Review HVAC dehumidification systems for thermal comfort: a critical review Pietro Mazzei * , Francesco Minichiello, Daniele Palma DETEC, University of Naples ‘‘Federico II’’, P.le Tecchio 80, 80125 Napoli, Italy Received 29 November 2003; accepted 25 July 2004 Abstract This paper, on the basis of the main literature indications, deals with moisture control in buildings dur- ing the summer season; so, the dehumidification of the air is analysed. Dehumidification is considered as a key feature of HVAC systems for thermal comfort. Initially, the principles of mechanical and chemical dehumidification are shown. The first one utilises mechanical means—compression refrigeration sys- tems—to cool the air and so to dehumidify it; the latter removes the water vapour from the air by trans- ferring it towards a desiccant material (adsorption or absorption). In the mechanical dehumidification field, a proper control of ambient temperature and humidity can be obtained by means of an air handling unit (AHU) which treats outside air alone, while recirculating air is treated by a simple cooling coil. Various possible AHU configurations are examined. Afterwards, HVAC systems for a theatre and for a supermar- ket are analysed. The use of hybrid systems with desiccant wheel for these applications has provided the following main results: remarkable savings in operating costs and higher plant costs (a simple payback time of 2–3 years for supermarket); a notable reduction of the power electric demand; a better control of ambient humidity. Ó 2004 Elsevier Ltd. All rights reserved. Keywords: HVAC systems; Dehumidification; Chemical desiccants 1359-4311/$ - see front matter Ó 2004 Elsevier Ltd. All rights reserved. doi:10.1016/j.applthermaleng.2004.07.014 * Corresponding author. Tel.: +39 081 7682301; fax: +39 081 2390364. E-mail addresses: [email protected](P. Mazzei), [email protected](F. Minichiello). www.elsevier.com/locate/apthermeng Applied Thermal Engineering xxx (2004) xxx–xxx ARTICLE IN PRESS
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ARTICLE IN PRESS
www.elsevier.com/locate/apthermeng
Applied Thermal Engineering xxx (2004) xxx–xxx
Review
HVAC dehumidification systems for thermal comfort:a critical review
Pietro Mazzei *, Francesco Minichiello, Daniele Palma
DETEC, University of Naples ‘‘Federico II’’, P.le Tecchio 80, 80125 Napoli, Italy
Received 29 November 2003; accepted 25 July 2004
Abstract
This paper, on the basis of the main literature indications, deals with moisture control in buildings dur-
ing the summer season; so, the dehumidification of the air is analysed. Dehumidification is considered as akey feature of HVAC systems for thermal comfort. Initially, the principles of mechanical and chemical
dehumidification are shown. The first one utilises mechanical means—compression refrigeration sys-
tems—to cool the air and so to dehumidify it; the latter removes the water vapour from the air by trans-
ferring it towards a desiccant material (adsorption or absorption). In the mechanical dehumidification field,
a proper control of ambient temperature and humidity can be obtained by means of an air handling unit
(AHU) which treats outside air alone, while recirculating air is treated by a simple cooling coil. Various
possible AHU configurations are examined. Afterwards, HVAC systems for a theatre and for a supermar-
ket are analysed. The use of hybrid systems with desiccant wheel for these applications has provided thefollowing main results: remarkable savings in operating costs and higher plant costs (a simple payback time
of 2–3 years for supermarket); a notable reduction of the power electric demand; a better control of ambient
humidity.
� 2004 Elsevier Ltd. All rights reserved.
Keywords: HVAC systems; Dehumidification; Chemical desiccants
1359-4311/$ - see front matter � 2004 Elsevier Ltd. All rights reserved.doi:10.1016/j.applthermaleng.2004.07.014
Why dehumidify? Narrowing the field down to the air-conditioning of buildings and assumingthat the aim is to maintain comfort conditions, it is clear that the air to be supplied (for simplic-ity one can refer to the supply of the outdoor air flow alone in the minimum quantity, that is theamount indispensable to guarantee the requirements of air quality) should have a lower humid-ity ratio 1 compared with that of indoor air so as to balance indoor moisture production (prin-cipally due to the occupants� metabolism and thus to the main activity in which they areengaged). However, in summer the humidity ratio of outside air is normally higher than the in-door humidity ratio: and thus the need to dehumidify (see Appendix). The design of HVAC sys-tems for thermal comfort requires increasing attention, especially in the light of recentregulations/standardisation on ventilation [1–8], so that an optimal level of indoor humiditymay be reached and maintained to ensure a comfortable and healthy environment 2 and to avoidcondensation damage for building envelope and furnishings. Since the paper focuses on dehu-midification issues, the fundamental concepts referred to mechanical and chemical dehumidifica-tion are briefly exposed.
umidity ratio of a given moist air sample is defined as the ratio of the mass of the water vapour to the mass of dry
ntained in the sample.
SHRAE Standard 62–2001 [3] addresses the need to control indoor humidity: ‘‘relative humidity in habitable
s preferably should be maintained between 30% and 60% relative humidity to minimize growth of allergenic and
genic organisms’’.
Nomenclature
cp constant pressure specific heat, J/(kg K)Fbp by-pass factor, –h specific enthalpy of the humid air, J/kgDhlv vaporization specific enthalpy, J/kg_m air mass flow rate, kg/sMCDB mean coincident dry bulb, �CMCWB mean coincident wet bulb, �CN number of occupants, –O.A. outdoor air fraction in the mixed air flow, –p pressure, Paq specific thermal power, W/(kg/s)_Q thermal power or thermal load, WSHR sensible to total Heat load Ratio, –T temperature, �CDT temperature difference, �CTdb = DB dry bulb temperature, �CTwb wet bulb temperature, �CTdp = DP dew point temperature, �CTsur cooling coil surface mean temperature, �C_V air flow rate, m3/sv velocity, m/sD difference between values, –/ relative humidity of the humid air, –�q volumic mass (density) mean value, kg/m3
x specific humidity (humidity ratio), kgv/kga (or gv/kga)A state of the air leaving cooling coilAHU air handling unitCC cooling coilCOP coefficient of performanceCVAAS constant volume all air systemDBTCS control system based on dry bulb temperatureDC dehumidificator capacityDEC direct evaporative coolingDW desiccant wheelHC heating coilHTX heat exchangerHVAC heating ventilating air conditioningIAQ indoor air qualityIEC indirect evaporative coolingNC normally closed (damper)NO normally opened (damper)
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PID proportional, integral, differential controllerRHCS control system based on ‘‘re-heating’’VAV variable air volume
Subscriptsa referred to dry aircc referred to cooling coilDE referred to dehumidificatorhc referred to heating coilin referred to inlet conditionsLAT referred to latent loadm referred to mixed airnew referred to a new valueo referred to outdoor airold referred to an old valueout referred to outlet conditionsr referred to indoor/recirculation airs referred to supply airsat referred to saturated humid airSEN referred to sensible loadT referred to total loadu unitaryv referred to humid air vapourvent referred to ventilation air or ventilation loadw referred to waterz referred to zone1 referred to state 1
Superscriptd referred to design conditions
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1.1. Mechanical dehumidification
Mechanical dehumidification is widely known to be based on cooling [9]. The humidity of theair to be treated is characterised by dew point temperature, Tdp,1, Fig. 1. If the air comes into con-tact with a cooling coil whose mean surface temperature, Tsur, is less than Tdp,1, condensation of apart of the water vapour, and thus dehumidification, is obtained. With reference to Fig. 1, the uni-tary thermal power to be subtracted (ignoring condensate energy) is taken as:
For the cooling coil the following parameter is really important:
SHRcc ¼ qcc;SEN=qcc;T: ð2Þ
dry bulb temperature
2
SHR
3
1
cc
qcc,T
cc,SENq
qcc,LAT
1h
h2
3h
hum
idit
y ra
tio
Fig. 1. Mechanical dehumidification: cooling and dehumidification.
