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S i m u l a t i o n o f a d e s i c c a n t - e v a p o r a t i v e c o o l i n gs y s t e m f o r r e s i d e n t i a l b u i l d i n g s

NRCC - 5 0 5 9 1

Haddad , K . ; Ouaz ia , B . ; Ba rhoun , H .

A version of this document is published in / Une version de ce document se trouve dans:3

rdCanadian Solar Buildings Conference, Fredericton, N.B., Aug. 20-22, 2008, pp. 1-8

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SIMULATION OF A DESICCANT-EVAPORATIVE COOLING SYSTEM FOR

RESIDENTIAL BUILDINGS

Kamel Haddad1, Boualem Ouazia2, and Hayssam Barhoun2

1CANMET Energy Technology Centre, Ottawa, Canada2

National Research Council, Ottawa, Canada

ABSTRACT

One technology that can help reduce the electricity

consumption of conventional air-conditioning

technology is the coupling of active dehumidification

with evaporative cooling. In this case sensible cooling

and moisture removal from indoor and outside

ventilation air are decoupled. In this study simulationmodels are developed for a conventional vapor

compression based cooling system and a desiccant

evaporative cooling system installed in an R-2000

house. Electricity consumption and comfort indices are

then predicted for the two systems for three regions of

the country with varying sensible heat ratios. It is found

that, compared to a conventional system, the desiccant

evaporative cooling system can lead to significant

electricity consumption reductions and also reduce the

number of hours when conditions inside the space are

uncomfortable.

INTRODUCTION

According to the Energy Use Data Handbook (Office

of Energy Efficiency, 2006), the total electricity

consumption for space cooling of residential buildings

in Canada has increased from 19.7 PJ in 1998 to 36.5

PJ in 2005. The relative increase in cooling electricity

consumption during this period is especially significant

for Ontario (82%) and Quebec (186%). In the year

2005 about 86% of the Canadian residential buildings

with a cooling system installed use a central air-

conditioning system while the rest use a room air-

conditioner instead. This rapid increase in electricity

consumption for space cooling is putting an even

greater burden on electric power plants in someprovinces. In Ontario summer air-conditioning has to

be met using fuel sources with high Green House Gas

emissions such as coal. There is then a need for air-

conditioning technology that can help reduce electricity

consumption for the purpose of energy conservation,

electric peak shaving, and GHG emissions reduction.

Conventional central air-conditioning systems based on

the vapor compression cycle have a rated Sensible Heat

Ratio (SHR) of around 75%. The operation of these

systems is controlled using only a thermostat and any

moisture removed from the space is a by product of the

temperature control. In humid climates the required

design latent load of the space to maintain comfortable

conditions can be high with a space design Sensible

Heat Ratio significantly less than 75%. In this case

using a conventional air-conditioning system leads to

uncomfortable conditions with high indoor humidity

levels. In addition, if the air-conditioning system is

oversized and the supply fan is in continuous mode,

part of the condensate in the drain pan of the cooling

coil will evaporate back into the air stream and finds its

way back inside the conditioned space (Henderson and

Rengarajan, 1996).

A technology that can help address the previous

shortcomings of conventional vapor compression

technology is based on coupling active desiccant

dehumidification with direct evaporative cooling.

There is no need for a compressor in this case and the

electricity used in this system is for pumping waterthrough the evaporative cooler and for pushing the air

around the system. The desiccant wheel can be

controlled independently using a humidistat that senses

the wet-bulb temperature of the space. A thermostat is

used to activate the evaporative cooler when there is a

need for space sensible cooling. In addition to the

potential energy savings and peak reductions, this

technology decouples the latent and sensible loads

making it possible to condition spaces with a wide

range of design Sensible Heat Ratios for better comfort

conditions.

Nelson et al. (1978) developed a simulation model

using TRNSYS software for a recirculation modedesiccant evaporative system. The second system they

studied based on ventilation mode where the

regeneration side of the desiccant wheel uses 100%

return air and the process side draws in 100% outside

air. Results are generated for Miami Florida for the

system based on ventilation mode to supply the cooling

load for a house. It is reported that up to 95% of the

regeneration heat of the desiccant wheel can be met

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from solar energy with 45 m2 collector area. The study

does not deal with electricity consumption of the

system.

