Applied Energy 88 (2011) 1354–1365

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Applied Energy

journal homepage: www.elsevier .com/ locate/apenergy

Desiccant wheel regenerated by thermal energy from a microcogenerator:Experimental assessment of the performances

Giovanni Angrisani a, Alfonso Capozzoli b, Francesco Minichiello c,⇑, Carlo Roselli a, Maurizio Sasso a

a DING, Università degli Studi del Sannio, Piazza Roma 21, 82100 Benevento, Italyb TEBE Research Group, DENER, Politecnico di Torino, Corso Duca degli Abruzzi 24, 10129 Torino, Italyc DETEC, Università di Napoli Federico II, P.le Tecchio 80, 80125 Napoli, Italy

a r t i c l e i n f o a b s t r a c t

Article history:Received 19 April 2010Received in revised form 25 August 2010Accepted 26 September 2010Available online 29 October 2010

Keywords:Desiccant wheelHVAC systemMCHPExperimental analysisLatent loadsRegeneration temperature

0306-2619/$ - see front matter � 2010 Elsevier Ltd. Adoi:10.1016/j.apenergy.2010.09.025

⇑ Corresponding author. Tel.: +39 081 2538665; faxE-mail addresses: giovanni.angrisani@unisannio.it

zoli@polito.it (A. Capozzoli), minichie@unina.it (Funisannio.it (C. Roselli), sasso@unisannio.it (M. Sasso

Hybrid desiccant HVAC systems have shown several advantages, compared to conventional cooling anddehumidification systems. Therefore, their use is also spreading for tertiary and residential buildings,especially when the regeneration of the desiccant can be obtained by using available waste heat.

In this paper, an experimental analysis is presented on the performances of a silica-gel desiccant wheel,inserted in a test facility characterized by an advanced desiccant air handling unit, coupled to an electricchiller, a natural gas-fired boiler and a small scale cogenerator. The desiccant wheel is regenerated byusing low temperature thermal energy recovered from the microcogenerator.

The effects of the main thermal-hygrometric parameters (outdoor air humidity ratio and temperature,regeneration air temperature) on the desiccant wheel performances have been experimentally evaluated;in particular, the thermal-hygrometric properties of the process air exiting the rotor and the desiccantwheel effectiveness values have been obtained. Finally, fixing the regeneration temperature at its maxi-mum available value (65 �C), ventilation and internal latent loads that the desiccant wheel can handlehave been evaluated and compared to the required values, both for a set of cities all over the worldand as a function of the thermal-hygrometric outdoor conditions.

� 2010 Elsevier Ltd. All rights reserved.

1. Introduction

Desiccant cooling is based on air dehumidification by means ofa desiccant substance (solid or liquid), and on its subsequent cool-ing. In fact, the dehumidification process is exothermic and deter-mines a sensible heating of the airflow being dehumidified, whichin turn has to be cooled to the thermal-hygrometric conditionsnecessary to handle the thermal loads [1–3].

To obtain a continuous operation of such a system, the desic-cant material has to be periodically regenerated by means of ahot airflow, in order to evaporate the absorbed/adsorbed water va-pour. Regeneration temperature depends on the desiccant mate-rial, but it can be low enough for the last generation ofdesiccants (60–70 �C is often sufficient), thus allowing the use ofwaste heat from cogeneration devices or energy recovered from asolar collector.

Therefore, a desiccant cooling system comprises three main ele-ments (Fig. 1): the dehumidifier, the regeneration thermal energysource and the cooling device.

ll rights reserved.

: +39 081 2390364.(G. Angrisani), alfonso.capoz-. Minichiello), carlo.roselli@).

As regards the dehumidifier, one of the most common configu-rations is the desiccant wheel (DW) that is a rotor filled with a soliddesiccant material, which slowly rotates between the process air tobe dehumidified and the regeneration air [4].

With reference to the thermal energy source for regeneration,the desiccant wheel can obviously be regenerated by a gas-firedboiler or an electric resistance when waste heat is either not avail-able or not sufficient (in terms of power or temperature); this oc-curs in particular for high regeneration temperature (above 80 �C).

As regards the cooling device, it can be a direct or indirect evap-orative cooler, an air-to-air heat pump (in cooling mode), or an air-cooled water chiller.

The obtained air conditioning system can be called desiccant-based (or hybrid) heating, ventilation and air conditioning (HVAC)system.

