Hawaii Energy and Environmental Technologies (HEET) Initiative · Hawaii Energy and Environmental...

Hawaii Energy and

Environmental Technologies

(HEET) Initiative

Office of Naval Research

Grant Award Number N0014-11-1-0391

Assessment of Desiccant Dehumidification:

Literature and Technology Review

Prepared by:

Sustainable Design & Consulting LLC

Prepared for:

University of Hawaii at Manoa, Hawaii Natural Energy Institute

July 2016

ASSESSMENT OF DESICCANT DEHUMIDIFICATION

Project Deliverable 2:

Literature and Technology ReviewJuly 9, 2016

FinalFinal

Sustainable Design & Consulting LLC

Prepared by:

Manfred J. Zapka, PhD, PE

James Maskrey, MEP, MBA

Prepared for:

Hawaii Natural Energy institute

RCUH P.O. #Z10117197

ASSESSMENT OF DESICCANT DEHUMIDIFICATION

Project Deliverable No. 2:

Literature and Technology Review

FINAL

Prepared for

Hawaii Natural Energy Institute

RCUH P.O. #Z10117197

July 9, 2016

Prepared by:

Manfred J. Zapka, PhD, PE (1)

James Maskrey, MEP, MBA, Project Manager (2)

(1) Sustainable Design & Consulting LLC, Honolulu, Hawaii (2) Hawaii Natural Energy Institute Honolulu, Hawaii

ACKNOWLEDGEMENTS

This research was supported and funded by the Office of Naval Research grant award N00014‐11‐0391

in collaboration with the Hawaii Natural Energy Institute at the University of Hawaii at Manoa.

The authors believe that desiccant cooling applications can play a significant part making building

conditioning more energy efficient and foster the implementation of more environmentally friendly

ways to provide occupant comfort in buildings.

ABBREVIATIONS

AC Air‐conditioning

ACHR Air Conditioning, Heating and Refrigeration NEWS

AHU Air‐handling Unit

ASHRAE American Society of Heating, Refrigerating, and Air‐Conditioning Engineers

BET Brunauer–Emmett–Teller

CHP Combined Heat and Power

CA conditioned air (also CA’ after treatment)

COP Coefficient of performance

DP Dew point

DX Direct Expansion Air Conditioning System

DEAC Direct Evaporative Air Coolers

DEC Direct evaporative cooler

EER Energy Efficiency Ratio Engelard Titanium Silicate (ETS)

EAC Evaporative Air Cooling external source of energy (Q)

GTI Gas Technology Institute

GCI Green Cooling Initiative

HNEI Hawaii Natural Energy Institute

HR Humidity Ratio

HVAC Heating Ventilation Air Conditioning

IC Internal Combustion

IEAC Indirect Evaporative Air Coolers

LDDX Liquid Desiccant DX System

IDECOAS Indirect and Direct Evaporative‐Cooling‐Assisted Outdoor Air System

IRR Internal Rate of Return

LD Liquid Desiccant

LDAC Liquid‐Desiccant Air Conditioner

LD‐IDECOAS Liquid Desiccant Indirect and Direct Evaporative‐Cooling‐Assisted Outdoor Air System

IEC Indirect Evaporative Cooler

LiCl Lithium Chloride

LCST Lower Critical Solution Temperature

M‐cycle Maisotsenko (M) ‐Cycle

MIT Massachusetts Institute of Technology

ABBREVIATIONS (CONT.)

O&M Operations & Maintenance

OA Outdoor (or outside) Air (also OA’ after treatment)

pH. Numeric Scale Used to Specify the Acidity or Basicity (alkalinity) of an Aqueous Solution

PV Photovoltaic System, also solar‐PV Power System

RH Relative Humidity

SHR Sensible Heat Ratio

R Separation Factor

SAC Solar air‐conditioning

SDC Sustainable Design & Consulting LLC

SI International System of Units

TAT Thermally Activated Technologies

TSA Temperature Swing Adsorption

VAC Vapor‐Compression Air Conditioning

WBT Wet‐bulb temperature

VAV Variable Air Volume

DEVAP Desiccant Enhanced Evaporative Air Conditioning

UNITS

atm Atmosphere pressure

BTU British Thermal Unit

BTUH/sqft BTU per square feet

°C Degree Celsius

CFM Cubic feet per minute

F Fahrenheit

K Kelvin

kW kilo Watt

kJ kIlo Joule

MPa Mega Pascale

square feet Square Feet

FINAL Assessment of Desiccant Dehumidification – Literature and Technology Review

TABLE OF CONTENTS

RCUH P.O. #Z10117197 Project Deliverable No. 2.: Literature and Technology Review Hawaii Natural Energy Institute Sustainable Design & Consulting LLC

July 9, 2016 Page i

TABLE OF CONTENTS

SECTION 1 ‐ EXECUTIVE SUMMARY AND OVERALL CONCLUSIONS ............................................................ 1

SECTION 2 ‐ INTRODUCTION OF DESICCANTS IN REFRIGERATION AND COOLING .................................... 3

SECTION 3 ‐ BASICS OF THE DEHUMIDIFICATION PROCESS ........................................................................ 6

3.1 Methods of Dehumidification ............................................................................................................ 6

3.2 Review of Psychrometrics ‐ Principles ................................................................................................ 8

3.3 Application of Psychrometric Chart ................................................................................................. 15

3.4 Introduction to Desiccants ............................................................................................................... 21

3.5 Desiccant Cycle ................................................................................................................................. 22

3.6 Desiccant Material ........................................................................................................................... 24

SECTION 4 ‐ METHODS OF DEHUMIDIFICATION ...................................................................................... 29

4.1 Cooling‐based Dehumidification ...................................................................................................... 29

4.2 Desiccant Dehumidification ............................................................................................................. 32

4.3 Comparing Desiccant to Cooling‐based Humidification ................................................................... 41

SECTION 5 ‐ PHYSICAL PROPERTIES OF DESICCANTS ................................................................................ 44

5.1 Basic Considerations of Physical Properties in Desiccant Dehumidification ................................... 44

5.2 Isotherms ......................................................................................................................................... 45

5.3 Types of Isotherm Models ............................................................................................................... 48

5.4 Adsorption Energy ............................................................................................................................ 54

5.5 Chemical and Physical Stability ........................................................................................................ 55

5.6 Adsorption Rate and Performance.................................................................................................. 55

5.7 Basic Heat and Mass Transfer Modelling by Example of Desiccant Wheel ..................................... 59

5.8 Heat and Mass Transfer Characteristics of Desiccant Polymers ...................................................... 66

SECTION 6 ‐ DESICCANT REACTIVATION ENERGY SOURCES ..................................................................... 75

6.1 Desiccant Regeneration with Heat from Natural Gas or other Fossil Fuels ..................................... 75

6.2 Desiccant Regeneration with Waste Heat ....................................................................................... 78

6.3 Desiccant Regeneration with Solar Heat and other Renewable Heat Sources ................................ 83


TABLE OF CONTENTS

RCUH P.O. #Z10117197 Project Deliverable No. 2.: Literature and Technology Review Hawaii Natural Energy Institute Sustainable Design & Consulting LLC

July 9, 2016 Page ii

TABLE OF CONTENTS (CONT.)

SECTION 7 – ALTERNATIVE COOLING SYSTEMS INTEGRATED WITH DESICCANT DEHUMIDIFICATION .. 87

7.1 Thermal Cooling Technologies ......................................................................................................... 89

7.2 Evaporative Cooling – General Principle .......................................................................................... 90

7.3 Direct Evaporative Cooler – Working Process ................................................................................. 93

7.4 Indirect Evaporative Cooler – Working Process .............................................................................. 94

7.5 Maisotsenko (M) ‐Cycle Enhanced Indirect Evaporative Cooling ................................................... 97

7.6 Absorption Chillers ........................................................................................................................... 99

7.7 Adsorption Chillers ........................................................................................................................ 106

7.8 Comparison between Absorption and Adsorption Cooling Technology ....................................... 111

7.9 Magnetic Refrigeration as an Alternative to Vapor Compression Refrigeration .......................... 112

SECTION 8 ‐ SOLID DESICCANT COOLING SYSTEMS ................................................................................ 116

8.1 Solid Desiccant Cooling Systems using Different Cooling Devices ............................................... 116

8.2 Solid Desiccant Cooling with Evaporative Cooling ....................................................................... 124

8.3 Solid Desiccant Cooling to Increase Conventional AC System ...................................................... 129

8.4 Solid Desiccant Colling with Solar Heat ......................................................................................... 135

SECTION 9 ‐ LIQUID DESICCANT COOLING SYSTEMS ............................................................................... 139

9.1 General System Configuration of Liquid Desiccant Cooling Systems ............................................ 139

9.2 Liquid Desiccant Systems to Enhance Conventional AC System ................................................... 144

9.3 Liquid Desiccant Cooling with Evaporative Cooling and Solar Energy .......................................... 149

9.4 Desiccant Enhanced Evaporative Air‐Conditioning ........................................................................ 158

SECTION 10 ‐ GENERAL FINDINGS OF THE LITERATURE REVIEW ............................................................ 164

REVIEWED LITERATURE ............................................................................................................................ 167


SECTION 1 ‐ EXECUTIVE SUMMARY AND OVERALL CONCLUSIONS

RCUH P.O. #Z10117197 Project Deliverable No.2.: Literature and Technology Review Hawaii Natural Energy Institute Sustainable Design & Consulting LLC

July 9, 2016 Page 1 of 171


The present literature and technology review is the deliverable of Part 1 of the project “Assessment of

Desiccant Dehumidification” which Sustainable Design & Consulting (SDC), LLC is performing for Hawaii

Natural Energy Institute (HNEI). Part two of the project follows the literature review and will assess the

feasibility of using candidate desiccant dehumidification systems in conjunction with building

conditioning in the hot and humid sub‐tropical climate of Hawaii

Under Part 1 of the project, an extensive literature review has been performed on the subject of

desiccant dehumidification with an emphasis on cooling and dehumidification applications for buildings.

A further emphasis was placed on building and space conditioning in humid climates, such as is found

throughout the Hawaiian islands.

Since the literature review addresses many technical and applied scientific papers and publications

discussing intricate thermodynamic concepts, the literature review also presents some important

thermodynamic concepts to facilitate understanding by the non‐technical or the non‐HVAC expert

reader.

This literature and technology review presents a succession of dehumidification technology

developments over the past 20 to 30 years. Desiccant dehumidification is being extensively used in

specific industrial and commercial applications where dry air is required below the dew point that could

be achieved with cooling‐coil based dehumidification.

This literature and technology review focuses on the use of desiccant dehumidification technology in

conjunction with building conditioning. To date these desiccant applications have mostly used solid

desiccant material installed in vessel‐swing operation or rotary wheel structures. The use of desiccants

in conventional AC has been mostly concerned with energy savings by recovering sensible and latent

loads in central AC‐systems. Such desiccant applications have become standard applications for

conventional AC‐systems. Liquid desiccant dehumidification in conjunction with conventional has only

received limited attention in the literature.

The use of desiccants in conjunction with evaporative cooling, often referred to in the literature as

desiccant cooling, was introduced about two decades ago and has not seen a wide‐spread use in

building designs and operations, although such systems can generate significant energy savings. Both

solid and liquid desiccant systems are suitable for integration with evaporative cooling applications.

While solid desiccant dehumidification integrations have the advantage of longer track record in actual

building installations, it is liquid desiccants which have attracted much attention in recent developments

of desiccant cooling operations. In fact, since 2010 several groundbreaking liquid desiccant cooling

technologies have been proposed. These recent liquid desiccant technologies developments suggest

significant application potential for Hawaii, because of their favorable process performance even in




July 9, 2016 Page 2 of 171

Hawaii’s humid climate and their ability to work with low‐grade, unconventional heat sources, including

solar heat and waste‐heat from combined heat and power applications.


SECTION 2 ‐ INTRODUCTION OF DESICCANTS IN REFRIGERATION AND COOLING


July 9, 2016 Page 3 of 171


Over the past decade, the worldwide demand for residential and commercial cooling has been growing

at a strong pace. Most of the near‐term and long‐term growth will come from developing countries and

countries in transition.

Industry related assessments for the short‐term (five year) development of the world‐wide refrigeration

and air‐conditioning market (ACHR, 2016) indicate an overall annual growth of 6%. Under this

projection, the biggest growth is projected with eight percent in both the Americas and Asia and two

percent in Europe and the Middle East, India & Africa. Asia‐Pacific is reported as the largest market in

absolute sales, with China and Japan representing 82 percent of that market by value.

A long‐term prediction of the global market by the Green Cooling Initiative (GCI) predicts that between

2010 and 2030 the market volume of the refrigeration and air conditioning sectors for these countries is

projected to increase by factor of nearly four (GCI, 2016).

While the demand for cooling is increasing, conventional cooling technology causes significant

environmental impact mainly due to harmful refrigerants and high energy consumption. This fact

coupled with the market trend suggests that impacts from cooling applications will likely increase in the

years to come. The logical method for impact mitigation would be to reduce the two main causes by

significant reduction in energy demand and release of harmful process agents, meaning harmful

refrigerants. Another possible mitigation measure is to develop and implement environmentally friendly

sustainable cooling technologies as a way to mitigate negative impacts.

GCI promotes the use of “Green cooling”, an overall strategy that leads to impact mitigation for the

cooling sectors (GCI, 2016). Green cooling uses a combination of three approaches, which include

maximizing energy efficiency, encouraging the use of natural refrigerants, and fostering a sustainable

approach to private and commercial energy consumption. Green cooling helps to protect the

environment, resources and the climate and supports the use of renewable technologies within cooling.

It thereby contributes towards a sustainable reduction of fossil fuel consumption.

The Heating Ventilation and Air‐Conditioning (HVAC) industry has acted upon the need for impact

mitigation with the introduction of various technology innovations. Innovations include, but are not

limited to, variable capacity compressors, system improvements in humidity control, high efficiency fans

and building construction improvements (Kammers,2010). Among these innovations, improved humidity

control provides significant benefits, in particular for the residential and commercial building sector.

Innovative humidity control extends beyond the conventional humidity management technology of

providing cold surfaces for below dew point separation of humidity in conditioned spaces. This

innovative humidity control separates the processes of heat rejection and humidity control in space

conditioning applications, which means the removal of sensible and latent loads, respectively. In this




July 9, 2016 Page 4 of 171

application, desiccant humidification replaces in of cold‐coil dehumidification. Using conventional HVAC

technology, three space condition functions of ventilation, sensible heat and latent heat removal are

performed concurrently, by means of air flow forced over external cooled coils. This conventional

process invariably causes energy losses and frequent over cooling of spaces, as excess cooling must be

provided to remove humidity, especially in hot and humid climates. Desiccant dehumidification avoids

some of the drawbacks by providing a means to dehumidify spaces without the need of cold coils for

condensation. As fundamental shift in space condition, desiccant dehumidification allows for a

separation of sensible and latent heat removal in space conditioning, as desiccants remove humidity,

and therefore the latent load, and a separate cooling technology removes the sensible load.

Desiccant dehumidification is not a new technology used in the refrigeration and cooling industry. In

fact, numerous industrial and commercial processes rely on low humidity process conditions, which can

be achieved more economically with desiccants rather than with cooling‐based dehumidification. In the

building industry, desiccant dehumidification has seen a limited range of applications. The reason for the

limited acceptance of desiccant dehumidification can be attributed to initial installation costs, the fact

that operational benefits are not fully understood, lack of technology awareness, and company

priorities, which are not focused on benefits of new technology (ECW, 2000). The main benefits

reported by the desiccant industry are the ability for precise humidity control and for the utilization of

low grade heat sources, such as solar and waste heat.

One of the other main reasons why desiccant systems have not been used to a great extent in space

conditioning is the fact that in high‐dew point applications, vapor compression systems are usually more

energy efficient than desiccant systems. The typical dew point in space conditioning, e.g. air‐

conditioning, is about 50 F, for a standard conditioned space temperature and relative humidity of 72 F

and 50%, respectively. Therefore, dew point reduction applications in buildings favor cooling‐based

dehumidification, based on cost advantages and technology familiarity of conventional vapor

compression cooling technology by building management.

When efficiency gains through building system integration and the building industry movement towards

“green” building technologies become more important considerations, benefits for the use of desiccant

systems can outweigh reasons to stick with the conventional HVAC. Apart from the key advantage of

independent sensible and latent heat removal, the use of waste or solar heat for desiccant regeneration

can lead to very favorable operational conditions. In fact, desiccant systems can compete if large

amounts of outside air must be processed, if electric demand charges are high, or if large amounts of

low costs solar and waste heat are available.

The use of low cost solar energy and waste heat from industrial processes for regeneration of desiccant

material is making the dehumidification system more cost‐effective. The initial cost of solar energy can

be significant, but life cycle costs typically contribute to a favorable present value analysis. Solar




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radiation is intermittent and weather‐dependent. Therefore, back‐ up heat sources or heat storage is

required to continue the drying process when solar energy is not available.