Another noteworthy study is by Smith et al. (1993)

who developed a simulation model for a ventilation

mode Munters Environmental Control Cycle (MEC).

Simulation results were generated for three differentlocations: Pittsburgh (Pennsylvania), Macon (Georgia),

and Albuquerque (New Mexico). The results from this

study indicate that the desiccant system is able to meet

the cooling loads in all three locations. However, the

paper does not present specific data for electricity and

other energy consumption for the system. There is also

no specific information on the comfort level achieved

within the space associated with the use of the

proposed system. It is reported that in Pittsburgh about

73% of the regeneration heat needed for the desiccant

wheel comes from solar. This fraction falls to 18% for

Albuquerque. The authors indicate that the simulation

models used in their study can be improved by usingproduct specific performance data. This study does not

also provide any specific information on the electricity

consumption associated of the system

In the present study, simulations models are developed

for a conventional central vapor compression system

and a recirculation mode desiccant evaporative cooling

system. The house model connected to the two HVAC

systems is based on the characteristics of the test house

at the Canadian Centre for Housing Technology

(CCHT). Results are then generated for the electricity

and auxiliary energy consumption, and for comfort

indices for the two types of systems for three locationsin Canada. The results show that the proposed

desiccant evaporative cooling system can be effectively

used for electrical peak shaving and to improve

comfort conditions inside residential buildings. In

addition, solar energy can be used to provide a

substantial portion of the auxiliary thermal energy

needed to regenerate the active desiccant wheel.

SIMULATION TEST HOUSE

The results generated in the present study are obtained

using simulation model for the Canadian Centre for

Housing Technology (CCHT) test house shown in

Figure 1. This house is built to the R2000 standard withseveral of its characteristics listed in Table 1. Internal

sensible heat gains to the first floor zone from

occupants, lighting, refrigerator, stove, dishwater, and

other kitchen appliances add up to 5.3 kWh/day. For

the second floor the internal sensible heat gains from

occupants, lighting, and clothes washer and dryer add

up to 4.3 kWh/day. The two load profiles for these

gains are shown in Figure 2. It is assumed that these

gains are 50% radiative and 50% convective.

Figure 1. CCHT Test and Reference Houses

Table 1

CCHT house characteristics

FEATURE DETAILS

Livable Area 210 m2 (2260 ft2), 2 storeys

Insulation Attic: RSI 8.6, Walls: RSI 3.5, Rimjoists: RSI 3.5

Basement Poured concrete, full basement

Floor: Concrete slab, no insulation

Walls: RSI 3.5 in a framed wall.

Exposed floor

over the garage

RSI 4.4 with heated/cooled plenum air

space between insulation and sub-floor.

Windows Low-e coated, argon filled windows

Area: 35.0 m2 (377 ft2) total, 16.2 m2

(174 ft2) South Facing

Airtightness 1.5 air changes per hour @ 50 Pa (1.0

lb/ft2)

Sensible Internal Heat Gains

0.00

0.20

0.40

0.60

0.80

1.00

1.20

1.40

0.00 5.00 10.00 15.00 20.00

Hour

HeatGain(kWh

Second Floor First Floor

Figure 2. First and second floors internal sensible heat

gains

The simulated occupancy is for two adults and two

children with a total latent internal heat gain for the

first and second floors of 0.47 and 0.66 kWh/day,

respectively. The latent heat gain from the stove is

assumed to be 40% of the total heat gain from this

appliance at 0.66 kWh/day. In addition to these internal

latent heat gains, there is a latent load associated with

the introduction of outside ventilation air.

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The house simulation model has zones for the

basement, first floor, second floor, garage, and attic

space. It is assumed that 40% of the total infiltration is

attributed to the basement, 20% to the first floor, 20%

to the second floor, and 20% to the attic. The house

HVAC system responds to a thermostat on the first

floor to maintain the dry-bulb temperature at 250.5

C. But it is assumed that the first and second floors are

fully mixed through the use of a high ventilation flow

rate between the two zones representing these two

floors. The basement, attic, and garage are not cooled

by the HVAC system.

