performance analysis of desiccant dehumidification systems driven by low-grade heat source

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 4 ( 2 0 1 1 ) 9 2 8e9 4 5

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Performance analysis of desiccant dehumidification systemsdriven by low-grade heat source

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 27 October 2009

Received in revised form

3 September 2010

Accepted 4 October 2010

Available online 10 October 2011

Keywords:

Desiccant wheel

Optimization

Dehumidification

Desiccant

Air conditioning

Silica gel

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

0140-7007/$ e see front matter ª 2010 Elsevdoi:10.1016/j.ijrefrig.2010.10.001

a b s t r a c t

If a desiccant dehumidification system can be driven by a heat source whose temperature is

below 50 �C, exhaust heat from devices such as fuel cells or air conditioners can be used as

its heat source, thereby saving energy. Therefore, in this study, we used a previously vali-

dated simulation model to determine the minimum heat source temperature for driving

a desiccant dehumidification system. We considered four desiccant dehumidification

systems that can be driven by waste heatdconventional desiccant-type systems (wheel

type and batch type with only desiccant), a system with a precooler, double-stage-type

systems (a type with two desiccant wheels and a four-partition desiccant wheel type), and

a batch-type system with an internal heat exchanger. We found that among these systems,

the last system can be driven by the lowest heated air temperaturedapproximately

33 �Cdwhich is considerably lower than that of the conventional system.

ª 2010 Elsevier Ltd and IIR. All rights reserved.

Analyse de la performance des systemes a deshumidificationa deshydratant avec une source de chaleur a bassetemperature

Motscles : Roue deshydratante ; Optimisation ; Deshumidification ; Deshydratant ; Conditionnement d’air ; Gel de silice

1. Introduction

In order to achieve energy savings with the use of

a compression-type room air conditioner, its evaporation

temperature must be increased by increasing the setting

3259.Jeong).ier Ltd and IIR. All rights

temperature. However, due to this increase in the evapora-

tion temperature, not only the temperature but also the

humidity of the room increases, leading to a degradation of

the indoor environment. Hence, to maintain a comfortable

indoor environment, a dehumidification system is required.

reserved.

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http://dx.doi.org/10.1016/j.ijrefrig.2010.10.001



Nomenclature

Aal-in area of aluminium surface in contact with water

(m2)

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 (m)

dh hydraulic diameter (m)

h specific enthalpy (kJ kg�1)

jm mass flux between air and desiccant wall surface

(kgm�2 s�1)

Kh overall heat transfer coefficient (Wm�2 K�1)

Km overall mass transfer coefficient (kgm�2 s�1)

Le Lewis number (e)

L desiccant length along air flow path (m)

le entrance region length (m)

lh pitch distance between flat walls (m)

lp wavelength of corrugation (m)

lr tube length along water flow path (m)

ma mass fraction of water vapour in moist air

(kg kg�1)

mb mass fraction of water vapour of desiccant wall at

equilibrium (kg kg�1)

N rotational speed (rph)

Nu Nusselt number (e)

Nut nondimensionaloverall heat transfercoefficient (e)

Pr Prandtl number (e)

Q heat transfer rate (kW)

qs heat flux between air and desiccant wall surface

(kWm�2)

Re Reynolds number (e)

Sc Schmidt number (e)

Sh Sherwood number (e)

Sht nondimensional overall mass transfer coefficient

(e)

St switching time (s)

T temperature (�C)t time (s)

tb thickness of corrugated sheet (m)

u velocity (m s�1)

V volume (m3)

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

x humidity ratio (g kg�1 (DA))

y y axis (e)

z z axis (e)

aa heat transfer coefficient between air and

desiccant surface (Wm�2 K�1)

ain heat transfer coefficient between tube surface and

water (Wm�2 K�1)

b mass transfer coefficient between air and

desiccant wall surface (m s�1)

q angle of desiccant wheel (rad)

l thermal conductivity (Wm�1 K�1)

h fin efficiency (e)

r density (kgm�3)

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

Subscripts

a moist air in air channel

ads adsorption

al aluminium

b desiccant bed

co cooled

coi cooling water inlet

he heated

hei heating water inlet

in inside

ini initial

out outside

pi process air inlet

po process air outlet

r water

ri regeneration air inlet

vap vaporization

w water

i n t e rn 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 4 ( 2 0 1 1 ) 9 2 8e9 4 5 929

A desiccant dehumidification method can dehumidify

air by converting latent heat into sensible heat, making

it unnecessary to supercool and then reheat the air in

amechanical dehumidification systemsuchas a compression-

type air conditioning system (ASHRAE, 2001).

Recently, solid desiccant air conditioning systems have

been attracting attention because they can be driven by solar

energy, waste heat, etc. In the case of solar energy as the

driving heat source, Mavroudaki et al. (2002) andHalliday et al.

(2002) independently conducted two feasibility studies of solar

driven desiccant cooling in diverse European cities repre-

senting different climatic zones. Normally, a heat source with

a temperature of at least 60e80 �C is required to drive

a desiccant air conditioning system (Harriman, 1994; Meckler,

1994). However, this heat level cannot always be easily

obtained. For a hybrid air conditioning systemwith desiccant,

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

investigated the performance of a hybrid desiccant cooling

system comprising a conventional vapour compression-type

system coupled with a desiccant dehumidifier. However,

because a conventional single stage desiccant was used for

this system, the condensation temperature of the vapour

compression-type refrigerator increased greatly. Moreover,

sometimes, an electric heater is also used to compensate for

;the shortage of driving heat. Such a system cannot increase

the system performance. To achieve a high performance

desiccant air conditioning system that can utilize various

types of heat sources, it is necessary to decrease the driving

heat source temperature for the desiccant regeneration.

Exhaust heat such as that of a compression-type refrigerator,

whose temperature is about 40e50 �C, commonly exists

everywhere. Exhaust heat is generally considered to be waste

heat; the utilization of all this waste heat to drive a desiccant

dehumidification system would lead to large energy savings.

With this background, we previously investigated

approaches to reduce the temperature of the heat source of



Fig. 1 e Schematic flow diagram and psychrometric chart of conventional dehumidification systems (wheel type and batch

type with only desiccant).


a desiccant dehumidification system. For example, we devel-

oped a four-partition desiccant wheel and investigated the

performance of a desiccant dehumidification system using

this wheel (Inagaki et al., 2004; Shibao et al., 2006, 2007).

