Category Archives: EuroSun2008-7

Monitoring equipment and evaluation of raw data

The data acquisition system consists of internally integrating heat meters with matched Pt100 type temperature sensors. The integrator has a sampling rate of 1s and calculates cumulated energy amounts and mean temperatures and powers. This internal sampling rate assures a correct collection of energy data for the highly dynamic temperature patterns characteristic of adsorption systems. The integrator and further temperature sensors are read out by a computer with a sampling rate of 15s. The monitoring software further reduces these values to cumulated energies and mean temperatures which are stored with an interval of 5 minutes in the raw data measurement file. The storage interval can be set by the system operator and thus allows a flexible data management. The post processing of the raw data further reduces the values to hourly accumulated and mean values — depending on the quantity considered. For the hourly mean temperature values also a standard deviation is calculated in order to judge the stability of the temperature within the evaluated hour.

Experimental performance parameters

Experimental cooling capacity and COP is obtained from the external fluid temperatures in evaporators as follows:

 

QE, exp mchw Cw (tchw, in Phw, o )

From external fluid temperatures in generator:

QG, exp moil Coil (toil, in ^ oil, o )

Corresponding values per unit mass flow rate are:

Qe, exp

 

(16)

 

(17)

  image651

(18)

  image652

(19)

  image653

(20)

 

5. Results

Overview of operation period

The results presented here correspond to the operation period from June 4th, 2007 to December 1st, 2007. From these 180 days the system was in operation during 108 day. The days without operation were either weekends or days without monitoring data due to a lightning stroke into the data acquisition system. Within these 108 days 1034 hours of operation were registered. In some hours the machine was in operation only a few minutes, but in most of the registered hours a

continuous operation was observed. In Table 1 the number of hours with an operation time within given limits is shown.

Operation

mode

Number of hours with an operation time t (in minutes) within the given limits

<15 min

15<t<30 min

30<t<45min

45<t<60 min

60min

cooling

65

23

36

37

282

heating

56

72

28

73

398

solar

39

35

34

26

176

Table 1.Statistics of the operation hours

4. Operation results

Basic absorption model vs. experimental results. Factors affecting the facility performance

Подпись: to [°C] Fig. 3. Comparison of basic model and experimental results. (a) Ideal and experimental cooling capacity vs. tG (b) Ideal and. experimental COP vs. tG image655

Fig. 3 illustrates the comparison of experimental performance parameters with those obtained through the basic model, Eqs. (14) and (15), for 220 test points. The differences observed help identifying the influence of components performance, serving as a diagnosis. In the following, the causes of deviation from the ideal conditions are explained, evaluated and included in the model.

5.1. Influence of components performance. Modified basic model.

The evaporators in this facility layout show low performance [8]. Because of this, part of m ref

supplied to them is not evaporated, but overflows. Therefore, in order to measure the performance of evaporators (fraction of evaporated refrigerant), the cooling capacity obtained if all refrigerant is used in the evaporators Qref is compared with QE exp:

Пе = ; Qref = mref • (hEin — hEo) (21)

Qref

Energy balance in the generator yields:

m = m — m

ref weak strong (22)

The obtained average nE is 50% which is considerably low. With the current design, the excess of refrigerant is not recirculated to evaporators, and therefore it goes to the solution reservoir inside the absorber. Beside this, it has been detected that the distribution of mref (m f and m f) is not symmetrical in some cases, and therefore one of the evaporators runs dry. This situation also explains the low value of qexp and its tendency is not consistent with rise in temperature (Fig. 3). Corresponding values of COPexp show obviously the same behaviour.

Taking into account the evaporators limitations explained above, the modified cooling capacity will be:

4E, mod 4Ei ‘ ПE (23)

The facility performance is also affected by the solution heat exchanger efficiency phex and consequently is considered in the analysis. Beside this, a noticeable difference was identified between solution temperature at generator outlet and separator outlet, as the generation process continued and the refrigerant separation kept on in the path followed by the solution. The temperature drop associated ranged from 4°C to 10°C. The heat transfer associated is called qG.

