Studies on Solid Desiccant Based Hybrid Air-conditioning Systems

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Studies on solid desiccant based hybrid air-conditioningsystems

P.L. Dhar*, S.K. Singh

Department of Mechanical Engineering, I.I.T. Delhi, New Delhi 110 016, India

Received 3 August 1999; accepted 27 February 2000

Abstract

Desiccant based air-conditioning systems oer a promising alternative to conventional air-

conditioning systems using vapour compression refrigeration especially under conditions involving highlatent loads. The desiccant can be used either in a stand-alone system or coupled judiciously with a

vapour compression system to achieve high performance over a wide range of operating conditions. In

this paper, the results of a detailed study of solid desiccant-based hybrid air-conditioning systems are

presented. The literature review revealed that various authors dier in their evaluation of the ecacy of

these systems. This seems to be due to dierent methods of modelling of dehumidi®er and dierences in

the operating conditions of the cycles employed. Accordingly, the performance of four hybrid cycles

(which include a new proposed cycle) for typical hot-dry and hot-humid weather conditions has been

evaluated using a detailed procedure for the analysis of rotary dehumidi®er, the most commonly

employed industrial dehumidi®er, based on the analogy method of Maclaine-Cross and Banks [I.L.

Maclaine-Cross, P.J. Banks, Coupled heat and mass transfer in regenerators Ð predictions using an

analogy with heat transfer, Int. J. of Heat and Mass Transfer 15 (1972) 1225±1241]. Eect of room

sensible heat factor, ventilation mixing ratio, and regeneration temperature has also been studied. The

results show that solid desiccant-based hybrid air-conditioning systems can give substantial energysavings as compared to conventional vapour compression refrigeration based air-conditioning systems in

most commonly encountered situations. 7 2000 Elsevier Science Ltd. All rights reserved.

Keywords: Solid desiccant; Hybrid air-conditioning system; Rotary dehumidi®er

Applied Thermal Engineering 21 (2001) 119±134

1359-4311/01/$ - see front matter 7 2000 Elsevier Science Ltd. All rights reserved.

PII: S 1 3 5 9 - 4 3 1 1 ( 0 0 ) 0 0 0 3 5 - 1

www.elsevier.com/locate/apthermeng

* Corresponding author. Tel.: +91-11-685-7753; fax: +91-11-686-2037.

E-mail address: [email protected] (P.L. Dhar).



1. Introduction

In any air-conditioning system two types of loads have to be met, viz. the sensible load

arising from heat transfer into the space and the latent load arising from moisture generated

within the space. The conventional systems use vapour compression refrigeration to meet both

these loads. To meet the latent load, air must be dehumidi®ed by cooling it below its dew

point. Higher the latent heat load, lower is the evaporator temperature required. But there is a

limitation on the evaporator temperature too since it must not go below 08C, the freezing

point, and a very low supply temperature of air can create situations of draft in the air-

conditioned space.

Therefore, in conventional systems reheat is often required in high latent heat load

application, which implies very poor energy eciency. In addition to it, large mass ¯ow

rates are maintained because moisture removal per unit mass of air is limited.

Simple desiccant systems are well suited to meet latent heat loads. Here the process air

is brought in contact with a material with high anity for water. Moisture is absorbed/

adsorbed by this desiccant material and the heat of absorption/adsorption released in the

process heats the air. The air is thereafter cooled by heat exchanger(s) and evaporative

cooler(s). Now if the system has to meet sensible heat load too, the process air should be

over-dried to permit its further cooling by direct evaporation after its sensible cooling

through heat exchange with the surrounding air. This overburdens the dehumidi®er,

especially in hot and humid climates, and it therefore needs a large quantity of heat to

remove the moisture picked up resulting in poor performance of simple desiccant cycles in

these climates.