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Its maximum value is one (the cooling coil has no dehumidifying effect: only sensible cooling),the minimum is zero (the cooling coil has no cooling effect: only dehumidification) [10].For example, referring to an all-air system with constant flow rate serving one zone (CVAAS),
the calculation of thermal loads for design conditions leads to the evaluation of the parameter:
3 T4 A
rows (5 In6 T
high d
315–27 T
follow
SHRdz ¼ qz;SEN=qz;T: ð3Þ
These loads can be balanced according to the two processes reported in Fig. 2. In process (a) thecooling coil is placed downstream of the point where outdoor and recirculated airflow rates mix;in process (b) mixing of outdoor and recirculated airflow occurs, if necessary, after the outdoorairflow has been dehumidified. From Fig. 3 it may be deduced that the values for parameterSHRcc of the dehumidification coils employed in the two processes differ significantly: that ofthe cooling coil dedicated to outdoor air flow is much smaller, so that an ‘‘ad hoc’’ design wouldprobably be necessary. It should be remembered that to obtain low values of SHRcc cooling coilsare needed which may be: fed by refrigerants at low temperature and/or at high velocity, 3 and/orwith a high number of fins, 4 and/or with moderate air speed, 5 and/or with a not too excessive findensity 6 [11]. To increase dehumidification capacity it would be better to reduce the by-pass fac-tor 7 by increasing the depth of the cooling coil rather than fin density.
1.2. Chemical dehumidification
Desiccants are materials with a high affinity for water vapour and may be solid or liquid.Adsorption is when the physical or chemical nature of the desiccant, generally solid, remains un-
ypical values for refrigerated water cooling coils are Tin = 5–7 �C, Tout = 10–12�C, v = 1.2m/s.s depth increases (in the direction of flow) air moves closer to saturation. Depth increases with the number of
common values of number of rows are between 2 and 8).
the comfort field typical values are in the range 2.0–2.5m/s.
he density is often indicated as fins per inch (fpi): common values are between 8 and 14 fpi (315–551 fins/m). Too
ensity can obstruct the condensation: suggested values [11] to maximise the dehumidification are in the range
36 fins/m.
aking ‘‘sat’’ as the state of saturated humid air at mean coil temperature, the by-pass factor may be expressed as
s: Fbp = (hout�hsat)/(hin�hsat).
coil
Cooling/
dehumidificationcoil
Return air
Outside air Supply air
Re-heating
o m A
r - +
(a) dehumidification downstream of the mixing chamber
Supply airOutside air
Return air
coilRe-heating
o
r+-
A
coildehumidification
Cooling/
(b) dehumidification of the outdoor air before the mixing chamber
Fig. 2. Mechanical dehumidification: different positions of the cooling coil for an all air system.
r
A
m
o
hum
idit
y ra
tio
dry bulb temperature
Fig. 3. Mechanical dehumidification: cooling and dehumidification in the two cases of Fig. 2 (re-heating coil not
enabled).
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changed in the dehumidification process; absorption conversely is when a change occurs, generallywith liquid substances [12,13].A common adsorption solid is silica gel, which behaves like a sponge. In fact, its structure is
extremely porous; its internal surface per volume unit is immense, approximately 250m2/cm3.Its pores have a diameter of nanometers and their volume accounts for approximately half ofthe total volume. Attracting forces between vapour and solid depend on the particular solid–va-pour pair and on physical structure of the solid. An initial layer of adsorbed molecules is formedand the force of attraction decreases as surface density of adsorbed molecules increases. This proc-ess grows as the number of layers increases. As with normal liquefaction, thermal energy is re-leased also for adsorption; the two values are of the same size. The amount adsorbed atequilibrium conditions may be expressed as mass or (normal) volume of gas per unit of massof the solid without gas. Adsorption desiccants are typically chemical compounds, such as syn-thetic polymers, silica gels, titanium silicates, natural or synthetic zeolites, activated aluminas,‘‘silica +’’, etc. [9,14–18].Common absorbents are various solutions of water and ethylene glycol, LiCl, LiBr,
CaCl2.In the context under examination regenerative systems are being dealt with, that is those for
which the mechanism of moisture removal is continuous.Air to be treated before supplying in indoor ambient is called ‘‘process air’’. Chemical dehumid-
ification [9,12,19,20] is based on the migration of water vapour from process air towards the sur-face of the desiccant due to the difference in partial vapour pressure (the value of pv is greater inhumid air). It may be observed that the pressure gradient is orientated in this direction because thedesiccant is dry and cool; should the material become warm and moist the pressure gradient isinverted and water vapour migrates from the desiccant to humid air. The typical cycle of the des-iccant is made up by three steps, Fig. 4.
• A–B. In A the desiccant is cold and dry; removing water from the process air, the surface pvgrows reaching the pv value of the surrounding air; the equality is reached in B state: the migra-tion of the vapor stops.
• B–C. The desiccant is removed from the process air, heated and exposed to a different air flow(regeneration air, which is then discharged in atmosphere): the gradient changes its directionand the migration of the water vapour occurs from the desiccant towards the air current. Instate C the humidity content has the starting value (A), but the pv is much greater becauseof the high temperature reached by the material.
• C–A. The material is cooled until the starting temperature. The values of humidity content andpv are restored. The cycle can be repeated.
From the description of the cycle it may be deduced that the cyclic process A–B–C requiresthermal power: heating between B–C (regeneration or reactivation) and cooling between C–A.It can be noted that the typical desiccant cycle made up by three steps, shown in Fig. 4 and pre-viously described, is currently reported as representative of the phenomenon in literature refer-ences [9,12], but it could be represented also as made up by four steps, since desorption stepB–C can be subdivided, for some applications (see the absorbtion system shown in Fig. 5), intwo separated stages: heating and properly said desorption.
Dry air
HEATEXCHANGER
Processair
Coolingfluid
CONDITIONER REGENERATOR
fluidHeating
Wet air
airRegeneration
HEATEXCHANGER
Fig. 5. Absorption dehumidification system (liquid desiccant).
Desiccant temperature
Adsorption/Des
icca
nt s
urfa
ce v
apor
pre
ssur
e
Desiccant cooling
Desiccant moisture content
A Absorption
CDesorption
B
(Desiccantregeneration)
Fig. 4. Chemical dehumidification: typical desiccant cycle.