Desiccant HVAC systems have the following main advantagescompared to conventional cooling-based dehumidification systems[5,6]:

(1) they can separately control sensible and latent load;(2) the cooling machine only has to cool the process air (without

dehumidifying it), so it can operate with a higher chilledwater temperature compared to a cooling machine coupledto a conventional cooling and dehumidification system,

Nomenclature

h specific enthalpy (kJ/kg)L specific ventilation latent load [W/(m3/h)]T temperature (�C)g effectiveness (–)Dhvs latent heat of vaporization for water (kJ/kg)x air humidity ratio (g/kg)DT temperature variation (�C)Dx air humidity ratio variation (g/kg)AHU air handling unitCOP coefficient of performanceDB–MCWB dry bulb–mean coincident wet bulb (�C)DP–MCDB dew point–mean coincident dry bulb (�C)DW desiccant wheelHVAC heating, ventilation and air conditioning

MCHP Micro Combined Heat and PowerSHR sensible heat ratio, i.e. the ratio between the sensible

and total thermal load

Subscriptsad adiabaticdeh dehumidificationDW referred to desiccant wheelo outdoor airr roomreg regenerations supplyth thermal

G. Angrisani et al. / Applied Energy 88 (2011) 1354–1365 1355

where the process air has to be cooled below the dew pointtemperature to obtain the desired humidity ratio. Hence, thecooling machine interacting with the hybrid HVAC systemhas a higher COP;

(3) due to the higher value of COP, electric energy requirementof the cooling machine is reduced;

(4) as the cooling machine only has to cool the process air with-out dehumidifying it, a reduction of the chiller size and con-sequently of the refrigerant fluid mass is obtained, and so alower environmental impact, both in terms of direct impact(ozone layer reduction and greenhouse effect due to refriger-ant fluids) and of indirect one (the reduced electric energyuse determines lower equivalent CO2 emissions of the powerplants);

(5) in the conventional cooling and dehumidification process,air commonly has to be re-heated (for example, by a gas-fired boiler or an electric resistance), in order to reach thedesired supply temperature, with a further primary energyconsumption. On the contrary, the re-heating process isnot required in the desiccant dehumidification;

(6) they can give very low dew point temperatures of processair, lower than �6.0 �C, while conventional systems usuallydo not reach dew point temperatures lower than about4.0 �C. So, the desiccant dehumidification technology is par-ticularly used when very dry air is needed in the operatingprocesses, such as in the chemical, pharmaceutical and foodindustries, or when a very low indoor humidity ratio isneeded in order to preserve or manipulate hygroscopic orhumidity sensitive materials;

(7) a better indoor air quality can be obtained too, due to sani-tizing effects of desiccant [7,8].

Thanks to these advantages, the use of desiccant technology isalso spreading for tertiary and residential buildings. On the other

Cooling device Dry air Desiccant dehumidifier

Regeneration heat source

Process air

Moisture out

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Heat

supply

Fig. 1. Simplified scheme of a desiccant cooling system.

hand, it is necessary to highlight both the low familiarity with suchtechnology and the frequent lack of knowledge about perfor-mances and cost/benefit ratio.

As a result of the importance of desiccant wheel in solid desic-cant cooling systems, many investigations have been performed onits design, modeling and optimization.

A mathematical modeling of desiccant wheel using a commer-cial software for solving mass and heat transfer equations has beendeveloped by Esfandiari Nia et al. [9]. The model is used to obtainsimple correlations for calculating humidity and temperature ofthe air exiting the wheel as a function of the physically measurableinput variables and to determine the optimal rotational speed.

Jia et al. [10,11] developed a mathematical model to predict theperformance of a rotary solid desiccant cooling system using a novelcompound desiccant material; this model was then validated bymeans of experimental tests. Stabat et al. [12] developed a desiccantwheel model to be adapted to building simulation code, which wasin good agreement with the experimental and manufacturer data.Hamed Ahmed et al. [13] developed a mathematical model to eval-uate the effect of the operating conditions on the performance of arotating dehumidification system using a liquid desiccant. The ef-fects of regeneration air temperature, process air and regenerationair inlet humidity ratio, rotational speed, process and regenerationair velocity on the amount of water absorbed are investigated. Xionget al. [14] investigated a novel two-stage liquid desiccant dehumid-ification system using an exergy analysis method. The proposed sys-tem is characterized by higher coefficient of performance and exergyefficiency compared to a basic liquid desiccant dehumidificationsystem. The transient and steady state transport phenomena in theDW have also been analyzed by Z. Gao et al., who developed a math-ematical model based on the one-dimensional Navier–Stokes equa-tions [15]. The model evaluates the humidity ratio and temperaturein both the airflow channels as a function of time. The predicted re-sults are validated by means of data taken from experimental results.

Beccali et al. [16,17] presented some models to evaluate theperformance of rotary desiccant wheels composed of different so-lid desiccant materials. The models were derived from the interpo-lation of experimental data obtained from the industry. Somecorrelations were developed for predicting outlet temperatureand humidity ratio.

In some cases, the performances of the DW were investigatedonly by means of experimental laboratory tests. For example,Mandegari and Pahlalavanzadeh [18] presented an experimentalstudy considering different climates (hot dry and hot humid) andvarious operating conditions, in terms of regeneration temperatureand wheel speed. The desiccant wheel effectiveness values in eachoperating condition were calculated and a new definition for effec-tiveness was introduced.

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Yao et al. [19] investigated a new silica-gel regeneration meth-od using power ultrasound. The experimental study proves thatthe proposed method can help to improve the regeneration effi-ciency and reduce regeneration energy.

