Category Archives: EuroSun2008-7

Operation of the adsorption chiller

Performance of the chiller

In order to evaluate the performance of the chiller a comparison of experimentally measured cooling capacities and COPs with expected values from calculations is shown in Fig. 2. The experimental values were selected according to the following criteria:

1. a continuous operation during the whole hour as well as in the hour before and after,

2. stable temperatures with a small standard deviation has been measures in all three circuits,

3. only cooling operation was considered.

TCooling operation

For the evaluation of the cooling operation only hours with constant and steady operation were considered. Operation periods less than 60 minutes within one hour were not considered in this evaluation in order to avoid effects of transient states. From Table 1 it can be seen that 282 hours could be considered.

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).

Main characteristics of Demo SP6b

Subproject SP6b is similar to SP6a. It is designed to condition a lecture room (17 people, 95 m2) and an office room (10 people, 80 m2) in the new office building, Solar XXI, at INETI and to warm up a DHW tank (0.4 m3) for lab washing purposes.

The main differences from SP6a are:

• The fuel used for the CHP — biodiesel (consequent lower thermal and electrical CHP output)

• The use of a chilled water tank in the cooling circuit

• The load demand (specific heating and cooling as well as DHW — consequent different number of running period, number of on/off cycles)

• Some details in the hydraulic circuit like, for example, a heat exchanger physically separating the chilled water loop from the main CHCP system circuit.

Energy equations

image039 image040
image041

The energy equations related to the adsorber, which will be given next, correspond to a multi-tubular system, whose inner surface exchanges heat with the water coming from the hot storage tank or from the water supply network, depending on the stage of the cycle. The adsorbent occupies the space delimited by the external wall of the tube and the corrugated fins.

Fig. 2 — Fin-tube heat exchanger and the simplification by annular fins

For the heat transfer in the adsorbent medium, the following model assumptions have been considered: (a) the pressure is uniform; (b) the heat conduction is two-dimensional (axial and radial) as detailed in Fig. 2; (c) the adsorbent-adsorbate pair is treated as a continuous medium in relation to thermal conduction; (d) the convective effects and pressure drops are negligible; (e) the condenser and evaporator are ideal, i. e. they have a constant temperature during the isobaric phases; (f) all the adsorbent particles have the same properties (including shape and size); they are uniformly distributed throughout the adsorbent, and in local thermal equilibrium with the adsorbate and the surrounding gaseous phase; (g) the gaseous phase behaves as an ideal gas; (h) the properties of the metal and the gaseous phase are assumed to be constant; (i) the properties of the heat transfer fluid, as well as those of the adsorbate, are considered as temperature dependent. It results the following equation

[Pi (CPі + aCP2 )] ] = к V2T + qst p ^ (2)

d t d t

Подпись: da dt Подпись: b image044 Подпись: qst dT RT2 dt Подпись: ,with image047 Подпись: (3)

where Cp is the specific heat (indices 1 and 2 refer to the adsorbent and the adsorbate, respectively), p the specific mass and к, the conductivity of the adsorbent. The total derivation of the concentration, a, is given by

image049 Подпись: dT . (d 2T 1 dT d2T ' —к — + + — dt ^d r2 r dr d z2 у Подпись: (4)

The da/dt term depends on the process that occurs in the adsorber. In the case of an isosteric process it is zero and for adsorption or desorption process, the term d lnp/dt is zero. Then, the energy equation for the adsorbent can be written as

image052 image053 Подпись: (5)

where u is a function of the process, 0 for isosteric and 1 for adsorption or desorption process. The condition in the middle of the adsorbent material, between two fins or two tubes comprises the adiabatic boundary condition. Other boundary conditions are in the interface between the adsorbent and the wall of the tubes and the fins. To solve the Eq. (4), the temperature on the wall of the tubes and on the fins are considered known, recalculated for each simulation step by the energy equation

Подпись: PtCPt Подпись: dt image057 Подпись: +—(T - T) A v f! Подпись: (6)

to the tubes and the following to the fins

image060 image061 image062 Подпись: mwCpw dTw A d x Подпись: (7)

where P is the perimeter, A is the area, Tt is the tube temperature, Tw is the water temperature, h is the conductance at the interface tube/adsorbent, and hfi is heat coefficients between the fluid and the tubes. The hfi is evaluated as the method described in [3]. The boundary conditions of Eq. (5) are adiabatic in both extremities and, to Eq. (6), they are adiabatic in the middle of the fin and known in the interface tube/fin. The temperature of the water is given by

were mw is the mass flow of water. To solve the system of equations formed by Eq. (1), (4), (5), (6) and (7) a mixed finite-difference and finite-volume method was used and the input data is chosen to be the temperature and the mass flow of the hot water, the number of tubes, the number and the thickness of the fins and the material of the fin. Additionally, the porous medium properties must be known, especially к and h. According to [4], for the AC-35 activated carbon: к = 0.19 W/mK and h = 16.5 W/m2K.

