Category Archives: EuroSun2008-7

DESICCANT COOLING SYSTEM WITH A REGENERATIVE EVAPORATIVE COOLER

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 building and its systems

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.

image389

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:

Thermo-economic analysis

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.

Table 1. Analysis parameters

Site parameters

Efficiencies

Investment parameters

Operation

parameters

Living area

200 m2

Gas burner

85%

NH3 chiller (8 kW)

6.000 €

Gas price

0,73€

Heating

demand

52

kWh/m2a

Electric grid

33%

Comp. chiller (8 kW)

4.570 €

Power price

0,17€

Cooling

demand

34

kWh/m2a

COP

ammonia

0,6

Installation chiller

400 €

Gas price escalation

14%

DHW (4 persons) Solar fraction

3300

kWh/a

67%

COP compr.

2,5

Storage tank

Gas boiler (installed) CPC collectors (installed)

1.500 €

1.300 € 400 €/m2

Power price escalation

10%

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.

Table 2. All-year system for heating, cooling and DHW supply.

Ammonia-based system

Comp.-based system

Solar collector field

30

m2

0

m2

Total investment

21200

7770

Gas consumption

6842

kWh/a

16102

kWh/a

Power consumption

873

kWh/a

2977

kWh/a

Net present Value 20 a

75067

143616

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.

image457

NH3 — Comp.

Years

Fig. 6. Differential net present value between ammonia and compression chiller. Payback time.

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.

The T-s diagram

Usually, three groups of fluids are distinguished: dry (positive slope), isentropic (zero slope) and wet (negative slope) fluids. For low temperature organic Rankine cycle systems and particularly with low power output levels, wet fluids like water, methanol or ethanol are not suitable. This category of fluids requires a superheat in order to avoid moisture after the expansion process. Fortunately, there are dry and isentropic fluids that do not exhibit excessive moisture after the expansion of the vapor through the expansion device. When isentropic fluids are used, isentropic expansions of saturated organic vapors result in saturated or superheated vapor regions so that erosion of the blades is avoided. Dry fluids with high positive slopes yield low efficiencies. In order to overcome this drawback, a regenerative heat exchanger should be used in parallel with the evaporator and condenser to raise the temperature of the liquid entering the evaporator. An ideal working fluid for ORCs is likely to have a vertical vapor saturated line and a vertical liquid saturated line so that all the heat input transfer should take place during the phase change and the heat rejection at the condenser occurs at the minimum temperature [1, 2, 6]. Fig.1 shows the T-s diagram for an ideal working fluid.

ik

Подпись: 1

Подпись: 3 Подпись: 2

T [K] [15]

s [kJ/kg. K]

Fig. 1: T-s diagram for the ideal working fluid

Substance

Physical data

Safety data

Environmental data

Molecular Mass (kg/kmol)

a T

1bp

(°C)

b t

crit

(°C)

cP

crit

(MPa)

ASHRAE 34 safety group

Atmospheric life time (yr)

dODP

eGWP (100 yr)

RC318

200.03

-6.0

115.2

2.778

A1

3200

0

10250

R600a

58.12

-11.7

135

3.647

A3

0.019

0

~20

R114

170.92

3.6

145.7

3.289

A1

300

1.000

10040

R600

58.12

-0.5

152

3.796

A3

0.018

0

~20

R601

72.15

36.1

196.5

3.364

0.01

0

~20

R113

187.38

47.6

214.1

3.439

A1

85

1.000

6130

Cyclohexane

84.16

80.7

280.5

4.075

A3

n. a

n. a

n. a

R290

44.10

-42.1

96.68

4.247

A3

0.041

0

~20

R407C

86.20

-43.6

86.79

4.597

A1

n. a

0

1800

R32

52.02

-51.7

78.11

5.784

A2

4.9

0

675

R500

99.30

-33.6

105.5

4.455

A1

n. a

0.738

8100

R152a

66.05

-24.0

113.3

4.520

A2

1.4

0

124

R717 (Ammonia)

17.03

-33.3

132.3

11.333

B2

0.01

0

<1

Ethanol

46.07

78.4

240.8

6.148

n. a

n. a

n. a

n. a

Methanol

32.04

64.4

240.2

8.104

n. a

n. a

n. a

n. a

R718 (Water)

