Category Archives: EuroSun2008-7

Experimental set-up description

A single characterisation apparatus, described in figure 2, was designed and built to characterize the changing porous media. During synthesis reaction the salt reacts with steam and swells. The mass transfer through this porous media will be affected if the overall volume of the reactant remains constant. Moreover, the low working pressure (from 1000 to 10000 Pa) can lead to significant mass
transfer limitation if the permeability is very low. Therefore, to avoid mass transfer limitations, the working volume must grow as the reactive block swells. On the other hand, the overall kinetics, X, depends on heat and mass transfers. Thanks to this single apparatus, the overall kinetics, the swelling and the transfer coefficients will be linked together.

image204
This experimental set up is described in figure 2. The disc located on the upper face of the reactive composite block, is a gas diffuser and a heat conductor as it includes a heating wire. On the lower face, there is a fluxmeter and a heat exchanger wall which can move. This very special device allows the reactive composite volume to swell during the reaction. The gas diffuser and the heat exchanger wall are connected to a displacement sensor. Several rigid wedges or springs are located between the gas diffuser and a fixed plate above it. Four thermocouples are in the reactive bed at different depth, and a pressure sensor measures the reactor pressure. The gaseous part of this reactor is connected to a gas tank and to an evaporator/condenser, whose pressure is measured by another pressure sensor. The evaporator/condenser is also connected to the liquid water tank, which includes a level sensor to measure the amount of reacting water.

Another liquid tank is used to control the location of the heat exchanger wall, under the lower face of the composite. The coolant fluid in this tank is connected on one side to this heat exchanger plate and on the other side to a compressed nitrogen bottle. The level of coolant in this tank gives information on the position of the exchanger plate. The coolant pressure is constant and it is given by a pressure sensor; the working volume of the reactant is not constant.

Simulation Based Optimisation of a Newly. Developed System Controller for Solar Cooling. and Heating Systems

Dirk Pietruschka1. Uli Jakob[8], Vic Hanby[9], Ursula Eicker1
1 Centre for Applied Research of Sustainable Energy Technology — zafh. net
Stuttgart University of Applied Sciences, Schellingstrasse 24, D-70174 Stuttgart, Germany
Tel.:+49/711/8926-2674, Fax: +49/711/8926-2698, Email: pietruschka@zafh. net
2 SolarNext AG, Nordstrasse 10, 83253 Rimsting, Germany
Tel.:+49/8051/6888-403, Fax: +49/8051/6888-490, Email: uli. jakob@solarnext. de
3 Institute of Energy & Sustainable Development, De Montfort University, Leicester LE1 9BH, U. K.

1. Introduction

The performance of solar driven cooling systems strongly depends on the implemented control strategies of the absorption cooling system including the chiller, the cooling tower, the installed cold distribution system and the solar collector field [1-4]. High electricity consumption caused by suboptimal control in combination with low solar fractions through insufficient system design are critical for the environmental and economical performance of installed absorption cooling systems (ACM), especially if they are compared to highly efficient electrical driven compression chillers [1, 4]. For the control of such complex systems often different independently operating component controllers are combined in one solar cooling installation resulting in a more ore less optimal system control. To overcome these problems and to allow the implementation of advanced control strategies, the development of combined system controllers are required, which are able to control all components of a solar cooling and heating system. Such a combined system controller has been developed by the SolarNext AG in Rimsting, Germany. A detailed dynamic simulation model in INSEL [5] has been used for the development of the advanced control strategies. To ensure optimal system operation the implemented advanced control strategies have been implemented and tested in the dynamic simulation environment. In the present paper this method is demonstrated for the optimisation of the start-up process through improved storage charge management of a solar driven chillii® Solar Cooling System, which has been installed in an office building in Rimsting, Germany. The design methods used for the developed control strategies, the controller algorithms found and the results of the controller tests in the simulation environment are presented and discussed in detail.

Table 1: Analysed storage charge and discharge cases

 

Case 1

 

Case 2

 

Case 3

 

Case 4

 

Control mode description

 

2000 l hot water storage, full volume always used

 

2000 l hot water storage with bypass Control modes:

1. Bypass storage completely; 2. Bypass hot water supply, feed in generator return; 3. Use full storage volume; 4. Charge storage in case of ACM shut down

 

Combination of case 2 and 3: Partitioned 2000 l storage with bypass

Control modes:

First control modes of bypass applied to upper storage part, then control modes of partitioned storage used

 

Partitioned 2000 l hot water storage for early system start-up, with 300 l on top and 1700 l below

Control modes:

1. Use upper storage part only

2. Use full storage volume

 

image277

Bypass mode:

^ see case 2

feed return flow of generator in storage tank if Tgin > 85°C

Switch to partitioned storage mode (case 3) IF Tsti, top > 75°C

-> Partitioned storage mode see case 3

Switch back to bypass mode at the end of the day if Tstitop < 75°C

 

-Collector pump

a) primary circuit

Bypass mode

Switch on IF Tcoll > 80°C;

Switch off IF THXp, in LT.