In terms of thermal energy source to regenerate the DW, greatattention is taken in the literature to solar energy. The evaluationand optimization of a solar assisted DW performance have beencarried out by Fouad et al. [20]. They developed a numerical modelto study the effect of parameters such as wheel thickness, speedand porosity, regeneration to adsorption area ratio, process airflowrate and humidity ratio and regeneration air temperature. A solarair collector is used as thermal energy source to regenerate thedesiccant material.

Vitte et al. [21] showed that a correctly controlled solar desic-cant cooling system can allow interesting energy savings in build-ing air conditioning. The influence of indoor and outdoor airconditions is characterized using numerical simulations, adoptinga validated model of the desiccant wheel.

A numerical and experimental study of a solar assisted desic-cant cooling system for air conditioning applications in Pakistanwas presented by Khalid et al. [22]. Tests were conducted on agas-fired hybrid desiccant cooling test rig. Using sets of measureddata from the test rig, a validation of a numerical model of thecooling system, in which the gas-fired heater is substituted by a so-lar air collector, was undertaken.

On the other hand, few research investigations were carried outto experimentally evaluate the performance of a desiccant wheelregenerated by means of low temperature thermal energy recov-ered from a cogenerator [23–27]. Moreover, with reference to thecoupling DW–MCHP (Micro Combined Heat and Power), littleattention has been paid on the DW effectiveness variation and onthe DW capability in handling ventilation and internal latent loads.

So, in this paper, the effects of the main thermal-hygrometricproperties (outdoor air humidity ratio and temperature, regenera-tion air temperature) that influence the DW performance havebeen experimentally evaluated; the silica-gel desiccant wheel ispart of a hybrid HVAC system, and the DW regeneration is carriedout by using thermal energy recovered from a microcogenerator. Inparticular, when varying the aforementioned parameters, the ther-mal-hygrometric properties of process air exiting the rotor and theDW effectiveness values were evaluated. Finally, fixing the regen-eration temperature at its maximum available value (65 �C, byusing only the MCHP for regeneration), ventilation and internal la-tent loads that the DW can handle were evaluated and comparedto the required values, considering both a set of cities all overthe world and the entire range of climatic conditions which oc-curred in the experimental tests.

2. Desiccant HVAC system coupled to a microcogenerator

2.1. Experimental apparatus

The main components of the experimental apparatus, located atSannio University (in Benevento, Southern Italy), are the following:

- a HVAC system, based on an air handling unit (AHU) with a des-iccant wheel, an electric chiller and a natural gas-fired boiler;

- a microcogenerator (MCHP), based on a natural gas-fired recip-rocating internal combustion engine.

The nominal characteristics of these components are:

- microcogenerator: electric power 6.0 kW, thermal power11.7 kW, electric efficiency 28.8%, thermal efficiency 56.2%;

- chiller: cooling capacity 8.50 kW, COP = 3.00;- boiler: thermal power 24.2 kW, thermal efficiency 90.2%;

- AHU: process airflow rate is 800 m3/h. During summer opera-tion, this is handled to achieve the supply conditions for theroom (Ts = 13–19 �C, xs = 7–11 g/kg), in order to maintainindoor thermal and humidity comfort.

In Fig. 2, the layout of the test facility is shown. The followingthree airflows are present (entirely drawn from the outdoor, stateO - therefore no recirculation is carried out):

- process air: dehumidified in the DW, (O-2), pre-cooled by inter-acting with the cooling air stream in an air-to-air cross flowheat exchanger (2–3), and finally cooled to the supply temper-ature by a cooling coil connected to the chiller (3–4);

- regeneration air: heated by the heating coil connected with theMCHP (O-5) and, when necessary, by a further heating coil con-nected with the boiler (5–6), in order to regenerate the DWmaterial (6–7);

- cooling air: cooled by a direct evaporative cooler (O-8) and thenused to pre-cool the process air exiting the DW (8–9).

The volumetric rates of the three airflows can be controlled bymeans of manual shutters. When these are at their maximumopening, the airflow rates get their nominal values (800 m3/h).

The position, the measured parameter, the measuring range andthe accuracy of the sensors are reported in Fig. 2 too.

In Fig. 3, the transformations relative to process, regenerationand cooling air are reported on a psychrometric chart.

2.2. Main characteristics of the desiccant wheel

The weight of the DW is 50 kg; its dimensions are700 mm � 700 mm � 440 mm (height �width � thickness), witha diameter of 700 mm. In reality, the frontal area of the rotor effec-tively exposed to process and regeneration airflows is relative to adiameter of about 600 mm, since a circular crown of the total areais obstructed by the metallic frame in the AHU. The rotor has thefollowing configuration: 60% of the rotor area is crossed by the pro-cess air, while the remaining 40% by the regeneration air. This is of-ten used when there is low temperature regeneration thermalenergy.

In fact, the DW is filled with silica-gel, a desiccant material thatcan be effectively regenerated at temperatures as low as 60–70 �C,values obtainable by using the thermal recovery from the MCHPand, only when necessary, the natural gas boiler. The rotor matrixis composed of alternate layers (smooth and wavy) of silica-gelsheets and metallic silicate, chemically bound into an inorganic fi-ber frame.

The so realized ‘‘honeycomb” frame has several advantages,such as the maximization of the superficial contact area, low pres-sure drops, and low weight but high structural durability.

