Category Archives: EuroSun2008-7

FIRST RESULTS OF A SOLAR-THERMAL LIQUID-. DESICCANT AIR CONDITIONING CONCEPT

B. M. Jones* and S. J. Harrison

Department of Mechanical and Materials Engineering, Queen’s University, K7L 3N6, Kingston, Canada

* Corresponding Author, jones@me. queensu. ca

Abstract

A thermal low-flow liquid-desiccant air handling machine was procured, installed, and field tested. The goal of the present investigation is to evaluate the field performance of the machine and characterize its operation for the temperature range of a solar thermal array. The system studied includes a natural gas boiler supplying the heat, and a cooling tower for heat rejection. System performance was evaluated for the 50 to 90°C temperature range, the operating range of solar thermal collectors. Cooling power varied between 4.3 kW and 22.8 kW for this range of temperature, with a latent heat ratio between 1.1 and 1.9, confirming that the unit is significantly dehumidifying the process air stream. Electrical COP varied between 0.58 and 4.48. Performance data indicates higher temperature solar collectors such as evacuated tube or double glazed flat plat collectors would be optimum in a solar cooling application with this system. These performance figures and methods will be used in future work to simulate and optimize a solar thermal driven dehumidification system for dedicated outdoor air systems.

Keywords: HVAC, Liquid Desiccant, Solar Thermal, Experiment

1. Introduction

An effective approach to the problem of latent cooling demand is to decouple the latent and sensible cooling loads. This is accomplished by employing a dedicated outdoor air system (DOAS). A DOAS is designed to precondition the outside ventilation air required for the building by removing the moisture. The remaining sensible load is handled by a conventional AC system in the return air stream, avoiding overcooling and reheat. The conventional electric chiller is put in parallel with the dry outside air stream [1]. This decoupling allows each sub-system to be efficiently designed for the type of cooling load (latent or sensible). Not only does this improve energy efficiency of the overall system, but the system is now able to properly handle a wider range of cooling loads. Figure 1 shows a DOAS configuration, with RA, OA, EA, and SA referring respectively to the Return, Outside, Exhaust, and Supply Air streams.

With increasing importance placed on climate change, energy conservation, energy security, and air quality, the low flow liquid desiccant dedicated outdoor air system driven by solar thermal energy is an attractive technology. It has significant potential advantages over many alternative building energy schemes. The important advantages of the solar thermal low flow liquid desiccant concept are summarized [2-5];

• Sustainable clean solar thermal energy source

• Solar combi-system arrangements possible, increasing solar thermal array usability

• No global warming potential associated with common refrigerants

• Low system pressure drop

• Zero carryover of corrosive desiccant material

• Air cleaning and filtering

• Lossless storage of solar cooling power in liquid salt solution

image657

2. Apparatus

Figure 2 is a photograph of the air handling apparatus and the cooling tower. Prominent features of the system are labelled. The primary components of the dehumidification system under investigation include the heat rejection system, the air handling system, and the thermal source system. The process flows for these systems are presented in Fig. 3. The air handling system is a self contained pre-commercial prototype, and includes blowers, pumps, and data acquisition and control equipment. Figure 3 indicates the positions of sensors within the overall apparatus. Sensors are placed such that an energy balance on the sub components can be performed, and to evaluate the performance of the system. An energy balance is calculated for the system components using flow meters and inlet/outlet temperature probes. Table 1 lists the system operational variables provided by the manufacturer and those used in experiment [6].

Manufacturer

Experiment

Air Flow

1500 L/s

1700 L/s

Hot water flow

80 L/min

80 L/min

Cold water flow

110 L/min

85 L/min

Desiccant charge

43% LiCl

43% LiCl

Desiccant flow

0.126 L/s

0.126 L/s

Regeneration thermal COP

0.60 0.83

0.70 — 1.00

Hot water temperature

70 — 100 C

50 — 70 C

Total Cooling

30 — 60 kW

4.2 — 22.8 kW

Latent Cooling Ratio

0.7 — 1.1

1.0 — 1.9

 

image658

Figure 2 — Site photo and apparatus layout

 

O* Blower

Collector

 

Gas Boi er

 

Regenerator

 

Absorber

 

Cooling Tower

 

Scavenging Air

 

image659

Ventilation Air

 

Desiccant Sump

 

Heat Exchanger

 

image660

image661

Figure 3 — Process flows and instrumentation

3. Experiment

The system was operated over 20 days in July in Kingston, Ontario, Canada. The range of water temperatures expected for a solar thermal application was the primary independent variable. Table 2 is a summary listing of the daily operation of the air handling system under 4 types of regenerator heating profile. The 10th, 13th, and 20th correspond to the 50, 70, and 90°C heating set point. Average conditions for daily operation are listed.

