Evaluation of liquid dessicant based evaporative cooling cycles for typical hot and humid climates

Heat Recovery Systems & CHP Vol. 14, No. 6, pp. 621-632, 1994 Elsevier Science Ltd 0590-4332(93)E0013-V Printed in Great Britain 0890-4332/94 $7.00 + .00

Pergamon

EVALUATION OF LIQUID DESSICANT BASED EVAPORATIVE COOLING CYCLES FOR TYPICAL HOT AND HUMID CLIMATES SANJEEVJAIN*, P. L. DHAR* and S. C. KAtlSmK~ *Mechanical Engineering Department and tCentre of Energy Studies, Indian Institute of Technology, Hauz Khas, New Delhi-ll0016, India (Received 24 August 1993)

Abstract--This communication presents an evaluation of various liquid dessicant cycles for air conditioning in hot and humid climates. Psychrometric evaluation of seven potential cycles for achieving standard comfort conditions (25°C/10 g/kg) in rooms has been carried out for 16 typical Indian cities. A computer simulation model is based on the constant effectiveness of heat exchangers (HX)/evaporative coolers (EC) and wet surface heat exchangers (WSHE). The absorber-cum-dehumidifier is assumed to provide air at the specified humidity level, while the outlet air temperature is taken to be equal to the cooling water temperature. The effect of various outdoor conditions and the effectiveness of HXs/ECs on cooling COP and volumetric air flow rate per unit cooling capacity have been investigated. It was found that a combination of dehumidifier and WSHE is better, in terms of COP, for a wide range of outdoor conditions. The results should be useful in the design of liquid dessicant based air conditioning systems suitable for the monsoon season in tropical countries like India.

NOMENCLATURE AM CI

c~ CMH COP E h~ h,~ hfg h~ h~ HCVA HRE

Q,~s RT

T. L,

TR

r~ u

v. w. REM

Mass transfer area The ratio of the total heat required during regeneration to the heat actually used for vaporization of moisture in regenerator The ratio of the solution temperature increase in solution heater to the temperature decrease in regenerator Volumetric air flow rate Coefficient of performance Specific heat of moist air Volumetric air flow rate per unit tonne of refrigeration Effectiveness of solution-solution heat exchanger Convective heat transfer coefficient Convective mass transfer coefficient Latent heat of vaporization of water Enthalpy of air at room conditions Enthalpy of air at supply air conditions Heat convected to air from solution in the regenerator Heat required for vaporization of moisture from solution Amount of heat required for regeneration Regeneration temperature required for solution Temperature of air Temperature of solution at outlet of regenerator Tonne of refrigeration = 3.5167 kW Temperature of solution Temperature of solution at outlet of absorber/regenerator Temperature of solution at inlet to the solution heater Overall heat transfer coefficient from solution to air Specific volume of the supply air Absolute humidity of air Equilibrium humidity of solution Amount of moisture removed from air in the absorber

!.

INTRODUCTION

Depleting energy resources and increasing environmental pollution have shifted the attention of researchers all over the world to alternative air conditioning systems; this is because conventional vapour compression systems use CFCs causing ozone layer depletion and contributing to the greenhouse effect, as well as demanding high grade electric energy. 621

622

S. JAIN et al.

One of the promising alternatives for comfort conditioning applications is the evaporative cooling system. This system has been widely used for hot and dry climates, but for humid climates and even, during the rainy season, in regions where summers are hot and dry, humidity levels are quite high, rendering evaporative cooling ineffective and restricting the widespread acceptance of evaporative cooling in countries like India. This deficiency can be overcome by using desiccants to remove the bulk of the moisture and evaporative cooling to reduce temperature. While both solid and liquid dessicants can be used, the latter have the advantage of a lower air pressure drop and simultaneous cooling during dehumidification, also, the required regeneration temperatures are much lower. Many liquid desiccant cycles have been previously reported. The earliest known liquid dessicant system was suggested and experimentally tested by L6f [1]. In this system return air from a room was dehumidified and simultaneously cooled in an absorber and then evaporatively cooled. The system used TEG as the absorbent and solar heated air for regeneration purposes. Sheridan [2] proposed a similar system using lithium chloride solution.

