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Energy 37 (2012) 336e345

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Energy

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Simulation and energy efficiency analysis of desiccant wheel systems for dryingprocesses

Stefano De Antonellis*, Cesare Maria Joppolo, Luca Molinaroli, Alberto PasiniPolitecnico di Milano, Dipartimento di Energia, Via Lambruschini 4, 20156 Milan, Italy

a r t i c l e i n f o

Article history:Received 1 July 2011Received in revised form21 October 2011Accepted 11 November 2011Available online 12 December 2011

Keywords:Desiccant wheelDehumidificationDryingEfficiencyEnergy savingHVAC

* Corresponding author. Tel.: þ39 223993823; fax:E-mail address: stefano.deantonellis@polimi.it (S.

0360-5442/$ e see front matter � 2011 Elsevier Ltd.doi:10.1016/j.energy.2011.11.021

a b s t r a c t

In drying processes it is necessary to appropriately control air humidity and temperature in order toenhance water evaporation from product surface. The aim of this work is to investigate several HVACconfigurations for product drying based on desiccant wheels, in order to find systems which reach highprimary energy savings through the appropriate integration of refrigerating machines, adsorptionwheelsand cogenerative engines. Simulations are carried out for different values of sensible to latent ambientload ratio and the effect of ambient and outside air conditions is evaluated for each configuration. It isshown that primary energy savings can reach 70e80% compared to the reference technology based ona cooling coil. With respect to works available in literature, the results of this study keep a generalapproach and they can be used as a simple tool for preliminary assessment in awide range of applications.

� 2011 Elsevier Ltd. All rights reserved.

1. Introduction

In many food drying processes, for example of rawmeat, cheeseand vegetables, products are kept in a ventilated room at controlledtemperature and humidity: an air handling unit (AHU), whichworks in recirculating mode, treats air in order to balance thesensible and the latent load (Fig. 1). In conventional AHU thedehumidification process is based on a cooling coil: the air is firstlycooled below its dew point temperature and then it is heated untilthe desired supply temperature is reached. This technology is themost diffused and it is considered as the reference one in mostapplications [1].

These processes can be often realized through more efficientHVAC systems based on desiccant wheels. Compared to conven-tional systems, the use of adsorption wheels can lead to thefollowing advantages:

- a reduction of the refrigerating group electricity consumptionand an increase of its evaporation temperature and EER(Energy Efficiency Ratio), because the cooling coil balancesonly the sensible load;

þ39 223993913.De Antonellis).

All rights reserved.

- a reduction of the heat consumption because the post heatingcoil is not necessary anymore;

- a reduction, or elimination, of the energy consumption requiredfor the evaporator defrosting in low temperature applications;

- a reduction of the presence of microorganisms like bacteria andfungi due to the absence of condensed water;

- the possibility to use low temperature sources to activate thedehumidification process (50e60 �C).

Many research works about performance analysis of desiccantsystems are available in literature and a detailed review of the mainHVAC dehumidification systems is carried out by Mazzei et al. [2].Since the first desiccant evaporative cooling cycle has been intro-duced by Pennington [3], many desiccant cooling and dehumidifi-cation systems based on sorption wheels have been developed.HVAC systems that integrate several desiccant wheels in the airhandling unit have been discussed by Kodama et al. [4,5]. Thesesolutions are particularly suitable for low regeneration temperatureapplications but they are bigger and more complicated than theconfigurations based on one desiccant wheel. Similar consider-ations can be drawn for the one rotor e two stage desiccant wheelsystems investigated from Ge et al. [6,7].

Many works analyze the performance of hybrid vapourcompressione desiccant wheels systems [8e13] and show that thistechnology can save significant energy compared to the referenceone based on a conventional cooling coil.

Fig. 1. Room and air flow scheme.

S. De Antonellis et al. / Energy 37 (2012) 336e345 337

In addition to the technology based on sorption wheels, otherchemical dehumidification solutions are possible: water vapour canbe absorbed through a liquid solution [14] or adsorbed througha solid desiccant bed [15,16]. The first one is not appropriate due tothe problem of desiccant solution dragging into the air stream,which can be a serious problem in many drying processes and inparticular in the food industry. The second one is characterized bythe temperature and humidity swinging of the supply air stream.For these reasons sorption wheels are typically preferred aschemical dehumidification technology.