Besides desiccant dehumidification as a means to augment conventional cooling technology, another

technology approach is the integration of desiccant dehumidification with evaporative cooling. This

combination offers significant energy saving and environmental impact mitigating potential. Evaporative

cooling is the adiabatic increase in absolute humidity with a simultaneous reduction in dry bulb

temperature. Here, air is subjected to free water surfaces which enhances evaporation while keeping

the enthalpy of the air constant. Typically, evaporative cooling is only effective at high temperatures and

low humidity levels, since the driving forces for evaporation are reduced in humid climates. Since the

Hawaii climate is the representative environmental condition for this literature review, stand‐alone

evaporative cooling would not be considered as applicable. Desiccant dehumidification can, however,

favorably enhance conditions for evaporative cooling by reducing the dew point of the external air

upstream of the evaporative cooler.

Apart from the purely technical considerations, the success of solar/desiccant cooling technologies will

depend on the encouragement and promotional schemes offered by the policymakers, and the efforts

undertaken by the manufacturers to improve the cost efficiency as well in developing better

technologies.


SECTION 3 ‐ BASICS OF DEHUMIDIFICATION PROCESS


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SECTION 3 ‐ BASICS OF THE DEHUMIDIFICATION PROCESS

This section introduces basic processes of dehumidification. Dehumidification is an essential process

with a wide range of industrial and commercial applications, most of them with specific

dehumidification technologies. The focus of this literature and technology review is, however, on

dehumidification in conjunction with building conditioning. Therefore, only processes and technologies

which are pertinent for space conditioning are discussed in more detail.

3.1 Methods of Dehumidification

There are two basic methods of dehumidification, cooling‐based and desiccant dehumidification. Apart

from these there are more specialized niche methods, which use pressure and electric field properties.

These dehumidification technologies have, however, only very limited application potential in building

conditioning and are not further discussed in this review

Cooling‐based Dehumidification

The method of dehumidification in conventional air‐conditioning systems uses cooling coils which are

maintained at a temperature below the desired dew point of the supply air. As air flows past the cooling

coils the air temperatures drop and water vapor condenses on the coils. Lowering air temperature

decreases the air’s ability to hold moisture. Thus, the air can be made drier by cooling it.

The water condensate drains into a pan by means of gravity and is thereby removed from the system.

The amount of moisture removed from the air is a function of the temperature of the coil. The coils are

cooled by chilled water, refrigerants, glycol solutions, or engineered fluids circulating through the coil.

Since this method reduces the temperature of the HVAC‐system supply air, spaces could be subcooled if

the rooms do not have sufficient internal heat gain to offset the low‐temperature air. Typically, the

supply air is therefore reheated downstream of the cooling coils to provide the desired indoor air supply

temperature. Reheating the supply air requires a significant amount of energy, depending on the target

room temperature. In applications that require low air temperature for dehumidification purposes,

while the requirement for sensible heat removal would not require such low air supply temperatures

overcooling can be a problem. In these applications a heat energy recovery strategy could be effectively

used to lower the energy input for reheat the air downstream and for pre‐cooling the air up‐stream of

the cooling point. A suitable strategy would be a heat‐pipe run‐around system, where coils on either

side of the cooling coil and use a refrigerant to economically pre‐cool and reheat the air stream.

Another limitation to this technique is the freezing point of water. When air is dried via refrigeration, the

cooling surfaces of the coils may attain sub‐freezing temperatures. Under these conditions ice can form

on the coils, which reduces the efficiency of the cooling system.




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Desiccant Dehumidification

Desiccant dehumidification provides an effective method for dehumidification through the use of

sorption of water vapor to and into a material that has a natural affinity to water. Desiccant

dehumidification uses adsorption and absorption, which use physical processes to take up the additional

moisture given up by the air without changing the size or shape of desiccants. Absorbents take up

humidity by allowing water vapor to enter into the desiccant material. Adsorbent materials, on the

other hand, attract and hold water molecules in pores at their surface. Absorbents generally can attract

and hold greater quantities of water per pound of desiccant material.

The desiccant based dehumidification does not require elaborate cooling equipment as used in the

cooling‐based dehumidification process. In typical technical dehumidification application, the desiccant

agent takes up moisture until an equilibrium state is reached between the desiccant material and the

air. At this point the desiccant is regenerated, typically through thermal processes, after which the

desiccant is again available for the dehumidification process.

There are two technical processes under which desiccant dehumidification is carried out, liquid and solid

desiccant systems.

Liquid‐desiccant systems: In liquid desiccant dehumidification, a liquid desiccant solution is brought into

contact with the air stream and water vapor is entering the desiccant solution through a difference

in vapor pressure. The method of providing free surfaces of liquid desiccant is dependent on the

design of the process apparatus, which can be based on spray, packed column or bubble bed. When

an unsaturated desiccant solution is exposed to air, the desiccant absorbs moisture from the

gaseous phase. The process conditions can be varied by changing the concentration of the liquid

desiccant, pressure and temperature. Upon saturation, which means a process threshold in the

diminishing ability to take up more moisture, the liquid desiccant is regenerated by heating it and

subsequent desorption of captured moisture to a waste air stream. An additional benefit of using

selected liquid desiccant systems is that the desiccant solutions can act as a biocide for the

conditioned air, which is beneficial in applications for which bacteria or viruses are the least

desirable.

Solid‐desiccant systems: In solid desiccant dehumidification, a solid desiccant material is brought into

contact with the air stream and water vapor attaches to the surface of the desiccant material. Moist

air is drawn through the spaces around desiccant material, which absorbs the moisture. As the

desiccant reaches its capacity, the material is moved into a warmer reactivation air stream. Moisture

is rejected from the desiccant surface through a difference in vapor pressure and is expelled by a

waste air stream. The regeneration of the desiccant material requires a heated or dry air stream.

There are different configurations of solid desiccants devices, such as the often used desiccant

wheel.




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Other dehumidification methods

There are other dehumidification methods for specialized industrial and commercial applications in use.

One of these methods is dehumidification of air by means of compression. As the pressure in air is

increased, the dew point or temperature at which water will condense is raised. In order to get dry air

within the vessel, a heat sink is needed to cool the compressed air. Typically, costs can be prohibitive

when used outside the specialized process, because of the equipment, space, and auxiliary equipment

required for the process. If compressed air is already available and only very small amounts of dry air are

needed for humidity control, compression could be feasible route to dry air. Dehumidification by

compression of air is, however, not feasible for application of space conditioning.

One new humidification method is electrostatic dehumidification technology. Rather than removing

humidity from the air by means of cooling‐based phase change or physical sorption processes, moisture

is removed through electrostatic fields which separate moisture vapor from air streams. The reported

benefit is a significant reduction in energy used for dehumidification, by eliminating energy needed for a

phase change or the regeneration of desiccants. This literature review has not identified advanced

prototypes or even commercially available of electrostatic dehumidifiers at this time.

3.2 Review of Psychrometrics ‐ Principles

Psychrometrics describes the physical and thermodynamic properties of gas‐vapor mixtures, whereby

the properties of moist air, e.g. a mixture of water vapor and air, is of significance to human comfort and

conditions in the built environment. In the process of humidification and dehumidification properties of

the moist air are changed, sometimes with or without transfer of heat or matter between the air and its

surroundings, which is referred to as isothermal or adiabatic processes. The review of dehumidification

methods necessitates a basic understanding of psychrometric principles. The following provides an

overview of some of the most important topics and metrics of moist air processes related to

dehumidification. A more comprehensive introduction to moist air properties and processes used in the

HVAC is given by ASHARE (2009).

Psychrometric Chart

The analysis of moist air properties in HVAC processes is facilitated by the psychrometric chart and

related property tables. There are multiple psychrometric charts available, each providing graphical

interpretations for constant pressures. Figure 3.1.1 shows the ASHRAE psychrometric chart for sea‐level

(ASGRAE, 2009). The chart is in SI units. If any two properties of the moist air are known, the chart

allows for a quick assessment of all other properties of moist air.




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Dry‐Bulb Temperature

The dry‐bulb temperature is also referred to as the “sensible” temperature of the air — the heat which

can be sensed by a dry thermometer. In the chart below, the dry bulb temperature of the air is displayed

at the bottom, increasing from left to right. Constant dry bulb temperatures are depicted as vertical

lines. (refer to Figure 3.1.2).

Figure 3.1.1: Psychrometric

Chart for pressure at sea

level (ASHRAE, 2009)

Presented in SI (metric

units)

Figure 3.1.2: Dry‐bulb temperature

displayed in the psychrometric chart,

(ASHRAE, 2009)

In the depicted psychrometric chart

temperature are given in degrees Celsius

(°C).

Base psychrometric Chart by ASHRAE;

modified for illustration example




July 9, 2016 Page 10 of 171

The dry‐bulb temperature is usually referred to as “air temperature”. The dry‐bulb temperature is

measured with a thermometer that is exposed to the air but shielded from radiation, and refers basically

to the ambient air temperature. It is called "Dry Bulb" because the air temperature is indicated by a

thermometer not affected by the moisture of the air.

Wet‐Bulb Temperature

The Wet Bulb temperature is the adiabatic saturation temperature. This means the temperature at

which water, by evaporating into moist air at a certain dry bulb temperature and absolute humidity

ratio, can bring air to saturation adiabatically, at a constant pressure. The adiabatic evaporation of water

and resulting the cooling effect causes the wet bulb temperature to be lower than the dry‐bulb

temperature.

The wet‐bulb temperature can be measured by a psychrometer, which consists of two thermometers,

where one thermometer’s sensor is covered by a wetted wick. Placing the instrument in an airstream

promotes evaporation and cooling, with the equilibrium temperature being the wet‐bulb temperature.

While this procedure is not strictly adiabatic saturation, the resulting error compared to the exact

thermodynamic wet‐bulb temperature is sufficiently small enough to allow use of the experimental

procedure for technical applications. Figure 3.1.3 illustrates wet‐bulb temperatures as lines in the

psychrometric graph.

Figure 3.1.3: Wet‐bulb temperature

displayed in the psychrometric chart






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Relative Humidity ‐ Percent of Saturation

In the technical literature relative humidity is often referred to as the moisture content of air, expressed

as a percent value, which air "can hold". This definition is not exact since it does not express a

temperature dependency, and sometimes causes confusion with the absolute humidity. As can be seen

in Figure 3.1.4. On the psychrometric chart, relative humidity is displayed as a series of curves,

increasing from the bottom of the chart. The “saturation curve” is the left boundary and represents

100% relative humidity. (see Figure 3.1.4)

Figure 3.1.4: Relative humidity of air‐

water vapor mixtures



The more exact definition of relative humidity in air can be expressed by vapor and air partial pressure,

density of the vapor and air, or by the actual mass of the vapor and air. Since partial vapor pressure is an

important property for all dehumidification processes, we define relative humidity as the ratio of vapor

partial pressure in the air to the saturation vapor partial pressure of the air at a certain dry bulb

temperature. The following equation calculates the relative humidity (The EngineeringToolBox).

ϕ = pW / pWS * 100% (1)

where: ϕ = relative humidity (%) pW = vapor partial pressure (mbar) pWS = saturation vapor partial pressure at the actual dry bulb temperature (mbar)




July 9, 2016 Page 12 of 171

Dew Point Temperature:

With decreasing temperatures moist air can “hold” less moisture than at higher temperatures. The dew

point refers to the temperature at which water vapor starts to condense out of the air. The dew point

depends on the absolute amount of water vapor in the air. The higher the amount of water vapor in the

air the higher is the dew point. Figure 3.1.5 shows an example of a dew point for a sample of moist air,

whose properties are defined by dry‐bulb and wet‐bulb temperature of 20°C and 16°C, respectively.

Consequently, dew‐point temperature tdp is the temperature of moist air saturated at pressure p, with

the same absolute humidity as that of the given sample of moist air. Cooling‐based dehumidification

systems remove moisture from air by cooling it below its dew‐point temperature, at which water vapor

condenses and is separated from the air as liquid phase draining from the cold coils.

Figure 3.1.5: Dew point temperature

The example shows a sample of moist air

with dry‐bulb and wet‐bulb temperatures

at 20°C and 16°C, respectively (A). The

value of absolute moisture of the sample

is determined by the horizontal line (A) to

(C). As the sample is cooled (sensitive

cooling) the state point of the sample

moves along the horizontal line towards

the saturation line (B). The dry‐bulb

temperature of the intersection of the

horizontal line of constant absolute

humidity and the saturation curse is the

dew point for the sample.



Humidity Ratio

The humidity ratio can be defined as the ratio between mass of water vapor present in moist air to the

mass of dry air.

x = mw / ma where x = humidity ratio (kg/kg or g/Kg) mw = mass of water vapor (kg or g) ma = mass of dry air (kg)




July 9, 2016 Page 13 of 171

Figure 3.1.6 illustrates the use of humidity ratio for a sample of moist air, whose properties are defined

by dry‐bulb and wet‐bulb temperature of 20°C and 16°C, respectively.

Figure 3.1.6: Humidity Ratio



at 20°C and 16°C, respectively (A).

The resulting humidity ratio can be read

from the graph as approximately 9.7 g/kg,

which indicates that are 9.7 g of water

vapor per kg of dry air. The humidity ratio

can also be expressed as kg/kg. In our case

9.7 g/kg converts to 0.0097 kg/kg.



Enthalpy

Enthalpy is the measure of the total energy in air. The enthalpy of moist air includes sensible and latent

heat, which is the enthalpy of the dry air and the enthalpy of the evaporated water in the air,

respectively. The total enthalpy, e.g. the sum of sensible and latent, is used when calculating cooling and

heating processes. The specific enthalpy of moist air is defined as the total enthalpy of the dry air and

the water vapor mixture per unit mass (of dry air. Therefore, the specific enthalpy has the units of

kJ/Kgda.

Specific enthalpy of moist air can be expressed as (ASHRAE, 2009):

h = ha + x hw

where

h = specific enthalpy of moist air (kJ/kg, Btu/lb)

ha = specific enthalpy of dry air (kJ/kg, Btu/lb)

x = humidity ratio (kg/kg, lb/lb)

hw = specific enthalpy of water vapor (kJ/kg, Btu/lb)

The determination of difference in enthalpy when going between state points in the psychrometric chart

is of important in determining energy requirements in heating and cooling processes. Figure 3.1.7

illustrates the use of enthalpy in moist air applications for a sample of moist air, whose properties are

defined by dry‐bulb and wet‐bulb temperature of 20°C and 16°C, respectively.




July 9, 2016 Page 14 of 171

Figure 3.1.7: Enthalpy




The resulting specific enthalpy is shown as

44.5 kJ/kgda



Specific Volume

Often it is required to convert between mass and volumetric flow rates or to calculate energies required

to move between state points. Therefore, the specific volume v of a moist air mixture is expressed in

terms of a unit mass of dry air:

v = V/Mda [m3/kgda]

where

V is the total volume of the mixture and Mda is the total mass of dry air

Figure 3.1.8 illustrates the determination of specific volume for a sample of moist air, whose properties

are defined by dry‐bulb and wet‐bulb temperature of 20°C and 16°C, respectively.

The use of psychrometric chart in solving mass and heat transfer processes in cooling is best carried out

with illustrative examples. Section 3.2 presents three representative cases, which helps in the

understanding of cooling and dehumidification processes presented in this literature review. For a more

analytical explanation of the examples refer to ASHRAE (2009).




July 9, 2016 Page 15 of 171

Figure 3.1.8: Humidity Ratio




The resulting specific volume is shown as

0.864 m3/kgda



3.3 Application of Psychrometric Chart

This section presents three case applications of the psychrometric chart, which helps the reader to

understand cooling and dehumidification processes presented in this literature and technology review.

Application of psychrometric chart ‐ Illustrative Case A; Sensitive Cooling

Statement: Moist air at dry‐bulb and wet‐bulb temperature of 30°C and 18°C, respectively, enters a

cooling coil assembly at a rate of 5 m3/s. The chilled air leaves the coil at a dry‐bulb temperature of 20C.

Determine the rate of heat that has to extracted from the air flow.

Solution: Figure 3.2.1 shows the heat and mass flow diagram of the example. Figure 3.2.2 shows the

psychrometric properties of the moist air at state points (1) and (2).