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Liquid dessicant based evaporative cooling

623

Bolzan and Lazzarin [3] suggested a system which used calcium chloride water mixture and, for the summer data, simulation of system performance for a week gave an average COP of 0.2. Robison [4, 5] designed a liquid dessicant system using large quantities of cold well water at 17.8°C. Turner [6] proposed a system which generated cooling water internally by evaporatively cooling the recirculated water with a portion of processed dry air; this system of operation was termed the Process Recirculation Mode (PRM). Peng and Howell [7] studied Turner's system and suggested an alternative operating mode called Exhaust Recirculation Mode (ERM). Mathematical modelling and simulated studies were carried out on these cycles. Johannsen [8] performed a computer simulation of a solar operated liquid dessicant system. The results, for five cities in South Africa, showed summer COP variation of 0.36-0.47.

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Agrawal e t al. [9] have suggested a dessicant cycle with a plate heat exchanger before the evaporative cooler. Jain and Singh [10] analysed the two cycles: PRM and ERM using lithium bromide as the absorbent, and found that the performance degraded rapidly in humid climates. They proposed a new cycle, discussed in the next section. Godara [11] proposed two new cycles and carried out a comparative analysis for six cycles. Jain [12] designed and fabricated a system based on Godara's cycle but reported no performance results. The effect of various parameters on the system performance was predicted. Scalabrin and Scaltriti [13] proposed a system with a large number of heat exchangers and studied system performance by varying the load and air supply conditions. Kumar [14] compared four liquid dessicant cycles for hot climates, assuming a constant value of effectiveness for heat exchangers and evaporative coolers. There have been a few experimental studies [15-22] related to liquid dessicant systems. In these studies, the results have mainly been reported for different absorber and regenerator configurations. The overall COP of the cycles is around 0.24).3. 2. A DESCRIPTION OF LIQUID DESSICANT CYCLES In general all liquid dessicant cycles consist of four major components: namely, dehumidifier, air-air and air-water heat exchangers and an evaporative cooler. Cycle 1 proposed by Godara [l l] is shown in Fig. 1. In this cycle, room air (after being evaporatively cooled) cools the dehumidified air, which, in turn, is sent to the room after evaporative cooling. Outdoor air is mixed with the room air just before the dehumidifier, as shown in the psychrometric chart.

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Psyehrometrie chart for cycle 2 Fig. 4. Cycle 4 psychrometric chart.

In cycle 2 (Fig. 2), proposed by Jain and Singh [10], there is no evaporative cooling of the room air. Otherwise it is similar to cycle 1. In cycle 3 (Fig. 3), proposed by Bolzan and Lazzarin [3], outdoor air is mixed with room air and the mixed air is evaporatively cooled. The rest of the cycle is similar to cycle 1. Cycle 4, proposed by Godara [11], is an extension of cycle 1 (Fig. 4). In this, part of the cooled, dehumidified air cools the rest of the air further, after itself being evaporatively cooled. The cooled air is then supplied to the room after evaporative cooling. Cycle 5 is a new cycle which makes use of the idea of regenerative evaporative cooling, so as to achieve cooling of air below its wet bulb temperature. In this cycle, room air mixed with outdoor air is dehumidified and then cooled in a wet surface heat exchanger before being supplied to the room (Fig. 5).

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5 = DBT Proposed new cycle

Fig. 5. Schematic diagram and psychrometric chart for cycle 5.

626

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Cycle 6, known as the PRM and proposed by Turner [6], is shown in Fig. 6. Here the mixture of outdoor and room air is dehumidified and cooled in an absorber to obtain the supply air. The cooling water is obtained in an evaporator by using part of the supply air. In cycle 7, called ERM (Fig. 7) and proposed by Peng and Howell [7], a portion of return air is used for evaporatively cooling the water required in the absorber. All these cycles have been analysed and studied in this paper for different outdoor conditions. In cycles other than these, it is not always possible to maintain comfort conditions. 3. C O M P U T E R S I M U L A T I O N OF CYCLES A computer program has been developed to simulate the performance of the aforementioned dessicant cooling cycles. The psychrometric data are represented by moist air property correlations obtained from refs [23-25]. It has been assumed that the absorber dehumidifies the air to the specified minimum humidity level whatever the inlet humidity. Also, that the temperature of air after dehumidification is equal to the inlet cooling water temperature. The supply air states for the PRM and ERM cycles have been assumed to be the same as those for cycle 1.