Tst1,bot + 5K; Tmin > 180 s Full storage mode

Switch on IF Tcoll > T st, bot + 10K; Switch off IF THXp, in. LT.

Tst, bot + 5K; Tmin > 180s

b) secondary circuit

Bypass mode:

Switch on generator pump IF THXp, in > 75°C and Tcst, mid > Tc, sup; Switch off IF Tg, in. LT. 65 or IF Te, out < 10°C ’

Charge hot storage tank in case of generator pump off

Feed return flow of generator in storage tank if Tg, in > 85°C

Switch to storage mode (case

1) IF Tst, top > 75°C

Switch back to bypass mode if

THXs, out < 68°C and generator

pump on

— ACM, evaporator and absorber/condenser pump

Switch on IF Tg, out > 65°C

and IF Tcst, mid > Tc, sup

Switch off IF generator pump

off and Te, out < Tcst, top — 2K or IF Teout <10 °C

 

image278

image279image280image281image282

Подпись: 21st June 2008 Hot summer Day with single clouds Average ambient temperature 22.1°C Average ambient humidity 60.9 %rH Total solar radiation 7 016.9 Wh/m2 Total cooling energy demand 81.5 kWh 10th July 2008 Hot summer day with almost clear sky Average ambient temperature 23.3 °C Average ambient humidity 50.7 %rH Total solar radiation 8 235.1 Wh/m2 Total cooling energy demand 115.1 kWh 11th July 2008 Hot summer day, some clouds in the early morning and thunderstorm in the afternoon Average ambient temperature 21.7°C Average ambient humidity 61.9 %rH Total solar radiation 5 461.4 Wh/m2 Total cooling energy demand 75.9 kWh image284 image285

Table 2: Analysed typical hot summer days

Подпись: 1000 900 800 700 § 600 C o 500 400 300 200 100 0 Подпись:Подпись:Подпись:Подпись:Подпись:Подпись:Подпись: tc,out meas [°C] -tst.top [°C] -Gt [W/m2] Подпись:image295100

90

80

„ 70

О

з

го 50

q3

40

<D

H 30 20 10 0

5. Results and discussion

The results for all analysed storage charge and discharge cases (case 1-4) and daily weather conditions are summarised in Table 3 and 4. If always the full storage volume is used (case 1) the start-up time of the absorption chiller varies between 11:03 and 11:38am depending on the summer day regarded. This perfectly corresponds with the commonly expected performance of solar driven absorption cooling systems. However, as a result of the late start-up 27 to 39% of the cooling load of the building can’t be covered by the system. As expected for case 1, the best results are obtained for the 10th July (cloudless day) and the worst result for the 11th July (some clouds in the early morning and thunderstorm in the early afternoon). In Case 2 with bypass of the hot storage, the start-up times can be significantly reduced by more than 1h 40 min for all of the three analysed days. The earliest start-up time of 9:23am is obtained on 10th July followed by 9:26am on 11th July and 9:40am on 21st June (single clouds). Due to the early system start-up, the cooling load which can’t be covered is also significantly reduced and varies between 11% and 14%. The partition of the storage tank in case 3 leads also to much earlier system start-up times but does not reach a higher coverage of the cooling load compared to case 2 on any of the analysed days. The best overall performance is reached for case 4 with combined bypass and partitioned storage control. Compared to case 1 the part of the cooling load which can’t be covered is reduced to 14 % (10th June), 10% (21st June) and 16% (11th July) with almost the same start-up times as in case 2. If the produced cooling energy is regarded, the advantage of the improved storage charge and discharge control in case 4 compared to case 1 becomes even more visible with 18% more cooling energy production on 21st June, 23% more cooling energy production on 10th July and 33% more cooling energy production on 11th July.

The thermal and electrical COP is very slightly influenced by the analysed storage charge and discharge control cases but is significantly influenced by the type of summer day. The highest thermal COP of the absorption chiller of 0.69 and the highest electrical COP of complete solar cooling system of 8.5 are reached on 10th July with almost clear sky due to the highest driving temperatures. For 21st June and 11th July significantly lower thermal COP of 0.64 to 0.66 and electrical COP of 7.1 to 7.7 are reached, which is mainly caused by the overall lower driving temperatures at generator inlet. The electrical COP of the solar cooling system includes the electricity consumption of all components and pumps (ACM, cooling tower (fan speed control) and

all pumps (generator, evaporator, absorber/condenser and primary and secondary collector pump), only the distribution pump and fan coils are not considered.