Table 2 — Experimental results

Day (July 2008)

Quantity

Symbol

Note

Unit

10

13

20

25

Start time

Time

hr

9.0

9.0

11.0

9.0

Time Period

At

Time

hr

7.0

6.0

5.0

7.0

Hot water inlet

Th.„,in

Avg.

°С

50.5

68.7

86.6

Solar

Hot water change

AThu„n

Avg

°С

-1.7

-3.4

-5.0

-3.5

Hot water flow

ІЩш

Avg.

kg/br

6295.9

6250.2

6793.5

6779.2

Cold water in

T„.„ i..

Avg.

°С

18.7

22.3

24.4

23.1

Cold water change

АТгш

Avg.

°С

l. S

3.1

4.4

3.2

Cold water flow

_

Avg

kg/br

5148.5

5188.6

5210.2

5004.2

Ambient temperature

Tair. in

Avg.

°С

21.0

21.2

22.9

23.8

Ambient change

ДT„ir

Avg.

°С

0.1

3.7

5.2

2.4

Ambient humidity

wamb

Avg.

g/kg

8.7

12.4

14.3

11.9

Ambient humidity change

AW

Avg.

s/k£

-1.0

-4.0

-6.5

-3.6

Air How rate

ihAir

Avg.

kg/liL-

7417.9

7511.5

7365.2

7415.3

Total cooling rate

Haul

Avg.

kW

4.9

13.4

20.0

13.5

Latent cooling rate

Q latent

Avg.

kW

5.1

21.5

30.4

18.4

Latent heat ratio

UIR

Avg

1.0

1.6

1.5

1.4

Heating rate

Qreg?..

Avg

kW

12.9

24.8

40.3

28.0

Water desorbed

г

Tot,

kg

91.3

214.5

214.1

190.2

Desorption rate

a..-..

Avg.

kW

8.9

25.0

30.4

19.1

Water absorbed

Tot.

kfi

50.8

181.4

208.9

0.0

Absorption rate

Q——-

Avg.

kW

5.1

21.5

30.4

18.4

Thermal COP

COrthrr„,a,

Avg.

0.70

1.01

0.75

0.68

Cooling COP

corcooltng

Avg.

0.38

0.54

0.50

0.48

Electrical COP

COP’Uc

Avg.

0.96

2.00

3.86

2.68

Regen. rone, in

Сіпл

Avg

wt. %

26.88

32.18

37.30

32.78

Regen. cuuc. change

ACc

Avg.

Wt. %

-1.00

-1.85

-2.97

-2.21

Coud. cone. In

Ci n. r

Avg.

Wt. %

26.57

31.16

35.20

31.80

Cond. cone, change

Д tv

Avg.

wt. %

0.72

2.36

3.22

1.88

System operation plots for the 25th of July are presented in Figs. 4 and 5. The 25th was operated on a time varying simulated solar heating profile, simulating a 70m2 evacuated tube collector array. Figure 4 is a plot of the conditioner performance for the 25th of July. The system begins operation when the temperature of the solar thermal array reaches 50°C, at 9:30 in the morning. The temperature profile can be seen in the top section of Fig. 4. The temperature profile approaches 80°C as the day progresses and the incident radiation on the simulated vacuum tube collector increases. The temperature profile is adjusted by the programmable logic controller every 30 minutes. From 2, the average cooling rate is 13.5 kW with a latent heat ratio of 1.4. The middle section of Fig. 4 displays the temperature at the process air inlet and outlet. The temperature rises an average of 2.4°C for this day. The bottom section of Fig. 4 shows a significant dehumidification effect, a change in absolute humidity averaging 3.6 g/kg. This amounts to a total absorption of 208.9 kg of water for this day. From the middle section of Fig. 5, it is observed that the absorption and desorption rates are similar. This indicates that the machine is approaching steady state operation even over a varying temperature profile The total amount of water desorbed for this day is 214.1 kg. The bottom section plots the coefficients of performance for this day. The average thermal COP for the 25th is 0.68, the average cooling COP is 0.48, and the average electrical COP is 2.68 for this solar temperature profile.

