performance analysis of four-partition desiccant wheel and hybrid dehumidification air-conditioning...

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Performance analysis of four-partition desiccant wheel andhybrid dehumidification air-conditioning system

Jongsoo Jeong*, Seiichi Yamaguchi, Kiyoshi Saito, Sunao Kawai

Department of Applied Mechanics and Aerospace Engineering, School of Fundamental Science and Engineering, Waseda University, 3-4-1-58-

210 Okubo, Shinjuku-ku, Tokyo 169-8555, Japan

a r t i c l e i n f o

Article history:

Received 14 July 2009

Received in revised form

8 October 2009

Accepted 2 December 2009

Available online 6 December 2009

Keywords:

Air conditioning

Desiccant system

Desiccant wheel

Geometry

Modelling

Simulation

Performance

Silica gel

* Corresponding author. Tel.: þ81 3 5286 324E-mail address: [email protected] (J.

0140-7007/$ – see front matter ª 2009 Elsevidoi:10.1016/j.ijrefrig.2009.12.001

a b s t r a c t

A desiccant dehumidification system with air can decrease energy consumption because it

can be driven by low-grade waste heat below 80 �C. If this system can be driven by low-

temperature heat sources whose temperature is below 50 �C, exhausted heat from fuel

cells or air conditioners that exist everywhere can be used as heat sources. This could lead

to considerable energy saving. This study provides a detailed evaluation of the perfor-

mance of a four-partition desiccant wheel to make a low-temperature driving heat source

possible and achieve considerable energy saving by the simulation and experiment.

Further, the study investigates the in-depth performance of a hybrid air-conditioning

system with a four-partition desiccant wheel by simulation. As a result, it was clear that

there exists an optimum rotational speed to maximize the dehumidification performance

and that the hybrid air-conditioning system improves COP by approximately 94% as

compared to the conventional vapour compression-type refrigerator.

ª 2009 Elsevier Ltd and IIR. All rights reserved.

Analyse de la performance d’une roue deshydratante a quatresegments et d’un systeme de conditionnement d’air adeshydratant hybride

Mots cles : Conditionnement d’air ; Systeme a deshydratant ; Roue deshydratante ; Geometrie ; Modelisation ; Simulation ; Performance ;

gel de silice

7; fax: þ81 3 5286 3259.Jeong).er Ltd and IIR. All rights reserved.

mailto:[email protected]

http://www.iifiir.org

http://www.elsevier.com/locate/ijrefrig

Nomenclature

Ab area of desiccant surface except for curve-shaped

desiccant surface in contact with air (m2)

Af area of curve-shaped desiccant surface in contact

with air (m2)

D diffusion coefficient (m2 s�1)

d diameter of desiccant wheel (m)

dh hydraulic diameter (m)

G mass flow rate (kg s�1)

h specific enthalpy (J kg�1)

hads adsorption heat (J kg�1)

hvap latent heat (J kg�1)

jm mass flux (kg m�2 s�1)

Kh overall heat transfer coefficient (W m�2 K�1)

Km overall mass transfer coefficient (kg m�2 s�1)

L desiccant length along air flow path (m)

le entrance region length (m)

lp length of plane in control volume (m)

lh pitch distance between flat walls (m)

ma mass fraction of water vapour in moist air

(kg kg�1)

mb mass fraction of water vapour in equilibrium of

the desiccant wall (kg kg�1)

N rotational speed (rph)

Nu Nusselt number (–)

qs heat flux (W m�2)

Q heat transfer rate (kW)

Pr Prandtl number (–)

Re Reynolds number (–)

RH relative humidity (–)

Sc Schmit number (–)

Sh Sherwood number (–)

T temperature (�C)

t thickness of corrugated sheet (m)

u velocity (m s�1)

V volume flow rate (m3 h�1)

Va volume of air flow path (m3)

Vb volume of desiccant bed (m3)

W power (kW)

X mass fraction of water in the desiccant (kg kg�1)

x humidity ratio (kg kg(DA)�1)

a heat transfer coefficient (W m�2 K�1)

b mass transfer coefficient (m s�1)

q angle of desiccant wheel (rad)

l thermal conductivity (W m�1 K�1)

hcom compressor efficiency (–)

hS content ratio of silica gel (–)

r density (kg m�3)

u angular speed of desiccant wheel (rad s�1)

z z axis of air flow path (m)

Subscripts

a moist air

ad adiabatic

ads adsorption

b desiccant bed

co cooling

com compressor

des desorption

eva evaporator

hyb hybrid system

i inlet

mde mechanical dehumidification

o outlet

p pure

pi process air at inlet

pro process air

r refrigerant

reg regeneration air

ri regeneration air at inlet

i n t e r n a t i o n a l j o u r n a l o f r e f r i g e r a t i o n 3 3 ( 2 0 1 0 ) 4 9 6 – 5 0 9 497

