performance analysis of a new developed a thesis in

PERFORMANCE ANALYSIS OF A NEW DEVELOPED

DESICCANT IN A PACKED TOWER

by

SEDEF KAVASOGULLARI, B.S.

A THESIS

IN

MECHANICAL ENGINEERING

Submitted to the Graduate Faculty of Texas Tech University in

Partial Fulfillment of the Requirements for

the Degree of

MASTER OF SCIENCE

IN

MECHANICAL ENGINEERING

Approved

Accepted

May, 1991

ACKNO\"'LEDGEMEJ\'TS

The author would like to thank Dr. Atila Ertas for his direction, counseling, valu

able advice, and support throughout this research, and to Dr. P. G andhldasan for

his help, advice, and direction. The author would like to express her appreciations

to Kham.is R. Al-Balushl for hls help during this study and to ~orton Company for

their continuous support. Special thanks are also extended to Dr. Herbert Carper

and Dr. Edward E. Anderson for serving as members of this thesis committee.

The author expresses her sincere gratitude to her Mother, Beyhan Kavasogullari,

for her support, and love. Special thanks for Michael J. Benson, for his support,

help and encouragement.

Financial support for this study has been provided by the state of Texas under

Energy Research in Application Program, Project # 309.

11

TABLE OF CONTENTS

ACKNOWLEDGEMENTS ................................................... u

ABSTRACT ............................................................. vii

LIST OF TABLES ...................................................... viii

LIST OF FIGURES ...................................................... ix

NOMENCLATURE ...................................................... Xl

CHAPTER

1. BACKGROUND AND LITERATURE SURVEY .................... 1

1.1 Introduction .................................................... 1

1.2 Desiccants ...................................................... 2

1.2.1 Solid Desiccants . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2

1.2.2 Liquid Desiccants . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 4

1.3 Gas Liquid Contact Devices ..................................... 6

1.3.1 Regular Packings ......................................... 10

1.3.2 Random Packings ......................................... 10

1.3.3 Selection of Packing Material . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11

1.3.3.1 Raschig Rings ....................................... 12

1.3.3.2 Snowflake Packings .................................. 12

1.4 Proposed Dehumidification System ............................. 13

2. COMPARISON OF HEAT AND MASS TRANSFER COEFFICIENTS OF LIQUID DESICCANT MIXTURES IN A PACKED COLUMN ............................ 19

2.1 Introduction ................................................... 19

2.2 Evaluation of Heat and Mass Transfer Coefficients . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 21

lll

2.3 Results and Discussion ......................................... 23

2.3.1 0.5 in. Ceramic Raschig Ring .............................. 24




3. EVALUATION OF THE PERFORMANCE OF THE DEHUMIDIFICATION PROCESS IN A PACKED COLUMN ........................................... 36

3.1 Introduction .................................................... 36

3.2 Evaluation of Heat and Mass Transfer Coefficients ........................................... 37

3.2.1 Holdup Calculations ...................................... 37

3.2.2 Interfacial Surface Area Calculations ...................... 37

3.2.3 Volumetric Heat and Mass Transfer Coefficients ........... 39

3.3 Simultaneous Analysis of Heat and Mass Transfer in the Packed Column ................................. 40

3.4 Computational Procedure ...................................... 44


4. APPLICATIONS OF DEHUMIDIFICATION SYSTEMS FOR AIR CONDITIONING UNITS ................................. 62

4.1 Introduction ................................................... 62

4.2 Air-Conditioning in Supermarkets . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 63

4.3 Concept of the Proposed Cooling System ....................... 64

4.4 Description of the Model ....................................... 65

4.5 System Analysis for the Supermarket ........................... 66

4.5.1 Supermarket .............................................. 66

4.5.2 Mixing of Air . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 66

IV

4.5.3 Dehumidification Tower ................................... 67

4.5.4 Vapor-Compression Unit .................................. 68

4.5.5 Solar Regenerator . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 69

4.5.6 Heat Exchangers .......................................... 69

4.5. 7 Auxiliary Heater .......................................... 70


4. 7 Economic Analysis . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 71

4.7.1 Targeted Energy Savings .................................. 71

4.7.2 Savings ................................................... 71

4.7.2.1 Pumps .............................................. 72

4. 7 .2.2 Fans . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 72

4.7.2.3 Vapor-Compression System .......................... 73

4.7.2.3 The Conventional AirConditioning Unit .............. 73

4.7.2.4 Comparison of Energy Requirements ................. 73

5. CONCLUSIONS AND RECOMMENDATIONS .................... 79

REFERENCES .......................................................... 83

APPENDICES

A. COMFORT ZONE ................................................ 87

B. CALCULATIONS ................................................. 88

Mixing of Air . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 88

Vapor Compression Unit ....................................... 89

Solar Regenerator .............................................. 91

Heat Exchangers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 92

Auxiliary Heater . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 94

v

Pump Energy Calculations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 95

Fan Energy Calculations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 97

Vapor Compression Unit Energy Calculations .................. 98

Energy Calculations for Conventional System . . . . . . . . . . . . . . . . . . . . . . . 99

Saving Calculation ............................................... 100

C. REGENERATOR COVERED "WITH GLASS BUT BOTH ENDS ARE OPEN .................................. 103

D. MATERIAL COST .............................................. 104

E. CONVENTIONAL SYSTEM PROCESSES ON THE PSYCHROMETRJC CHART ........................... 106

VI

ABSTRACT

Dehumidification of air using desiccants is a developing topic in the last decade

as result of energy shortage considerations. In this study, a theoretical method is

employed to analyze a liquid desiccant packed tower system. A new desiccant (Cost

Effective Liquid Desiccant, 'CELD') is introduced which is a 50% LiCl and 50%

CaCl2 mixture. The heat and mass transfer coefficients of LiCl, CaCl2 , and CELD

solutions are compared. The packed column performance of the CELD is calculated

using a new snowflake packing material. After showing that the CELD is a stable

liquid desiccant, a numerical model of a dehumidification system is combined with

a vapor compression unit to form a hybrid cooling unit to study the energy savings

for a supermarket.

In this study, it is shown that the CELD is promising to be a cost effective sta

ble liquid desiccant. The snowflake packing is an efficient packing material. Conse

quently, liquid desiccant hybrid cooling systems is a new alternative for conventional

air conditioning units with economical benefits.

Vll

LIST OF TABLES

2.1 Constants for calculating static and total holdup .......................... 30

2.2 Constants for calculating interfacial surface area. . . . . . . . . . . . . . . . . . . . . . . . . . 30

4.1 Additional. cost of ~he proposed hybrid cooling system on conventional uru t. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7 4

Vlll

LIST OF FIGURES

1.1 Packed Tower ........................................................... 15

1.2 Some random packings . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 16

1.3 Plastic lntalox® snowflake packing ....................................... 17

1.4 The proposed system . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 18

2.1 Schematic of the closed cycle drying system .............................. 31

2.2 Effect of liquid flow rate on volumetric liquid phase mass transfer coefficient for 0.5 in. ceramic raschig ring . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 32

2.3 Effect of desiccant concentration on volumetric liquid phase mass transfer coefficient for 0.5 in. ceramic raschig ring . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 33

2.4 Effect of liquid inlet temperature on volumetric liquid phase mass transfer coefficient for 0.5 in. ceramic raschig ring . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 34

2.5 Effect of liquid fl.ow rate on volumetric liquid phase heat transfer coefficient for 0.5 in. ceramic raschig ring . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 35

3.1 Total holdup versus liquid inlet flow rate for lntalox® snowflake packing .. 49

3.2 Differential section of a packed tower .................................... 50

3.3 Dehumidification performance of the packed column with lntalox® snowflake packing .............................................. 51

3.4 Change of air outlet temperature with the packing height for the lntalox® snowflake packing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 52

3.5 Change in the exit humidity of air with the liquid flow rate for Intalox® snowflake packing . . .. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 53

3.6 Change in the liquid and air temperature with the desiccant flow rate for Intalox® snowflake packing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 54

3.7 Change in the air outlet humidity with the increase in the desiccant inlet temperature for lntalox® snowflake packing ........................ 55

3.8 Change in the desiccant and air outlet temperature with the increase in the desiccant inlet temperature for Intalox® snow:A.ake packing .............. 56

lX

3.9 Change in the air outlet humidity with the change in desiccant inlet temperature for 40% and 45% concentrations, with 3 and 4 GPM flow rate of desiccant for Intalox® snowflake packing . . . . . . . . . . . . . . . . . . . . 57

3.10 Change in the air humidity with the increase in the air inlet temperature for Intalox® snowflake packing ............................. 58

3.11 Change in the desiccant and air temperatures with increase in the air inlet temperature for Intalox® snowflake packing with 100% saturated air ..... 59

3.12 Change in the outlet humidity with the change in the inlet humidity of air for Intalox® snowflake packing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 60

3.13 Change in the outlet temperatures of desiccant and air with the change in air inlet humidity for Intalox® snowflake packing ..................... 61

4.1 Schematic of the proposed cooling system . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 75

4.2 Processes on the psychometric chart . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 76

4.3 Air-desiccant and desiccant-desiccant heat exchangers . . . . . . . . . . . . . . . . . . . . 77

4.4 Desiccant-water heat exchanger . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 78

X

a

c

ds

E

Energy

Fc,FL

g

G

G'

h' a G

H,

NOMENCLATURE

specific interfacial surface area for absorption, desorption, (m2 /m3 )

specific interfacial surface area for vaporization, (m2 /m3)

molar density, (kmoljm3 )

heat capacity of dry air, (J/kmol K)

heat capacity of water, (J/kmol K)

heat capacity of liquid, ( J /kmol K)

specific heat, (kJ /kg°C)

diffusivity of liquid (m2 /s)

diameter of the sphere that has the same surface, (m)

effectiveness, dimensionless

energy, (kW)

mass transfer coefficient for air and liquid respectively, (kmol/m2 s)

gravitational acceleration, (m/s2)

superficial molar gas mass velocity, (kmol/m2s)

superficial gas mass velocity (kgdryair/m2s)

molar velocity of dry air, (kmol/m2s)

enthalpy, (kJ /kg)

saturation enthalpy for gas phase, (kJ /kg)

latent heat of evaporation, (kJ /kg°C)

heat transfer coefficient for air and liquid respectively, (W /m2K)

volumetric heat transfer coefficient for the gas phase corrected for

the mass transfer, [Ackermann Correction], (W /m3 K)

integral heat of solution per mole of solution, (kJ /kmol)

Xl

HL

HG

JD

JH

K'

KG

kG,kL

kth

L'

L

m

p

p·

Pt

PBM

PrG,PrL

Power

liquid phase molar enthalpy, (kJ /kmol)

gas phase molar enthalpy, (kJ /kmol)

mass transfer group, dimensionless

heat transfer group, dimensionless

overall mass transfer coefficient on mass basis, (kg/m2 s)

overall mass transfer coefficient on molar basis, (kmol/m2 atm s)

mass transfer coefficient for gas and liquid

respectively (kg/ m 2 s)

thermal conductivity of liquid (W /m2K)

superficial liquid mass velocity (kg/m2s)

superficial liquid molar velocity (kmol/m2 s)

mass flow rate, kg/s

molecular weight of water vapor, kg/kmol

molar flux of water relative to a fixed surface, (kmol/m2 s)

pressure, (Pa)

vapor pressure, (Pa)

total pressure, (Pa)

mean partial pressure, (Pa)

Prandtl number for air and liquid respectively

power, (kWh)

heat flow rate, (k W)

heat flow in the evaporator, (kJ /kg)

heat flux for the gas phase, (kJ /m2 kg)

Schmidt number for air and liquid respectively

air temperature (°C)

Xll

T.-1

w

X

Yc

Yci

Yc

z

p

Subscripts

A

cd

com

D

liquid desiccant temperature (°C)

interfacial temperature, (°C)

ambient temperature, (°C)

humidity ratio, (kg/kg)

desiccant flow rate, (kg/hr)

concentration of water in liquid phase, (kmol H20 /kmol liquid mixture)

value of Xc at the interface, (kmol H20/kmol liquid mixture)

concentration of water in the gas mixture, (kmol H2 0/kmol gas mixture)

value of Yc at the interface, (kmol H20/kmol gas mixture)

concentration of water in gas phase, (kmol H2 0/kmol dry air)

height of the packing from the bottom, (m)

liquid desiccant concentration (%)

fractional void volume in a dry packed tower, dimensionless

operating void space in the packing, dimensionless

viscosity of air and liquid, respectively, (kg/m s)

molar latent heat of vaporization of water at the

reference temperature, (kJ /kmol)

liquid operating, static, and total holdup respectively, (m3 /m3)

density, (kg/m3)

efficiency

ambient air

condenser

compressor

desiccant

Xlll

E vapor compression unit

£1" evaporative cooler

F fan

M air mixture

n new system

0 old system

p pump

PD desiccant pump

PW water pump

R room

ref refrigerant

WB wet bulb

XlV

CHAPTER 1

BACKGROUND AND LITERATURE SURVEY

1.1 Introduction

Increasing energy costs in the last decade have resulted in an increased attention

to dehumidification systems. Dehumidified air has potential uses in many fields, e.g.,

air-conditioning, crop-drying, meat and fish drying, etc.

Dehumidification of air can be achieved in many different ways, such as refrig

eration, mechanical compression, and using desiccants. The refrigeration method

is widely used in air-conditioning units. In conventional air conditioning units, air

is cooled below the saturation temperature, which is below the comfort conditions,

to condense the excess moisture. For comfort conditions, air is reheated over the

condenser. This is a very expensive way of dehumidifying the air. Mechanical com

pression is another way of dehumidification. When air is compressed mechanically,

the moisture in the air can be condensed due to high vapor pressure even at high

temperatures. However, this is only suitable for dehumidifying small amounts of

air and the need for a large amount of cooling water makes the system impractical

and costly. The third method of uses desiccants. This was studied in detail in this

project. Desiccants are materials that have the ability of extracting moisture from

different media.

The goal of this thesis was to conduct a theoretical study of the performance of

a liquid desiccant packed tower system using a newly developed desiccant called

1

2

'Cost Effective Liquid Desiccant' (CELD). CELD is a mixture of 50% lithium chlo

ride and 50% calcium chloride. The heat and mass transfer coefficients for CELD

will be calculated, and the results will be compared with those of lithium chloride,

and calcium chloride. Then the performance of CELD in the packed tower will

be evaluated, and the results will be discussed. Two hybrid cooling systems that

use CELD as the desiccant for dehumidification were modelled, and an economic

analysis of the model was undertaken.

In this chapter, a background for desiccants, towers and packing materials will

be presented and the proposed dehumidification system will be discussed.

