then increasing the performance of the solar thermal field. In fact, the temperature in the generator can be as low as 60°C for adsorption and 85°C for absorption in the winter period.

Performance in summer

In Figure 6 temperature trends of the storage system and the average seasonal yield of the collectors varying the capturing surface and storage volume are reported. The average seasonal temperatures are always between 85°C and 100°C inclusive which correspond to average seasonal values of collector efficiency which varies from a minimum of 65% to a maximum of 75%.

In Figure 7 the trend of the average seasonal solar fraction is illustrated, defined as the seasonal refrigerant energy supplied by the absorption chiller compared to the seasonal refrigerant energy required by the building, equal to 31500 kWh cooling energy, varying the collectors surface

Подпись: Fig.6. Storage system temperature and average seasonal yield of the collectors varying the capturing surface and the storage volume Подпись: Collectors Surface [m2] Fig. 7. Average seasonal solar fraction varying the capturi ng surface and storage volume

and the storage volume. Such a parameter grows both with the capturing surface and with the storage volume, since it is possible to obtain higher inlet temperatures at the absorption chiller generator with consequent greater refrigerant energy provided. It is important to observe that the increase marked by the solar fraction is obtained by changing from 25 m2 to 100 m2 of collectors, while for surfaces greater than 100 m2 the solar fraction hardly increases. Despite the higher temperatures supplied to the generator, the solar fraction is in fact penalised by the operative limits of the absorption machine.

Finally, in Figure 8 the average seasonal performance coefficient trend of the absorption machine varying the capturing surface and storage volume is illustrated. The average seasonal COP varies between a minimum of 0.391 to a maximum of 0.437, from the reference value provided for the absorption chiller considered and equal to 0.60. Such a value is actually attainable in nominal operative conditions, with adequate inlet temperatures to the generator obtained by means of an auxiliary heater. The performance coefficient of the machine is penalised since the plant considered uses a traditional chiller instead of an auxiliary heater for the distribution of the refrigerant load which is not distributable by the absorption chiller.

Подпись: 0 50 100 150 200 250 Collectors Surface [m2] Fig.8. Average seasonal COP of the considered absorption chiller varying the capturing surface and storage volume 0,44

C. O.P.

0,43 0,42 0,41 0,40 0,39 0,38

3. Conclusions

The energy performance of a solar energy system which uses a simple effect absorption chiller for the production of a refrigerated water flow rate used for the cooling of an open space environment by means of a radiant ceiling was analysed. The same solar collectors are used during winter for heating the same building, by means of the same distribution terminals. Heat pipe solar collectors were simulated by virtue of the elevated temperatures required for correct functioning of the absorption chiller Control logics were proposed which regulate the inlet temperature of the radiant ceiling, which permitted the evaluation of solar fractions obtained by the plant, based on
the capturing surface area and storage volume. It was observed that during winter, due to the high temperatures reached in the storage system, the collectors operate with limited efficiency yet the system is capable of supplying solar fractions close to 100% for capturing surfaces greater than or equal to 100 m2, independently of the storage tank volume. The high temperatures reached in the tank are opportunely regulated to supply the radiant ceiling, by means of a three way valve blending system which operates for long periods of time with high recirculation flows. The total yield of the system, due to the losses of thermal energy in the tank, assumes values equal to 30% for capturing surfaces of 200 m2, independently from the storage volume.

In summer, the system yield assumes values that are variable between 75% and 13%, due to the operative limits of the absorption machine which requires the use of an auxiliary chiller. It was decided to use an electric auxiliary refrigerating machine to supply the load that was not distributable by the absorption machine in that preliminary evaluations of primary energy consumption showed convenience compared to the use of an auxiliary heater usable to raise the temperature of the generator.

The solar fractions reached increase slightly for capturing surfaces greater than 100 m2, since they exceed the operative limits of the absorption chiller, principally represented by its insufficient flexibility, which makes the use of an auxiliary system necessary. For an capturing surface of 50 m2 and a storage volume of 25 m3 an average seasonal solar fraction equal to 30.6% is attained; the maximum value is 35.7% for a plant with 200 m2 of collectors and 25 m3 of storage. The refrigerant energy supplied by the absorption chiller hardly varies for capturing surfaces greater than 100 m2; for a storage volume of 25 m3 it is 10643 kWh with a surface area of collectors equal to 100 m2 while it is equal to 11185 kWh changing to an capturing surface of 200 m2, with an increase of 5% with a doubling of the capturing surface.

Finally, it is possible to observe the slight variability of almost all the analysed parameters compared to the storage volume; storage volumes greater than 10 m3 do not bring about benefits to plant performance but it is not possible to go below such values in order to guarantee system stability.