The use of sorptionmaterials for product drying and consequentadvantages are investigated and discussed in several works. Mad-hiyanona et al. [17] use a solid desiccant to dry coconut, Nagayaet al. [18] propose a desiccant wheel to dry vegetables and Pasini[19,20] to dry salami. Thoruwa et al. [21] and Dai et al. [22] proposea solar driven adsorption system for grain drying. Wang et al. [23]evaluate performance of a surface drying hybrid device based ona desiccant wheel.

A detailed comparison among different hybrid systems fordrying purposes based on desiccant wheels is not available inliterature. Furthermore most studies are related to a specificapplication, mainly different from product drying, and for thisreason results are not easily applicable to other contests.

The aim of this work is to evaluate primary energy consumptionof different HVAC systems for drying purposes, based on sorptionwheels, and to compare the obtained results with the referencetechnology based on a cooling coil. The analysis is carried out forambient temperature between 2 �C and 25 �C and relative humiditybetween 50% and 75%: this operating range represents typicalconditions of raw ham, salami and cheese drying processes. Simu-lations are carried out for different sensible to latent ambient loadratios and for several ambient and outside air conditions. In thisway, compared to available literature, the results can be adapted todifferent fields and they are not limited to a specific application.

2. Air handling units (AHU) description

Air handling units based on desiccantwheels can be combined indifferent ways, each of them leading to specific advantages. A largecombination of systems is possible and, in addition, they can berealized both through separated or cogenerative power production.For this reason only representative systems are considered. In Figs. 1and 2 and Table 1 the schemes and the components of eachanalyzed configuration (denoted with a capital letter) are reported.

Up to three air streams are present in the investigated airhandling units:

- The supply air flow, which is the main one treated in the airhandling unit.

- The process air flow crossing the desiccant wheel (configura-tion B, C, D, E and F), that is a fraction of the supply air stream.

- The regeneration air flow that is introduced to remove theadsorbed water from the desiccant wheel (only in configura-tion B, C, D and E).

As already introduced in Section 1, the drying system works inrecirculation mode. In fact in this kind of applications air change isnot required and, maybe, it should be even avoided in order topreserve bacteria that are necessary for the product drying process(such as in the cheese and raw ham industry). Therefore, in all theinvestigated configurations, it is assumed the AHU is fed by an airstream at ambient conditions. If present, the regeneration air flowis fed by air at outside conditions.

Hereinafter the reference and the desiccant wheel basedconfigurations discussed in this works are briefly described:

- Configuration A is the reference and conventional solution: thesupply air flow is cooled and dehumidified through the coolingcoil a, in order to satisfy the latent load, and then it is heatedthrough coil c until the sensible load is balanced.

- Configuration Arec is similar to the reference one (A): in thiscase heat rejected from the condenser is partially recovered forthe post heating process of the supply air stream.

- Configuration B is the simplest solution based on a desiccantwheel: the process air is first dehumidified and heated throughthe desiccant wheel, and then it is cooled (or heated) throughcoil b (or c). The wheel is regenerated through a highertemperature air stream which is heated through coil f.

- Configuration C is similar to the B one and it integratesa sensible heat wheel which reduces the power required toheat the regeneration air stream.

- In configurationD and Eadirectexpansion refrigerating group isused to cool the process air stream through the evaporator h or land to heat the regeneration air stream through the condenser iorm. While in the first case the evaporator cools the process airat the desiccant wheel outlet, in the second one it cools anddehumidifies the air stream before crossing the wheel.

- In configuration F a direct expansion refrigerating group is alsointroduced. In this case there is not a second regeneration airflow: adsorbed water is directly removed from the desiccantmatrix through the process air stream. The process air is firstheated through the condenser n, it is cooled and humidifiedthrough the desiccant wheel, it is cooled and dehumidified

Fig. 2. Schemes of the investigated HVAC systems.

S. De Antonellis et al. / Energy 37 (2012) 336e345338

through the evaporator p and finally heated and dehumidifiedagain through the wheel.

Quite obviously in configuration B, C, D, E and F the cooling andheating coils b and c are introduced in order to balance the ambientsensible load.

3. Components modelling

Referring to Fig. 2 and Table 3, the modelling of the componentsof the investigated air handling units is described.

3.1. Desiccant wheel

In the present work, desiccant wheel performance is calculatedthrough a one dimensional gas side resistance model based on heatand mass transfer equations, that is able to accurately predict thebehaviour of the wheel under different operating conditions. The

model calculates average air temperature, air humidity and pres-sure drop of each flow leaving the component.

In the simulations the process to regeneration area ratio Apro/Areg

and the revolution speed N are always optimized in order tomaximize the moisture removal capacity of the component. Only inconfiguration F balanced air flows are assumed (Apro/Areg ¼ 1). Adetailed description of the model, its validation and the differentoptimization criteria are reported in a previous work [23].