Table 3: Results Related to Heating Energy Consumption and Cooling Energy Production

Case

Start-up time ACM

QS

Qh, st_in

Ps

Qh, ACM

Qh, stored

Qc, req

Qc, ACM

Qc, stored

Cooling energy not covered

[h:min]

[kWh]

[kWh]

[%]

[kWh]

[kWh]

[kWh]

[kWh]

[kWh]

[kWh]

[%]

10th July, hot summer day with almost clear sky

1.1

11:04 AM

592.7

223.4

38%

135.7

77.8

114.7

92.1

7.7

31.2

27%

2.1

9:23 AM

592.7

228.8

39%

148.0

70.8

114.7

101.4

1.4

15.5

14%

3.1

9:43 AM

592.7

233.0

39%

154.9

70.9

114.7

105.4

7.6

17.9

16%

4.1

9:23 AM

592.7

232.5

39%

150.3

73.2

114.7

103.1

3.0

15.5

14%

21st June, hot summer day with single clouds

1.2

11:38 AM

481.5

166.6

35%

99.8

57.3

81.3

65.6

6.3

22.9

28%

2.2

9:40 AM

481.5

172.2

36%

120.3

43.3

81.3

79.9

6.7

9.0

11%

3.2

9:52 AM

481.5

174.0

36%

119.3

48.3

81.3

79.2

6.9

10.0

12%

4.2

9:40 AM

481.5

177.6

37%

121.9

47.3

81.3

80.8

6.9

8.3

10%

11th July, hot summer day, some clouds in the early morning and thunderstorm in the early afternoon

1.3

11:07 AM

399.3

134.9

34%

80.5

45.2

75.9

51.1

3.6

29.2

39%

2.3

9:26 AM

399.3

138.8

35%

102.3

28.4

75.9

67.0

2.9

12.6

17%

3.3

9:45 AM

399.3

141.6

35%

95.6

36.5

75.9

62.2

3.1

15.2

20%

4.3

9:26 AM

399.3

136.0

34%

103.9

22.4

75.9

68.0

3.6

12.4

16%

Legend:

Qs Solar irradiation on the collector plane (gross area)

Qh, st_in Heating energy stored in the hot storage at the end of the day

Qc, acm Cooling energy produced by the ACM

Qc, stored Cooling energy stored in the cold storage at the end of the day

Qh, st_in Solar heating energy storage input

ns Efficiency of the solar system

Qc, req Required cooling energy

Qh acm heating energy used by the ACM

Table 4: Results Related to ACM Operation

Case

ACM

operation

time

[h:min]

Qh

[kWh]

Qel

[kWh]

Qc

[kWh]

Qh

[kW]

Qc

[kW]

COPth

[-]

COPel

[-]

tg, in [°C]

tg, out

[°C]

te, in

[°C]

te, out

[°C]

ta, in

[°C]

о О о

-і—»

10th July, hot summer day with almost clear sky

1.1

5:54

135.7

10.8

92.1

85.2

15.6

0.68

8.51

86.4

74.6

17.6

11.4

25.9

32.1

2.1

7:30

148.0

11.9

101.4

22.6

14.8

0.68

8.49

79.4

69.4

17.6

11.7

25.4

31.2

3.1

7:10

154.9

12.4

105.4

23.5

14.8

0.68

8.47

80.4

70.0

17.5

11.6

25.5

31.3

4.1

7:48

150.3

12.3

103.1

21.9

14.5

0.69

8.39

78.2

68.5

17.6

11.8

25.3

31.0

21st June, hot summer day with single clouds

1.2

4:44

99.8

9.3

65.6

22.1

13.7

0.66

7.07

78.3

68.5

17.1

11.6

25.5

31.0

2.2

6:07

120.3

11.0

79.9

20.0

13.1

0.66

7.26

74.4

65.5

17.1

11.9

25.1

30.2

3.2

6:18

119.3

11.2

79.2

19.3

12.6

0.66

7.07

73.0

64.4

17.0

12.0

24.9

29.9

4.2

6:26

121.9

11.4

80.8

19.1

12.6

0.66

7.06

72.5

64.0

17.0

12.0

24.9

29.9

11th July, hot summer day, some clouds in the early morning and thunderstorm in the early afternoon

1.3

3:43

80.5

6.8

51.1

23.6

13.8

0.64

7.50

79.7

69.3

17.4

11.8

25.9

31.4

2.3

5:11

102.3

8.7

67.0

20.1

13.0

0.65

7.69

75.0

66.0

17.1

11.9

25.4

30.6

3.3

5:03

95.6

8.4

62.2

19.8

12.9

0.65

7.42

74.7

65.9

17.0

11.9

25.4

30.5

4.3

5:28

103.9

9.5

68.0

19.3

12.5

0.65

7.18

72.8

64.2

17.2

12.2

25.2

30.1

Legend:

Qel

Electricity used by the whole solar cooling system, including the ACM, cooling tower and all pumps