1. Introduction

Since a desiccant dehumidification method can dehumidify

air by converting latent heat into sensible heat, it is unnec-

essary to supercool and to reheat the air on a mechanical

dehumidification system such as a compression-type refrig-

erator (ASHRAE, 2001). Therefore, this method attracts atten-

tion as the air-conditioning system from the view point of

energy saving and amenity. In the normal desiccant dehu-

midification system (Harriman, 1994; Meckler, 1994), an

approximately 80 �C heat source is directly required to dehu-

midify the outdoor air. If an electric heater or boiler is used as

this driving heat source, it runs counter to energy saving. In

order to solve this problem, we have investigated the desic-

cant wheel dehumidification method with the multistage

adsorption and regeneration process to make it possible to use

the low-temperature heat source that is approximately

40–50 �C from the compression-type refrigerator (Inagaki

et al., 2004; Shibao et al., 2006).

As a result, it turned out that we could construct a highly

efficient air-conditioning system by combining the multistage

desiccant dehumidification system with the compression-

type refrigerator. However, the system increased in size

because of the multiple desiccant wheels. Accordingly, in

a previous report (Shibao et al., 2007), we suggested the four-

partition desiccant wheel to downsize multiple desiccant

wheels. This wheel can realize the double-stage adsorption

and regeneration process in only one desiccant wheel. Lately,

Ge et al. (2008) also experimentally investigated its perfor-

mance under various operation conditions for a one-rotor

two-stage rotary desiccant cooling system. However, as it was

not discussed theoretically, the performance of the four-

partition desiccant wheel system was not totally evaluated

considering driving heat source.

For the hybrid air-conditioning system with desiccant,

Yadav (1995), Dhar and Singh, 2001 and Jia et al. (2006) inves-

tigated a performance of a hybrid desiccant cooling system

comprising the conventional vapour compression-type


refrigerator coupled with a desiccant dehumidifier. However,

as the conventional single stage desiccant was used for this

system, the condensation temperature of the vapour

compression-type refrigerator increases greatly. And, some-

times, an electric heater is also used to compensate the

shortage of the driving heat. This system can’t increase the

system performance. To accomplish high performance hybrid

desiccant air-conditioning system with only vapour

compression-type refrigerator, it needs to decrease the driving

heat source temperature for the desiccant regeneration.

Hence, the four-partition desiccant wheel has been investi-

gated to decrease the driving heat source temperature as

hybrid system.

To construct the hybrid system with the four-partition

desiccant wheel, this study investigates the in-depth perfor-

mance of the four-partition desiccant wheel with the simu-

lation and experiment to clearly evaluate the characteristics

of a four-partiton desiccant wheel; effect of the regeneration

temperature, rotational speed of desiccant wheel, cooling

temperature, process air flow rate, and humidity of the

ambient air. Moreover, to realize a high efficiency for the air-

conditioning system and amenity in the room at the same

time, we further discuss the performance of a new type of

hybrid air-conditioning system that combines the four-parti-

tion desiccant wheel with the compression-type refrigerator

by simulation as well.

2. Four-partition desiccant wheel and hybridair-conditioning system

Fig. 1 shows a hybrid air-conditioning system that combines

two conventional desiccant wheels with the compression

cycle, and Fig. 2 shows the hybrid air-conditioning system

with a four-partition desiccant wheel is divided by the equal

area ratio on each flow path. The four-partition desiccant

wheel used in the experiment is shown in Fig. 3. Fig. 4 illus-

trates the driving points of each cooling and dehumidification

system, such as mechanical dehumidification

including a reheat process with the

compression-type refrigerator, the normal type of desiccant

dehumidification that has one desiccant wheel

, and a two-stage dehu-

midification. The four-partition desiccant wheel that can

realize the double-stage adsorption and regeneration process

Fig. 1 – Hybrid air-conditioning syst

with only one desiccant wheel has the same air flow path as

two desiccant wheels – process air and

regeneration air , as shown in Fig. 4.

In the process , the process air precools before the

moisture of the process air is removed. Further, after the first-

stage dehumidification by the desiccant wheel process

, the process air at the process is cooled by

evaporator again. In the second-stage dehumidification

process , the moisture is removed by the desiccant

again, and then, the air cools down to the target process air

outlet temperature at point . Otherwise, the regeneration air

preheats in the process before the regeneration air

flows into the desiccant wheel. After the first-stage process

, the regeneration air heats up in the condenser of the

refrigerator in process as shown in Fig. 2. Then, the

regeneration air goes through the second-stage regeneration

process .

Surplus condensation heat is rejected to the ambient air, as

shown in Fig. 1 or Fig. 2 considering the entire heat balance of

the system. Thus, the temperature of the process air can be

lowered on the dehumidification air flow path so that the

regeneration temperature at point and can also be lower.

Consequently, the heat source temperature can be reduced

drastically. Further, as it is unnecessary to cool the process air

below the dew point temperature, this dehumidification

method with the four-partition desiccant wheel can save the

energy greatly. The refrigerant of the compression-type

refrigerator is R134a.