Clearly, a more ecient process would result if a vapour compression machine and a

desiccant system were combined. This is termed as desiccant-based hybrid air-conditioning

system (Fig. 1). Here the desiccant removes the moisture ingress into the space, and the heat

Nomenclature

CMMSUP supply air ¯ow rate (m3/min)

MOD.VEN-HTX ventilation±heat exchanger cycle

NO. OF DEHUM number of dehumidi®ers

QHTR heater input (kW)

RECIR-CON recirculation±condenser cycleRSHF room sensible heat factor

TEVAP evaporator temperature (8C)

VC CYCLE conventional system using vapour compression refrigeration for the

cooling coil

VEN-CON ventilation condenser cycle

WCOMP compressor work (kW)

W.E. CONS. weighted energy consumption (kW, Heat)

P.L. Dhar, S.K. Singh / Applied Thermal Engineering 21 (2001) 119±134120



exchanger/indirect evaporative cooler and vapour compression system share the sensible heat

load.

For the purpose of comparison all the above mentioned systems are plotted together on a

psychometric chart (Fig. 2). In the conventional system the air is directly cooled and

dehumidi®ed from state M (i.e., a mixture of return air and fresh ventilation air) to state O by

passing it over a coil through which chilled water or refrigerant are circulated. However, in

Fig. 1. Schematic diagram of typical hybrid desiccant cooling system.

Fig. 2. Plot of simple desiccant cooling cycle, simple vapour compression cycle and typical hybrid cycle on

psychrometric chart.

P.L. Dhar, S.K. Singh / Applied Thermal Engineering 21 (2001) 119±134 121



high latent load applications, the state O does not lie on room sensible heat factor (RSHF) line

and therefore reheating of this cold air (O±S) has to be done to meet the load requirements

properly.

In the simple desiccant system the mixed air at M is dehumidi®ed to a very low moisture

content level D2 in the dehumidi®er. The air which gets heated in the dehumidi®cation process

is then sensibly cooled to state C using indirect evaporative cooling (IEC) and ®nally brought

to the same supply state S using direct evaporative cooling. No reheating is needed.In the hybrid system the air is dried to a smaller extent (point D1) and thereafter sensibly

cooled upto the supply state S, partly by IEC (D1±E) and partly by a refrigerant circulating

through a cooling coil (E±S). Clearly, here the dehumidi®er duty is much lesser than that of a

simple desiccant system, but there is a load on the refrigeration coil, though much lesser than

that of a simple refrigeration system. Since the evaporator temperature required to do sensible

cooling (E±S) is higher than that needed for conventional cooling and dehumidi®cation (O±S),

the power consumption per ton of cooling is also lower. But there is an additional requirement

of heat energy for regenerating the dehumidi®er.

It is obvious that without detailed calculations it is not possible to predict which of these

systems would be superior from the point of view of energy consumption and initial cost. This

would depend on factors like performance of the dehumidi®er, the indirect evaporative cooler,

the vapour compression system, the type of hybrid cycle employed, the RSHF, the indoor andoutdoor DBT, and speci®c humidity values.

2. Literature review

One of the earliest comprehensive studies on solid desiccant-based hybrid cooling systems

was done by Burns et al. [1]. They studied three hybrid system con®gurations for supermarket

applications (high latent load) and a comparison of their performance with conventional air-

conditioning system was made. The cycles termed as ventilation±condenser cycle, recirculation±

condenser cycle and ventilation±heat exchanger cycle are shown in Figs. 3±5, respectively.

In the ventilation±condenser cycle only the ventilation air (at state 1) is passed through thedehumidi®er. The dehumidi®ed air (2) is cooled by indirect evaporative cooling (2±3) before

mixing it with the recirculated room air (6). The mixed air (4) is then further cooled sensibly in

a chilled water coil (4±5) till it reaches the RSHF line. To regenerate the desiccant, waste

condenser heat is used to preheat the ambient air (1±7). Any further heating needed (7±8) is

provided by some auxiliary heat source. The regenerative air stream is cooled and humidi®ed

as it passes through the desiccant and then exhausted to the outside, state (9).

The recirculation±condenser cycle (Fig. 4) diers from the ventilation cycle in that, here the

mixture of recirculated air and the ventilation air (state 2) is passed through the dehumidi®er.

Consequently, the dehumidi®er size is larger since it has to handle larger volumes of air but as

is evident on comparing Figs. 3 and 4, the quantity of moisture to be removed per kg of air is

much smaller. Therefore, the regeneration temperature needed is lesser than that needed in theventilation±condenser cycle.