Consequently, we experimentally determined that this

system could be driven by a lower-temperature heat source

of approximately 40e50 �C. We also successfully constructed

an extremely high efficiency air conditioning system by

combining the desiccant dehumidification system with the

Fig. 2 e Schematic flow diagram and psychrometric c

four-partition desiccant wheel with a compression-type

refrigerator. However, performance analyses of desiccant

dehumidification systems driven by low-grade heat sources

have not yet been clarified. Hence, we wanted to investigate

some other types of desiccant dehumidification systems

that can be driven by lower-temperature waste heat.

In this study, we analyzed the performances of four

desiccant dehumidification systems that can be driven by

a low-grade heat source: conventional desiccant-type systems

hart of dehumidification system with precooler.




(wheel type and batch type with only desiccant), a system

with a precooler, double-stage-type systems (a type with two

rdesiccant wheels and a four-partition desiccant wheel type)

and a batch-type system with an internal heat exchanger. For

these analyses, we used a simulation model whose validity

was confirmed experimentally by Yamaguchi et al. (2007).

2. Low-temperature-driven desiccantdehumidification systems

Figs. 1e4 show schematic flow diagrams and psychrometric

charts of the four desiccant dehumidification systems. For

calculation purposes, we assumed that the process air of these

systems was obtained directly from the room and not from

ambient air. We also assumed that the process air was cooled

by a compression-type refrigerator.

2.1. Conventional dehumidification systems (wheel typeand batch type with only desiccant)

Fig. 1(a) and (b) shows the schematic flow diagram and

psychrometric chart of a wheel-type system, respectively.

This type consists of a desiccant wheel, a cooler, and a heater.

Fig. 3 e Schematic flow diagram and psychrometric chart of doub

a four-partition desiccant wheel type).

A sensible heat exchanger is usually employed with this

system. This heat exchanger does not affect the heated air

temperature while driving the system. Thus, we ignored this

heat exchanger in the calculations.

Fig. 1(c) and (d) respectively shows the schematic flow

diagram and psychrometric chart of a batch-type systemwith

two fixed desiccant beds. These beds are alternately switched,

thereby causing the process air and regeneration air to alter-

nately flow into each bed. When the total heat and mass

transfer area of these two beds are the same as those of the

desiccant wheel, the time-averaged simulation results of

this system are almost the same as those of the wheel type.

We define these two types of desiccant systems as conven-

tional dehumidification systems.

2.2. Dehumidification system with precooler

Fig. 2(a) and (b) respectively shows the schematic flow

diagramand psychrometric chart of a systemwith a precooler.

The process air cools down in the precooler before entering

the desiccant wheel. Due to this cooling action, the relative

humidity of the process air increases; thus, the relative

humidity of the dehumidified process air increases and

its temperature decreases.

le-stage-type systems (type with two desiccant wheels and



Fig. 4 e Schematic flow diagram and psychrometric chart of batch-type dehumidification system with internal heat

exchanger.


2.3. Double-stage-type system (a type with twodesiccant wheels and a four-partition desiccant wheel type)


diagram and psychrometric chart of a system with two

desiccant wheels. A characteristic of this system is that in

addition to precooling, the other cooling processes occur after

the dehumidification processes. On the regeneration side, the

heating processes occur before the regeneration processes.

The heated air temperature can be reduced further by these

heating and cooling processes.

Fig. 3(c) and (d) respectively shows the schematic flow

diagram and psychrometric chart of a four-partition desiccant

wheel system. The four-partition desiccant wheel is divided

into equal areas for each flow path. The results of the simu-

lation performed in this study showed that the performance

and heated air temperature of this system were almost the

same as those of the abovementioned two desiccant wheels.

The advantage of this type is that its size and cost can be

reduced to less than the type with two desiccant wheels,

because the double-stage process can be carried out using

just one desiccant wheel.

Fig. 5 e Photo of de

2.4. Batch-type dehumidification system with internalheat exchanger


diagram and psychrometric chart of a batch-type dehumidi-

fication system with an internal heat exchanger. Generally,

the relative humidity of the process air after the dehumidifi-

cation process decreases but the temperature of this air

increases. Thus, the heated air temperature increases.

Otherwise, in this system, the process air in the dehu-

midification process is cooled down directly by the internal

heat exchanger and the regeneration air in the regeneration

process is also heated up directly by the internal heat

exchanger. This internal heat exchanger decreases the heated

air temperature considerably. The heat exchanger has corru-

gated fins and tubes to which the desiccant is attached.

3. Description of test facility

Fig. 5 shows a photo of the desiccant wheel used in the

experiment. The experimental facility was divided into three

siccant wheel.



Fig. 6 e Test facility layout (T: Thermocouple, X: Dew point meter, F: Flow meter).


main parts, as shown in Fig. 6: the condition generator for

the process air, the condition generator for the regeneration air

and the test section. The condition generators for the process

and regeneration air are composed of cooling coils for dehu-

midification, vapour tubes for humidification, heating coils for

temperature regulation, and cooling coils for temperature

regulation. In addition, the condition generator for the regen-

eration air needed a heater for temperature regulation. For the

process and regeneration air paths, ambient air sent by blowers

flows into the air temperature and humidity condition gener-

ators, which subsequently achieve the target air temperature

and humidity as the inlet conditions of the test section. In the

test section, the process air and regeneration air are respec-

tively dehumidified and humidified by the desiccant bed. After

the test section, the air is released to the atmosphere. Heat

sources for the cooling and heating are provided by a chiller

and boiler, respectively. Therefore, these condition generators

make it possible to supply air under stable conditions all year.

The analysis of the performance of each type of desiccant

system under consideration was based on the heat and mass

balances in the desiccant media, so the following measure-

ments were performed:

� Inlet and outlet air temperature (�C) of process air;

� Inlet and outlet air humidity (g kg�1 (DA)) of process air;

� Inlet and outlet air temperature (�C) of regeneration air;

� Inlet and outlet air humidity (g kg�1 (DA)) of regeneration air;

� Outlet volumetric flow rates (m3 s�1) of process and regen-

eration air;

� Rotational speed (rph) or switching time (t) of desiccant;

� Inlet and outlet water temperature (�C) of cooler;� Inlet and outlet water temperature (�C) of heater;� Outlet water volumetric flow rate (m3 s�1) of cooler;

� Outlet water volumetric flow rate (m3 s�1) of heater.