3. Detailed structure and mathematicalmodel of desiccant wheel, and model of otherelements

Fig. 5 shows the detailed structure and model of the desiccant

wheel. The structure of the wheel is usually realized by arrayal

made up of a corrugated lamina and a plane that are sheets of

the glass fiber impregnated with desiccant. The diameter of the

desiccant wheel that is made of silica gel is 250 mm, and the

thickness is 200 mm. These are the same as those of the actual

desiccant wheel; HY-SG. Specifications of this desiccant wheel

and property of desiccant are listed in Table 1. A mathematical

model of the desiccant wheel is based on the static simulation

model constructed by Yamaguchi et al. (2007).

em with two desiccant wheels.

Fig. 2 – Hybrid air-conditioning system with four-partition desiccant wheel.

Fig. 3 – Four-partition desiccant wheel.


3.1. Assumptions

Some assumptions are required to build the simulation

model. These assumptions are as follows:

� As the desiccant walls move along the q direction, the air

that flows in the channel also moves along the q direction.

However, the movement of the air along the q direction is

considerably slower than that in the z direction. Therefore,

the movement of the air in the q direction can be neglected.

� The resistance of heat and mass transfer in the z and q

directions inside the desiccant walls can be neglected. In

other words, heat and mass transfer are that by forced

convection transfer and the diffusion, respectively, between

air and desiccant wall, and are considered on only direction

of air – desiccant.

� The entrance region length can be neglected because it is

considerablyshorterthanthelengthof thedesiccantwheel (L).

3.2. Model of desiccant

For the air side, the continuity equation of the entire moist air is

VavðrauaÞ

vzþ jm

�Af þAb

�¼ 0 (1)

The continuity equation of the moisture vapour is

0 20 40 60 800

0.01

0.02

0.03

Temperature T oC

102030406080

5

15

1

Hum

idit

y ra

tio

x kg

/kg(

DA

) Four-partition dehumidification Process air Regeneration air

Conventional dehumidification Process air Regeneration air

Mechanical dehumidification

Cooling temperature Regeneration temperature

''

'

'

Fig. 4 – Air flow of aixr-conditioning systems.

12

NN-1 1 2 M

Z

zz

Inlet air

Inlet wheelOutlet wheel

Outlet air

0

Desiccant bed

qs jmAir

auaxa

b xb

Ab Af

Control volume

Va

VbVa

Vb

Air channelDesiccant wall

lp

Air channel

t

l h

t /2

Fig. 5 – Detailed structure and model of desiccant wheel.


VavðrauaxaÞ

vzþ jm

�Af þ Ab

�¼ 0 (2)

The energy equation of the air is

VavðrauahaÞ

vzþ�jmhvap þ qs

��Af þ Ab

�¼ 0 (3)

For the desiccant, the continuity equation of the desiccant is

Vbuvrb

vq� jm

�Af þ Ab

�¼ 0 (4)

Table 1 – Specifications of this desiccant wheel andproperty values of desiccant.

Length of desiccant wheel L m 0.2

Diameter of desiccant wheel d m 0.25

Thickness of corrugated sheet t mm 0.2

Length of plane in control

volume

lp mm 3.8

Pitch distance between

flat walls

lh mm 1.9

Channel hydraulic

diameter

dh mm 1.5

Surface area per volume

of air flow path

(AbþAf)/Va m2 m�3 3000

Surface area per

volume of desiccant

(AbþAf)/Vb m2 m�3 6000

Area ratio of wheel on

regeneration air side

% 50 (each

part 25%)

Area ratio of wheel

on process air side

% 50 (each

part 25%)

Desiccant – Silica gel

Content ratio of

silica gel

hS – 0.7

Apparent density

of desiccant

kg m�3 800

Specific heat of

desiccant

J kg�1 K�1 980

The continuity equation of the water content in the desic-

cant is

VbuvðrbXbÞ

vq� jm

�Af þAb

�¼ 0 (5)

The energy equation of the desiccant is

VbuvðrbhbÞ

vq��jmhads þ qs

��Af þAb

�¼ 0 (6)

The mass and heat transfer rates are

jm ¼ Kmðma �mbÞ (7)

qs ¼ KhðTa � TbÞ (8)

The overall mass and heat transfer coefficients are

Km ¼ ShraDa

dh(9)

Kh ¼ Nula

dh(10)

When the simulation is carried out, the finite volume

method is adopted using the above model. The heat and mass

transfers between the air and the desiccant walls are divided

into the forced convection transfer and the diffusion,

respectively, on desiccant walls. Shah (1975) investigated the

forced convection heat transfer in a small air channel shown

in Fig. 5. According to his results, a Nusselt number is given for

laminar and fully developed conditions with a constant

surface temperature.