1.2 Desiccants

Desiccants are the materials that have a high affinity for water vapor. As a

result of absorbing moisture, some desiccants change their chemical structure or

become diluted. To be able to reuse the desiccant, regeneration is necessary. The

regeneration of the desiccant can be accomplished by using solar energy or fossil fuel.

Regeneration temperatures can be as low as 50°C or as high as 250°C depending on

the type of the desiccant. Desiccants have many areas of applications. Desiccants

can be broadly classified as solid and liquid desiccants.

1.2.1 Solid Desiccant

Lithium chloride, silica gel, activated alumina, and activated carbon are some

examples of solid desiccants. The solid desiccant that has a porous structure picks

up the moisture from the air and forms a hybrid of the desiccant. Regeneration

is accomplished either by reheating the solid desiccant or by passing the heated

air through the desiccant bed to absorb the moisture in the desiccant. A model

that uses a solid desiccant in the dehumidification chamber of an air-conditioning

3

unit (hybrid cooling system) in a supermarket has been designed and operated by

Manley et al. (1985). In this system, the solid desiccant is gel tapes that are covered

with solid lithium chloride which is shaped like a rotary wheel that is casted in a

net. This shape allows the air to pass through the desiccant. One third of the bed

is isolated from the other part, and this portion is used for integrated regeneration

of the desiccant. In the regeneration part, air that is heated by a natural gas heater

passes through the desiccant wheel to pick up the moisture.

In another study, the recovered heat from the condenser is used for the regen

eration of the desiccant (Macdonald, 1983). Whitehead (1985) suggested a similar

system for a hot and humid climate. In this system, outside air is dehumidified

below comfort conditions and then mixed with the returning air from the room.

This system decreases the amount of air that needs to pass through the desiccant

wheel and requires less power for operation. Hybrid cooling systems that use solid

desiccants promise up to a 60% savings over conventional vapor compression units

(Macdonald, 1983). In the hybrid cooling system, the only power required is to

blow the air through the desiccant and sensibly cool the air a few degrees. In many

of these studies, the desiccants are highly stable, but the stability is very much de

pendent on the contaminants in the air used (Burns et al., 1985). For this reason,

air passing through the desiccant is filtered. This creates additional pressure drops,

which mean higher power blowers. Solid desiccant hybrid cooling systems have been

analyzed theoretically in many studies. Systems are shown to have "coefficient of

performance" (COP) as low as 0.8 and as high as 2.48(Kettleborough et al., 1986).

But the number of experimental studies are limited, and there is not enough com

mercialization. These are a few of the reasons why hybrid cooling systems are not

widely used.

4

1.2.2 Liquid Desiccants

Moisture can be removed from the air by using a solution of lithium chloride,

calcium chloride or bromide. "Then the air and liquid desiccants are brought into

contact, depending on the concentration of the liquid desiccant, moisture tends

to move from the gas phase to the liquid phase. The driving force in this phase

change is the difference between the vapor pressure of the liquid desiccant and vapor

pressure of the air. So in such a contact system, maintaining a proper desiccant

concentration will result in dehumidification regardless of the moisture in the air.

Many chemicals can be used as liquid desiccants, e.g., lithium chloride, calcium

chloride, monoethylene, sodium or potassium hydroxide, zinc chloride or bromide,

sulfuric acid, phosphoric acid, etc. There are some characteristics which need to be

considered for the choice of liquid desiccants. First, liquid desiccants should have

a low vapor pressure for effective drying. It should be regenerated easily at low

temperatures. There should be a large solubility range on the temperature scale to

prevent precipitation and it should be nontoxic. Most importantly, it should also

be inexpensive.

Some advantages of liquid desiccants over their solid counterparts are their abil

ity to be regenerated separately from the dehumidification process and liquid des

iccants can be cooled or heated through heat exchangers. Moreover, the liquid

desiccant will clean up the air from contaminants and disinfect it. This has not

only the advantage of the lower blower power but also the advantage of clean air

(Kettleborough et al., 1986).

In the present study, a liquid desiccant system was chosen because of the fore

going advantages and lack of research in this area. The liquid desiccants have many

potential areas of application. Drying crops is one of the primary examples. Air

during the dehumidification process is warmed up to a few degrees above the am-

5

bient temperature, as a result of released latent heat of vaporization. This warm

and dry air can be used for batch-drying operations for the post-harvest handling

of crops. The process has considerable economical advantages over conventional

natural gas systems in which the air is only sensibly heated.

Liquid desiccants can also be used as heat pumps ( G andhidasan, 1989). Heat

can be stored and recovered by diluting and reconcentrating the liquid desiccant.

Power generation is another potential area for the use of liquid desiccants. Solar

energy and liquid desiccants can be combined to produce high pressure steam to

keep a turbine running.

Another field of application is in the food-processing industry ( Gandhidasan,

1989). Foods spoil mainly because of their high water content. Liquid desiccants

can be used for drying red meats or fish, concentration of citrus juices, and the

cooling and freezing of solid foods. Biswal et al. (1982) developed a mathematical

model of a packed tower for a solar-driven food-processing application.

The regeneration of liquid desiccants can be achieved by either solar panels or

by using natural gas and a regeneration tower. Gandhidasan (1983) suggested a

parallel-flow solar regenerator for the reconcentration of the liquid desiccant in his

theoretical study. It is found that preheating the liquid will increase the performance

of the regenerator. The other controlling parameters in the system are insolation

of solar panel, ambient temperature, water vapor pressure in the ambient air, and

the ratio of liquid flow rate to air flow rate. Peng et al. (1984) studied the different

types of solar regenerators and concluded that in hot and humid climates open

regenerator performance is very low. Leboeuf et al. (1980) used a packed tower to

reconcentrate a lithium chloride solution for absorption cooling using solar heated

air. Gandhidasan (1990) derived a simple expression to calculate the mass of water

evaporated from the weak desiccant in a packed tower using hot air.

6

In the present study, a new desiccant is proposed, which is called CELD. It is a

mixture of 50% lithium chloride and 50% calcium chloride. Calcium chloride is the

cheapest desiccant (around 15 cents per pound) and easily available. It promises

dehumidification up to 25% for concentrations around 40%. However it is very

unstable with the inlet conditions of the air and desiccant. Lithium chloride is

a very strong and stable liquid desiccant, but it is very costly (around $8.00 per

pound). It promises dehumidification up to 60% for concentrations of 40%. The

mixing of these two substances produces a optimal desiccant (i.e., effectiveness and

cost saving of around 50% is achieved compared to lithium chloride). The 50%

combination of the two desiccants was chosen according to the vapor pressure of

the mixture. Detailed information about CELD solutions can be found in Ertas et

al. (1990a). CELD is used for the first time in this project.

1.3 Gas Liquid Contact Devices

The contact of liquid and gas can be achieved in parallel or counter flow towers

such as the wetted-wall towers, spray towers or chambers, fin tube surfaces in a

column, and packed towers.

Wetted-wall towers are used for controlled mass transfer operations. The liquid

runs down inside a vertical pipe where the gas may flow in either direction. Al

though the gas phase pressure drop is very low, the interfacial surface area is just

equal to the inner surface of the pipe. This surface area is insufficient for effective

dehumidification.

Liquid is sprayed into the gas stream by a nozzle in the spray to·wers. The

flow can be a countercurrent in a vertical column, or cocurrent in a horizontal

column. Even though the gas phase pressure drop is small, it requires a large

amount of pumping power to disperse the liquid into fine droplets. In addition,

7

liquid tends to entrain with the gas in large amounts. Mist eliminators are needed

for most of these operations. Grifford (1957) suggested a dehumidification system

for an air-conditioning unit. In this system, two spray towers are used-one for

dehumidification and one for regeneration. The cooling water circulated in the

dehumidifying tower keeps the operation temperature around 80°F. Similarly, hot

water inside the regeneration tower keeps the temperature of regeneration between

200°F and 270°F. Even though the economic savings are not a considerable amount,

the author suggests the system because of the sterile atmosphere it provides. Gupta

et al. (1978) proposed a 3-ton air-conditioning unit that uses a spray tower, solar

regenerator, and direct type evaporative cooler. The system was designed for hot

and humid climates. The only drawback of the system was the corrosive liquid

desiccant. Corrosion is especially important in the solar panels. The system COP

was calculated to be around 0.2. Queiroz et al. (1984) have used the Markel

integral approach to calculate the mass transfer coefficient for the dehumidification

operation in a spray tower.

Peng et al. (1981) studied a fin tower. The individual heat and mass transfer

equations were derived. The interfacial surface area is assumed to be equal to the fin

surface as long as the liquid flow occupies less than 20% of the volume. However, in

the experimental studies there were many problems because of the flooding which

is coupled with choking. The reason for this is the structure of the fins and the

liquid desiccant ( triethylene glycol, TEG). They modelled a dehumidifying tower

suitable for a 3-ton air conditioning load. The theoretical and experimental results

were in close agreement. However, the theoretical model needs to be reevaluated

for different initial conditions. The liquid is sprayed over the tubes of the fin tube

surfaces in a column where the air is blown upward through the column. The cooling

water that circulates inside the fin tube keeps the operation isothermal. However, in

8

some cases, extremely large air flow rates are required, and it is difficult to control

the liquid film over the fin tubes.

Packed towers are useful for cases where continous contact between the gas and

liquid phases are required. Towers can be filled \vith packings or devices that will

provide the required interfacial surface area, as shown in Fig. 1.1. The liquid is

distributed over the packing and covers it as a thin liquid film. For the dehumidifi

cation processes, the recommended design is the countercurrent flow column. The

desiccant falls down through the packing because of gravity while the gas moves

upwards due to the pressure difference created by the blower. There is a consid

erable pressure drop in the gas phase, whereas for the liquid phase the only power

requirement is for pumping the liquid up to the top of the tower. Sherwood et al.

(1939) correlated the heat and mass transfer coefficients of the packed tower us

ing 0.5, 1.0, and 1.5 in. ceramic raschig rings with water and air. The experiments

were conducted for different absorption and desorption cases and different heights of

packing. McAdams et al. (1949) studied a packed column with 1 in. carbon raschig

rings, theoretically and experimentally. They calculated the heat and mass transfer

coefficients in both modes and concluded that overall transfer coefficients should

not be used in the evaluation of tower performance. Treybal (1969) calculated the

mass and heat transfer performance of a packed column, using individual phase

coefficients. The heat and mass transfer is calculated through the mass and energy

balance through the tower. The tower height is calculated for the desired level of

performance. Factor et al. ( 1980) calculated the performance of a packed tower the

oretically and experimentally with ceramic lntalox® saddle packing. The difference

between the methods of Treybal (1969) and Factor et al.'s study is the interfacial

surface temperature. Treybal calculated the interfacial surface temperature using

an iterative method. Factor et al. assumed the interfacial surface temperature was

9

equal to the water bulk temperature. Monoethylene glycol (MEG) and lithium bro

mide (LiBr) were used in the experimental studies. MEG gave very unstable results

because of its structure. LiBr gave not only stable results but it was also consistent

with theoretical calculations. The only problem they had was the large pressure

drop across the packed column. They suggested the use of different packing mate

rials. Lof et al. (1984) measured the heat and mass transfer coefficients of a packed

column with 1 in. ceramic raschig rings and compared their experimental results

with the ~vlcAdams' equation {1949). They give different reasons for the inconsis

tency of the results. However, McAdams' equations were correlated using carbon

raschig rings. Gandhidasan et al. (1987) used a method similar to Treybal (1969)

for the calculation of performance of the packed tower. In this study, they assumed

that the resistance to mass transfer exists in both phases and that the interfacial

surface temperature was equal to the liquid bulk temperature. Ullah (1986) used

the same method to evaluate the dehumidification effectiveness of a packed column

with calcium chloride. This method was developed for the performance analysis

of an air-conditioning unit which used a dehumidifying tower. The effectiveness

method reduces the high computational cost. Other components of the system are

an evaporative cooler, a solar roof pool regenerator, and a cooling water system. An

absorption cooling system column mechanism was analyzed by Collier {1979). In

this system instead of heating the regeneration environment, liquid desiccant was

heated in the open solar regenerators. The system performance was calculated ac

cording to the performance of the solar regenerators. The amount of water that can

be removed in the solar panel "\vill correspond to the amount of latent load that will

be absorbed in the absorption tower. The open solar regenerators brought many

problems to the system. Gandhidasan, {1990), derived a simple expression which

determines the performance of a solar-desiccant air-conditioning unit with the given

10

initial conditions. The unit consists of an evaporative cooler, solar regenerator, de

humidifying tower, and cooling water system. This approximation provides a broad

idea about the performance of the system under given conditions.

The packed columns are chosen to work with in this study because they provide

large interfacial surface areas and ease of operation. Due to the insulation on the

walls oft he column, the heat and mass transfer processes can be treated as adiabatic.

The packing material for columns are classified into two major groups according to

their structure.

1.3.1 Regular Packings

This type entails a large group of packing. Counter flow trays, stacked raschig

rings, structured packings, and expanded-metal-lath packings are some examples of

regular packing. The material can be metal, plastic, ceramic or the combination of

these types. They usually provide a lower pressure drop than random packings and

controlled interfacial surface areas. The controlled surface area provides controlled

mass transfer in the volume. Their structure changes according to design purposes.

They provide an efficient media for countercurrent liquid gas flow. However, their

prices are very high.

1.3.2 Random Packings

Random packings are simply materials that are dumped inside the tower irreg

ularly. In the past, the packing materials used were broken stones, gravel, or lumps

of coke. Even though they have lower cost, they cause larger pressure drops in

the gas phase as a result of their structure. Today various packings with differ

ent structure and materials are available. Some examples of widely used packings

are: raschig ring, berl saddle, super Intalox® saddle, maspac packing, Tellerette®

11

packing, pall ring, hy-pak packing, and IMTP® packing. Some of the packings are

shown in Fig. 1.2 (Strigle, 1987). Most of these packings are available in metal,

plastic, and ceramic form and in many sizes, e.g., raschig rings are available from 6

to 100 millimeters (mm) in diameter (Treybal, 1980). The packing materials that

have been designed in the last five years are primarily available in plastic. The

reason for this can be summarized as: new plastics are resistant to heat and heavy

loads and inert to most chemicals; they have light weight and low cost; and they

can be cast giving larger void space and interfacial surface areas.

1.3.3 Selection of Packing Material

Packings are chosen according to following properties:

1. Packing material should provide a large interfacial surface area for

the contact of liquid and gas.

2. Packings should have sufficient void space for the liquid and gas to

move along the tower.

3. Packings material should be inert to liquid which is being processed.

4. Packings should resist the structural loads and also should be easy to

handle.

5. Packings should have a low price.

The packing material choice in this study was made according to the above

mentioned factors and the availability of data. Ceramic raschig rings and Intalox®

high-performance snowflake packings are studied.