The following sensors and instruments were used:

� Thermocouple (type T, class 1, accuracy �0.2 �C) and plat-

inum resistance temperature sensors (Pt100, grade A, 100

UG, 0.10% at 0.8 �C) were used for air and water temperature

measurements, respectively.

� Dew point meter (�0.2 �C accuracy in a temperature range

of �80 to 100 �C, with a 4e20 mA output) for humidity

ratio measurement.

� Annubar flow meter (�1.0% accuracy in a range ability of

10:1 for air) and oval flow meter (�1.0% accuracy at a flow

rate of 3e20 lmin�1 for water) for volumetric flow rate

measurements.

� Electrical parameter measuring apparatus (0.20% accuracy)

for electrical energy output measurements, including

voltage, power and frequency.

A data acquisition logger read and stored data about every

10 s.



Fig. 7 e Detailed structure and model of desiccant wheel.


4. Simulation model of desiccant wheel

Fig. 7 shows the detailed structure and model of a desiccant

wheel. This desiccant wheel consists of corrugated and flat

walls that contain silica gel as a desiccant. The structure of the

desiccant media is usually realized by an arrayal made up of

a corrugated lamina and a plane consisting of sheets of glass

fibre impregnated with the actual desiccant media, HY-SG.

The air flows through the channels between these walls. The

detailed configurations of all types of desiccant media are

listed in Table 1. The size of the desiccant wheel and the

surface area per air flowpath and desiccant volume are shown

in Table 2. This simulation model was based on the static

model of Yamaguchi et al. (2007), the validity of which they

confirmed experimentally. Here, we used this model in order

to reduce the calculation time considerably.

4.1. Assumptions

We made the following assumptions when constructing the

simulation model:

Table 1 e Structural specifications and property values ofdesiccant media.

Parameters Symbols Units Values

Thickness of corrugated sheet tb mm 0.2

Wavelength of corrugation lp mm 3.8

Pitch distance between flat walls lh mm 1.9

Channel hydraulic diameter dh mm 1.5

Desiccant e Silica gel

Apparent density of desiccant kgm�3 800

Specific heat of desiccant J kg�1 K�1 980

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

ing into the channel also moves along the q direction.

However, the movement of air along this direction is

considerably slower than that in the z direction. Therefore,

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

� As the desiccant wall is thin, the heat and mass transfer

resistances in the z and q directions for the inner desiccant

walls can be neglected. In other words, the heat and

mass transfers are those by forced convection transfer and

diffusion, respectively, between the air and desiccant wall,

and are considered only in the direction of airedesiccant.

� The entrance region length (le) can be neglected because it

is considerably shorter than the length of the desiccant

wheel (L).

� The pressure drop of the air in the air channel is not

considered.

4.2. Mathematical model

Fig. 7 shows the simulation model of the desiccant wheel.

For the air side, the continuity equation for all of the moist

air is:

VavðrauaÞ

vzþ jm

�Af þ Ab

� ¼ 0 (1)

where Va is the volume of the air side and AfþAb is the total

surface area of the desiccant wall in direct contact with the air

in the control volume shown in Fig. 7. The air velocity, ua,

implies the mean velocity in the air channel.

The continuity equation of moisture is:

VavðrauaxaÞ

vzþ jm

�Af þ Ab

� ¼ 0 (2)



Table 2 e Specifications of desiccants used in each type of system.

Type of system Parameters Units Values

Conventional

desiccant-type

system

Wheel type Size of desiccant wheel m 40.25� 0.20

Surface area per volume of air flow path (AbþAf)/Va m2m�3 3000

Surface area per volume of desiccant (AbþAf)/Vb m2m�3 6000

Batch type Size of desiccant bed m 0.16� 0.20� 0.11



System with precooler Size of desiccant wheel m 40.25� 0.20



Double-stage-type

system

Two desiccant wheels

type

Size of each desiccant wheel m 40.25� 0.10



Four-partition desiccant

wheel type

Size of desiccant wheel m 40.25� 0.20



Batch-type system

with internal heat

exchanger

Size including desiccant and internal heat exchanger m 0.16� 0.20� 0.23



Surface area inside tube per volume of internal heat

exchanger Aal_in/Val

m2m�3 1700

Aluminium heat transfer area ratio Aal_in/(AfþAb) e 0.43

Hydraulic diameter of flow inside tube mm 1.2

Thickness of tube mm 0.4

Thickness of corrugated fin mm 0.2

Density of aluminium channel kgm�3 2700

Heat capacity of aluminium channel J kg�1 K�1 921


And the energy equation of air is:

VavðrauahaÞ

vzþ �

jmhvap þ qs

��Af þAb

� ¼ 0 (3)

The continuity equation of the entire desiccant is:

Vbuvrbvq

� jm�Af þ Ab

� ¼ 0 (4)

where Vb is the volume of the desiccant wall in the control

volume shown in Fig. 7.

The continuity equation of water in the desiccant is:

VbuvðrbXbÞ

vq� jm

�Af þAb

� ¼ 0 (5)

The energy equation of the desiccant is:

VbuvðrbhbÞ

vq� �

jmhads þ qs

��Af þ Ab

� ¼ 0 (6)

The mass and heat transfer rates are:

15 20 25 30 350.5

1

5

10

50

Non

-dim

ensi

onal

he

at tr

ansf

er c

oeff

icie

ntN

u t

Rotational speed N rph

Regeneration temperature: 80 : 70 : 60

2.12

Fig. 8 e Nondimensional heat and mass transf

jm ¼ Kmðma �mbÞ (7)

qs ¼ KhðTa � TbÞ (8)

where Ta and Tb represent the average temperature of the

control volume in Fig. 7.

We derived the overall mass and heat transfer coefficients

Km and Kh experimentally. Fig. 8 shows the nondimensional

heat transfer coefficient, Nut, and the mass transfer coeffi-

cient, Sht. These coefficients are defined as follows:

Nut ¼ Khdh

la(9)

Sht ¼ Kmdh

raDa(10)

where dh is the hydraulic diameter. The heat and mass

transfers between the air and desiccant walls were divided

15 20 25 30 350.5

1

5

10

50

Non

-dim

ensi

onal

m

ass

tran

sfer

coe

ffic

ient

Sh t

Rotational speed N rph

Regeneration temperature: 80 : 70 : 60

2.12

er coefficients determined experimentally.