image662

Figure 4 — Conditioner performance under simulated solar load

image663

Figure 5 — Regenerator solar heating profile and system COP’s

Figures 6 and 7 display the average daily coefficients of performance for the data set. The thermal regeneration COP is the energy of desorbed water divided by the heat absorbed. The cooling COP is the energy change of the air stream divided by the heat absorbed. The electrical COP is defined as the energy change of the air stream divided by the electricity consumption rate, which was 5.5 kW for this system.

Overall, the regenerator COP averaged 0.81 for the days of constant heating temperature operation, while the cooling COP is an average of 0.53. The cooling COP is lower due to the enthalpy gained by the desiccant as it condenses and absorbs water vapor from the process air. The cooling COP can be improved with a higher capacity heat rejection system. The electrical COP can be improved with variable speed drives on pumps and blowers.

image664

Figure 7 — Daily electrical COP

4. Conclusions

A liquid desiccant air handling system driven by thermal energy was procured. Piping, valves, motors, and thermal sub-components were connected. Transducers included flow meters, thermistors, and relative humidity sensors. Desiccant solution concentration was monitored using a batch process density meter and a density-concentration correlation. A data acquisition system was designed and installed. The system is controlled using a programmable logic controller using relay ladder logic controlling the drive units for pumps and fans, and monitoring system faults. The system commissioning included troubleshooting and government safety inspections.

The system was tested in a field environment. The data set was processed giving enthalpy flows and performance figures for daily operation of the machine. An enthalpy balance on the system was used to check sensors and assumption of adiabatic operation. The assumption of an
adiabatic control volume around the conditioner and regenerator was found to be valid. Regenerator thermal COP was calculated to be 0.76 to 0.99, and conditioner thermal COP was in the range of 0.38 to 0.61 for the entire data set over 50°C to 90°C. Electrical COP increased from 0.58 at 50°C to 4.48 at 90°C. The average latent heat ratio was 1.1, 1.3, and 1.5 respectively for the three temperatures evaluated, showing that the heat load on the conditioner increases as the desiccant is heated in the regenerator. The average cooling power was 4.3, 14.1, and 21.3 kW for the temperature range. Data indicates that high temperature solar thermal collectors would achieve better performance based on their ability to maintain a desired cooling power and their increased electrical COP.

Future work in this project will include coupling a solar thermal array incorporating liquid desiccant storage, and upgrading drive units to improve the system electrical COP.

References

[1] Mumma, S. A. (2007). DOAS and desiccants Engineered Systems, 24, 37 — 49.

[2] Mei, L. & Dai, Y. (2008) A technical review on use of liquid-desiccant dehumidification for air­conditioning application Renewable and Sustainable Energy Reviews, 12, 662 — 89

[3] Hwang, Y.; Radermacher, R.; Al Alili, A. & Kubo, I. (2008) Review of solar cooling technologies HVAC and R Research, 14, 507 — 528

[4] Katejanekarn, T. & Kumar, S. (2008) Performance of a solar-regenerated liquid desiccant ventilation pre­conditioning system Energy and Buildings, 40, 1252 — 1267

[5] Jain, S. & Bansal, P. (2007) Performance analysis of liquid desiccant dehumidification systems International Journal of Refrigeration, 30, 861 — 72

[6] Lowenstein, A.; Slayzak, S. & Kozubal, E. A zero carryover liquid-desiccant air conditioner for solar applications International Solar Energy Conference, 397 — 407

Cooling power and COP

In Fig. 3 and Fig. 4 a histogram of the measured COP and cooling power is shown. COPs between 0.48 to 0.62 were obtained. The measured power shows a much larger dispersion: cooling powers between 3.1 and 5.9kW have been obtained.