3. Results

3.1. Experimental analysis of the desiccant wheel performances

The desiccant wheel performances, in terms of humidity reduc-tion, process air outlet temperature and effectiveness, depend onseveral operational parameters. In particular, the influence of theregeneration air temperature, as well as the temperature andhumidity ratio of the outdoor air entering the desiccant wheel, isanalyzed in this paper.

Increasing the regeneration air temperature, Treg = T6, with fixedoutdoor humidity ratio, Dx, that is the difference in humidity ratioof process air between upstream and downstream of the DW, in-creases, Fig. 4a. In fact, the moisture desorption process from the

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Fig. 3. Air processes on the psychrometric chart.

1358 G. Angrisani et al. / Applied Energy 88 (2011) 1354–1365

desiccant matrix (on the regeneration side) is endothermic, thusfavored by high temperatures. Increasing the regeneration temper-ature, the DW section being regenerated is subjected to a deeperdrying process. Consequently, the desiccant matrix can attractmore moisture from process air during the successive dehumidifi-cation process.

In Fig. 4a, it can be also noted that the best performance, interms of dehumidification capability (Dx), is obviously obtainedconsidering the isoenthalpyc process (the ideal adsorption pro-cess). Among the three tests considered, at the same regenerationtemperature, the best performance is obtained in test #3, duringwhich xo attains its maximum value, and so the desiccant materialcapability to catch water vapour droplets on its surface is maxi-mized. In fact, the higher the water vapour content in outdoorair, the higher the difference in terms of vapour partial pressurebetween outdoor air and desiccant material surface. Consequently,diffusion of the water vapour droplets from the former to the latteris higher.

When comparing tests #1 and #2 (Fig. 4a), characterized bysimilar values of xo, the dehumidification capability is higher fortest #1, due to the lower outdoor air temperature. This is so be-cause the adsorption process is exothermic, hence favored by lowtemperatures.

In Fig. 4b, Dx as a function of xo is reported, for different To andfor Treg = 65.0 �C. It monotonically increases with xo, while, forfixed values of xo, it decreases when To rises, as already explainedfor Fig. 4a. These results are in good agreement with data suppliedby the manufacturer, where the experimental Dx is lower than theindicated one by about 15%.

In Fig. 5a, DT, the process air temperature difference betweendownstream and upstream of the DW, is reported as a functionof Treg, for three tests characterized by different values of To andxo. Evidently, DT increases with Treg, because of the heating ofthe desiccant matrix on the regeneration side and, as a conse-quence of the rotation of the wheel, on the process side. Obviouslythe increase in DT is only due to the rise in T2, To being constant.Moreover, the lowest values of DT occur in test #1, due to the low-est values of To and xo. As regards the remaining tests, test #3 ischaracterized by a lower To but a higher xo compared to test #2.This last condition is predominant such that the highest DT occursfor test #3. In fact, DT increases with process air inlet humidity ra-

tio (see also Fig. 5b, referred to a Treg = 65.0 �C) as the wheel re-moves a greater quantity of water vapour when xo grows(Fig. 4b). Therefore, DT rises due to the increase in the adsorptionheat released during the process. Moreover, DT is lower in the caseof isoenthalpyc process.

In Fig. 5b, DT as a function of xo is reported, for different valuesof To. It can be seen that DT increases when xo rises and/or To de-creases, because in these cases the adsorption process is enhanced(Fig. 4b) and so the adsorption heat rises.

The results reported in Figs. 4 and 5 are in good agreement withsimilar analysis reported in literature [4,28].

Some definitions of desiccant wheel effectiveness are used[18,29–31]:

- the thermal effectiveness (1), the conventional effectiveness def-inition for a heat exchanger:

gth ¼ ðT2 � ToÞ=ðT6 � ToÞ ð1Þ

- the regeneration effectiveness (2):greg ¼ ðxo �x2ÞDhvs=ðh6 � hoÞ ð2Þ

It expresses, for unitary mass flow rate, the latent load handledby the DW with respect to the regeneration thermal power re-quired for the adsorption process;

- the dehumidification effectiveness (3):

gdeh ¼ ðxo �x2Þ=ðxo �x2;idealÞ ð3Þ

where x2,ideal is the ideal humidity ratio of process air stream at thedesiccant wheel outlet. If its value is zero, the process in the DW isideal and the process air is completely dehumidified. So, gdeh ex-presses the comparison between the real dehumidification capabil-ity and the ideal one.

- the adiabatic effectiveness (4):gad ¼ 1� ðh2 � h0Þ=h0 ð4Þ

The real adsorption process is not isoenthalpyc (air enthalpy in-creases), so the comparison between the real adsorption processand the isoenthalpyc one becomes significant.

In Fig. 6a, the values of the above-mentioned effectiveness as afunction of Treg are shown, for fixed values of To and xo. When the

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Fig. 4. Difference between process air humidity ratios at inlet and outlet of the desiccant wheel as a function of regeneration temperature (a) and outside air humidity ratiowith Treg = 65.0 �C (b).