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For the regeneration of liquid absorbents either low pressure vapour or warm liquid water isused; for solid adsorbents thermal energy is used made available generally by combustion (director indirect gas-fired heaters). In both cases, if possible, waste heat may be used (heated water fromsolar panels, from condensers of traditional refrigerating machines, from cogenerators, fromunder-used heaters, etc.).With reference to Fig. 5, a typical apparatus for the dehumidification of an air flow based on
absorption is described. The liquid desiccant is distributed on fill material similar to that used in
0
2
4
6
8
10
12
14
16
18
20
-10 -5 0 5 10 15 20 25
Water saturation curve
20% LiCl
30% LiCl
40% LiCl
Dry bulb temperature [˚C]
Equ
ilibr
ium
hum
idity
ratio
[g/
kg]
Fig. 6. Absorption dehumidification: equilibrium chart for a water–lithium chloride solution.
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cooling towers, whose function is to increase the contact surface area where air and desiccant meetin order to improve thermal and mass transfer. Process air flows countercurrently to liquid des-iccant. Beside the air conditioning tower, the heat exchanger cools the solution to remove absorp-tion heat and any other thermal input towards the solution. A second tower is present forregeneration. As the solution absorbs the vapour it dilutes and thus the pv of the water increases.Fig. 6 [9] shows the equilibrium diagram, at standard atmospheric pressure, for a water–lithiumchloride solution, in a format similar to that of the psychrometric chart. It shows, for each tem-perature, the drop in equilibrium specific humidity as the concentration rises, and, for each con-centration value, a rise in equilibrium specific humidity as temperature rises. Hence the necessityto re-concentrate the solution. This occurs in the regeneration tower: the solution is heated by theheat exchanger beside the tower in order to increase the pv of the water and force the vapour tomigrate from the solution towards an air current, which is then discharged in the atmosphere. Aclosed circuit continuously recirculates the solution between the drain pans of the two towers.When temperature and concentration of the solution are well controlled, process air leaves the
tower in the psychrometric conditions required for it to be supplied into the room. As we shall see,the same does not apply to the process of adsorption, which requires cooling after dehumidifica-tion. Various absorption dehumidification lines can be obtained, with an outlet air temperaturenearly equal than inlet, unlike mechanical and adsorption dehumidification.The pressure gradient may be inverted by diluting the solution (so humidification of the air may
be obtained).Systems of the type described have been used for large size buildings in the tertiary sector [19],
the extra costs of which are reasonably compensated by energy saving. As humidity control in thebuilding becomes established a significant growth, also on a lower scale, is expected: smaller units(some thousands of m3/h) have already entered the market.Also as regards adsorption dehumidification, as mentioned earlier, the desiccant must be peri-
odically regenerated. This technology, long employed in the industrial sector, has now provokedrenewed interest and has been extended towards the non-industrial sector, both because of more
Fig. 7. Air dehumidification by adsorption and post-cooling.
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stringent IAQ 8 requirements, and the gradual reduction of the regeneration temperature requiredby new desiccant materials.The moisture contained in humid air partially condenses in the desiccant: it is adsorbed because
of the vapour partial pressure difference between process air and desiccant surface. So the processair temperature increases because of the conversion in sensible heat of both condensation heat andheat due to the adsorption process. After all, outlet process air specific humidity decreases whiletemperature increases, Fig. 7. Before supplying into the room, process air must be cooled, Fig. 7,by means of one or more of the following components: direct expansion or chilled liquid (gener-ally water) cooling coil (CC); indirect evaporative cooling (IEC); rotary or static heat recuperator(HTX).A first type of system is the dual desiccant bed adsorption dehumidifier, Fig. 8: the air to be
dehumidified passes through a bed of granular desiccant and is dehumidified; as the desiccantmust be periodically regenerated with the introduction of warm air to make the process continu-ous a second bed is necessary to substitute the first during regeneration. The functions of the twodesiccant beds are alternated simply by valves, as shown in Fig. 8. The equipment is simple thoughrather bulky. 9 Moreover the conditions of the dehumidified air at the outlet are less and less uni-form as the desiccant approaches saturation point.These drawbacks have been overcome by a system with solid desiccant inserted in a rotating
exchanger [19–22], called ‘‘desiccant wheel’’ (DW), more common than the previous system, espe-cially in the non-industrial air-conditioning sector. In Fig. 9 a typical desiccant 10 wheel is shown[13]. The structure, very similar to that of an enthalpy wheel, is usually realized by wrapping anelement made up of a corrugated lamina and a plane. The lamina is fibrous and impregnated 11
8 Comparing the outdoor air requirements of the two versions of the ASHRAE Standard (62–1981 and 62–2001)
‘‘Ventilation for Acceptable IAQ’’ [2,3], it can be observed an increase of 2–4 times; for example: (units in l/s per person)
from 7.5 to 15 for hotel rooms, from 3.5 to 8 for auditorium, from 3.5 to 10 for meeting rooms, from 2.5 to 10 for
offices.9 To avoid the pulverisation of the desiccant and a not uniform distribution of the air, the air speed must be
moderate. So, this type of system is not able to handle high flow rates, as usually required in typical HVAC
applications.10 Typically the thickness varies between 5 and 45cm, the diameter between 10 and 420cm.11 Stronger the connection between substratum and desiccant, less the discharge of desiccant pieces in process air; so,
the pollution of the process air and the reduction of the wheel performance are avoided.
Fig. 9. Typical desiccant wheel.
+
Inlet regeneration air
Exitprocess air
Heatingcoil
Coolingcoil
Packedtower
regeneration airExit
process airInlet
Adsorption Desorption
towerPacked
Fig. 8. Packed tower adsorption system.
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with desiccant. Air passes through the channels that are formed between the lamina layers, ar-ranged parallel to the wheel axis. The device rotates slowly (6–20rev/h) [23] between the twoair fluxes: moisture is removed from process air through the desiccant; after a partial rotation,the portion of saturated wheel is regenerated by warm, dry air (‘‘regeneration air’’) so it may
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be reused. A fixed element with a flexible gasket separates the two airflows. Occasionally a smallflow (purge) is removed from process air and cools the regenerated desiccant, after which it ismixed with incoming regeneration air.
2. HVAC dehumidification systems
2.1. HVAC systems with mechanical dehumidification
The importance of the parameter SHRcc (2) for the cooling coils can be stressed by consideringthe ventilation load:
12 T
_Qvent ¼ _moðho � hrÞ: ð4Þ
Considering a unitary volumetric airflow rate (4) may be taken as
_Qvent;u ¼ �qoðho � hrÞ: ð5Þ
Fixing the indoor conditions Tdb = 25�C and / = 50%, in Fig. 10, for six European sites and fortwo different outdoor design conditions, the unitary summer ventilation loads, subdivided intosensible and latent components, are reported. It may be observed that the latent component isvery often the greater, particularly when peak dew point temperature data (ASHRAE 1% DP-MCDB) are employed, with values up to 90% of the total load. What this means is that a dehu-midification coil capable of dealing with such loads would need a particularly low values ofparameter SHRcc (2), which seldom occurs.