Qh /Qc

heating energy used / cooling energy produced by the ACM

COPel

Electrical coefficient of performance of the whole solar cooling system

COPth

Thermal coefficient of performance of the ACM

tg, in/tg, out

Average Generator inlet / outlet temperature

te, in/te, out

Average Evaporator inlet / outlet temperature

ta, in/tc, out

Average Absorber inlet / condenser outlet temperature

Qh ‘ Qic

Average heating / cooling power

image296

To visualise the behaviour of the developed control algorithm, one example of detailed simulation results are shown Figure 3 for the most complex control in case 4 with combined bypass and partitioned storage control for the 21st of June. This chart clearly shows the different control phases, with the start-up of the primary collector pump at 80°C collector temperature which is then switch off again due to a sudden drop of the collector outlet temperature caused by the cold water stored in the tubing. The second start up of the collector pump is also followed by a start-up of the generator pump in bypass mode which is turned off again as the generator inlet temperature drops below 65°C. The primary collector pump remains in operation since the collector outlet temperature at heat exchanger inlet is above the storage bottom temperature. The generator pump is switched on again as soon as the collector outlet temperature at heat exchanger inlet reaches 75°C. After the generator outlet temperature reaches 65°C the ACM with evaporator and absorber pump is switched on and the ACM starts to produce cold water which is fed into the cold storage tank. After more than one hour of operation the generator inlet temperature decreases below 65°C due to low solar radiation and the generator pump is switched off. The ACM with evaporator and absorber pump remains in operation as long as the evaporator outlet temperature is 1 K below the cold storage top temperature. After the generator pump is turned off, the upper part of the partitioned storage is charged by the collector and reaches a temperature above 75°C within a short time period of half an hour. Therefore, at the second start up of the ACM the control switches from bypass mode to partitioned storage mode. The ACM is then in operation until the evaporator outlet temperature decreases below 10°C which leads to a switch off of the generator pump followed by a switch off of the ACM including evaporator and absorber pump as soon as the evaporator outlet temperature is less than 1 K below the cold storage top temperature. Due to the switch off of the heating load the temperature in the upper partitioned part of the storage volume increases fast above 90°C which results in a switch from portioned storage mode to full storage mode for the rest of the day.

Figure 3: Detailed simulation results of case 4 for the 21st June 2008

Despite of the detailed analysis of the performance of the solar cooling system, the full dynamic simulation model was also very useful in the development phase of the controller algorithms. Several not obvious control errors could be detected and directly eliminated in the controller code without the necessity of time consuming on site or laboratory tests.

6. Conclusions

The focus of the present work is on the optimised control of solar driven absorption cooling systems for early system start-up in the morning and sufficient heat supply during operation. For the analysis of the behaviour of the developed control algorithms and their effect on the overall performance of the solar cooling system a detailed full dynamic simulation model developed in INSEL has been used. Four different control options for the storage charge and discharge control of a 15 kW chillii® Solar Cooling System installed in an office building of the SolarNext AG in Rimsting Germany have been analysed for three different typical summer days. In case 1 always the full hot water storage volume is used for the solar system, in case 2 a storage bypass is integrated for an early system start-up. In case 3 the storage is partitioned in an upper 300 l part and a lower 1700 l part and in case 4 the storage bypass is combined with a the partitioned storage tank. The developed control algorithms have been tested and improved in the dynamic simulation environment for different weather conditions. Several not obvious control errors could be detected and removed from the control code without the necessity of time consuming on site or laboratory tests. The proofed control codes of the four cases mentioned above were then used to analyse the overall performance of the solar cooling system for three different typical hot summer days. It could be shown, that through the implementation of a storage bypass the start up time can be significantly reduced by 1h 40 min in the worst and nearly 2 h in the best case. For the partitioned storage case, the start-up time is between 12 and 20 minutes later. On a cloudless summer day the start-up of the solar cooling system is at 9:23am (Bypass), 9:43am (partitioned storage) instead of 11:04am (full storage). The overall best performance is reached for case 4 with the combined bypass and partitioned storage control. If the produced cooling energy is regarded, the advantage of the improved storage charge and discharge control in case 4 compared to case 1 becomes clearly visible with 19% more cooling energy production on 21st June 2008 (day with some clouds) and 10th July 2008 (cloudless day) and 33% more cooling energy production on 11th July 2008 (some clouds in the early morning and thunderstorm in the early afternoon).

The thermal and electrical COP of the solar cooling system is only very slightly influenced by the different control modes but strongly depends on the analysed type of summer day. For a hot cloudless summer day (10th July) the thermal COP is a 0.69 and the overall electrical COP is 8.5 in the best case. Due to the lower available driving temperatures, the thermal COP is reduced on cloudy days (21st June and 11th July) to 0.64 in the worst case. The lower thermal COP results in longer operational hours at low cooling capacity and thereby reduces also the electrical COP which is 7.06 in the worst case.

References:

[1] Pietruschka, D., Jakob, U., Eicker, U., Hanby, V. “Simulation based optimisation and experimental investigation of a solar cooling and heating system”, Solar air conditioning, 2nd international conference, Tarragona, Spain, 2006

[2] Henning, H.-M. “Solar-assisted air-conditioning in buildings — a handbook for planners”, Springer — Verlag 2004, ISBN 3-211-00647-8

[3] Mendes, L. F, Collares-Pereira, M., Ziegler, F. “Supply of cooling and heating with solar assisted heat pumps: an energetic approach”, Int. J. Refrig. Vol 21, No.2, pp 116-125, 1998

[4] Kohlenbach, P. “Solar cooling with absorption chillers: Control strategies and transient chiller performance”, Dissertation Technische Universitat Berlin, 2006