Nuhadh=la ¼ 2:12 (11)

Sherwood number is given by the analogy between the heat

and the mass transfers. The value is

Shhbdh=Da ¼ 2:12 (12)


For laminar flow, the entrance region length (le) may be

expressed as follows:

le=dh ¼ 0:05Re ðPr or ScÞ (13)

Under the experimental conditions of this study, the

entrance region length is approximately 5% of the length of

the desiccant wheel. Therefore, the entrance region length

can be neglected.

Furthermore, for the heat and mass transfers, we investi-

gated the non-dimensional overall heat and mass transfer

coefficients that include both of the convective and conduc-

tive resistance derived from the experiment (Yamaguchi et al.,

2007) illustrated in Fig. 6. Hence, Fig. 6 also shows the only

convective heat and mass transfer performance for the air

side from the calculation. These are almost 2.12 as shown in

Eqs. (11) and (12) and almost agree with the convective

performances. This implies that the resistance of the heat and

mass transfers is controlled by the air side of the desiccant

wheel. This is attributed to the fact that the thickness of the

wall is considerably smaller than the size of the air channel.

3.3. Boundary conditions

For boundary conditions, the conditions of the regeneration

and process air (flow rate, humidity ratio, and temperature)

are given at the inlet of the desiccant wheel.

0 � q <p

2; p � q <

3p

2; z ¼ 0 ðProcess air at inletÞ

Tair ¼ Tpi; xair ¼ xpi; uair ¼ upi (14,15,16)

p

2� q < p;

3p

2� q < 2p; z ¼ L; ðRegeneration air at inletÞ

Tair ¼ Tri; xair ¼ xri; uair ¼ uri (17,18,19)

3.4. Properties of silica gel

Barlow (1982) described the adsorption heat and the adsorp-

tion isotherm of a silica gel.

The adsorption heat is expressed as a function of the mass

fraction Xp (pure silica gel) of water in the desiccant and the

latent heat hvap of evaporation. It follows that

Xp ¼ Xb=hS (20)

15 20 25 30 350.5

1

5

10

50

Non

-dim

ensi

onal

over

all h

eat t

rans

fer

coef

fici

ent

Rotational speed N rph

Regeneration temperature: 80 : 70 : 60

Fig. 6 – Non-dimensional overall hea

Xp < 0:1

hads=hvap ¼ 1:3� 1:75Xp (21)

Xp � 0:1

hads=hvap ¼ 1:14� 0:15Xp (22)

Adsorption isotherm is given by

RH ¼

�0:616238Xp þ 16:7916X2

p � 74:34228X3p þ 116:6834X4

p

�

f1� ðT� 40Þ=300g(23)

The desiccant wall does not contain only pure silica gel. The

mass fraction Xb of water in actual desiccant gives relation to

Xp of pure silica gel by the content ratio (hS) of silica gel as

shown in Eq. (20). Therefore, Eq. (23) can be expressed as pure

desiccant adsorption isotherm considering the content ratio

of the pure silica gel. Content ratio (hS) of silica gel is 0.7.

3.5. Model of other elements

Model of hybrid desiccant system and mechanical dehumid-

ification system with same elements is as below.

For the model of compressor, expansion valve, evaporator

and condenser in hybrid desiccant system, the continuity

equation of the refrigerant is

Gr�o � Gr�i ¼ 0 (24)

In compressor, the energy equation of the refrigerant is

Grho � Grhi �Wcom ¼ 0 (25)

The compressor efficiency is as follows

hcom ¼had � hi

ho � hi(26)

In expansion valve, the energy equation of the refrigerant is

Grho � Grhi ¼ 0 (27)

For the energy equation of evaporator and condenser, the

energy equation is

Grho � Grhi � Qr ¼ 0 ðin case of condenser :

þ Qr; in case of evaporator : � QrÞ (28)

15 20 25 30 350.5

1

5

10

50

Non

-dim

ensi

onal

over

all m

ass

tran

sfer

coe

ffic

ient


Regeneration temperature: 80 : 70 : 60

t and mass transfer coefficients.

Fig. 7 – Flow diagram of experimental setup.


Heat transfer performance of heat exchanger is simply

decided by approach temperature, degree of superheat and

degree of subcooling. The approach temperature of evapo-

rator is decided with temperature difference between evapo-

ration temperature and heat exchanger outlet air

temperature. The approach temperature of condenser is

defined as temperature difference between condensation

temperature and heat exchanger outlet air temperature.

Degree of superheat in the evaporator and degree of sub-

cooling in the condenser are shown in Table 2.

4. Experimental method of four-partitiondesiccant wheel

We have conducted an experiment and simulation in order to

examine theactual performance of the four-partition desiccant

wheel and investigated the validity of the simulation model

with the experiment in detail by fully examining the effect of

many parameters – the regeneration temperature, rotational

speed, cooling temperature, process air volume flow rate, and

ambient air humidity ratio – on the performance of the four-

partition desiccant wheel. This is because the model and the

detailed performance of the four-partition desiccant wheel

have not been clarified yet. This experiment is carried out in

order to justify the assumption that the four-partition desic-

cant wheel can be used for the hybrid air-conditioning system.