3.2. Direct expansion refrigerator

The vapour compression refrigerating groups RGmain,S, RGmain,S,L,RGD, RGE and RGF are modelled through experimental data availablefrom a manufacturer [25]. The system is based on a semi-hermeticreciprocating compressor, refrigerant is R-407C, the liquid sub-cooling and the vapour super-heating are assumed equal to 5 �C.From the evaporation and condensation temperature the EER of thesystem is determined.

Table 3Components input data.

Case 1 Case 2 Case 3 Case 4 Case 5

Design conditionsQL [kW] 1 1 1 1 1QS [kW] �3/3 �3/3 �3/3 �3/3 �3/3msupply [kg/s] 1 1 1 1 1

Evaporator a and bTeva,a [�C] 5 5 5 �10 10Teva,b [�C] 10 10 10 �1 12,5

Refrigerating groups RGmain,S,L, and RGmain,S

Tcond,RGmain[�C] 30 20 40 30 30

Tcond,RGmain(G) [�C] 40 30 50 40 40

Heating coil (f)Treg,f [�C] 60 60 60 60 60

Condenser i, m, n and qTcond,i [�C] 60 60 60 60 60Tcond,m [�C] 40 40 40 e 40Tcond,n [�C] 30 30 30 30 30Tcond,q [�C] 25 25 25 e 30Dpi,m,n,q [Pa] 100 100 100 100 100

Evaporator h, l, p and rTeva,h [�C] 10 10 10 0 10Teva,l [�C] 5 5 5 e 10Teva,p [�C] 5 5 5 0 10Teva,r [�C] 5 5 5 e 5Dph,l,p,r [Pa] 100 100 100 100 100

Heat wheel gDpg [Pa] 50 50 50 50 50xS [e] 0,7 0,7 0,7 0,7 0,7

Filters d and eDpd,e [Pa] 100 100 100 100 100

Other dataDTres,eva [�C] 1 1 1 1 1DTres,cond [�C] 2 2 2 2 2DTpinch,cond [�C] 5 5 5 5 5DTw,cond [�C] 5 5 5 5 5

Table 1Components of each investigated air handling unit.

Configuration

A Arec B C D E F

Desiccant wheel e unbalanced flows (dw) X X X XDesiccant wheel e balanced flows (dw0) XEvaporator (a) X XEvaporator (b) X X X X XHeating coil (c) X X X X X XHeating coil (crec ¼ c þ crec,cond) XFilter (d) X X X X XFilter (e) X X X XHeating coil (f) X XSensible heat wheel (g) XEvaporator (h) XAir condenser (i) XEvaporator (l) XAir condenser (m) XAir condenser (n) XEvaporator (p) X

S. De Antonellis et al. / Energy 37 (2012) 336e345 339

In the study only electric power consumption related to the useof the compressor is considered: power related to auxiliaries is nottaken into account.

It is assumed that condensers of the main refrigerating groupRGmain,S and RGmain,S,L are cooled through water from a coolingtower. Therefore the condensation temperature depends onoutside air conditions and in particular onwet bulb temperature, asresumed in Tables 2 and 3.

Note that the main refrigerating groups are named RGmain,S andRGmain,S,L respectively when only sensible or both sensible andlatent heat are exchanged at the evaporator.

DTpinch,HC [�C] 10 10 10 10 10v [m/s] 1,5 1,5 1,5 1,5 1,5hfan [e] 0,6 0,6 0,6 0,6 0,6hdefrost [e] e e e 0,4 e

hth,sep [e] 0,9 0,9 0,9 0,9 0,9hel,sep [e] 0,45 0,45 0,45 0,45 0,45hth,cog [e] 0,45 0,45 0,45 0,45 0,45hel,cog [e] 0,35 0,35 0,35 0,35 0,35

3.3. Evaporator

Dry evaporators are well known devices [1]: performance ofcomponents a, b, h, l and p is evaluated in terms of by-pass factor(BPF) and air side pressure drop (Dpeva). Both values are assumedconstant according to experimental data available in open litera-ture [19].

The BPF is defined in this way:

BPF ¼ Xair;out;eva � XADP

Xair;in;eva � XADPy

Tair;out;eva � TADPTair;in;eva � TADP

(1)

and the apparatus dew point temperature is assumedTADP ¼ Teva þ DTres,eva, where Teva is the working fluid evaporationtemperature and DTres,eva is related to the conductive and convec-tive fin resistance.