July 9, 2016 Page 16 of 171

Figure 3.2.1: Heat and mass flow diagram

Moist air with properties at (1) and at 5 m3/s flow rate enters the system interface cooling coils assembly (1) and is cooled by an external heat sink. The rejected heat through the cooling coils results in the air properties at system interface (2).

mda is the mass flow in m3/s h is the specific enthalpy in kJ/kgda W is the humidity ratio in g/kgda 1q2 is the heat rate in kW that is extracted

from the air flow between interfaces (1) and (2)


property for illustrative case A

The following properties are read

from the chart:

h1 = 50.5 kJ/kgda

h2 = 40.0 kJ/kgda

v1 = 0.869 m3/kgda

W1 = W2 = 8 g/kgda

Base psychrometric Chart by

ASHRAE; modified for illustration

example

Since there is no phase change the moisture ratio of conditions (1) and (2) remain constant at 8 g/kgda

The air mass flow rate is calculated with mda = 5.0 m3/s / 0.869 m3/kgda = 5.75 kgda/s.

The rate of heat that is extracted from the air is calculated as follows:

1q2 = mda *( h1 ‐ h2) = 5.75 [kgda/s] * (50.5 – 40.0) [kJ/kgda] = 60.4 [kJ/s] = 60.4 kW




July 9, 2016 Page 17 of 171

Application of psychrometric chart ‐ Illustrative Case B; Moist air cooling and dehumidification:

Statement: Moist air at 32°C dry‐bulb temperature and 60% RH enters a cooling coil at 10 m3/s and is

processed to a final saturation condition at 10°C. Find the required rate of heat extraction

(refrigeration).


psychrometric properties of the moist air at system interfaces (1) and (2).

Figure 3.2.3 Heat and mass flow diagram for illustrative case C (ASHRAE, 2009)

Moist air with properties at (1) and at 10 m3/s flow rate enters the system interface (1) and is cooled by an external heat sink. The rejected heat through the cooling coils results in the air properties at system interface (2) and a mass of water (condensate)

mda is the mass flow in m3/s h is the specific enthalpy in kJ/kgda W is the humidity ratio in g/kgda hw is the specific enthalpy of water at

saturation at final saturation temperature

mw is the mass flow rate of condensed water

1q2 is the heat rate in kW that is extracted from the air flow between interfaces (1) and (2)

Figure 3.1.11 indicates the energy equation as follows:

Heat flux incoming Heat flux outgoing

mda * h1 mda * h2 + mw * hw + 1q2




July 9, 2016 Page 18 of 171


property for illustrative case

B

The following properties

are read from the chart:

h1 = 70.5 kJ/kgda

h2 = 29.5 kJ/kgda

v1 = 0.885 m3/kgda

W1 = 15.0 g/kgda

W2 = 7.6 g/kgda

Base psychrometric Chart

by ASHRAE; modified for

illustration example

The air mass flow rate is calculated with mda = 10.0 m3/s / 0.885 m3/kgda = 11.3 kgda/s.

The tabulated value of specific Enthalpy, of Water at Saturation at (2) is 42.02 kJ/kgw

The rate of heat that is extracted from the system is calculated as follows:

mda * h1 = mda * h2 + mw * hw + 1q2

1q2 = mda (h1 ‐ h2) ‐ mw * h hw

With: mw = mda * (W1 ‐W2)

1q2 = mda (h1 ‐ h2) ‐ mda * (W1 ‐W2) * hW

1q2 = mda ((h1 ‐ h2) ‐ (W1 ‐W2) * hW)

1q2 = 11.3 [kgda/s] * ((70.5‐29.5) [kJ/kgda] – ((15.0‐7.6)/1000) [kgw/kgda] * 42.02 [kJ/kgw])

1q2 = 459.8 kJ/s = 459.8 kW




July 9, 2016 Page 19 of 171

Application of psychrometric chart ‐ Illustrative Case C; Adiabatic mixing of two moist air streams

Statement: Two streams of moist air are adiabatically mixed. Stream (1) has 4 m3/s with 6°C and 4°C

dry‐bulb and wet‐bulb temperatures, respectively and Stream (2) has 10 m3/s with 25°C dry‐bulb

temperature with 60% RH. Find the dry‐bulb temperature and RH of the resulting moist air mixture.


psychrometric properties of the moist air at system interfaces (1) and (2).

Figure 3.2.5: mass and energy flow

diagram for illustrative case C

Two streams (1) and (2) of moist air are

adiabatically mixed. The resulting mixed

stream (3) is a combination of the two

psychrometric conditions.

mda is the mass flow in m3/s

h is the specific enthalpy in kJ/kgda

W is the humidity ratio in g/kgda

Under the adiabatic conditions conservation of energy and mass results in the following three

relationships:

Heat flux incoming Heat flux outgoing

mda1 * h1 + mda2 * h2 mda3 * h3

Dry air mass incoming Dry air mass outgoing

mda1 + mda2 mda3

Water vapor mass incoming Water vapor mass outgoing

mda1 * W1 + mda2 * W2 mda3 * W3




July 9, 2016 Page 20 of 171


properties for illustrative case C

The following properties are read

from the chart:

h1 = 17.2 kJ/kgda

h2 = 55.7 kJ/kgda

v1 = 0.776 m3/kgda

v2 = 0.861 m3/kgda

W1 = 4.3 g/kgda

W2 = 12.0 g/kgda

Base psychrometric Chart by

ASHRAE; modified for illustration

example

The three relationships can be rearranged:

h2 – h3 =

W2 – W3 =

mda1

h3 – h1 W3 – W1 mda2

This results in a weighted distribution of properties, where the resulting properties of condition (3) as

follows:

mda1 = 4.0 [m3/s] / 0.776 [m3/kgda] = 5.2 [kgda/s]

mda2 = 10.0 [m3/s] / 0.861 [m3/kgda] = 11.6 [kgda/s]

mda1 / mda2 = 5.2 / 11.6 = 0.444 [‐]

h3 = (h2 + 0.444*h1) / 1.444 = 43.9 [kJ/kgda]

W3 = (W2 + 0.444 * W1 / 1.444 = 9.6 [g/kgda]

Figure 3.2.7 shows the results depicted in the psychrometric chart




July 9, 2016 Page 21 of 171

Figure 3.2.7: Resulting

psychrometric properties for

illustrative case C

h3 = 43.6 kJ/kgda

W3 = 10.0 g/kgda

Tdb3 = 18.5 °C

RH3 = 74%

The property of (3) is on line between (1) and (2); the distance (2) to (3) is 44.4% and (3) to (1) is 55.6% of the distance (1) to (2)

Base psychrometric Chart by ASHRAE; modified for illustration example

3.4 Introduction to Desiccants

Desiccants are types of sorbents, where sorbents are materials characterized by their ability to attract

and hold gases or liquids. Desiccants have a particular affinity to water.

Virtually any material acts as desiccants and attracts water. Building materials can attract sizable

amounts of water, such as woolen carpets, which can hold up to 23% of their dry mass in water vapor.

Commercial desiccants must have much higher ratios of water attraction, as high as 1000% their dry

mass in water vapor, depending on the type and on the moisture available in the environment. In

addition, desiccants continue to attract moisture even when the surrounding air is quite dry and the

driving forces for sorption are small, a characteristic that other materials do not share.

The process of water sorption is basically the same in all types of commercial desiccants. The

conditioned desiccant material is subjected to a moist air stream and attracts moisture until saturated,

reaching equilibrium with the surrounding air. The desiccant is then regenerated by applying heat to

drive out moisture. Since the process of sorption generates heat due to latent heat of water vapor and,

to a smaller amount, due to additional heat of sorption, the desiccant material has to be cooled before

the next cycle of sorption can commence.

There are two basic processes of sorption, referred to as adsorption or absorption, depending on

whether the desiccant undergoes a chemical change as it takes on moisture. Adsorption does not

change the desiccant, except by adding water vapor mass. Absorption changes the desiccant as water

enters into and changes properties of the material.




July 9, 2016 Page 22 of 171

Typical Desiccant Applications

Desiccants have been used in the HVAC industry to a certain extent, to dry moist ambient air. Their

market penetration is, however, well behind cooling‐based dehumidification, due to cold coil

dehumidification being a relatively simple and low cost alternative. While lacking in the number of

installed units compared to conventional HVAC applications, desiccant dehumidification can offer

tangible benefits under the following operating conditions:

• The latent load in the moist air is large in comparison to the sensible load; using only cooling‐

based dehumidification could overcool the building.

• Energy cost to regenerate the desiccant is low compared to the energy costs of refrigeration

dehumidification, including the energy for reheating of low dew point air supply.

• If too low dew‐point air is required, refrigeration dehumidification coils would need to be

operated at subfreezing dew points resulting in icing problems.

3.5 Desiccant Cycle

The driving force of desiccant dehumidification is the difference between water vapor pressures at the

desiccant’s surface and of the surrounding air. Sorption of moisture occurs with a positive vapor

pressure gradient at the desiccant surface, which means a higher vapor pressure in the surrounding air

than at the exposed desiccant surface. Desorption occurs with an opposite pressure gradient, when the

desiccant surface vapor pressure is higher than that of the surrounding air.

The nonlinear relationship between desiccant moisture content and vapor pressure at the desiccant

surface is depicted in Figure 3.4.1. As the moisture content of the desiccant rises, so does the water

vapor pressure at its surface. At some point, an equilibrium condition occurs, when the vapor pressure

at the desiccant surface is the same as that of the surrounding air. At this point neither sorption nor

desorption occurs, since the pressure difference has reached a minimum threshold which is an

insufficient driving force. At this point the desiccant has to be regenerated to restore the desiccant’s

ability for sorption of water vapor. Figure 3.4.2 illustrates the effect of temperature on vapor pressure at

the desiccant surface. Both higher temperature and increased moisture content increase surface vapor

pressure. This indicates that desiccants can attract moisture better at lower temperatures, since the

pressure gradient between the surrounding air and the desiccant surface is larger. At higher

temperatures the vapor pressure at the desiccant surface is higher, which results in less sorption or, in

case of desiccant regeneration, in better desorption when moisture is driven from the desiccant to the

surrounding air.

Figure 3.4.2 illustrates the sorption and desorption cycle used in desiccant dehumidification. The cycle

starts at (1) with a cold and dry desiccant coming into contact with a moist air stream. During process




July 9, 2016 Page 23 of 171

step (1) to (2) moisture is attracted and the moisture content of the desiccant increases. The desiccant

surface temperature increases due to latent heat and heat of sorption. As the desiccant enters

equilibrium conditions at (2) sorption stops. During process steps (2) to (3) external heat is added and

the vapor pressure at the desiccant surface increases, creating a pressure gradient to drive moisture out

of the desiccant. At (3) the moisture content of the has fallen to a desired process condition. During

process step (3) to (1) the desiccant is cooled to attain the lower vapor pressure of the initial state at (1),

where a new cycle starts.

The effectiveness of desiccant cycle depends on the energy required to move through this cycle. The

process step (1) to (2), when dehumidification occurs, requires only relatively small fan or pump energy

to bring the desiccant and the air stream into physical contact. The process steps (2) to (3) and (3) to (1)

requires significant heating and cooling energy, respectively.

Figure 3.4.1: Desiccant Water Vapor Pressure as Function of Moisture Content and Temperature (ASHRAE, 2009)

Figure 3.4.2: Desiccant Cycle (ASHRAE, 2009)

During process step (2) to (3) heating requires the sum of the following heat input:

• Sensible heat input to raise the desiccant to a temperature high enough to make its surface vapor

pressure higher than that of the surrounding air, the higher the

• Latent heat input to evaporate moisture




July 9, 2016 Page 24 of 171

• Latent heat for desorption of water from the desiccant (a small amount)

During process step (3) to (1) cooling requires sensible heat rejection to the surrounding, where the

required cooling capacity is proportional to the desiccant mass properties and difference between its

temperature at (3) and (1). Regeneration in the desiccant cycle also works with pressure differences.

This is used in certain industrial and commercial applications but not in HVAC applications.

Summarizing, the greater the difference between the air and desiccant surface vapor pressures, the

greater is the ability of the desiccant material to attract moisture from the air. The selection of the most

effective desiccant for a particular application depends on the range of water vapor pressures in the air,

the temperature of the regeneration process, and moisture sorption and desorption characteristics of

the desiccant.

3.5 Desiccant Material

There are two forms of desiccant materials, solid and liquid desiccants.

Liquid Desiccants act as absorbents to attract and hold moisture. In a liquid absorption dehumidifier, air

is brought into contact with liquid desiccant solution. If the water vapor pressure of the liquid desiccant

surface is lower than that of air, moisture is absorbed at the surface film surface and therefore moisture

migrates from the air to the desiccant; this process is called dehumidification. Conversely, if water vapor

pressure at the desiccant surface is higher than that in the air moisture migrates to the air; this process

is called desorption.

The vapor pressure of the liquid desiccant is directly proportional to its temperature and inversely

proportional to its concentration. Figure 3.5.1 illustrates the effect of temperature and concentration on

the vapor pressure of various solutions of water and lithium chloride (LiCl), a common liquid desiccant.

In the sample condition shown in Figure 3.5.1 the solution that is 25% lithium chloride has a vapor

pressure of 1.25 kPa and 3.3 kPa at temperatures of 21°C and 37.8°C, respectively. This indicates that

the warmer the liquid desiccant is, the less moisture it can absorb from the air.

In standard practice, performance of liquid desiccants is controlled by adjusting process temperature,

concentration, and pressure. The desiccant process temperature is controlled by simple heaters and

coolers to adjust the temperature of the desiccant. Liquid desiccant concentration is controlled by

heating the desiccant to drive moisture out into a waste airstream or directly to the ambient.

Commercially available liquid desiccants have an especially high water‐holding capacity. Selected

products have water absorption capacity of more than 1,000 % on a dry‐mass basis. In order to achieve

effective sorption and desorption processes, the two‐phase flow through the process reactor has to

provide sufficient free surfaces of desiccant and a good flow distribution. High surface renewal rates and




July 9, 2016 Page 25 of 171

maximum contact time between the gaseous and liquid phase allow the desiccant agent to approach its

theoretical sorption and desorption capacity.

Solid desiccants attract moisture as a surface active sorption process. The moisture does not change the

nature of the solid adsorbent. The condensed adsorbed moisture is confined into capillaries, and the

individual adsorbents attract moisture through electrical field charges at the desiccant surface. The

electric field is not uniform in magnitude and polarity, therefore specific sites on the desiccant surface

attract water molecules that have a net opposite charge. When the complete surface is covered with

moisture, the adsorbent can hold still more moisture because vapor condenses into the first water layer

and fills the capillaries throughout the material.

The ability of solid adsorbents to attract moisture is a function of vapor pressure difference between the

adsorbent surface and the surrounding air. Generally, solid adsorbents have a smaller capacity to attract

water per dry weight than liquid absorbents. ASHRAE (2009) suggests that, a typical molecular sieve

adsorbent can hold 17% of its dry mass in water when the air is at 21°C and 20% RH, whereas a LiCl‐

solution can hold 130% of its mass at the same process conditions.

Solid adsorbents have an advantage with respect to temperature dependency. Molecular sieves, for

example, can adsorb moisture in warm airstreams, making dehumidification possible at elevated

outdoor temperatures. Solid adsorbent can be manufactured with specific pore diameters to increase

adsorption capacity for selected ingredient in the air stream. For example, water has an effective

molecular diameter of 0.32 nm. A molecular sieve adsorbent with an average pore diameter of 0.40 nm

adsorbs water but has almost no capacity for larger molecules, such as organic solvents. This selective

adsorption characteristic is useful in many applications. For example, several desiccants with different

pore sizes can be combined in series to remove first water and then other specific contaminants from an

airstream.

Intrinsic parameters that affect adsorption performance of different desiccant material include total

surface area, total volume of capillaries, and range of capillary diameters. A large surface area gives the

adsorbent a larger capacity at low relative humidity. Large capillaries provide a high capacity for

condensed water, which gives the adsorbent a higher capacity at high relative humidity. A narrow range

of capillary diameters makes an adsorbent more selective in holding vapor molecules. Figure 3.5.1

illustrates the effects of the three parameters shown by three noncommercial silica gel adsorbents. Each

of the adsorbents has a different internal structure. Gel 1, shown in Figure 3.5.1 has large capillaries,

making its total volume large but its total surface area small. Gel shows large and small adsorption rates

at high and low relative humidity, respectively.

For example, materials with large capillaries necessarily have a smaller surface area per unit of volume

than those with smaller capillaries. As a result, adsorbents are sometimes combined to provide a high

adsorption capacity across a wide range of operating conditions.




July 9, 2016 Page 26 of 171

Figure 3.5.1: Adsorption and Structural

Characteristics of Some Experimental

Silica Gels (ASHRAE, 2009)

The figure illustrates the effect of

capillary diameters on the capacity of

adsorbents to attract and hold

moisture.

The adsorbent Gel 1 has a high capacity

to attract condensation, which means at

high relative humidity.

Types of solid desiccants: Sorbent Systems (2016) presents five types of solid desiccants which are often

used dehumidification applications:

Montmorillonite Clay (“Clay”) is a naturally occurring porous adsorbent with good regeneration

characteristics at low temperatures without substantial deterioration or swelling. Clay is inexpensive

and effective within normal temperature and relative humidity ranges. Variations in adsorption

performance are due to differences in obtaining this naturally occurring material.