12

1.3.3.1 Raschig Rings

These are thin walled hollow cylinders. Their sizes are determined by the diam

eter of the cylinder. Even though they are available in many different sizes from 6

mm to 100 mm, in the present study only 13 mm (0.5 inches, in.) 25 mm (1 in.),

38 mm (1.5 in.), and 50 mm (2 in.) will be used. Ceramic raschig rings of these

sizes were studied previously by Shulman et al. (1955) and the available data was

reorganized in Treybal (1980). The availability of the data was one of the reasons

these packing materials were used in this study. The comparison of heat and mass

transfer coefficients for lithium chloride, calcium chloride, and CELD was done us

ing these packing materials. As raschig rings are made of ceramics, in addition to

their high weight, they are liable to be broken during the dumping into the tower.

1.3.3.2 Snowflake Packings

This is a new product of the Norton Company. Its nominal diameter is 94 mm,

and height is 30.5 mm. Snowflake packing provides approximately 95% \'oid space

and a surface area of 4.54 m 2 /m3• Packing elements are shaped like a snowflake

as shown Fig. 1.3 (Norton, 1987). Larger void space allows more gas and liquid

to move along the tower with smaller pressure drops and also results in smaller

holdups. This is an advantage for dehumidification systems, because the stagnant

liquid inside the tower gets saturated in a short time and this causes a decrease in

the mass transfer performance of the system. In addition to the above mentioned

properties, polypropylene snowflake packings are easy to handle and are resistant

to corrosion and high temperatures and have a low price.

13

1.4 The Proposed Dehumidification System

The proposed system is formed by a dehumidifying tower, a droplet separator,

two heat exchangers, a water cooling tower, an auxiliary heater, two desiccant

tanks, and a desiccant regeneration system as shown in Fig. 1.4. The figure will

be completed in the following chapters by adding the pumps and fans that are

necessary for circulating the liquid and air, respectively.

The dehumidifying tower is a packed column that is filled with plastic snowflake

packing material. The tower is a 24 in. diameter thin walled PVC pipe. The height

of the tower will be decided according to the performance of the CELD solution.

Two packing holders restrict the packing from the upper and lO\ver ends and also

control the gas and liquid flow distribution. The liquid distributor distributes the

liquid on the packing material evenly. The height of the tower is suggested to be at

least 2 feet longer than the packing height. If desired an additional few feet below

the packing can be used as a desiccant tank.

The droplet separators are needed because the liquid tends to be entrained the

leaving gas. The heat exchangers, cooling water and the cooling tower, and the

auxiliary heater are necessary for heating the weak desiccant corning from the de

humidifying tower and cooling the desiccant returning from the regeneration system.

The regeneration system is needed to concentrate the diluted desiccant back to

the desired concentrations. Regeneration can be achieved by using solar energy

or fossil fuel. Solar panels that are covered with glass, with the upper and lower

ends open to the atmosphere, can be used in the regeneration of the desiccant. A

detailed study of this type of solar panel is given by Mullick et al. (1974). As the

liquid falls down from the upper end of the panel, it is heated by the solar radiation.

The heating process increases the vapor pressure of the desiccant. The air 'vhich

has a lower vapor pressure than liquid desiccant, picks up moisture in the panel. :\

14

liquid temperature around 60°C is sufficient for the performance of the solar panel.

A natural gas auxiliary heater can be used during the time periods when the solar

radiation is not available.

Another way of regeneration is to use packed columns. The mass transfer at the

of liquid and gas contact is a result of the vapor pressure difference of the two phases.

If the vapor pressure of the liquid desiccant increases above the vapor pressure of

the ambient air, water will be transferred from the liquid phase to the gas phase.

This can be achieved in two methods. The first method is to heat the air. When

air is heated, its relative humidity is decreased. During the contact in the column

air transfers heat to the liquid phase. This method has been studied previously by

Leboeuf et al. (1978). A second mode of dehumidification is suggested where the

liquid desiccant is heated. By this way, the vapor pressure of the desiccant solution

rises above the ambient air vapor pressure.

After studying the performance of the new liquid desiccant CELD, the above

dehumidification system will be analyzed as a part of a hybrid cooling system for a

supermarket in a hot and humid climate.

15

Tower

Liquid Distributer

~~;~~m~~~~~~~- Packing Holder

Packing Suppon

Liquid Out -

Figure 1.1 Packed tower.

,

(a) (b) (c) (d) (e)

(f) (g)

Figure 1.2 Some random packings (a) raschig ring, (b) lessing ring, ' c 1

partition ring, (d) berl saddle, ( e )Intalox® saddle, (f) telleret te. (g) pall nng.

16

17

Figure 1.3 Plastic Intalox® snowflake packing.

Hct;cncralinn Sys1cm -

Li!juid Dcsiccanl ====

Wa1cr

-Ueal

l

-=;!JE"hm'"I-J j ~--~:ling : f Tuwer l f

·--l &nr~A l

-

~-AirOu1

:~6)!t!m~~ ~~:~~. ~~ ;~ ~~ ~!-~.. . . ~. -• .!,..!~.!·.!~~:~!-·!-P-!e;.~~-~~.¥.~:!-~-!! ;ic Packed Tower:;; h~5~!;i~!~i!i!!!! .!~6~--~ .!:!~~a;~-!-;.! .!~~?~a!i!~:.!~~;e ._a;a!(~;~..A" ~.!~ e;.! ~9!%

.MJn

l>roplcl Scpcralnr

Figure 1.4 The proposed system.

t--'

00

CHAPTER 2

COMPARISON OF MASS AND HEAT TRANSFER

COEFFICIENTS OF LIQUID DESICCANT

MIXTURES IN A PACKED COLUMN

2.1 Introduction

Dehumidification processes have important applications in batch-bin drying sys

tems. A liquid desiccant closed cycle drying system is shown in Fig. 2.1. The basic

concept of the system is to reduce the moisture content and warm up the air which

can be used for drying, a few degrees above the ambient temperature. This system

essentially consists of three processes: a batch drying process, a dehumidification

process, and a regeneration process. The liquid-phase falls downward due to gravity,

while the gas-phase moves upward through the packed tower due to the pressure

differential created by the blower. During the air-liquid contact, the air assumes

the vapor pressure of the desiccant solution. The vapor pressure of the desiccant

solution is lower than that of the humid air; hence, moisture transfer will occur

from the air to the solution. During the process there is a rise in air temperature,

due to the release of the latent heat of vaporization. Meanwhile, since moisture is

being absorbed, the desiccant is diluted. To reuse the desiccant in the process, it

must be regenerated to an acceptable level of concentration. The heat required for

regeneration can be supplied by any solar regenerator or an auxiliary heater. Then

19

20

the concentrated, hot desiccant transfers heat to the dehumidified air in a liquid

to-air heat exchanger and returns to the storage tank. The warm, dry air is sent

to the batch-bin dryer, and the concentrated desiccant is cooled and supplied from

the tank to the top of the tower.

The system mentioned above is an example of a dehumidifying packed column

application. In order to analyze the performance of packed columns for dehumidifi

cation processes, calculation of the mass and heat transfer coefficients of the packing

material used in the column is necessary. Queiroz et al. {1984) have used the Markel

integral approach to calculate the mass transfer coefficient for the dehumidification

operation in a spray tower. Recently, Lof et al. {1984) conducted experiments to

calculate the overall heat and mass transfer coefficients in a packed tower using

a lithium chloride solution. A recent report by Gandhidasan (1985) is related to

the present study. For an adiabatic operation in packed towers, he si:lggested the

use of individual phase-coefficients rather than overall coefficients. In this chapter,

calculations are performed according to this method.

This chapter is concerned with the interface transfer of heat and mass v.·hen

air is brought into contact with liquid desiccant mixtures. A theoretical study of

heat and mass transfer analysis in an air-desiccant dehumidification contact system

(packed column) employing liquid desiccants, namely calcium chloride (CaCl2),

lithium chloride (LiCl), and a new liquid desiccant mixture CELD (Ertas et al.,

1990a) is studied. This new liquid desiccant assures no solidification, and it is

stable. To be able to compare the behavior of the CELD solution with the other

two desiccants, calculations will be done using ceramic raschig rings. As mentioned

in the previous chapter, the characteristics of 0.5 in., 1.0 in., 1.5 in., and 2.0 in.

ceramic raschig rings are studied by Shulman et al. {1955) and Treybal (1980).

21

2.2 Evaluation of The Heat and Mass Transfer Coefficients

Many investigations of heat and mass transfer in fluid-solid systems have indi-

cated that the void space, fLo, is an important variable for the evaluation of the

heat and mass transfer coefficients. A knowledge of liquid holdup and the effec-

ti\·e interfacial areas is necessary to determine the void space in a packed tower.

"Holdup" refers to the liquid retained in the tower as film wetting the packing and

as pools caught in the crevices between packing particles (Treybal, 1980). There

are three different types of liquid holdup. The total holdup, cPLt, defined as the total

liquid in the packing under operating conditions, is expressed as the volume of the

retained liquid per unit volume of packing. The static holdup, cPL., is defined as the

liquid in the packing which does not drain from the packing when the liquid supply

to the column is discontinued. The operating holdup, cPLo, defined as the difference

between the total and static holdups, represents the liquid that will drain from the

packing and is also a measure of the liquid flowing through the packing when the

tower is in operation (Shulman et al., 1955).

The relation between the three holdups is given by

(2.1)

The void space, fLo, available for air flow in the packing is the difference of the void

space, f 1 , for the dry packing material and the total holdup. It is given as (Treybal,

1980)

(2.2)

For the purpose of estimating the total and static holdups the following empirical

equations are recommended (Shulman et al., 1955)

(2.3)

22

(2.4)

where /3 = 1d.s 9 and d.s is the diameter of a sphere with the same surface area as

a single packing particle. The constants, a, /, 8, 8, and ). are given in Table 2.1

(Treybal, 1980). The gas-phase mass transfer coefficient is given by (Shulman et

al., 1955)

Fe = 1.195G[d.sG' I ~c(1 - ELa)r0-36

(ScG)0.667

The liquid-phase mass transfer coefficient is given by (Shulman et al., 1955)

or

where C is the molar density of the desiccant.

(2.5)

(2.6)

(2.7)

For vaporization, the interfacial area is proportional to the total holdup and the

relationship is given by (Shulman et al., 1955)

<PLt a= 0.85aA-,

<PLo (2.8)

where aA is the interfacial area for the absorption or desorption with water or

aqueous solution and given by

_ [808G'] n(L')P aA- m p~5 , (2.9)

where m, n, and pare empirical constants, given in Table 2.2(Treybal,1980). Through

the heat-mass transfer analogy, the gas-phase heat transfer coefficient can be cal

culated assuming ~ = 1. (This assumption is reasonable because, during the

23

dehumidification process, the water vapor evaporates from the air stream and is

absorbed by the aqueous solution. Also, during the regeneration process, the wa

ter vapor evaporates from the aqueous solution and is absorbed by the air stream.

Hence, this can be approximated as a water-air system and a Lewis number of unity

can be assumed for this purpose.) The result is

hG _ l.195G'Cs[dsG' I JLG(1 - t:Lo)r0

.36

- ( PrG )0.667 ' (2.10)

where J D is the mass transfer group and is equal to (Shulman at al., 1955)

FG( ScG )0.667

JD = G ' (2.11)

and J H is the heat-transfer group given by

(2.12)

Similarly, the liquid-phase heat transfer coefficient is given by

h kth[dsL'Jo.4s(P )o.s L = 25.1- -- TL .

ds J.LL (2.13)

The corresponding volumetric coefficients can be obtained by multiplying equations

(2.5), (2.i), (2.10), and (2.13) with a, as obtained by equation (2.8).

2.3 Results and Discussion

As indicated in the above analysis, mass and heat transfer coefficients were calcu-

lated theoretically for the air-phase and liquid phase using 0.5, 1.0, 1.5, and 2.0 in.

ceramic Raschig rings. Calcium chloride, lithium chloride, and CELD solutions were

used as liquid desiccants and the results are correlated in the following equations

24

(for correlation, a 'Statistical Analysis System', (SAS), non-linear regression model

is used and the mean square error in the analysis is found to be approximately

0.0079).

2.3.1 0.5 in. Ceramic Raschig Rings

Calcium Chloride Solution

hLa _ 292575.38( L')o.s2 e-o.oo132TL eo.oo24e ( G')O.l37.

FLa 0.04 7( L')o.s2 e-o.oonsTL eo.oo64e ( G')o.137.

hGa 4072.53(L')O.l89eO.oolosTG( G')0.777.

0.158( L')O.l89 eo.ooo9sTG ( G')o.111.

CELD Solution

hLa 28321 1.25( L')o.s2 e -o.oo1TL eo.oo29e( G')0.137.

FLa _ 0.0809( L')o.s2 e-o.oousTL eo.oos1e ( G')0.137.

hGa _ 4072.53(L')0.189e0.00105TG(G')o.777.

0.158( L')o.1s9 eo.ooo9sTG ( G')o.111.

Lithium Chloride Solution

hLa 273489.14( L')o.s2 e-o.oo1TL eo.oo41e( G')0.137.

FLa 0.0960( L')o.s2 e-o.oon7TL eo.oosse ( G')O.l37.

hGa 4072.53( L')O.l89 eo.oo1osTG ( G')o.111.

0.158( L')o.1ss eo.ooossTG ( G't777.

(2.14)

(2.15)

(2.16)


- Calcium Chloride Solution

hLa _ 261963.44( L')0.63 e-0.00132TL eo.oo24e( G')o.o74.

FLa _ 0.042(L')0.63e-o.oonsh eo.oo64e( G')o.o14.

haa 3668.29( L')0.2eo.ooo9sTG( G')o.n.

0.142( L')o.2 eo.ooossTG ( G')o.n.

CELD Solution

hLa 253580.59( L')0.63 e -o.oo1TL eo.oo29e ( G')o.o14.

FLa _ 0.072(L')0.63e-o.oonsheo.oos7e(G')o.074.

haa 3668.29( L')o.2 eo.ooogsTG ( G')o.11.


hLa 244873.51( L')0.63e-o.ooiTL eo.oo4te( G')o.o14.

FLa _ 0.086( L')0.63e-o.oon7heo.oo6se( G')o.o14.

haa 3668.29( L')o.2 eo.ooo9sTG ( G')o.n.


Calcium Chloride Solution

hLa 149066.3( L')o.73 e-o.oot32TL eo.oo24e ( G')o.oo9.

FLa 0.024( L')0.73 e -o.oonsTL eo.oo64e ( G')o.oo9.

haa 2294.20( L')0.29 eo.ooossTG ( G')o.6s.