10.2

100

374

22.064

A1

n. a

0

<1

R134a

102.03

-26.1

101

4.059

A1

14.0

0

1430

R12

120.91

-29.8

112

4.114

A1

100

1.000

10890

R123

152.93

27.8

183.7

3.668

B1

1.3

0.020

77

R141b

116.95

32.0

204.2

4.249

n. a

9.3

0.120

725

R245fa

134.05

15.3

154.1

3.64

B1

8.8

0

820

R236fa

152.0

-1.4

124.0

3.20

209

0

6300

R227ea

170.0

-17.5

102.0

2.95

36.5

0

2900

a Tbp : Normal boiling point; b Tcrit: Critical temperature; c Pcrit: Critical pressure;

1

2

3

4

5

6

7

8

9

10

11

12

13

14

15

16

17

18

19

20

21

22

23

dODP: Ozone depletion potential, relative to R-11; eGWP: Global warming potential, relative to CO2. n. a: non available

Compact self-supporting construction

Подпись: Fig. 3. Internal construction of the adsorption chiller. All heat exchangers are made of the widespread tube and fin construction. About 50 % is empty in order to provide enough space for vapour flow.

One of the major problems of the construction of adsorption chillers is the need to design a vacuum vessel containing the adsorber heat exchangers, the evaporator and the condenser. Typical design rules for vacuum vessels result in very heavy and voluminous constructions due to the demands of stability against atmospheric pressure. In order to overcome this disadvantage we developed a design, where the heat exchangers and the internal construction of the machine are used as a support for the vacuum-tight envelope. This dual-use allows to applying thin stainless steel metal sheet as vacuum containment (See fig.3). This light and very simple construction results in very high savings of material and volume and is a major step for the construction of an economical adsorption chiller.

Model description

Due to the alternating batch process the operation of the two sorptive heat exchangers is characterized by their transient behaviour. Therefore, a dynamic model of the sorptive heat exchanger was developed applying the modelling language Modelica and using the simulation environment Dymola. The dynamic model of the sorptive heat exchanger is using the Modelica Standard Library (version

2.2.2. ) in combination with the Modelica_Fluid (Beta version) library which is the commonly used library for thermofluid modelling in Modelica at Fraunhofer ISE. For the description of the moist air properties the Modelica. Media. MoistAir model was applied.

The two-dimensional numerical model is based on the following assumptions and simplifications:

• Modelling of one pair of adsorption and cooling channels

• Lumped, constant heat and mass transfer coefficients for adsorption and evaporation processes • No heat transfer and thus no heat losses to the surrounding

The sorptive heat exchanger model may be discretized in a desired number of volume elements. Each element comprises an adsorption channel model, a cooling channel model as well as a model of the heat exchanger wall.

Adsorption channel model

Both the adsorption and cooling channel models are based on finite volume elements. Applying the Modelica_Fluid standardized connectors, volume elements may be connected by fluid ports in which the medium variables of state pressure and enthalpy, the medium mass flow rate and in case of multi­substance media such as moist air the component mass fraction are handed over.

Подпись: dmw dt image078 Подпись: [kg/s] (equation 1)

The sorption material and therefore the physical description of the sorption process is included in the adsorption channel model. The driving force of the adsorption process is the pressure difference between the partial pressure of vapour in the air volume pv and the equilibrium pressure of the sorption material pequ. The sorption process is mathematically described by the following linear interrelation:

in which flsorp (in the order of 10-8 kg/Ns) is the global mass transfer coefficient describing the

kinetics of water uptake by the sorption material. In order to solve the equation the equilibrium pressure pequ must be determined from the sorption equilibria. To this purpose a generalized form of Dubinin’s theory of volume filling is applied and implemented in a sorption material package for the use in Modelica [2]. The interaction between the sorption material and the air volume is described by postulating the balance equations of mass and energy via a set of differential equations. In the adsorption channel model the heat of adsorption is released in the sorption material, raising its temperature. Cooling of the sorption material is then calculated by convective heat transfer to the process air bulk flow and via conduction through the separating heat exchanger wall. To this purpose, the heat exchanger wall and the adsorption model volume element are connected via a heat port in which the state variable temperature and the flow variable heat flow rate are handed over.