Evaluation of Solar Desiccant Cooling System for Field Test Office Building

The annual energy consumption for space heating and cooling in a office building of Japan is 359GJ/m2 [4]. About 40% of the annual energy consumption for space heating and cooling will be expected to be reduced using this solar thermal system for the similar type of the office building. Therefore, the reduced annual energy consumption for space heating and cooling in the field test office building is assumed by 57.6GJ/year. The reduced crude oil and the CO2 emission are estimated by 1.5kL and 4ton using the rates of 38.7GJ/kL and 2,649kg-CO2/kL, respectively.

The initial cost of this solar thermal system was 20 million JPY included the construction fee. If this solar system is used for 20 years, the CO2 reduction cost is 250 JPY/kg-CO2.

3. Conclusion

The solar desiccant cooling system was developed and the system performance was described in this paper. This system was developed as the passive solar thermal system using the renewable energy without the heat source equipment and the dehumidification cooling system for the fresh air.

From the field test results, it was found that the solar desiccant cooling system for office building

was effective throughout a year.

Acknowledgement

This study was supported by the research funds of NEDO project, Research and Development of

Technologies for New Solar Energy Utilization Systems for FY2005-F2007, and Grant-in-Aid for

Scientific Research (C)(19560598). The authors would like to express their sincere thanks to the

support.

References

[1] H. Roh, K. Suzuki, Research and Development of Air-based Passive Solar Dehumidification Cooling System, Part 1 Operation Test of Solar Dehumidification Cooling System, Proceeding of JSES/JWEA Joint Conference 2006 (Renewable Energy 2006 Japan Day), pp.309-312. (in Japanese)

[2] H. Roh, K. Suzuki, Research and Development of Air-based Passive Solar Dehumidification Cooling System, Part 2 Field Test of Solar Dehumidification Cooling System in Summer, Proceeding of JSES/JWEA Joint Conference 2007, pp.405-408. (in Japanese)

[3] S. Song, K. Suzuki, H. Roh, Study on the Performance of Dehumidification Cooling System with Solar Thermal and Well Water, Summaries of Technical Papers of Annual Meeting Architectural Institute of Japan 2008, D-2, pp.1209-1210. (in Japanese)

[4] Heat Pump & Thermal Storage Technology Center of Japan, White Paper of Heat Pump and Thermal Storage, pp.335, Ohmsha, 2007. (in Japanese)

Principal of Operation

Three external circuits are connected to the Millennium MSS Air Conditioning:

Thermal heat source (e. g. MSS solar collectors)

Air conditioning distribution system for cooling and heating (e. g. radiant floor, fan-coil units)

Heat sink for charging and discharging (e. g. swimming pool, cooling tower, air cooled condenser or geothermal holes)

Millennium Mss Air Conditioning System is a modular absorption machine that differs from the “standard” Lithium Bromide type absorption machines in three main aspects:

It has internal storage in each of the two accumulators. This allows the machine to store chemical energy with a very high density. This energy can subsequently be used both for cooling and heating. It is important to emphasize that this is chemical energy, not thermal energy that is stored.

It works intermittently with two parallel accumulators (Barrel A and Barrel B).

It is designed to use relatively low temperatures and is hence optimized for usage with solar thermal collectors. It also works with a stable temperature inside the accumulators, which in turn allows for an effective use of solar thermal collectors.

Millennium MSS Air conditioning system made up of two “barrels” each consisting of a reactor and condenser/evaporator. The two barrels can operate in parallel.

Cooling

The water returns from the distribution system at a higher temperature than when it left the condenser / evaporator (we have cooled the building). This heat causes the water in the evaporator to boil and the steam passes down to the reactor, where it condenses, since the reactor is relatively cooler. Steam that condenses into water in the reactor will dilute the LiCl solution. The diluted LiCl solution is then pumped through the filter basket, where it mixes with the salt and regains its saturation. The saturation is needed to continuously provide a temperature difference between the condenser/evaporator and the reactor.

image273

680 mm 680 mm

Barrel ABarrel B

Mode

Storage Capacity *

Maximum Output Capacity **

Electrical

COP[7]

Thermal

Efficiency

Cooling

60 kWh

10/20 kW

77

68%

Heating

76 kWh

25 kW

96

160%

* Total storage capacity (i. e. including both barrels)