Подпись: 20%Подпись: 0%Подпись: COPimage629Подпись:image631120%

100%

80% « 0) □

60% I"

■o

0)

40%

з

Є

20%

0%

Fig. 3. Frequency diagram of thermal cooling COPs Fig. 4. Frequency diagram of cooling powers. Taking into account the total cold production and heat input over the whole cooling period including hours with on-off behaviour an overall COP of 0.574 with a standard deviation of 0.066 for the hourly values is obtained. The mean cooling power for the hourly values was 4.38kW with a standard deviation of 0.67kW. Nevertheless it is not clear, if this variation of cooling powers is due to a limitation in the capabilities of the chiller or to the requirements of the load. Although the cooling load of the canteen’s kitchen is almost always above 6kW because of the high internal loads, the control system may limit the chillers output power in order to avoid too low inlet air temperatures which may be uncomfortable.

A Sensitivity Analysis Of A Desiccant Wheel

P. Bourdoukan1[16], E. Wurtz2, P. Joubert1 and M. Sperandio1

1 LEPTIAB, Universite de La Rochelle La Rochelle, Avenue Marillac, 17000 La Rochelle, France
2 Universite de Savoie, Campus Scientifique, 73376 Le Bourget du Lac, France
* Corresponding Author, paul. bourdoukan@univ-lr. fr

Abstract

Desiccant cooling powered by solar energy and using water as a refrigerant has a low environmental impact and appears as an important technique to reduce energy consumption in buildings. The cooling potential of the system is based on the performance of the desiccant wheel that removes humidity from outside air to increase the potential of the humidifier. In this paper a sensitivity analysis of the desiccant wheel dehumidification is performed using the design of experiments. The impact of outside temperature, outside humidity ratio, the regeneration temperature and the regeneration humidity ratio is studied on the dehumidification rate of the wheel using experimental and numerical results.

Keywords: solar desiccant cooling, sensitivity, experiments, simulation

1. Introduction

Solar desiccant cooling is a heat driven technique powered by solar collectors. It is based on evaporative cooling and utilizes a desiccant wheel to remove humidity from outside air. When adsorbing the humidity the desiccant needs to be regenerated by moderately hot air stream provided by solar collectors. This technology presents the advantages of being friendly environmental since its electrical consumption is limited to the auxiliaries (fans and pumps), beside it use water as a refrigerant in opposition to vapor compression technique using refrigerants with high environmental impact. The general scheme of a solar desiccant cooling plant is shown in the figure 1 below:

image665

Fig. 1. Desiccant cooling installation and evolution of air properties in the psychometric chart

With reference to Fig. 1 the conventional cycle operates as follows: first, outside air (1) is dehumidified in a desiccant wheel (2); it is then cooled in the sensible regenerator (3) by the return cooled air before undergoing another cooling stage by an evaporative process (4), finally, it is
introduced into the building. The operating sequence for the return air (5) is as follows: it is first cooled to its saturation temperature by evaporative cooling (6) and then it cools the fresh air in the rotary heat exchanger (7). It is then heated in the regeneration heat exchanger (8) and finally regenerates the desiccant wheel (9) by removing the humidity before exiting the installation.

The task of the desiccant wheel is first to reduce the humidity of outside air in order to match indoor air standards and second to provide an extra dehumidification to increase the potential of supply humidifier. The desiccant wheel appears as a key component. The dehumidification performance of the desiccant wheel depends on the operating conditions [1] e. g. the wheel rotation speed, the air flow rate, the outside temperature and outside humidity, the regeneration temperature and regeneration humidity. Usually an optimum rotation speed and optimum air flow rate are recommended by the manufacturer thus these two parameters are constant during the operations but outside and regeneration conditions are not. So the performance of the desiccant wheel depends intrinsically on these parameters.

In this paper a sensitivity analysis is conducted on a desiccant wheel using the silica gel as an adsorbent to investigate the effect of outside and regeneration conditions on the performance of the desiccant wheel. The method used in this analysis is design of experiments (DOE) [2]

2. Sensitivity analysis

2.1. Design of experiments

In the DOE the response (y) of the studied phenomena influenced by different parameters or factors (xi) is expressed using a polynomial form [2]:

image666

N N N •

the lower limit is the minimum required for the silica gel. For the regeneration humidity ratio range it is taken with the consideration of the outlet conditions of the return humidifier.