Table 3 e Simulation and experimental conditions forvalidity of simulation model for single desiccant wheel.

Parameters Units Values

Dehumidification

inlet

Process air temperature �C 32.5

Process air relative

humidity

% 63.7

Process air humidity

ratio

g kg�1

(DA)

19.5

Process air mean

velocity

m s�1 2.8

Dehumidification

outlet

Cooled air temperature �C 18.0

Regeneration

inlet

Heated air temperature �C 50e80

Heated air humidity

ratio

g kg�1

(DA)

11.9

Heated air relative

humidity

% 4e15.4

Heated air mean

velocity

m s�1 2.8

Rotational speed of desiccant wheel rph 60



into forced convection transfer and diffusion, respectively,

inside the desiccant walls.

Shah (1975) investigated the forced convectionheat transfer

in a small air channel. In his study, he derived the Nusselt

number,Nu, for laminar and fully developed flow conditions at

a constant surface temperature. This number is given by:

Nuhaadh=la ¼ 2:12 (11)

The Sherwood number, Sh, is the same as Nu when the Lewis

number, Le, is taken as 1.0.

Sh ¼ Nu (12)

Thus, Sh becomes:

Shhbdh=Da ¼ 2:12 (13)

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

expressed as:

le=dh ¼ 0:05Re� ðPr or ScÞ (14)

Under normal driving conditions for the desiccant wheel, le is

approximately 5% of the length of the desiccant wheel. Thus,

le can be neglected.

As shown in Fig. 8, the values of Nut and Sht are almost

identical to those of Nu and Sh. This implies that the resis-

tance of the heat and mass transfers is governed by the air

side of the desiccantwheel. This is considered to be the reason

that the thickness of the desiccant wall is considerably less

than the size of the air channel. Therefore,

Nut ¼ Nu (15)

Sht ¼ Sh (16)

We adopted Shah’s Nu and Sh in the simulation in this paper

because we can say that we confirmed Nut and Sht by experi-

ment, as shown in Fig. 8.

For the properties of silica gel, the heat of adsorption, hads,

is expressed as a function of the mass fraction of water in

the desiccant. Barlow (1982) described the ratio of the heat of

adsorption to the heat of vaporization, hvap, and the adsorp-

tion isotherm of silica gel. The desiccant wall does not contain

just pure silica gel. Therefore, its adsorption isotherm is

corrected by the content ratio of this impure silica gel con-

taining othermaterials. In this study, 0.7 is used as the content

ratio of the silica gel in the desiccant. Other property values of

the desiccant are listed in Table 1.

4.3. Boundary conditions

Conditions related to the regeneration air and process air (flow

rate, humidity ratio and temperature) at the inlet of the

desiccant wheel are given as boundary conditions.

0 � q < p, z¼ 0 (process air at inlet):

Tair ¼ Tpi (17)

xair ¼ xpi (18)

uair ¼ upi (19)

p � q < 2p, z¼ L (regeneration air at inlet):

Tair ¼ Tri (20)

xair ¼ xri (21)

uair ¼ uri (22)

4.4. Validity of simulation model

The detailed specifications of the tested single desiccant media

are shown in Table 1.We conducted experiments using a desic-

cantwheelwithadiameterof350mmandathicknessof50mm.

The surface area per volume of air flow path and desiccant are

shown in Table 2. Table 3 shows the experimental and simula-

tion conditions for the validity of the simulation model of

a single desiccant wheel. The condition generator prepares the

experimental conditions for the desiccant bed inlet. The

temperature and humidity ratio at the dehumidification inlet

and regeneration inlet respectively represent the outdoor and

indoor conditions in Tokyo during the summer season.

Fig. 9 shows the validity results for the simulation and

experiment for the single desiccant wheel. As shown in this

figure, the simulation results agreed well with the experi-

mental results, with errors of approximately 2% for the

temperature and humidity. This confirms the validity of the

model for the desiccant wheel. We also confirmed the validity

of thismodel for the simulation of the four-partition desiccant

wheel, with errors of approximately 4% for temperature and

7% for humidity ratio between the simulation and experi-

ment, as shown in Fig. 10, using the simulation and experi-

mental conditions shown in Table 4. The configuration of the

four-partition desiccant wheel is described in Table 2.

5. Dynamic model of fixed desiccant bed

We used a dynamic model to simulate a fixed desiccant bed in

Fig. 11, which shows the detailed structure of the desiccant



45 50 55 60 65 70 75 80 8510

12

14

16

18

20

Air

hum

idity

rat

iox

g/kg

(DA

)

Heated air temperature T oC

45 50 55 60 65 70 75 80 8530

40

50

60

70

80

Air

tem

pera

ture

To C


Simulation, Outlet Experiment, Outlet

Diameter of desiccant wheel : 350mmThickness of desiccant wheel : 50mmMean velocity of air in the air channel :2.8m s-1

Rotational speed of rotor : 60 rph Simulation, Inlet Experiment, Inlet

Fig. 9 e Simulation and experimental results for validity of simulation model for single desiccant wheel.


bed. This desiccant bed also consists of corrugated and flat

walls that contain silica gel, just as with the desiccant wheel

type. The structure of the desiccant bed is usually realized

by an arrayal made up of a corrugated lamina and a plane

consisting of sheets of glass fibre impregnated with the actual

desiccant, HY-SG. The air flows through the channels between

these walls. The structural specifications of the desiccant

media are listed in Table 1. The size of the desiccant bed

and the surface area per air flow path and desiccant volume

are shown in Table 2.

5.1. Assumptions

We made the following assumptions when constructing the

simulation model:

� The heat and mass transfer resistances in the z and y

directions inside the desiccant walls are not considered.

� The entrance region length can be neglected because it is

considerably shorter than the length of the desiccant bed (L).

� The pressure drop of the air in the air channel is not

considered.