** Cooling capacity per barrel: 10 kW cooling is the maximum capacity. If both barrels are used in parallel (double mode) the maximum cooling output is 20 kW and the maximum heating output is 25 kW.

image274

Heating

Подпись: The example below shows one of the two barrels discharging heating

Heating is just cooling in reverse, meaning that the charged energy is extracted as heat by connecting the condenser/evaporator to the heat sink and the reactor to the distribution system. Water returns from the distribution system at a lower temperature than when it left the reactor (we have heated the building). This water boils the water in the condenser/evaporator and steam passes down to the reactor. Steam condenses into water which dilutes the LiCl solution in the reactor. The diluted LiCl solution is pumped through the salt filter basket where it mixes with the salt and regains its saturation. The saturation is needed to continuously provide a temperature difference between the condenser / evaporator and the reactor. During discharging, the heating energy is extracted by connecting the evaporator to the heat sink and the reactor to the distribution system. Under charging, heat can also be extracted by connecting the condenser to the distribution system under charging mode.

5 m3 solar ammonia-carbon adsorption refrigerated container for food preservation

R. E. Critoph

University of Warwick, Coventry, CV4 7AL, UK.

R. E. Critoph@warwick. ac. uk
Abstract

An ‘alpha-version’ solar adsorption refrigerator for chilled food preservation is being developed jointly by the University of Warwick and Advanced Technology Materials Inc. The requirement is to cool a 5m3 insulated container in temperatures up to 40°. The prototype uses an ammonia — active carbon pair in a 2-bed cycle with heat and mass recovery. An ice-bank is created within the container during the day to act as a thermal store. The driving heat is supplied by 10m2 of evacuated tube collectors via a pressurised water loop. Simulations suggest that between 1 and 2 kW of cooling can be supplied given reasonable levels of insolation and that the adsorption cycle time may be made a simple function of insolation alone.

Keywords: Solar Adsorption Refrigeration Ammonia

1. Introduction

Previous research at the University of Warwick on a mobile air conditioning system [1] has resulted in a patented concept [2] for a highly compact solid sorption reactor. In the work reported here the technology is applied to solar powered refrigeration. There is a requirement for maintaining chilled food at 0-5 °C in transportable containers in remote areas away from grid electricity. The conventional technology solution is to use vapour compression refrigeration powered from motor-generator sets. The University and Advanced Technology Materials Inc. (USA) are collaborating in the development of a solar thermal powered system, which will have parasitic power for controls etc. delivered by PV’s. The ‘alpha’ version, is due for field testing in Arizona from November 2008.

2. Specification

The standard insulated container, manufactured by CMCI (Figure 1) has external dimensions 2.4 m x 1.5 m x 2.1 m and internal volume of 4.7m3 . It is normally cooled by a conventional vapour compression chiller, rated at about 2kW cooling at 2°C. It is required to maintain normal use at ambient temperatures of 40°C using a solar thermal cooling system.

3. Design

Naturally, any solar powered system requires thermal storage and it has been decided to use an ice bank integrated with the flooded evaporator of the refrigerator. Approximately 50 kg of ice is needed and this is incorporated into a vertical wall within the container. The wall has enough fins extending into the cold space so that cooling within the container is achieved by natural convection. Figure 2 shows the complete evaporator/ice-bank assembly and Figure 3 shows the flooded evaporator alone.

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The evaporator consists of approx. 40 vertical half inch diameter tubes with a large reservoir above and parallel feed below. Later versions will have direct expansion evaporators which will have the advantages of lower mass, lower cost and reduced refrigerant charge, but the overwhelming advantage of a flooded evaporator is that requires less development and is comparatively risk free. Each of the vertical tubes fits tightly between the fins of an aluminium extrusion that forms part of the ice-bank. Without this heat transfer enhancement, towards the end of the process of freezing the water, the evaporating temperature would drop significantly as heat from the freezing front had

Подпись: Figure 3: Flooded Evaporator Schematic to be conducted through an increasing thickness of ice, thereby reducing the system COP (cooling power / driving heat input).

The same aluminium extrusion is used on the outside of the ice tank to transfer heat to the cold space.