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The effectiveness of air-air and air-water heat exchangers, evaporative coolers, cooling towers and degree of saturation in the evaporator (for PRM and ERM cycles) have been assumed to be 0.8, unless otherwise varied to study their influence. The effectiveness of the wet surface heat exchanger was kept to 1.2. The amount of ventilation was fixed at 10% and the minimum humidity at 0.004 kg/kg d.a. Two performance indices: COP and volumetric air flow rate per unit tonne of refrigeration (CMH/TR) were predicted from the computer simulation results. COP values have been calculated as follows: COP = Specific cooling effect (enthalpy difference between supply and room air states) Regeneration heat supplied per kg of dry air Air circulation rate per unit TR can be given by: CR = C M H / T R =

3.5167,3600.0, Vs hr - hs

where V, is the specific volume of the supply air. Regeneration heat has been evaluated using the relation: Qreg -- WREM,hrg,C] *C2, HRS 14/0---D

628

S. JAIN et al. Table I. COP and Cm values of cycles for different outdoor conditions Ambient S. No. 1. 2. 3. 4. 5. 6. 7. 8. 9. 10. 11. 12. 13. 14. 15. 16.

City Ahmedabad Amritsar Bhopal Bombay Calcutta Coimbat Delhi Hyderabad Jodhpur Lucknow Madras Nagpur Patna Roorkee Trivand Visakhapatnam

Cycle 1

Cycle 2

Cycle 3

Cycle 4

DBT

WBT

COP

CR

COP

CR

COP

CR

COP

CR

42.8 42.5 41.7 34.5 39.5 36.7 43,0 39,5 43.5 42.8 39.2 42.9 42.4 42.5 32.9 38.4

27.6 27.9 25.3 28.4 29.3 28.3 28.1 25.3 27.9 28.3 28.5 27,5 28.1 27.8 27.2 30.4

0.39 0,38 0.41 0,37 0.37 0.37 0.38 0.41 0.39 0.38 0.38 0.39 0.38 0.39 0.38 0.35

556 557 544 550 561 552 559 542 559 560 556 556 558 557 543 565

0.39 0.38 0.42 0.37 0.36 0.37 0.38 0.42 0.38 0.38 0.37 0.39 0.38 0.38 0.38 0.34

742 744 720 731 750 735 747 715 747 749 742 742 746 743 717 758

0.35 0.35 0.38 0.35 0.34 0.35 0.35 0.38 0.35 0.35 0.35 0.35 0.35 0,35 0,37 0.33

587 589 567 579 596 582 592 563 59t 594 589 587 591 589 566 605

0.41 0.40 0.43 0.39 0.38 0.39 0.40 0.43 0.40 0.40 0.39 0.41 0.40 0.40 0.40 0.37

522 523 510 516 526 518 525 507 524 525 522 522 524 523 508 530

The units of D B T and WBT are °C and of CR are m 3 h ~TR

where WREM is the amount of moisture to be removed in the regenerator; hcg is the latent heat of vaporization; C~ is the ratio of the total heat required during regeneration to the heat actually used for vaporization of moisture in the regenerator and C2 is the ratio of the temperature increase of the solution in the solution heater to the temperature decrease in the regenerator. The method of estimating C~ and C2 is discussed in the Appendix.

4. RESULTS AND DISCUSSIONS The performance of the seven cycles discussed earlier was evaluated at 16 typical summer design conditions [26] taking 1% DBT and WBT values. The results obtained are tabulated in Tables 1 and 2. It is found that: (i) PRM and ERM cycles give higher values of COP when followed by cycle 5 consisting of WSHE. (ii) The performance of both PRM and ERM deteriorates sharply with a decrease in wet bulb depression. (iii) Cycles l, 3 and 4 require lower air circulation rates when compared to other cycles. (iv) For cycles 1-4, COP decreases with increasing air humidity but the decrease is marginal. (v) The increase in COP of cycle 4 over cycle 1 is not significant although additional equipment, like HX and EC, is required. Table 2. COP and C m values of cycles for different outdoor conditions Ambient S. No. 1. 2. 3. 4. 5. 6. 7. 8. 9. 10. l 1. 12. 13. 14. 15. 16.