REFERENCE CONVENTIONAL

CENTRAL VAPOR COMPRESSION AIR-

CONDITIONING SYSTEM

Description

Figure 3 shows the layout for a conventional residential

unitary-split air-conditioning system consisting of an

outdoor condenser, condenser fan, indoor evaporator,

compressor, and expansion valve. Electricity

consumption in this system is due to the operation of

the outdoor fan, indoor fan, and compressor. The

outdoor air fraction of the total supply air is 10%.

Figure 3. Conventional unitary split air-conditioning

system

Table 2 shows the various parameters that characterize

this system using performance data for a commercial

product documented by Neymark and Judkoff (2002).

The data in Table 2 is at A.R.I. rating temperatures of

To,db = 35 C, Ti,db = 26.7 C, and Tin,wb = 19.4 C. The

system total supply flow rate is 425 L/s which are

equally split between the first and second floor. The

supply fan cycles on and off to maintain the house at

250.5 C. All the power input to the fan motor goes

toward heating the air flow.

Table 2

Characteristics of conventional vapor compression

system

VARIABLE VALUE

Total Cooling Capacity 8174 W

Compressor Power 1860 W

Condenser Fan Power 108 W

Indoor Fan Power 154 W

C.O.P. 4.15

Sensible Heat Ratio 0.75

Simulation Model

The simulation model used accounts for the variation

of the total cooling capacity and the compressor power

input of the air-conditioner with the inlet evaporator

coil temperature and outdoor dry-bulb temperature. In

addition, the sensible heat ratio of the direct expansion

coil is a function of the inlet wet-bulb and inlet dry-

bulb temperatures to the coil. The model predicted

electricity consumption is corrected for part-load

degradation of the system.

SOLAR ASSISTED DESICCANT-

EVAPORATIVE COOLING SYSTEM

Description

The proposed recirculation cycle desiccant-evaporative

cooling system is shown in Figure 4. Return air is

mixed with outside air and then dehumidified in an n

active rotary desiccant wheel. Warm supply air from

this wheel is then cooled through a sensible heat

exchanger before entering a direct evaporative cooler

and then delivered to the conditioned zones. On the

regeneration side outside air is first cooled through adirect evaporative cooler and then used to cool the

supply air in the sensible heat exchanger. This air is

then heated to regenerate the desiccant wheel.

Figure 4. Recirculation cycle desiccant evaporative

cooling system using solar energy

The exhaust regeneration air from the desiccant wheel

is used to preheat the regeneration air at the inlet of the

wheel. This air can be further heated through a fan coil

using warm water from a solar storage tank. An inline

heater is used to provide any additional heat needed to

get the desired air temperature at the inlet of the

Direct

evaporative

cooler

Outside

air

Supply

air

Direct

evaporative

cooler

Sensible

wheel

Heater

Desiccant

wheelReturn

air

Fresh

air

Fan coil

Sensible

wheel

Water from

solar tankWater to

solar tank

Ambient

air inlet

Ambient

air exhaust

Compressor

fan

Condenser

coil

Supply air

Indoor space

Return air

Supply air duct

Fresh air

Evaporator

coil

Indoor fan

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desiccant wheel on the regeneration side. A schematic

of the solar system is shown in Figure 5. The collector

performance data used in the simulation model is based

on actual performance data for a commercially

available solar collector.

Figure 5. Solar collector and storage tank system

All the desiccant system components are cycled on and

off to maintain the space wet-bulb temperature at

180.5 C. In addition, the process side fan and

evaporative cooler are also cycled on and off to keep

the space dry-bulb temperature at 25 C. In effect, this

control strategy decouples the sensible and latent

cooling functions of the system.

Simulation Models

The heat exchangers of the system (sensible wheels and

fan coil) are modeled using a constant effectiveness

approach where the total heat transfer is given by:

( coldinhotinp TTcmQ ,,min)( = && ) (1)

Similarly a constant effectiveness is assumed for the

direct evaporative cooler:

inwbin

inout

TT

TT

,

= (2)

An empirical model is used for the desiccant wheel

based on actual performance data for a commercially

available component. The model is based on two

correlations for the outlet regeneration temperature and

outlet regeneration humidity ratio:

inr

inrTinrTinrT

inp

inp

TinpTinpTToutrT

wawaTa

T

wawaTaaT

,

,7,6,5

,

,4,3,21, ++++++=

(3)

inr

inrwinrwinrw

inp

inp

winpwinpwwoutrT

wawaTa

T

wawaTaaw

,

,7,6,5

,

,

4,3,21, ++++++=

(4)