The refrigeration system is based on an adsorption cycle using ammonia as refrigerant and active carbon adsorbent. The plate heat exchanger (sorption generator) developed for this and other applications is shown in Figures 4 and 5. It is of low thermal mass (for good COP) and allows rapid cycling which reduces the physical size of the sorption generator for a given cooling power. Two of these generators, about 200mm on one side have been used in ‘TOPMACS’, an EU project to create a car air conditioning system that is driven by the waste heat of the engine. Laboratory operating results are given in Figure 6, which shows a mean cooling power of 1.6kW. The solar ice-making application is similar and design work is based on two similar but improved units. In the original generator, granular carbon in 4mm thick layers was sandwiched between stainless steel shims containing numerous water channels for heating and cooling the

carbon. The new design will utilise a more highly conductive (~2.0 W m K-1) carbon developed by ATMI

Подпись: Figure 4: Plate Sorption generator from ‘TOPMACS’ project with manifolds and flanges

which will enable the use of 12mm carbon layers, reducing cost, complexity and thermal mass. The two beds will be operated in a simple cycle with both mass and heat recovery, with typical cycle times of 2 minutes.

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Slots for adsorbent

 

Water channels

 

Figure 5: Solar refrigerator plate sorption generator core with 12 mm slots

 

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Подпись: Figure 6: Experimental cooling output of two bed sorption machine (TOPMACS)

The original design for car air conditioning was heated or cooled by unpressurised water. The solar collectors are expected to operate at well above 100°C and so the choice had to be made between using a heat transfer fluid or pressurised water. The heat transfer properties of water are so superior that pressurised water was selected.

A schematic of the whole refrigeration system is shown in Figure 7. Hot water from the solar collectors is pumped to either G1 or G2, heating the carbon within and desorbing ammonia. The ammonia flows through a check valve to the condenser where heat is rejected, through the float type expansion valve to the flooded evaporator / ice bank, where it boils and produces useful cooling. From there it passes through a check valve and into the other generator, where it is adsorbed. It is necessary to remove the heat of desorption via another pressurised water loop to an

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air cooled heat exchanger. At a suitable time (optimised for maximum cooling power) the cycle is reversed and the bed that was desorbing becomes the adsorber and vice versa. At the change-over, the performance can be improved, firstly by briefly opening the mass recovery line to equalise pressures in the two generators and then by using the heat recovery loop to pre-heat one bed with the reject heat of the other.

Given the high collector temperatures, the only commercially available options are evacuated tubes. The collectors used in our simulation models are Thermomax DF100 2m2 panels which feature direct flow of the fluid (water at up to 8 bar pressure) through the tubes. Figure 8 shows the manufacturer’s performance data. 10m2 collector area will be used to obtain a peak cooling power of up to 2 kW at an ambient temperature of 40°C in a desert environment.

A critical area of design is the waste heat rejection from the condenser and adsorbers. This is done using conventional fan coils and with attention being paid to minimising the fan power. The design compromises are critical. A small compact heat exchanger may have higher temperature differences which lead to lower COP and hence more heat to be rejected. It may also need more fan power and since parasitic electrical power will be met from PV, this must be minimised. Conversely, very large heat exchangers could be both impractical and costly. The compromise chosen uses a direct condenser measuring 650 h x 900 l x 570 w and with a 66W fan motor and a cooler measuring 650 h x 900 l x 470 w with a 102W fan motor.

Collector efficiency, G=1000W/m2, Tamb=40°C

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Figure 8: Performance curves of chosen evacuated tube collectors

4. Simulation

The operation of the complete system has been modelled in Matlab to assist the design. The operation of the chiller has to be modelled at a timestep of about 0.001 s which is obviously impractical for modelling several days of operation. This problem has been overcome by deriving a pseudo-dynamic model in which the chiller is assumed to respond much more quickly (within minutes) than changes in the load or ambient conditions.

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Figure 9 : Performance envelope for ambient temperature 30°C and evaporating temperature -10°C

An example of the approach is given in Figure 9 in which each point (derived from detailed simulation every 0.001 s) corresponds to a balance between the heat input from the collectors at the particular insolation and ambient temperature, together with a particular cycle time (control

parameter) and evaporating temperature (corresponding to the state of the load). The envelope of the points (linear and quadratic are shown) gives the instantaneous cooling power corresponding to the best control strategy for that particular evaporating temperature and ambient temperature for the full range of insolation.

Подпись: Figure 10 : Optimum performance envelope and that based on a simplified control algorithm for ambient temperature 30°C and evaporating temperature 10°C
A set of these correlations for a range of evaporating and ambient temperatures may be combined empirically to yield a polynomial function for optimum cooling power under any conditions which can act as input to the model of the ice-bank and cold box. Preliminary examination of these results implies that the cycle time can be made a simple function of insolation only (i. e. ignoring ambient and evaporating temperature) with comparatively little penalty. Figure 10 illustrates this for the particular case of an ambient of 30°C and evaporating at 10°C.