Fig. 7 shows the flow diagram of the experimental facility for

the four-partition desiccant wheel and the measurement

points. For the regeneration air inlet temperature to the

desiccant wheel, the temperature at point is equivalent to

the temperature at point , and the process air inlet temper-

ature at point flowing into the desiccant wheel agrees with

temperature at point in Fig. 2. Fig. 8 shows the condition

generator of the air. Table 3 gives experimental conditions.

Simulation conditions are of course the same as these.

5. Simulation and experimental results forfour-partition desiccant wheel

5.1. Validity of experiment accuracy and definition ofparameters

Fig. 9 for example shows the total dehumidification rate of the

moisture in the dehumidification side and the regeneration

rate in the regeneration process. From this figure, the

Table 2 – Simulation conditions of conventionalcompression-type refrigerator.

Evaporator Approach temp. �C 5.0

Degree of superheat �C 5.0

Condenser Approach temp. �C 5.0

Degree of subcooling �C 0.0

Compressor Adiabatic efficiency – 0.8

Rotational speed rpm 2000

Process air Mass flow rate kg(DA)/s 0.04

Outdoor air Mass flow rate kg(DA)/s 0.04

experimental accuracy for the moisture balance of four-

partition desiccant wheel is satisfactory level in less than

approximately 10% error.

To show the simulation and experimental results, the

following parameters are used: cooling temperature Tco is the

temperature of the process air at the entrance of the desiccant

wheel (points 2 and 4 in Fig. 4). Regeneration temperature Treg

is a temperature of the regeneration air at the entrance of the

desiccant wheel (points 8 and 10 in Fig. 4). Dehumidification

performance Dx is defined by the difference between the inlet

and the outlet humidity ratio of the process air.

5.2. Effect of regeneration temperature

Fig. 10 shows the effect of the regeneration temperature on

the dehumidification performance and air states at each

measuring point with the fixed desiccant wheel rotational

speed, cooling temperature, ambient air inlet humidity ratio,

and process and regeneration air inlet velocity shown in Table

3. The most important point from this experiment is that the

target of dehumidification performance can be achieved by

the heat source below 50 �C.

5.3. Effect of rotational speed

Fig. 11 shows the effect of the rotational speed on the

dehumidification performance and air states at each

measuring point with the fixed cooling temperature, regen-

eration temperature, ambient air inlet humidity ratio, and

process and regeneration air inlet velocity shown in Table 3.

In this experiment, the humidity difference is maximized as

3.2 [g/kg(DA)] when the rotational speed is approximately 4

[rph]. Therefore, this system should be designed at the point

of maximum humidity difference.

Fig. 8 – Condition generator.

Table 3 – Simulation and experimental conditions.

Process air inlet (1 in Fig. 2) temp. �C 28.0

Process air inlet humidity ratio g/kg(DA) 11.9

Process air inlet relative humidity % 50.0

Regeneration air inlet (7 in Fig. 2) temp. �C 32.3

Regeneration air inlet humidity ratio g/kg(DA) 19.5

Regeneration air inlet relative humidity % 63.0

Cooling temp. (2 and 4 in Fig. 2) �C 18.0

Mass flow rate of process air kg(DA)/s 0.04

Mass flow rate of regeneration air kg(DA)/s 0.04

Rotation speed of desiccant wheel rph 4.0

Process and regeneration air inlet velocity m/s 2.0

0.00 0. 05 0. 10 0.15 0.20 0.25 0. 300.00

0.05

0.10

0.15

0.20

0.25

0.30+10%

-10%

+20%

-20%

0%

Adsorption mass flow rate Gads g s-1

Des

orpt

ion

mas

s fl

ow r

ate

Gde

sg

s-1

Wheel thickness: 200 mmRegeneration temp.: 35 - 55ºCRotational speed: 5rphSuperficial velocoty: 4m s-1

Fig. 9 – Moisture balance between the dehumidification

rate and the regeneration rate.

10

20

30

40

50

Tem

pera

ture

To C T4: : Fixed value

T5

6

8

10

12

14

Hum

idity

rat

iox

g/kg

(DA

)

10

20

30

40

50

Tem

pera

ture

To C

T2: Fixed valueT3

x2: Fixed valuex3

4

6

8

10

12

Hum

idity

rat

iox

g/kg

(DA

)

x4x5

0

2

4

6

8

10

Hum

idity

dif

fere

nce

xg/

kg(D

A)

0

2

4

6

8

Rot

atio

nal s

peed

Nrp

h

Simulation

Experiment

Rotational speed: Fixed value

30 40 50 60 7030

40

50

60

70

Regeneration temperature Treg regoC

Tem

pera

ture

To C T10

T11

18

20

22

24

26

Hum

idity

rat

iox

g/kg

(DA

)

30

40

50

60

70

Tem

pera

ture

To C

T8T9

x8: Fixed valuex9

30 40 50 60 7020

22

24

26

28

Regeneration temperature T oC

Hum

idity

rat

iox

g/kg

(DA

)

x10x11

Δ

Fig. 10 – Effect of regeneration temperature on the dehumidification performance and air states at each point.