It is found that these correlations fit the actual

performance data of the desiccant wheel very well with

R2 > 97%. Once the humidity ratio at the exit of the

regeneration side is determined from Equation 4 using

inlet conditions to the wheel, a moisture mass balance

is used to find the humidity ratio at the exit of the

process side. The enthalpy at the regeneration side

outlet is based on temperature and humidity ratio

values from Equations 3 and 4, respectively. An energy

balance on the wheel is then used to find the enthalpy

at the outlet of the process side. As a result, two

properties are now known for moist air at the exit of

the process side.

A stratified tank model is used for the solar system

storage tank with a total of 15 stratified layers. Hot

water from the solar collector enters at the top of thetank while cold water from the fan coil enters at the

bottom of the tank. At the same time cold water from

the bottom of the tank is delivered to the solar

collectors and hot water from the top of the tank is

delivered to the fan coil. The tank sits in the basement

and skin losses are accounted for as internal gains to

this zone.

The solar collector model predicts the thermal

performance of a flat plate solar collector. The model is

based on the collector thermal efficiency equation:

t

oincoll

G

TT +=

,

10 (5)

The model corrects the efficiency for non-normal

incident solar radiation and for any difference between

actual operating flow rate and that used at rating

conditions.

A summary of the parameters used in the simulations

of the desiccant evaporative cooling system is shown in

Table 3.

ASHRAE COOLING COMFORT ZONE

The ASHRAE comfort zone for cooling, as specified in

the ASHRAE Handbook of Fundamentals (1997), isbounded at the top by the constant wet-bulb

temperature line Twb of 20 C and the bottom with the

constant dew-point temperature line Tdp of 2 C. The

maximum allowable operative temperature on the x-

axis is 27 C at Tdp = 2 C and 26 C at Twb = 20 C.

The operative temperature of 25 C and a relative

humidity of 50% (Twb = 18 C) are right at the centre of

the summer comfort zone. In this study it is assumed

that discomfort inside the occupied space occurs when

either of the following conditions is true:

Tdb > 26.5 C

(Twb > 18.5 C) and (Top > 23 C)

(Tdp < 2 C) and (Top > 23 C)

Storage Tank

Collector Array

PumpPump

HX

Pump

Toward

fan coil

From

fan coil

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Table 3

Characteristic parameters for desiccant evaporative

cooling systemCOMPONENT VARIABLE VALUE

Direct evaporative cooler Saturation efficiency 0.9

Circulation fans Motor efficiency 0.8

Air-to-air sensible HX Effectiveness 0.7

Effectiveness 0.6Flow rate (kg/s) 0.227

Fan coil

Pump power (W) 75

0 0.694

1 -4.85

Slope 45

Flow rate (kg/s-m2) 0.0066

Area (m2) 14.4

Solar collector

Pump Ton; Toff(C) 5.5; 1.0

Volume (Liters/m2

collector area)

75Solar tank

Blanket U-value (W/m2-

K)

0.83

Collector side pump

power (W)

50

Tank side pump power

(W)

25

Flow rate (kg/s-m2) 0.0066

Collector and

solar tank pumps

Fraction of power

transferred as heat

1.

Desiccant wheel Motor power (W) 10

Solar system liquid-to-

liquid HX

Effectiveness 0.8

SYSTEM PRESSURE DROPS

A detailed comparison of the performance of the

conventional vapor compression system to the

desiccant evaporative cooling system requires a good

estimate of the pressure drops of the two systems.Tables 4 and 5 list pressure drop data used in the

simulations for various components on the regeneration

and process sides, respectively, of the desiccant

evaporative cooling system. All of the pressure drop

data, except for the house pressure drop, reported in the

Tables is based on actual performance data of

commercially available products. The house pressure

drop is based on actual measurements taken at the

CCHT. The fan power required to overcome the

pressure drop is based on the following equation:

fan

PVPower

=

&

(6)

Where a fan efficiency of 80% is assumed.

For the conventional system, the supply fan has to

overcome flow resistances from the existing house and

duct system (144 Pa), a filter (83 Pa), and DX coil (62

Pa) for a total fan power requirement of 154 W.