30 40 50 60 7010

20

30

40

50

Air

tem

pera

ture

To C T4

T5


10

20

30

40

50

Ait

tem

pera

ture

To C T2

T3 Simulation

Experiment

Fig. 10 e Simulation and experimental results for validity o


The continuity equation of moist air and the mass conserva-

tion of water vapour in the moist air are expressed as:

Vavravt

þ VavðrauaÞ

vzþ jm

�Af þ Ab

� ¼ 0 (23)

VavðraxaÞ

vtþ Va

vðrauaxaÞvz

þ jm�Af þAb

� ¼ 0 (24)

The energy equation of moist air is

VavðrahaÞ

vtþ Va

vðrauahaÞvz

þ �jmhvap þ qs

��Af þ Ab

� ¼ 0 (25)

The continuity equation and energy equation in the desiccant

wall are:

Vbvrbvt

� jm�Af þ Ab

� ¼ 0 (26)

VbvðrbXbÞ

vt� jm

�Af þAb

� ¼ 0 (27)

6

8

10

12

14

Air

hum

idity

rat

iox

g/kg

(DA

)

x2x3

30 40 50 60 704

6

8

10

12

Air

hum

idity

rat

iox

g/kg

(DA

)

x4x5


f simulation model for four-partition desiccant wheel.



Table 4 e Simulation and experimental conditions forvalidity of simulation model for four-partition desiccantwheel.


Dehumidification

inlet

Process air

(No. 1 in Fig. 3)

temperature

�C 28.0

Process air

humidity ratio

g kg�1 (DA) 11.9


humidity

% 50.0

Process air mean

velocity in air channel

of desiccant.

m s�1 2.8

Dehumidification

outlet

Cooled air

temperature

(Nos. 2, 4 and 6

in Fig. 3)

�C 18.0

Regeneration

inlet

Heated air (Nos. 8

and 10 in Fig. 3)

temperature

�C 35.0e64.0

Heated air humidity

ratio

g kg�1 (DA) 19.5

Heated air relative

humidity

% 63.0

Process air mean

velocity in air channel

of desiccant

m s�1 2.8

Rotation speed

of desiccant

wheel

rph 4.0

Specific heat of

desiccant

J kg�1 K�1 980


VbvðrbhbÞ

vt� �

jmhads þ qs

��Af þAb

� ¼ 0 (28)

The heat flux, qs, between the air and desiccant wall is

zOne air channel

Air flow

y

Air flow

Adsorption

zy

0

Desorptio

zy

0

Fig. 11 e Dynamic mod

qs ¼ KhðTa � TbÞ (29)

where Ta and Tb represent the average temperatures of

each part in the control volume in Fig. 11.

The mass flux, jm, between the air and desiccant wall is


The overall heat and mass transfer coefficients, Kh and Km,

are derived from Nut and Sht, which are identical to those

of the desiccant wheel.


Conditions related to the regeneration air and process air

(flow rate, humidity ratio and temperature) at the inlet of

the desiccant bed are given as boundary conditions.

When the desiccant bed is used in the adsorption process,

z¼ 0 (process air at inlet):

Tair ¼ Tpi (31)

xair ¼ xpi (32)

uair ¼ upi (33)

When the desiccant bed is used in the desorption process,

z¼ L (regeneration air at inlet):

Tair ¼ Tri (34)

xair ¼ xri (35)

uair ¼ uri (36)

The initial conditions were

t¼ 0:

Tb ¼ Tb ini (37)

Air flow

n

Desiccant bed

qs jm

z Δ z

auaxa

Air Va

Vb

el of desiccant bed.




Xb ¼ Xb ini (38)

5.4. Validity of dynamic simulation model of fixeddesiccant bed

The detailed specifications of the tested single desiccant

media are shown in Table 1 and the desiccant size and surface

area per volume of air flow path and desiccant are shown in

Table 2. Table 5 shows the simulation and experimental

conditions for the validity of the simulation model for a fix-

ed desiccant bed. Fig. 12 shows the validity results for the

simulation and experiment obtained using the dynamic

model for the fixed desiccant bed. We believed that the

simulation model could be validated by a good agreement,

with errors of approximately 5% for the temperature and 2%

for the humidity ratio between the simulation and experiment

results in Fig. 12. As shown in this figure, the outlet temper-

ature and humidity ratio were not constant over time.

6. Model of fixed desiccant bed with internalheat exchanger

Fig. 13 shows a fixed desiccant bed with an internal heat

exchanger; this exchanger is the same as that described in

Section 2.4. The internal heat exchanger was made of

aluminium. The density and specific heat of aluminium are

2700 (kgm�3) and 900 (J kg�1 K�1), respectively. The air flows

through the channels between the desiccant walls. Structural

specifications of the desiccant side are listed in Table 1.

More details of the internal configuration of the aluminium

channel are shown in Table 2.

6.1. Assumptions

Most of the assumptions that were made when constructing

this model were the same as those mentioned in Section 5.1.

The other assumptions were as follows:

Table 5 e Simulation and experimental conditions forvalidity of simulation model for fixed desiccant bed.


Dehumidification

inlet



humidity

% 50.0

Process airmean velocity m s�1 3.0

Dehumidification

outlet


Regeneration

inlet

Heated air temperature �C 62.2

Heated air relative

humidity

% 14.0

Mean velocity m s�1 3.0

Switching time

for each

desiccant bed

T 420

Specific heat of

desiccant

J kg�1 K�1 980

� The temperature of the corrugated fins and tubes is the

same as that of the desiccant bed.

� The heat transfer performance of the corrugated fins and

the thermal contact resistance between the corrugated

desiccant bed and aluminium fins is supposed to be

included in the fin efficiency.

� The temperature distributions of the tubes in the thickness

directions can be neglected.


The continuity equation of moist air and the mass conserva-

tion of water vapour in the moist air are expressed as:

Vavravt

þ VavðrauaÞ

vzþ jm

�hAf þAb

� ¼ 0 (39)

VavðraxaÞ

vtþ Va

vðrauaxaÞvz

þ jm�hAf þAb

� ¼ 0 (40)

The energy equation of moist air is

VavðrahaÞ

vtþ Va

vðrauahaÞvz

þ �jmhvap þ qs out

��Afhþ Ab

� ¼ 0 (41)

The continuity equation and energy equation in the desic-

cant wall and tube are expressed as:

Vbvrbvt

� jm�hAf þAb

� ¼ 0 (42)

VbvðrbXbÞ

vt� jm

�hAf þAb

� ¼ 0 (43)

VbvðrbhbÞ

vtþ Val

vðralhalÞvt

� �jmhads þ qs out

��AfhþAb

�þ qs inAal in

¼ 0 ð44Þ

The continuity equation of water is expressed as:

Vrvrrvt

þ VrvðrrurÞvy

¼ 0 (45)

The energy equation of water is:

VrvðrrhrÞ

vtþ Vr

vðrrurhrÞvy

� qs inAal in ¼ 0 (46)

Heat and mass transfers occur between the air and desiccant

wall, and heat is transferred between the aluminium and

water, as shown in Fig. 13. AfþAb is the surface area of

the desiccant wall that is in direct contact with air. Val is

the volume of the entire aluminium part in the control

volume. The surface area, Aal_in, of the aluminium tubes is

that which is in direct contact with the water.