0.089( L')0.29 eo.ooo1sTG ( G')o.6s.

25

(2.17)

(2.18)

(2.19)

(2.20)

26

- CELD Solution

hLa 144294. 77( L')o.13e-o.oo1TL eo.oo29e( G')o.oo9.

FLa 0.041 ( L')0.73e-o.ooush eo.oos7( ( G')o.oo9.

he a _ 2294.20( L')0.29 eo.ooossTG ( G')o.6s.

0.089( L')0.29 eo.ooo7sTG ( G')o.6s. (2.21)

- Lithium Chloride Solution

hLa 139341.84( L')0.73e-o.ooiTL eo.oo41e( G')o.oo9.

FLa _ 0.049(L')o.73e-o.oou7TLeo.oo6se(G')o.oo9.

he a 2294.20( L')0.29 eo.ooossTG ( G')o.6s.

0.089( L')o.29 eo.ooo1sTG ( G')o.6s. (2.22)


- Calcium Chloride Solution

hLa 128416_56( L')o.66 e-o.ooi32TL eo.oo24e.

FLa - 0.021 ( L')o.66 e-o.oousTL eo.oo64e.

he a 2013.24( L')0.233 eo.ooos4TG ( G')0.64.

Fe a - 0.078( L')0.233eo.ooo77TG ( G')0.64. (2.23)

- CELD Solution

hLa 124306( £')0.66 e-O.OOlTL e0.0029e.

FLa 0.0355( L')0.66 e -O.OOliSTL eo.oosn.

he a 2013.24(L')0.233eo.ooos4TG( G')0.64.

Fe a 0.078( L')0.233 eo.ooonTG ( G')0.64. (2.24)


hLa - 120040.99( L')o.ss e-o.oo1TL eo.o041e.

FLa - 0.042( L')o.sse-o.oon7TL eo.oosse.

hGa - 2013.24( L')0.233 f0.00084TG ( G')0.64.

FGa - 0.078( L')0.233 f0.00077TG ( G')0.64. (2.25)

The following parameters were used to find the above correlation equations

G' = 0.3 to 1.1 kg/m2 .s;

L' = 0. 7 to 6. 7 kg/m2 .s;

TG = 25 to 70°C;

TL = 26.66 to 60°C;

~ = 32.5 to 40%.

27

It is believed that these parameters are reasonably free of end effects and can be

used in further calculations.

In the present study, data for the viscosity and density of lithium chloride and

calcium chloride, which are necessary for the evaluation of heat and mass transfer

coefficients, are obtained from Ertas et al. (1990b ).

The effect of some of the liquid desiccant parameters, such as concentration, ~,

temperature, TL, and flow rate, L', have been presented in the correlation equations

2.14-2.25. To better understand the effect of these parameters, the results are also

presented in Figs. 2.2 to 2.5.

The typical plot of Fig. 2.2 shows that lithium chloride has the higher rate of

change of liquid phase mass transfer coefficient, when compared with the calcium

chloride and CELD solutions. For a particular air flow rate, the volumetric liquid

phase mass transfer coefficient also increases as the solution flow rate increases. It

28

is apparent from this figure that the rate of the mass transfer coefficient increase

for the three desiccants is greater over the first 1 kgf(m2 x s) of liquid desiccant

mass flow rate. Although the CELD solution is a mixture of 50% lithium chloride

and 50% calcium chloride, a considerable increase in liquid mass transfer coefficient

is evident compared with the calcium chloride.

The effects of desiccant concentration on the liquid phase mass transfer coeffi

cient are shown in Fig. 2.3. The mass transfer coefficient increases with the increase

of desiccant concentration. However, as shown in Fig. 2.2, the mass transfer coef

ficient of the liquid desiccants is a stronger function of desiccant flow rate rather

than the concentration.

The liquid phase mass transfer coefficient as a function of liquid desiccant inlet

temperature is given in Fig. 2.4. From the figure, the mass transfer coefficient

decreases as the desiccant inlet temperature increases. This is expected because, as

the liquid desiccant temperature increases, the desiccant vapor pressure increases,

thereby reducing the performance of the liquid desiccant. Although the traces

from this figure revealed that the effect of inlet temperature on the mass transfer

coefficient is not significant, it is known that desiccant inlet temperatures should

be kept at a certain level for the best system performance.

The change of the liquid phase heat transfer coefficient with mass flow rate of

the liquid desiccant is given in Fig. 2.5. This figure indicates that the heat transfer

coefficient is a strong function of the desiccant flow rate. However, as can be seen

from the figure, the change of the heat transfer coefficients is almost the same for

all three desiccants.

From correlation equations 2.14-2.25, it is clear that the liquid phase mass and

heat transfer coefficients decrease as the ceramic raschig ring size increases for the

29

same operating parameters. This is due to the decrease in total surface area per

unit volume.

The CELD desiccant solution in a packed column greatly improves the mass

transfer as compared to a calcium chloride desiccant solution. This improvement is

nonlinear which suggests that the lithium chloride and calcium chloride do not act

independently in the solution. It is also apparent that the solution inlet temperature

and concentration do not significantly affect the mass transfer in the ranges expected

for air drying. The CELD desiccant solution demonstrates the same liquid phase

heat transfer as the pure lithium chloride and calcium chloride solutions. This

suggests that the solvent dominates the liquid phase heat transfer rather than the

solute.

In this chapter, we demonstrated the behavior of the CELD solution in compar

ison to its two components. The tower analysis of the CELD solution will be made

using the new Intalox® snowflake packing material.

30

Table 2.1: Constants for calculating total and static holdups

RASCHIG RINGS Porcelain Ceramic Carbon

a 2.09 x10 6 2.09 X 10-6 7.34x1o-6

I 1.508 1.508 1.104 b 2.47x10 4 2.47x1o- 4 5.94x10-4

() 0.376 0.376 0.376 ,\ 1.21 1.21 1.21

Table 2.2: Constants for calculating interfacial surface area.

Nominal size of Packing Range of L'

mm 1n. kg/ m2 • s m n p

13 0.5 0.68-2.0 28.01 0.2323 L' - 0.30 -1.04 2.0-6.1 14.69 0.01114 L' + 0.148 -0.111

25 1.0 0.68-2.0 34.42 0 0.552 2.0-6.1 68.2 0.0389 L' - 0.0793 -0.47

38 1.5 0.68-2.0 36.5 0.0498 L' - 0.1013 0.274 2.0-6.1 40.11 0.01091 L' - 0.022 0.140

50 2.0 0.68-2.0 31.52 0 0.481 2.0-6.1 34.03 0 0.362

Debumldltler

Wet Air -......... . . .. .. . . .. . . . . .. .. . . . . . . . . . . . ...... . .. . . . . .. . . . . . ...... . . . . . . . . . . . . . . . . . . . . . . . . ...... . .. . . . . . . . . . . . .. . .. .

I

Pump. Batcb-BID 1

Dryer L~-

--Dry, Warm air

.Pump

Heat Exchar~~er

Weak Desiccant r-- '':

Strong Desiccant

Figure 2.1 Schematic of the closed cycle drying system.

31

0.29524

0.0

- I i /"1 -~": i

~~)~· -< :· I """"--(.~:: . v

I : ... ~ ~-. I/

rv~··.- / I ,c',-W

/ v ·,\.e. -/ / r'l: ~ t-.o:,.... 4~r·~ ,.. ,.,e,, ~ "/ ~

;~~ !' v TL. = 31°C :/ v 1/ "(

0.0

l; m = 40% ,. G' = 0.7 kg/ m2 s

I I

Superficial Liquid Mass Velocity, kg I m2 s

I

l

7.00

Figure 2.2 Effect of liquid flow rate on volumetric liquid phase mass transfer coefficient for 0.5 in. ceramic raschig ring.

32

0.22869 -

0.10509

32.5

I Lithi ~mC ~lorid~

I I

I

rl=:T n I i

~ I

I

TI..· = 31°C L m = 3 kg/m2 s G' = 0.7 kgfm2s I

r~lr 11m C hlori be

Desicant Concentration, Weight%

I j

I

40.0

Figure 2.3 Effect of desiccant concentration on volumetric liquid phase ma.:>s transfer coefficient for 0.5 in. ceramic raschig ring.

33

0.17593 IL

.:;;: [C

-J\...

0.08276 I

25.0

~h-l I l I

IOl Ide

l

:L l[)

I i

I ~ =40% L' =) kglm2s a· = 0. 7 kg!m2s

I IICIU n< :h orin~ I

I

Desicant Inlet Temperature, degree C

I ,-

I I f-

:-

I

60.0

Figure 2.4 Effect of liquid inlet temperature on volumetric liquid phase

mass transfer coefficient for 0.5 in. cerarrric raschig ring.

34

894587

0.0

0.0 Superficial Liqutd Mass Velocity, kg I m2 s

7.0

Figure 2.5 Effect of liquid flow rate on volumetric liquid phase heat transfer

coefficient for 0.5 in. ceramic raschig ring.

35

CHAPTER 3

EVALrATIOI\ OF THE PERFORMANCE OF THE DEHUMIDIFICATION

PROCESS IN A PACKED COLUMN

3.1 Introduction

The dehumidification process in a packed tower is a result of simultaneous heat

and mass transfer. The difference in the vapor pressure of the desiccant solution

and the air, at the interfacial surface, is the driving force for mass transfer. As the

moisture transfers from the gas phase to the liquid phase latent heat is liberate.

This increases the temperature of the flow stream. The heat and mass transfer

coefficients that are calculated in the second chapter, F Ga, F La, hGa, and hLa are

the individual phase coefficients. The second chapter deals with the comparison of

heat and mass transfer coefficients for CELD and its two components, namely CaC12

and LiCl with four different size raschig rings. The individual phase coefficients are

recommended for use in packed column performance evaluations (McAdams, 1949).

The performance of the packed column will be evaluated by using Intalox® snowflake

packings. This new material is promising as an efficient new packing material.

In this chapter, the heat and mass transfer coefficients for CELD, using lntalox®

snowflake packing material will be calculated, and the performance of the dehumid

ification process in a packed column will be evaluated. Similar studies have previ

ously been done in which the liquid phase resistance to mass transfer was assumed

to be negligible (Lof et al., 1984). In this study, the resistance of mass transfer

36

37

will be taken into consideration, and evaluations will be made using the method

suggested by Gandhidasan, (1985).

3.2 Evaluation of Heat and Mass Transfer Coefficients

The Intalox® snowflake packing material is a relatively new packing material.

The data for the characteristics of the packing are not available in an organized

form. The properties of the packing are calculated from the information which is

obtained from the packing manufacturer.

3.2.1 Holdup Calculations

The total holdup plot for Intalox snowflake packing obtained from Norton,

(1987), is shown in Fig. 3.1. As seen in the figure, the pressure drop across the

packing is an important criteria for determining the holdups. Approximately 0.5

in. of water column pressure head is anticipated for every foot of packing of the de

humidification process (Ertas et al., 1990c). The total holdup equation is (Norton,

1987)

cPLt = 2.9 X 10-3 L'l.07,

where L' is in kg/(m2 x s) and cPLt is in m3 /m3•

(3.1)

The snowflake packing is especially designed for lower static holdups. The static

holdup is assumed to be 2% of the total holdup.

3.2.2 Interfacial Surface Area Calculations

The snowflake packing material provides 91.8 m2 /m3 of dry surface area (Norton,

1987). However, during the operation, the effective interfacial surface area for the

mass and heat transfer is actually much less than this value. Shulman et al. (1955)

38

calculated the effective interfacial surface area by taking the ratio of kGa and kG,

where these values are experimentally measured. They conducted these tests for

ceramic raschig rings and berl saddle packings. Treybal (1980) correlated these data

from the figures in the form of non-linear equations (i.e., effective interfacial surface

area is given as a nonlinear function of liquid and air flow rate).

In this study, a similar procedure will be followed for the calculation of the

effective interfacial surface area for snowflake packing material. The overall mass

transfer coefficient on a mass basis, K'aA, as obtained from the snowflake packing

manufacturer (obtained by personal contact from Norton Chemical Process Prod

ucts Company, John R. Sauter, Chamberlain Labs. Stow, Ohio 44224) , is

K' a A = 0.209205( L')0.4364( G')0.695' (3.2)

where L' and G' are in terms of lb/(ft2 x hr) and K'aA is in terms of lb/{ft3 x hr).

The above equation is obtained through a chilled water dehumidification experi

ment. In this study, it is assumed that during the dehumidification process desiccant

air contact system behaves like a water-air contact system. K'aA also satisfies the

equation

J dHG = K'aA z. H(;-HG G'

(3.3)

The overall transfer coefficient, K'aA, is related to the KGaA, which is a overall

transfer coefficient on a molar basis, by the equation (McAdams, 1954; Kern, 1950)

(3.4)

39

where Pt is the total pressure and Mv is the molecular weight of water vapor.

Even though KcaA is the overall mass transfer coefficient, it can be converted to

individual phase transfer coefficients by the relationship (Shulman et al., 1955)

1 1 m --=--+--K caA kcaA kLaA'

(3.5)

where m is Henry's constant. It can be calculated using Henry's law as (Treybal,

1980)

(3.6)

where the p· is the gas phase vapor pressure, and Pt is gas phase total pressure. kc

can be calculated from (Shulman et al., 1955)

Fe kc = --,

PsM (3.7)

where F G is given by equation (2.5) and kL is given by equation (2.6). The interfacial

surface area is calculated from the equation (3.5). For the calculations of PsM and

m, air phase is assumed to be dry but at the interface, it is assumed to be saturated

with water vapor. The effective interfacial surface area, a, can be calculated from

equation (2.8). The heat and mass transfer coefficients are obtained from equations

(2.5), (2. 7), (2.10), and (2.13). The corresponding volumetric coefficients can be

obtained by multiplying these equations by the result of equations (3.5) and (2.8).

3.2.3 Volumetric Heat and Mass Transfer Coefficients

The results of the above equations are calculated for CELD solution and are

correlated using a SAS non-linear regression model. The mean square error for the

40

calculations are: for FLa is 0.00345; for Fca is 0.00107; for hLa is 0.00344; and for

hca is 0.00105. The correlated equations for snowflake packing with CELD solution

are

hLa 130507( L')0.44e-o.ooo1sTL eoms~ ( G')o.69.

FLa - 0.0373( L')0.44e -o.ooo3TL eo.o2n~ ( G')o.s9.

he a - 187 .5( L')O.o4eo.osTG ( G')l.33.

Fe a - 0.00727( L')o.o4eo.osTG ( G')l.33.

The above equations are valid for the following ranges

G' = 0.3 to 1.1 kg/m2 s;

L' = 0. 7 to 6. 7 kg/m2s;

Tc = 25 to 70°C;

TL = 26.66 to 60°C;

~ = 32.5 to 40%.