The Heat Exchanger cycle (Fig. 5) is a variant of the ventilation cycle. The hot

dehumidi®ed air is cooled with the help of ambient air in a heat exchanger; no indirect




Fig. 3. Schematic diagram and plot of ventilation/condenser cycle.




Fig. 4. Schematic diagram and plot of recirculation condenser cycle.




Fig. 5. Schematic diagram and plot of heat exchanger cycle.




evaporative cooler is used and the condenser heat is also not utilised. This ambient air is

further heated to the regeneration temperature in the auxiliary heater.

Burns et al. [1] reported that these cycles would give energy savings, in comparison to

the conventional air-conditioning systems, ranging from 56.5 to 66% at moderate ambient

conditions of 308C, 0.016 kg/kg da, RSHF of 0.35 and space conditions of 248C, 0.0104

kg/kg da. These calculations were based on the concept of weighted energy consumption,

with one unit of electrical energy weighted twice that of thermal energy.Sheridan and Mitchell [5] investigated the performance of a hybrid desiccant cooling system

for hot-humid and hot-dry climates. The system studied was similar to recirculation±condenser

cycle. In high sensible heat load, SHF was over 0.9 and in high latent load, SHF varied from

0.3 to 0.5. In high sensible heat load applications the energy savings ranged from 24 to 40%

for these two climates. However, they also found that the hybrid cycle saved more energy in a

hot and dry climate than it did in a hot and humid climate, where it may even use more energy

than a conventional system. Here, 1 kW of electrical energy was taken to be equivalent to 3.33

kW of thermal energy.

Maclaine-Cross [4] studied the feasibility of gas ®red hybrid desiccant cooling systems for

medium to large general air-conditioning projects. It was suggested that engine drive for

vapour compression plant could halve the energy costs for Australian conditions if waste heatwas recovered to regenerate the desiccant.

Worek and Moon [8] investigated the performance of a ®rst generation prototype desiccant

integrated hybrid system. The results showed that at the same level of dehumidi®cation, 60%

performance improvement over vapour compression system was obtained at ARI design

conditions. The performance of hybrid system decreased as the outdoor humidity ratio was

increased. Nevertheless, over the range investigated, the performance improvement varied from

74 to 44%. The performance improvement increased as outdoor temperature was increased

keeping humidity constant; over the range of outdoor temperature investigated the hybrid

system had a performance improvement of 20±80% over the conventional vapour compression

system.

Singh et al. [7] in a recent work have analyzed the afore-mentioned three hybrid cyclesfor Indian climatic conditions. Modelling of the dehumidi®er operating at a ®xed

regeneration temperature of 1358C, and regeneration to process air mass ¯ow rate of 0.33,

is done using the performance data from a manufacturer. It is reported that energy

savings ranging from 30% to 50% can be easily achieved at higher latent heat load

applications.

On the basis of literature review, it is evident that in high sensible load applications, large

energy savings are possible by using hybrid systems instead of conventional air conditioning

systems. But for high latent heat load application, there is inconsistency in the results of

dierent authors. Some have claimed that considerable energy saving would be possible but

others have concluded that a hybrid system might use more energy than a vapour compression

system. Further, most of the studies have been done at ARI design conditions which are very

``mild'' compared to the ambient conditions actually encountered in tropical climates. Singh et

al. [7] have studied the performance for such conditions, but have used a very simple empirical

model of the dehumidi®er which does not permit investigation of the in¯uence of basic design




parameters (like regeneration temperature, ratio of the mass ¯ow rates of process to

regeneration air, etc.) on the system performance.

Accordingly, an attempt has been made in the present work to investigate the suitability of

solid desiccant-based hybrid systems for air-conditioning applications in typical hot-dry and

hot-humid climatic conditions by incorporating a detailed model for simulating the

performance of the rotary desiccant wheel.