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regeneration temperature rises, there is an increase in dehumidifi-cation effectiveness, gdeh, thanks to the enhancement of the corre-sponding desorption process.

Moreover, the rise of Treg causes an increase in the heat lossesfrom the hot side of the DW (regeneration section) to both the coldside (process section) and the outdoor environment, due to en-hanced convective – conductive heat transfer mechanisms, and astronger heating of the matrix and of the desiccant material is alsocaused. Thus, the regeneration effectiveness decreases (the aug-mentation of the latent load handled by the DW does not balancethe increase in the regeneration thermal power).

Furthermore, the enthalpy h2 increases, causing a light fall ofthe adiabatic effectiveness. Finally, the consequent augmentationof T2 is not able to balance the increase of the regenerationtemperature, so the thermal effectiveness has a descendingbehavior.

In Fig. 6b, the four effectiveness trends are reported as a func-tion of the outdoor air temperature, for fixed values of Treg andxo. As already noted in Fig. 4a, the growth of To leads to a reduction

of the dehumidification wheel capability, thus to a decreasing ofdehumidification and regeneration effectiveness.

With higher air temperatures at the DW inlet, To, the adsorptionprocess is penalized. Consequently, the reduction of the releasedheat of adsorption leads to the enhancement of the adiabaticefficiency.

As regards gth (1), at constant regeneration temperature, pro-cess air temperatures at the inlet and the outlet of the rotor bothincrease. The thermal effectiveness passes through a decliningtrend, in accordance with the manufacturer’s data.

The DW effectiveness trends shown in Fig. 6a and b are in goodagreement with the trends reported in [18].

In Fig. 6c the DW effectiveness values are reported as a functionof outdoor air humidity ratio, for fixed values of Treg and To.

As regards greg, increase in xo causes an augmentation of thedehumidification capability, xo �x2, cfr. Fig. 4b, while Treg is fixed,and so greg grows. With reference to gth, the growth of dehumidifi-cation capability determines a rise in the released heat of adsorp-tion, therefore both T2 and gth rise.

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Fig. 5. Difference between process air temperatures at outlet and inlet of the desiccant wheel as a function of regeneration temperature (a) and outside air humidity ratiowith Treg = 65.0 �C (b).

1360 G. Angrisani et al. / Applied Energy 88 (2011) 1354–1365

At constant outdoor temperature, the increase in the specifichumidity strongly affects the dehumidification capability in theideal process, xo �x2,ideal, leading to the decreasing trend of thedehumidification effectiveness.

Finally, at higher outdoor humidity ratio, adiabatic effective-ness improves, according to data supplied by the DWmanufacturer.

Fig. 7 refers to data obtained in the absence of thermal energydelivered to regeneration air, in order to highlight the process ofdesiccant saturation. The graph shows the time trend of Dx fortwo different tests, characterized by similar temperatures, but dif-ferent moisture contents of outdoor air. The two curves start fromthe same initial value of Dx, and then continue differently. In par-ticular, in the test #2, characterized by a higher value of xo, thesaturation process is more rapid. In fact, at the same time, theDx residual is less than in the test #1, and this is due to the higherquantity of water vapour in the air for the test #2, which tends tosaturate more rapidly the surface of the desiccant. After about

76 min during which the wheel dries the process air without beingregenerated, the saturation process may be considered almost en-tirely completed, as Dx residual at the end of the test is about 1/4and 1/8 for test #1 and test #2 respectively, compared to the initialDx.

3.2. Latent load handled by the desiccant wheel

On the basis of the experimental data previously reported, ananalysis on the silica-gel desiccant wheel has been carried out toestimate its performance in handling the ventilation and internallatent loads of a conditioned space. The process air volumetric flowrate has a negligible influence on the desiccant wheel performance(in terms of dehumidification capability, i.e. Dx) if the followingconditions are satisfied:

(a) the face velocity of the process air at the DW inlet remainsconstant;

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90

100

45 50 55 60 65 70

Regeneration temperature [°C]

[%]

thermal

dehumidification

regeneration

adiabatic

ωo = 13.0 g/kg

To = 32.7°C

20

30

40

50

60

70

80

90

25 27 29 31 33 35

Process air inlet temperature [°C]

[%]

thermal

dehumidification

adiabatic

regeneration

ωo = 11.2 g/kg

Treg = 66.4°C

20

30

40

50

60

70

80

90

7 9 11 13 15

Process air inlet humidity ratio [g/kg]

[%]

thermal

adiabatic

dehumidification

regeneration

To = 31.6°C

Treg = 66.2°C

(a)

(b)

(c)

Fig. 6. Various types of desiccant wheel effectiveness as a function of regenerationtemperature (a), process air inlet temperature (b) and humidity ratio (c).

0.5

1.0

1.5

2.0

2.5

3.0

3.5

4.0

1 16 31

Time

ω [g

/kg]

Fig. 7. Difference between process air humidity ratios at inlet and outlet of the desicca

G. Angrisani et al. / Applied Energy 88 (2011) 1354–1365 1361

(b) the ratio between process air and regeneration air flow rateis equal to 1, as in the previously reported experimentalanalysis.