The sensible and latent power exchanged from the coil isrespectively:

PS;eva ¼ mair;evacpair�Tair;out;eva � Tair;in;eva

�(2)

PL;eva ¼ mair;evall�v

�Xair;out;eva � Xair;in;eva

�(3a)

or

Table 2Ambient and outside air conditions.

Case 1 Case 2 Case 3 Case 4 Case 5

Ta [�C] 15 15 15 2 254a [%] 75 75 75 75 50Te [�C] 25 15 35 25 254e [%] 50 50 50 50 50

PL;eva ¼ mair;evaðll�v þ ls�lÞ�Xair;out;eva � Xair;in;eva

�(3b)

in case of frost formation on fins.The total power exchanged by the coil is calculated in this way:

��PL;eva þ PS;eva� ¼ PcompEER (4)

3.4. Air condenser

The air condensers i, m and n of the direct expansion refriger-ating groups are simply modelled through the followinghypothesis:

- the air temperature leaving the component is assumedTair,out ¼ Tcond � DTres,cond, as already introduced for theevaporator;

- the air pressure drop is assumed constant.

The sensible power exchanged by the component is:

PS;cond ¼ mair;condcpair�Tair;out;cond � Tair;in;cond

�(5a)

or

PS;cond ¼ Pcompð1þ EERÞ (5b)

S. De Antonellis et al. / Energy 37 (2012) 336e345340

3.5. Water condenser

Water condensers of refrigerating groups RGmain,S and RGmain,S,L

are modelled in this way:

- the outlet water temperature is assumed Tw,out,cond ¼ Tcond� DTpinch,cond.

- The inlet water temperature is Tw,in,cond ¼ Tw,out,cond � DTw,cond.

The sensible power exchanged by the coil is:

PS;cond ¼ mwcpw�Tw;out;cond � Tw;in;cond

�(6a)

or

PS;cond ¼ Pcompð1þ EERÞ (6b)

3.6. Heating coil

Referring to Fig. 2, heating coils c and f are assumed fed bya constant water flowand the supplied water temperature Tw,in,HC iscontrolled through a 3-way valve. The water that passes throughthe component is heated through a boiler or a cogenerative engine.The desired air temperature at the coil outlet is always reachedthrough the variation of the supply water temperature, and it iscalculated as Tair,out,HC ¼ Tw,in,HC � DTpinch,HC. The pressure dropacross the component is constant.

The sensible power exchanged by the heating coil is:

PS;HC ¼ mair;HCcpair�Tair;out;HC � Tair;in;HC

�¼ mw;HCcpw

�Tw;in;HC � Tw;out;HC

� (7)

3.7. Heating coil with condenser heat recovery

Heating coil with condenser heat recovery (denoted in Fig. 2with crec e config. Arec) is simply assumed split in two parts: thefirst one is fed by heat rejected from the condenser (crec,cond) andthe second one by heat supplied from the boiler or engine (c). Themaximum power that can be recovered from the condenser is theminimum between:

PMAXS;HC;rec;cond ¼ min

�Wel

�EERRGmain

þ 1�;mair;HCcpair

��Tw;out;cond � DTpinch;HC � Tair;in;HC

�� (8)

If the term mair,HCcpair (Tair,out,HC � Tair,in,HC) is lower thanPMAXS;HC;rec;cond, the heating coil is entirely fed by the heat rejected from

the condenser and therefore the thermal power supplied to the airstream is:

PS;HC;rec ¼ PS;HC;rec;cond ¼ mw;HC;rec;condcpw� �Tw;in;HC;rec;cond � Tw;out;HC;rec;cond

� (9a)

Otherwise the thermal power is integrated by the boiler orengine and therefore it is:

PS;HC;rec ¼ PS;HC þ PMAXS;HC;rec;cond ¼ mw;HCcpw

� �Tw;in;HC � Tw;out;HC�þ PMAX

S;HC;rec;cond

(9b)

In both cases it is:

PS;HC;rec ¼ mair;HC;reccpair�Tair;out;HC;rec � Tair;in;HC;rec

�(9c)

3.8. Sensible heat wheel

The efficiency and pressure drop of the sensible heat wheel areboth assumed constant and in agreement with experimental dataavailable from a manufacturer [26]. The sensible efficiency of thecomponent is defined in this way:

εS;HW ¼ mair;regcpair�Tair;out;reg � Tair;in;reg

�ðmaircpairÞmin

�Tair;in;reg � Tair;in;ex

� (10)

3.9. Defrosting apparatus

If the evaporation temperature is below the air dew point andbelow zero Celsius degrees, frost formation can occur on the coilsurface. For this reason a defrosting apparatus should be intro-duced. In the present work only an electric defrosting system isconsidered and the electric power required to remove the frost is(frost sub-cooling is neglected):

Pel;defrost ¼ Qlls�lhdefrostll�v

(11)

3.10. Filter

Filters d and e are simply characterized by a constant pressuredrop in the air circuit. The increase of their pressure drop due todust accumulation is not taken into account.