[5] Schumacher, J. “Digitale Simulation regenerativer elektrischer Energieversorgungssysteme”,

Dissertation Universitat Oldenburg, 1991 www. insel. eu

System Description

The schematic of the integrated system is shown in Fig. 1. The main components of the system are: (1) solar collector and hot water tank, (2) absorber or dehumidifier, (3) desorber or regenerator, and (4) indirect evaporative cooler.

image403

(a) MODE 1: air conditioning and hot water heating

image404

(b) MODE 2: space and hot water heating

HWT — hot water storage tank, R — regenerator, A — absorber,
DHW — domestic hot water, CSS — concentrated solution storage,
DSS — diluted solution storage
Fig. 1. Schematic of the Integrated Thermal System

The system will provide hot water all year round, space heating during the heating season and cooling / dehumidification during the cooling season. The hot water service should be available all year round and the system operation can be divided into two different modes, Fig. 1: (a) cooling / dehumidification and water heating and (b) space and domestic hot water heating.

Mode 1: Space cooling / dehumidification and Water Heating

In this mode, the air is dehumidified by contacting it with the concentrated desiccant solution in the absorber before being cooled by an indirect evaporative cooler. The regenerator is used to reconcentrate the liquid desiccant solution using the hot water from the tank. The dehumidification system is assumed to operate when the air outlet temperature from the indirect evaporative cooler alone exceeds 27°C.

Mode 2: Space Heating and Hot Water Heating

For domestic water heating, the hot water stored in the tank is used to indirectly heat the mains water on demand. Space heating is provided by circulating water from the storage tank into the regenerator (used as a heat exchanger) in order to indirectly heat the air entering the building. Heat in the exhaust air from the building can be recovered by the absorber to preheat the outside air to be used for heating. In this mode, the absorber is inactive.

TNRSYS [7] was used as a simulation tool. For the purpose of analysis, an array of evacuated tube collectors were assumed to provide the heating required to satisfy the cooling (CL), space heating (HL) and domestic hot water loads (DHW). The system has the following instantaneous efficiency curve coefficients: a0 = 0.717, a1 = 5.472 kJ/hr. m2K, and a2 = 0.0306 kJ/hr. m2K2 [8]. The system has the following values of incident angle modifier (IAM):

Angle (deg).

0

10

20

30

40

50

60

70

80

90

Trans. IAM

1.0

1.0

0.99

.98

0.96

0.93

0.87

0.74

0.38

0

Long. IAM

1.0

1.02

1.08

1.18

1.37

1.4

1.34

1.24

0.95

0

Control strategies of solar cooling systems

1.1. Conventional control by hot water adjustment

The conventional control strategy includes a three-way valve in the hot water circuit between solar field and chiller or storage tank and chiller to mix the return flow from generator with the supply flow from the hot water source. It can be necessary to decrease the hot water inlet temperature when solar irradiation is high as with rising driving temperature and constant cooling water temperature, chilled water temperature would decrease — constant flow rates of all external water circuits assumed. This for example could cause condensation of water at chilled ceiling panels. In case of a low chilled water temperature level (e. g. 6/12°C) a drop of the chilled water temperature can cause a shut down of the chiller due to the danger of ice formation in the evaporator. In Figure 2, a scheme of this control strategy is presented.

image502

Fig. 2. Chilled water control by hot water adjustment.

This control strategy can only prevent chilled water temperature from too low values. Cooling water temperature is kept constant.

Simulation of an absorption based solar cooling facility using a. geothermal sink for heat rejection

R. Salgado*, A. Burguete, M. C. Rodriguez, P. Rodriguez

ITEA Research Group. Universidad Carlos III de Madrid, Departamento de Ingenieria Termica y de Fluidos.
Address: Avda. Universidad 30, 28911, Leganes (Madrid), Spain; Tel: (+34)916248884
Corresponding author: rsalgado@,ing. uc3m. es

Abstract

An important issue of solar cooling facilities based on absorption cycles and sometimes not giving the necessary attention is the recooling process of the absorber and condenser. This is critical in the overall behaviour of the facility because the condensation and absorption temperatures will affect the COP and cooling capacity of the chiller. Most of the time, the recooling process is made by using a wet cooling tower in a closed loop through the absorber and condenser. The use of a wet cooling tower gives good results in terms of cooling capacity and COP, but presents some health risk, like legionella, and its use is restricted to the industrial sector and places where water scarcity is not present. This paper presents the modification of the already validated TRNSYS simulation of a solar cooling facility, implementing a geothermal heat sink instead of the wet cooling tower in order to dissipate the heat generated internally in the absorption chiller. Simulation results shows that a geothermal heat sink composed of 6 boreholes of 100 meters of depth should be sufficient in order to substitute the wet cooling tower, for a typical Spanish single family dwelling.