In this case, the DW dehumidification performance is substan-tially independent of volumetric flow rate. In the following analy-sis, the ventilation and internal latent loads are expressed perunitary volumetric air flow rate, considering indoor thermal-hygro-metric conditions characterized by Tr = 25 �C and xr = 10.5 g/kg(relative humidity = 54%), and supply air temperature equal to17 �C; Treg is fixed to 65 �C (maximum value obtainable by usingthe selected MCHP).

The outdoor design conditions herein considered are based onthe following ASHRAE data [32]:

- design for cooling: 0.4% DB–MCWB (dry bulb–mean coincidentwet bulb), 1.0% DB–MCWB and 2.0% DB–MCWB;

- design for dehumidification: 0.4% DP–MCDB (dew point–meancoincident dry bulb) and 1.0% DP–MCDB.

It is helpful to use an example for understanding these designdata. For Istanbul, 0.4% DB = 31.1 �C and MCWB = 21.4 �C: as re-gards DB temperature, this means that the value 31.1 �C is ex-ceeded on average by the indicated percentage (0.4%) of the totalnumber of hours in a year (8760), i.e. by 35 h per year, for the per-iod of record; as regards MCWB temperature, the value 21.4 �C de-fines the average wet bulb temperature that was observed whenthe air was at the extreme high dry bulb temperature (31.1 �C).In other words, the value 21.4 �C is not the average wet bulb tem-perature during the entire warm season, but just its average valuewhen the air is very hot [33].

The designer has to choose the set of conditions and probabilityof occurrence that can be conveniently applied to the specific situ-ation. The DB–MCWB data represent outdoor conditions of hot,mostly sunny days. They are therefore commonly used in sizingcooling equipment, such as chillers and cooling coils. On the con-trary, design conditions based on dew point temperatures (DP–MCDB) are directly related to the highest values of humidity ratio,which represent peak moisture loads from the weather. These val-ues are especially useful for humidity control applications, hencein desiccant cooling and dehumidification, cooling-based dehu-midification and ventilation system design.

46 61 76

[min]

test #1 (To = 26.7°C,ωo = 5.37 g/kg)

test #2 (To = 26.1°C, ωo = 7.56 g/kg)

test #1

test #2

nt wheel as a function of the time, in the absence of regeneration thermal energy.

(a)

(b)

0

1

2

3

4

5

Wie

n/C

ity

Ath

inai

Bud

apes

t/F

erih

egy

I

Beo

grad

Sar

ajev

o/B

utm

ir

Zag

reb/

Mak

sim

ir

Buc

ures

ti A

fum

ati

Bol

ogna

Fir

enze

Ista

nbul

Cas

abla

nca

Bos

ton

Ott

awa

Toro

nto

But

tonv

ille

A

Lim

a

Syd

ney

Air

port

Am

o

Spe

cifi

c ve

ntila

tion

lat

ent l

oad

ha

ndle

d b

y th

e D

W [

W/(

m3 /

h)]

0.4 % DB-MCWB 1.0 % DB-MCWB 2.0 % DB-MCWB

93%84% 90%

90%95%

90%91%

98%

2

3

4

5

6

Wie

n/C

ity

Ath

inai

Bud

apes

t/F

erih

egy

I

Beo

grad

Sar

ajev

o/B

utm

ir

Zag

reb/

Mak

sim

ir

Buc

ures

ti A

fum

ati

Bol

ogna

Fir

enze

Ista

nbul

Cas

abla

nca

Bos

ton

Ott

awa

Toro

nto

But

tonv

ille

A

Lim

a

Syd

ney

Air

port

Am

o

Spe

cifi

c ve

ntila

tion

lat

ent l

oad

ha

ndle

d b

y th

e D

W [

W/(

m3 /

h)]

0.4 % DP-MCDB 1.0 % DP-MCDB

Fig. 8. Specific ventilation latent load handled by the desiccant wheel for various cities and for different outdoor design thermal-hygrometric conditions (Treg = 65 �C).

1362 G. Angrisani et al. / Applied Energy 88 (2011) 1354–1365

In Fig. 8, the specific ventilation latent load that can be handledby the desiccant wheel for different cities around the world is re-ported. The cities selected for the analysis are characterized by val-ues of the outdoor humidity ratio (derived from DB–MCWB andDP–MCDB data) included within the range of experimental dataavailable for the desiccant wheel performance (see Section 3.1).

Fig. 8a refers to DB–MCWB data (for cooling purposes), whileFig. 8b refers to DP–MCDB data (for dehumidification purposes).For those cities where the DW is able to handle the entire requiredspecific ventilation latent load (L, proportional to xo �xr), it is re-ported on the ordinate. This occurs for all the cities in Fig. 8a andonly for some of them in Fig. 8b. On the contrary, for those citieswhere the DW cannot entirely balance L, only the fraction of L cov-ered by the DW (proportional to xo �xs, with xs experimentallyevaluated) is shown on the ordinate and reported in percentagetoo. This occurs for some cities (such as Athinai and Bologna) inFig. 8b. This result strictly depends on the above-mentioned differ-ence between DB–MCWB and DP–MCDB data. Moreover, note thatfor some cities and for certain percentiles (Athinai, Beograd, Sara-jievo, Sydney), the ordinate is equal to zero (Fig. 8a) because the in-door humidity ratio is higher than outdoor one (i.e., there is noventilation latent load).