Heat exchanger wall model

The heat exchanger wall model is a predefined Modelica model including the thermal mass of the aluminium wall and the heat conduction through the wall. It is connected to the cooling channel and the adsorption channel models via heat ports.

Cooling channel model

Подпись: dmw dt

Подпись: вevap • A PV (T, P) - Psat (Tmat )]
Подпись: [kg/s] (equation 2)

The cooling channel model is structurally similar to the adsorption channel model. The evaporation model again follows a linear driving force approach:

The cooling channel wall is always covered by a thin water film from which water evaporates into the air flow. Water is introduced to the volume element at a constant temperature and assumed to instantaneously reach the surface temperature which is a valid assumption for a thin water film [3]. Knowing the evaporation rate determined by equation 2 and taking account of the convective heat transfer rate between the cooling channel air flow and the heat exchanger wall, the balance equations of heat and mass conservation are solved.

Protocol for the identification of transfer parameters

The reactive bed is compacted in the reactor between the gas diffuser which is, in that case, fixed by rigid wedges, and the heat exchanger wall. A controlled pressure is applied to the coolant and therefore under the heat exchanger wall. During reaction, the coolant level and the displacement sensor give information on the relative position of the wall.

The reactions can occur after compacting. Thermodynamic constraints were fixed, i. e. the heat exchanger and evaporator/condenser temperatures were chosen according to thermodynamic equilibrium of solid/gas reaction and liquid/gas reaction [1].

The overall reaction advancement X is the number of mole of hexahydrated salt (which is deduced from the level in the liquid reacting water tank) divided by the total number of salt moles in the block. X is time dependant, and it is include between 0 and 1.

To simplify the identification of heat and mass transfer parameters, some constraints were applied:

• Temperature and pressure are fixed in the stability domain of the reactive salt i. e. at the boundaries of the reaction (X = 0 and X = 1).

• Transfers are assumed 1D because of the small ratio thickness / diameter of the sample, from 10 to 50 mm / 300 mm. Moreover, the external boundary is thermally insulated and airtight.

The measurements from the heating wire, thermocouples in the reactive bed and fluxmeter were used to identify the effective thermal conductivity. This identification uses the Fourier’s law in steady state.

Owing to the range of particle size and working pressure, the diffusion through the porous media is controlled by Darcy’s law. Nevertheless, the Knudsen diffusion could be envisaged. There are 3 categories of flow according to the mean free path of the fluid through the pores of the porous media.

In our case, the mean free path is close to the mean diameter of pores. The convective flow and the diffusive flow are not negligible because of Darcy’s law and Knudsen’s law respectively, equation 2.

Подпись: (2)к dp Dk dp p dz p dz

Подпись: dt Подпись: A. P fi+b 1 dz { p I p )dz) Подпись: (3)

The Darcy’s law is modified by the introduction of the Klinkenberg’s coefficient b which depends on the Knudsen diffusivity Dk, the dynamic viscosity g, and the permeability k. The pressure evolution through the reactive porous media is given by this equation combined with the mass balance equation, equation 3. A numerical solution has been implemented using the commercial software: COMSOL®.

The determination of the permeability and the Klinkenberg’s coefficient was carried out in transient state. First, the gas tank volume and the gas volume above the reactive block are disconnected. Each of them is filled by steam up to the pressure pg and pd, respectively for the gas storage tank and the dead volume in the reactive bed. Then, the two volumes are connected together. The experimental evolution of the two pressures versus time is compared with the simulated one. The permeability and the Klinkenberg’s coefficients are determined by these pressure evolutions. Each volume pressure is simulated according to equation 3. The typical simulations are presented in figure 3. The first equalization of pressure Peg characterizes the gas tank volume and the dead volume above the reactive bed. As this pressure equalization occurs in less than 1 ms, we assume that it does not involve the porous volume of the reactive bed. After this first step, the gas diffuses through the porous volume of the reactive bed, and the pressure evolves from (peg, teg) to (pec, tec). The identification of k and b is carried out between these two points.