Table 4

Component pressure drops on regeneration side of

desiccant evaporative cooling system

COMPONENT PRESSURE

DROP (Pa)

POWER

(W)

Direct evaporative cooler 53 28

Sensible air-to-air HX 52 28

Sensible air-to-air HX 52 28Fan coil 67 36

Desiccant wheel

WSG550x200

350 186

Sensible air-to-air HX 52 28

Total: 626 334

Table 5

Component pressure drops on process side of desiccant

evaporative cooling system

COMPONENT PRESSURE

DROP (Pa)

POWER

(W)

Filter 83 44

Desiccant wheel 350 186

Sensible air-to-air HX 52 28

Direct evaporative cooler 53 28

Existing duct and house

pressure drop

144 76

Total: 682 362

RESULTS AND DISCUSSION

The simulation models are used to generate results for

Halifax, Ottawa, and Calgary using actual weather data

files for the year 2001 for the period June 1st August

31st. These locations are chosen because they represent

a good diversity in the fraction of the total load that is

sensible. All the results are generated using a 5-minute

time step.

Total Cooling Demand using Ideal Controller

Table 6 shows the sensible, latent, and total cooling

demands predicted for the three cities using an ideal

controller to maintain the first and second floor dry-

bulb temperature below 25 C and relative humidity

below 50%. These cooling demands do not include the

contribution of the ventilation air and circulation fan

heat toward the space cooling requirements. As

expected, for a location such as Halifax, the latent

loads represent a larger portion of the total cooling

load. This trend is reversed in the dryer climate of

Calgary.Table 6

Cooling demand using ideal controller for CCHT

houseCITY

sQ

(kWh)

lQ

(kWh)

tQ

(kWh)

SHRHalifax 695.4 226.9 922.3 0.754

Ottawa 1202.5 232.3 1434.8 0.838

Calgary 620.2 39.3 659.5 0.94

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Conventional System Results

Figure 6 shows the variation of the second floor

temperature (Operative, Dew-point, and Wet-bulb) for

Ottawa when the house is cooled using a conventional

air-conditioning system. In this case the wet-bulb

temperature of the floor is continuously changing with

the control of the dry-bulb temperature. The threetemperatures in Figure 6 for the second floor are very

close to those of the first floor due to the high mixing

simulated between the two floors. The mean radiant

temperature of the second floor is higher because of the

high solar gains of this floor.

CCHT Second Floor Temperatures (Ottawa June-August 2001)

0

5

10

15

20

25

30

3624 3792 3960 4128 4296 4464 4632 4800 4968 5136 5304 5472 5640 5808

Hour

Temperature(DegC

Operative Dew-Point Wet-Bulb

Figure 6. Second floor temperature variation with

conventional system for Ottawa

Table 7 lists the electricity consumption of the

compressor, outdoor fan, and indoor circulation fan of

the cooling system. Table 8 shows the predicted total

hours of discomfort for the second floor of the house.

The number of hours of discomfort for the first floor islower because of the lower solar gains and the cooling

effect from the basement. As the total cooling demand

(sensible and latent) increases, the number of hours

with uncomfortable conditions inside the space

increases as expected.

Table 7

Electricity consumption of vapor compression systemLocation Ec

(kWh)

Efan,i(kWh)

Efan,o(kWh)

Et(kWh)

Ottawa 331 22 30 383

Halifax 173 12 16 201

Calgary 143 10 13 166

Table 8Total hours of discomfort for second floor using vapor

compression system

Locatio

n

NhoursTdb > 26.5 C

NhoursTwb > 18.5 C

and

Top > 23 C

Total Hours

of

Discomfort

Ottawa 18 137 155

Halifax 4 165 169

Calgary 3 0 3

Desiccant-Evaporative Cooling System Results

Figure 7 shows the variation of the second floor

temperature when conditioned with a desiccant cooling

system. In this case the controlled wet-bulb

temperature is maintained around the set point of 18

C. In contrast to the conventional system, in the case

of the desiccant cooling system, activedehumidification is activated only when the wet-bulb

temperature reaches a maximum threshold.