As regards the experimental evaluation of xs (the supply airhumidity ratio, equal to the process air humidity ratio after the

desiccant wheel), it is obtained for each city by the experimentalknowledge of the maximum desiccant wheel dehumidificationcapability (DxDW) relative to a regeneration temperature of65 �C, i.e. xs = xo � DxDW.

In Fig. 9a and b, the maximum specific internal latent load thatcould be handled by the desiccant wheel (independently of thetype of internal latent load) is presented, considering the ASHRAEcooling and dehumidification design data, respectively. This loadis proportional to xr �xs, with xs experimentally evaluated as re-ported above.

In Fig. 9b, note that for the cities in which the ventilation latentload cannot be completely handled by the desiccant wheel (be-cause xs > xr, as occurs, for example, in Bologna), no internal la-tent load can be balanced by the wheel, such that negativevalues are shown. In these cases, the lower the ventilation latentload covered by the DW with respect to the required value(Fig. 8b), the higher the absolute value of the internal latent loadreported in Fig. 9b.

Thus, in these cases, a higher Treg is necessary and the onlywaste heat from MCHP is not sufficient for regeneration process.

The results reported in Fig. 9, for given cities, can be generalizedfor any climatic condition, which is the main purpose of Fig. 10,where the following parameters have been reported as a functionof outdoor air humidity ratio:

0

1

2

3

4

5

Wie

n/C

ity

Ath

inai

Bud

apes

t/F

erih

egy

I

Beo

grad

Sar

ajev

o/B

utm

ir

Zag

reb/

Mak

sim

ir

Buc

ures

ti A

fum

ati

Bol

ogna

Fir

enze

Ista

nbul

Cas

abla

nca

Bos

ton

Ott

awa

Toro

nto

But

tonv

ille

A

Lim

a

Syd

ney

Air

port

Am

o

Spe

cifi

c in

tern

al l

aten

t lo

ad

hand

led

by

the

DW

[W

/(m

3 /h)

]

0.4 % DB-MCWB 1.0 % DB-MCWB 2.0 % DB-MCWB

-1.0

-0.5

0.0

0.5

1.0

1.5

2.0

2.5

3.0

Wie

n/C

ity

Ath

inai

Bud

apes

t/F

erih

egy

I

Beo

grad

Sar

ajev

o/B

utm

ir

Zag

reb/

Mak

sim

ir

Buc

ures

ti A

fum

ati

Bol

ogna

Fir

enze

Ista

nbul

Cas

abla

nca

Bos

ton

Ott

awa

Toro

nto

But

tonv

ille

A

Lim

a

Syd

ney

Air

port

Am

o

Spe

cifi

c in

tern

al l

aten

t lo

ad

hand

led

by

the

DW

[W

/(m

3 /h)

]

0.4 % DP-MCDB 1.0 % DP-MCDB

(a)

(b)

Fig. 9. Specific internal latent load handled by the desiccant wheel for various cities and for different outdoor design thermal-hygrometric conditions (Treg = 65 �C).

G. Angrisani et al. / Applied Energy 88 (2011) 1354–1365 1363

- the specific latent load that can be globally handled by the des-iccant wheel, for different outdoor temperatures (Fig. 10a, dot-ted lines). This is proportional to xo �xs;

- the required specific ventilation latent load (Fig. 10a, continu-ous line). This is proportional to xo �xr. Therefore, for fixedindoor conditions, it depends only on the climatic conditions;

- the specific internal latent load that can be balanced by the des-iccant wheel, for different outdoor air temperatures (Fig. 10b).This is proportional to xr �xs.

Analyzing Fig. 10a, it can be noted that, for values of outdoorhumidity ratio lower than about 15.5 g/kg, the specific latent loadthat can be handled by the desiccant wheel is higher than the spe-cific ventilation latent load required by the outdoor air humidityratio. This means that in these conditions the desiccant wheel is al-ways able to balance at least the ventilation latent load. On thecontrary, for outdoor humidity ratio higher than 15.5 g/kg, the des-iccant wheel can handle the required ventilation latent load onlyfor low outdoor air temperatures. This confirms that the processair inlet temperature also strongly affects the DW dehumidificationcapability.

As a consequence, in Fig. 10b, only for outdoor humidity ratiolower than about 15.5 g/kg, the desiccant wheel can also balancea part or all the internal latent load, such that positive values occur.

In order to check if the wheel is able to entirely cover the inter-nal latent load for a given application, Fig. 11 (coupled to Fig. 10b)

can be useful. In Fig. 11, the specific internal latent load required bya given application is presented as a function of the room SHR (sen-sible heat ratio, i.e. the ratio between the sensible and total ther-mal load). This allows to identify the value of the specificinternal latent load to be balanced (Fig. 11) through the knowledgeof both the sensible and latent load of the considered indoor ambi-ent. Successively, by the diagram of Fig. 10b, the user can check ifthe desiccant wheel is able to balance the required internal latentload for certain outdoor design conditions.