2.2. Experimental setup

The experimental installation of La Rochelle [1] is used for the measurements of the performance of the desiccant wheel. This experimental installation consists of a silica gel desiccant wheel and aluminum sensible regenerator and two rotating humidifiers. At the inlet and outlet of each component a psychrometer is used to measure the dry and wet bulb temperature. The local atmospheric pressure is measured too thus using the dry and wet blub temperature the humidity ratio is then measured accurately. At the desiccant wheel outlet, the dry bulb temperature and the humidity ratio are not uniform. In order to have an accurate measurement of the mean outlet temperature and humidity, three humidity measurements and 6 temperature measurements are performed simultaneously.

The major parts of the dehumidification rates (wi-w2) used in the sensitivity analyses are experimental measurements but when a combination of parameters is not possible experimentally numerical results of the desiccant wheel model are used to complete the required combinations.

The model used will be introduced in the following section.

Driving, heat rejection and chilling temperatures

The control system start the chiller only if a driving temperature above 72°C is available, and turns the machine off when it falls below 68°C (Fig. 5). The measured heat rejection temperatures were very favourable for the chiller: in 95% of the hours an inlet temperature to the machine between 18°C and 20°C was registered. These numbers show the effectiveness of the installed boreholes for the heat rejection in the present system. Chilled water temperatures (inlet temperature to the cooling coil) between 7°C and 14°C were measured. As a conclusion, the temperature lift defined as temperature difference between chilled water outlet and heat rejection temperature inlet to the machine was about 7.9K (Fig. 6).

Desiccant wheel model. Model description

The heat and mass transfer model for the desiccant wheel used below is based on the analogy method with heat transfer that occurs in the sensible heat regenerator. It was first introduced by Banks [3] and Maclaine cross [4] then Jurinak [5] and Stabat [6] improved the model. The following assumptions are considered:

• The state properties of the air streams are spatially uniform at the desiccant wheel inlet

• The interstices of the porous medium are straight and parallel

• No leakage or carry-over of streams

• The interstitial air velocity and pressure are constant

• Heat and mass transfer between air and porous desiccant matrix is considered using lumped transfer coefficients

• Diffusion and dispersion in the fluid flow direction are neglected

• No radial variation of the fluid or matrix states

• The sorption isotherm does not represent a hysteresis

• Air reaches equilibrium with the porous medium

image667

The heat and mass conservation equations:

image668 image669 image670

Heat and mass transfer equations:

image671

Equations (2), (3), (4) and (5) are coupled hyperbolic non-linear. With the assumption of the Lewis number (Le), equal unity and the desiccant matrix in equilibrium with air means (Td= Teq and weq= wa). Banks [2] used matrix algebra and proved that these equations can be reduced using potential function Fi(T, w) to the following system:

(7)

9(C*)

(10)

Where

C * = Mm. c pm N

Подпись:Подпись: = 0

image674

Подпись: cПодпись: pm

image677

Introducing the equations of heat transfer alone in the sensible regenerator as stated in [3]:

(12)

ncf is the efficiency of the counter flow heat exchanger for balanced flow.

F = h

(13)

Подпись: F = (273.15 + T) F 2 =

image679

6360 (14)

The model described above, was implemented to SPARK [8] a general simulation environment based on equation. In the next section the model is validated experimentally.

ROOF INTEGRATED 10kW PV STATION AND ABSORPTION. CHILLER FOR MEDICAL CENTER IN YEREVAN

Victor Afyan1 and Arsen Karapetyan2

1 SolarEn, LLC, 2/2 Shrjanayin St., Yerevan 0068, Armenia.

2 SolarEn, LLC, 2/2 Shrjanayin St., Yerevan 0068, Armenia.

* Corresponding Author, victor afyan@solaren. com

Abstract

10 kW PV power station is integrated into the roof of the five-store medical center building and connected to the 3-phase grid on the net metering basis. For cooling needs of the building there is a hot water driven lithium bromide absorption chiller with 420kW cooling capacity. The chiller is powered by a gas boiler and supplies chilled water to the fan-coils system of the building. PV station is installed and tested. Absorption chiller is installed and commissioned. Technical data for PV and cooling systems are presented.

Keywords: BIPV, net metering, cooling, absorption chiller.