[1] Arcuri, N., Bruno, R., Ruffolo, S., 2005. Prestazioni termiche di sistemi di riscaldamento a soffitto radiante alimentati da collettori solari, Proceedings from the 2nd International CLIMAMED Conference, Madrid, March

[2] Lazzarin, R., Crose, D., 2000. Il soffitto radiante nella climatizzazione ambientale, SG Editoriali, Chap. 1

[3] TRNSYS, Reference Manual, AA. VV., 2001. A transient system simulation program, Solar Energy Laboratory, Madison, Wisconsin, USA

[4] UNI 10349, 1994 — “Heating and cooling of buildings; climatic data”

[5] Gansler, R. A., KleinS. A., 1993 “Assessment of the Accuracy of generated Meteorological Data for Use in Solar Energy Simulation Studies”, Proceedings of the 1993 ASME International Solar Energy Conference, Washington D. C.

[6] Oliveti G., Arcuri N., Bruno R., De Simone M., 2007 “Energy Performances Of A Radiant Floor Heating System Supplied By Solar Collectors With Ventilation Stream Heating By An Air To Air And An Air To Water Heat Exchanger”- REHVA International Congress — Helsinki [7] Oliveti G., Arcuri N., Bruno R., Mazzuca A., 2005 “Energy performance of an absorption chiller supplied by solar collectors in Mediterranean area”, SWC 2005, Orlando, USA

[8] UNI 10339, 1995, Air-conditioning systems for thermal comfort in buildings. General, classification and requirements. Offer, order and supply specifications.

[9] Lazzarin R., Castelletti F., Busato F., 2006 Soffitti radianti e aria primaria, Condizionamento dell’aria Riscaldamento e Refrigerazione N° 6

[10] Oliveti G., Arcuri N., Bruno R., 2008, Caratterizzazione di Soffitti Radianti che Impiegano Tubi Capillari per il Riscaldamento degli Ambienti” AICARR International Convention

[11] Arcuri N., Bruno R., 2005, Prestazioni termiche di sistemi di raffrescamento a soffitto radiante e relative strategie di controllo, 60° ATI National Congress, Rome


Marco Beccali, Pietro Finocchiaro, Massimiliano Luna, Bettina Nocke

Dipartimento di Ricerche Energetiche ed AMbientali (DREAM) — Universita degli Studi di Palermo
Via delle Scienze bld. 9 — 90128 Palermo, Italy — tel +39091236211 — Fax +39091484425
Corresponding Author: marco. beccali@dream. unipa. it

The paper concerns the first monitoring results of a solar DEC system that has been installed at the end of 2007 at the DREAM Department, Universita degli Studi di Palermo, Italy. The system is mainly composed of a desiccant cooling unit for primary air treatments specially designed for application in humid climate, a liquid solar collector system and a radiant ceiling. The system is fully operating since March 2008 and first summer operation results and evaluations are already available. The monitoring results have been aggregated in terms of COP of the desiccant cycle, fraction of the total cooling energy covered by the desiccant cycle and primary energy ratio. They are fairly good and reasonably comparable to the simulation results previously conducted in the design phase. In particular, the monitoring of the system has shown interesting results especially for what concerns the heat recovery from the condenser side of the water chiller used to feed the radiant ceiling and the auxiliary cooling coils. On the other side, some improvements are still necessary, such as air leakages reduction in the heat exchanger, check of some relative humidity measurement points, optimization of set points for auxiliary coils and supply humidity control.

Keywords: Desiccant cooling, radiant ceiling, heat rejection, monitoring results

1. Introduction

Several studies show that desiccant cooling system have a limited dehumidification potential for given characteristics of the desiccant rotor, regeneration temperature, flow rates and so on and that auxiliary cooling power for dehumidification is required to fulfil the desired supply air conditions. In addition, the humidity ratio of outside air for the considered location can be very high in summer season (over 20 g/kg). Therefore, particular interest was set to the dehumidification performances of the system. Moreover, the system can be tested in different internal latent load conditions by means of adjustable humidifiers, in order to simulate different occupation patterns.

Another key issue of the set-up is that the heat rejection of the chiller is partially utilized in a heating coil displaced in the AHU return air path for regeneration purposes, allowing a reduction of the required solar field area.

It is known that the energy saving potential of solar cooling systems can be very dramatically affected by higher electricity consumption for auxiliary equipment (fan, pumps, etc) in comparison to the one of conventional systems. With this aim, a detailed monitoring of electricity consumption for ventilation, chiller and auxiliaries was performed.