12
Regarding the CVAAS system it is worth highlighting the problem of ambient moisture controlin partial load conditions. In actual fact HVAC plants are designed to deal with peak loads, whilefor most of the working time the loads are smaller.From the energy balance equation in steady state for a general zone, the followings are
obtained:
_Qz;SEN ¼ _macpðT db;z;r � T db;z;sÞ; ð6Þ
_Qz;LAT ¼ _maDhlvðxz;r � xz;sÞ: ð7Þ
As the plant works at a constant air flow rate, a reduction of the sensible load requires a corre-sponding increase in supply air temperature.Should the control system be based (DBTCS) on a comparison between dry bulb temperature
of the local thermal sensor and the setpoint, and on the consequent modulation of the power ofthe coil (i.e. of the temperature of supply air), as the ambient load diminishes the mean surfacetemperature of the coil increases and its dehumidification capacity drops. To understand the sizeof this problem Table 1 shows the number of hours in which outdoor temperature [24] is less thanindoor design temperature (Tdb,r = 25�C), for three Italian sites, while the opposite occurs with
ypical values for refrigerated water cooling coils are greater than 0.75.
-10
-5
0
5
10
15
20
25
30
1%D
B-M
CW
B
1% D
P-M
CD
B
1%D
B-M
CW
B
1% D
P-M
CD
B
1%D
B-M
CW
B
1% D
P-M
CD
B
1%D
B-M
CW
B
1% D
P-M
CD
B
1%D
B-M
CW
B
1% D
P-M
CD
B
1%D
B-M
CW
B
1% D
P-M
CD
B
Uni
tary
ven
tilat
ion
load
[kW
/(m³/s
)]
LatentSensible
MILAN GENOA BARCELONA
PARIS
LONDON
ATHENS
Fig. 10. Unitary summer ventilation loads for six European sites (Milan, Genoa, Barcelona, Paris, London, Athens)
and for two different outdoor design conditions.
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humidity ratio (xr = 9.9gv/kga). In Rome, for example, on a total of 2928h for the summer sea-son, this happens for the 60% of the hours.ASHRAE 1% DB-MCWB design data [9] on outside air (Table 2) show that for Rome the cal-
culation of thermal loads of the building is performed assuming Tdb,o = 30�C. Since, at the peakof dew point temperature (25 �C), the corresponding dry bulb temperature (MCDB) falls to 27�C,in these conditions a substantial reduction of the sensible thermal load of the building will occur,while latent load (due mainly to the occupants) remains unchanged. This suggests that supply airtemperature must increase and that factor SHRz diminishes. Thus what follows for the DBTCSsystem is an increase in ambient relative humidity.A control technique aimed at overcoming these obstacles is based on re-heat (RHCS). The
cooling coil is designed in such a way as to ensure that the air outflow can satisfy simultane-ously maximum sensible and latent loads in the ambient. As the sensible load diminishes anambient temperature sensor progressively activates a heating coil 13 (re-heat), which heats theair in such a way as to balance this load. The temperature sensor works on the heating coilalone, while the cooling coil is designed to balance maximum loads and is regulated to balancelatent load.Naturally this control system is not energetically efficient. With reference to the building-plant
whole, steady state energy balance may be given as
13 W
ducts
_Qcc ¼ _Qz;SEN þ _Qz;LAT þ _Qhc þ _Qvent: ð8Þ
Thus energy use related to the re-heating coil weighs heavily not only for the generation of hotthermal carrier fluid, but also for the generation of cold thermal carrier fluid. The increase in costs
ith more than one zone, instead of a single re-heating coil placed in the AHU, the number of coils (placed in the
near the rooms) should correspond to the number of zones.
Table 1
Number of hours during which To < Tr and xo > xr for three Italian sites Tdb,r = 25 �C; /r = 50%; xr = 9.9g/kg
Period: 1 June–30 September Total hours: 2 0928
x [g/kg] Tdb [�C] 13.5 14.5 15.5 16.5 17.5 18.5 19.5 20.5 21.5 22.5 23.5 N. hours
Site: Crotone
18.5 0 0 0 0 0 0 0 0 0 0 2 2
17.5 0 0 0 0 0 0 0 0 0 0 5 5
16.5 0 0 0 0 0 0 0 0 1 11 11 23
15.5 0 0 0 0 0 0 0 3 22 24 13 62
14.5 0 0 0 0 0 0 1 9 30 30 21 91
13.5 0 0 0 0 0 4 23 36 34 37 24 158
12.5 0 0 0 0 2 45 46 52 49 41 27 262
11.5 0 0 2 10 41 42 54 34 47 40 37 307
10.5 0 2 10 23 26 35 40 41 41 36 42 296
9.5 4 9 12 14 27 16 22 33 30 29 37 233
Total 1 0439Percentage on total 49.1%
Site: Rome
18.5 0 0 0 0 0 0 0 0 0 0 0 0
17.5 0 0 0 0 0 0 0 0 0 0 2 2
16.5 0 0 0 0 0 0 0 0 2 9 6 17
15.5 0 0 0 0 0 0 0 1 8 12 9 30
14.5 0 0 0 0 0 0 2 17 26 25 13 83
13.5 0 0 0 0 0 5 32 51 38 21 31 178
12.5 0 0 0 0 16 54 71 75 61 35 33 345
11.5 0 0 1 32 93 94 81 60 43 34 32 470
10.5 0 5 35 75 42 44 47 36 39 28 32 383
9.5 7 32 46 24 31 21 17 16 15 25 26 260
Total 1 0768Percentage on total 60.4%
Site: Milan
18.5 0 0 0 0 0 0 0 0 0 0 0 0
17.5 0 0 0 0 0 0 0 0 0 0 2 2
16.5 0 0 0 0 0 0 0 0 1 3 5 9
15.5 0 0 0 0 0 0 0 4 10 14 15 43
14.5 0 0 0 0 0 0 12 31 32 31 12 118
13.5 0 0 0 0 0 12 44 59 52 40 25 232
12.5 0 0 0 0 37 68 76 57 33 33 24 328
11.5 0 0 12 89 102 62 43 47 37 36 21 449
10.5 0 48 83 75 32 40 30 26 21 24 23 402
9.5 86 82 34 17 22 18 13 12 11 6 14 315
Total 1 0898Percentage on total 64.8%
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Table 2
Summer design conditions of outdoor air for some Italian sites by UNI 10339 and ASHRAE data
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does not regard operating costs alone but also investment costs (an increase in power of the refrig-erating machine, for example). It is worth pursuing energy recovery (thermal energy from conden-sation, for example) to avoid heating coil energy use.Modulation of cooling coil power may also be related to a moisture sensor in the ambient: as
long as local / is kept below setpoint value, cooling coil power is controlled by the local thermo-stat; should / exceed setpoint value, the humidistat takes over and the thermostat, if necessary,requires re-heat.It is clear that the best results would be obtained if the control parameter of the cooling coil was
the dew point temperature, rather than relative humidity. The former is actually a measure ofhumidity ratio while the latter varies, moisture content being equal, with Tdb. Any Tdb fluctuation,moisture content being equal, may lead to unreliable control actions. Dew point temperature sen-sors aside, modern automated systems enable the Tdp calculation to be made from the usual Tdband / sensors.Two techniques of energy recovery are worth citing here.