3.11. Fan

The fan is modelled through a constant efficiency device and itselectric power consumption is calculated in this way:

Pel;fan ¼ mairDprairhfan

(12)

where Dp is the air pressure drop in the circuit fed by the fan andhfan is its efficiency.

It is assumed the fan works all the time, except during thedefrosting period.

4. Energy balances

Each AHU configuration is simulated through equations re-ported in paragraph 3 and input data summarized in Tables 2 and 3.Equations used to calculate power balances are presented below.

The overall latent and sensible ambient loads are:

QL ¼Xi

qL;i (13)

QS ¼Xi

qS;i (14)

where qL,i and qS,i are respectively the latent and sensible contri-bution of each heat source. In the analysis discussed in section 5 ofthis work, QL is assumed constant and QS is varied.

If there is frost formation on cooling coil fins (case 4), in order totake into account also the liquid to solid water phase change, anadditional latent load term Q 0

L is introduced:

Q 0L ¼

Xi

qL;i

�ls�lll�v

�(15)

S. De Antonellis et al. / Energy 37 (2012) 336e345 341

On the other side, the sensible ambient load increases due to thepresence of the defrost apparatus (assumed electric) and thefollowing sensible load Q 0

S is added:

Q 0S ¼

1

hdefrost� 1

!ls�lll�v

Xi

qL;i (16)

Referring to Table 1 and Fig. 2, the generic form of powerbalances is reported below. The latent heat balance of the supply airstream is:

PL;a þ PL;dw;pro þ PL;l þ PL;p þ QL þ Q 0L ¼ 0 (17)

and the sensible heat one is:

PS;a þ PS;b þ PS;c þ PS;c;rec;cond þ PS;dw;pro þ PS;h þ PS;l þ PS;nþ PS;p þ QS þ Q 0

S ¼ 0 ð18ÞIf there is not frost formation on evaporator, the terms Q 0

L and Q 0S

are assumed equal to zero. Thermal load related to fans work isneglected.

The electric power consumption due to the compressor and fanwork are calculated through the following equations:

Pel;comp ¼ jPL;a þ PS;ajEERRGMAIN;S;L

þ jPS;bjEERRGMAIN;S

þ jPS;hjEERRGD

þ jPL;l þ PS;ljEERRGE

þ jPL;p þ PS;pjEERRGF

(19)

Pel;fan;pro ¼mpro

�DpdþDpdw;proþDphþDpiþDpnþDppþ2Dpdw0

�rairhfan

(20)

Pel;fan;reg ¼mreg

�Dpe þ Dpf þ Dpdw;reg þ 2Dpg þ Dpi

�rairhfan

(21)

Therefore the overall electric power consumption is:

Pel ¼ Pel;comp þ Pel;defrost þ Pel;fan;pro þ Pel;fan;reg (22)

Electric power consumption due to pressure drop across maincoils (a, b, c and crec) and ducts is not considered because it iscommon to all the analyzed configurations.

The thermal power consumption necessary to heat the regen-eration air stream and the supply air is:

Pth ¼ PS;c þ PS;f (23)

Finally the power consumption referred to the primary source,in case of separated thermal and electric power production, iscalculated in this form:

Ppr;sep ¼ Pthhth;sep

þ Pelhel;sep

(24)

and in case of cogenerative power production, it becomes:

Ppr;cog ¼ Pthhth;cog

þ Pel � Pth hel;cog=hth;coghel;sep

(25)

It is assumed that the cogenerator (an internal combustionengine) is controlled in order to provide the required thermalpower Pth. If the electric power contextually produced by theengine is lower than Pel, the difference is supplied by the electricgrid; vice versa, the surplus is sent to the grid.

5. Simulation results

All the introduced equations have been implemented in Matlab.The different HVAC configurations (A, Arec, B, C, D, E, and F)described in the previous paragraph are compared in terms ofthermal, electric and primary power consumption. The comparisonis carried out for different ambient and outside air conditions, assummarized in Table 2.