Keywords: Solar cooling, geothermal heat sink, TRNSYS simulation, Experimental solar plant

1. Introduction

Solar thermal cooling facilities based on absorption cycles, when applied to the domestic sector are being extensively investigated these days because of their great potential on lowering the overload on electricity grids during summer season, but also for their energy and environmental advantages. Most commercially available water fired absorption chillers use the BrLi-H2O working solution. This solution is widely known and its performance has been well documented by many researchers [1, 2]. These types of chillers commonly need wet cooling towers for recooling, in order to dissipate the heat generated internally in the absorber and condenser, thus limiting their use in both the industrial and residential sector for its cost and bulk but also posing the health risk of legionella. Also the operating cost of a wet cooling tower gets incremented by the need of running the cooling fan, consuming a considerable amount of electric energy, and for replenish the water that gets evaporated and dragged out of the cooling tower. These drawbacks are very impeding for the domestic sector.

There are some commercially available absorption chillers recooled by an air stream, even being fired by hot water [3]. In hot days the ambient temperature result too high for them; as a result their COP and cooling capacity substantially diminishes. This effect is more pronounced when the machine is driven by the limited temperature of hot water produced in solar collectors.

An option for absorption cycles recooled by water is the use of a geothermal heat exchanger, taking advantage of the lower soil temperature. The installation of this kind of heat sink is more complex than installing a wet cooling tower because of the previous excavation of the borehole heat exchangers (BHE). The properties and type of soil underneath will dictate the dimensioning and hence the total cost of the facility, but once installed, its maintenance cost is low. Making use of this type of heat sink could promote the use of these facilities in the more sensitive domestic sector.

The Universidad Carlos III de Madrid (UC3M) counts with an experimental solar thermal cooling facility using a wet cooling tower coupled to an absorption chiller for heat rejection purposes. A numerical simulation using the TRNSYS tool has been accomplished and validated over the current experimental set-up.

This paper presents the modification of the already validated TRNSYS simulation, implementing a geothermal heat sink instead of the wet cooling tower. Different BHE connected in parallel and different borehole depths are analyzed in a trial an error way in order to select the best configuration for supplying the necessary heat rejection rate of the absorption chiller.

Cooling Load Profile

In this phase of the work the aim was to create the hourly cooling load profiles of the diary Best Milk (Marrakech, Morocco). The necessary information has been provided by the factory personnel and through audits on the plant Best Milk uses raw milk provided by a vast number of farms in the area of Marrakech. Cows are milked early in the morning in small dairy farms in villages around Marrakech, then milk is collected in several collecting centres and transported by truck tanks to the diary factory in the city. Generally milk tanks reach the factory by midday and the milk temperature ranges between (9-28 °C) depending on the availability of refrigeration systems at the collecting centre and the ambient temperature. According to the process engineers, the milk has to be cooled after the first milking to 7.3°C or less within 4 h of the start of the first milking. In the existing system milk is cooled via two plate heat exchangers, supplied on their cold side with a chilled water at about 2°C. Moreover, even though with the dedicated financial resource for the project which allowed only to size the pilot system far smaller than the existing conventional one, the system design was done in a way to increase replication potential and it has been decided to work on a portion of the industrial process (i. e., a portion of the milk volume flow) which allows to speculate on the possible behaviour of the a solar refrigeration system correctly sized for the application (i. e., about seven times in capacity). A scheme of the process is given in

image593

fig. 2. In order to investigate the opportunities of utilizing the solar cooling system in the fresh milk cooling process, both, the cooling load profile and the solar radiation for one typical summer day were studied; it was concluded that due to the mismatch between the two profiles an inertial component was needed. A cold storage has been selected to store cold energy in the phases when the availability of solar radiation allows cold production, without having concurrent cooling demand. Also, due to the existing system configuration — where there are two heat exchangers to cool milk, and they are used alternatively — it was decided to connect the solar cooled heat exchanger to play a role on cooling the return chilled water to the existing chiller as this is common between the two existing milk heat exchangers. The load profiles worked out — an example of a daily profile is presented in fig. 3 — highlighted, on a daily basis, that four operation phases take place as further described.

Operation phase:

1. Direct cooling: This is the default mode of operation of the system when there are simultaneous solar energy and cooling load.

2. Charging the cold storage: This basically happens early in the day when there is enough solar radiation to run the chiller but there is no cooling load as the milk didn’t arrived yet to the factory.

3. Discharging the cold storage: As soon as the supply energy by the solar radiation is not sufficient to run the chiller, the system turns to discharge mode, where cold energy started to be discharge to

4. image594System Off: Obviously, when there is no load and there is no solar radiation the system goes off, this mainly happens in the very late hours of the night and before sunrise.

4 System components and their models

4.1 Solar field

An essential requisite, to obtain good system efficiency, is the choice of the Solar Collector type and the working fluid, so that should be guaranteed the desired chiller operating conditions (feeding temperature, fluid flow rate etc.). As a result Roof mounted parabolic troughs RMT produced by IST and the diathermic oil were selected. In these collectors, solar energy focused and concentrated on a liquid-filled receiver dramatically reduces convection and conduction thermal

loses. The receiver/absorber is a steel tube coated with a selective blackened nickel surface and surrounded by glass. A single motor drives the collectors to track the sun continuously during the day.