(3.8)

The physical properties of the CELD that are used in these calculations are

obtained from Ertas et al. (1990a).

3.3 Simultaneous Analysis of Heat and Mass Transfer in the Packed Column

Vilhen air and desiccant come in contact in the packed column, because of the

difference in their vapor pressure, heat and mass transfer occurs, until a certain

equilibrium is reached. As the liquid falls down through the tower, it absorbs the

moisture from the air and its vapor pressure increases. In contrast, as the gas

moves up in the column, its vapor pressure decreases with its water vapor content.

The following equations obtained from Treybal (1969) are used to simulate the

41

performance of a packed column. The following assumptions are necessary for the

calculations of an adiabatic desiccant-air contact system:

1. Heat loss through the walls is negligible.

2. Interfacial surface temperature is equal to the temperature of the bulk

liquid. The order of error due to this assumption is very small and

can be seen in Treybal, (1969).

3. Packing.is adequately irrigated and, therefore, the interfacial surface

area is the same for both heat and mass transfer.

4. No axial dispersion occurs in the packed tower. Co et al. (1971)

studied the mixing in the packed column by using adequate liquid

distribution and showed that assuming constant properties in a hori

zontal section of the packing is reasonable.

A differential section of a packed tower is shown in Fig. 3.2. The following equa

tions are derived using basic heat and mass balances in the section. The enthalpy

of the liquid is given by

HL CL(TL- To)+ ~H.,

dHL - CLdTL + d~H •.

(3.9)

(3.10)

In the calculations, the change of ~H. with the concentration is neglected, since its

effect on the performance of the column is very small. Similarly for air

Ha - CB(Ta- To)+ Yc[Cc(Ta- To)+ >.to],

dHa - CBdTa + YcCcdTa + dYc[Cc(Ta- To)+ >.to]·

(3.11)

(3.12)

42

The direction of the heat and mass flux is taken positive in the direction of air to

liquid. The mass transfer at the interface of air and liquid is

Ncadz = Fe ln C1 adz= -GBdYc, [ 1- y '] 1- Yc

(3.13)

or

dYe=_ Feazn[1- Yci]. dz GB 1- Yc

(3.14)

The heat transfer at the interface is

(3.15)

Applying the Ackermann correction for simultaneous heat and mass transfer gives

dl' h' GB[Cc7z] ea = -1- exp[GB(C/r" )fheaJ"

(3.16)

The transfer of moisture to the liquid from the interface is

Ncadz = FL ln c adz. [ 1- X ]

1- Xci (3.17)

A mass balance in the control volume in Fig. 3.2 gives the result

(3.18)

An energy balance on the air side of the Fig. 3.2 gives

43

Combining equations (3.12) and (3.19) reduces them to

(3.20)

or in reduced form

h(;a(TG- Ti) --=-Gs(Cs + YcCc)"

dTG dz

(3.21)

The liquid temperature gradient given is derived in a similar way.

(3.22)

The interfacial condition can be obtained by equating equations (3.13) and (3.17).

Yci = 1 - (1 - Yc) [ 1

- Xc] ~ · 1- Xci

(3.23)

The above equation can be solved simultaneously with vapor-liquid equilibrium

data for CELD solution. Using a SAS linear regression model the available vapor

pressure data is correlated in the following equation. (The root mean square error

is 0.0000327.)

0.05720066 Yci = 0.07983109 + 0.00172330877: - ----

Xci (3.24)

Equations (3.23) and (3.24) are solved simultaneously for Yci and Xci, for every

section of the column. The analysis presented above will evaluate the performance

of the column with the given: liquid inlet temperature and concentration; air inlet

temperature and concentration; exiting air humidity or desired level of dehumidi-

fication; flow rates of liquid and air. The output will be the height of the packed

44

column, exiting au temperature and humidity, and liquid temperature and con

centration. Equations (3.14), (3.21), and (3.22) are used to define liquid and air

properties in every section of the column. This method is called the 'incremental

research method'. The calculation starts at the bottom of the tower where z=O.

3.4 Computational Procedure

The computation starts with reading the inlet conditions of air and liquid, and

the exit humidity of the air. A reference temperature is assumed. Calculations start

by assuming the exit temperature of the liquid . Then the packing height is divided

into N sections. N is chosen as 100. The liquid and gas exit conditions are calculated

from the mass and heat balance in every section. The program is initialized from

the bottom of the tower with the guessed liquid exit temperature. First the heat

and mass transfer coefficients are calculated and then interface moisture condition

is calculated. The concentration and temperature gradients are calculated along

the z direction. After this step the incremental tower height is calculated for the

given change in the absolute humidity. Then for the next incremental tower section,

the air and liquid properties are updated.

The same calculation of interface moisture conditions are repeated for each sec

tion. When the ( N + 1) cross section or Nth tower section is processed, the program

reaches the top of the column. Here the calculated liquid inlet temperature is corn

pared with the given value. If the difference of the two values is larger than a given

error, a new appropriate liquid exit temperature is assumed and iterations are re

peated until the convergence is reached. An over relaxation method is used for this

iterative calculation. The 'Successive Over Relaxation' (SOR) is equal to 1.5.

45

The results are liquid and gas temperatures, liquid concentration, air exit hu

midity and tower height. The liquid and air humidity profile can also be determined,

if desired.


As explained in the procedure, the computer simulation gives the height of the

packed column as the output for the desired level of performance. The performance

of a column is analyzed in order to find the optimal design for the tower. An

important criterion for the optimal column is the packing height. Fig. 3.3 shows

the performance of packed columns for different packing heights. As the packing

height increases, the outlet humidity of the air decreases. However the relationship

between the two is not linear. As the packing height increases, the rate of decrease

of humidity decreases. This is a result of a decrease of the vapor pressure of the air.

When the tower reaches a certain height, the air which is close to the top of the

tower will be dehumidified, and as a result of this, its vapor pressure will decrease.

Even though the desiccant is strong (i.e., has a low vapor pressure), the driving

force (i.e., the difference between the vapor pressure) will be very low. This will

result in a lower mass transfer. A 0.5 m. packing height is chosen to work with

for constant height calculations; because after this height, the rate of decrease of

humidity becomes smaller.

Most of the applications of dehumidification processes have special air outlet

temperature requirements. The change of outlet air temperature with the packing

height is shown in Fig. 3.4. The air temperature increases with the packing height

for the first 1.3 m. of the packing. After that the air temperature starts to decrease.

It is simply because after the first meter of packing, there is a very low mass transfer

rate (i.e., the latent heat production is very low). Air which is about to leave the

46

tower loses heat to the entering liquid. As a result of low latent heat production, air

loses heat to the liquid desiccant whlch is at a lower temperature. Liquid and air flow

rates, liquid and air inlet temperatures, air inlet humidity, and liquid concentration

at the inlet are other factors that affect the packing tower performance.

The effect of some of these parameters will be discussed to be able to understand

the behavior of the tower under different operating conditions. Fig. 3.5 shows the

change in the performance of the column for a fixed height of packing. The air outlet

moisture keeps on decreasing with the increase in the liquid flow rate. However,

similar to the packing height, the rate of decrease of humidity decreases with the

increasing liquid flow rate. The reason for thls is that the larger liquid flow fills

up the void space available for air flow, w hlch causes a decrease in the interfacial

surface area. As a result, the mass transfer decreases.

The increasing desiccant flow rate causes a decrease in the exit temperatures of

liquid and gas phases (Fig. 3.6). Thls is due to the increased amount of liquid which

absorbs the produced latent heat. The decreasing rate of latent heat production

also affects the change in the temperature of both phases. As seen from the figure,

air temperature is affected less than liquid temperature. The increasing liquid

temperature is an advantage for the regeneration process.

Fig. 3. 7 shows another important factor of analysis, the column performance

versus increasing liquid inlet temperature. As the liquid inlet temperature increases,

the air outlet humidity also increases. Increasing liquid temperature causes an in

crease in the vapor pressure of the solution which results in a lower mass transfer

rate. The effect of increasing liquid temperature on air and desiccant outlet tem

peratures can be seen in Fig. 3.8. Even though there is a larger difference between

the liquid and air temperatures, the rate of increase in both phases are similar.

47

The liquid phase absorbs the latent heat that has been produced during the phase

change.

The comparison of the performances of a 0.5 m. packed column with 3 and 4

GPM flow rates for 40% and 45% desiccant concentrations are shown in Fig. 3.9. As

seen for both concentrations, increasing liquid inlet temperatures causes a decrease

in the performance of the packed column. All four cases show a decrease in the

performance; however, 4 GPM flow rate with 45% concentration has the highest

slope. The 4 GPM flow rates are affected by the increase in the liquid temperature

more than 3 GPM flows. This behavior can be observed from the slope of the

curves.

Fig. 3.10 shows the effect of increasing air inlet temperature on the change of

humidity for 100% saturated air. As shown in the figure, increasing air temperature

causes an increase in the change of humidity. However, the exit humidity also

increase. An increase in the air inlet temperature causes a decrease in the change

of air temperature and an opposite effect on the change of liquid temperature, as

seen in Fig. 3.11.

The performance of the packed column for different climates with different hu

midities is another important factor. Fig. 3.12 shows the performance of the col

umn in three different climates in the United States; Florida, Texas, and Okla

homa. The performance of the column decreases for increasing humidities. How

ever, the amount of the extracted moisture increases (~humidity(Oklahoma)=3.21

10-3 kg/kg, ~humidity(Texas)=4.88 10-3 kg/kg, ~humidity(Florida)=6.57 10-3

kg/kg). The effect of increasing moisture in air on the outlet liquid and air tem

peratures can be seen in Fig. 3.13. The increase in temperature in both phases is

evident. However, the rate of increase in liquid phase is considerably larger than

the rate of increase in the gas phase.

48

Some of the parameters that affect the performance of the packed column are

discussed above. However the change in the air flow rate is another important

factor. Computer runs show that air flow rate has an opposite effect to liquid flow

rate. The increasing flow rate requires a longer tower for a desired level of operation.

10 0 0 0 e o

;::;- 0 E ; 0 ~

=?

8 c.

s. 4

.3

2

.I ::lo ~08 a: .06 w ..... ~ 0

0

0

4

3

2

00 1

WATE~ L!OUID RATE. lbff11• h

891000 2 :: J S6 79910000 2 3 4 567 0 ~-I

I

I

I I

: I

I :

! j

I ' ~ ~ '

' .J i""L_

'

, v I "' // I

' ~ / I I I .I llv

' ? i ! '

'

~E

t::=g ./8 /A

0 ~-

8 6 5 4.., 3=

-v

l.n

I~ ~

2' ~ :;,

0 c ~a

OS::t 06a::

w ..... 04~ 3 0

2 0

0. 2 3 .; : 6 78910

01 2 3 ~ 56 789100

WATEF= LIQUID RATE. kgfm2 • s

AP

mm H20/m in. H20/ft

A-0 0.0

8-42 0.5

C-83 1.0

0-125 1.5

E- 167 2.0

Figure 3.1 Total holdup versus liquid inlet flow rate for lntalox~ snowflake

packing.

49

'

Interfacial Surface

Area adz

Ga Yc+ dYe TG +dTG

•I'

50

--~-~-~~~:~:~-~--~-~:~:~-~-~:~:~--~-~:~:~-~--~-=:~:~-~~=-~--~-~-~-~-~-~-~-~-~-~-~-~-~-~-~-~-~----z+dz ~ ----------------

I

Ti _..!, -I

Liquid Desiccant Air I

Xci I _ .... "j ..:, Yci I I I

I ~-----------------1 ~--------------------~-------------------------------~~~~~~~~~~~~~---z --------- ____________________ i ________ _

'If L+dL

TL + dTL Xc + dxc

Ga Yc TG

Figure 3.2 Differential section of a packed tower.

.... :0: ..... 38 Tc = 34.44°C 0 t:ll)

TL = 32.22°C ~ - L' = 4.0GPM .... ~ ct:

G'= 1091 CFM ~ 33 ~ = 40% ..... 0

Wj = 38.6 10·3 kg/kg t:ll) ~ ..., c >.. 28 -"'0 .... .... .... -.... t> 23 -= 0 .... <

18 4-------~-----------r---.--~------~--~--~ 0.0 0.5 1.0

Packing Height, m.

Figure 3.3 Dehumidification performance of the packed column "·ith Intalox® snowflake packing.

51

u c Q.) .... ::l --1:.: Q.) Q. .... .... . o I:"'"

u -::l 0 .... <

37 ~-----------------------------------------,

36

35

34 0 J 0.5

Ta = 34.44°C TL = 32.22°C L' = 4.0GPM G' = 1091CFM ~ = 40% wi = 38.6 w-3 kg/kg

1.0 1.5

Packing Height, m.

figure :).-1 Change of air outlet temperature with the packing height for the I nt alox:8 snowflake packing.

52

...: co:: Tc = 34.44°C '-0 TL = 32.22°C Cll 28.5

...:..:: G' = 1091 CFM - ~ = 40% ... ~ Wj = 38.6 w-3 kg/kg eo: ~ Packing Height = 0.5 m '-0 Cll

...:..:: 27.5 "" ----i-:-9 ,.. c: :::::

:I: 26.5 -~ -6 ... -<

25.5~~~--~~~~~~~~~~~~~~--~~--~

2 3 4 5 6 7 8

Desiccant Flowrate, GPM.

Figure 3.5 Change in the exit humidity of air with the liquid flow rate for Intalux® snowflake packing.

53

44

cJ 0

~ :l ~ ... 0 c. ,.. c: ~ 39

~ :l 0

34

2 4

Ta = 34.44°C TL = 32.22°C G' = 1091 CFM

~ = 40% Wj = 38.6 w-3 kg/kg Packing Height = 0.5 m

6

Desiccant Flowrate, GPM.

8

Figure 3.6 Change in the liquid and air temperature with the desiccant flow rate for lntaloxiB snowflake packing.

54

32 ...

Tc = 34.44°C c-:: '- L' = 3.0GPM 0 c..o 31 G' = 1091 CFM ~

~ = 40% -... wi = 3R.6 1 o-3 kg/kg ~

ee Packing Height = 0.5 m ~ 30 .._ 0 c..o ~

....., 29 ----

;;:., -=s! 28 ,... = :I

::r: ts - 27 6 ... <

26 i-~~~----r-~~----~--~------~--------~ 25 30 35 40 45

Desiccant Inlet Temperature, °C.

figure 3.7 Change in the air outlet humidity with the increase in the desiccant inlet temperature for lntalox® snowflake packing.

5.5

TG = 34.44°C 48 L' = 3.0GPM

G' = 1091 CFM ~ = 40%

cJ Wi = 38.6 } Q·3 kg/kg 0 Packing Height = 0.5 m ~ ::::;

43 i:o ... a) Q..

