Category Archives: EuroSun2008-7

International protocols and Environmental criteria

In 1974 Nobel Prize winners Rowland and Molina proposed that chlorofluorocarbons were stable enough to reach the stratosphere, where, under intense solar radiation they released chlorine atoms that could destroy stratospheric ozone layer protecting the earth’s surface from UV rays. This was confirmed by an extensive worldwide program that aimed at monitoring the ozone layer. As a consequence, a series of national and international agreements calling for the phase-out of the production, sale and use of ozone depleting substances such as chlorofluorocarbons (CFCs) and hydrochlorofluorocarbons (HCFCs) [3] were formulated. Another major environmental concern is the climate change also called global warming. Our planet is already facing the consequences of this phenomenon: polar ice caps melting, flooding from high precipitations, increase of the earth global mean surface temperature, plants and birds migration towards the poles, etc. The increasing awareness conducted to the Kyoto Protocol in 1997 [11]. The impact of a substance on the environment is characterized by parameters like ozone depletion potential (ODP), global warming potential (GWP), atmospheric lifetime (ALT). Table 1 lists environmental data for potential ORC working fluids. For the sake of a safety environment, the designers when selecting a working fluid for a particular application should insist on substances for which the ODP, GWP and ALT are as low as possible.

Demonstration systems

Two demonstration subprojects (Demo SP6a and Demo SP6b), among a total number of 12 in Polysmart project, are implemented in Portugal. Their description as well as the main differences between them is presented below.

3.1. Main characteristics of Demo SP6a

Subproject SP6a is designed to condition (heating and cooling) some office rooms (175 m2) equipped with fan coils at Ao Sol’s two-storey office building, and to heat up a DHW tank (0,4 m3) for the lavatory (showers included). A showroom (50m2) equipped with a small radiant floor as well as two fan coils (4-pipe distribution system) will also be supplied by the CHCP plant whenever needed during guided visits.

3.1.1. Functional Block Diagram and main Equipment

The two main parts of the Combined Heat Cool and Power system — pCHCP are the Combined Heat and Power unit — CHP (cogeneration) and the Thermally Driven Chiller unit (TDC).

The former is composed by a LPG fuelled, single-cylinder 4-stroke engine (approx. 580 cm3) which drives a 5.5 kWe nominal active power, water-cooled asynchronous generator [5]. Using this type of fuel, the standard CHP unit, operating at nominal conditions, has a thermal output of

around 12.5 kWth and the heat recovery circuit reaches a temperature of 84°C. Since this level of temperature is not high enough to drive the TDC at full capacity, the CHP unit is equipped with an additional condensing exhaust heat exchanger (condensing unit) that enables to further raise the TDC driving temperature to around 90 °C by recovering the heat from the flue gas (at around 150°C). Thus, in this particular case, the overall efficiency of the plant can rise to over 100% (in relation to LHV for the fuel used) depending on the environmental conditions and conditions of use.

The latter is an 8 kWcold ammonia/water prototype machine [6] with a nominal COP of 0.6 and direct air-cooling (no need for external heat rejection devices — cooling tower).

A hot water storage tank (1m3) is used as a buffer between the CHP and the TDC / heat sink. This buffer also allows modulating the water temperature of the heating distribution circuit, since the CHP has a fixed thermal output.

Since the return temperature from the TDC hot water circuit is higher than the maximum allowed inlet temperature of the CHP heat recovery circuit, the DHW tank can be heated up by the surplus heat trough an internal heat exchanger. Once the DHW is fully charged, an Emergency Heat Rejection unit (re-cooler) guaranties the inlet flow temperature, preventing the CHP from damage.

The chilled water circuit is not physically separated from the CHCP main circuit by any heat exchanger.

Подпись: Fig.2. Functional Block Diagram

The following figure shows the functional block diagram of the pCHCP plant.

The system can also use solar energy as thermal input, both in winter and in summer. The collectors will be CPC AO SOL and 30m2 are already installed for this purpose. The objective is to measure the contributions of the different sources in different operating modes, analysing the results and proposing the best operational and integration procedures.

Both yearly heating and cooling loads and peak power demand have been estimated using TRNSYS simulation tool [7]. The goal is to select the zones to acclimatize that lead to an optimum operating period of the CHCP plant.

The conditioned area of the building is around 175 m2 and the yearly specific cooling load is estimated to be circa 25 kWh/m2.