1. Introduction

Building integrated photovoltaic (BIPV) and grid-tied solar power stations are well known and implemented in developed countries due to the favorable feed-in tariff and rebates. Although PV technology has a long history in CIS countries, particularly for autonomous power supply applications, its BIPV implementation is hampered by absence of grid connection legal mechanisms and financial stimuli.

PV power plant as well as other renewable energy (RE) power plants connection and parallel operation with the grid has been ensured and regulated by the net metering mechanism adopted in Armenia in 2005. According to the regulation all RE power plants and even combined heat and power (CHP) units with capacity up to 150kW can benefit parallel operation with the grid based on the net metering.

The PV power station and tri-generation energy efficient system has been designed for the renovated building of the Armenian-American Wellness Center in Yerevan. The design is based on 10kW roof-integrated grid-connected PV power station, hot water driven lithium bromide absorption chiller with 420kW cooling capacity and cooling tower, gas genset with 110kWe and 150 kWth capacity and a back-up gas boiler. The CHP issue is still pending but other components of the system are already installed and described below

A novel material for desiccant wheels: Performance testing results

P. Kohlenbach*, D. Rossington and A. Weigand

1 CSIRO Energy Technology, PO Box 330, Newcastle, NSW, 2300, Australia
Corresponding Author, paul. kohlenbach@csiro. au

Abstract

CSIRO Energy Technology is developing a small-scale desiccant-based air-conditioning system for residential applications. In this context, a desiccant wheel made of a novel material has been experimentally tested for its dehumidification performance. The material is an iron-alumino-phosphate zeolite with an AFI structure and traded under the name of FAM Z-01. A 300mm diameter desiccant wheel was tested under varying inlet conditions of temperature and humidity with regard to its dehumidification performance. It was found that for constant regeneration humidity the maximum moisture removal capacity of the material is 17 grams of water per kg dry air at 50°C regeneration temperature and 24 grams of water per kg dry air at 80°C regeneration temperature from an inlet air stream of 40 °C and 95% relative humidity. At supply inlet temperature between 10 and 30°C and supply inlet relative humidity between 20 and 50% it was found that the difference in moisture removal at a regeneration temperature of 50 °C and at 80 °C is around 1 g/kg d. a.. At varying regeneration humidity (matching ambient conditions) it was found that the moisture removal is considerably lower, even though the regeneration air is supplied at the same temperature. Maximum moisture removal was 5.1 g/kg d. a. and 14.5 g/kg d. a. for supply inlet conditions of 40°C/95% RH at 50 degC and 80 degC regeneration temperature, respectively.

Keywords: Dehumidification, desiccant wheel, zeolite, FAM Z-01

1. Introduction

CSIRO is currently developing a solar-powered air-conditioning unit for residential houses, using a desiccant-evaporative process to provide cool and dehumidified air. This process is very well suited for the recovery of low-grade solar energy or waste heat. Thermal energy and water are used to provide air-conditioning, hence consuming only a very small amount of electrical power. As part of the development CSIRO is testing novel desiccant wheel materials for dehumidification purposes. The two most common materials for desiccant wheels are silica gel and LiCl due to their low cost and ease of handling. They are however limited in their moisture removal capacity, especially at regeneration temperatures below 80 degC. Recently researchers and manufacturers have been developing advanced materials to increase the moisture removal capacity and hence to allow for smaller unit size.

Jia et. al. [1, 2] describe a comparison between a novel composite desiccant wheel made of silica gel and lithium chloride and a conventional wheel made of silica gel only. They found that the composite wheel has a greater moisture removal capacity compared to the silica gel wheel, especially at lower inlet humidity. The regeneration temperature of the composite wheel was found to be lower than that of the pure silica gel wheel. Tokarev et. al. [3] have analysed a composite sorbent based on CaCl2 as an
impregnated salt and MCM-41 as a host matrix. At regeneration temperatures between 70 and 120 degC the moisture removal capacity of the composite was greater than of silica gel. Cui et. al. [4] investigated the properties of DH5, DH7 and 13x adsorbents with regard to their use in desiccant cooling systems. Their results show that DH5 and DH7 adsorbents have greater moisture removal capacity than silica gel when tested at a regeneration temperature of 100 degC. Restuccia et. al [5] also investigated a composite sorbent SWS-1L, a mesoporous silica gel KSK impregnated with CaCl2. It showed a promising moisture removal capacity of up to 0.7 g of water per 1 g of dry sorbent at regeneration temperatures of 90-100°C. Kakiuchi [6] presented the FAM-Z01 material used in this work as an application for adsorption heat pumps and proposed the application for desiccant wheels. This application was further investigated by Oshima et. al [7] who evaluated the performance of a desiccant rotor containing FAM-Z02 zeolite material. Various regeneration temperatures and air inlet conditions have been investigated in a parameter study. The moisture removal of the FAM-Z02 rotor was found to be 11-22% higher than that of a silica gel rotor at regeneration temperatures of 50-70°C.