Silica gel has an amorphous micro‐porous structure with a distribution of pore opening sizes of roughly

3‐60 angstroms, which results in a very large surface area that will attract and hold water by

adsorption and capillary condensation. Silica gel is efficient at temperatures below 77°F, but loses




July 9, 2016 Page 27 of 171

adsorbing capacity at higher temperatures, similar to Clay. Silica gel is non‐corrosive and nontoxic

and has received approval for use in food and drug packaging.

Molecular sieves (also Zeolite) has a larger adsorption capacity than either silica gel or clay. A molecular

sieve is synthetic rather than naturally occurring, and therefore higher in cost per unit, but due to its

extremely large range of adsorptive capabilities, it might often be the best value.

Calcium oxide (CaO) can adsorb much greater amounts of water at low relative humidity than other

materials. It is also effective in retaining moisture at high temperatures. This desiccant is relatively

inexpensive as compared to many other desiccants.

Calcium sulfate (CaSO4), also known commercially as Drierite®, is an inexpensive manufactured

desiccant alternative general‐purpose desiccant geared mainly toward laboratory use. It is

chemically stable, non‐disintegrating, nontoxic, non‐corrosive, and does not release its adsorbed

water when exposed to higher ambient temperatures. CaSO4 is a low cost desiccant but has a

comparably smaller small adsorbent rate compared to dry weight. CaSO4 has regeneration

characteristics that tend to limit its useful life. Table 3.5.1: shows important properties of the five

solid desiccants.

Table 3.5.1: Properties of the five desiccants. (Sorbent Systems, 2016)

Property Molecular Sieve

Silica Gel Clay CaO CaSO4

Adsorptive Capacity at low

H20 Concentrations Excellent Poor Fair Excellent Good

Rate of Adsorption Excellent Good Good Poor Good

Capacity for Water @77° F,

40% RH High High Medium High Low

Separation by Molecular

Sizes Yes No No No No

Adsorptive Capacity at

Elevated Temperatures Excellent Poor Poor Good Good

Figure 3.5.2 shows the moisture adsorption capacity of the five desiccants as a function of relative

humidity. Figure 3.5.3 illustrates activation conditions of several desiccants. Activation conditions are of

importance for regeneration. Figure 3.5.5 illustrates the adsorption rates as a function of time.




July 9, 2016 Page 28 of 171

Figure 3.5.2: Adsorption capacity for moisture as a function of relative humidity (Sorbent Systems, 2016)

Figure 3.5.3: Temperature dependence of activation conditions for regeneration (Sorbent Systems, 2016)

Figure 3.5.4: Equilibrium Capacity (H2O) of solid

desiccants (Sorbent Systems, 2016)


SECTION 4 ‐ METHODS OF DEHUMIDIFICATION


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This section presents main methods of dehumidification. While the topic of this literature report is

desiccant dehumidification and related technology applications, a brief summary of cooling based

dehumidification is provided. There are potential applications where desiccant and cooling‐based

dehumidification can be combined to achieve effective cooling and space conditioning.

4.1 Cooling‐based Dehumidification

The basis for cooling‐based dehumidification is the compression refrigeration cycle. Figure 4.1.1 shows a

simple process diagram of the cooling‐based dehumidification, where humid air passes over cooling

coils, moisture condenses and is removed by gravity separation. The low temperature in the cooling

coils is maintained by a refrigeration system. Typically, cooling‐based dehumidification occurs in

conjunction with an air conditioning systems.

Figure 4.1.2 illustrates the psychrometric process of cooling the air below the dew point. The figure

indicates that there are two thermal processes, sensible and latent cooling, where latent cooling is the

process of dehumidification.

Figure 4.1.1: Basic process cycle for cooling‐based

dehumidification. (Harriman, 2002)

Humid air passes over cooling coils which are at a

temperature below the dew point of air. The

humidity condenses on the coils

Figure 4.1.2: Psychrometric air path (Harriman, 2002)

Cooling systems first chill the air to its dew point (100%

relative humidity). After that point, further chilling

removes moisture. The more the air is cooled, the more

will it be dried.

Figure 4.1.3 shows the type of refrigeration cycle which is used in most commercial and residential air

conditioning systems. The system basically conveys and rejects heat from the conditioned space to the

outdoor environment, using the refrigerant to transport the heat energy.




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Heat is removed from the conditioned space by first transferring its heat energy to the expanding and

vaporizing refrigerant in the evaporator. The latent heat of the refrigerant evaporation in the evaporator

creates a significant heat sink. From the evaporator, the refrigerant vapor flows to the compressor,

where pressure in the refrigerant is increased. Further downstream the high pressure refrigerant vapor

enters the condenser where condensation occurs creating high temperatures resulting in heat rejection

to the external environment or another external heat sink. The now liquid refrigerant flows to an

expansion valve which creates a pressure differential across the valve as the in the refrigerant travels

downstream to the low pressure evaporator. As the refrigerant enters the evaporator, the refrigeration

cycle is completed.

Figure 4.1.3: Basic refrigeration cycle used in

The type of refrigeration cycle shown is used in most

commercial and residential air conditioning systems.

Cooling‐based dehumidification systems can have different component configuration. There are three

basic working principles to provide the cooling capacity and heat sink for water vapor condensation:

Direct expansion cooling system

Chilled liquid cooling system

Dehumidification‐with‐reheat system

For the direct expansion cooling and chilled liquid cooling systems dehumidification occurs in

conjunction with air conditioning systems where cooling and dehumidification occurs. The

dehumidification‐with‐reheat system does not contribute to space cooling, but only contributes to

dehumidification.

Direct expansion cooling: This type of cooling system is often referred to as “DX” or direct expansion

system. This system is mostly used in residential or roof top cooling system installations. Figure

4.1.4 illustrates the DX‐system diagram. Moist supply air flows over cooling coils, which serve as

the evaporator.

Chilled liquid cooling: Chilled liquid systems use a secondary cooling loop to cool and dehumidify air. In

the secondary cooling loop chilled water, or other secondary cooling fluid, is used to transport heat

from the cooling coils to the evaporator. The secondary cooling loop adds complexity to the system




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and therefore it is typically used in larger system installations, where they can gain advantages of

installed cost and operating efficiency over DX systems. Figure 4.1.5 illustrates the chilled liquid

cooling system diagram.

Dehumidification‐with‐reheat: It is common practice that air‐conditioning systems use reheat to

increase the temperature of the supply air downstream of the cooling coils. This is done to assure

occupant comfort, since cold air can lead to draft and overcooling of conditioned spaces. This

section presents a dehumidification configuration that most residential, but some commercial,

dehumidifiers use. This system is depicted in Figure 4.1.6, and only controls humidity level in spaces

humidity without cooling. The dehumidification‐reheat systems uses combined cooling and to

achieve great efficiency. As illustrated in Figure 4.1.6, the dehumidification‐reheat system has a

condenser coil for reheat immediately downstream of the evaporator, which is the cooling coil

where moist air sheds humidity. This configuration results in an efficient dehumidification process

performance, since the low air temperature after the cooling coil makes the refrigerant condenser

very efficient. The reheat energy from the condenser is essentially free, since it is heat rejected by

the cooling process.

While for most dehumidification applications cooling‐based process provides good system performance,

there are problems with frozen condensate that can occur with very low dew point conditions. In these

cases, frozen condensate reduces heat transfer effectiveness and can clog coils, thus reducing the

airflow. For systems that cool air below 32°F dew point defrost coil systems and suitable controls can

provide some mitigation.

Figure 4.1.4: Direct expansion

cooling process diagram

(Harriman, 2002)

The DX‐cooling cycle is mostly

used in commercial and

residential cooling applications.

Here dehumidification occurs

in conjunction with space

cooling. In the DX system the

cooling coils over which the

moist air flows function as the

evaporator.




July 9, 2016 Page 32 of 171

Figure 4.1.5: Chilled liquid

cooling process diagram

(Harriman, 2002)

The chilled liquid cooling cycle

uses a secondary cooling loop

to transport heat between the

cooling coils and the

evaporator. The secondary

cooling loop complexity and

but offers benefits. The chilled

liquid cooling loop is typically

used in larger installations.

Figure 4.1.6: Dehumidification‐reheat

system (Harriman, 2002)

This dehumidifier configuration uses a cooling‐reheat scheme

to remove moisture from air. The arrangement is energy efficient because it uses condenser reheat downstream of the evaporator. The reheat energy is essentially free, and the condenser is most efficient in the low temperature air that comes from the evaporator coil.

The source image (Harriman, 2002) was

edited by adding comments

4.2 Desiccant Dehumidification

This section introduces several process characteristics of desiccant dehumidifying. Section 5 will present

more detail about the physical properties of desiccants and the desiccant dehumidification process. The

process of desiccant dehumidification can be separated in three main process steps. These three main

process steps are described in Figure 4.2.1.




July 9, 2016 Page 33 of 171

Desiccant Dehumidification process‐step A: Sorption

This process steps describes the process from point (1) to point

(2).

The surface vapor pressure of the desiccant material is low

because it is dry and cool and has a low moisture content

(conditions at point (1)). As the desiccant draws moisture from

the surrounding air, the desiccant surface becomes saturated

and hot, because of the heat of condensation and sorption

(conditions at point (2)). At point (2) vapor pressure is equal to

that of the surrounding air and the desiccant cannot collect

more moisture because there is no pressure difference

between the surface and the vapor in the air.

Desiccant Dehumidification process‐step B: Desorption


(3).

The desiccant to conditions of state point (2) requires energy

(e.g. heat) input to go to point (3). For this the desiccant is

taken out of the moist air, heated, and placed into a different

airstream. Due to an increase of temperature the vapor partial

pressure at the desiccant surface increases to a point where

the vapor pressure is higher than in the surrounding air.

Therefore, moisture moves off the desiccant surface to the air

to equalize the pressure differential. At the conditions at point

(3) the desiccant is dry and hot. Therefore, although the

desiccant is dry, its vapor pressure is still too high to collect

moisture from the air, due to the desiccants high temperature.




July 9, 2016 Page 34 of 171

Desiccant Dehumidification process‐step C: Cooling


(1). This completes the desiccant basic process cycle.

The desiccant is cooled to restore a low vapor pressure,

returning it to conditions at point (1).

Figure 4.2.1: Three main process steps A through C of the basic desiccant dehumidification cycle,

(Harriman, 2002)

These three basic process steps are used in basically all desiccant equipment configurations. Harriman

(2002) presents five typical process equipment configurations for desiccant dehumidifiers:

1. Liquid desiccant packed column

2. Solid desiccant packed tower

3. Solid desiccant rotating horizontal bed

4. Solid desiccant multiple vertical bed

5. Solid desiccant rotating Honeycomb

Liquid desiccant packed columns: In the packed column dehumidifiers liquid desiccant solution is

brought into contact with moist air. Figure 4.2.2 shows a typical configuration with two process vessels.

The larger of the two vessels is used for the dehumidification (sorption) process step and the smaller

vessel for the regeneration (desorption) of the desiccant solution. In both vessels, liquid desiccant

solution is either sprayed or distributed on the packing inside the vessel (e.g. packing of rings and other

geometries). The liquid desiccant solution runs down on the packing material and shear forces stimulate

surface renewal. Moist air is flowing upwards and is brought into contact with the liquid desiccant

solution. The desiccant solution drips down into a pan and from where a portion is recirculated and the

remaining portion is pumped to the smaller regeneration. Control indicates how much desiccant has to

be recirculated. In the smaller vessel liquid desiccant is regenerated by adding heat, letting it run down

over packing like in the larger vessel and blowing a separate air stream, the reactivation air, through the

smaller vessel. Moisture leaves the desiccant and moves to the reactivation air. An important process

parameter is the creation of active surfaces, also referred to as surface renewal. In this process surfaces

of droplets renew frequently as droplets break up, form new drops and are mixed into the mass of the

liquid desiccant. Some inert packing or baffles inside the sorption process vessel increases the




July 9, 2016 Page 35 of 171

magnitude of surface renewal. The numbers on the diagram indicates how the process equipment

perfumes in regard to the basic desiccant dehumidification process.

Figure 4.2.2: Liquid desiccant packed column – process vessel ‐ These units are like air washers, except they

spray liquid desiccant into the process air instead of simply water. The heat and moisture from the

dehumidification process is transferred to the desiccant. Heat is rejected through an external cooling system

and moisture is rejected in the desiccant regenerator, which re‐concentrates the diluted desiccant solution.

(Harriman, 2002)

As the desiccant returns from the regenerator to the sump on the dehumidification vessel, it is dry but

still hot, which results in a high vapor pressure. The desiccant is circulated through a cooler, which is

connected to a chilling system or cooling tower. Downstream of the cooler, the vapor pressure of the

desiccant is lowered since it is both dry and cool and is ready to absorb more moisture. Among the

advantages of the packed column is good control of heating and cooling. As an added benefit,

independent of the thermal treatment, the liquid desiccant can eliminate particles from the supply air

flow.

Disadvantages of the system include process inertia, which makes adaptation of rapidly changing

humidity conditions difficult. Larger quantities of desiccant solution may be distributed throughout a

large piping system with large reserve sump. This can result in significant time to respond to fast

changing internal moisture loads or different necessary outlet conditions. The slow response on the




July 9, 2016 Page 36 of 171

outlet conditions, however, can also be a benefit since to creates “process inertia” which means slow

responses to inlet changes. A large mass of recirculating desiccant is used to protect the internal process

from rapid changes in weather moisture.

Typically, liquid spray dehumidifiers are employed in large, central systems rather than small, free‐

standing units. The main reason is the somewhat higher complex of the liquid spray dehumidifier in

comparison to solid desiccant units. Some large systems connect several dehumidification process

vessels to a single regenerator reactor. This can have advantages of first cost, at the expense of

complexity of controls. Maintenance varies between the type of liquid desiccants used, since some

desiccants solutions are more corrosive than others.

Solid desiccant packed tower: The solid desiccant packed tower dehumidifier system consists of pairs of

process vessels which are filled with solid granular desiccant material, such as silica gel or molecular

sieve. The pairs of vessels operate in a batch wise, or process swing mode.

Figure 4.2.3 illustrates the working principles. One process vessel operates in dehumidification mode

while the other regenerates the desiccant material after it becomes saturated with moisture. In Figure

4.2.3, the two vessels are identified as process vessels A and B. In dehumidification mode, moist process

air starts to pass through process vessel A. The desiccant material in vessel A is cool and is effective to

adsorb humidity. The moisture content and temperature of the desiccant increases as moist process air

flows through vessel A. At the same time when moist process air flows through vessel A, vessel B is in

desorption mode. Reactivation air, heated by the so‐called desiccant heater, is passed through vessel B

driving out the moisture from the desiccant and transferring the moisture to the reactivation air. After

the desiccant in vessel A becomes saturated and/or the desiccant in vessel B is fully regenerated, the 4‐

way valves switch the air flow and the functions of vessels A and B are reversed.

Since drying and reactivation take place in separate, sealed compartments, the packed tower

dehumidifier is frequently used to dry pressurized process gases. The process can achieve very low dew

points, as low as below ‐40°F. Desiccant dehumidifiers for compressed air are frequently the packed‐

tower type.

Since the system operates batch wise and not continuous, conditions of the process air change during

one cycle of the swing operation. At the start of dehumidification, dry and cool desiccant is exposed to

the process airstream. Since the vapor pressure is at a maximum, the process air can be deeply dried. As

its moisture capacity fills up, the air is not dried as efficiently. In order to limit the changes in the

desiccant, controls are provided to ensure the process vessels are changed before the process air

condition become too wet. The air flow velocities inside the process vessels are kept low in order to

avoid uneven air flow distribution and the disturbance of the granulated desiccants. As a result, the solid

packed tower units can be quite large.




July 9, 2016 Page 37 of 171

Figure 4.2.3: Solid packed tower dehumidifier (Harriman, 2002) with annotations added

Moist air is dried in pairs of process vessel, which operate batch wise. The process vessels are filled with

granulated solid desiccant. The two process vessel alternate as between dehumidification and desiccant

regeneration mode. The desiccant is regenerated by hot airstream. The system is used frequently for

compressed air and pressurized process gasses. It is less common in ambient‐pressure applications.

Rotating horizontal bed

The rotating horizontal bed dehumidifier consists of granular desiccant that are held in a series of

shallow, perforated trays which rotate continuously between sections the process and reactivation

airstreams. In dehumidification mode, trays rotate through a section where moist process air passes

vertically through the trays. The desiccant adsorbs moisture. The trays then rotate into a position where

reactivation airstream heats the desiccant, raising its vapor pressure and releasing the moisture into the

reactivation air. The hot desiccant material is cooled by cold air flowing through the trays before the tray

reaches the section where the desiccant is again exposed to the moist process air. Figure 4.2.4 illustrates

a typical system configuration.