Regeneration temperature and improved system configuration

Figure 7 shows that a rising regeneration temperature causes a sharp increase in exergy destruction and a drop in the exergetic efficiency. Regeneration at lower temperatures is favourable, which is also a well known fact for rotary type DEC systems. At higher regeneration temperature the temperature differences during the sensible heating and cooling stages are high, therefore increasing irreversibilities of the heat exchange. The decrease in exergetic efficiency is further due to an increase in exergy losses as the air leaving the heat exchanger during desorption is of a temperature significantly higher than ambient (fig. 2), thus of significant exergy content.

An improved system configuration would maximize the exergetic product and minimize exergy losses, increasing the exergetic efficiency. This could theoretically be realized a) during adsorption by further cooling the dehumidified process air (state 2) by an additional indirect evaporative cooler and b) during desorption by exploiting the exergy of the air leaving the sorptive heat exchanger for the pre­heating of the regeneration air. Simulations were performed for an enhanced system configuration including both improvement measures. The heat exchanger for heat recovery was characterized by an efficiency of 80% and simulations were performed for a regeneration temperature of Treg=60°C. Due to the lower resulting process outlet temperature T2 during adsorption the exergetic product can be increased by 17%, whereas due to pre-heating of regeneration air the exergetic input is decreased by 9% compared with the simple configuration at equivalent regeneration temperature. Therefore, the exergetic efficiency rises from 0.18 to 0.23, which is an increase by 28%. To put these values in perspective, the thermal COP, calculated as the ratio of removed heat to heat input during desorption, increases from 0.99 for the simple to 1.05 for the improved configuration. Due to the high COP values a good energetic performance is expected, but the exergetic assessment characterized by comparatively low exergetic efficiencies suggests that the improvement potential is still high. Detailed analysis of the exergetic performance for the individual cycle stages and the sub-processes evaporative cooling and adsorption/desorption is within the scope of future work.

6 Conclusions

Simulation results from a dynamic model of a sorptive heat exchanger were presented in an exergetic framework. Results show that the exergetic efficiency reaches a maximum at specific combinations of ambient temperature and humidity ratio, suggesting that the heat exchanger performs best in a certain range of climatic conditions. The exergetic performance is highest at low regeneration temperature, thus underlining the system’s good applicability with low grade heat. Further simulations showed that the exergetic efficiency can be increased by an enhanced system configuration comprising heat recovery during desorption and an additional indirect evaporative cooler.

7 References

[1] Motta, M., Henning, H.-M. (2005): A novel high efficient sorption system for air dehumidification (ECOS), International Sorption Heat Pump Conference, June 22-24, Denver, CO, USA.

[2] Schicktanz, M., Nunez, T. (2008): Modelling af an adsorption chiller for dynamic system simulation, International Sorption Heat Pump Conference, Seoul, Korea, September 2008, in press.

[3] Bosnjakovic, F. (1997): Technische Thermodynamik, Teil II, 6. Auflage, Steinkopf, Darmstadt.

[4] Szargut, J., Stryrylska, T. (1969): Die exergetische Analyse von Prozessen der feuchten Luft, Heizung, Lueftung, Haustechnik, 5, 173-178.

[5] Wepfer, W. J., Gaggioli, R. A., Obert, E. F.(1979): Proper evaluation of available energy for HVAC, ASHRAE Transactions, 667-677.

[6] Taufiq, B., Masjuki, H., Mahlia, T., Amalina, M., Faizul, M., Saidur, R. (2007): Exergy analysis of evaporative cooling for reducing energy use in a Malaysian building. Desalination, 209, 238-243.

[7] Chengqin, R., Ninping, L. Guanga, T. (2002): Principles of exergy analysis in HVAC and evaluation of evaporative cooling schemes. Building and Environment, 37, 1045-1055.

[8] Taufiq, B., Masjuki, H., Mahlia, T., Amalina, M., Faizul, M., Saidur, R. (2007): Exergy analysis of evaporative cooling for reducing energy use in a Malaysian building. Desalination, 209, 238-243.

[9] Kanoglu, M., Carpinlioglu, M., Yildirim, M. (2004): Energy and exergy analyses of an experimental open — cycle desiccant cooling system. Applied Thermal Engineering, 24, 919-932.

[10] Camargo, J., Ebinuma, C., Silveira, J. (2003): Thermoeconomic analysis of an evaporative desiccant air conditioning system. Applied Thermal Engineering, 23, 1537-1549.