3. Computer modelling of hybrid systems

The most important component of a hybrid air-conditioning system using solid desiccants is

the rotary desiccant wheel. Its mathematical modelling needs solution of coupled partial

dierential equations, but Maclaine-Cross and Banks [3] showed that these equations can be

transformed into two sets of equations of the same form as those of a rotary sensible heat

regenerator. The performance of the rotary desiccant wheel has been determined by using this

analogy, following the formulation given by Jurinak [2]. The model has been validated over a

wide range of operating conditions by comparing the predictions of the computer programme

based on this method with the rated performance of two types of wheel manufactured in India.

A satisfactory agreement was observed, with the predicted outlet humidity values generallyexceeding the actual values by about 1 gm/kg da (for details, see [6]).

The performance of all other components (various heat exchangers and evaporative coolers)

has been predicted by assuming suitable values for their eectiveness. The condenser

temperature of the vapour compression system has been assumed to be 158C higher than the

ambient DBT, and, following Burns et al. [1], its COP has been taken as 46% of the COP of

the corresponding Carnot cycle between the condenser and evaporator temperatures. Besides

the afore-mentioned three cycles, the analysis has also been done for a modi®ed heat exchanger

cycle, shown in Fig. 6, where the exhaust from the air-conditioned space is used, instead of the

ambient air, for pre-cooling the dehumidi®ed air in the heat exchanger.

The analysis has been done for typical comfort conditions (258C DBT and 10 g/kg da

moisture content) with: hot-dry outdoor conditions of 43.28

C DBT and 7.26 g/kg da moisturecontent, and hot-humid conditions of 38.68C DBT and 16.14 g/kg moisture content; two

RSHF values of 0.35 and 0.75; and two values of ventilation air to supply air ratio: 0.1 and

0.2. The analysis presumes that the latent heat load has to be met by the dehumidi®er and the

Fig. 6. Schematic diagram of proposed cycle.




sensible heat load by the IEC and the vapour compression refrigeration system. A provision

has been made in the computer program that if one dehumidi®er is unable to lower the process

air speci®c humidity below room humidity, then additional dehumidi®ers are used with inter-

cooling in between them. Also, provision for direct evaporative cooling has been made

wherever necessary. This type of situation arises when the dehumidi®er exit humidity is low,

and the RSHF is high, as illustrated in Fig. 7. Here the supply air state needed to reach the

RSHF line only through sensible cooling (i.e., point 5) has a DBT lower than the dew pointtemperature, which is not physically possible. Therefore, a suitable check has been

incorporated in the computer program and a provision is made for direct evaporative cooling

(4±4 ') of the air leaving the cooling coil. The evaporator temperature is assumed to be 5 8C

higher than the dew point temperature and the state of air after the evaporator (4) is calculated

as usual from its eectiveness.

4. Results of analysis

The performance parameters at the design regeneration temperature of 1358C, for dierent

combinations of parameters mentioned above, are given in Tables 1 and 2 for hot-dry

conditions. The weighted energy consumption has been calculated by giving a weightage of 3to the electrical energy in comparison to the thermal energy. The following conclusions can be

drawn from these results:

1. For ventilation±condenser, ventilation±heat exchanger and modi®ed ventilation±heat

exchanger cycles, the supply air ¯ow rate required is very high compared to that in the

recirculation±condenser cycle and conventional systems using the vapour compression cycle

(VC cycle). In these cycles, only ventilation air is passed through the desiccant wheel, which

Fig. 7. Plot of recirculation condenser cycle at high sensible heat factor and summer design conditions.




is only 10% or 20% of the supply air. In contrast, in the recirculation±condenser cycle all

the supply air is passed through the desiccant wheel and hence a much larger amount of moisture is removed, even if the change in the air speci®c humidity in the dehumidi®er is

small.

2. Auxiliary heater input is the highest for the recirculation±condenser cycle because the ¯ow

rate of regeneration air is much more than that in other hybrid cycles.

3. As the SHF increases, the load on the desiccant wheel diminishes, and so the supply air ¯ow

rate reduces for all the cycles. However, there is a corresponding increase in the sensible

heat load which re¯ects in the increase in load on the refrigeration system.