Absorption chiller efficiency and operation mode free cooling

In [13] the COP is considered as key figure to characterise the energy performance of a refrigeration machine. For thermal driven refrigeration machines, the COP, which indicates the required heat input for the cold production, can be defined as follows:

Qlow Heat flux extracted at low temperature level

(eq. 2) COP = &— = .

Qhigh Driving heat flux to cooling equipment

Based on the measurement data of the years 2004 to 2006, the COP depending on the daily generated cold work by the absorption chiller is given in the diagram of figure 5.

Подпись: 0.9 daily cold work |к\Ъ d] Fig. 5. Coefficient of performance depending on the daily produced cold work.
COP [-1

The COP of the absorption chiller increases with increasing amount of daily generated cold work. The absorption chiller reaches its nominal COP, when it generates more than 200 kWh cold work per day. Below a daily amount of 200 kWh/d cold work, part load is governing the operating time with timing operation of the absorption chiller. The total cold work generated by the absorption chiller in the observed five year period is 31,365 kWhth.

Besides the operation of the absorption chiller, some cold work was generated by the operation mode free cooling. Free cooling does not only occur in spring and autumn, but also under German weather conditions in summer time. The diagrams in figure 6 established the amount of cold work generated by the absorption chiller and the operation mode free cooling in relation for the months May to September for the years 2005 and 2006.

image007

The operation mode of free cooling ranges from 1.2 up to 70 % in some months. In the years 2002 to 2007, 25% of the cold demand of the solar cooling plant was covered by free cooling, which means 10,499 kWhth. These results show the extensive potential of the operation mode free cooling. The total supplied cold work of the solar cooling plant is 41,864 kWhth.

Подпись: Fig. 7. Photos of the inner wall of the pipelines of the re-cooling cycle before and after treatment.

In 2003 leaks in the re-cooling cycle were found. The reason of the leakages was corrosion of the welding seam of the steel pipes, although a re-cooling water treatment was running. A refurbishment of the re-cooling loop was carried out in the year 2004. The pipelines of the re-cooling cycle were pickled by hydrochloric acid to remove corrosion products, afterwards the pipelines were rinsed and existing leakages repaired. Figure 7 shows two photos of the inner wall of the pipelines before and after the treatment.

To avoid further corrosion of the pipelines, the re-cooling water treatment was adjusted and a regularly water analysis and maintenance have been done to ensure the quality of the re-cooling water. Since the refurbishment, which took three weeks, the re-cooling cycle is operating well.

3. Conclusion

This study presents the operational experiences of the solar cooling plant of the Fraunhofer Institute in Oberhausen, Germany, which has operated for space cooling since 2002. The plant includes a single­effect lithium bromide-water absorption chiller with a cold capacity of 58 kWth, vacuum tube collector with an aperture area of 108 m2, a hot water storage with the capacity of 6.8 m3, a cold water storage with the capacity of 1.5 m3 and a 134 kWth cooling tower. The solar cooling plant is integrated into the supply infrastructure of the institute, so that solar surplus heat can be supplied to the heating system

and vice versa. Furthermore the operational concept of the cooling plant enables the operation mode free cooling. In the observed time period from 2002 to 2007, the total cold work provided by the solar cooling plant is 41,864 kWhth. Free cooling covers 25 % of the cold demand. The average solar collector efficiency is 0.28 and the COP of the absorption chiller varies from 0.37 to 0.81. In the five year period, 60 % of the driving heat of the absorption chiller is produced by the solar collector field.

Acknowledgments

The author Ahmed Hamza H. Ali wishes to acknowledge the Alexander von Humboldt Foundation, Germany, for the fellowship grant during this work. The authors acknowledge the funding of the German Ministry of Economy within the project “Study on Measurements of Optimization of Operation of Solar-thermal Driven Plants for Cold Generation (FKZ 0327406D). The project is furthermore integrated within the IEA Task 38.

References

[1] K4RES-H., (2006). Solar Assisted Cooling-State of the Art, Key Issues for Renewable Heat in Europe, Solar Assisted Cooling — WP3, Task 3.5, Contract EIE/04/204/S07.38607, 21p.

[2] Balaras, C. A., Grossman, G., Henning, H. M., Ferreira, C. A.I., Podesser, E., Wang, L. and Wiemken, E. (2007). Solar air conditioning in Europe—an overview. Renewable and Sustainable Energy Reviews, Vol.11, pp.299-314.

[3] Ward D. S. and. Lof G. O. G., (1975). Design and construction of a residential solar heating and cooling system. Solar Energy, Vol. 17, No.1, pp. 13-20.

[4] Ward D. S., Lof G. O. G. and Uesaki T., (1978). Cooling subsystem design in CSU solar house III Solar Energy, Vol. 20, No. 2, pp. 119-126.

[5] Ward, D. S., Weiss, T. A. and Lof G. O. G., (1976). Preliminary performance of CSU Solar House I heating and cooling system. Solar Energy, Vol. 18, No. 6, pp. 541-548.