One important aspect of using adsorbents in a solar cooling system is the long-term stability of the desiccant. Earlier investigations by Belding et. al [8] have shown that silica gel and 13x adsorbents can lose up to 63% and 13%, respectively, of their original moisture removal capacity after 50,000 cycles. The FAM Z-01 material used in this work has been tested by the manufacturer and has shown a 5% capacity loss after 50,000 cycles [6].

2. System and methodology

The experimental testing has been undertaken at CSIRO’s Energy Technology site in Newcastle, Australia. Figures 1 and 2 show the test rig used for experimental purposes.

image148

Figure 1. Schematic diagram of experimental test setup. (1) Intake Filter, (2) Fan, (3) Medium Temperature Coil, (4) Low Temperature Coil, (5) Primary Heater Bank, (6) Steam Injection Humidifier, (7) Secondary Heater Bank. RH=relative humidity, T=temperature, DP=differential pressure, V=volumetric air flow.

image149

Figure 2. Photo of the experimental test rig at CSIRO (desiccant wheel not shown).

As shown in Figure 1 and 2, the test rig consists of two separate conditioned air streams, one for desiccant wheel supply, and the other for desiccant wheel regeneration. Each of these two air streams enter via an intake filter (item 1 of Figure 1), and is then pressurised by a centrifugal fan (item 2) which is controlled by a variable speed drive to enable air volume control. Peak flow of 1000m3/hr is achievable with a 300Pa pressure drop across the desiccant wheel. The air stream then passes through two cooling coils. The first coil (item 3) is cooled with a 2°C chilled glycol solution capable of dehumidification of the leaving air stream to a moisture ratio of approximately 7 g/kg dry air. The second cooling coil (item 4) is cooled with a -5°C chilled glycol solution capable of further dehumidification of the leaving air stream to a moisture ratio of approximately 4 g/kg dry air. The dehumidified air stream then passes through a primary electric heater bank (item 5). This heater bank is capable of heating the air stream to 90°C in the case of the regeneration air stream, and 40°C in the case of the supply air stream. The air stream is then humidified as required using a steam injection lance (item 6). Low pressure steam at 1.5 bar is injected in the air stream via nozzles at a rate of up to 45 g/kg dry air. Finally the air stream passes though a secondary electric heater bank (item 7). The secondary heater bank is of similar capacity as the primary heater bank allowing for load sharing and fine temperature control. The supply and regeneration air streams are then ducted to the test desiccant wheel. This can be done in counter flow and parallel flow arrangement. The supply and regeneration air streams leaving the desiccant wheel are ducted from the wheel and exhausted outside. Temperature and Humidity are measured and recorded after each of the control elements described above. The temperature and humidity entering and exiting the desiccant wheel is calculated by averaging a number of sensors distributed over the cross section of the ductwork entering and leaving the desiccant wheel. Volume flow rate of the supply and regeneration air streams is calculated from velocity measured at the entering side of the desiccant wheel.

Basic configurations

The basic control schemes have been applied on the following three basic configurations for the installation:

2.2. Direct use configuration

The basic scheme used to execute the simulations in this case is the one on the following figure.

image252

Fig.2. Configuration for direct use

This corresponds to the habitual configuration for installations with absorption chiller. In this case to start the chiller working they must be fulfilled two requirements: to have solar energy enough stored on the tank and to have cold demand.

Expected results — 1st level

The main results within the first level are the

Подпись: • PER systemПодпись: • PERref same… primary energy ratio of the installed solar assisted heating and cooling ((Equation 1)

. primary energy ratio of an assumed reference system working under the conditions as the installed SHC system ((Equation 2).