E ~ 4) -::::; 0 38

33 ,_------~----~--~--~--------~------~--~ 25 35 45


Figure 3.8 Change in the desiccant and air outlet temperature with the incrPase in the desiccant inlet temperature for Intalox® snmdlake packing.

56

<

22 ~-----------------------------------------------.

21

20

19

18

:7

16

Tc = 32.22°C G' = 1091 CFM Wj = 24.18 1 o-3 kg/kg Packing Height = 0.5 m

15 ~--~----~--~----~--------~--~--~----~--~ 28 30 32 34 36 38


Figure 3.9 Change in the air outlet humidity with the change in desiccant inlet temperature for -!0% and 45% concentrations, ,,·ith 3 and 4 GPI\J flow rate of desiccant for Intaloxl3 snowflake packing.

.... C':l '-0 eLl ~ -.... ~

~ ~ '-0 eLl ~

,...., -._.

>.. -:-2 E :::: -.... c: ~ Oll c: co:

..::: u

l2.5

11.5

10.5

9.5

8.5

7.5

31 32 33

TL = 32.22°C L' = 3.0GPM G' = 1091 CFM

~ = 40% Packing Height = 0.5 m

34 35

Air Inlet Temperature, °C.

36

Figure 3.10 Change in the air humidity ·with the increase in the air inlet temperature for lntalox~ snowflake packing.

58

cJ c

c:

13 ~------------------------------------------~

11

9

3 Air

TL = 32.22°C L' = 3.0GPM G' = 1091 CFM ~ = 40% Packing Height = 0.5 m

1 C:==:===::::::::~ 31 32 33 34 35 36

Air Inlet Temperature, °C.

figure 3.11 Change in the desiccant and air temperatures with increase in ti1t' air inlet temperature for lntaloxl8 sno,\·fiake packing with lOOo/c saturated air.

59

~ ~

...; .... ~

eo: '- TL = 32.22°C c eJ) To = 32.22°C .:.t:. - 21 L' = 4.0GPM ...

G'= 1091 CFM ~ ;:

~ = 40% ~ '- Packing Height = 0.5 m 0 eJ) 20

.:.t. ,..., -._.

>.. -~ 19 .... = ....

::c .A Florida 0 ::I 18 0 Texas 0 ... < c Oklohama

17 20 22 24 26 28

Air Inlet Humidity, 10·3 kg of water I kg of air.

Figure 3.12 Change in the outlet humidity with the change m the inlet humidity of air for Intalox® snowflake packing.

60

3E

TL = 32.22°C TG = 32.22°C

37 L' = 4.0GPM G'= 1091 CFM ~ = 40%

cj 36 Packing Height = 0.5 rn c

~ ::; ::; 35 ... ~ c.. ,.. c: ~ ~ 34

-::; 0 Air

33

32 +-------~~----~----~--------~------~~ 20 22 24 26 28

Air Inlet Humidity, 10·3 kg of water I kg of air.

Figure 3.13 Change in the outlet temperatures of desiccant and air with the change in air inlet humidity for Intalox~ sno"·fiake packing.

61

CHAPTER 4

APPLICATIONS OF DEHUMIDIFICATION SYSTEMS

FOR AIR CONDITIONING UNITS

4.1 Introduction

In many southern states, the summer electrical load is significantly influenced

by the use of air-conditioners. The energy required to operate these units create a

daily late afternoon peak load. These electrical demand curves reach even higher

peaks during the hottest summer days. One of the most attractive features of

the proposed project, which uses solar energy to provide comfort cooling, is the

incidence of maximum solar radiation when the need for cooling is greatest.

Different types of solar cooling systems are described in the literature and they

can be broadly classified as closed-cycle and open-cycle systems. Much attention

has been given to closed-cycle systems, particularly absorption refrigeration and

vapor compression systems while comparatively little attention has been given to

open cycle space cooling systems.

Absorption refrigeration systems have been extensively studied by many in

vestigators and successful combinations are LiBr-H 2 0 and H20-NH 3 systems. In

the absorption system, solar energy need not be converted into mechanical energy.

Hence this system seems to be more attractive. However, the major disadvantages

of this system are the requirement of a large amount of cooling water for remov

ing the waste heat from the condenser and absorber and the low vacuum pressure,

62

63

which must be maintained in the evaporator and absorber at all times. The vapor

compression system consists of the Rankine cycle heat engine, which produces me

chanical output to drive a conventional air-conditioning compressor. It has been

studied and the conclusion is that this system would out perform the absorption

system when coupled with a focussing collector. The conversion efficiency would

be low and the requirement of a sun tracking concentrator makes the system more

expensive and cumbersome.

The proposed hybrid desiccant cooling systems combine a desiccant dehumidi

fier with a vapor-compression unit to meet the supermarket air-conditioning load.

The proposed hybrid cooling system utilizes the desiccant to meet the latent load

and a vapor-compression unit to handle the sensible load. The system uses solar

energy for the regeneration of the liquid desiccant. One of the major advantages

of this system is that the required temperature of 50°-65°C for regeneration of the

absorbent solution can be easily obtained by using a simple solar regenerator. In

this chapter, the concept, design, and cycle analysis of the proposed cooling system

is described.

4.2 Air Conditioning in Supermarkets

Supermarkets have special requirements for air-conditioning. Because of open

refrigeration stands and a high moisture production rate, the latent load to sensible

load on the air-conditioning system is very high. In most cases, a conventional

system uses central air-conditioning for sensible cooling and refrigerated cases for

the latent load, moisture removal. To be able to condition the air to the required

level, a large amount of air flow rate is needed. This is a very inefficient way of

using energy. Supermarkets consume approximately 4% of U.S (Burns et al., 1985).

electrical energy. The proposed hybrid cooling system promises a very efficient way

64

of air-conditioning in supermarkets. The proposed system, first dehumidifies the air

through a packed tower which handles the latent load, and then sensibly cools the

air through the vapor-compression unit that has a high COP due to a low cooling

rate. As the air is not being reheated in the proposed system, there is no waste

energy.

4.3 Concept of the Proposed Cooling System

In southern parts of the United States, both temperature and humidity are high

in summer months and hence both dehumidification and cooling are necessary. In a

vapor compression unit, the humid air is first drawn over an evaporator, chilled to

the dew point, often to a temperature below that is desired in the occupied space.

Then the chilled air is drawn over the condenser coil, where it is reheated. In the

proposed system there is no need for reheating. The primary energy input into the

system is in the form of heat required to regenerate the absorbent solution. This

temperature is considerably lower than that required for operation of vapor com

pression systems. The proposed hybrid desiccant cooling system is an alternative

to a conventional vapor compression system as a means of reducing electrical en

ergy consumption and energy costs. Since the vapor compression unit in a hybrid

system only has to remove a portion of the load, the electrical energy consumption

is substantially reduced over that of a conventional system. Further, since the va

por compression unit no longer has to cool the air below dew point temperatures,

the evaporator temperature may be raised. This increases the COP of the vapor

compression unit and reduces energy input to the compressor. By separating the

performance of latent and sensible cooling into two different components, it is now

possible to reduce humidity without simultaneously reducing the supply air temper

ature below desired levels. Dehumidification is no longer dependent on the limits

65

of the Yapor-compression unit. Various designs for liquid desiccant dehumidifiers

have been proposed and studied. However the present work considers a packed bed

dehumidifier and CELD as the liquid absorbent for the proposed cooling system.

4.4 Description of the Model

The proposed system is designed to provide cooling for a supermarket. The

schematic of the system is shown in Fig. 4.1, and the processes are shown on the

psychrometric chart in Fig. 4.2. The system consists of a conventional vapor

compression unit, an auxiliary heater, a solar regenerator, three heat exchangers,

a dehumidifier, a cooling tower, pumps for the solution and water, two fans and

the ducts. The assumed ambient conditions are 32.22°C and 70%RH. The indoor

design conditions of the conditioned space are 23°C and 0.0086 kg of water /kg

of air (Appendix A). The mixture of return air from the supermarket and fresh

air from the atmosphere enters the dehumidifier at 'd' as shown in the Figure.

After the dehumidification process, the air is warmed up and dried, which is not

suitable for comfort conditions. Hence, after removing the latent cooling load in

the dehumidifier, the dried air at condition 'a' is passed over the evaporator coil

in the conventional vapor compression unit. The air is sensibly cooled to condition

'b', and the conditioned air is then delivered to the building.

A simple system using atmospheric air through the solar regenerator is con

sidered in the proposed cooling system. An open solar regenerator is simple and

effective, but in humid climates the evaporation rate is low. The weak desiccant

from the dehumidifier is pumped through heat exchangers to the solar regenerator.

The regenerator consists of a single galvanized iron sheet painted black, over which

the weak absorbent solution flows as a thin film. The collector is insulated on the

underside. To reduce heat loss, it is covered by a single glazing of 3 mm thickness.

66

·water evaporating from the liquid surface is removed by wind gusts. The heat re

jected by the condenser of the vapor compression unit is used to preheat the weak

desiccant after the dehumidification process.

4.5 System Analysis in the Supermarket

The proposed system for the supermarket will be analyzed, component by com

ponent, for complete performance.

4.5.1 Supermarket

For an average supermarket, one can assume a surface area of 2,800 m2, and

it produces moisture at a rate of 11.3x 10-3 kg/s (Manley, 1985). However, the

supermarket temperature needs to be kept below 24°C and 55% relative humidity

or 0.010244 kg of water/kg of air absolute humidity (Burns, 1985). Because of the

open refrigeration stands, the latent load to sensible load ratio is very high. In order

to maintain the comfort in the supermarket, air is cooled down to 23°C and dried

to 0.0086 kg of water/kg of air (Burns, 1985). Air is drawn from the building at a

rate of 6.95 kg/s and is mixed with 10% of the outside air (0.695 kg/s).

4.5.2 Mixing of Air

The air coming from inside the building mixes with the outside air. The following

conditions were used.

Air leaving the building

mass flow rate - 6.95 kg/s (1314 7 cfm),

temperature - 24°C,

absolute humidity - 0.010244 kg of water/kg of air.

Outside air

mass flow rate- 0.695 kg/s (1315 cfm),

temperature - 32.22°C,

absolute humidity - 0.02142 kg of water/kg of air.

Mixed air

mass flow rate- 7.645 kg/s (1315 cfm),

temperature - 24. 70°C,

absolute humidity- 0.01126 kg of water/kg of air.

The detailed calculations are given in Appendix B.l.

4.5.3 Dehumidification Tower

67

The dehumidification tower has a diameter of 48 in. and is packed with snowflake

plastic packings to a height of 1 ft. The dehumidifier is modeled according to

calculations in Chapter 3. The inlet and outlet conditions of the to·wer are as

follows.

Air inlet

TA - 24.7°C,

wA - 0.01126 kg of water/kg of air,

G' 7.645 kg/s = 14462 cfm.

68

Air exit

TA - 27.262°C,

WA 0.0086 kg of water/kg of air,

G' - 7.6246 kg/s = 14424 cfm.

Desiccant inlet

TD 32.22°C,

~D - 40%,

L' 0.5258 kgjs.

Desiccant exit

TD - 41.638°C,

~D 38.5%,

L' 0.5462 kg/s.

4.5.4 Vapor-Compression Unit

Air passes over the coils of the evaporator which uses refrigerant-12 as the re

frigerant. Air enters the evaporator at

T 27.262°C,

69

w - 0.0086 kg water /kg air,

m 7.6246 kg/s.

and leaves the evaporator at

T 23.ooooc,

w 0.0086 kg water/kg air,

m - 7.6246 kgjs.

The heat absorbed in the evaporator is Q EV = 32.64 k Vv (kilowatt) and the heat

rejected at the condenser is Qed= 36.11 kW. The refrigerant flow rate is 0.2834 kg/s.

Detailed calculations are given in Appendix B.2.

4.5.5 Solar Regenerator

The amount of water that should be removed from the desiccant is 73.44 kg/hr,

and the solar panel surface area required for this operation is 95.37 m 2 • Detailed

calculations are given in Appendix B.3.

4.5.6 Heat Exchangers

From the information given in Appendix C, we know that the solution leaves

the regenerator at 60°C. The hot desiccant at the outlet of the regenerator needs

to be cooled down prior to entering the dehumidifier. The weak desiccant which

comes out from the dehumidifier can first be heated with the warm condenser air in

the air-to-desiccant heat exchanger, and then with hot desiccant in the desiccant

to-desiccant heat exchanger. The heat rejected from the condenser produces air at

45.0°C and at 2.81 kg/s. This air is used for heating the weak desiccant at 41.64°C

70

to 43.80°C in an air-desiccant heat exchanger. Detailed calculations are g1ven m

Appendix B.4.1.

Further, the cooler weak desiccant and warmer strong desiccant transfer heat in a

desiccant-to-desiccant heat exchanger, shown in Fig. 4.3. After passing through the

desiccant-to-desiccant heat exchanger, the weak desiccant is warmed up to 56.76°C

and is sent to the solar regenerators. The strong desiccant is cooled to 46.53°C, and

it needs to be cooled further to 32.22°C. Detailed calculations are given in Appendix

B.4.2. The cooling of strong desiccant can be achieved with a cooling water system.

The cooling water system carries water at 0.6 kg/s mass flow rate. The inlet

temperature of the water is 28.70°C and the outlet is 36.50°C, as shown in Fig. 4.4.

A cooling tower is used to cool the warm water. Detailed calculations are given in

Appendix B.4.3.

4.5.7 Auxiliary Heater

An auxiliary heater is used during the time in which the performance of the solar

regenerator is lower than given in Appendix C. It is assumed that the regenerator

only performs at the highest operating conditions for four hours every day and at

half of the highest performance for another four hours. During the rest of the day

the contribution of sun is negligible. Then, the auxiliary heater should heat the

desiccant from 56.76°C to 60.00°C which is in heat terms, 4.59 k'V. The average

energy requirement for the auxiliary heater is 3.44 k W.


From the above analysis, it is seen that the proposed hybrid desiccant cooling

system, a combination of liquid desiccant dehumidifier with a vapor-compression

unit, will provide comfort conditions in hot, humid climates. Only by adding the

71

dehumidification cycle, the existing system can be converted to a hybrid cooling

system. For the new supermarket buildings, this system promises additional sav

ings with the lower air flow rate (e.g., smaller blower and duct size). The solar

regenerator is simple in construction and in operation. The additional cost of the

system on top of the conventional vapor-compression unit is given in Table 4.1. The

references for the price list are given in Appendix E.

4. 7 Economic Analysis

The econorruc analysis of the system will include the energy savmgs, yearly

money savings, and the payback period of the proposed system.