CCHT First Floor Temperatures (Ottawa June-August 2001)

0

5

10

15

20

25

30

3 62 4 3 79 2 3 96 0 4 12 8 4 29 6 4 46 4 4 63 2 4 80 0 4 96 8 5 13 6 5 30 4 5 47 2 5 64 0 5 80 8

Hour

Temperature(DegC)

O perative D ew -Point W et -Bulb

Figure 7. Second floor temperature variation for

Ottawa with desiccant evaporative cooling system for

Ottawa

Table 9 shows the electricity consumptions of the

process side and regeneration side fans. The Table also

shows the total electricity and electric auxiliary energy

consumptions. These results are obtained with an inlet

regeneration temperature of 70 C. The results show

that the total electricity consumption Et for Ottawa,

Halifax, and Calgary decrease by 22.7, 30.8, and

53.6%, respectively, compared to the conventional

cooling system. These results show that the desiccant

evaporative cooling system has very good potential in

reducing the peak electrical load of residential air-

conditioning compared to a conventional system. The

relative reduction in the electricity consumption of the

desiccant system improves in locations with either high

or low latent loads such as Halifax and Calgary,

respectively.

Table 9

Electricity consumption of desiccant evaporative

cooling system without solar energy (Tr,in=70 C)City E

fan,p(kWh)

Efan,r(kWh)

Ewheel(kWh)

Et(kWh)

Qother(kWh)

Ottawa 165 127 4.3 296 5305

Halifax 80 57 1.9 139 2371

Calgary 53 23 0.8 77 974

A high latent load degrades the performance of a DX

coil of a conventional system, which then consumes

even more electricity to achieve a certain sensible

cooling load. While a low latent load indicates that

most of the cooling load of the space can be met by

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using mostly the evaporative cooler of the desiccant

system, which operates on much less electricity than a

compressor-based system.

The downside of the desiccant system is that auxiliary

thermal energy has to be supplied to heat the air at the

inlet of the regeneration side of the desiccant wheel.

Ideally this auxiliary energy would come from a freesource such as waste heat. Alternatively, a solar system

using solar collectors and a water storage tank (Figure

5) can be used to provide most of the thermal energy

needed to heat the regeneration air of the desiccant

wheel.

Table 10 shows the electricity and auxiliary energy

consumptions of the desiccant system using a 14.4 m2

solar collector system. In this case the total electricity

consumption of the system Et increases significantly

due to the operation of the circulation pumps of the

solar system. In the case of Ottawa and Halifax there is

now a slight increase in electricity consumption

compared to a conventional system. A modest decrease

in electricity consumption is predicted for Calgary.

These results indicate then that the use of a solar

system, of a similar size to the one simulated in this

study, can potentially wipe out any benefit of the

desiccant cooling system as a peak shaving or energy

saving technology. Under such circumstances it can not

be justified to operate the desiccant system using the

solar system.

Table 10

Electricity and auxiliary energy consumption of

desiccant evaporative cooling system with solar energy

(Tr,in=70 C)City Efan,p(kWh)

Efan,r(kWh)

Epumps(kWh)

Et(kWh)

Qother(kWh)

Ottawa 178 153 94 429 3287

Halifax 91 71 68 232 997

Calgary 62 31 53 147 95

However, it is important to keep in mind that the COP

of the reference vapor compression cycle is previously

assumed to be 4.15, which is more typical of highly

efficient air-conditioners. If we assume a more

moderate COP for the conventional system, the

electricity reduction potential of the desiccant system

will be even greater.

If a solar assisted desiccant cooling system is installedin a house, it is expected that it will be used to satisfy

part of the energy input needed for space heating,

domestic hot water, and space cooling. So a complete

life cycle cost analysis of the desiccant evaporative

cooling system using solar energy has to account for all

the functions that the solar system can be used for in

the home.

Table 11 shows the predicted total number of hours of

discomfort for the second floor with the desiccant

system installed. For Ottawa and Halifax the desiccant

system has less number of hours of discomfort

compared to the conventional system. Most importantly

the results for Halifax show a very large decrease in the

number of hours of discomfort from 169 to 32. For

Calgary there is actually an increase in the number of

hours but the total number is still very small relative to

the duration of the cooling season. It is evident that the

desiccant system is especially suited for areas where

the latent loads are high.