• An extremely rational and simple technique involves the use of a regenerative thermalexchange 14 (recuperator) as shown in Fig. 11. Two gas–liquid exchangers transfer heat fromair flowing into the dehumidification coil to air flowing out. The liquid employed is actuallywater, slightly glycolated if necessary, which is pumped around a closed circuit. The energy sav-ing obtained is clear: the power needed by the cooling coil is less, its dehumidifying capacityimproves, re-heat rarely requires an external energy source. Control may be exerted via thepump or a three-way valve.
• Enthalpy wheel. 15 It is a rotary exchanger (regenerator) normally made up of a cylinderfilled with a desiccant-coated aluminium matrix. 16 Return air flows across half the wheeland outside air (countercurrent) over the other half. High-speed rotation (700–2400rev/h)[23] causes thermal and mass transfer. In summer, heat and moisture pass from outsideair towards return air; thus the former undergoes initial cooling with dehumidification;the reverse occurs in winter. If the two flow rates are equal, indicating with ‘‘o’’ and ‘‘r’’the states of the two currents, outdoor air downstream of the wheel will fall, on the psychr-ometric chart, within the segment which joins ‘‘o’’ to ‘‘r’’. A wheel has an 80% latent effi-ciency when the variation of humidity ratio of outside air equals 80% of Dxor. Similarlysensible efficiency (based on DTor), enthalpic or total efficiency (based on Dhor) are defined.Regulation is simply obtained varying rotation speed (using a variable speed motor). If thealuminium matrix is bare, thermal transfer alone occurs and the wheel operates as a sensiblerecuperator.
With variable air volume (VAV) systems the comparison between the signal of the local sensorof Tdb and setpoint modulates airflow supplied at constant temperature: as sensible load
14 Currently indicated as ‘‘run-around-coil’’.15 Also called passive desiccant wheel.16 For more details on the vapour transfer it can be seen the following paragraph on the adsorption
P. Mazzei et al. / Applied Thermal Engineering xxx (2004) xxx–xxx 17
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diminishes compared with design value the airflow supplied also decreases thanks to the action ofthe VAV diffuser. In addition, the Air Handling Unit (AHU) supply fan must be modulated tomaintain static pressure constant as airflow varies.If, at partial load, the air is supplied at constant dew point temperature, it is evident that the
reduction of airflow supplied, latent load being equal, will cause an increase in local humidity ra-tio (see Eq. (7)). One solution might be to modulate the dew point temperature of the supply airusing a moisture sensor; constant dry bulb temperature of the air in the supply duct may beachieved with a re-heating coil or a regenerative exchanger.To offset the inconvenience of a possible reduction in pollutant dilution when airflow falls sig-
nificantly, it is possible to combine the system of motorised coupled dampers of the outside air/recirculating air mixing chamber of the AHU with a CO2 sensor (as an IAQ indicator). The out-side air/recirculating air ratio must be increased as supply airflow decreases due to the action ofthe VAV diffusers.Direct control on ambient Tdb and / is also possible with an AHU to treat outside air alone,
while recirculating air is treated by a simple cooling coil (without dehumidification) [23,25–28].Control of the two devices is independent. The power of the dehumidification coil, which workson outside air, is controlled by a room humidistat to keep humidity values below setpoint. Cool-ing coil power, which works on recirculation air, is controlled by a local thermostat to keep tem-perature values below setpoint.Referring to the ambient, from the water mass balance in steady state, it follows:
xs ¼ xr �_mv_ma
; ð9Þ
which provides the humidity ratio value of the air to be supplied in order to balance latent load, atask for only the AHU dedicated to outside air. In normal circumstances (moisture is largely gen-erated by the occupants alone; conditions to be maintained inside: Tdb = 25�C and / = 50%) thedew point temperature of the supply air is about 12�C (xs � 9gv/kga).In a multi-zone area xs values would be different for each zone:
17 I
flow t
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ARTICLE IN PRESS
xs;z1 ¼ xr �_mv;z1_ma;z1
; ð10Þ
xs;z2 ¼ xr �_mv;z2_ma;z2
; ð11Þ
. . .
xs;zn ¼ xr �_mv;zn_ma;zn
: ð12Þ
However, since the AHU treating the outside air is generally one, the value of xs at the entranceto the n zones must be one as well: of the n values obtained for xs, the higher value is selected(xs,z2, for example). In this way, setting this xs also at the entrance of the other zones, it is nec-essary that the other flow rates _ma to be supplied increase over and above the value initially esti-mated on the basis of air exchange needs alone. For example:
_ma;z1;new ¼ _mv;z1xr � xs;z2
> _maz1;old ¼_mv;z1
xr � xs;z1: ð13Þ
Thus, selecting maximum humidity ratio for supply air means maximising airflow rate in eachzone: if this solution was not adopted, in some zones to balance latent load the supply airflow ratewould be less than required for needs of air exchange.The results of an extensive simulation with an HVAC system based on the employment of an
AHU dedicated to outside air have been presented in [25]. Comparing the three configurations out-lined in Fig. 12, the third configuration turns out to be more profitable, as regards energy use andpower installed. The first configuration includes: a liquid regenerative exchanger placed near thecooling coil, a heating coil for heating in winter and co-operating with the exchanger in summerif necessary, a humidifier forwinter. In the second configuration, an enthalpywheel placed upstreamof the cooling coil dehumidifies and cools the outside air in summer; in winter it heats and humidifiesthe outside air, and so the humidifier is unnecessary. In the third configuration, the sensible rotaryexchanger takes place of the liquid regenerative exchanger and makes the heating coil unnecessary.Air-and-water systems use centrally cooled water to feed the terminals, designed to balance sen-
sible load, installed within the building. Sensible load may be removed: 17
(a) cooling and distributing recirculation air by means of fan-coils;(b) by means of convection and thermal radiation, maintaining at low temperature the floor orceiling radiating panels.
The second method is widespread in Europe and is gaining ground in the US [29].
2.2. Hybrid HVAC systems with chemical dehumidification
The diagram of Fig. 9 shows that the humidity of outlet process air is a function of its inlet stateand of regeneration air temperature; its value decreases:
nducers are terminals that, through nozzles supplying primary air at high pressure, drive the recirculating air to
hrough the cooling coil. The mixing air is supplied in ambient. These terminals are practically not more in market.
CC
100% O.A.
HumidifierHC
- +
Supply airUpstream
coilrecovery
Downstreamrecovery
coil
Configuration 1: conventional cooling, heating and humidification with run-around heat recovery
-100% O.A.
+Pre-heating
+Enthalpy
wheel
Return air
coil
Supply air
HCCC
Upstream
coilrecovery
coilrecovery
Downstream
Configuration 2: enthalpy wheel heat recovery and run-around re-heat
+
100% O.A.