8 Nomenclature and Subscripts




area [m2]


molar mass [g/mol]






specific heat capacity [J/kgK]


mass [kg]






exergy [J]


specific volume [m3/kg]


dry air




specific exergy [J/mol, J/kg]


mass transfer coefficient [kg/Ns]




reference state


pressure [Pa]


efficiency [1/100]




state points


specific gas constant [J/molK]


relative humidity [%]




temperature [K, °C]


humidity ratio [kg/kg]


constant pressure


time [s]



9 Acknowledgement

The project support of the German Federal Ministry of Economics and Technology and the support of the Reiner Lemoine Stiftung for the Ph. D. research of Constanze Bongs is gratefully acknowledged by the authors.

Market availability

During the last few years especially in Europe many new small-scale sorption chillers have been developed. Many of these absorption and adsorption chillers have now passed from the prototype phase to field tests and into the production. Today absorption chillers with capacities from 4.5 kW to 30 kW and adsorption chillers with capacities from 7.5 kW to 15 kW cooling capacity are available [3]. Table 1 shows the different applications for the different chillers, the adsorber chillii® STC8 with a cooling capacity of 7.5 kW is mainly for residential buildings, the 12 kW ammonia/water absorption chiller chillii® PSC12 is for office buildings or process cooling like e. g. milk cooling and the 15 kW water/silica gel adsorber chillii® STC15 as well as the water/lithium bromide absorber chillii® WFC18 (17.5 kW cooling capacity) are for air-conditioning e. g. of office buildings, hotels, banks, bakeries, public and administration buildings.

0.79 x 1.06 x 0.94 m3

Подпись: Company Product name Technology Working pair

image221 image222 image223 image224

Подпись: Cooling capacity Heating temperature Recooling temperature Cold water temperature COP

Подпись: (source: Pink) 12 kW 85 / 78 24 / 29°C 12 / 6°C 0.62
Подпись: 7.5 kW 75 /68°C 27 / 32°C 18 /15°C 0.56 Подпись: 15 kW 75 / 69 27 / 32 °C 18/ 15°C 0.56
Подпись: 17.5 kW 88 / 83°C 31 / 35°C 12.5 / 7°C 0.70
Подпись: 0.80 x 0.60 x 2.20 m3 Подпись: 0.79 x 1.35 x 1.45 m3 Подпись: 0.60 x 0.80 x 1.77 m3

Table 1. Small-scale sorption chillers for solar cooling systems.

Dimensions (LxDxH)


260 kg

350 kg

510 kg

420 kg

Electrical power

20 W

300 W

30 W

72 W

Structure of the method

image303 image304 Подпись: DDL image306

Подпись: VLПодпись: £> Which size and technical features for the solar cooling system ?

image309 Подпись: Which budget (invest + running costs) ? Annualized costs ? Costs of PE savings ?

Подпись: П

Подпись: Conclusions (presentation of all the results)
Подпись: How can I summarise all the £> results of my predesign ?
Подпись: Qualitative (web presentation)

The fast predesign method for the selection and predesign of solar cooling systems in buildings is made of several steps which will be described in the following chapters. Some of them are didactic whereas others are more technical. The Figure 1 is aimed at presenting the structure with the different steps constituting the method.

Fig. 1. Structure of the method.

Influence of collector area and storage tank volume

Several simulations were performed for the different buildings situated in Lisbon. For the office building, the best results occur for a collector area and a storage volume between 120-180 m2 and

0. 05-0.13 m3/m2, respectively; that represents a storage capacity between 1-5.5 hours. For the hotel, the best results occur for a collector area and a storage volume between 200-300 m2 and 0.01-0.07 m3/m2, respectively; that represents a storage capacity between 0.3-3.1 hours. For the single-family house, the best results occur for a collector area and a storage volume between 15-30 m2 and 0.05-0.13 m3/m2, respectively; that represents a storage capacity between 2.5-13 hours.

4.2 Energy, economics and emission savings

Considering a solar thermal system size that provides an annual solar fraction of 60%, for the office building and hotel, the solar fraction is higher for March and December; for the single­family house, the solar fraction is higher for April, May and October. The higher solar fraction is related to the lower energy need that occurs in those months.

4.4.1 Office building

Table 4 presents the energetic, economic and emission analysis for the system configuration with lower total cost of produced energy for a solar fraction of 60%.