2. Description of the system

The system is composed of a desiccant cooling unit for primary air handling (AHU) equipped with liquid solar collectors and coupled with a radiant ceiling. The AHU provides fresh air to a room of 450 m3 used as a laboratory for modelling and testing of solar cooling systems. Also the performance of the desiccant AHU in combination with the radiant ceiling can be investigated, with the aim to maximize the specific cooling power of the radiant ceiling in different load conditions. Figure 1 shows the system configuration and its main component.










Подпись: AIR Подпись: CC1 Подпись: CC2


Подпись: EXTERNAL CONDENSER image323


Fig. 1. Functional scheme — summer operation

The desiccant cooling unit provides air change and dehumidification by means of a desiccant wheel regenerated both from solar collectors and rejected heat of the chiller. If the humidity ratio and/or temperature set point of supply air is not met, further dehumidification and temperature decreasing can be achieved by means of two auxiliary cooling coils. In particular, the first coil CC1 is utilized for pre­dehumidification purposes, whereas the second one controls the air temperature if the indirect evaporative cooling is not sufficient to reach the desired supply temperature. The radiant ceiling provides to meet the sensible loads.

image324The control of the AHU in cooling cycle has four operation modes. The system starts in the ventilation mode, where no active air handling is performed, but only the fans are switched on (mode 0). In mode 1, only indirect evaporative cooling is performed by means of a heat recovery wheel HX and the humidifier on the return side. In mode 2, the solar desiccant cooling cycle is operated by means of the desiccant wheel regenerated by the heat coming from solar collectors and the condenser of cooling

machine. In mode 3, the auxiliary back-up coils are activated to meet the cooling loads in case the desiccant cycle is not sufficient to reach the desired supply air conditions.

As already mentioned, a thermodynamic heat recovery is achieved by means of the condensation heat of the water chiller feeding the radiant ceiling and the auxiliary cooling coils. The heat rejected by the chiller can be used on the return side process to preheat the regeneration air stream [1, 2, 3]. The condensation coil HC1 allows a preheating of the regeneration air stream increasing its temperature of about 12-15°C. An external condenser connected in series to this coil provides for the remaining condensation of the refrigerant of the vapour compression cycle. A further advantage of this configuration is also the good time correlation between the cooling power demand and the heat rejection at the condenser. Further temperature rise up occurs in the solar collectors loop in order to reach regeneration temperatures of 65 — 70°C at the inlet of the desiccant rotor. Due to the higher chilled water temperature required both by the chilled ceiling and auxiliary cooling coils (about 12°C), an increase of the cooling capacity and COP of the chiller can be achieved. A heat storage tank of 0.6 m3 balances the heat produced by the solar system and the heat supplied to the coil HC2. The following table shows some relevant technical data.

Table 1: Design data

Absorber area of solar flat plate collectors



Azimuth of solar field



Slope of solar field



Heat storage capacity



Design ventilation air flow rate



Specific absorber area


[m2/1000 m3/h]

Active area of the radiant ceiling



Max cooling capacity of the cooling machine



Rated COP of cooling machine



A more detailed description of the plant in terms of control issues and monitoring system can be found in a previous work of the same authors [4].

Adaptation of Type 177 to EAW Wegracal SE 15

Подпись: Fig. 2. Determination of averaged and variable parameters for the characteristic equation.

In Fig. 2 the cooling capacity of the ILK test data is shown as function of AAt*. The modified characteristic temperature difference AAt* has been calculated from the arithmetic means of the measured inlet and outlet temperatures assuming a constant Duhring parameter (5=1.15). In addi­tion Fig. 2 depicts the results of a linear data fit with constant and variable characteristic parameters sE and rE. For the determination of the variable parameters sX and rX the data have been grouped into five classes between the minimum and maximum temperature lift included in the ILK data base. In Fig. 2 only the groups with lowest and highest temperature lift (AtACE = 13±2 K and AtACE = 29±2 K) are high-lighted with grey symbols to illustrate the variation of sE, rE and, AAtmin, respectively. If only these two groups are used, the coefficients sE1 and sE0 for the linear dependency of sE(AtACE), in equation (6) can be found from the upper two equations in Fig. 2. However, for the parameter list of Type 177 all five groups of AtACE-intervals have been used as supporting points in a linear least square fit to determine all the values sX,1, sX, o and rX1, rX,0.

As seen from Fig. 2 the general trend of the cooling capacity is predicted correctly when cal­culated by the characteristic equation with constant (solid line) or variable (closed symbols) parameters. If the variable parameters are used, also the scatter of the measured data is reproduced slightly better. Nevertheless, a relatively large deviation remains. This may be an effect of measurement uncertainties and / or different methods and time intervals of chiller evacuation during the test procedure. These effects are not included in the method of characteristic equations.