July 9, 2016 Page 38 of 171

The system configuration can operate in modular fashion in order to increase drying capacity. Increased

capacity can also be obtained by increasing the tray diameter to hold more desiccants or the number of

beds stacked on top of one another is increased. The rotating horizontal bed can provide constant outlet

moisture level.

A disadvantage can be an uneven settlement and resulting distribution of the granulated desiccant

material, which increases the possibility of uneven flow and leakage of air flow through the beds. A

typical remedy is to direct the process and reactivation airflow in parallel, rather than generating a

counter flow. With regard to energy efficiency, a parallel arrangement of process and reactivation

airflows does not perform as well as a counter flow arrangement. As a result, reactivation energy

consumption can be higher than compared to other designs. A significant benefit of rotating horizontal

bed systems are the low first costs. The design is simple, expandable and easy to manufacture.

Operating costs of the system can, however, be higher than with other designs, especially when energy

costs are high.

Figure 4.2.4: Rotating horizontal bed dehumidifier ‐ (Harriman, 2002)

Horizontal solid desiccant material is slowly rotated between process and reactivation airstreams. Leakage

between moist and dry airstreams can affect the performance.




July 9, 2016 Page 39 of 171

Multiple vertical bed: This dehumidification system has solid desiccant sitting in a vertical bed

configuration. The air streams, either the process or activation air streams, pass vertically through the

relatively high cells. This vertical bed configuration combines feature of a packed tower and a rotating

horizontal bed design in an arrangement that is well‐suited to atmospheric pressure dehumidification

applications. Figure 4.2.5 illustrates the assembly.

Figure 4.2.5: Multiple vertical bed dehumidifier ‐ (Harriman, 2002) ‐ Granular desiccant beds are installed vertically on a rotating assembly. Multiple cell like segments of the circular assemble pass through process and activation air streams. The system combines benefits static packed towers and rotating tray dehumidifiers. The design includes more complex parts due counter flowing process and reactivation air streams, which offer more energy efficient operation. The increased installation cost is offset by a lower operating cost than either packed tower or rotating horizontal bed type units

Rather than single or double packed tower systems which operate batch‐wise, the multiple vertical beds

are installed on a circular carousel with eight or more towers. These towers are exposed to alternative

process and reactivation air streams. This design can achieve low dew points since leakage between

process and reactivation air circuits is minimized. Due to separation between adjacent towers, process

and reactivation airstreams can be arranged in the more efficient counter flow pattern for better heat

and mass transfer. Different from the stationary packed tower with discrete process steps (e.g. swing

operation), the rotating multiple vertical bed assembly operates more or less continuously and can

provide a relatively constant outlet moisture condition on the process air. This can be a significant

advantage to the undulating performance of the packed tower units.




July 9, 2016 Page 40 of 171

In comparison to the horizontal bed dehumidifier, the system is mechanically more complex, has a

higher first cost and requires more complex operation. Generally, however, these are minor limitations

compared to the large savings in energy and performance improvements at low dew points.

Rotating Honeycomb or Solid Desiccant Dehumidification Wheel: The dehumidifier design uses a

rotating wheel which hold structures of desiccant material about a horizontal axis. This assembly is often

referred to as either “desiccant wheel” or rotating honeycomb. The wheel is a light weight structure

with corrugated and desiccant impregnated material. The structure allows ready horizontal passes of

the air streams through the slowly rotating wheel, resulting in low pressure losses. Figure 4.2.6 shows a

typical corrugated assembly, where the desiccant material adheres to the inner surfaces of the

horizontal small cells.

Figure 4.2.6: Honeycomb Structure of the Solid Desiccant Dehumidification Wheel , (Carnegie Mellon University, 2016)

Desiccant material covers a lightweight, open structure through which air can flow with small pressure losses.

Figure 4.2.7 illustrates the working principle of the desiccant wheel. Process air flows through the

horizontal cells of the corrugations in a major section of the desiccant wheel (or process) section. The

desiccant adsorbs moisture from the moist air stream until it becomes saturated (attains process point

(2)). As the wheel turns further, the desiccant material enters a section of the wheel that is exposed to

the reactivation section of the wheel enclosure, where a heated reactivation air stream dries out the

desiccant material in a counter flow process (attain process point (3)). After rotating through the

reactivation section the desiccants in the corrugated wheel structure are cooled by small portion of the

process air to lower their surface vapor pressure and bring the desiccant material to point (1). From

here, a new process cycle continues.

The rotating desiccant wheel design has several advantages over other dehumidifier configuration. The

structure is lightweight and offers an effective moisture removal capacity to overall weight ratio. As a

result, mechanical power to drive the system can be kept low. The corrugated wheel structure offers

few air flow obstructions and therefore pressure losses and flow resistance remain low. All of these

performance parameters result in good energy performance and relatively low maintenance efforts and

costs.




July 9, 2016 Page 41 of 171

Figure 4.2.7: Rotating Honeycombe® or Solid Desiccant Dehumidification Wheel ‐ (Harriman, 2002) ‐ The

design combines high surface area with low total desiccant mass, making these units especially efficient. The

small number of parts reduces maintenance to a minimum.

4.3 Comparing Desiccant to Cooling‐based Humidification

The selection of the desiccant dehumidifier for a specific application requires a careful optimization of

the entire system installation and operation. Harriman (2002) suggested that the following criteria have

to be addressed.

Installed cost: The first cost of the dehumidifier itself typically is only a small fraction of the cost

of the entire installation, which includes heating and cooling systems and overall site

infrastructure, including piping, electricity and other process components

O&M cost: Operating and maintenance costs are usually significantly larger than the first costs,

since dehumidification is an energy intensive process. The leading O&M cost items are heat for

reactivation and cooling for the desiccant and process air. Harriman (2002) suggested that

available low‐cost energy sources, such as solar or waste heat can often offset installed cost

differences in a relatively short time period.

Demonstrated operational reliability: Dehumidifier configurations differ between applications and

weather conditions. The installation of the reliable equipment for the condition is of importance.




July 9, 2016 Page 42 of 171

Design assumptions: The selection of the type and size of dehumidification equipment

configuration is dependent on the design assumption. Reasons for varying equipment selections

could include incomplete or erroneous data available to the system designer. It impossible to

state universally accurate comparisons because the working conditions of the desiccant

dehumidification systems can differ between applications.

With regard to selecting between desiccant and cooling‐based dehumidification, Harriman (2002)

suggested the following additional design considerations:

Cooling and desiccant‐based dehumidification technologies should complement each other and

costs can be lowered by using integrated systems.

The selection about using desiccant or cooling‐based dehumidification depends on cost of

electrical power and thermal energy. When using solar or waste heat and high electricity costs,

desiccant systems outperform cooling based systems. If electricity is inexpensive and activation

thermal energy costly, cooling‐based dehumidification can outperform desiccant systems.

Conventional design recommends cooling‐based dehumidification at high air temperatures and

moisture levels. At low target dew point, e.g. below a 40F dew point, the cooling‐based

dehumidification process becomes inoperative because condensate freezes on the coil. In these

cases, only desiccant dehumidification is the viable option.

Desiccants have advantages when treating ventilation air for building HVAC systems which also

use ice storage. In these cases, desiccants can offer significant cost savings.

Figure 4.3.1 presents data that compares first and O&M cost of various desiccant and cooling‐based

dehumidification technologies as a function of target dew point of dried air.




July 9, 2016 Page 43 of 171

Figure 4.3.1: First and

O&M costs of various

desiccant and cooling‐

based dehumidification

technologies as compared

to a Honeycombe®

desiccant wheel

dehumidifier

(Harriman,2002)


SECTION 5 ‐ PHYSICAL PROPERTIES OF DESICCANTS


July 9, 2016 Page 44 of 171


This section discusses important physical properties of desiccants, which affect the underlying processes

and the performance of desiccant dehumidification technology.

5.1 Basic Considerations of Physical Properties in Desiccant Dehumidification

Xing (2000) suggested that predicting the performance of desiccant material is difficult, because physical

properties of desiccant material typically varies between vendors and is significant manufacturing‐

process dependent. The author suggested that even for the same type of desiccant, different

manufacturers have different property data, which are sometimes considered proprietary, since specific

dehumidification applications call for different desiccant.

Desiccant materials are hygroscopic substances that create significant affinity for water vapor. Several

authors pointed out that “technically speaking”, any material would qualify as a desiccant since they can

attract some amount of water from the air. In order to qualify for commercial use in space conditioning,

however, a desiccant must be able to hold much larger amounts of water; some commercial solid

desiccant materials can hold up to 50% of their weight in water Xing (2000).

Collier et al (1981) suggested the following desirable physical properties for any desiccant material used

in desiccant cooling applications are:

1. Mechanical and Chemical Stability: Solid desiccant material does not deliquesce (become liquid).

Desiccants should not undergo hysteresis when cycled.

2. Large Moisture Capacity per Unit Weight: It is desirable to cycle as much water as possible for a

given amount of desiccant. This reduces the amount of desiccant required and the size of the

cooling system.

3. Large Adsorption/ Absorption Capacity at Low Water Vapor Pressures: The moisture capacity

should not deteriorate at very low water vapor pressures. This increases the relative dryness

achievable, which will have a strong effect on reducing fan power requirements.

4. Low Heat of Adsorption/Absorption: A low heat of adsorption increases cooling capacity and COP.

5. Ideal Isotherm Shape: The isotherm shape should be close to the “Ideal” function of the

application. This lowers regeneration temperature and thus increases COP.




July 9, 2016 Page 45 of 171

5.2 Isotherms

An isotherm is defined as the relation between relative humidity of the air in equilibrium with the

desiccant particle and the water content of the particle at a specified temperature. Isotherms are

considered important physical properties of desiccant materials, to assess heat and moisture transfer

interactions during adsorption and desorption processes in various kinds of desiccant dehumidification

systems.

Using the example of silica gel Ramzy et al (2015) presented a procedure of developing isotherm

equations of desiccant material that can be used in predicting heat and mass transfer interactions. The

optimization of the desiccant bed design requires mathematical assessment of the heat and mass

transfer interactions in the desiccant bed. Ramzy et al (2015) pointed out that several investigators have

presented different procedures to identify the isotherm equations required for their theoretical

simulations. All of the isotherm equations are based on experimental data. Two examples of isotherm

expressions for silica gel are presented in the following:

Ng et al. (2001) found that silica gel water vapor isotherms can be estimated by a detailed Henry's law's

correlation for silica. The authors provide the following equation:

Where:

W = Gel water content [kgw/kgdry silica gel]

Ts = Temperature in oC

Pv = Partial pressure of vapor in the air [Pa],

R = Gas constant of air in [J/(kg*K)],

K10 and Qst are tabulated values for specific silica gel products.

San et al.(2002) presented the isotherms for water vapor‐silica gel in terms of absolute humidity and

particle temperature




July 9, 2016 Page 46 of 171

Where:

W = Gel water content [kgw/kgdry silica gel]

Ts = The temperature in oC

C = The air absolute humidity [g/m3]

Ramzy et al (2015) suggested that obtaining experimental data which can be used in developing

equations can represent a significant effort. The authors presented their own simple procedure to

identify the water vapor‐ silica gel isotherm, which requires less time effort and cost and provides

acceptable predictions for the isotherm equation.

In this procedure the water content of silica gel is identified by drying the silica gel sample with infrared

heat. The authors pointed out that although silica gel has a very high melting temperature (1,600 ºC), it

will lose its chemically bound water and hygroscopic properties if heated above 300 °C. By determining

the initial weight (before heating process) and the dry weight (after heating process) the initial water

content of the silica gel is obtained.

The representative isotherm curve was obtained by analyzing 20 samples with the same 10 % initial and

different final water content. The samples remained in the fluidized bed for 48 hours so that water

content changing adsorption and desorption processes could obtain final equilibrium.

The results were curve fitted which resulted in the following isotherm relationship:

[Ramzy et al (2015)]

The graph of this isotherm relationship is depicted in Figure 5.2.1.




July 9, 2016 Page 47 of 171

Figure 5.2.1: Relative humidity of inter particle air at different silica gel water content [Ramzy et al, 2015)]

Figure 5.2.2 shows a comparison between the simplified procedure proposed by Ramzy et al (2015) and

several other researchers. The results reported by Ramzy et al are labeled (7) in Figure 5.2.2.

Figure 5.2.2: Isotherms of water vapor‐ silica gel collected from literature along with the isotherm obtained from the present work (Ramzy et al,2015)




July 9, 2016 Page 48 of 171

Figure 5.2.3 shows the results of a sensitivity analysis comparing the theoretical isotherm equations with

the same experimental data. Figure 5.2.3 suggests that the simple procedure for developing the

equilibrium isotherm equation of water vapor and silica gel lies in the same range of the isotherms

obtained in literature.

Figure 5.2.3: Experimental and theoretical variations of exit air humidity ratio and temperature during adsorption process using different isotherm equations [Ramzy et al (2015)], Annotated

5.3 Types of Isotherm Models

Generally speaking, the adsorption process creates a film of the adsorbate on the surface of the

adsorbent. Adsorption is a surface phenomenon, and the adsorption process is controlled by surface

active energy. The basic premise of adsorption is bonding capacities of atoms at the surface of the

adsorbent. Atoms on the surface of the adsorbent are not wholly surrounded by other adsorbent atoms

and therefore can attract adsorbates. The exact nature of the bonding depends on the desiccant

material. The adsorption process is generally classified as electrostatic, chemical or physical sorption.

Foo and Hameed (2010) presented a comprehensive investigation into types of adsorption isotherm,

which they characterized as an “invaluable graph describing the phenomenon governing the retention

(or release) or mobility of a substance to a solid‐phase at a constant temperature and pH”. Adsorption

equilibrium (the ratio between the adsorbed amount with the remaining in the solution) is established

when an adsorbate containing phase has been contacted with the adsorbent for sufficient time to

achieve dynamic balance.




July 9, 2016 Page 49 of 171

Foo and Hameed (2010) reported that over the years, a wide variety of equilibrium isotherm models

have been formulated, included the often used isotherm models by Langmuir, Freundlich, Brunauer–

Emmett–Teller, Redlich–Peterson, Dubinin–Radushkevich, Temkin, Toth, Koble–Corrigan, Sips, Khan, Hill,

Flory–Huggins and Radke–Prausnitz. According to Foo and Hameed (2010) these models can be grouped

into three fundamental approaches, namely two parameter isotherm model, the three parameter

isotherm model and multilayer physisorption isotherms.

Each of the isotherm models describe different adsorption characteristics. In conjunction with desiccant

adsorption processes the Brunauer–Emmett–Teller (BET) isotherm model is widely used. The BET model

was developed to derive multilayer adsorption systems with relative pressure ranges from 0.05 to 0.30

MPa.

Brunauer et al (1938) categorized sorption isotherms depending upon their sorption mechanisms. The

sorption mechanisms are defined as types 1 through 5. Figure 5.3.1 through 5.3.4 show five types of

Brunauer isotherms. Mei et al (1992) presented some commercially available materials (Figure 5.3.5).

Generally speaking, molecular sieves are considered type 1 materials, silica gels are considered type 2

materials, and wool is considered a type 3 material. Hydratable salts vary between type 2 and type 3.

Mei et al (1992) suggested an “ideal” isotherm for desiccant cooling applications (Figure 5.3.6).

Figure 5.3.1: Brunauer isotherms of types Figure 1 and 3 compared to a linear sorption isotherm (Brunauer, 1938) reproduced

Figure 5.3.2: Brunauer type 2 isotherm (Brunauer, 1938) reproduced




July 9, 2016 Page 50 of 171

Figure 5.3.3: Brunauer type 4 isotherm; (Brunauer, 1938) reproduced

Figure 5.3.4: Brunauer type 5 isotherm; (Brunauer, 1938) reproduced

Figure 5.3.5: Adsorption isotherms of typical desiccants (Mei, 1992) modified

Figure 5.3.6: Ideal isotherm for desiccant cooling applications, (Mei, 1992) modified




July 9, 2016 Page 51 of 171

Pesaran (1993) reported on research in conjunction with a desiccant cooling ventilation cycle that

identified the desiccant isotherm shape (characterized by separation factor) which would result in the

optimum performance. The author applied different regeneration temperatures ranging from 65°C to

160°C to identify the corresponding optimum isotherm shape at each. The research used Brunauer

Types 1 and 3 isotherms because previous studies had shown their potential over other desiccant types

for the investigated application. The author described the relationship between moisture uptake and

relative humidity for Type I and III desiccants by the equation:

W* = RH*

R + RH* ‐ R • RH*

where:

R = Separation factor

RH* = Temperature adjusted relative humidity

W* = Moisture loading fraction, w /W max

The isotherm shapes of various Type 1and 3 desiccants used by the author are shown in Figure 5.3.7.