Preliminary simulations have been run for each analyzed systemin order to identify the optimal evaporation, condensation andregeneration temperature which minimizes primary powerconsumption, as reported in Table 3. A simplified sensitivity anal-ysis is carried out in order to evaluate the influence of the mostcritical parameters, as discussed in the followings:

- The variation of the evaporation temperature slightly leads tohigher primary power consumption (in most cases if thetemperature increases/reduces by 3 �C, around 5% variation ofpower consumption occurs);

- The lower is the regeneration temperature, the lower is theprimary power consumption and the higher is the desiccantwheel diameter (although a minimum temperature is neces-sary to achieve the regeneration process). To avoid non senseconfigurations characterized by extremely large wheels, theregeneration temperature has been selected in order to havea reduction of the process air humidity ratio across thecomponent at least equal to 3.5 g/kg [24]. Furthermore, if theregeneration temperature changes by 10 �C, the primary powerconsumption varies around 5e10%, depending on the config-uration and the Qs/Ql ratio;

- Heat exchangers pinch points and conductive resistances areset in agreement with typical values of real applications [1];

- Condensation temperature is directly determined from thepinch point and from the air temperature or the desiccantwheel regeneration temperature.

In the first part of the study thermal, electric and both separatedand cogenerative primary power consumption of investigatedsolutions are evaluated in case 1 conditions. Then it is analyzed theeffect of different outside conditions (case 2 and 3) and ambientconditions (case 4 and 5) on primary power consumption of themost efficient configurations. All the input data used in the simu-lations are summarized in Table 3.

In this work sensible and latent ambient loads are assumedpositive respectively when it is required cooling and dehumidifi-cation of the supply air stream.

In Figs. 3 and 4 it is shown the electric and thermal powerconsumption of each system for different sensible to latent loadratios (Qs/Ql) at constant dehumidification load (Ql ¼ 1 kW). It ispossible to state that:

- in the reference configuration A, the thermal powerconsumption Pth is only related to the heating coil (c). WhenQs/Ql increases the power exchanged by the heating coildecreases until it is switched off at Qs/Ql ¼ 2. At the sametime the electric power consumption Pel keeps constantbecause it is related only to the latent load which does notvary. When Qs/Ql is higher than 2 the cooling capacity, andtherefore the electric power consumption, increases in orderto satisfy the sensible load. In this condition it is assumedthat the latent load is not controlled and a higher dehumid-ification occurs.

- in configuration Arec, electric power consumption is higherthan the reference one (A) due to the higher condensationtemperature and, therefore, the lower EER. In addition, the

Fig. 3. Case 1 (Ta ¼ 15 �C, 4a ¼ 75%, Te ¼ 25 �C, 4e ¼ 50%): electric power required bythe HVAC systems for different Qs/Ql ratios.

S. De Antonellis et al. / Energy 37 (2012) 336e345342

thermal power consumption is lower due to the heat recoveredfrom the condenser.

- in configuration B, Pth is constant when Qs/Ql is higher than �1because it is related only to the regeneration coil f. If Qs/Ql islower than�2, Pth increases due to the use of the heating coil c.Pel is constant whenQs/Ql is lower than�2 and it increases withQs/Ql due to the use of the compressor.

- configuration C behaves as configuration B: in this case Pth isalways lower than the previous one because a heat recoverysystem across the regeneration air stream is introduced. On theother side, due to the increase of pressure drop in the circuit,electric power consumption is slightly higher.

- in configuration D, Pth is only related to the heating coil c: whenQs/Ql is higher than 0 no thermal power for air post heating isrequired. In the same way Pel is constant when Qs/Ql is lowerthan 0 and then it increases as the cooling coil b is switched ondue to the cooling load.

- configuration E behaves in a similar way of configuration D: inthis case the evaporator cools and partially dehumidifies theprocess air before crossing the desiccant wheel. Compared toconfiguration D, the dehumidification capacity increasesbecause the process air relative humidity at the componentinlet is higher [17]; therefore the process air flow decreases andboth Pth and Pel reduce.

Fig. 4. Case 1 (Ta ¼ 15 �C, 4a ¼ 75%, Te ¼ 25 �C, 4e ¼ 50%): thermal power required bythe HVAC systems for different Qs/Ql ratios.

- in configuration F, when Qs/Ql is higher than �2, Pth is equal tozero because the heating coil c does not work. Instead Pelincreases because the main evaporator b is switched on.

In Figs. 5 and 6 primary power consumption is shown in case ofseparated and cogenerative power production. It is put in evidencethat:

- the reference configuration A is efficient only when Qs/Ql ishigher than 2. In the other conditions the configurations basedon the desiccant wheel are more efficient.