The heat flux, qs_out, between the air and desiccant wall is

qs out ¼ KhðTa � TbÞ (47)

where Ta and Tb represent the average temperatures of each

part in the control volume in Fig. 13.

The mass flux, jm, between the air and desiccant wall is




0 200 400 600 800 100020

30

40

50

60

70

Air

tem

pera

ture

To C

Humidification

Dehumidification

Time t s

Inlet Outlet

0 200 400 600 800 10000

10

20

30

40

Time t s

Humidification

Dehumidification

Air

hum

idity

rat

iox

g/kg

(DA

)

Inlet Outlet

Fig. 12 e Simulation and experimental results for validity of simulation model for the fixed desiccant bed.


The overall heat andmass transfer coefficients, Kh and Km, are

derived from Nut and Sht, which are identical to those of the

desiccant wheel.

The heat transfer coefficient of thewater flowing inside the

tube is assumed to be almost constant and is derived from

the conventional Hausen correlation (1943). For the average

value of heat transfer coefficient ain over the entire length of

the tube, Hausen (1943) presented the following empirical

relation for fully developed laminar flow in tubes:

Nur ¼ 3:66þ0:0668

�drlr

�RerPrr

1þ 0:04h�

drlr

�1=2RerPrr

i2=3 (49)

The height of the interior of the tube in Fig. 13 is 1.2 mm,

and the hydraulic diameter, dr, of the tube is also 1.2 mm due

to a regular square cross-section. The hydraulic diameter, dr,

of 1.2 mm was decided by conditions to suppose an Nur of

approximately 3.66, thermal conductivity of 0.66Wm�1 K�1

forwater and a heat transfer coefficient of 2000Wm�2 K�1. dr/lr

Fig. 13 e Desiccant bed with

is 0.0075 (¼0.0012/0.16) in Table 2. Water temperature Tr

shifts by heat transfer between the aluminiumwall and water.

The value of Nur approaches a constant value of 3.66 when

the tube is sufficiently long. Hence, the temperature profile

is fully developed when Nur approaches a constant value.

We consider that water flow inside the tube is laminar flow

because the hydraulic diameter of tube is very small. The

heat transfer rate of the water side is:

qs in ¼ ainðTb � TrÞ (50)

where Tb and Tr represent the average temperatures of

each part in the control volume in Fig. 13.


The conditions related to the regeneration air and process

air (flow rate, humidity ratio and temperature) at the inlet

of the desiccant wheel are given as boundary conditions.

When the desiccant bed is used in the adsorption process,

internal heat exchanger.




z¼ 0 (process air at inlet):

Ta ¼ Tpi (51)

xa ¼ xpi (52)

ua ¼ upi (53)

When the desiccant bed is used in the desorption process,

z¼ 1a (regeneration air at inlet):

Ta ¼ Tri (54)

xa ¼ xri (55)

ua ¼ uri (56)

When the tube is used in the adsorption process,

y¼ 0 (process air at inlet):

Tr ¼ Tcoi (57)

When the tube is used in the desorption process,

y¼ lr (regeneration air at inlet):

Tr ¼ Thei (58)

The initial conditions were:

t¼ 0:

Tb ¼ Tb ini (59)

Xb ¼ Xb ini (60)

6.4. Examples of simulation results obtained usingdynamic model for fixed desiccant bed with internal heatexchanger

Fig. 14 shows examples of simulation results over time

obtained using the dynamic model for the fixed desiccant

bed with the internal heat exchanger. The simulation condi-

tions show that a heated air temperature of 33 �C and

switching time of 69.5 s were optimized with a cooled air

temperature of 18 �C and humidity ratio difference of 4 g kg�1

(DA) between the inlet and outlet of the dehumidification side.

At that time, the cooling water temperature was 16.1 �C and

the heating water temperature was 34.7 �C. The configuration

of the fixed desiccant bed with an internal heat exchanger

0 20 40 60 80 100 120 14010

20

30

40

50

60

Air

tem

pera

ture

To C

Humidification

Dehumidification

Time t s

Air inlet and in Fig.4 Air Outlet and in Fig.4 Heating and cooling inlet

Fig. 14 e Examples of simulation results obtained using a dynam

exchanger.

is shown in Table 2, and other simulation conditions and

results are shown in Table 6.

For a reliability of simulation results, we believe that we

had a highly reliability simulation in this step. For the desic-

cant bed, we investigated the characteristics of a batch-type

desiccant bed based on a fixed only desiccant bed by experi-

ment and simulation, and for the heat transfer characteristics

of single phase flow inside the tube, we adopted ain, which

was calculated from Hausen’s empirical relation in equation

(49) and validated for fully developed laminar flow. Moreover,

for the thermal contact resistance between the corrugated

desiccant bed and aluminium fin, the effect of the thermal

contact resistance was supposed to be included in the fin

efficiency. Hence, we supposed that simulation results could

achieve a higher reliability by above explanation. We will

report the detailed characteristics of the fixed desiccant

bed with the internal heat exchanger experimentally in the

next step.

7. Experiment accuracy, and simulationconditions and procedure

7.1. Experiment accuracy

We investigated the moisture and heat balance under the

conditions of a wheel thickness of 20e400 mm, regeneration

temperature of 50e80 �C, rotational speed of 0e200 rph and

mean velocity of 1e4 m s�1.

Fig. 15(a) and (b) shows the accuracy of the experiment

for a desiccant wheel. The moisture balance and heat

balance were found to be in good agreement between the

adsorption process air and desorption regeneration air, within

approximately 15%.