-
CC
Supply air
Return air
recoverySensible heat
Supply fan
Return fan
coilPre-heating Enthalpy
wheel
Configuration 3: two wheels system (enthalpy wheel and sensible one)
Fig. 12. Mechanical dehumidification: dedicated outdoor air handling unit configurations. Configuration 1:
conventional cooling, heating and humidification with run-around heat recovery. Configuration 2: enthalpy wheel
heat recovery and run-around re-heat. Configuration 3: two wheels system (enthalpy wheel and sensible one).
P. Mazzei et al. / Applied Thermal Engineering xxx (2004) xxx–xxx 19
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• when humidity and temperature of entering process air diminish,• when the temperature of regeneration air increases,• when the process air velocity decreases.
0
2
4
6
8
10
12
14
16
18
20hu
mid
ity ra
tio [g
/kg]
Angelantoni
MILAN GENOA BARCELONA PARIS LONDON ATHENS
ASHRAE
in
out
ω
ω
Fig. 13. Typical performances of a desiccant wheel, for six European sites and for ASHRAE 1% DP-MCDB outdoor
design conditions (data obtained from performance charts of Fig. 14, cap. 22, 2004 ASHRAE Handbook—HVAC
Systems and equipment [64], and from performance charts kindly provided by Angelantoni Industrie S.p.A. for the
desiccant wheel model RU-060 6/4, for a regeneration temperature of 60 �C).
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In Fig. 13 the performances of a desiccant wheel, for six European sites and for ASHRAE 1%DP-MCDB outdoor design conditions, are reported. Fig. 13 is based on standard data and ondata related to a particular wheel working at a low regeneration temperature (60 �C). It has beenfound that the level of humidity reached by the second wheel is not sufficiently low for usual air-conditioning applications. This could be solved by increasing slightly regeneration temperature orby pre-cooling the process air.Fig. 14 shows an interesting use of the desiccant wheel combined with a vapour compression
refrigerating machine: process air is first cooled and dehumidified as it passes through the evap-orator (A–B), and then dehumidified further through adsorption (B–C); regeneration air is heatedby the condenser (D–E), and then flows through the wheel (E–F). Energy input within the systemis thus only electric. The desiccant wheel must be made of desiccant regenerable at low tempera-tures, available on the market.In the field of the summer air conditioning for non-industrial applications, the chemical dehu-
midification can be applied instead of the traditional mechanical dehumidification; 18 often thetwo technologies are integrated (hybrid HVAC systems): the first to balance the latent load,the second one to balance the sensible load.Hybrid HVAC systems with chemical dehumidification distinguish themselves essentially for
the following reasons.
18 For Italian climates, in winter season the latent load balancing can be obtained also by supplying outdoor air,
usually characterised by a humidity ratio less than the value required in ambient.
Fig. 14. Example of utilization of a desiccant wheel inserted in a cooling cycle (Econosorb—for kind permission of
Angelantoni Industrie S.p.A.).
P. Mazzei et al. / Applied Thermal Engineering xxx (2004) xxx–xxx 21
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(a) Traditional refrigeration systems are not suitable to control separately latent and sensiblethermal loads: often, in order to adequately control ambient relative humidity, it is necessaryto cool air up to low temperatures and then to reheat. Consequently COP reduction and highenergy use take place. Handling sensible loads by means of a traditional refrigeration plantand latent loads by means of desiccant, especially in presence of a low sensible/total loadratio, can significantly enhance the system efficiency. With this kind of hybrid system re-heat-ing is not necessary. This advantage is particularly evident in partial-load conditions [22].
(b) In operating conditions, hybrid systems based on chemical dehumidification permit to controlseparately both temperature and humidity (the DW is connected to a humidity sensor, the CCto a temperature sensor). On the contrary, in traditional cooling systems only temperature isgenerally directly controlled (DBTCS), while humidity can vary.
(c) Systems based on chemical dehumidification allow to reduce humidity even when requireddew point temperature is very low, so allowing an easier balance of high latent loads. Onthe contrary, conventional systems can dehumidify air stream generally only for requireddew point temperatures higher than 4�C.
(d) Hybrid systems assure a better thermal comfort, since humidity can be accurately controlled,and also a better air quality [30–33]. In fact, the absence of condensed water strongly reducesthe presence of microorganisms as bacteria, viruses and fungi. So these systems are particu-larly recommended in application in which severe hygienic conditions must be maintained(medical facilities and laboratories).
(e) Since in hybrid systems the CC task is only sensible cooling of the air stream, the coolingfluid temperature can be higher (for example, the typical 5–7�C of the chilled water canbe changed up to 14�C and over), with a consequent increase of the refrigerating machineCOP.
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(f) Hybrid systems allow to reduce the vapour compression refrigerating machine power becauselatent load is already balanced by the desiccant system. The smaller size allows to reduceenergy use, required electric power and starting investment capital.
(g) The technology based on chemical dehumidification, reducing electric power and energyrequirements and the CFC and HCFC refrigerant fluids use, is characterised by a low envi-ronmental impact.
(h) The chemical dehumidification can be applied also to existing traditional HVAC systems(‘‘retrofitting’’) which are not able to balance latent load, for example when outdoor air per-centage is increased in order to conform the plant to the present standards.
(i) Installing cost of an hybrid systemwith chemical dehumidification is generally higher than a tra-ditional system, but it can be balanced, in some applications, by lower operating costs [34–40].
(j) It is possible to use available thermal energy [41,42] to regenerate the desiccant.
The main disadvantages connected to hybrid HVAC systems with chemical dehumidificationare the following:
• it is possible that, in presence of solid adsorbent materials, solid particles could be dragged bythe air stream, but such inconvenience is decreasing while technology improves;
• thermal energy necessary to the regeneration process is a considerable amount and it increaseswith the dehumidification requirements and with the regeneration temperature. Only the lastgeneration of adsorption desiccants allows to obtain regeneration temperatures between40�C and 80 �C, so it is possible to satisfy the reactivation needs with low temperature thermalrecoveries. However, the payback not always attains acceptable values [43];
• the scarce familiarity with such technology and the lack of information about performancesand cost/benefit ratio hamper the spreading of hybrid desiccant systems [44], even if todaychemical dehumidification technology can compete with conventional systems also for residen-tial and commercial applications.