City Ahmedabad Amritsar Bhopal Bombay Calcutta Coimbatore Delhi Hyderabad Jodhpur Lucknow Madras Nagpur Patna Roorkee Trivandrum Visakhapamam

Cycle 5

Cycle 6

Cycle 7

DBT

WBT

COP

CR

COP

CR

COP

CR

42.8 42.5 41.7 34.5 39.5 36.7 43.0 39.5 43.5 42.8 39.2 42.9 42.4 42.5 ~32.9 38.4

27.6 27.9 25,3 28.4 29.3 28.3 28.1 25.3 27.9 28,3 28,5 27.5 28.1 27.8 27.2 30.4

0.55 0.53 0.69 0.44 0.44 0.46 0.52 0.65 0.54 0.51 0.47 0.56 0.52 0.54 0.48 0.39

714 715 698 706 719 709 717 694 717 718 714 713 716 715 696 723

0.74 0.68 0.92 0.18 0.40 0.31 0.71 0.75 0.73 0,65 0,43 0.75 0.66 0.68 0.17 0.26

722 740 706 1302 856 978 725 756 725 743 849 721 741 739 1440 1000

0.63 0,57 0.80 -0.25 0.14 0.60 0.61 0.62 0.55 0.28 0.64 0.56 0.58 -0.11

776 800 760 -1083 1657 780 856 780 803 1075 776 801 799 -1694

The units of ambient D B T and W B T are °C and of CR are m 3 h - i T R - ~. - - indicates that the cycle is unfeasible for that condition.

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Fig, 9. The effectof heat exchangereffectivenesson the Cm of cycle 4 for different outdoor conditions.

To study the influence of different components on system performance; effectiveness of heat exchangers, cooling tower, amount of ventilation, minimum humidity level and degree of evaporator saturation (for PRM and ERM cycles) have all been varied and plotted in Figs 8-16 for a typical cycle 4. The findings are as follows: (i) The COP of the cycle increases and the air circulation rate reduces with an increase in heat exchanger effectiveness [Figs 8 and 9]; this is due to a reduction in supply air temperature. (ii) A higher cooling tower effectiveness lowers the cooling water supply temperature and thus the outlet temperature of air from the absorber is lowered; this improves the COP of the cycle, while volumetric air flow rates are reduced [Figs l0 and 11]. (iii) An increase in percentage ventilation increases the dehumidification load on the absorber, thus decreasing its COP [Fig. 12]. The air flow rates remain the same for all cycles except cycle 3. For cycle 3 mixed air is used for heat exchange. Therefore the supply air temperature and, thus, the C M H / T R changes. 0.50

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Fig. l 1. The effect of cooling tower effectivenesson Cmof cycle 4 for different outdoor conditions.

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Fig. 12. The effect of ventilation on the COP of cycle 4 for different outdoor conditions,

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Fig. 13. The effect of the degree of evaporator saturation on the Cm of the PRM cycle for different outdoor conditions.

(iv) An increase in the degree of saturation of the air in the e v a p o r a t o r (i.e. the ratio of the actual humidity of the air leaving the e v a p o r a t o r to the saturation humidity o f the air leaving at the outlet temperature) for P R M and E R M cycles reduces the air flow requirement in the e v a p o r a t o r assuming that the water temperature is the same (Fig. 13). This, in turn, decreases the volume of o u t d o o r air to be mixed with r o o m air. The humidity level of mixed air decreases, thus increasing the C O P o f cycles (Fig. 14). (v) On one hand, a decrease in the m i n i m u m humidity level in the absorber increases the water removal rate and thus the regeneration heat requirement. On the other hand, the specific enthalpy difference between supply and r o o m air will increase. It can be seen that the C O P value increases and the air flow requirement goes down (Figs 15 and 16).

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250

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Fig. 15. The effect of minimum humidity level on the Cmof cycle 4 for different outdoor conditions.