Pec

 

image209

Pg

Подпись: tec < 1minteg< 1 ms

Figure 3 : Typical evolution of pressure versus time in transient state

Operational Outcomes of the DESODEC Solar Driven. Cooling/Heating system

A. Hambarian* and W. Akunyan

American University of Armenia (AUA), Engineering Research Center (ERC), 40 Marshall Baghramian,

Yerevan 0019, Armenia
Corresponding Author, arthambr@aua. am

Abstract

Being one of the early ones in the world, this is the first solar driven desiccant cooling system in Armenia and NIS. The article discusses issues related with the operation of the system since its commencement in February 2002. Topics involve: A. Efficiency issues, covering influence of the perception of humidity, quality of the windows of the space cooled, monitoring results; B. Reliability, including malfunctions related to heat recovery wheel drive, fluid (antifreeze) circulation, overheating, sediment formation on three-way valves; solar heat exchanger breakage due to ice formation; C. other operational issues: PV system integration; filters problem.

Keywords: solar hot water, solar driven cooling, desiccant cooling, solar PV.

1. Introduction

“Design And Installation of a Solar Driven Desiccant Cooling Demonstration System” (DESODEC) project funded by INCO Copernicus Program of the EU involved more than three years of collaborative work of five partners: the DER/INETI (Portugal), ISE/Fraunhofer (Germany), InterSolarCenter (Russia), Contact-A and Engineering Research Center of the American University of Armenia. As a result, with the energy provided by the Solar Water Heating (SWH) collector field installed on the rooftop of the AUA building and operation of the Desiccant Evaporative Cooling (DEC) machine, heating and cooling capability is achieved in a 154 seats auditorium of AUA.

In this article we summarize 6 years experience of the operation of the DESODEC system.

2. Description of the system

The solar field, fig. 1 consists of 32 flat plate hot water collectors with total surface area of 64 m2 providing up to 40 kW of power equivalent. The solar field and the panels are designed by DER/INETI together with Contact-A and manufactured by SolarEn LLC company in Armenia. Desiccant cooling machine comprising desiccant wheel, heat recovery wheel, humidifiers, fans and control and monitoring devices are designed by ISE/Fraunhofer and manufactured mostly in the USA and Germany. The system can deliver up to 8000 m3 cooled air with cooling power equal to 60 kW in the desiccant cooling mode to the Auditorium (fig. 2) [1-4].

The system has a 3 ton stratified storage tank designed to provide a heat buffer and a temperature difference for efficient thermodynamic cycle.

image297

Figures 2. a. and b. describe the solar driven desiccant cooling well-known operating principle employed: air that is pre-desiccated by the desiccant absorption (evaporative cooling) wheel is cooled by humidifiers. The cooling efficiency of humidifiers is considerably elevated due to the air desiccation. Solar energy is used for regeneration of the desiccant absorption wheel. To increase the efficiency of the process, a heat recovery wheel is used to return useful energy (cold or warmth) from the exhaust channel.

image298

image299

Fig. 3. a. Components of the DEC machine; b. Temperature-humidity diagram of the system operation.

Thus, in this system, a cooling or heating function is performed by a solar driven machine, that uses up to 40kWheat low potential solar energy provided by solar hot water panels, and an average of 5kWelectric, for driving the balance of the system.

3. Operation

Figures 4. A. and b. show the screenshots of the system controls interface that have comprehensive information on the operation. One can see that water temperature of 94°C is required to output 16.6°C air with relative humidity equal to 52%, and at 70.7°C yields 23.8°C and 56% respectively.

image300
b.

Space Heating, Cooling and Hot Water Load Profiles

The house has two storeys with a total floor area of 213 m2 and has well insulated reverse brick veneer external walls. The windows of the house are double glazed 4 mm clear with 12 mm argon filled and low E film. It has solid construction external doors.

image405

Figures 2 — 4 show monthly profiles of cooling, space heating and domestic water heating loads for the house exposed to Adelaide, Brisbane and Sydney typical year weather data.

image406

Fig. 3. Monthly cooling, DHW and SH load profiles — BRISBANE

image407

Fig. 4. Monthly cooling, DHW and SH load profiles — SYDNEY

Monthly cooling and space heating loads for the house were calculated using AccuRate package [9]. The house exposure to different weather data results in 6.9, 7.7, and 7.6 star ratings of the house in Adelaide, Brisbane and Sydney. The star rating is a measure of energy performance of the house based on the fabric and form, not including the effect of the cooling / heating system used in the house. This rating system is mandatory for new Australian houses.