4. 7.1 Targeted Energy Savings

According to the Public Utility Commission of Texas, Electric Division, "End

Use Modeling Project Interim Report", through the years 1990-1999, the average

yearly commercial end-use electricity consumption for air-conditioning is 18,048

GWH (Gega Watt-Hours). The total energy consumption through these years is

180,480 GWH/year. It is assumed 40% of energy will be used in supermarkets (Pub.

Ut. Com., 1990). By using ERAP conversion factors, this is equal to 8.37 x 1014

Btu's targeted for energy saving.

4.7.2 Savings

The saving that are introduced by the new system in energy and in money will

be calculated. The energy saving factor, yearly saving of money, and payback period

of the proposed system will be calculated. The savings factor is the percent energy

savings of the proposed system over the conventional system. Energy required for

the proposed system

Energy required to run the pumps - Energyp

Energy required to run the fans - EnergyF

Energy required to run the vapor- compression unit EnergyE

72

Energy required for auxiliary heater Energyauxil.heater

Energy = Energyp + EnergyF + Energyauxil.heater + EnergYE·

The overall energy needed for the system will be calculated from the energy

requirements for each part. The energy requirement for the conventional unit will

be calculated from the energy required for the state changes of air.

4.7.2.1 Pumps

In the system there are three pumps-two for the desiccant and one for water

circulation. One of the desiccant pumps is placed on the strong desiccant line, and

the energy required to run this pump is 0.01768 kW. The other desiccant pump is on

weak desiccant line and the energy required for this pump is 0.0185 kW. The \vater

pump which is on the cooling water line requires 0.0138kW. Detailed calculations

are given in Appendix B.6.

4. 7.2.2 Fans

In the system, there are two fans-one is a blower to circulate the air corning

from the building and the other is a fan for the cooling tower. The energy required

for the blower is 15.9 kW. The energy required for the cooling tower fan is 0.0062

kW. The detailed calculations are given in Appendix B.7.

73

4. 7.2.3 Vapor-Compression System

The energy required for the compressor can be calculated from the assumption

that the condenser is working at 45°C and the evaporator at 18.0°C. The energy

required for the unit is 3.46 kW. Detailed calculations are given in Appendix B.S.

4.7.2.4 The Conventional Air Conditioning Unit

Conventional air-conditioner operation can be investigated in two steps. The

first step is cooling the air to the saturation state and condensing the excess humid

ity. The second step is warming up the air to a comfortable condition. The energy

required for the unit is 44.32 kW. The energy required for the blower is 15.9 kW.

The detailed calculations are in given Appendix B.9.

4.7.2.5 Comparison of Energy Requirements

Energy needed to run conventional system

Energyo = 60.22 kJ/s = 60.22 kW

Energy needed for the proposed system

Energyn = 22.85 kJ /s = 22.85 kW

Energy saving = 60.22 - 22.85 = 37.37 kW.

( 4.1)

( 4.2)

The savings factor, SF, is the percentage of the energy that will be saved with

the installation of the proposed system. The proposed system's saving factor is

62%. The yearly savings are $5712 and the payback period for the unit is 18.75

months. Detailed calculations are given in Appendix B.10.

Table 4.1: Additional cost of the proposed hybrid cooling system on conventional unit.

I Item I Unit I Unit Cost I Total Cost $ + Tax j Solar Panel 2952.57 2952.57 Pumps; Desiccant 2 108 216

Water 1 108 108 P.V.C. pipe 338 338 Dehumidification Tower 1 77 77 Cooling Tower Fan 1 48.08 48.08 Cooling Tower Frame 1 50 50 Droplet Separator 1 75 75 Diffuser, Packing Supporter, Packing holder, Tower table 150 150 Heat exchanger

Desiccant-Desiccant 1 600 600 Desiccant-Water 1 550 550 Desiccant-Air 1 350 350

Desiccant Tanks 1 113 113 Desiccant 1363 1363 Packing Material 80 80 Auxiliary Heater 190.77 190.77 Miscellaneous 300 300 Labor to Manufacture (54 hr) $ 8 per hr 432 Labor to Assemble 150 150

Total Retail Cost 8143.42 From Retail to Wholesale Conversion 20% -1626.68 Wholesale Conversion Total Cost 6514.73 Assuming 15% Profit +977.21

I Tax +456.03

I Retail Sale Price 7947.97

74

States:

1- Air G'= 6.95 kg/s t = 240C w = 0.0102441eg water/kg air

2-Air G'= 7.645 kg/s t = 24.70C w = 0.01126 leg water/kg air

3-Air G'=7.625 legis t=27.2620C

w = 0.0086 leg water/kg air 4-Air

G'=7.625 legis t = 230C

w = 0.0086 leg water/leg air

5- Desiccant L' =0.5258 legis t= 32.220C ~=40%

6-Desiccant L' = 0.5462 kgls t = 41.6380C

X= 38.5%

7-Desiccant L' =0.5462 legis t = 43.SOC ~=38.5%

8- Desiccant L' =0.54621egls t = 56.760C ~= 38.5%

9- Desiccant L' =0.5258 legis t = 600C ~=40%

10-Desiccant L' = 0.5258 legis t = 46.530C ~=40%

11-Cooling Water L'=0.6q/s t =28.70C

12-Cooling Water L' = 0.6 leg/s t= 36.50C

13-Ambient air G' =0.695 t= 32.220C

{;)

Solar Regenerator

Refrigerant

Air

- Desiccant Water

w = 0.02142 kg water/kg air

fiuure 4.1 Schematic of the proposed cooling system. 0

60%

23.00 24.00 24.70

Temperature, OC

76

.. ·;::; '-

Cl)

~ tJ

3 11.260 <;;

·c; E '-

10.244 0

E Cl)

~ 8.600

:.0 ·e ::I

:X: u

;..=

t en

27.26 32.22 a-b, Vapor-Compression System

b-e, Room

c-d-e, Mixing with Ambient Air

d-a, Dehumidifier

Figure 4.2 Processes on the ps~rchometric chart (not to scale).

Waste Heat Desiccant-desiccant from heatexchanger

the Condenser

~~28~~siccant I• M1 :tW Desiccant-air 43.SOOC

heat exchanger Weak Desiccant

60°C Strong Desiccant

Figure 4.3 Air-desiccant and desiccant-desiccant heat exchangers.

I I

32.22°C Strong Desiccant

Water-desiccant heat exchanger

46.530C Strong Desiccant

Figure 4.4 Desiccant-\'l.·ater heat exchanger.

78

CHAPTER 5

CONCLUSIONS AND RECOMMENDATIONS

In this project, dehumidification of air using liquid desiccants has been carried

out. The heat and mass transfer coefficients are calculated for the packed column

using CaCl2 , LiCl, and a newly introduced liquid desiccant CELD solution for four

different sizes of ceramic raschig rings. The performance of the CELD solution in a

packed column filled with plastic snowflake packing, which was a new packing, was

also evaluated. The performance of the CELD solution in a packed column filled

with 1.5 in. ceramic raschig rings was also calculated and the performance of the

two was compared. A hybrid cooling system, which uses a CELD solution and a

packed column with snowflake packing to absorb the latent load of air is modelled.

Regeneration is accomplished in a semi open solar regenerator. A basic economic

analysis of the model was performed.

The following conclusions were drawn from this study:

1. CELD is a 50% CaCl2 and 50% LiCl mixture, and its performance

is above the pure CaCl2 solutions and below the LiCl. However the

behavior is not related to the concentration linearly.

2. CELD is a promising, inexpensive, and stable liquid desiccant for

dehumidification and hybrid cooling systems.

79

3. Heat and mass transfer coefficients for the CELD-air contact system

using snowflake packing are calculated. The gas phase coefficients

are strong functions of gas phase mass velocity, and liquid phase

coefficients are strong functions of both liquid and gas phase mass

velocities.

4. The effect of liquid and gas temperatures and liquid concentration on

the heat and mass transfer coefficients are comparably smaller than

the liquid and gas flow rates, for the range in which the coefficients

are evaluated.

5. The performance of the CELD solution in the packed tower filled with

snowflake packing material showed that a 40% CELD solution in a

0.5 m packing height will create 11.9 x 10-3 kg of water I kg of air

change in the humidity. For the same flow rate, 1.0 m packing height

will create 16.6 x 10-3 kg of water I kg of air change in the humidity.

For optimum operation, it may be desirable to the increase packing

height to 1.0 m. However increasing the height above 1.0 m will not

create any significant change in the outlet humidity of air. It should

be kept in mind that these values are liable to change with the change

in the air and liquid flow rates.

6. The calculations showed that for the same level of dehumidification

of air, plastic snowflake packings require a higher tower than 1.5 in.

ceramic raschig rings. However, snowflake packing is found to be

superior because of its low price and weight, lower pressure drops

and ease of installment.

80

7. After observing the stable performance of CELD solutions, a hybrid

cooling system is modelled for a supermarket in a hot and humid cli

mate which uses a snowflake packing material with the CELD solution

as a desiccant in a column. Additional dehumidification and regen

eration cycles are proven to introduce savings over the conventional

units up to 62% with a payback period of less than 2 years. This

shows that liquid desiccant hybrid cooling systems are cost-effective

alternatives for conventional air conditioning units.

The following are recommendations for future studies in this area:

1. The mass and heat transfer coefficients that are calculated in this

study can be compared with the experimental evaluations. The per

formance of the tower can also be compared with the experimental

results.

2. The properties of snowflake packing material, v.-hich is available in

this study, can be reorganized and evaluated.

3. The liquid desiccant hybrid cooling system model can be restudied

for different air and liquid temperatures to optimize the efficiency of

the system.

4. The solar regenerator can be replaced by a packed regeneration col

umn where the heat to the system is obtained from a natural gas

heater which heats the liquid. Even though this system may have

lower energy savings as the heat for evaporation is obtained from the

heater, it will increase the stability of the regeneration cycle.

81

5. Similar liquid desiccant hybrid cooling units can be developed for

residential or smaller size commercial business uses (e.g., fast food

restaurants)

6. An experimental setup for a liquid desiccant hybrid cooling system

can be developed and the performance of the system can be analyzed.

Similar studies have been conducted using solid desiccants. An ex

perimental setup can bring more attention to liquid desiccant hybrid

cooling systems.

7. Similar models can be developed for different areas in which liquid

desiccant dehumidification systems may be used (e.g., low tempera

ture batch drying operations, etc.)

8. The noncorrosive materials are recommended to use for hybrid cooling

system, specially for heat exchangers. However, this will increase the

payback period.

82

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Mullick, S. C. and Gupta, M. C., 1974, "Solar Desorption of Absorbent Solution," Solar Energy, vol. 16, pp. 19-24.

Norton Company, 1987, Intalox High Performance Snowflake Packing, 2M-132200302-8/88 Bulletin ISPP-1R.

Peng, C. S. P. and Howell, J. R., 1981, "Analysis and Design of Efficient Absorbers for Low Temperature Desiccant Air Conditioning," Transaction of ASME Journal of Solar Energy Engineering, vol. 103, pp. 67-74.

Peng, C. S. P. and Howell, J. R., 1984, "The Performance of Various Types of Regenerators for Liquid Desiccants," Transactions of ASME Journal of Solar Energy Engineering, vol. 106, pp. 133-141.

Public Utility Commission of Texas, Electric Division, 1990, End-Use Modeling Project Interim Report.

Queiroz, A. G., et al., 1984, "Performance Analysis of an Air Drier for a Liquid Dehunildifier Solar Air Conditioning System," ASME ~7inter Annual Meeting, 84-WA/Sol-6.

Sherwood, T. K. And Holloway, F. A. 1., 1939, "Performance of Packed TowersExperimental Studies of Absorption And Desorbtion," American Institute of Che:mlcal Engineers, pp. 21-37.

Shulmann, H. 1., Ullrich, C. F., Proulx, A. Z., and Zimmerman, J. 0., 1955, "Performance of Packed Columns, Part I,II, and III," A.I.Ch.E Journal, vol. 1,no. 2, pp. 247-264.

Strigle, Jr. R. F., 1987, Random Packings and Packed Towers, Design and Application, Gulf Publications.

Thuesen, G. J. and Fabrycky, W. J., 1989, Engineering Economy, Prentice-Hall International Series in Industrial and Systems Engineering.

Treybal, R. E., 1969, "Adiabatic Gas Absorption and Stripping in Packed Towers," Industrial and Engineering Chemistry, vol. 61, no. 7, pp. 36-41.

Treybal, R. E., 1980, Mass Transfer Operations, Me Graw-Hill.

86

rllah, M. R., 1986, "Analysis of a Counterflow Indirect Evaporative Cooling and Liquid Desiccant Dehurrildification System,'" a Ph.D. dissertation submitted to Texas A&M University.

'\\:"hitehead, E. R., 1985, "Outdoor Air Treatment for Humidity Control in Supermarket," ASHRAE Transactions, vol. 91, part lb, pp. 434-440.

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APPENDIX A

COMFORT ZONE

I !/ H

r r E lO E

l! . 0 5t ..

f "' " 0

i

\0

Sl•nd•rd Err~cliv~ T~mpenlurt •nd lhr ASHR.-'f. Cnmfnrl Lonn

87

APPENDIX B

CALCULATIONS

B.1 Mixing of Air

The humidity ratio is calculated as

WA-ambient - WM 10 WAf- WR 1

0.02142- WM 10 WM - 0.010244 1 '

and

WM = 0.01126 kg of water/kg of air.

Temperature of the mixture is calculated from the enthalpy equation

H = Cp X t + w X HE.

The ambient air enthalpy

HA-ambient 1.005 X 32.22 + 0.02142 X 2560.26

HA-ambient 87.22 kJ /kg.

The room air enthalpy

H R - 1.005 X 24.00 + 0.010244 X 2545.37

HR 50.19 kJjkg.

The mixed air enthalpy can be calculated from the ratio

88

(B.1)

(B.2)

H A -ambient - H M

HM-HR 87.22- HM HM- 50.19

The enthalpy of mixed air

H M = 53.55 kJ /kg.

10 1

10

1

89

(B.3)

Hence from the enthalpy equation the temperature of the mixed air is calculated.

53.55 = 1.005 X TM + 0.01126 X 2546.64,

The mixture enters the dehumidifier at 24.7°C and 0.01126 kg of \Yater/kg of

dry air.

B.2 Vapor Compression Unit

Air passes over the coils of the evaporator which carries freon-12 as refrigerant.

Air enters the evaporator at

w 0.0086 kg water /kg air,

m 7.6246 kg/s,

90

and leaves at 23°C. The heat flow is

(B.4)

Q - 7.6246 X 1.005 X 4.262 = 32.65 kW.