Liquid dessicant based evaporative cooling

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Fig. 16. The effect of minimum humidity level on the COP of cycle 4 for different outdoor conditions. CONCLUSIONS In the present study seven potential liquid desiccant cycles have been analysed. Liquid desiccant systems have the advantage of a lower pressure drop and simultaneous cooling during dehumidification. The m a i n conclusions emerging from the study are: (i) A cycle with a wet surface heat exchanger gives a consistently high C O P but the air circulation rate needed per tonne o f refrigeration is also higher. (ii) The p e r f o r m a n c e o f P R M and E R M cycles fluctuates widely for different o u t d o o r conditions. These are thus unsuitable for tropical weather. (iii) Higher effectiveness values of air-air heat exchangers, lower cooling water inlet temperatures and dehumidification of air to lower humidity levels reduces air circulation rates and improves the C O P of cycles. (iv) An increase in the a m o u n t o f ventilation reduces the cycle COP; it should, therefore, be kept to a m i n i m u m value, as was found necessary from indoor air quality considerations. (v) An increase in the degree o f saturation of the e v a p o r a t o r of the P R M cycle also improves the C O P and reduces air requirement. (vi) In comparison, the C O P values of v a p o u r absorption systems are in the range 0.3-0.8 [27, 28]. Also, these systems operate either under v a c u u m or high pressure, in contrast to the atmospheric pressure operation of desiccant systems. The results obtained herein should be useful for the design and development of liquid desiccant based air conditioning systems.

REFERENCES 1. G. O. G. Lof, Cooling with solar energy. Congress on Solar Energy, Tucson, Arizona, pp. 171-189 (1955). 2. N. R. Sheridan, Solar air conditioning. J. L E. Australia, 47-52 (1961). 3. M. Bolzan and R. Lazzarin, Open cycle absorption cooling devices with spray chamber regeneration and air solar collectors. Proc. of XVth Int. Congress of Ref., Venice, Vol. IV, pp. 697-704 (1979). 4. H. I. Robinson, Liquid sorbent air conditioner. In Alternative Energy Sources (edited by T. N. Veziroglu). Hemisphere, New York (1978). 5. H. I. Robinson and S. H. Houston, Absorption/desorption solar cooling system performance. Fund. Appl. Solar Energy, AICh.E Symposium Series, 76, 198, pp. 139-143 (1980). 6. N. C. Turner, Cooling method and system. U.S. Patent, 4,171,620, 23 October (1979). 7. C. S. P. Peng and J. R. Howell, Optimization of liquid desiccant systems for solar/geothermal dehumidification and cooling. J. of Energy, 5, 6, 401-408 (1981). 8. P. Johannsen, Performance simulation of a solar air conditioning system with liquid desiccants. Int. J. of Ambient Energy, 5, 2, 59-88 (1984).

632

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APPENDIX In the present study, C~ and C 2 are assumed to be constant, although the actual values may vary depending on the operating conditions. C~ is defined as the ratio of the total heat required during regeneration to the heat actually used for the vaporization of moisture in the regenerator: C~ = 1 + HCVA/HREM = 1 + [U.AM.(T~- TD]/[AM.h,,.(Woq- l~,)hrg] 1 + [hc/hm].[(T s - Ta)/Weq- Wa)].[1/hfg] ~- 1 +[Cpm/hfg].[(rs - ~a)/(Wcq- Wa)1.

For typical values of parameters: C t = 1 + [(1.01/2500).(15/12,10 3)] -~ 1.5 (approx). C 2 is defined as the temperature increase of the solution in the solution heater to the temperature decrease in the regenerator: C2 = ( R T -

T s 2 ) / ( R T - T~t)

where T~2 is given by the effectiveness relation: E = (T~2- T~)/(Tr~- Tst),

or Ts2 = E*T~t + (I -- E)*Tsl.

Thus, C 2 = 1 + [(1 - E ) ( T r , -

T~I)I/(RT - T~l ).

For typical values of parameters, C2 = 1 + (0.2,40/16) -~ 1.5 (approx).

Thus both C~ and C 2 are approximately the same and equal to 1.5.