[6] Bong, T. Y., Ng, K. C., and Tay, A. O., (1987). Performance study of a solar powered air conditioning system. Solar Energy, Vol. 39, No. 3, pp. 173-182.

[7] Al-Karaghouli, A., Abood, I., Al-Hamdani, N. I., (1991). The solar energy research center building thermal performance evaluation during the summer season. Energy Conversion and Management, Vol. 32, pp. 409-417.

[8] Yeung, M. R., Yuen, P. K., Dunn, A., and Cornish, L. S., (1992). Performance of a solar powered air conditioning system in Hong Kong. Solar Energy, Vol. 48, No. 5, pp. 309-319.

[9] Best, R., Ortega, N., (1999). Solar refrigeration and cooling, Renewable Energy, Vol. 16, pp. 685- 690.

[10] Syed, A., Izquierdo, M., Rodriguez, P., Maidment, G., Missenden, J., Lecuona, A. and Tozer, R., (2005).

A novel experimental investigation of a solar cooling system in Madrid, Int. J. Refrigeration, Vol. 28, pp. 859-871.

[11] Pongtornkulpanich, A., Thepa, S., Amornkitbamrung, M., and Butcher C. (2007). Experience with fully operational solar-driven 10-ton LiBr/H2O single-effect absorption cooling system in Thailand. Renew Energy, doi:10.1016/j. renene.2007.09.022.

[12] Ahmed Hamza H. Ali, Peter Noeres and Clemens Pollerberg. (2008). Performance assessment of an integrated free cooling and solar powered single-effect lithium bromide-water absorption chiller. Solar Energy. doi:10.1016/j. solener.2008.04.011.

[13] Henning, H. M. (Ed.) (2007). Solar-Assisted Air-Conditioning in Buildings, A Handbook for Planners. ISBN 978-3-211-73095-9. Second Ed., Springer Wien New York.

Summary of Test Results

Our first hypothesis was that the bubble pump would not function below 100°C. It can be seen from Figure 12 below that while the bubble pump was below 100°C the cool side temperature was rising, but when the bubble pump temperature exceeded 100°C, the cool side temperature began to fall.

image102

Secondly, we wanted to see how cool we could get the cool side of the refrigeration system using the solar collector of limited size. As can be seen from Figure 13, a temperature difference of 20°C between the atmosphere and the cool side was maintained with the bubble pump being at 120°C. It is worth mentioning that all tests were done between 10 am and 2 pm, for optimal solar heating.

Figure 13: Relationship between bubble pump and the cool side temperatures Economic considerations

Preliminary financial calculations showed that if the refrigeration system is to use electricity, it will cost about 17000 QR in the span of the next twenty years (not including the rise of gas prices). On the hand, the solar-operated refrigeration system will incur some initial cost, but this cost will decline to almost zero after all capital is paid for.

Conclusions

The intention of the present work was to study the viability of using solar thermal heat absorption refrigeration technology in Qatar. From the preceding it is established that cooling using solar thermal heat is indeed possible to accomplish in Qatar, but the initial installation costs may currently be prohibitive. It is believed that this could be a viable application, but it is need of further attention and development to reduce installation costs.

Recommendations

As a result of the experience gained during the present experimental work, it is recommended that research in this area continue in the state of Qatar. Specifically, it is recommended that the project be taken to a further level, and a new student be assigned to perform the following tasks.

1. A heat absorption refrigeration system be constructed that is specifically designed to accommodate larger solar panels.

2. Further investigation is done to determine possible cost advantages of building a hybrid compressor/thermal absorption refrigeration system for day/night use.

3. A large scale application be built and tested to determine sizing requirements for typical room sizes in Qatar.

References

[1] Shipman, B. C.; 1936, Air Cooling and Conditioning Apparatus and System, U. S. Patent #2 058 042.

[2] Erickson, Donald C (2007) Extending the Boundaries of Ammonia Absorption Chillers, ASHRAE Journal.

[3] Uli Jakob et al, (2007) Experimental Investigation of Bubble Pump Performance for a Solar Driven 2.5 KW Diffusion Absorption Cooling Machine, Heat Transfer in Components and Systems for Sustainable Energy Technologies, Chambery, France.

[4] David W. Bridger and Diana M. Allen, 2005, “Designing Aquifer Thermal Energy Storage Systems” ASHRAE Journal, Vol. 47, No. 9, September 2005.

[5] Klein, S. A., Reindl, T. D., 2005, “Solar Refrigeration” ASHRAE Journal, Vol. 47, No. 9, September 2005 Acknowledgment

This publication was made possible by a grant from the Qatar National Research Fund. Its contents are solely the responsibility of the authors and do not necessarily represent the official views of the Qatar National Research Fund.

Control strategy of the solar plant

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

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

image241

Fig. 2. Characteristic curve of the considered absorption chiller

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

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

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

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

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

Qcool = K2 ‘ (tIA — 26) (6)

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

Fifth step : economical analysis

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

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Fig. 3. Example of graph presenting economical results.