In the technical literature various hybrid system with desiccant wheel configurations are pro-posed [12,13,20,30,35,37,39,42,43,45–62].A detailed description and some possible classifications of these configurations are reported in a
previous authors� paper [63]. For example, in Fig. 15 a possible system configuration (with partialrecirculation) is shown. It can be observed the presence of an enthalpic economizer (linked tothree coupled dampers) that varies the outdoor air stream percentage: when outdoor air condi-tions allow it, partial recirculation mode gives the place to the temporary more convenient all-external air mode. In Fig. 16, considering the hybrid system of Fig. 15, qualitative treatmentsof process and regeneration air on the psychrometric chart are reported for a typical Italian sitein summer design conditions.Various control schemes may be implemented for systems based on the desiccant wheel, for
example (a) coupled dampers to by-pass the wheel—process air side—and (b) modulation ofthe power delivered by the regenerator [62]. The signal from the humidity sensor, placed in theambient or, even better, in the return air duct, is sent to the controller (possibly a PID controller);if necessary, the controller output goes to the actuator of a by-pass system of the desiccant wheelwith opposed blade coupled dampers. One is a front damper, the other by-pass. As / increases
spaceReturn from
+
+
H C
DesiccantwheelOutdoor air
(ventilation air)Outdoor air
(economizer)
o hx
Sensible heat
dw
recovery
m
C C
-
-
To
ro
coolerEvaporative
heaterRegeneration
rg
xl
outdoorsRelief air to
r
space
Humidifier
sr
ConditionedSupply fan
Return fan outdoors
Fig. 15. Partial recirculation desiccant hybrid system with or without evaporative cooler.
Fig. 16. Psychrometric chart related to the system of Fig. 15 (without evaporative cooler).
P. Mazzei et al. / Applied Thermal Engineering xxx (2004) xxx–xxx 23
ARTICLE IN PRESS
compared with setpoint, the controller increases the degree of aperture of the front damper andreduces that of the by-pass damper. As / decreases compared with setpoint, the opposite happens:the front damper closes and the by-pass damper opens. Should / continue to decrease with thefront damper closed it may be necessary to intervene on the dehumidification capacity of thewheel (varying its angular velocity, if, for example, a frequency control for the wheel�s electricmotor is present, or by regulating the on/off on the regenerator and the wheel).The output of the controller may be sent (case b) to a three-way servo valve placed on the
regeneration heating coil, instead of to the by-pass system actuator described earlier, so as tomodulate regeneration power and hence the dehumidification capacity of the wheel as / variesin the ambient.Less effective control of / may be exerted by acting on the regeneration air side. A temperature
sensor is placed in the regeneration airflow downstreamof the wheel: in response to this signal a con-trollermodulates the regenerator. If the temperature signal falls below setpoint, the controller opens
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the damper which regulates regeneration airflow rate; the opposite occurs if the temperature signalexceeds setpoint. Alternatively, the controller may act on the three-way valve of the regenerationheating coil in such a way as to modulate regeneration power on the basis of the temperature signal.
3. Some applications
The applications of HVAC systems based on mechanical and chemical dehumidification areseveral: as regards the second one, the adsorption technique will be specifically analysed, becauseit is the most developed; specific literature [19,20,64] can be examined for the applications basedon absorption.An interesting application of HVAC systems with mechanical and adsorption dehumidification
regards retail store [63] and, above all, supermarkets [19,20,34,35,65,66], in which the air condi-tioning is necessary both for the comfort of the occupants and for the correct operating of theopen refrigerators, which operate well only if the ambient humidity is maintained low.Moreover, for the supermarkets, the ratio between the sensible and latent components of the
ambient thermal load is in favour of the latent one. In fact, the presence of the open refrigeratorsreduces much more the sensible load than the latent one [67]. So, it is clear that if the dehumid-ification is mechanical, it is necessary the oversizing of the cooling coil, and also the successivere-heating of the handled air, with remarkable energy use. Besides, in the stricter conditions ofpartial load, the traditional system can lose the moisture control in ambient, with important neg-ative consequences [19,68–71]: (a) the load of the open refrigerators increases, and so their oper-ating costs; (b) because of the greater frost formation, longer defrost periods are required, theshelf-life of the exposed goods shortens, their aspect gets worse, the paper containers deterioraterendering illegible the labels; (c) with the increasing of the dew point temperature of the air, sur-face condensation problems appear on the walls, on the structures, on the goods [71,72].A possible solution consists in resorting to hybrid HVAC systems with dehumidification by des-
iccant wheel, that offers numerous advantages. These are the ones of the chemical dehumidifica-tion (Section 2.2) and furthermore:
(a) a consisting energy saving can be obtained: for this aim the desiccant dehumidification sys-tem, that balances the latent load, can usefully employ the condensation heat of the vapourcompression chiller in order to pre-heat the regeneration air [73], so reducing the thermalpower demand;
(b) the anti-sweat heaters systems can be eliminated or reduced.
In [66] an application of traditional system and desiccant wheel hybrid system, both of the roof-top type, for a supermarket of approximately 3700m2 sited in Rome, 19 is reported: the obtainablesavings with the hybrid systems with respect to the traditional ones have been estimated, 20 interms of operating costs, considering that the most part of these is connected to the open refrig-erators and that the proper operation of the HVAC system indirectly induces remarkable savings
19 Italian electric energy and gas fares have been considered.20 By means of the program DesiCalcTM [50].
P. Mazzei et al. / Applied Thermal Engineering xxx (2004) xxx–xxx 25
ARTICLE IN PRESS
also on the refrigeration system. The simple payback time, connected to the extra cost of the hy-brid system, resulted of 2–3 years. Moreover, it has been found that relative humidity in ambientis maintained to lower levels, therefore the frost formation is reduced and the problems related tothe shelf-life are exceeded, as well as those connected to the aspect and the integrity of the exposedgoods. Finally, a reduction of approximately 30% of the air flow rate supplied to the rooms hasbeen noticed, with consequent reduction of system and operating costs and of the space requiredby the ducts. The savings on the operating costs increase significantly if it is imposed that both thesystems guarantee in ambient the same low relative humidity values.In [63] an application of traditional system and desiccant wheel hybrid system, for a theatre of
1200m2 sited in Rome 19, is reported: different configurations (among which also that reported inFig. 15) of traditional and hybrid systems, both centralized and of the roof-top type, have beenconsidered. The obtainable savings 20 with the hybrid systems with respect to the traditional ones,in terms of operating costs, resulted between 23% and 38%, however always greater than 35% forall the roof-top configuration systems. The savings of primary energy are very similar to those rel-ative to the operating costs. In the various cases, the reduction of the power electric demand re-sulted remarkable, up to approximately 44–50%. It also resulted that the hybrid systems, unlikethe traditional ones, perfectly control in every condition the relative humidity in ambient (in fact,the percentage of hours, with respect to total, in which the ambient relative humidity is greater than60%, is around 20% for the traditional systems and is instead less than 0.6% for the desiccant wheelsystems). Finally, it has been verified that the savings on the operating costs obtainable with thehybrid systems increase (up to a maximum of approximately 45%) with the increasing of the levelof occupation and the outdoor air flow rate per person. This tendency confirms that the HVACsystems with dehumidification by desiccant wheel are particularly suitable to balance high latentloads related to the increase of the required outside air flow rate and of the level of occupation.Besides, the analysis reported in [63] also shows a great variation of the savings depending onthe configuration of hybrid system considered; therefore it is recommended to examine, case forcase, which of the configurations can be the most suitable to maximize the obtainable savings.
0
5
10
15
20
25
30
0 2 4 6 8 10 12 14 16 18
Process air flow rate [103 m3/h]
Cur
rent
sel
ling
pric
e [
/(m
3 /h)]
Two wheels systems
Industrial dehumidifiers (∆ω = 8 g/kg)
Commercial dehumidifiers
(∆ω = 6 g/kg)
Target price for commercial dehumidifiers
Fig. 17. Range of the normalized current selling prices for industrial and commercial dehumidifiers with solid desiccant
and heat recovery.