For the office building there is no economic viability in any of the locations considered — see Figure 3.

In situations where an electric backup is not possible (only gas backup), the solar system installed in Rome with a solar fraction of 60% is economically more profitable.

A gas boiler as backup solution, instead of an electric compression chiller, allows a reduction in solar collector area between 3-9%, for the same solar fraction.

In the Mediterranean cities, the flat-plate collector compared to the vacuum tube collector, allows a reduction in the total cost of produced energy between 0-1.6 c€/kWh — see Figure 3. In Berlin, both collector technologies show similar results. Vacuum tube technology has the advantage of allowing a reduction of collector area between 30-50% — see Figure 4.

4.4.2 Hotel

For the hotel, in Rome it is possible to achieve economic viability for a solar system with gas backup and solar fractions between 20-40% — see Figure 3. Comparing with a conventional electric air-conditioning, this system leads to a reduction of 5 c€/kWh of produced energy. In situations where an electric backup is not possible (only gas backup), solar systems installed in Rome and Lisbon with solar fractions between 20-60% are economically interesting. A gas boiler as backup solution, instead of an electric compression chiller, allows a reduction in solar collector area between 0-20% for the same solar fraction.

In the Mediterranean cities, the flat-plate collector compared to the vacuum tube collector, allows a reduction in the total cost of produced energy between 0-2.8 c€/kWh — see Figure 3. In Berlin, both


collector technologies show similar results. Vacuum tube technology has the advantage of allowing a reduction of collector area between 20-50% — see Figure 4.

Table 4. Energy, economic and emission analysis for systems in Lisbon with gas backup and 60% solar




Energy and emission data

Economic data

Cooling: gas absorption chiller Heating: gas boiler

Offices building

CO? emissions saved in 20 Years: 182 ton Elect. Energy (MWh/ano)




Elect. 1% .



19 Gas saving

2% 454 56%






Investment: Solar: 48 k€ / Backun: 35 k€ Exploitation cost (€/year)

Electricity 204 €

7% ж Gas 931C зг%

1 Ж Maintenance Sav. ngin^^B ^ 598 C lrstyear x—,. I Water 21%

962 € ^————————————— 215 €

33% 7%


CO? emissions saved in 20 Years: 500 ton elect. Energy (MWh/ano)




Consump. ^


4 3 VGas saving 2% 125.0







Investment: Solar: 97 k€ / Backun: 58 k€ Exploitation cost (€/year)

Electricity-. _

^ ззГзс

7* 34%


Saving in 1061C lrstyear^—14%

2 889 Є ^^^Щ^ЛЛ/ater

38% — 575 €


Single-family house

CO2 emissions saved in 20 Years: 28 ton Consump. Energy (MWh/ano)







0 2 n Gas saving

1% 71 Ж 58% Gas





Investment: Solar: 6 k€ / Backup: 13 k€ Exploitation cost (€/year)

Electricity 38 €

5% GaS 245 €




Saving in


Experimentation for pool 2: desiccant process


In order to propose an alternative to traditional desiccant cooling systems, using desiccant wheel, the CEA — INES, in collaboration with the CNAM-IFFI, developed a multifunctional exchanger (Figure 1a) able to couple heat and mass transfers of the desiccation.

The process air enters the heat exchanger by the centre and then flows in a radial direction between the desiccant bed and the periphery where the air is collected. To heat or cool the desiccant during regeneration or adsorption mode, a copper heat exchanger made of 30 tubes of 10 mm diameter with fins is immersed in the desiccant bed. The exchanger is filled with 50 kg of Engelhard KC — Trockenperlen-N type silica gel. The particles are of spherical shape with a mean diameter of 3.4

mm. The total mass of the prototype is 60 kg. The Figure 1b presents the scheme of a solar desiccant cooling installation using desiccant heat exchangers. The traditional two wheels (desiccant and heat recovery) are replaced by two desiccant exchangers (air and desiccant/water) functioning heated or cooled; the cooling corresponds to the drying step while heating is used for the regeneration. During the regeneration phase, hot water, flowing directly from solar collectors, is used. During the adsorption phase, cold water circulation, resulting for example from a cooling tower or geothermal sinks, maintains the adsorbent under optimal sorption conditions. The prototype of heat exchanger filled with adsorbent was tested and it appeared that the system is able to cool air from 33°C and 40% HR (outside air) to 24°C and 90% HR (blowing air) [1]. A TRNSYS numerical model of desiccant heat exchanger was also developed to simulate the heat and mass transfer for the adsorption and regeneration processes in the prototype. The results of the numerical model are in agreement with the experimental results [2], and systems performances could be evaluate.