The Type 1 isotherms correspond to desiccants with a separation factor of R < 1, and Type 3 isotherms

correspond to desiccants with a separation factor of R > 1. Separation factors were varied between R =

0.05 (Type I moderate) and R = 1.5 (Type III moderate).

Collier (1981) discussed the characteristics of an “ideal” desiccant material, which he contended are

almost impossible to meet in practical applications. Collier suggested that the requirement of an "ideal"

desiccant material would be that adsorption/absorption capacity are independent of water vapor

pressure. The author points out that, "unfortunately", the heat of adsorption is not independent of the

isotherm shape but is defined by the Clausius‐Clapeyron relationship. This relationship defines that a

material with a weak pressure dependence must also have a weak temperature dependence in order to

keep the heat of adsorption low.

It is important to keep the heat of adsorption of the desiccant material reasonably low. Even though

most of this energy can be recaptured by the sensible‐heat exchanger, a high heat of adsorption leads to

the requirement of a high regeneration temperature.




July 9, 2016 Page 52 of 171

Figure 5.3.7: Normalized

Moisture Capacity for Various

Type I and III Isotherms used

for the determination of

adsorption in a desiccant

cooling ventilation cycle

Pesaran (1993)

The figure shows isotherms

for Type I and III desiccant

material with different

separation factors.

Figures 5.3.8 through 5.3.10 show the adsorption isotherms for natural zeolite, molecular sieve, and

silica gel. Silica gel shows considerable vapor pressure dependence, whereas natural zeolite and

molecular sieve show much smaller water vapor pressure dependence. Natural zeolite and molecular

sieve must undergo very large temperature variations in order to experience minimal changes in water

content. Silica gel, on the other hand, can experience large variations in water uptake for much smaller

temperature swings, because of the relatively low heat of adsorption of water. In actual practice,

molecular sieve, natural zeolite, and silica gel produce about the same net cooling effect per unit weight

of material with nearly the same adsorption energy.

What silica gel loses in its pressure dependence, it seems to gain in its temperature dependence. The

temperature insensitivity of molecular sieve and natural zeolites is advantageous during the

dehumidification process; it is a disadvantage during the regeneration process. For silica gel, the

temperature sensitivity is a disadvantage during dehumidification but an advantage during

regeneration.




July 9, 2016 Page 53 of 171

Figure 5.3.8: Adsorption Isotherms for

Natural Zeolite; Collier (1981)

Figure 5.3.9: Adsorption Isotherms for

Molecular Sieve; Collier (1981)




July 9, 2016 Page 54 of 171

Figure 5.3.10: Adsorption Isotherms

for Silica Gel; Collier (1981)

5.4 Adsorption Energy

The adsorption energy is referred to as the heat which is liberated or absorbed during the adsorption or

desorption process, respectively. Mei et al (1992) suggested that it is very important to note that

temperature dependence of the isotherm shape and the adsorption energy are related. Therefore, the

shape of the isotherm will be uniquely determined by the heat of adsorption at that temperature and

loading. The author pointed out that the physics that ties these properties together is defined by the

Clausius‐Clapeyron relationship.

During the adsorption processes the heat of adsorption is not constant over the entire range of

desiccant loadings. The first molecules that attach themselves to the surface of the adsorbent are

bonded by the strongest forces available to the system and they have the highest adsorption energies

(strongest bonds) associated with them. As the process of molecules accumulating at the surface of the

desiccant progresses, the available sites for bonding fill and the desiccant loading increases. This results

in less energetic sites being occupied. As the desiccant approaches saturation conditions, the binding

energies approach energy levels of condensation, or evaporation. Therefore, heat of adsorption will be

highest when loading is zero and will be lowest at saturation loading. The rate at which this energy

changes with water loading will vary depending upon the material.




July 9, 2016 Page 55 of 171

5.5 Chemical and Physical Stability

As specified in the literature, desiccant material can suffer from performance losses if it is not stable

under certain process conditions. Mei et al (1992) presented two important aspects of physical and

chemical stability that deserve special attention. These are (1) the ability of the desiccant material to

withstand high regeneration temperatures and (2) to be reliably regenerated by ambient‐source air.

Many commercial desiccants rely on molecular sieve structures which are considers as not stable at high

temperatures. Desiccant dehumidification cycles, however, require that desiccant sorption properties

should be stable at temperatures between 300 F and 400 F. Some desiccants are "fouled" by the

presence of certain molecules that have a greater affinity for the active sites than water vapor does.

Others change chemical composition in the presence of certain compounds that change the desiccants

water sorption characteristics.

5.6 Adsorption Rate and Performance

The adsorption process is dependent on a number of properties, such as isotherm shape, heat of

adsorption, heat capacity, moisture capacity, and on other operating conditions, including temperature

and method of regeneration.

Pesaran and Hoo (1993) investigated the characteristic shape of the desiccant isotherm as a function of

regeneration temperature and process which affect the performance for a solar desiccant cooling cycle

operating in ventilation mode. The authors used thermal coefficient of performance and cooling

capacity of the system as metrics to describe the process.

The authors suggested that solar heat derived from flat‐plate collectors was initially considered to be a

good low‐temperature heat source available for regeneration. However, lower regeneration

temperatures typically resulted in decreased cooling performance with several candidate systems

investigated. As a consequence, size and cost of the solar cooling system had to be increased. As a

remedy, high‐temperature regeneration desiccant materials and cycles were proposed to improved

thermodynamic performance at higher regeneration temperatures. The main objective of the research

reported by Pesaran and Hoo (1993) was the determination of how the performance of a solar desiccant

cooling system would improve when the regeneration temperature was increased.

The system used for the investigation resembled a desiccant cooling ventilation cycle with a rotary

desiccant dehumidifier, a heat exchanger, two evaporative coolers, a desiccant regeneration heater, and

ancillary equipment such as fans and pumps. Figure 5.6.1 illustrates the system schematic where outside

air was dried in the dehumidifier and then cooled by regenerative evaporative coolers and supplied to

the conditioned space. The regeneration heater (powered by solar energy) heated the air, which

reactivated the desiccant by driving the moisture from it.




July 9, 2016 Page 56 of 171

Figure 5.6.1: Schematic Diagram of a Desiccant Cooling System; Pesaran and Hoo (1993)

Based on previous investigations, Pesaran and Hoo (1993) used Brunauer Types 1 and 3 desiccant

isotherms because these desiccant characteristics had shown better performance than other desiccant

types. In addition, in order to investigating the effects of regeneration temperatures and dynamics two

regeneration methods were used, referred to as “staged” and “no‐staged” regeneration methods. The

authors revealed that the differences of the staged and no‐staged methods had a significant effect on

the mass and heat transfer performance.

Figure 5.6.2 shows a schematic of the staged and no‐staged regeneration schemes. The two methods

differ in their regeneration process configuration. In the staged regeneration method, only a fraction of

the regeneration airstream is heated to the design regeneration temperature. The remainder of the

regeneration airstream is composed of warm air returning from the conditioned space. A portion of the

heat of adsorption is added to the return air via heat exchange between the downstream dehumidifier

flow and the upstream regeneration flow. Under the staged method, only a small portion of the

regenerated area of the rotary desiccant bed (30% of the desiccant exposed to the high temperature

airstream), at the end of the regeneration cycle, is exposed to high temperature airstream from the

regeneration heater. Thus makes up about 30% of the total regeneration airstream. The 30% value was

selected based on the previous experiments published by Collier (1989). Collier had suggested that the

maximum thermal COP is achieved with a separation factor of R = 0. 1 and a regeneration temperature

of 95°C.

The results of the no‐staged regeneration are shown in Figure 5.6.3 and 5.6.4. Figure 5.6.3 correlates the

cooling capacity and coefficient of performance (COP) with the regeneration temperature. Figure 5.6.4

indicates that thermal COP performance improves with lower regeneration temperature, with the

maximum COP value encountered at the lowest temperature investigated. This performance is due to

the fact that under the no‐staged method the entire regeneration area is subjected to the regeneration

temperature and the desiccant material is exposed to the drying airstream for a much longer period




July 9, 2016 Page 57 of 171

than necessary. This causes desiccant over‐regeneration and a drier and hotter desiccant bed at the

start of the dehumidification cycle. While this increases the cooling capacity, it results in a lower thermal

COP because of the inefficient use of the thermal energy in regeneration. This process especially applies

to high regeneration temperatures as over‐regeneration becomes more pronounced and the COP

decreases accordingly. On the opposite side, at low regeneration temperatures the over‐regeneration of

the desiccant is less pronounced and one could would expect the thermal COP to improve.

The Pesaran and Hoo (1993) pointed out that if cooling capacity and size of equipment is of higher

concern than the cost of thermal energy the no‐staged regeneration is a good solution. The cost of

thermal energy in a solar desiccant cooling system, however, is a major concern and has a significant

impact on the overall size of the system. Under these circumstances a tradeoff between cooling capacity

per unit size versus equipment size and cost of the solar collectors must be closely examined.

Figure 5.3.4 shows that the associated separation factors varies between Type I moderates (R=0.1) and

nearly linear isotherms (R = 1). For no‐staged regeneration, the two performance criteria of thermal COP

and system cooling capacity do not correspond to the same value of R. Maximizing cooling capacity

favors a Type I moderate isotherm, while emphasizing COP performance favors a more linear isotherm.

Figure 5.6.2: Schematic Diagram of Staged Regeneration and No‐staged Regeneration Schemes, (Pesaran & Hoo,1993)




July 9, 2016 Page 58 of 171

Figure 5.6.3: Maximum Cooling Capacity Performance with Corresponding COP, No‐staged Regeneration (Pesaran & Hoo,1993)

Figure 5.6.4: Optimum Separation Factors Resulting in Maximum Thermal COP and Maximum Cooling Capacity Performance, No‐staged Regeneration (Pesaran & Hoo,1993)

The results of the staged regeneration are shown in Figure 5.6.5 and 5.6.6. Figure 5.6.5 correlates the

cooling capacity and Coefficient of performance (COP) with the regeneration temperature. Figure 5.6.6

suggests that the maximum thermal efficiency and maximum cooling capacity are reached at

regeneration temperatures of 110 °C and 160 °C, respectively. Figure 5.6.6 shows that for staged

regeneration, optimum performance of the system in terms of thermal COP seems to favor two

different values of R, one for low‐temperature applications and a second for high‐temperature

applications. A similar trend can be observed for cooling capacity.

Pesaran and Hoo (1993) concluded that No‐staged regeneration, while inherently simpler than a staged

regeneration method, requires significant compromises in the performance of the overall system.

Maximum thermal COP and maximum cooling capacity performance were achieved at completely

different separation factors, especially at regeneration temperatures above 80°C. In addition, the

regeneration temperature has opposing effects on thermal COP and cooling capacity, increasing one

while decreasing the other. As a result, significant tradeoffs must generally be accepted when optimizing

for either thermal COP or cooling capacity. On the other hand, Pesaran and Hoo (1993) concluded that

staged regeneration, while adding complexity to the desiccant cooling system, provides an excellent

combination of cooling capacity and thermal COP performance. At high regeneration temperatures, a

low separation factor maximizes both COP and cooling capacity. At low regeneration temperatures, the

value of R resulting in maximum COP performance.




July 9, 2016 Page 59 of 171

The authors suggested that staged regeneration, when used in high‐temperature applications (T >

120°C), generally reduces the size and cost of a solar desiccant cooling systems. For most low‐

temperature applications (T < 80°C), it would appear that no‐staged regeneration is preferred.

Figure 5.6.5: Maximum Cooling Capacity Performance with Corresponding COP, Staged Regeneration (Pesaran & Hoo,1993)

Figure 5.6.6: Optimum Separation Factors Resulting in Maximum Thermal COP and Maximum Cooling Capacity Performance, Staged Regeneration (Pesaran & Hoo,1993)

5.7 Basic Heat and Mass Transfer Modelling by Example of Desiccant Wheel

Ruivo, C. et al (2011) suggested that desiccant dehumidification lowers O&M costs through a reduction

in peak electricity demand and associated electricity infrastructure costs. The drawback is the higher

initial cost compared with equivalent conventional systems. The authors identified opportunities of cost

reduction at the design stage through careful cycle selection, flow optimization and size reduction.

Designers typically can use only data given by the manufacturers of desiccants, which are usually

restricted to particular sets of operating conditions. This design data is based on experimental or

numerical approaches which by‐and‐large provide tools to assess steady‐state, but no dynamic heat and

mass transfer performance of the desiccant applications. The authors pointed out the need to establish

numerical models to describe the dynamic aspect of heat and mass transfer in desiccant applications.

Ruivo, C. et al (2011) described the desiccant wheel as a compact and mechanically resistant heat and

mass transfer system composed of a high number of channels with porous desiccant walls. The

hygroscopic matrix is submitted to a cyclic sequence of adsorption and desorption of water molecules.




July 9, 2016 Page 60 of 171

The regeneration process of the matrix (desorption) is imposed by a hot airflow. In each channel of the

matrix, a set of physical phenomena occurs, which include heat and mass convection on the air side and

heat and mass diffusion and water sorption processes in the desiccant lined wall.

Figure 5.7.1 shows the schematics of the desiccant wheel system that was used to define the dynamics

of the combined heat and transfer processes. The airflow 1 (process air) and airflow 2 (regeneration air)

cross the matrix in a counter‐flow configuration, with equal or different mass flow rates. The desorption

zone is generally equal or smaller than the adsorption zone.

(a) Desiccant wheel with porous structure of the matrix

(b) Desiccant wheel, definition of adsorption and desorption zones

Figure 5.7.1: Desiccant wheel system, (Ruivo, C. et al, 2011), annotated

Ruivo, C. et al (2011) pointed out that approaching airflows in each zone can present instabilities and

heterogeneities and are generally turbulent. The flow conditions inside the matrix, however, are laminar

mainly because of low values of hydraulic diameter of the channels. In very short matrixes with larger

hydraulic diameters of the channels, the entrance effects can be relevant. During the adsorption/

desorption processes there are non‐uniform distributions of adsorbed water content and temperature,

inside the matrix.

The authors provided examples of steady‐state and dynamic heat and mass transfer processes. Figure

5.7.2 (a) and (b) illustrate the differences between steady‐state and dynamic performance of sorption

observed in a desiccant wheel. The annotations depict the psychrometric evolutions in both air flows 1

and 2. Figure 5.7.2 (a) describes the steady‐state conditions of the airflow and (b) the dynamic

conditions. Figure 5.7.2 (b) indicates the influence of the channel length and of the cycle duration on the

psychrometric evolutions. The authors suggested that for each channel length there is an optimum




July 9, 2016 Page 61 of 171

value of the cycle duration, i.e. the optimum rotational speed that maximizes the dehumidification rate.

The effect of optimal rotation diminishes with increasing channel length.

(a) Steady‐state conditions (b) Dynamic conditions, with Influence of the channel length and of the cycle duration

Figure 5.7.2: Psychrometric evolutions of the airflows in a desiccant wheel (Ruivo, C. et al, 2011), annotated

The authors utilized a numerical model of the channel flow with a set of conservation equations. Fig.

5.7.3 illustrates the physical domain of the channel model used in the investigations. Two phases co‐

exist in equilibrium inside the desiccant porous medium, the equilibrium being characterized by sorption

isotherms. Using a simplification, the authors used two mechanisms of mass transport, surface diffusion

of adsorbed water and Knudsen (mean free path) diffusion of water vapor, with the wall considered to

be a homogeneous desiccant porous material. Other boundaries of the domain were considered

impermeable and adiabatic.




July 9, 2016 Page 62 of 171

Figure 5.7.3: Physical domain of the modelled channel (Ruivo, C. et al, 2011) Here Hc is hydraulic radius of the channel, x = distance from channel entrance.

The authors reported on the dynamic heat and mass transfer responses in the desiccant matrix. Figure

5.7.4 (a) and (b) show heat and mass transfer characteristic of the bulk air flow inside a channel of the

desiccant matrix. The data plotted in Figure (a) and (b) illustrate heat and mass transfer behavior of a

desiccant wheel with a channel length of 0.3 m and a cycle time of 500 sec. The x‐values in the figure

describe the distance from the channel entrance. The desiccant wheel used was divided into two equal

parts, the adsorption and desorption zones with counter‐current airflows.

The desiccant material used was silica gel. The inlet temperatures of the process and regeneration

airflows were 30 °C and 100 °C, respectively. For the analysis the two airflows had the same inlet water

vapor content (0.01 kg kg–1) and mass flow rate. Figure 5.7.4 (a) and (b) show the dependencies of heat

and mass flow on the time step in the process, respectively. Figure 5.7.4 (a) and (b) suggest that the

airflow and the desiccant lined wall were close to thermodynamic equilibrium throughout most of the

wheel cycle in most of the rotor domain. This suggests that it is possible to optimize the

dehumidification performance of the rotor through a selection of rotation speed and of the extent of

the adsorption and desorption zones.