- configuration C, compared to configuration B, presents similarenergy consumption: the integration of an extra heatexchanger in the AHU is not relevant and, therefore, suggested.

- configuration Arec, quite obviously, is more efficient thanconfiguration A at negative values of the Qs/Ql ratio.

- configuration B and C are particularly efficient when Qs/Ql < 0in case of cogenerative power production.

- configuration E is very interesting in a wide range of the Qs/Ql

ratio investigated.- when thermal power is required, the cogenerative powerproduction, compared to the separated one, leads to relevantenergy savings.

In Fig. 7 the best solutions are compared to the reference system(configuration A e separated power production) in terms ofprimary energy savings (PES), defined as:

PES ¼ Ppr � Ppr;sep;APpr;sep;A

$100% (26)

where Ppr is the primary power consumption of the consideredconfiguration.

The most efficient configurations that have been analyzed arerespectively:

- configuration E: themost efficient configuration approximatelyin the entire investigated range of Qs/Ql ratio (Figs. 5 and 6).

- configuration B: the best one in case of cogenerative powerproduction when Qs/Ql < �1. Note that this solution ispreferred to configuration C due to its simplicity, despite itsslightly higher power consumption.

- configuration D: this solution is less efficient than the previousones, but in this case frost formation on evaporators coil can’t

Fig. 5. Case 1 (Ta ¼ 15 �C, 4a ¼ 75%, Te ¼ 25 �C, 4e ¼ 50%): primary power required bythe HVAC systems in case of separated power production for different Qs/Ql ratios.

Fig. 6. Case 1 (Ta ¼ 15 �C, 4a ¼ 75%, Te ¼ 25 �C, 4e ¼ 50%): primary power required bythe HVAC systems in case of cogenerative power production for different Qs/Ql ratios.

Fig. 8. Case 2 (Ta ¼ 15 �C, 4a ¼ 75%, Te ¼ 15 �C, 4e ¼ 50%): primary power required bythe best HVAC systems for different Qs/Ql ratios.

S. De Antonellis et al. / Energy 37 (2012) 336e345 343

occur and therefore it is suitable for low temperatureapplications.

It is possible to state the following considerations:

- configuration A e separated power production: it is particu-larly efficient for positive values of Qs/Ql, when the heating coilis switched off. In this working condition the cogenerativepower production does not give advantages because the HVACsystem does not require heat.

- configuration B e cogenerative power production: it is thesimplest desiccant wheel based configuration and it leads toa significant reduction of primary power at negative values ofQs/Ql. Due to the relevant thermal power consumption, thecogenerative asset is preferred to the separated one.

- configuration D e separated power production: it is efficientwhen Qs/Ql is around zero. In this condition the electric powerconsumption is dominant and therefore cogeneration is notnecessary.

- configuration E: this configuration keeps great performance ina wide range of sensible to latent load ratio, and in particularwhen Qs/Ql < 0.5. It is the best desiccant wheel based AHU.

Fig. 7. Case 1 (Ta ¼ 15 �C, 4a ¼ 75%, Te ¼ 25 �C, 4e ¼ 50%): optimal HVAC systemconfigurations for different Qs/Ql ratios and Primary Energy Savings Ratio (PES)compared to the reference solution.

Finally outside and ambient conditions are varied, accordingto Table 2, and HVAC systems primary power consumption isevaluated and compared. Thus the effect of boundary conditionsis not negligible, it is possible to state they are not relevant inorder to determine the optimal HVAC system for each Qs/Ql ratio,as already shown from Bourdoukan et al. [13] in case of recircu-lation mode.

The effect of outside air conditions variation is shown in Figs. 8and 9. It should be considered that, in all the analyzed configura-tions, the higher is the outside air temperature, the lower is the EERof the refrigerating groups RGmain,S,L and RGmain,S due to the increaseof the condensation temperature. For this reason electric powerconsumption of the main refrigerating group is higher in case 3conditions and lower in case 2 conditions. In addition outsidetemperature and humidity ratio variation slightly affect thermalpower supplied to the regeneration air stream.

Similar considerations can be drown when ambient conditionsare varied. Although the aim of the analysis is to evaluate howambient temperature and humidity ratio variation affects primarypower consumption, it should be remarked that case 4 and 5 stillrepresent real operating conditions (i.e. fish or raw meat drying incase 4, automotive paint booth in case 5). As shown in case 4(Fig. 10), the lower is the ambient temperature, the lower is theevaporating pressure of the refrigerating groups and therefore their

Fig. 9. Case 3 (Ta ¼ 15 �C, 4a ¼ 75%, Te ¼ 35 �C, 4e ¼ 50%): primary power required bythe best HVAC systems for different Qs/Ql ratios.