Modelling of a solar cooling system

We are going to present the simplified model of our installation whose principle scheme is shown on the Fig. 1. The aim is to simulate the behaviour of our installation in order to get a better understanding of the comfort in the building hour by hour, taking account of external demands such as sunshine and attendance of the building. Therefore each component is modelled in the SPARK environment and integrated to the general model.

3.1. The solar collector field

image559The solar thermal collector field is composed of 36 double glazing collectors (SchucoSol-U5 DG) specially designed for solar cooling systems. The collectors are spread over three loops of ten collectors plus a loop of six collectors. The four loops are installed in parallel. Before modelling the solar collector field, we have to carry out a “macro class” which will represent one solar collector. Then, we have to duplicate and to couple this macro class to meet the distribution. The modelling of our solar collector is based on the efficiency method. The efficiency equation [1] is given by the manufacturer and depends on the inlet and outlet temperature of the collector, climatic conditions (the outside temperature (Toutside) and the sunshine (G)).

V. .

dU — ^raialvad



«W* = 111 ^ Cp « (Tto — Tout)

iyalct’jgL *

* ™"i8L = Чипашг * * Гі ~ ™ * * (^I«J — TL, l)








To model the solar collector we apply the first law of thermodynamics [3]:


Fig. 4 : solar captor scheme








3.2. The hot and the cold tanks

To model the tow tanks, we are going to use the same method by applying the first law of thermodynamics. Example for the hot water tank:

dT .



The assumptions are:

The generator pump starts when Twater tank > 80°C.

The pipes of the solar loop are well insulated (no heat losses).

The flows into the solar and the generator loops are constant when the pumps are working.

Investigation of the mass flow rate in the discharging heat exchanger


Since only part of the return fluid stream of the chilled water is cooled by the solar driven absorption chiller, the influence of the system performance of the quantity of this mass flow rate was investigated. Because the mass flow on the “cold side” was more or less fixed due to the flow rate required by the absorption chiller only the “hot side” flow rate, which is the side of the chilled water of the existing system, was changed. Three different mass flow rates were simulated: 2000, 3000 and 4000 kg/h. Results are presented, in fig. 12 and fig. 13, for the systems with 2 m3 storage volume and varying number collectors. As expected the solar fraction SF as well as the total system performance increases with higher flow rates on the “hot side” of the heat exchanger (see fig. 12).

This is because the storage can be discharged faster (with larger temperature differences) in discharge mode. Moreover the chiller performs better in direct mode due to the higher inlet temperatures in the evaporator heat exchanger (i. e., less part load operation). Increasing the mass flow rate from 2000 kg/h to 4000 kg/h results in increment of about 1% of the solar fraction for a specified number of collectors, However, higher flow rate means larger pump, higher costs and

electricity consumption. And thus, a middle value of 3000 kg/h was selected as the flow rate for the cold side of the system.

Storage tanks loop

The results show a tendency of over prediction of the energy going out of the storage tanks. This is mainly due to the model used for the storage tanks which simulates the “ideal case” in order to maintain the stratification in the tank (the hot water coming from the collectors goes in the layer with the similar temperature). In reality, the hot water enters in a fixed inlet in the middle of the first tank which limits the stratification in the tanks. The temperature at the top of the first tank (to supply) is therefore smaller than the simulated one, which explains the difference with simulated values. This problem can be solved by using another type of tank model using fixed inlet positions.

2.2. Adsorption chillers

Подпись: (2) Подпись: C0P = Подпись: (3)

Table 3 summarizes the different cases simulated and compares the measured and simulated results in terms of mean cooling power and COP of the chillers defined by equations (2) and (3).

n is the number of time steps in the considered period i is the considered time step.

Table 3: Summary of the measured and simulated results for one of the adsorption chiller



P_cool, m

P_cool, sim













Case 1









Case 2









Case 3









Case 4









NB: The operating conditions for the two others water circuits (chilled and recooling) can be seen on the figures 5-8.

The results show good agreement between simulation and measurements for the different operating conditions studied. Therefore the model will be used for further investigations.