4. Compressor work is lower for ventilation±condenser and recirculation±condenser cycles than

Table 1

Performance of hybrid and conventional VC cycles, at hot-dry conditions, mixing ratio of 0.1 and the design

regeneration temperature of 1358C

RSHF VC CYCLE VEN-CON VEN-HTX MOD.VEN-HTX RECIR-CON

CMMSUP 0.35 15.08 73.94 73.94 74.31 9.51

0.75 11.49 28.43 28.43 28.58 8.14

QHTR (kW) 0.35 3.262 2.962 2.77 2.966 3.81

0.75 0 1.139 1.066 1.1406 3.26

TEVAP (8C) 0.35 6 23.35 21.44 23.192 12.69

0.75 9.3 17.13 14.92 17.00 8.879

NO. OF DEHUM 0.35 ± 1 1 1 1

0.75 ± 1 1 1 1

WCOMP (kW) 0.35 2.90 0.4818 1.268 0.5817 0.5997

0.75 1.425 0.8829 1.288 0.9368 0.7092

W.E. CONS. (kW) 0.35 11.96 4.407 6.574 4.711 5.609

0.75 4.275 3.788 4.93 3.951 5.3876

Table 2

Performance of hybrid and conventional VC cycles, at hot-dry conditions, mixing ratio of 0.2 and the designregeneration temperature of 1358C

RHSF VC CYCLE VEN-CON VEN-HTX MOD.VEN-HTX RECIR-CON

CMMSUP 0.35 13.71 36.97 36.97 37.15 2.247

0.75 12.31 22.59 27.63 23.274 7.93

QHTR (kW) 0.35 2.629 2.96 2.77 2.96 3.704

0.75 0 1.81 2.0714 1.8574 3.1762

TEVAP (8C) 0.35 6.0 21.7 18.27 21.38 12.34

0.75 9.0 16.74 16.74 16.759 18.502

NO. OF DEHUM 0.35 ± 1 1 1 1

0.75 ± 1 1 1 1

WCOMP (kW) 0.35 2.7594 0.5076 1.381 0.615 0.6056

0.75 1.5518 0.7839 1.321 0.8504 0.71362

W.E. CONS. (kW) 0.35 10.91 4.483 6.913 4.807 5.5208

0.75 4.6553 4.1617 6.0344 4.408 5.317




the ventilation±heat exchanger and modi®ed ventilation±heat exchanger cycles. The reason is

that in the ®rst two cycles indirect evaporative cooling is employed while in the latter two, a

heat exchanger is used. In hot-dry summer, cooling potential of IEC is very high and so the

load on the refrigerated coil is reduced.

5. Weighted energy consumption is lowest for the ventilation±condenser cycle. It is due to the

lower heater input and lower compressor work.

6. Energy saving is highest in high latent heat load application (RSHF = 0.35); maximumsaving at mixing ratio of 0.1 is 63.15% and at mixing ratio of 0.2 is 58.9% with ventilation±

condenser cycle. On the other hand for modest latent loads characterized by a high value of

RSHF (0.75), the corresponding energy savings are 11.4 and 10.63%, respectively. Other

cycles show poorer performance than ventilation±condenser cycle under these hot-dry

conditions irrespective of the RSHF and the ventilation air to supply air ratio.

The predicted performance under hot-humid ambient conditions is given in Table 3. It can be

seen that the ventilation±condenser, ventilation±heat exchanger and modi®ed ventilation±Heat

exchanger cycles need two dehumidi®ers each to bring process air humidity below the room air

humidity. The recirculation±condenser cycle, however, needs only one dehumidi®er. It also has

the lowest weighted energy consumption of all the cycles with an energy saving (in comparison

to the conventional system) of 46.6% at high latent load conditions (RSHF = 0.35) whenventilation to supply air ratio is 20%.

At low latent heat load (RSHF = 0.75), however, the performance of all the hybrid cycles is

poorer than that of the conventional cycle, although the work of compression needed for

refrigerated cooling is substantially lower than in the vapour compression cycle. This is mainly

due to the fact that reheating is no longer needed in the conventional system, resulting in

drastic reduction in its total energy requirements.