0

10

20

30

40

Tem

pera

ture

To C

T4: Fixed valueT5

8

10

12

14

16

Hum

idity

rat

iox

g/kg

(DA

)

0

10

20

30

40

Tem

pera

ture

To C

T2: Fixed valueT3

x2: Fixed valuex3

Simulation

Experiment

6

8

10

12

14

Hum

idity

rat

iox

g/kg

(DA

)

x4

x5

0

2

4

6

8

10

Hum

idity

dif

fere

nce

xg/

kg(D

A)

0 5 10 15 2020

30

40

50

60


Tem

pera

ture

To C

T10: Fixed value T11

18

20

22

24

26

Hum

idity

rat

io

xg/

kg(D

A)

20

30

40

50

60

Tem

pera

ture

To C

T8: Fixed value T9

x8: Fixed value x9

0 5 10 15 2020

22

24

26

28


Hum

idity

rat

iox

g/kg

(DA

)

x10

x11

Fig. 11 – Effect of rotational speed on the dehumidification performance and air states at each point.


5.4. Effect of cooling temperature

Fig. 12 shows the effect of the cooling temperature on the

dehumidification performance and air states at each

measuring point with the fixed desiccant wheel rotational

speed, regeneration temperature, ambient air inlet humidity

ratio, and process and regeneration air inlet velocity shown

in Table 3. The process air can be dehumidified by the

regeneration air whose temperature is approximately 43 �C,

even though the cooling temperature is 18 �C; this tempera-

ture is significantly higher than the cooling temperature

(approximately 10 �C) of the mechanical dehumidification

system. In addition, the tendency of the performance

changes when the cooling temperature is approximately

17 �C. This is because the temperature of the process air

reaches the dew point below approximately 17 �C so that

some water is removed from the air as the condensed water

in the first cooling process.

5.5. Effect of process air volume flow rate

Fig. 13shows the effect of the processair volumeflow rate on the

dehumidification performance. When the process air volume

flow rate of the process air path and regeneration air path is

10 15 20 25 3020

30

40

50

60

Cooling temperature TcooC

Tem

pera

ture

To C T10: Fixed value

T11

18

20

22

24

26

Hum

idity

rat

io

xg/

kg(D

A)

x8: Fixed valuex9

20

30

40

50

60

Tem

pera

ture

To C

T8: Fixed valueT9

10 15 20 25 3018

20

22

24

26

Cooling temperature TcooC

Hum

idity

rat

io

xg/

kg(D

A)

x10

x11

0

10

20

30

40

Tem

pera

ture

To C T4

T5

6

8

10

12

14

Hum

idity

rat

io

xg/

kg(D

A)

0

10

20

30

40

Tem

pera

ture

To C

T2

T3

x2

x3

4

6

8

10

12

Hum

idity

rat

io

xg/

kg(D

A)

x4

x5

0

2

4

6

8

10

Hum

idity

dif

fere

nce

xg/

kg(D

A)

0

2

4

6

8

Rot

atio

nal s

peed

Nrp

h

Simulation

Experiment

Rotational speed: Fixed value

Fig. 12 – Effect of cooling temperature on the dehumidification performance and air states at each point.


changed simultaneously, the rotational speed, cooling

temperature, regeneration temperature, ambient air inlet

humidity ratio shown in Table 3 is fixed, the humidity difference

has a maximum point as the effect of the fixed rotational speed.

5.6. Effect of ambient air inlet humidity

Fig. 14 shows the effect of the ambient air inlet humidity ratio

on the dehumidification performance with the fixed the

rotational speed, cooling temperature, regeneration temper-

ature, and process and regeneration air inlet velocity shown in

Table 3. The humidity difference between the process air inlet

and outlet state becomes wide as the inlet humidity ratio of

the ambient air decreases.

5.7. Validity of the model

The simulation results are in complete agreement with the

experimental results. Therefore, the validity of the model is

confirmed.

6. Simulation method of conventionalcompression-type refrigerator and hybridair-conditioning system

The performance of the hybrid air-conditioning system

with the four-partition desiccant wheel is compared with

the conventional vapour compression-type refrigerator by

0 50 100 150 2000

2

4

6

8

10

Process air volume flow rate V m3/h

Simulation Experiment

Hum

idity

dif

fere

nce

x g/

kg(D

A)

Δ

Fig. 13 – Effect of process air volume flow rate on the

dehumidification performance.