Then, considering the real efficiency of the solution heat exchanger and the heat transfer losses in the system generator-separator, the modified generation power will be:

qG, mod = qG, i + qhex. i (1 — Vhex ) + qG-sep (24)

As a result, the modified COP becomes:

COPmod = ^mo1 (25)

qG, mod

Подпись:
Results obtained incorporating the specific behaviour of components to the basic model can be appreciated in Fig. 4. The predicted cooling capacity and COP shows good agreement with experimental results. This indicates that the main causes of deviation have been included in the model, thus verifying the preliminary diagnosis.

(a) (b)

6. Conclusions

A facility representing a single effect adiabatic LiBr/water chiller implementing adiabatic absorption facility has been operated under different conditions, corresponding to both design and off-design operational conditions. Performance parameters have been experimentally determined for every test condition.

The differences observed between ideal and experimental results helped to diagnose and identify the influence of components performances on the overall performance of the facility. The

evaluation of the ideal and real cooling powers allowed detecting a low performance of evaporators caused by both dry operation and overflow. Another influence factors which causes the deviation from ideal behaviour are the solution heat exchanger efficiency and the heat transfer losses in the system generator — separator.

Taking into account the particular layout and operation features tested, a good agreement with experimental performance parameters and those obtained through a modified basic absorption model has been achieved. This model incorporates a quantified loss and/or efficiency separating from ideal and has been fitted to experimental data. Experimental results demonstrate the operational possibilities and flexibility of the design, showing a great potential for further work.

Acknowledgements

The financial support of this study by the Ministry of Education, Science and Technology through CLIMABCAR project DPI 2003-01567, TRANSMACA project DPI 2002-02439 and MINICOM project (FIT 0204-2004-68 and FIT 020100-2003-233), is greatly appreciated. The authors express their gratitude to the technicians of Universidad Carlos III de Madrid Mr Manuel Santos and Mr Carlos Cobos for their invaluable help in this work.

References

[1] Venegas M., Izquierdo M., Rodriguez P., Lecuona A. Heat and mass transfer during absorption of ammonia vapour by LiNO3-NH3 solution droplets, Int. J. Heat Mass Transfer, 47 (12-13) (2004) 2653­2667.

[2] Venegas M., Rodriguez P., Lecuona A., Izquierdo M. Spray absorbers in absorption systems using lithium nitrate-ammonia solution, Int. J. Refrigeration, 28 (4) (2005) 554-564.

[3] Arzoz D., Rodriguez P., Izquierdo M. Experimental study on the adiabatic absorption of water vapor into LiBr-H2O solutions, Applied Thermal Engineering, 25 (5-6) (2005) 797-811.

[4] Warnakulasuriya F. S.K., Worek W. M. Adiabatic water absorption properties of an aqueous absorbent at very low pressures in a spray absorber, Int. J. Heat Mass Transfer, 49 (9-10) (2006) 1592-1602.

[5] Elperin T., Fominykh A., Orenbakh Z. Coupled heat and mass transfer during nonisothermal absorption by falling droplet with internal circulation, Int. J. Refrigeration, 30 (2) (2007) 274-281

[6] Wang L., Chen G. M., Wang Q., Zhong M. Thermodynamic performance analysis of gas-fired air-cooled adiabatic absorption refrigeration systems. Applied Thermal Engineering. 27 (8-9) (2007) 1642-1652.

[7] Gutierrez G., Venegas M., Rodriguez P., Izquierdo M., Lecuona A. Experimental characterization of a single stage LiBr-H2O absorption test rig. In Proc. ECOS 2006, Vol. 3, Crete, Greece, July 12-14 (2006) p. 1331-1316.

[8] Gutierrez G., Zacarias A., Venegas M., Rodriguez P. Cooling power evaluation of a water lithium bromide absorption test rig. In Proc. ECOS 2007, Vol. 2, Padova, Italy, June 25 -28 (2007) p. 1183-1190.

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).

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.