Adelaide is a semi arid region with low humidity, mild winter and hot summer. Brisbane is warmer with relatively high humidity, mild summer and short heating period. Sydney is cooler and less humid and has a slightly longer heating season than Brisbane. These are clearly illustrated in Figs. 2 — 4 above and Fig. 5 which shows the montly latent load as a percentage of the total load.

image408

J FVIAMJ J ASOND Fig. 5. Latent load as a percentage of total cooling loads

The monthly figures for DHW loads were derived from Australian Standard 4234 [10] for a medium size system which provides seasonal (monthly) load profile, hourly load profile and monthly cold water temperatures. Monthly space heating (SH) and cooling loads were estimated using an Australian building energy rating software, AccuRate.

Monthly profile total loads are shown in Fig. 6. As seen, Adelaide monthly total loads are much higher than Brisbane and Sydney loads which are very similar. As expected, monthly loads peak in peak summer and winter months and level during spring and autumn.

Подпись:4000

І 3000

о"

<

З 2000

I—

£ 1000

о

2. Results and Discussions

For the analysis, the main metrics used to illustrate the system performance are the solar fraction (SF) and total useful heat. Solar fraction is defined as the fraction of solar energy used to satisfy the load. The load not met by solar energy is handled by an auxiliary system.

Подпись: Fig. 7. Solar Fraction as a function of collector area (Tank size: 315 L, 450 L, collector slope: 25° and 45°)

Figure 7 shows the solar fractions as functions of collector area for systems serving houses in the three cities. Two hot water tank sizes (315 L and 450 L) were studied; the collector were tilted at 25° and 45 °C.

As shown, for Brisbane and Sydney, a collector area of 6 m2 can provide more than 80% of the heat required for cooling, heating and hot water. On the other hand, for Adelaide, much lower solar fractions are delivered by the same system size without much improvement in the solar fraction when a larger collector area is used. Two other interesting findings revealed by Fig. 7 are: (1) a tank size of 315 L performs as well as 450 L tank, and (2) collector slope between 25 — 45 is insensitive to system performance.

Fig. 7 shows clearly that for Brisbane and Sydney, application of an integrated thermal system is appropriate. On the other hand, due to drier climate (Fig. 5), the application of such a system in Adelaide is not leading to large energy savings.

Подпись: ■ BRI-COMBI ■ BRI-INTEGRATED

Futher analysis shows that thermal performance of an integrated thermal system is as good as the combisystem performance (heating and hot water only) in terms of solar contribution to handle the total load (Figs. 8 — 9).

2 3 4 6 8 10

Collector Area, m2

Fig. 8. Solar Fraction for Combisystem (heating and hot
water only) and Integrated System — Brisbane

4 image4126 8

Collector Area, m2

Fig. 9. Solar Fraction for Combisystem (heating and hot
water only) and Integrated System — Sydney

image413

The system performance can also be measured in terms of monthly useful energy delivered by the system to satisfy the loads. This is illustrated in Figs. 10 — 11 which compare the useful heat delivered by an integrated thermal system and a combisystem with a 6 m2 collector area installed in Brisbane and Sydeny. As shown, the integrated system delivers much higher useful energy than the combisystem, particularly during the summer months. For a collector area of 6 m2 and a 315 L storage tank, the annual useful heat delivered by the intgerated system for Brisbane and Sydney are 17.0 GJ and 11.6 GJ, respecticely, compared to 15.7 GJ and 11.9 GJ delivered by the combisystems.

3. Conclusion

An analysis of the thermal performance of an integrated heating system for the provision of hot water, space heating, cooling and dehumidification has been presented. It was observed from numerical investigation that for a location like Adelaide with semi arid climate, such a system will work with a minimum role for the solar liquid desiccant cooling sub-system. In such a case, a combisystem or the use of an absorption cooling system is more suitable. On the other hand, for locations such as Sydney and Brisbane, such systems are appropriate. An integrated thermal system with a 315 L tank, and a 6 m2 evacuated tube collector area can serve typical house in either Brisbane or Sydney.