Processair

Exhaust air

Compressor

Condenser 1Condenser 2

Expansion valve

Supply air

Evaporator

Outdoorair

Fig. 15 – Schematic representation of compression-type

refrigerator as mechanical dehumidification.

hyb–eva hyb–proQ G h h (32)


the simulation. Fig. 15 shows the flow diagram of the

conventional compression-type refrigerator as mechanical

dehumidification, which reheats the process air by its

condensation heat. The numbers in Fig. 15 correspond with

the numbers in Fig. 4. To evaluate the performance of each

system, pressure drops in the evaporator and the condenser

of the compression-type refrigerator are not considered.

Further, the adiabatic efficiency of the compressor stays

constant at 0.8. Simulation conditions and input air condi-

tions of the compression-type refrigerator are given in Tables

2 and 3, respectively. The COP of the conventional

compression-type refrigerator and the hybrid air-condi-

tioning system is defined as

COPmde ¼Qmde�eva

Wmde�com(29)

COPhyb ¼Qhyb�eva

Whyb�com(30)

For the hybrid air-conditioning system, the regeneration

and the cooling temperatures of the desiccant wheel corre-

spond with the condensation and evaporation temperatures

with the approach temperature, respectively.

The cooling heat of the conventional compression-type

refrigerator as shown in Figs. 4 and 15 is

( )mde–eva mde–proQ G h h (31)

12 15 18 21 240

2

4

6

8

10

Ambient air humidity ratio x g/kg(DA)

Simulation Experiment

Hum

idity

dif

fere

nce

x g/

kg(D

A)

Δ

Fig. 14 – Effect of ambient air inlet humidity ratio on the

dehumidification performance.

The cooling heat of the hybrid air-conditioning system as

shown in Figs. 2 and 4 is Silica gel is used as the desiccant.

The rotational speed of the desiccant wheel has the optimum

point to maximize the system performance. The conditions of

the ambient air are assumed to be the one in summer in

Tokyo. Moreover, the approach temperature between the

refrigerant and the air in the heat exchanger remains

constant, for instance, the evaporator approach temperature,

evaporator superheat temperature, and condenser approach

temperature are 5 �C. SHF (sensible heat factor) is the sensible

heat load per all the heat load in the room. Exhaust air

temperature in Fig. 2 is also decided adopting the approach

temperature by ambient air.

7. Simulation results for hybridair-conditioning system

7.1. Effect of SHF

Fig. 16 shows the effect of SHF on the characteristics and

performance of the system, for instance, the refrigerant

condensation temperature, the regeneration air temperature,

the refrigerant evaporation temperature (the refrigerant evap-

oration temperature needed for the mechanical dehumidifica-

tion is also shown), the heat in the condenser and the

regeneration air, the compressor power, and COP (COP of the

mechanical dehumidification is also indicated). In this simula-

tion, SHF is changed by the humidity ratio of the process air

outlet, point . The temperature of points , , and is equal;

this temperature is 18 �C. When SHF is more than 0.59 (Region

A), the condensation temperature remains constant. This is

because the regeneration air temperature at points and is

lower than the setting temperature of the exhaust air at point .

Consequently, the condensation temperature cannot be

reduced even though the regeneration temperature decreases.

The evaporation temperature also remains constant to make

the temperature, point , of the process air constant. Hence, the

COP becomes constant.

30

40

50

60

Tem

pera

ture

To C Regeneration air ,

Exhaust air Condensation

0

5

10

15

20

Eva

pora

tion

tem

pera

ture

To C

Hybrid Mechanical

dehumidification

0.4 0.5 0.6 0.70

0.5

1

1.5

2

Hea

t tra

nsfe

r ra

teQ

kW

SHF

Condenser Regeneration air

Region ARegion B

0

0.1

0.2

0.3

0.4

0.5

Com

pres

sor

pow

er

Wco

mkW

2

3

4

5

6

7

8

CO

P

Hybrid Mechanical

dehumidification

0.4 0.5 0.6 0.720

40

60

80

100

SHF

Eff

icie

ncy

impr

ovem

ent

%

Region ARegion B

Fig. 16 – Effect of SHF.


In contrast, the COP of the mechanical dehumidification

increases with increasing SHF. Therefore, the efficiency

improvement of the hybrid air-conditioning system decreases

with increasing SHF. When SHF is less than 0.59 (Region B),

owing to the increase in the latent heat to regenerate the

desiccant, the regeneration and the condensation temperature

need to be raised. Thus, the compressor power increases, and

the COP decreases with decreasing SHF. The evaporation

temperature becomes lower on the mechanical dehumidifi-

cation as compared to the hybrid system when SHF decreases;

therefore, the COP of the mechanical dehumidification

decreases with decreasing SHF. Further, the efficiency

improvement of the hybrid system decreases. All the

condensers have sufficient heat to regenerate the desiccant.