Herewith the annual savings of primary energy through the utilisation of the SHC system can be expressed.

Further through the calculated cost per installed cooling capacity ((Equation 4), it will be possible in the following years to draw a learning curve for SCH systems.

3.1. Expected results — 2nd level

The main result within the second level is the calculation of the solar thermal energy which could not be exploited by the SHC system — this shows the “efficiency” of the SHC system ((Equation 12 and (Equation 13, see Fig. 5)

Pre-commercial development of a cost-effective solar-driven absorption chiller

Pedro Adao1*, Manuel Collares Pereira1, Andrea Costa2

1 Ao Sol SA (SunCool SA), Lugar da Sesmaria Limpa, 2135-402 Samora Correia, Portugal
2 ACE, Pfalzerstr. 75, 83109 Grosskarolinenfeld, Germany
* Corresponding Author, pedro. adao@aosol. pt

Abstract

This work presents the development results of a novel absorption chiller for solar cooling applications built at Ao Sol (SunCool laboratories) in Portugal. The chiller is at pre­commercial stage at the time of writing. Focus of the ongoing development work is the optimisation of technical and economic parameters such as overall dimensions. A thermo­economic analysis will give insight on the commercial potential and challenges.

Under present conditions, dictated by a new European legislation and current fuel price, the use of small absorption chillers coupled to solar thermal technology for residential applications turns out to be potentially both environmentally sound and economically viable. Necessary requirements are cost-efficient solar collectors and chillers, and an economically optimized combined system.

Relying on previous work performed at the University of Lisbon, a compact absorption chiller working with a solution of ammonia and water has been built and tested. The device is mainly made of plate heat exchangers and it is directly cooled by air, thus avoiding the need of a wet cooling tower. The control strategy is aimed to an overall concept that includes collector field, storage, gas back-up, chiller and building.

The thermodynamic design is presented along with experimental results of the thermal performance — in terms of chilled water production and coefficient of performance. The economic impact of the developed chiller for a common residential application in Portugal is compared with a reference system and discussed.

Keywords: solar cooling, air-cooled chiller, ammonia-water absorption, net present value

1. Introduction

The present European Directive on Energy establishes ambitious goals for the year 2020: 20% contribution of renewable energies to the final energy demand and 20% reduction of the final energy consumption through energy efficiency measures [1]. This creates a very strong drive for the solar penetration in the Heating and Cooling of Buildings, responsible for about 40% of the total final energy consumption in the EU [1].

Combining this directive with the ESTIF solar thermal installed capacity goals for 2020 [2 ESTIF Newsletter July 2007] an upper limit of about 4 million new solar driven machines can be anticipated for the residential and services market.

If the more ambitious ESTIF expectations are taken into consideration this figure will be multiplied by a factor of 5 to 10 to 2030.

The present situation in Europe shows a huge potential for renewable energies. In the building sector, with its highest emissions share, the potential for energy savings is over proportional high. Building cooling is rising to become a crucial issue in the next years. Solar energy is indisputably assessed as a major contributor to these savings and it is being fully backed by policy makers through long-term directives at EU-level.

Trends in Southern Europe: Southern European countries are in a particularly dire situation with a growth of energy consumption in the buildings sector precisely because of the widespread use of electrically-driven air-conditioning units; a situation which will even worsen with the warming trends associated with global warming. In the Portuguese residential and services sector the average increase rate of electricity consumption over the last three years is more than 4 times the average rate of the European Union [3]. This is mainly due to the rise in air-conditioning demand traditionally achieved with highly inefficient small-size chillers (window units). This generates a tendency, which is against the goal of 20% energy use reduction and requires even stronger action than in Northern European countries.

In the past, INETI and some of the people involved in the present development participated in a European funded project [4] to analyze the technical and economic potential of solar absorption cooling in the Mediterranean region. Essentially the conclusion was that — through savings induced by solar energy — an all-year system providing heating, cooling and domestic hot water would compare favourably with a conventional system with a boiler for heating and an electric chiller for cooling with a payback time of less than 10 years. The conclusions and ideas developed in that study were truly influential on the decision of AO SOL to take up the present R&D effort.