Table 3

Performance of hybrid and conventional VC cycles, at hot-humid conditions, mixing ratio of 0.2 and the designregeneration temperature of 1358C

RSHF VC CYCLE VEN-CON VEN-HTX MOD.VEN-HTX RECIR-CON

CMMSUP 0.35 18.85 68.08 68.08 65.23 13.206

0.75 11.97 26.185 26.185 25.088 12.085

QHTR (kW) 0.35 3.962 11.54 11.66 11.15 5.596

0.75 0 4.4385 4.486 4.2887 5.1213

TEVAP (8C) 0.35 6.0 22.79 20.31 22.73 14.15

0.75 8.1 16.04 13.56 15.69 12.59

NO. OF DEHUM 0.35 ± 2 2 2 1

0.75 ± 2 2 2 1

WCOMP (kW) 0.35 3.588 0.6224 1.4356 0.6051 0.7571

0.75 1.7137 0.9097 1.3367 0.9103 0.79676

W.E. CONS. (kW) 0.35 14.728 13.407 15.967 12.96 7.867

0.75 5.141 7.1676 8.496 7.0196 7.51




5. In¯uence of wheel design parameters

As mentioned above, in many situations, for recirculation±condenser cycle, the dehumidi®er

outlet humidity was much lower than the room speci®c humidity, and therefore, direct

evaporative cooling had to be incorporated in the system under high RSHF conditions. This

clearly suggests that the regeneration temperatures should be lesser than the present design

value. Accordingly, the in¯uence of regeneration temperature on the performance of recirculation±condenser cycle has also been studied.

Very interesting results emerge from this study (Table 4). It is seen that under hot-dry

conditions the weighted energy consumption reduces drastically with reduction in regenerationtemperature. In fact, it seems possible to do wheel regeneration by directly using the air leaving

the condenser. Consequently, there is no regeneration heat requirement and the weighted

energy consumption decreases by 98% for high latent heat load (RSHF = 0.35) and by about

66% for low latent load (RSHF = 0.75) with 10% ventilation. Of course, energy reduction is

not without a price for the supply air ¯ow rate increases considerably, which necessitates a

larger dehumidi®er.

Under hot-humid conditions also, the weighted energy consumption decreases with reduction

in regeneration temperature but not so drastically (Table 5). At 20% ventilation rate there is

now an energy saving over vapour compression cycle of 38.2% at high latent load (RSHF =0.35). However, at low latent load (RSHF = 0.75) energy consumption of recirculation±

condenser cycle is still slightly higher than that of a conventional system using vapour

compression refrigeration (Table 5).

At this high ventilation ratio an interesting result appears. It is seen that the weighted energy

consumption falls as the regeneration temperature is reduced from 135 to 808C, but increases

(for high latent loads) when regeneration temperature is further reduced to 608C. This is clearly

Table 4

Performance of hybrid recirculation±condenser cycle at reduced regeneration temperatures and hot-dry conditions

for 10% ventilation air

RSHF VC CYCLE Regeneration temperature

1358C 808C 608C 43.28C

CMMSUP 0.35 15.08 9.51 12.79 17.44 27.43

0.75 11.49 8.14 10.22 12.82 16.87

QHTR (kW) 0.35 3.262 3.81 1.99 0.8063 0

0.75 0 3.26 1.588 0.592 0

TEVAP (8C) 0.37 6.0 12.69 17.265 19.86 22.42

0.75 9.3 8.879 12.28 14.25 15.97

NO. OF DEHUM 0.35 1 1 1 1

0.75 1 1 1 1

WCOMP (kW) 0.35 2.90 0.5997 0.362 0.256 0.0837

0.75 1.425 0.7092 0.548 0.512 0.485

W.E. CONS. (kW) 0.35 11.96 5.609 3.074 1.574 0.251

0.75 4.275 5.386 3.232 2.128 1.456




because the supply air ¯ow rate increases drastically as the regeneration temperature is reduced

from 80 to 608C (Table 5). It clearly brings out the need for optimisation of the regeneration

temperature.