Other cases with different cycle durations were simulated. The registered influence of cycle τ on the

global heat and mass transfer rates is shown in Fig. 5.7.5.

Ruivo, C. et al (2011) investigated the internal resistance to heat and mass transfer of the desiccant

matrix. The authors pointed out the shortcomings of simplified methods describing air stream processes

and related interaction with the desiccant medium by assuming a fictitious bulk flow pattern, as well as

fictitious heat and mass convection coefficients for the gas side.




July 9, 2016 Page 63 of 171

Figure 5.7.4 (a): Heat

transfer characteristic

inside a channel of the

desiccant matrix.

Ti = Temperature at the

matrix interface

Tf = Temperature of the

bulk airflow

(Ruivo, C. et al, 2011)

Figure 5.7.4 (b): Mass

transfer characteristic

inside a channel of the

desiccant matrix.

Wi = Water ratio at the

matrix interface

Wf = Water ratio of the

bulk airflow


Figure 5.7.6 illustrates the physical model used to describe heat and mass transfer inside a desiccant

wall element, which is reduced to an element of the channel wall consisting of a homogeneous

desiccant medium. The heat and mass transfer phenomena inside the porous medium are considered




July 9, 2016 Page 64 of 171

only in the lateral direction (e.g. y – direction). For simplification, the bulk air flow was considered fully

mixed, which means it was characterized by constant and uniform properties.

Figure 5.7.5:

Heat and mass transfer rates

per unit of transfer area of

the desiccant wheel as a

function of the wheel cycle

time


Figure 5.7.6: Schematic

representation of the

channel wall element

HP = Thickness of

channel wall element.


Figure 5.7.7 (a) and (b) show time‐varying profiles of dependent variables along the desorption process

in a desiccant layer of HP = 1 mm, for temperature and adsorbed water content, respectively. Figs. 5.7.7

(a) and (b) depict results obtained with the “normal” internal resistances. Figure 5.7.7 (a) shows small

temperature gradients, which indicate that the internal resistance to heat diffusion is almost

insignificant. Figure 5.7.7 (b), on the other hand shows significant mass diffusion gradients. The

gradients of the adsorbed water content are significant during almost all the transient process, mainly

near the convective surface.




July 9, 2016 Page 65 of 171

Figure 5.7.7 (a): Time‐varying

profiles of temperature in a

desiccant layer of Hp = 1 mm


Figure 5.7.7 (b): Time‐varying

profiles of adsorbed water

content X` in a desiccant

layer of Hp = 1 mm


In order to account for varying assessments of heat and mass transfer coefficients Ruivo, C. et al (2011)

introduced two additional scenarios. These are referred to as “A” = negligible internal resistances to the

heat and mass diffusion (approach A‐“null resistances”) and “B” = thermal lumped capacitance method

(approach B‐“null thermal resistance”). Figure 5.7.8 (a) and (b) suggest that significant inaccuracies may

result when the internal resistance to mass diffusion is neglected, leading to unrealistic estimation of

the convection fluxes at the interface.




July 9, 2016 Page 66 of 171

Figure 5.7.8 (a): Predicted

temperature at the convective

interface, Hp = thickness of

desiccant film


Figure 5.7.8 (b): Predicted

adsorbed water content at the

convective interface,

Hp = thickness of desiccant film


5.8 Heat and Mass Transfer Characteristics of Desiccant Polymers

Staton (1998) discussed heat and mass transfer characteristics of advanced desiccant material in

conjunction with air conditioning applications. The author described process characteristics of total

energy exchangers, where the heat and mass transfer processes occur in a two phase process, and at

different rates.




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As an illustration, a limiting process factor could be the desiccant moisture uptake rate, if a desiccant

material has a slow moisture uptake rate. As a consequence, an energy transfer wheel would have to

rotate much slower to transfer moisture than heat. The author indicates that the difference in wheel

rotation could be significant, with a desiccant wheel and a heat transfer wheel spinning at typical speeds

of 6‐10 rph and 60 rpm, respectively. As a consequence, slow adsorption and desorption transfer rates

of the desiccant material used in an air conditioning systems could necessitate installing two phase

processing wheels for latent and sensible load reduction.

On the other hand, the use of a desiccant material with an increased rate of absorption and desorption

and a low desorption temperature could enable the designer to incorporate the sensible heat exchanger

component with the desiccant dehumidifier to produce a total enthalpy exchanger. The author pointed

out that total enthalpy exchanger introduces possibilities for increased energy transfer efficiency,

reduced material costs, a decrease in the number of components required to properly condition the air,

and a decline in the demand for refrigerant‐based components within the conditioner. Figures 5.8.1 (a)

and (b) illustrate the one and two step conditioning process of the supply and discharge air in air

conditioning systems, respectively.

(a) One Step Air Conditioning Process (b) Two Step Air Conditioning Process

Figure 5.8.1: Schematic of one and two step air conditioning process, Staton (1998)

Staton (1998) suggests the use of polymer desiccants as an alternative to traditional desiccants within

energy exchanger applications, since polymers have reported higher sorption rates and require lower

regeneration temperatures. The author refers to five desirable characteristics for these new desiccant

materials:

1. Mechanical and chemical stability

2. Large maximum moisture capacity

3. High sorption rates at low vapor pressures

4. Low heats of adsorption

5. Ideal Isotherm shape




July 9, 2016 Page 68 of 171

Polymers are large molecules consisting of repeating chemical units, or mers. One polymer molecule, or

’macro‐molecule’, can consist of 50 to 3000 mers. Polymers are capable of being cross‐linked, which is

the condition of the polymer chains being bonded together to form a network. The bonds responsible

for cross‐linking are covalent bonds, which will neither dissolve when put in solution nor flow when

heated. Super‐Absorbent Polymers (SAP) are hydrogels, water‐insoluble hydrophilic homopolymers or

copolymers, which are able to swell and absorb 10 to 1000 times their own weight in water. Super‐

Absorbent Polymers are crosslinked in order to avoid dissolution. There are three main classes of SAPs:

1. Cross‐linked polyacrylates and polyacrylamides

2. Cellulose‐ or starch‐acrylonitrile graft copolymers

3. Cross‐linked maleic anhydride copolymers

The most important property of hydrogels is their high affinity for water. The capability of hydrogels to

attract water can be attributed to the presence of groups like hydroxyl groups which are highly water‐

soluble. Even when saturated, hydrogels can remain insoluble and structurally stable due to the three‐

dimensional cross‐linked network.

Staton (1998) identified five candidate polymer desiccant materials and conducted scoping tests to

identify the best polymer desiccant for further studies. The five polymer desiccant materials and two

traditional desiccant materials were as follows:

Cross‐linked polyacrylamide copolymer: Typically used as a soil additive to keep moisture content

more constant. Known to maintain strength when swollen and readily absorbs and desorbs

water.

Polyacrylate: Acrylic acid cross‐linked with sodium acetate; typically used in hygiene products.

Known for selective absorbency as a result of the pH of the water.

Polyvinyl alcohol: Porous open‐cell sheet foam capable of absorbing up to twelve times its weight in

water. Also available in powder and liquid forms. Typically used as sponges in cosmetic and

industrial applications.

Polyethylene glycol: Copolymer exhibiting LCST characteristics with water in lower molecular

weights. Lower molecular weight materials are in liquid form. These can be used as thickeners in

cosmetics and food products.

Polypropylene glycol: Copolymer exhibiting lower critical solution temperature (LCST) characteristics

with water in lower molecular weights. Typically used in food, cosmetic, personal care and

pharmaceutical applications.

Silica Gel: Ceramic desiccant popular for use in dehumidification applications. Used for comparison

with the polymer desiccants.

Calcium sulfate: Ceramic desiccant popular for use in dehumidification applications. Used for

comparison with the polymer desiccants.




July 9, 2016 Page 69 of 171

For the above mentioned desiccant materials experiments were conducted (Staton, 1998) to determine

each material’s ability to absorb water and to determine each materials maximum sorption capacity.

Similar experiments were also conducted to determine how quickly the material released the stored

moisture (desorption rates). Figures 5.8.2 and 5.8.3 show the results of these scoping tests. Table 5.8.1

indicates the equilibrium adsorption rates.

Figure 5.8.2: Candidate polymer desiccant adsorption rates; (Staton, 1998)

Table 5.8.1: Equilibrium moisture content values for the desiccant material candidates; (Staton, 1998)




July 9, 2016 Page 70 of 171

Figure 5.8.3: Candidate

polymer desiccant desorption

rates; (Staton, 1998)

The results of the tests suggested that polymer desiccant polyvinyl alcohol had the best process

performance, which included the fastest rate of adsorption and the fact that that desorption could be

accomplished at low regenerative temperatures. In addition, film and foams, the desiccant material

forms of interest, were proven safe to use in heat recovery ventilation application. Therefore, a more

detailed analysis was conducted with this desiccant material only.

Staton (1998) analyzed three different forms of polyvinyl alcohol coating applications for total energy

exchanger in a desiccant wheel and fixed plate total enthalpy exchanger. Figures 5.8.4 through 5.8.6

summarize some main results and comparisons of the desiccant exchanger used in the tests. The three

polyvinyl alcohol material coatings evaluated above were compared to a commercially available

desiccant dehumidifier manufactured by LaRoche Air Systems. The LaRoche desiccant material is a

ceramic coating on a fibrous paper substrate. Similar to the pure polyvinyl alcohol foam wheel structure,

the LaRoche paper wheel needs no supporting substrate, which means that can be made entirely of the

ceramic‐coated paper.




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Polyvinyl Alcohol Film‐Coated Rotary Wheel Total

Enthalpy Exchanger

Polyvinyl Alcohol Foam Coated Rotary Wheel Total

Enthalpy Exchanger

Polyvinyl Alcohol/Ceramic Composite Coated Rotary

Wheel Total Enthalpy Exchanger

Figure 5.8.4: Wheel Total Enthalpy Exchanger Efficiency vs. Channel Depth; (Staton (1998)




July 9, 2016 Page 72 of 171


Enthalpy Exchanger


Enthalpy Exchanger



Figure 5.8.5: Wheel Total Enthalpy Exchanger Efficiency vs. Desiccant Coating Thickness; (Staton (1998)

Figure 5.8.7 compares the efficiencies of the three polymer desiccant materials and the LaRoche coating

option. The performance of a fixed counter flow plate exchanger was determined for polyvinyl alcohol

foam coating and La Roche coating. The results are presented in Figure 5.8.8. The results suggest a

significantly higher efficiency of polyvinyl alcohol foam coating for the plat exchanger than for desiccant

wheel applications.




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Enthalpy Exchanger


Enthalpy Exchanger



Figure 5.8.6: Wheel Total Enthalpy Exchanger Efficiency vs. air stream velocity; (Staton (1998)




July 9, 2016 Page 74 of 171

Figure 5.8.7: Comparison

between different rotary

wheel total enthalpy

exchanger coatings;

(Staton ,1998)

Figure 5.8.8: Comparison

between fixed plate total

enthalpy exchanger desiccant

materials; (Staton ,1998)


SECTION 6 ‐ DESICCANT REACTIVATION ENERGY SOURCES


July 9, 2016 Page 75 of 171


Daou (2006) described the process of desiccant thermal regeneration as part of desiccant cooling. The

regeneration heat source supplies the thermal energy necessary for driving out the moisture that the

desiccant had taken up during the sorption phase. Several possible energy sources were considered.

Those included solar energy, waste heat, and natural gas heating, and the possibility of energy recovery

within the system. (Figure 6.1)

In the case of solid desiccant, the heat of regeneration is furnished by blowing the scavenger air stream

concurrently through the desiccant matrix. For a liquid desiccant system being used, the heat of

regeneration is furnished to the desiccant solution inside the regenerator process vessel where a

scavenger air stream is concurrently blown to carry away the moisture desorbed under the desiccant

heating. The scavenger air can also be a hot air stream brought into contact with the dilute desiccant

solution inside the regenerator thereby heating it.

Figure 6.1: Process schematics of desiccant cooling, (Daou 2006)

6.1 Desiccant Regeneration with Heat from Natural Gas or other Fossil Fuels

There are numerous desiccant applications where the regeneration heat is provided by fossil fuel. A

typical application is natural gas dehydration systems using desiccants remove water vapor from natural

gas during production, transmission, and distribution. Natural gas flowing from production wells or

underground storage requires dehydration to protect the distribution system from corrosion and

hydrate formation. In the commonly used solid desiccant absorption dehydration process, wet gas flows




July 9, 2016 Page 76 of 171

through a pressure vessel tower filled with solid desiccant. Natural gas is used to gently heating the

regeneration gas, and sending it to the wet desiccant tower to dry out the desiccant.

Farang (2011) conducted research on a natural gas drying process using molecular sieves, which is

considered as one of the most important desiccant materials in industrial natural gas dehydration. The

authors stated that for the regeneration of weak desiccants, such as molecular sieve, the higher the

regeneration temperature, the smaller are the required quantities of regeneration gas. In the

experiments, the drying operation resulted in dew points of minus 30 °C or lower (pressure 20 MPa).

With a regeneration temperature equal or close to the adsorption temperature, the quantity of purge

gas depended on its pressure. For example, when the regeneration pressure is increased from 0.1 to 0.5

MPa, the flow rate of purge gas is increased from 10% to 20%. Regeneration of the sorbent at a

temperature above the adsorption temperature was leading to a reduction of the purge gas quantity.

The authors identified a regeneration temperature as 200–230 °C.

Netušil and Ditl (2012) described natural gas dehydration method water by a solid desiccant. They

suggested molecular sieve, silica gel or alumina as the desiccant material of choice. The authors

identified recommended regeneration temperatures of 230, 240 and 290 °C for silica gel, alumina and

molecular sieves. Figure 6.1.1 shows a typical scheme of the temperature swing adsorption dehydration

process used by the authors. The amount of adsorbed water increases with the pressure in the process

vessel and decreases with its temperature. The swing adsorption dehydration columns always work

batch wise. A minimum of two vessels are used, where one vessel dries the gas while the other is being

regenerated.

This method is known as temperature swing adsorption (TSA). The TSA heater was an ordinary burner or

a shell and tube heat exchanger warmed by steam or other hot process fluid. The regeneration gas

warms in the heater and flows into the column, where it passes through the adsorbent and the water

desorbs into the regeneration gas. The water saturated regeneration gas then flowed into the cooler.

The cooler used cold air to decrease the temperature of the regeneration gas. When the water

saturated regeneration gas is cooled, partial condensation of the water occurs.

Kozubal et al (2012) introduced the new concept of Desiccant Enhanced Evaporative Air‐Conditioning

(DEVAP). The system is a compact AC‐system for residential and commercial applications. The

regeneration heat for the liquid desiccant can be supplied by a variety of heat sources including natural

gas, waste heat or even solar heat. Figure 6.1.2 show illustrations of a DEVAP used in a commercial

application. Person (1994) presented a two stage desiccant conditioner. Figure 6.1.3 illustrates the

process where the desiccant wheel is regenerated by a natural gas burner heating an air flow.




July 9, 2016 Page 77 of 171

Figure 6.1.1: Typical scheme of the temperature swing adsorption dehydration process, (Netušil and Ditl, 2012)

Figure 6.1.2: Process schematics of the DEVAP system, (Kozubal et al, 2012)

Figure 6.1.3: All‐Air VAV System

Retrofitted with a Desiccant

Preconditioning System, (Pesaran,

1994)




July 9, 2016 Page 78 of 171

6.2 Desiccant Regeneration with Waste Heat

Jalalzadeh‐Azar (2005) described desiccant cooling as an important part of the wide range of Thermally

Activated Technologies (TAT) which use conversion of heat for the purpose of indoor air quality control.

For regeneration of the desiccant, the vapor pressure differential is reversed in the regeneration process

that drives the moisture from the desiccant. A burner or a thermally compatible source of waste heat

can provide the required heat for regeneration.

When implemented in the context of combined heat and power (CHP), also known as cogeneration, TAT

can significantly improve fuel efficiency/utilization through heat recovery from the on‐site power

generators and leads to a significant reduction in emissions. Desiccant materials with regeneration

source temperatures typically ranging from about 160 F to 300 F are compatible with various types of

on‐site generators, including reciprocating IC engines, micro‐turbines, and certain types of fuel cells.

Cascading desiccant cooling with another compatible technology, such as absorption or adsorption

cooling, further enhances the overall system fuel efficiency and utilization. Cascading processes allows

sequential heat recovery from a single heat source for driving two or more thermally activated systems

with different operating temperatures to achieve higher efficiencies.

Jalalzadeh‐Azar (2005) described a CHP system which incorporated a 60‐kW micro‐turbine for onsite

power generation and a gas‐to‐liquid heat exchanger for heat recovery from the exhaust gas for space

heating and desiccant dehumidification, depending on the season. A gas compressor was added to the

systems to boost the natural gas pressure to the operating pressure of the micro turbine combustor.