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Finally, in Fig. 17, data recently appeared in literature [74] about the selling prices of dehumid-ifiers with desiccant wheel and heat recovery wheel, designed for industrial and commercial fields,are reported.
4. Conclusions
The HVAC systems design requires more and more attention, above all considering the recentventilation standards, in order to reach and maintain in ambient the optimal level of humidity,particularly in partial load conditions.In the mechanical dehumidification field, the direct control of ambient Tdb and / can be prop-
erly obtained by means of an air handling unit (AHU) which treats outside air alone, while recir-culating air is treated by a simple cooling coil; control of the two equipments is independent.Among the possible AHU configurations, that one with two wheels, an enthalpy wheel and a sen-sible one, respectively upstream and downstream of the cooling coil, turns out to be more conven-ient as regards energy use and power installed.In the chemical dehumidification field, absorption systems have been used successfully for large
size buildings in the tertiary sector; as humidity control in the buildings becomes established, theirsignificant growth, also on a lower scale, is expected. Controlling opportunely temperature andconcentration of the solution different dehumidification lines can be obtained, with an outletair temperature nearly equal to inlet, unlike mechanical and adsorption dehumidification; so, trea-ted process air can have psychrometric conditions required for it to be supplied into the roomwithout re-heating or post-cooling.Adsorption dehumidification systems, long employed in the industrial sector, has now pro-
voked renewed interest and has been extended towards the non-industrial sector, both becauseof more stringent IAQ requirements, and the gradual reduction of the regeneration temperaturerequired by new materials.In the field of the summer air conditioning for non-industrial applications, the adsorption dehu-
midification is alternative to the traditional mechanical dehumidification; often the two technol-ogies are integrated (hybrid HVAC systems): the first to balance the latent load, the latter tobalance the sensible one.Hybrid HVAC systems with adsorption dehumidification allow a better moisture control and
offer numerous other advantages. However it has to be pointed out that: (a) the payback not alwaysattains acceptable values; (b) the scarce familiarity with such technology and the lack of informa-tion about performances and cost/benefit ratio hamper the spreading of hybrid desiccant systems.Among the various possible applications of hybrid HVAC system with desiccant wheel, a
supermarket and theatre have been considered.Comparing a traditional system and a hybrid system with desiccant wheel, both of the roof-top
type, for a supermarket of approximately 3700m2 sited in Rome, resulted: (a) a simple paybacktime of 2–3years; (b) relative humidity in ambient is maintained to lower levels, so the frost for-mation is reduced and the problems connected to the shelf-life, as well as to the aspect and theintegrity of the exposed goods, are exceeded; (c) a reduction of approximately 30% of the supplyair flow rate, with consequent reduction of system and operating costs and of the space requiredby the ducts.
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The same comparison, for a theatre of 1200m2 sited in Rome, has provided the following re-sults: (a) the obtainable savings, in terms of operating costs, resulted between 23% and 38%;(b) the reduction of the power electric demand is meaningful, up to approximately 44–50%; (c)the hybrid system, unlike the traditional one, perfectly controls in every condition the relativehumidity in ambient; (d) the savings on the operating costs obtainable with the hybrid systemsincrease (up to a maximum of approximately 45%) with the increasing of the level of occupationand the outdoor air flow rate to be assured for person.
Acknowledgement
The authors would like to thank Mr. Malaterra of Angelantoni S.p.A. for his detailed informa-tion about the desiccant wheel.
Appendix A
In summer the humidity ratio of outside air is normally higher than the indoor humidity ratio:and thus the need to dehumidify. If, for example, the indoor comfort conditions are Tdb = 25�Cand / = 50%, (i.e. xr = 10gv/kga) and the outdoor air design conditions are Tdb = 32�C and /= 55%, (i.e. xo = 16.5gv/kga), the humidity ratio of the supply air must be:
21 F
xs < xr ¼ 10 gv=kga < xo ¼ 16:5 gv=kga
and the dehumidification system must be able to reduce the x = 16.5gv/kga to xs. In steady state,from the mass balance for water, referred to a conditioned ambient with N occupants, 21 it isobtained:
xs ¼ xr �N � _mvuN � _mau
¼ xr �_mvu
�qa � _V au: ðA:1Þ
The humidity ratio of the airflow to be supplied must be equal to indoor airflow humidity ratiominus the ratio between unitary mass flow rates (referred to each occupant) of water vapourand air. Referring to the case under examination, and assuming the following values as indicative[1]:
_mvu ¼ 60g=h �qa ¼ 1:17kg=m3 _V au ¼ 10dm3=s
it follows that:
xs ¼ 10� 1:4gv=kga ¼ 8:6gv=kga ! T dp;s ¼ 11:8 C:
So, the dehumidification capacity, in terms of humidity ratio variation, must be
Dx ¼ 16:5� 8:6 ¼ 7:9gv=kga:
or simplicity it is assumed that vapour in the ambient is generated by the occupants alone.
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It can be noted that only the 18% (1.4/7.9) of Dx is due to the indoor load, while 82% is due tothe outdoor air.The dehumidification capacity (DC) usually expresses the water mass flow rate removed from
the handled air flow [75]. In steady state, from the water mass balance referred to the dehumid-ifier, one can obtain:
DC ¼ _mw ¼ _maDxDE ¼ �qa _V aDxDE: ðA:2Þ
To express, as usual, DC in [kg/h] units, in Eq. (A.2) the following units are used: �qa [kg/m3], _V a
[m3/h], x [kgv/kga]. It can be observed that the unitary capacity is equal to the humidity ratiovariation:
DCu ¼_mw_ma
¼ DxDE: ðA:3Þ
Many building and furniture materials are known to be hygroscopic, thus the way in which airis supplied may give rise to condensation and mould growth. It would be advisable to mix treatedair with ambient air in order to increase temperature, thus preventing extremely cold air fromcoming into contact with surfaces, so avoiding condensation.
References
[1] UNI 10339, Impianti aeraulici a fini di benessere—Generalita, classificazione e requisiti—Regole per la richiesta
d�offerta, l�offerta, l�ordine e la fornitura, 1995.[2] ASHRAE Standard 62-1981, Ventilation for acceptable indoor air quality, 1981.
[3] ASHRAE Standard 62-2001, Ventilation for acceptable indoor air quality, 2001.
[4] CEN prENV 1752, Ventilation for buildings: design criteria for the indoor environment, 1996.
[5] CIBSE Guide A, revision section 2, Environmental criteria for design, Charter Institute of Building Service
Engineers, U, 1993.
[6] DIN 1946 Part 2, Ventilation and air conditioning: technical health requirements, 1994.
[7] NKB Report No.61, Indoor climate—air quality, Nordic Committee on Building Regulations NKB, 1991.
[8] B.W. Olesen, International development of standards for ventilation of buildings, ASHRAE Journal (April) (1997)
31–39.
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