This implies that the hybrid system can be driven by only the

waste heat from the compression-type refrigerator without

any other heat source. The COP of the hybrid system is higher

than the mechanical dehumidification shown in the psycho-

metric chart of Fig. 4 and flow diagram of Fig. 14. For example,

when SHF is 0.59, the COP of the hybrid system is approxi-

mately 94% higher than that of the conventional mechanical

dehumidification system.

7.2. Effect of process air cooling temperature

Fig. 17 shows the effect of the cooling temperature at points

and on the refrigerant condensation temperature, the

regeneration air temperature, the refrigerant evaporation

temperature, the heat transfer rate, the compressor power, and

the COP with SHF–0.5. When the cooling temperature at points

and increases to more than 18 �C (Region A) on the

condition that the evaporation temperature and the supply air

temperature at point remain constant, the condensation

temperature increases because the regeneration air tempera-

ture also rises by the influence of the cooling temperature at

points and . Thus, the compressor power increases and

COP decreases.

When the temperature at points and decreases, the

trend of each parameter in Fig. 17 shifts at the point of the

cooling temperature 18.0 �C. In region B, below the temperature

of the shift point, the cooling temperature is lower than the

supply air temperature to the room so that the evaporation

temperature is reduced with the decrease in the cooling

temperature. At the same time, the drop in the regeneration

temperature because of the influence of the fall in the cooling

temperature decreases the condensation temperature. Hence,

the COP and the efficiency improvement are almost constant. In

case ofregion C shiftedat the point of the cooling temperature of

approximately 17 �C, the cooling temperature at point rea-

ches the dew point temperature, and then, the moisture of the

process air is removed in evaporator 1. The regeneration

temperature at points and decreases because of the

reduction in the latent heat in the regeneration process of the

desiccant wheel. Therefore, the condensation temperature

further decreases, and the COP increases.

The trend of each parameter also changes at the point of

the cooling temperature of approximately 15.2 �C. The

regeneration air temperature is lower than the constant

30

40

50

60

Tem

pera

ture

To C

Regeneration air , Exhaust air Condensation

0

5

10

15

20

Eva

pora

tion

tem

pera

ture

To C

Hybrid Mechanical

dehumidification

14 16 18 20 22 24 26 280

0.5

1

1.5

2

Process air cooling temperature T ,oC

Condenser Regeneration air

Region A

Region CRegion B

Region D

Hea

t tra

nsfe

r ra

teQ

kW

0

0.1

0.2

0.3

0.4

0.5

Com

pres

sor

pow

erW

com

kW

2

3

4

5

6

7

8

CO

P

Hybrid Mechanical

dehumidification

14 16 18 20 22 24 26 2820

40

60

80

100

Process air cooling temperature T ,oC

Eff

icie

ncy

impr

ovem

ent

%

Region ARegion B

Region CRegion D

Fig. 17 – Effect of process air cooling temperature.


exhaust air temperature below the cooling temperature of

approximately 15.2 �C. When the regeneration temperature

at points and decreases, the condensation temperature

cannot be reduced and remains constant because of the

constant exhaust air temperature. The COP decreases

because of the reduction in the evaporation temperature

with the decrease in the cooling temperature. When the

cooling temperature is approximately 15.2 �C, the COP of

the hybrid air-conditioning system becomes the highest. At

this point, the COP of the hybrid system is approximately

94% higher than that of the conventional compression-type

refrigerator whose COP is constant. Therefore, the COP of

the hybrid system improves considerably when the evapo-

rator removes a certain amount of the latent heat load.

This implies that the reduction in the condensation

temperature has more impact on the COP than the reduc-

tion in the evaporation temperature.

8. Conclusions

This paper evaluates the characteristics of the four-partition

desiccant wheel by experiment and simulation, and the

performance of the hybrid air-conditioning system that

combines the four-partition desiccant wheel with the

compression-type refrigerator by the simulation. The results

are summarized as follows:

1) Experimental results of the four-partition desiccant wheel

were in good agreement with the simulation results even

though the experimental conditions changed greatly.

Therefore, the validity of the simulation was confirmed.

2) The performance of the four-partition desiccant wheel was

clarified indetailby investigatingtheeffectof theregeneration

temperature, rotational speed, cooling temperature, process

air volume flow rate, and ambient air inlet humidity ratio.

3) The dehumidification technique using the hybrid system

with the four-partition desiccant wheel can be driven more

efficiently than the mechanical dehumidification using the

compression-type refrigerator. When the evaporator

burdens a certain amount of the latent heat, the COP of the

hybrid system improves considerably. For example, the COP

of the hybrid system is approximately 94% higher than that

of the conventional mechanical dehumidification system.

9. Future plan

The pressure drop of the hybrid system with four-partition

desiccant wheel is considered larger than the conventional

system due to multiple air flow paths. Therefore, it is necessary

to consider the pressure loss for the efficiency of the hybrid

system. Now, we are investigating the influence of pressure

drop. We will report a detailed effect of pressure drop on four-

partition desiccant system as well as the conventional system.


r e f e r e n c e s

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