7.2. Simulation conditions and procedure

For the size of each type of system shown in Table 2, the

dimensions of the desiccant wheel used in the wheel-type

conventional systemwere 4 250 mm� 200 mm. The desiccant

bed of the batch-type conventional system and the wheels of

the double-stage-type system with two desiccant wheels had

two separate desiccant beds and wheels, respectively. Hence,

each size was half that of the desiccant wheel of the wheel-

Air

hum

idity

rat

iox

g/kg

(DA

)

0 20 40 60 80 100 120 1400

10

20

30

40

Time t s

Dehumidification

Humidification

ic model for the fixed desiccant bed with the internal heat



Table 6e Simulation conditions and results optimized forfixed desiccant bed with internal heat exchanger.


Dehumidification

inlet


Process air humidity ratio g kg�1

(DA)

11.9

Process air relative humidity % 50.0

Process air mean velocity m s�1 2.80

Dehumidification

outlet


Cooling water inlet

temperature

�C 16.1

Regeneration

inlet

Heated air temperature �C 33.0

Heating water inlet

temperature

�C 34.7

Heated air relative humidity % 61.0

Heated air mean velocity m s�1 2.80

Switching time of each desiccant bed T 69.5


Specific heat of aluminium J kg�1 K�1 921

Water mean velocity m s�1 0.07


type conventional system. Table 2 shows the size of one bed

or wheel for each desiccant type.

For each system, the front section area and the void frac-

tion (Va/(VaþVb)) of the desiccant for the flow of the process

air or regeneration air were 0.01765 m2 and 0.65, respectively,

except for the batch-type system with the internal heat

exchanger. Because this system has corrugated fins made

of aluminium, the front section area was 0.0368 m2 and the

void fraction was 0.31 (Va/(VaþVbþVal)). Such sizes for the

systems were assumed so that the quantities of desiccant

and the mean air velocities in the air channels for all types of

systems were the same. The regeneration air flow rate was

the same as the process air flow rate for all of the system

types. From these points, it is clear that all six types of

systems had the same heat and mass transfer performances

between the air side and the desiccant wall.

The common simulation conditions for all of the desiccant

types are shown in Table 7. The ambient air conditions

considered in this study were set on the basis of typical

summer ambient air conditions in Tokyo. For indoor condi-

tions, the temperature and relative humidity were set to 28 �Cand 50%, respectively. The target of the process air

Fig. 15 e Validity of experiment accura

temperature supplied to the indoors was 18 �C, and the

sensible heat factor (SHF) was 0.5.

In the simulation, the effects of the cooled air temperature

and heated air temperature on each system performance

were investigated, because these two parameters have

a considerable influence on the system performance. Cooled

air temperature Tco implies the temperature of the process

air at the outlet of the cooler, and heated air temperature The

implies the temperature of the regeneration air at the outlet

of the heater; the points at which these temperatures were

attained are shown in Figs. 1(b), 1(d), 2(b), 3(b), 3(d) and 4(b).

While calculating the effect of Tco, the difference between

the humidity ratios of the process air at the inlet and outlet,

Dx, was kept constant, at a value of 4 g kg�1 (DA).

While calculating the effect of The, the cooled air temper-

ature was maintained at a constant of 18 �C. To show the

simulation results of the batch-type systems, time-averaged

data were shown.

We defined the switching time, St, as the time at which one

air channel remained on either the dehumidification side or

the regeneration side. St for the desiccant wheel was 3600/

(2�N ).

In the subsequent calculations, St was optimized to mini-

mize the heated air temperature while calculating the effect

of Tco, as shown in Fig. 16, and to maximize Dx while calcu-

lating the effect of The, as shown in Fig. 17.

8. Results and discussion

Next, we determined the effect of Tco and The on the perfor-

mances of the four types of dehumidification systems

considered in this study by simulation. At that time, heated

air temperature The or dehumidification rate Dx were opti-

mized to find the minimum heated air temperature The for

cooled air temperature Tco and maximum dehumidification

rate Dx for heated air temperature The under the condition of

optimized St. We adopted a trial and error method to search

for the target point for the simulation method.

8.1. Effect of cooled air temperature

Fig. 16 shows the effect of Tco on The, the heat transfer ratesQco

and Qhe in the cooler and heater, respectively, and the

cy by moisture and heat balance.



Table 7 e The common simulation conditions for alldesiccant types.


Dehumidification

inlet

Process air

temperature

�C 28.0

Process air

humidity

ratio

g kg�1 (DA) 11.9

Process air

relative

humidity

% 50.0

Process air

mean

velocity

m s�1 2.80

Dehumidification

outlet

Cooled air

temperature

�C 10.0e28.0

Regeneration

inlet

Heated air

temperature

�C 40.0e80.0

Heated air

relative

humidity

% 63.0

SHF (Sensible

heat

factor)

e 0.50

Heated air

mean

velocity

m s�1 2.80

Ambient air temperature �C 32.3


Specific heat of aluminium J kg�1 K�1 921

Heat capacity ratio between

aluminium and desiccant bed

(mbedCpbed)/(malCpal) 4.30

Fin efficiency of batch-type system

with internal heat exchanger

e 0.85

30

40

50

60

70

80

90

Cooled air temperature TcooC

Hea

ted

air

tem

pera

ture

The

o CH

eat t

rans

fer

rate

in h

eate

rQ

hekW

Opt

imum

sw

itchi

ng ti

me

S ts

Hea

t tra

nsfe

r ra

te in

coo

ler

Qco

kW

Dehumidification rate target:x=4g/kg(DA)

28oC

Heated air Temp.(Mark " ") : 32.3

0

0.5

1

1.5

2

0

0.5

1

1.5

2

10 15 20 25 300

200

400

600

8002st desiccant wheel(Two desiccant type)

1st desiccant wheel(Two desiccant type)

Conventional systems System with pre-coolerDouble stage systems Batch type system with internal heat exchanger

Fig. 16 e Effect of cooled air temperature on various

parameters.


optimum switching time, St, to achieve the target Dx. The

target Dx indicates the humidity difference ratio between

the process air inlet humidity ratio of 11.88 g kg�1 under the

conditions of 28 �C and a relative humidity of 50% and the

process air at the outlet humidity ratio of 7.88 g kg�1 under

the conditions of 18 �C and an SHF of 0.5, as shown in Table 7.

When cooled air temperature Tcos of the batch-type system

with an internal heat exchanger, double-stage system and

systemwith precooler became 17.4 �C, 13 �C and 12.5 �C at the

point marked by ‘�’, respectively, we found that each system

except the conventional desiccant system of Fig. 1 could be

driven as a desiccant system with an outdoor temperature

of only 32.3 �C directly, without heating.