With reference to both Figs. 10 and 11, in all the cases in whichthe desiccant wheel cannot balance the required latent load, ahigher Treg is necessary.

4. Conclusions

An experimental analysis on the performances of a silica-geldesiccant wheel, inserted in a test facility characterized by a desic-cant air handling unit coupled to an electric chiller, a boiler and amicrocogenerator has been presented. The desiccant wheel isregenerated by using thermal energy from a microcogenerator,therefore at low temperature (lower than 70 �C).

Useful desiccant wheel performance curves were obtained, asregards the difference between process air inlet and outlet humid-ity ratio (i.e. the moisture removal capability of the wheel) andtemperature, as a function of regeneration temperature and pro-cess air inlet humidity ratio and temperature. Also the trends of

0.0

0.5

1.0

1.5

2.0

2.5

3.0

3.5

0.5 0.6 0.7 0.8 0.9 1

Spe

cifi

c in

tern

al l

aten

t lo

ad re

quir

ed [

W/(

m3 /

h)]

SHR [-]

Fig. 11. Specific internal latent load required as a function of sensible heat ratio(SHR).

0

1

2

3

4

5

6

78 9 10 11 12 13 14 15 16 17 18

Outdoor air humidity ratio [g/kg]Sp

ecif

ic la

tent

load

hand

led

by th

e D

W [

W/(

m3 /h

)]

To = 25.0°C 27.0°C 29.0°C 31.5°C 34.5°C

Req

uire

d sp

ecif

ic v

entil

atio

n

l

aten

t loa

d [W

/(m

3 /h)]

-2

-1

0

1

2

3

4

5

6

8 9 10 11 12 13 14 15 16 17 18

Outdoor air humidity ratio [g/kg]

Spec

ific

inte

rnal

late

nt lo

ad

hand

led

by th

e D

W [

W/(

m3 /h

)]

To = 25.0°C 27.0°C 29.0°C 31.5°C 34.5°C

(a)

(b)

Fig. 10. Specific latent load handled by the DW (a), required specific ventilation latent load (a) and specific internal latent load handled by the DW (b), as a function of outdoorhumidity ratio (Treg = 65 �C).

1364 G. Angrisani et al. / Applied Energy 88 (2011) 1354–1365

various types of DW effectiveness (regeneration, dehumidification,thermal and adiabatic effectiveness) were analyzed, as a functionof the abovementioned parameters.

The results obtained provided useful information on the perfor-mances of a silica-gel desiccant wheel regenerated by a microco-generator (low regeneration temperature range, i.e. 40–70 �C), asthis issue was poorly investigated in literature.

The performance curves considering an ideal isoenthalpycadsorption dehumidification process were presented too. In thisway, it is possible to compare and evaluate the deviation entityof the real evaluated performances from the ideal conditions.

Furthermore, an analysis was presented to evaluate the desic-cant wheel performance in handling the ventilation and internallatent load for various cities around the world. The results showedthat the selected outdoor thermal-hygrometric design datastrongly affect the possibility for the desiccant wheel to completelybalance the latent load. When using ASHRAE design dehumidifica-tion data (characterized by high dew point temperature) for thecities analyzed, the desiccant wheel is often not able to entirelyhandle even the ventilation latent load (due also to the low regen-eration temperature, equal to about 65 �C). On the contrary, whenusing ASHRAE design cooling data, the wheel always allows at leastthe covering of the ventilation latent load.

Then, the results obtained for given cities were generalizedfor any climatic condition. On the basis of the obtained experi-mental performance curves, useful diagrams were obtained toevaluate the specific latent load that the desiccant wheel canhandle as a function of outdoor humidity ratio, for different out-door temperatures. The results showed that, for outdoor humid-ity ratio higher than about 15.5 g/kg, the desiccant wheel canbalance at least the ventilation latent load only for low outdoortemperatures. This confirms that the process air inlet tempera-ture also strongly affects the DW dehumidification capability.Moreover, the specific internal latent load that can be handledby the desiccant wheel versus outdoor humidity ratio for differ-ent outdoor temperatures were evaluated. A simple chart toidentify the required specific internal latent load as a functionof SHR was obtained. In this way, the designer can easily evalu-ate the specific internal latent load for the analyzed indoorambient (characterized by a given SHR) and check if the silica-gel desiccant wheel with low regeneration temperature is ableto balance it, otherwise, higher regeneration temperatures are

G. Angrisani et al. / Applied Energy 88 (2011) 1354–1365 1365

necessary and the waste heat from MCHP could be not sufficientfor the desiccant regeneration.

Acknowledgments

This work has been financed by Italian research project PRIN2007 ‘‘Criteria and methodologies for the optimization of small/medium scale polygeneration systems”.

Moreover, this work was developed in a project promoted byInternational Energy Agency (IEA), Annex 54, Integration of Mi-cro-generation and Related Energy Technologies in Buildings.

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