Table 11

Total hours of discomfort for second floor using

desiccant evaporative cooling systemLocatio

n

NhoursTdb > 26.5 C

NhoursTwb > 18.5 C

and

Top > 23 C

Total

Hours of

Discomfort

Ottawa 10 129 129

Halifax 2 31 32

Calgary 5 8 13

One potential promising improvement to the desiccant

system in Figure 4 is the replacement of direct

evaporative coolers with indirect ones instead. This

modification to the system can help minimize the

moisture gain through the system and improve energy

and comfort performance. Such a system is currently

being tested at the Canadian Centre for Housing

Technology. Preliminary results from experimental and

simulation work on the desiccant system with indirect

evaporative cooling show an even greater potential for

electricity peak shaving and improving comfort level.

CONCLUSIONS

Detailed component based simulation models are

generated for a conventional system and a desiccant

evaporative cooling system, with and without solar

energy, are developed and then used to assess the

potential of the desiccant system for electric peak

shaving and improved indoor comfort. Based on the

results generated in this study it possible to make the

following conclusions:

When the desiccant evaporative cooling system is

used without solar energy, there is a significant

reduction in electricity consumption associated with

residential space cooling (30% for Halifax and 53%

for Calgary). The relative reduction in electricity

usage, compared to a conventional system,

increases for areas with either a high or low

Sensible Heat Ratio.

The desiccant evaporative cooling system has

significant potential in reducing the occurrence of

uncomfortable high indoor humidity levels that tend

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to occur in humid parts of the country such as

Halifax.

The use of solar energy for regeneration of the

desiccant wheel can provide a significant portion of

the auxiliary thermal energy needed. However, the

operation of the solar system increases the

electricity consumption of the cooling system,which can significantly reduce or completely

eliminate the potential of the desiccant system as a

peak shaving technology.

The desiccant system presented in this study can be

further enhanced through the use of indirect

evaporative coolers, which reduces the moisture

gain to the supply air. Such system is currently

being tested at the Canadian Centre for Housing

Technology. Preliminary experimental and

simulation results show that this system has an even

greater potential for peak shaving and improved

indoor comfort.

NOMENCLATURE

a Correlation coefficient

cp Fluid specific heat (J/kg-C)

E Electricity consumption (kWh)

P Component pressure drop (Pa)

Effectiveness

Collector efficiency

fan Fan motor efficiency

Gt Total solar radiation reaching collector (W/m2)

m& Fluid mass flow rate (kg/s)

Nhours Number of hours

Q&

Heat transfer rate (W)Q Cooling demand (kWh)

SHR Sensible heat ratio

T Temperature (C)

V& Volume flow rate (m3/s)

w Humidity ratio (kg/kg)

Subscripts

c Compressor

coll Solar collector

cold Cold side of heat exchanger

db Dry-bulb

dp Dew-point

hot Warm side of heat exchangerl Latent

i Inside

in Inlet

o Outdoors (ambient)

op Operative

other Non electric

out Outlet

p Process side

r Regeneration side

s Sensible

t Total

T Temperature

w Humidity ratio

wb Wet-bulb

wheel Desiccant wheel

REFERENCES

ASHRAE Handbook, Fundamentals Volume,

American Society of Heating Refrigerating and

Air-Conditioning Engineers Inc., Atlanta, GA.,

1997

Henderson, H.I., Rengarajan, K. 1996. A Model to

Predict the Latent Capaciy of Air-Conditioners and

Heat Pumps at Part-Load Conditions with Constant

Fan Operation. ASHRAE Transactions 102(1):

266:274.

Nelson, J., Beckman, W.A., Mitchel, J.W., Close, D.J.1978. Simulation of the Performance of Open

Cycle Desiccant Systems Using Solar Energy.

Solar Energy, Vol. 21, pp. 273:278.

Neymark, J., Judkoff, R. 2002. International Energy

Agency Building Energy Simulation Test and

Diagnosis Method for Heating, Ventilating, and

Air-Conditioning Equipment Models (HVAC

BESTEST), Volume 1: Cases E100-E200.

Office of Energy Efficiency, 2006. Energy Use

Handbook. Natural Resources Canada.

Smith, R.R., Hwang, C.C., and Dougall. R.S. 1993.Modeling of a Solar-Assisted Desiccant Air-

Conditioner for a Residential Building.

Thermodynamics and the Design, Analysis, and

Improvement of Energy Systems, ASME.

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