6. Eect of changing other design parameters

Jurinak [2] has made some recommendations regarding the wheel operating parameters for

optimum performance of desiccant cooling system. To gauge the in¯uence of these

recommendations on the system performance, appropriate changes have been made in the

values of the process to regeneration air ¯ow ratio (changed from 3 to 1.25), corresponding

wheel face area ratios (changed from 3:1 to 1:1) and the wheel speed (changed to 9 rph). The

regeneration temperature has been taken as 808C for all the three cycles, and keeping in view

the afore-mentioned results, the recirculation±condenser cycle analysis has also been done for

regeneration temperature of 608C. Tables 6 and 7 present a summary of these results.

It can be seen that the performance of most of the cycles under these operating conditions is

better (both weighted energy consumption and supply ¯ow rate are lesser) than those under the

wheel design operating conditions (Tables 2 and 3). The only exception is the proposed cycle

which shows slight deterioration in performance under hot-dry conditions but an appreciableimprovement of performance under hot-humid conditions. It is also seen that in general, the

weighted energy consumption for recirculation±condenser cycle at both regeneration

temperatures (80 and 608C) is higher than that obtained at the same regeneration temperatures

when other design conditions of the wheel are kept unaltered. This is because of the fact that

the ratio of regeneration to process ¯ow rate now is 0.8 while under the design operating

conditions this ratio is 0.33. Consequently, even though the total supply air ¯ow rates are

Table 5

Performance of hybrid recirculation±condenser cycle at reduced regeneration temperatures, hot-humid conditions

and 20% ventilation

RSHF VC CYCLE Regeneration temperature

1358C 808C 608C

CMMSUP 0.35 18.86 13.2059 26.67 88.95

0.75 11.97 12.085 20.71 34.21

QHTR (kW) 0.35 3.962 5.596 4.843 6.55

0.75 0 5.1213 3.76 2.52

TEVAP (8C) 0.35 6.0 14.15 18.94 22.14

0.75 8.1 12.59 15.88 16.976

NO. OF DEHUM 0.35 1 1 1

0.75 1 1 1

WCOMP (kW) 0.35 3.588 0.7571 0.798 1.372

0.75 1.714 0.79646 0.9017 1.22

W.E. CONS. (kW) 0.35 14.728 7.867 7.24 10.66

0.75 5.141 7.51 6.46 6.18




reduced under new operating conditions, the regeneration side ¯ow rates remain higher. The

only notable exception is the case of 608C regeneration temperature, 20% ventilation and high

latent heat load (RSHF = 0.35). Here, this supply air ¯ow rate reduction is so large (from

88.95 to 25.12 cmm) that the regeneration side ¯ow rate is lower than that under the wheel

design conditions (Table 7). Consequently, the weighted power consumption here is 7 kW,

which is much less than the corresponding value of 10.66 kW with wheel design conditions

unaltered.

7. Concluding remarks

It is clear from this study that considerable energy saving can be achieved by using solid

Table 6

Performance of hybrid cycles at altered wheel operating conditions in hot-dry weather with 20% ventilation

RSHF VEN-CON VEN-HTX MOD.VEN-HTX RECIR-CON (808C) RECIR-CON (608C)

CMMSUP 0.35 36.367 36.37 39.71 7.306 9.67

0.75 22.25 27.12 24.08 6.154 8.01

QHTR (kW) 0.35 2.713 2.41 3.12 2.726 1.073

0.75 1.6597 1.7955 1.892 2.296 0.888

TEVAP (8C) 0.35 21.727 19.43 21.62 9.70 14.01

0.75 16.71 16.71 16.9 4.71 9.11

NO. OF DEHUM 0.35 1 1 1 1 1

0.75 1 1 1 1 1

WCOMP (kW) 0.35 0.4827 1.2536 0.6111 0.5868 0.4704

0.75 0.766 1.29 0.855 0.712 0.622

W.E. CONS. (kW) 0.35 4.161 6.168 4.95 4.486 2.4846

0.75 3.8598 5.66 4.45 4.432 2.754

Table 7Performance of hybrid cycles at altered wheel operating conditions in hot-humid weather with 20% ventilation

RSHF VEN-CON VEN-HTX MOD.VEN-HTX RECIR-CON (808C) RECIR-COND (608C)