Fig. 10. Case 4 (Ta ¼ 2 �C, 4a ¼ 75%, Te ¼ 25 �C, 4e ¼ 50%): primary power required bythe best HVAC systems for different Qs/Ql ratios.

S. De Antonellis et al. / Energy 37 (2012) 336e345344

EER; vice versa in case 5 (Fig. 11). In case 4, primary powerconsumption Ppr of the reference configuration A is strongly higherbecause of the electric power consumption required by thedefrosting apparatus. For the same reason, Ppr of configuration E isnot investigated: this solution is not suitable for low temperatureapplications and arrangement B and D should be preferred.

In conclusion it is possible to state that the variation of ambientand outside air conditions within the range resumed in Table 2,slightly influence primary power consumption, except for config-uration A e case 4 due to frost formation on the cooling coil.

In addition it should be underlined that, in many real applica-tions, sensible and latent loads do not vary too much in the year, forexamplewhen loads due to product drying are strongly higher thanheat losses through the room envelope and due to air infiltrations.In these cases, the above considerations can be useful to prelimi-nary assess seasonal performance of the system.

Finally, it is well known that the analyzed configurations cannotbe evaluated only in terms of energy efficiency. Industry manage-ment mainly uses an economic approach in order to evaluate thebest solution, considering in particular the payback time and thenet present value of the investment. Therefore this study could becompleted with a further research activity that handles a detailedeconomic analysis of the investigated configurations.

Fig. 11. Case 5 (Ta ¼ 25 �C, 4a ¼ 50%, Te ¼ 25 �C, 4e ¼ 50%): primary power required bythe best HVAC systems for different Qs/Ql ratios.

6. Conclusions

In this work it is shown that desiccant wheel based air handlingunits for drying processes can lead to relevant primary energysavings compared to the reference technology based on a coolingcoil. The optimal configuration depends on the ambient Qs/Ql ratioand slightly on ambient and outdoor conditions. Quite obviously, thereference system based on a cooling coil is efficient only if contem-porary cooling and dehumidification is required (Qs/Ql > 0) and, inparticular, it is the best choice ifQs/Ql> 2. On the other side themostcommon desiccantwheel configuration (B) is recommendablewhenQs/Ql is lower than �1. If the sensible to latent ambient load ratio iswithin�2 and 2, otherHVAC systems based on a desiccantwheel areparticularly efficient. In these conditions primary energy savings canreach 70e80%, compared to the reference technology. Therefore theQs/Ql ratio is a relevant criterion that should be considered in order toselect the optimal HVAC system arrangement.

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Nomenclature

A: desiccant wheel cross sectional area for air flow [m2]AHU: air handling unitBPF: by pass factor of the cooling coil [e]cp: isobaric specific heat [J/(kg K)]EER: energy efficiency ratio [e]L: desiccant wheel length [m]m: mass flow [kg/s]MRC: moisture removal capacity [gv/s]N: wheel rotational speed [rev/h]P: average power [W]Dp: pressure drop [Pa]Qreg: heat of regeneration [kW]Qs: ambient sensible load [kW]Ql: ambient latent load [kW]Re: Reynolds number [e]T: air temperature [K]TADP: apparatus dew point [K]DTres: temperature difference due to conductive resistance [K]v: face air velocity on the wheel face [m/s]X: air humidity ratio [kgH2O/kgDA]

Greek lettersεS: sensible heat wheel efficiency [e]ll-v: water latent heat of vaporization [kJ/kg]ls-l: water latent heat of solidification [kJ/kg]h: efficiency [e]r: density [kg/m3]

Subscriptsa, b, ..: component a, b, ..A, B ..: configuration A, B, ..a: ambientad,w: adsorbed watercog: cogenerative power productioncond: condensationcomp: compressorDA: dry airdefrost: defrostdes: desiccantdw: desiccant wheel e unbalanced flowsdw0: desiccant wheel e balanced flowse: external (outside)el: electriceva: evaporation or evaporatorex: exhaust airfilter: filterfan: fanHC: heating coili: generic heat sourcein: inletl: liquid waterL: latent heatout: outletpro: process airpr: power referred to primary sourcerec: recoveredref: referencereg: regeneration airS: sensible heatsep: separated power productionth: thermaltot: totalv: water vapourw: water