Assuming the evaporator temperature is 18.0°C and the condenser temperature

as 45.00°C, heat absorbed by the evaporator is calculated as

(B.5)

qL 194.842-79.647 = 115.195 kJjkg.

Mass flow rate of the refrigerant is

m = 32·65 kJ/s = 0.2834 k /s. 115.195 kJ /kg g

The air leaves the evaporator at

T - 23.00°C,

w 0.0086 kg water/kg air,

m 7.625 kg/s.

The heat rejected by condenser can be calculated as

Q in~H,

Q 0.2834 X (207.074- 79.647),

Q 36.11 kW. (B.6)

91

B.3 Solar Regenerator

The amount of water to be removed from the solution can be calculated to obtain

0.5258 kg/s of 40% CELD solution.

Amount of Desiccant

0.4 x 0.5258 = 0.2103 kg/s.

Amount of Water

0.6 x 0.5258 = 0.3155 kg/s.

The amount of desiccant in the 38.5% CELD solution is

X= 0.2103 0.385

X = 0.5462 kg/ s.

Amount of water entering into the solar regenerator = 0.5462-0.2103=0.3359 kg/s.

Amount of water to be removed =0.3359-0.3155=0.0204 kg/s or 73.44 kg/hr.

From the solar regenerator calculations (Appendix C), the surface area needed

to remove 73.44 kg/hr water is

73.44 - 9 37 2 - 5. m. 0.77

B.4 Heat Exchangers

B.4.1 Air-Desiccant Heat Exchanger

92

The heat rejected by the condenser is 36.11 k\V. Assuming that air will pass

over the condenser and warm up to 45°C, this air will be used to warm up the weak

desiccant. The amount of air that will be needed can be calculated as

Q - (mC,D.t)ai,.

36.11 - mai,. X 1.005 X ( 45.00 - 32.22)

mA - 2.81 kg/s.

This air will be used to warm up the desiccant at 41.638°C, by assuming an

effectiveness of 0.65. (m x C,)ai,. > (m x C,)D the effectiveness equation is (Kreith,

1965)

E -

0.65

Tout-A

(m x C,)A x (Tin-A- Tout-A) (m X C,)D X (Tin-A- Tin-D) (2.81 X 1.005) X ( 45.00- Tout-A)

(0.5462 X 2.594) X ( 45.00- 41.638)

43.90°C.

The desiccant outlet temperatures can be calculated from the heat balance.

( mC,D.t)D - ( mC,D.t)ai,.

0.5462 X 2.594 X (Tout- 41.638) 2.81 X 1.005 X ( 45.00- 43.90).

(B.7)

The weak desiccant leaves the heat exchanger at 43.8°C. The recommended

design for the heat exchanger is a 36 in. diameter, 2-foot high cylinder where the

air is blown through the coils of the desiccant from the bottom of the cylinder.

93

The heat exchanger uses the blower of the condenser. Further, the cooler weak

desiccant and warmer strong desiccant transfers heat in a desiccant to desiccant

heat exchanger.

B.4.2 Desiccant-Desiccant Heat Exchanger

The strong desiccant enters the heat exchanger at 60°C and the weak desiccant

enters at 43.80°C. The effectiveness of the heat exchanger is assumed as 0.8. The

flow rates of the desiccant are close to each other and their C'P is assumed to be the

same. Thus the effectiveness of the heat exchanger is

~T --=0.8. l::t.Tma.z

(B.8)

The inlet solution temperature from the heat exchanger to solar regenerator can

be calculated as

T.t,.ong-de• - 43.80 = 0.8 60- 43.80 '

Solution enters the solar regenerator at 56. 76°C. The strong desiccant outlet

temperature can be calculated from the heat balance.

0.5462 X 2.594 X (56.76- 43.80) 0.5258 X 2.594 X (60 - T).

94

T:::46.53°C is the temperature at which the strong desiccant leaves the heat

exchanger. The recommended design is a tube-tube or a shell tube heat exchanger.

The strong desiccant needs to cool down to 32.22°C.

B.4.3 Cooling \Vater System

This system is designed to cool the desiccant from 46.53°C to 32.22°C. The

water flow rate is assumed as 0.6 kg/s and the heat exchanger effectiveness as

0.8. The recommended design is a plate-tube or a shell tube heat exchanger. As

(m x C,)wate .. > (m x C,)D, the heat exchanger effectiveness is (Kreith, 1965)

E ( m X C,)wate .. X (Tin-water - Taut-water)

( m X C,)D X (Tin-D - Tin-water)

q - ( m X C,)D X !lTD :::: 19.51 kJ /s

q EX (m X C,)D X (Tin-D- Tin-water)= 19.51 kJ/s

T. = 28.7°C in-water

q (m X C,)wate,. X !lTwate,. = 19.51 kJjs

(B.9)

The approach of the cooling tower is 1.5, and the range is 7.8. A redwood

counter flow cooling tower is used with spray diffusers and splash filling. The forced

draft is suggested. The air flow rate is 0.03 kg/ s with the air exit temperature of

34°C.

B.5 Auxiliary Heater

The auxiliary heater should heat the desiccant from 56.76°C to 60.00°C. In heat

terms

D.T

m

= 60.oooc- 56.76°C

2.594 kJ kg oc

0.5462 kg/s

Q 4.59 k'W.

Then, taking the average energy requirement of the auxiliary heater

Energyauzil.heater 4.59 X 4 X 0.5 + 4.59 X 16

24

Energyauzil.heater - 3.44 kW.

B.6 Pump Energy Calculations

Energy required to run the pump is

A 1 . 1 Energyp = I..J.p x - x m x -.

p ,

Assuming that the pressure head is around 10 ft. of water column

D.p = 10 ft. of water= 3023.87 kg/m2•

For the weak desiccant pump

m 0.5462 kg/s

p 1246.8 kg/m3•

95

(B.10)

(B.ll)

96

For the strong desiccant pump

m - 0.5258kg/s

p = 1259.0 kg/m3•

Energy needed for the weak desiccant pump

EnergyPDl 1 1

3023.87 X 1246

_8

X 0.5462 X O. 7

EnergyPDl - 1.893 kgm/ s

EnergyPDl 0.0185 kW.

Energy needed for the strong desiccant pump

1 1 EnergypD2 3023.87 x

1259_ x 0.5258 X O.

7

EnergypD2 - 1.804 kgm/s

EnergypD2 - 0.01768 kW.

For the water pump

m - 0.6 kg/s

p 996.9 kg/m3

~p - 1511.93 kg/m2•

Energy needed for the water pump

Energypw 1511.93 X -1

- X 0.6 X -1-

996.9 0.7

Energypw 1.299 kgm/s = 0.0127 kW.

as

97

B. 7 Fan Energy Calculations

Energy required to run the fan that circulates the air in the building is calculated

E . A 1 1 nergyF = m x ~..Jt.P x - x -.

Paif' TJ

m 7.645 kg s

D.p - 0.2 m of water

kg PA - 1.177 3

m

TJ - 0.8.

kgm EnergyF1 = 1618.8 -- = 15.9 kW.

s

Energy required for the fan of cooling tower

m - 0.03 kg s

D.p 0.02 m of water

kg Paif' - 1.177 3

m

TJ - 0.8.

kgm EnergyF2 = 0.635 -- = 0.0062 kW.

s

(B.12)

(B.13)

98

B.8 Vapor Compression Unit Energy Calculations

Energy required for compressor can be calculated from the assumption that the

condenser is working at 45°C and evaporator at 18.0°C. From the freon 12 tables

p - 1.0843 MPa

s - 0.68837 kJ/kg K,

and the state is calculated as

H - 207.074 kJ /kg.

Heat rejected by the condenser

Qed - 36.11 kJ js.

Heat absorbed by the evaporator

Qcooling - mair X !:lH

Qcooling - 32.65 kJ /s.

Energy needed for the compressor IS simply the difference between the heat

rejected at the condenser and heat absorbed at the evaporator.

. . Energycom Qed- Qcooling

Energycom - 3.46 kJ/s = 3.46 kW.

(B.14)

(B.15)

99

B.9 Energy Calculations for the Conventional System

The air that enters the evaporator will be cooled below the dew point temper-

ature. The dew point temperature of the entering air with the inlet humidity of

0.01126 kg of water/ kg of dry air, is 15.9°C. Enthalpy of this state can be read from

the psychrometric chart as 12.85 kcal/kg. The desired level of absolute humidity

is 0.0086 kg of water/ kg of dry air. The corresponding dew point temperature is

11.8°C. Enthalpy of this state is 8.00 kcal/kg. Appendix E shows the state changes

during the operation on the psychrometric chart. The air outlet temperature from

the evaporator is considerably low for comfort conditions (given in Appendix B).

The condenser heat is utilized for the reheat of the air which returns to the building.

The energy absorbed in the condenser can be calculated from the enthalpy changes

in two states, a and b.

The saturation state of the entering air can be found from steam tables.

EnergyEv - 155.13 kW.

(B.16)

(B.17)

The COP of the vapor compression units for high latent load applications (i.e.

supermarkets) in hot and humid climates is between 2.3 to 3.5 (information obtained

from AirTech Company). In this calculation COP will be taken as 3.5. The energy

required to run the compressor is the ratio of the energy absorbed in the evaporator

and COP.

Energycom

Energycom

EnergyEv

COP 44.32 kW.

(B.18)

(B.l9)

100

The energy required to run the fan needs to be added

EnergyF = 15.9 k\,V.

B.10 Saving Calculations

Saving Factor

Savings factor, SF, is the percentage of the energy that will be saved by the

installation of the proposed system.

SF - ( Energyn) 1- X 100 Energyo

(B.20)

SF ( 22.85) 1--- X 100 60.22

SF 62%.

Thus, the new system promises up to 62% energy savings.

Yearly Savings

Assuming that the price of natural gas is $3.84/million Btu (from Energas,

Industrial Prices, Lubbock, TX, 1990), and the price of electricity is $0.04 /kWh

(from Ranking of Electricity Prices), daily saving that the current system will bring

can be calculated. For the proposed system, natural gas is used in the auxiliary

heater. The power needed to run the auxiliary heater for a day

power auzil.heater 4.59 X 4 X 0.5 + 4.59 X 16

power auzil.heater 82.62 kWh/day= 281694.5 Btu/day

Costauzil.heater - $1.08/day. (B.21)

101

Power needed to run pumps, fans, and the vapor-compression unit for the pro

posed system is

Powerothe... - 19.41 X 24

Powerothe... 465.84 kWh/day

Costothe .. • $18.63/day.

The total cost of running the proposed system for a day is $19.71.

For the conventional system the power needed to run the vapor-compression

unit and the fans

Powero

Powero

Casto

60.22 X 24

1445.3 kWh/day

$57.80/day.

Thus, the new system will provide saving = $57.80- $19.71 = $38.08 per day.

Yearly savings can be calculated from the daily savings assuming that the air

conditioning unit will be used from May to September, for five months. The total

saving for a year is

A - 5 months/year x 30 days/month x $38.08 per day

A - $5712 /year.

(B.22)

102

Payback Period

Payback is the period of time required for the cumalitive cash in flows from a

project to equal the initial cash outlay. The interest rate during this period and the

salvage value of the equipment will be taken into consideration. The method used

is called 'Capital Recovery with Return' (Thuesen, 1989). The initial investment,

P, is $7947.97. The salvage value of the pumps, desiccant tank, auxiliary heater and

fans is calculated from 20% of the initial costs. The initial total cost of these pieces

are $ 675.85, and the salvage value, F, is $135.17. Currently, the interest rate, i,

for the commercial loans is 11% for large businesses ( N CNB Texas National Bank).

The payback period can be calculated from the following equation (Thuesen, 1989)

n -ln [ A-ix1P-F)]

ln(1 + i) n - 1.56 years = 18.75 months.

The total payback period is 1.56 years.

(B.23)

APPENDIX C

REGENERATOR COVERED WITH GLASS

BUT BOTH ENDS ARE OPEN

Assume ambient temperature is 32.22°C, desiccant temperature is 60°C, partial

pressure of water vapor in ambient air is 24.8 mm Hg. Assume convection heat

transfer coeff. is 5.25 x 10-3 kW /m2°C, based on Mullick et al. (1974), energy

transfer by convective heat transfer is 0.14595 kW /m2• Assuming h/k based on

Dropkin's measurement, convective mass transfer coef., k, is 19.5 kg/hr m2 (~y).

~pis 45.2 mm Hg, or !::l.y is 0.0393. Mass Transfer is k x /j.y, 0. 77 kg water/hr m2•

103

APPENDIX D

MATERIAL COST

1. \Vater pumps and desiccant pumps are obtained from Grainger cat

alog Chemical-Resistant Magnetic-Drive Centrifugal Pumps, manu

facturers model 3-MD-SC, stock no. 2P038.

2. Cooling tower fan from Grainger catalog 12 in. diameter ring fan

Dayton, stock no. 2C107.

3. Dehumidifying tower 48 in. diameter 90 psi pvc pipe from Diamond

Plastic Co., Lubbock, TX.

4. PVC pipe as ducks 12 in. diameter 50 psi pvc pipe from Diamond

Plastic Co., Lubbock, TX.

5. Liquid desiccant 0.5258 kg/s of solution is equal to 0.21032 kg/s

CELD flow rate. The LiCl is $2.99/lb, from LitCo Co., Gastonia

N. C. and CaCl2 is $0.2766/lb, from Mayfield Paper Co., Lubbock

TX. This is equal to Sl.633/lb or $3.6/kg for CELD. 30 minute supply

of CELD is

CostcELD = 0.21032 x 3.6 x 30 x 60 = $1363

The cost of CELD solution is $1363.

6. Desiccant Tank 200 gallons plastic tank, from Lee Co. Idalou TX.

7. Intalox® snow flake packing from Norton Co., Akron OH.

104

8. Nat ural gas liquid heater from Me MasterCarr Catalog, no. 3598K51.

9. Cost of each m 2 collector is calculated as: From General Steel Co,

Burch Glass Co, Selle Insulation, Lubbock, TX

Description Total Cost

16 Gage Galvanized Steel Sheet 2 x $6.17 per m2

16 Gage Window Glass $17.216 per m2

2" Fiberglass Insulation $1.4 per m2

Total Material Cost $30.956 per m2

Total Cost of Solar panel of 95.37 m 2 $2952.57

10. Heat exchangers approximate prices from Southwestern Refrigaera

tion Co., Lubbock, TX.

•

105

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APPEXDIX E

CO~VENTIONAL SYSTE:..IS PROCESSES

ON THE PSYCHROMETRIC CHART

PHYS1'ROMETRIC qHART rS~OitCUI'Cl Of,...IIJIWI".C It'• I..,,.. riO''"" llfJ

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performance analysis of a new developed a thesis in

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