For example, in Fig. 16, when cooled air temperature Tco

and the targetDx are 18 �C and 4 g kg�1respectively, the heated

air temperature The values of conventional desiccant dehu-

midification systems, the system with the precooler and the

double-stage-type systems are approximately 62.2 �C, 49.9 �Cand 43.2 �C, respectively. The is the lowest (33.0 �C) for the

batch-type system with the internal heat exchanger. There-

fore, heated air temperature The of the batch-type systemwith

the internal heat exchanger is approximately 29.2 �C lower

than that of the conventional desiccant-wheel-type systems.

For all types of systems, when cooled air temperature Tco

drops, the heat transfer rates of the heater and cooler decrease

and increase, respectively, because the decrease in the cooled

air temperature leads to a drop in the heated air temperature

along almost the same relative humidity line. Hence, the heat

transfer rate in the heater becomes smaller to desorbmoisture

in the desiccant due to the lower heated air temperature. In

the case of the heat transfer rate in the cooler, a lower cooled

air temperature requires a greater heat transfer rate in the

cooler.

The heat transfer rate of the cooler for the batch-type

system with the internal heat exchanger is higher than those

of the other types of systems; this high rate of the former

system is attributed to the thermal capacity of the aluminium

part. The trend of each parameter in the cases of the precooler

system and double-stage systems shifts at a certain cooled

temperature. Below the cooled temperature of the shift point,

the cooled temperature at the point reaches the dew point

temperature.

Theoptimumswitching times,St, of eachsystem,except for

the batch-type system with the internal heat exchanger,

increasewithadecrease in thecooledair temperature.The two



0

4

8

12

16

Heated air temperature TheoC

Hum

idity

rat

io d

iffe

renc

e x

g/kg

(DA

)H

eat t

rans

fer

rate

in h

eate

rQ

hekW

Opt

imum

sw

itchi

ng ti

me

S ts

Hea

t tra

nsfe

r ra

te in

coo

ler

Qco

kW

Cooled air temperature : 18

0

1

2

3

4

5

0

1

2

3

4

5

40 50 60 70 800

200

400

600

800

Conventional systems System with pre-coolerDouble stage systems Batch type system with internal heat exchanger

Fig. 17 e Effect of heated air temperature on various

parameters.


desiccant type has a different optimumswitching time in each

desiccantwheel.We suppose that thefirst desiccantwheel has

a high relative humidity and that moisture is fully adsorbed

even though the switching time is short. The St of the batch-

type system with the internal heat exchanger is faster than

those of other systems and remains almost constant because

of the thermal capacity of the aluminium part.

8.2. Effect of heated air temperature

Fig. 17 shows the effect of The on themaximumhumidity ratio

difference, heat transfer rates of the heater and cooler and

optimum switching time, St, between the adsorption and

desorption for each system with a cooled air temperature of

18 �C.

St is selected at an optimumpoint to achieve themaximum

dehumidification rate. The highest system performances, in

order of the humidity ratio differences between the process

air inlet and outlet for all system types, are the batch-type

system with the internal heat exchanger, double-stage-type

systems, system with the precooler, and conventional desic-

cant-wheel-type systems. Higher heated air temperatures

result in greater heat transfer rates in the heater, and thus the

relative humidity decreases at the point of heated air

temperature on the regeneration side. Hence, the humidity

ratio difference and heat transfer rate in the cooler increase

with a decrease in the relative humidity on the regeneration

side.

For the optimumswitching time, the increase in the heated

air temperature in each type of system, except the batch-type

system with an internal heat exchanger, shows that the

switching time becomes short enough to obtain themaximum

humidity ratio difference, whereas the batch-type system

with an internal heat exchanger almost has a constant

switching time due to the thermal capacity of the water flow

rate.

From these results, we conclude that the batch-type

system with the internal heat exchanger can be driven by

a heat source with the lowest heated air temperature and

can remove the maximum humidity. The use of this system

can result in large energy savings and aid in maintaining

a comfortable indoor environment.

As for the fin efficiency by mass transfer on the fin surface

of the internal heat exchanger, it has not been reported. Since

this type has fins covered with desiccant and the heat and

mass transfer occurs simultaneously on the fin surface, in

this paper, we conclude that mass transfer by temperature

variation on the fin is supposed as similar activity as the fin

efficiency of heat transfer. Therefore, the fin efficiency of

desiccant for mass transfer is also provided as fin efficiency

0.85 considering the common values of fin efficiency for heat

transfers.

9. Conclusions

In this study, we used previously validated simulation

models to analyze the performances of four dehumidification

systemsdconventional desiccant-type systems (wheel type

and batch type with only desiccant), a system with

a precooler, double-stage-type systems (a system with two

desiccant wheels and a four-partition desiccant wheel type),

and batch-type systemwith an internal heat exchanger. Using

these simulation models, we could successfully evaluate the

performance of all types of systems and especially investi-

gated the minimum heated air temperature along with the

cooled air temperature in Fig. 16 and the maximum humidity

ratio difference along with the heated air temperature in

Fig. 17 by the trial and error optimization method. The simu-

lation results showed that out of the four systems, the batch-

type systemwith the internal heat exchanger can be driven by

a heat source with the lowest heated air temper-

aturedapproximately 33 �C at a cooled air temperature of

18 �C; further, the heated air temperature can be reduced

by decreasing the cooled air temperature. As for the fin




efficiency by mass transfer on the fin surface of the internal

heat exchanger, this type has fins covered with desiccant and

the heat and mass transfer occurs simultaneously on the fin

surface. Furthermore, we will conduct experimental investi-

gation research of the effect ofmass transfer on the fin surface

of desiccant as fin efficiency of mass transfer alone.

r e f e r e n c e s

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Barlow, R.S., 1982. Analysis of the Adsorption Process and ofDesiccant Cooling Systems: a Pseudo-steady-state Model forCoupled Heat and Mass Transfer. Solar Energy ResearchInstitute, Golden, Colorado, USA.

Dhar, P.L., Singh, S.K., 2001. Studies on solid desiccantbasedhybridair conditioning systems. Appl. Therm. Eng. 21 (2), 119e134.

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performance analysis of desiccant dehumidification systems driven by low-grade heat source

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