CMMSUP 0.35 56.03 56.03 40.763 11.48 25.12

0.75 21.55 21.55 24.19 10.63 20.31

QHTR (kW) 0.35 9.766 9.395 7.215 5.00

0.75 3.756 3.61 4.28 4.63 4.443.59

TEVAP (8C) 0.35 22.4 21.064 21.80 12.5 18.36

0.75 14.2 12.86 16.96 11.16 15.7

NO. OF DEHUM 0.35 2 2 2 1 1

0.75 2 2 2 1 1

WCOMP (kW) 0.35 0.597 1.178 0.50148 0.7957 0.8567

0.75 0.4435 1.254 0.7431 0.82157 0.9403

W.E. CONS. (kW) 0.35 11.56 12.93 8.72 7.39 7.01

0.75 5.587 7.37 6.51 7.099 6.41




desiccant-based hybrid air-conditioning cycles instead of conventional systems using

refrigerated cooling coils alone, especially in hot-dry weather conditions. In hot-humid weather

conditions energy savings are possible only under high latent load conditions. However, there

is a great need to optimise the operating parameters of the desiccant wheel for getting the best

performance. The relative performance of various cycles depends very strongly on these

parameters. Thus, it is seen from the study that in hot-dry weather. The ventilation±condenser

cycle performs the best, consuming even lesser energy than a conventional system using vapourcompression cycle. In hot-humid weather, however, it is the ventilation±condenser cycle which

performs the best at high latent loads, but for low latent loads (RSHF = 0.75) again the

conventional system with refrigerated cooling coil performs better. This picture, however,

changes when we operate at reduced regeneration temperatures. The recirculation±condenser

cycle now performs better even in hot-dry weather. These results also give us a clue to the

reasons responsible for con¯icting statements made in the literature regarding ecacy of hybrid

system vis-a Á -vis vapour compression system. Most desiccant wheels being marketed today have

been optimized for dehydration duties, to achieve very low exit moisture content. The

requirements for hybrid air-conditioning applications being considerably dierent, the optimum

operating parameters for this application are bound to be quite dierent. Consequently, if a

desiccant wheel is operated at the usual design conditions (suitable for dehydration duties)

while being employed in a hybrid air-conditioning system, the performance would be quiteinferior to that obtained when the wheel's operating conditions are optimized. So a proper

comparison of the eectiveness of various cycles for given load and weather conditions

demands that proper optimisation of the wheel operating parameters be carried out. Maclaine-

Cross and Banks [3] analogy model used in this work seems to be quite suitable for such

optimisation.

References

[1] P.R. Burns, J.W. Mitchell, W.A. Beckman, Hybrid desiccant cooling systems in super market applications,

ASHRAE Trans 91 (Part-1B) (1985) 457±468.

[2] J.J. Jurinak, Open cycle solid desiccant cooling, component models and system simulations, Ph.D. thesis,

University of Wisconsin-Madison, 1982.

[3] I.L. Maclaine-Cross, P.J. Banks, Coupled heat and mass transfer in regenerators Ð predictions using an

analogy with heat transfer, Int. J. of Heat and Mass Transfer 15 (1972) 1225±1241.

[4] I.L. Maclaine-Cross, Hybrid desiccant cooling in Australia, Australian Refrigeration Air-conditioning and

Heating 41 (5) (1987) 16±25.

[5] J.C. Sheridan, J.W. Mitchell, A hybrid solid desiccant cooling system, Solar Energy 34 (2) (1985) 187±193.

[6] S.K. Singh, Some studies on solid desiccant-based hybrid air-conditioning systems, M. Tech Thesis, Department

of Mechanical Engg., I.I.T. Delhi, 1996.

[7] M. Singh, S. Jain, S.C. Kaushik, Energy conservation through hybrid air conditioning cycles: computer

modelling studies, unpublished paper, I.I.T. Delhi, 1996.

[8] W.M. Worek, C.J. Moon, Desiccant integrated hybrid vapour compression cooling Ð performance sensitivity to

outdoor conditions, Heat Recovery Systems and CHP 8 (6) (1988) 489±501.


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Studies on Solid Desiccant Based Hybrid Air-conditioning Systems

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