References

[1] Australian Greenhouse Office, Australian Residential Building Sector — Greenhouse Gas Emissions 1990 — 2010 — Executive Summary Report, 1999.

[2] Lee, T., Ferrari, D., Donnelly, E. (2003): Active Solar Space Heating for a Cool Temperate Climate: A Case Study, 41st ANZES Solar Energy Society Conference, Melbourne.

[3] Saman W. Y. and Mudge, L. 2002, Development, implementation and promotion of a scoresheet for household greenhouse gas reduction in South Australia, Australian Greenhouse Office.

[4] Henning, H.-M. (2004): Solar Air-Conditioning of Buildings, EuroSun2004, Freiburg, Germany, 24.­26.6.04.

[5] Krause, M., Saman, W. Y. Vajen, K. (2005), Regenerator Design for Open Cycle Liquid Desiccant Systems — Theoretical and Experimental Investigations, International Conference on Solar Air Conditioning, Staffelstein, Germany, 6.-7.10.2005.

[6] Krause M., Saman W. and Vajen K. (2006): Optimisation of a regenerator for open cycle liquid desiccant air conditioning systems, Proc. 2006 Eurosun Conference, Glasgow, Scotland.

[7] TRNSYS 16 Reference Manual, 2004.

[8] Halawa, E., Bruno, F., Saman, W. (2007): Potential Application of Combisystem for An Australian Climatic Region, ISES Solar World Congress 2007, Beijing, September.

[9] Hearne Website — http://www. hearne. co. nz/ — Residential Building Thermal Performance Assessment (2006) — viewed 20 June 2007.

[10] Australian Standard AS 4234-1994: Solar Water Heaters — Domestic and Heat Pump — Calculations of Energy Consumption.

New control by cooling water adjustment

Figure 3 shows a scheme of the new developed control strategy. In this case, driving temperature can rise as high as the chiller control allows. Thus, high irradiation can be used and no driving energy or exergy is wasted. Cooling water temperature is not kept constant anymore but it is used as a control parameter. Base of the control strategy is the method of the characteristic equation — a simple model which yields the cooling capacity of an absorption chiller as function of the so-called characteristic temperature function AAt, a function of the temperatures of the external water circuits (hot, cooling, chilled water). According to this equation, cooling capacity does not change, if the characteristic temperature function remains constant. From the characteristic temperature function follows immediately how changes of one external temperature can be compensated by control of the other external temperatures. That means for example a higher driving temperature can be compensated by increasing the cooling temperature without a change in the chilled water temperature.

Подпись: temperature tAin yields tAin image504 Подпись: Q E - 0,9 0,42 t tCout Подпись: If a constant

In [2] the method of the characteristic equation was presented. The adapted characteristic equation for the 10 kW absorption chiller is Q E = 0,42 • AAt + 0,9 with the adapted characteristic temperature function AAt’ = tG — 2.5 • tAC +1.8 • tE . Temperatures tG, tA, tC, tE are the arithmetic mean values of inlet and outlet of the external water loops at generator, absorber, condenser and evaporator, tAC is the arithmetic mean value of tA and tC. Solving of AAt’ for the cooling water inlet

chilled water outlet temperature tEout is required, QE is substituted by VE • p • cp • (Ein — tEout) and tEout is set to the required value, e. g. 15°C if a chilled ceiling is applied. The measured values of the other temperatures and the chilled water flow rate are continuously inserted and tAin recalculated hereupon. I. e., the characteristic equation model defines the necessary cooling water temperature. A standard PID-controller in a closed loop is used to control the fan frequency to provide this temperature. Via a frequency converter the speed of the cooling tower fan is set according to the output of the controller.

Summing up, this control strategy can manage unsteady driving temperatures. The cooling water inlet temperature is adapted to hold chilled water temperature constant. First successful laboratory tests of this control are presented in [8]. This paper focuses on comparison of the two control strategies by simulation.