Figure 6.2.1 illustrates the air‐handling unit (AHU) incorporating a desiccant wheel, a DX cooling coil, and

a heating coil. A glycol solution leaving the heat exchanger at about 180 F preheats the regeneration air,

which is further heated by a gas burner to about 275 F for regeneration of the desiccant. In the heating

season, the hot liquid is circulated through the space‐heating coil instead (Figure 6.2.2).

Monitoring of the system over a year revealed that the overall efficiency of the CHP system could

exceed 50% on humid summer days, compared to the corresponding net electrical efficiency of about

21% to 23%. These results emphasized the importance of TAT, without which the CHP systems cannot

attain high energy conversion efficiencies which usually justifies for higher first costs.




July 9, 2016 Page 79 of 171

CHP system, roof top installation

Schematic of air‐handling unit

Figure 6.2.1: CHP system with desiccant unit (Jalalzadeh‐Azar, 2005)

Kassem (2013) suggest that the use of heat to regenerate desiccant material in a drying system can

lower the potential of energy saving in cooling desiccant cooling applications. However, the use of low

energy or free available energy such as waste heat from industrial processes for regeneration of

desiccant material will make the system more cost‐effective.

Figure 6.2.2 shows the operational concept and diagram of a desiccant‐based ventilation and air‐

conditioning system. The temperature of the processed air from the desiccant dehumidifier increases

due to the heat of condensation and heat of sorption. Heat recovery devices are used to recover this

energy. Downstream of the heat recovery the air becomes warm and dry. Since the air is warmer than

required for space comfort temperature, an evaporative cooling process is applied by either direct

addition of air moisture or indirect addition of air moisture in secondary air stream. The




July 9, 2016 Page 80 of 171

application of evaporative cooling process reduces the air temperature with either slight increase of air

moisture content or constant air moisture content for the direct or indirect evaporative cooler,

respectively.

Figure 6.2.2: Operational

concept and diagram of the

desiccant‐based ventilation

and air‐conditioning system

(Kassem, 2013)

Kassem (2013) suggests that liquid desiccants are better suited for the use of waste heat than solid

desiccants. While the capacity to absorb moisture is generally greater in liquid than in solid desiccants,

liquid desiccants require lower regenerating temperature, mostly in the range of 40–70 °C while it is in

the range of 60–115 °C for a solid desiccant. Figure 6.2.3 shows an example of a liquid desiccant cooling

system which uses low grade regeneration heat, such as derived from waste heat.

In the depicted system (in Figure 6.2.3) concentrated solution is sprayed at point A over the cooling coil

at point B while supply air at point 1 is blown across the stream. The solution absorbs moisture from the

air and is simultaneously cooled down by the cooling coil. The results of this process are the cool dry air

at point 2 and the diluted solution at point C.




July 9, 2016 Page 81 of 171

Figure 6.2.3: A liquid

desiccant cooling system

using low grade heat for

desiccant regeneration,

(Kassem, 2013), annotated

An aftercooler reduces the air stream temperatures further down, if required. In the regenerator, the

diluted solution from the dehumidifier is sprayed over the heating coil at point E which derives its heat

supply from a waste heat source. Ambient air at point 4 is blown across the solution stream. Some water

is taken away from the diluted solution by the air while the solution is being heated by the heating coil.

The resulting concentrated solution is collected at point F and hot humid air is rejected to the ambient

at point 5. A recuperative heat exchanger preheats the cool and diluted solution from the dehumidifier

using the waste heat of the hot concentrated solution from the regenerator, resulting in a higher COP.

Another process application of waste heat in desiccant regeneration is the use waste heat from

condensers in vapor compression refrigeration systems. Misha (2012) suggested the system depicted in

Figure 6.2.4, which is a hybrid system combining a vapor compression system with liquid desiccant

dehumidification and regeneration. Such a system can obtain high thermal performance and energy

saving.

Zhang and Yin (2008) conducted simulations of the system depicted in Figure 6.2.4 and concluded that

the air and desiccant flow rates are the governing process parameters. Figure 6.2.5 (a) and (b) show the

regeneration efficiency as a function of air and desiccant flow rates. From these results the authors

conclude that:

For air‐side heating source regeneration models, there is an optimum solution flow rate to achieve

maximum regeneration efficiency, which greatly depends on the air flow rate;

The maximum regeneration efficiency decreases with the increase of the air flow rate;

Usually, the maximum regeneration efficiency can achieve at very low flow rate of the desiccant

solution.




July 9, 2016 Page 82 of 171

Figure 6.2.4: An energy‐

efficient air conditioner

with a liquid desiccant

regenerator using

exhausted hot air from the

condenser; (Misha, 2012)

(a) Air flow rate (b) Desiccants flow rate

Figure 6.2.5: Results of simulation of a liquid desiccant cooling system using low grade heat for desiccant

regeneration; (Zhang and Yin, 2008)

Process parameters indexed in Figure 6.2.5 are as follows:

G = mass flow rate, kg/s

X = mass concentration of desiccant solution, %

Ƞreg = regeneration efficiency, dimensionless

Vair = Volume flow rate, m3/s




July 9, 2016 Page 83 of 171

6.3 Desiccant Regeneration with Solar Heat and other Renewable Heat Sources

Solar heat is often mentioned as a main advantage of using desiccant dehumidification to remove latent

loads in air conditioning systems. While energy and cost savings can be realized during the operation,

solar heat from thermal collectors often also qualify for installation incentives (Chen et al, 1992).

Many authors point out benefits of using heat energy instead of utility electricity to energize

dehumidification. Heat sources required to regenerate/dry the desiccant for reuse can include various

solar energy and recovered waste heat as well as gas‐produced heat (direct‐fired or cogenerated). Off‐

peak electric resistance heaters or heat pumps can also be used.

Barlow (1982) performed a comprehensive analysis of solar heat used in desiccant dehumidification. The

author suggested that both solid and liquid desiccants are suitable for solar heat regeneration. Barlow

suggested that data and correlations of the properties of regular density silica gel appeared to be the

most suitable available solid desiccant for solar cooling systems.

Sahlot and Riffat (2016) suggested that the use of solar energy for desiccant applications in space

conditioning has several overlapping benefits. Desiccant cooling systems require electricity to operate

pumps and fans and heat energy to pre‐heat the desiccant solution for regeneration. The author stated

that a solar PV system can be used to derive electricity to drive the pumps and fans. Secondly, solar heat

can be used to regenerate weak desiccant solution. Figure 6.3.1 shows the schematic diagram of a

desiccant cooling system integrated with an evaporative cooler and a solar collector. Combined solar

collector and regenerator is used directly to pre‐heat the desiccant solution before introducing it in the

regenerator. Another type of solar collector uses water as a medium of heat transfer between collector

and liquid desiccant. The combined type of solar collector device is more efficient as all the heat energy

absorbed by the collector is directly transferred to the desiccant solution. However, as desiccant

solution is often corrosive, the solar collector would need to be corrosion resistant.

Sultan et al (2015) provided a comprehensive review of desiccant applications powered by solar heat.

The authors suggested that zeolite prepared from fly ash has shown significant potential for solar cooling applications. While solid desiccants have advantages because of their relative operational ease,

liquid desiccants are often aggressive agents and require suitable material and care when handled in

operation. Common liquid desiccants are lithium chloride (LiCl), lithium bromide (LiBr), calcium chloride

(CaCl2) and triethylene glycol [37,38,42]. Liquid desiccants have the advantage that they are requiring

lower regeneration temperature, typically 60 to 75 °C, which makes them very suitable to utilize solar or

low grade waste heat.




July 9, 2016 Page 84 of 171

Figure 6.3.1: Solar powered desiccant cooling system, (Sahlot and Riffat, 2016)

Kassem (2013) highlights the cost effectiveness of the use of low energy or free solar energy for

regeneration of desiccant material. The authors presented a review of solar energy as a “free energy”

for the regeneration and pointed to the extensive research which had been conducted. The initial cost

of solar energy system is considerable, but over time significant savings can be realized. Therefore, the

payback period should be considered. A disadvantage is that solar radiation is weather‐dependent;

therefore, back‐ up energy or energy storage would be required to continue the drying process when

solar energy is not available. Kassem (2013) pointed out that The use of evaporative coolers for mass

exchange equipment can be applied in liquid desiccant systems. This application can reduce the

installation cost of such systems because the evaporative coolers are available in the market with lower

cost when compared with designing and manufacturing specified mass exchange equipment for such

systems.

Kassem (2013) mentioned advantages of the liquid desiccant over solid desiccants as the lower

regenerating temperatures of liquid desiccants, typically in the range of 40–70 °C, while the higher

values in the temperature range of solid desiccants is as high as 115 °C. This allows the use of low grade

heat sources such as solar energy or waste heat for liquid desiccant regeneration. In addition, liquid

desiccants can be stored in form of concentrated solution for use during periods when solar energy is

absent, and thus offer more flexible operational characteristics. The liquid desiccants are attractive

because of their operational flexibility and their capability of absorbing pollutants and bacteria. Their

disadvantages include entrainment and subsequent carryover of desiccant droplets into the process air

stream during the dehumidification operation.




July 9, 2016 Page 85 of 171

Technologically, the equipment providing air/solution contact surface (contactor) for liquid desiccant

solutions include wetted wall/falling film absorber, a spray chamber or a packed tower. The packed

towers are subdivided into regular (structured) or irregular (random) packing modes. Kassem (2013)

concluded that liquid desiccant assisted air conditioning can achieve up to 40% of energy savings with

regard to traditional air conditioning system and those savings become even greater when the calorific

energy needed for regeneration is drawn from waste heat, solar energy or any other free energy

sources. Figure 6.3.2 shows a typical liquid desiccant cooling process, where liquid desiccant is

regenerated by solar heat.

Daho (2006) commented on feasibility studies of solar driven desiccant cooling in diverse European

cities representing different climatic zones on the continent. The conclusion reached by the author was

that primary energy savings were achieved in all climatic conditions. A decline in energy savings was

noticed in highly humid zones. This decline was attributed to the high temperature required to

regenerate the desiccant in the climates of high humidity. Sahlot and Riffat (2016) estimated that

potential of solar energy uses in desiccant cooling systems and concluded that desiccant systems can

save up to 50% of primary energy.

Figure 6.3.2: A liquid desiccant cooling system with solar collector; (Kassem, 2013)

Besides solar heat, other renewable heat sources can be used to regenerate desiccants. Jradi and Riffat

(2014) present a tri‐generation system which was driven by biofuel. The process, as illustrated in Figure

6.3.3, used a biofuel gasifier to provide fuel for the generator, and the waste heat used to heat domestic




July 9, 2016 Page 86 of 171

water supply as well as for use in absorption chiller. The authors pointed out that waste heat could drive

the regeneration heat instead of the absorption chillers.

Figure 6.3.3: An internal combustion‐based tri‐generation system with a biomass gasification unit ; (Jradi and

Riffat, 2014)


SECTION 7 – ALTERNATIVE COOLING SYSTEMS INTEGRATED WITH DESICCANT DEHUMIDIFICATION


July 9, 2016 Page 87 of 171


The main focus of this literature review is the application of desiccant dehumidification to the

conditioning of buildings. This section presents the technologies which are important to derive effective

cooling applications combined with desiccant dehumidification. It must be noted that often the term

desiccant cooling is used, which is somewhat inaccurate, since desiccants carry out only the latent heat

removal while other cooling technologies contribute sensible heat removal. Figure 7.1 shows the types

of AC systems discussed in this section.

1. Grid powered Conventional AC systems: A large percentage of all existing AC systems work with grid

supplied power and conventional vapor compression cooling. The three functions of the AC‐system are

carried out as follows:

A. Sensible heat removal: Supply air passing cooling coils

B. Latent heat removal: water vapor condenses in cooling oils

C. Ventilation: Air supply that passes the cooling coils is provided to the spaces

2. Grid powered alternative AC systems: This alternative system works with a conventional vapor

compression cooling systems which is augmented with a desiccant dehumidification system. The three

functions of the AC‐system are carried out as follows:


B. Latent heat removal: Water vapor is removed by desiccant sorption. The desiccant regeneration

uses process (waste) heat derived from either combined heat and power (CHP) systems, other

process heat or fossil fuel combustion. A combined dehumidification process of cooling based

dehumidification (condensation on conventional cooling coils) and desiccant dehumidification is

also possible, which can save energy.


3. Solar powered Conventional AC systems: This AC‐system uses PV‐solar derived electric energy to drive

a conventional AC‐system. The three functions of the AC‐system are carried out as follows:


B. Latent heat removal: water vapor condenses on cooling oils


4. Thermal solar powered Alternative AC systems: This alternative system obtains most of its energy

through thermal energy conversion for sensible cooling and dehumidification. The three functions of

the AC‐system are carried out as follows:

A. Sensible heat removal: Sensible heat is removed by supply air passing over surfaces, cooled by

closed systems, or be adiabatic evaporative cooling

B. Latent heat removal: Water vapor is removed by desiccant sorption. The desiccant regeneration

uses solar heat.

C. Ventilation: Air supply that passes the cooling surfaces is provided to the spaces




July 9, 2016 Page 88 of 171

Figure 7.1: Types of building AC systems

Note: Section 7 discusses only literature relative to closed and open thermal systems which supply sensible heat

removal (shaded block in point 4 of above bulleted list)




July 9, 2016 Page 89 of 171

7.1 Thermal Cooling Technologies

This section provides a brief review of several alternative cooling technologies which remove sensible

heat from conditioned spaces. Table 7.1.1 provides an overview of the technologies reviewed hereafter,

where the main distinction is made between closed and open systems.

Closed cycle systems are equipped with thermally driven chillers, which provide chilled water that is

either used in air handling units to provide sensible and latent (cooling‐based) load removal or that

is distributed via a chilled network to decentralized room installations such as fan coils or chilled

ceilings. Available thermally driven chillers on the market are absorption chillers, which are most

common, and adsorption chillers, offered currently by few manufacturers only. A component,

necessary in all chilled water systems, is a heat rejection system.

Open cycle systems are evaporative chillers which provide sensible heat removal. The “refrigerant”

is always water, which is brought into direct contact with the process air and adiabatically reduce

the dry bulb temperature of the moist air.

Among the thermally driven cooling systems, closed cycle systems currently occupy a central position.

According to a survey in 2003, approx. 71 % of the installations were equipped with chillers (63 % using

absorption technology and 8 % using adsorption chillers) (SOLAIR (2009). 29 % of the installations are

desiccant cooling systems

SOLAIR (2009) suggested that solar cooling systems represent a fast growing market with several new

international companies selling products. New interesting developments and concepts are appearing in

the different technologies. In the future, the main barriers to overcome in order to increase the market

share of thermal cooling technologies are the organization of the sector, cost reduction, and the

adaptation of the machines and the main components to technical specifications of solar cooling

systems.

Table 7.1.1: Overview of main air conditioning technologies using thermal energy, SOLAIR (2009)

Method Closed cycle Open cycle

Principle / Process

Providing chilled water for convective

(supply air) or radiative cooling

surfaces

Adiabatic heat removal through

evaporation of water in air supply

Name of cooling

technology

Adsorption

chiller

Absorption

chiller

Direct

evaporative

cooling

Indirect

evaporative

cooling

Adds humidity to air No No Yes No

Phase of sorbent Solid Liquid N/A N/A

Typical material pairs water ‐ silica gel water/Li‐Br water ‐ air water ‐ air




July 9, 2016 Page 90 of 171

7.2 Evaporative Cooling – General Principle

Evaporative cooling has been used for a long time to provide relief to humans exposes to high

temperature climate. The concept of using water for air cooling has been around for centuries (Watt and

Brown, 1997). While complex evaporative cooling systems have been invented and refined over the past

decades, the basic concept has not changed. Egyptian frescoes portraying large, porous jars of water

being fanned to force evaporation and subsequent cooling have shown the prevalence of evaporative

cooling since ancient times. In the Arizona desert in the 1920s, people would often sleep outside on

screened‐in sleeping porches during the summer. On hot nights, bed sheets or blankets soaked in water

would be hung inside of the screens. Whirling electric fans would pull the night air through the moist

cloth to cool the room.

That concept, slightly more refined, became the evaporative coolers that to this day provide a low‐cost,

low‐technology alternative to refrigerated air conditioning. An evaporative cooler produces effective

cooling by combining a natural process ‐ water evaporation ‐ with a simple, reliable air‐moving system.

Fresh outside air is pulled through moist pads where it is cooled by evaporation and circulated through a

building by a fan system. With the evaporative process, the outside air temperature can be lowered as

much as 30 F. (Palmer, 2002)

The term “swamp cooler” is frequently used for evaporative coolers, because of the standard design of

having a tray which collects the water that is not evaporating. Evaporative coolers work well, provided

the outside air they are drawing in is dry and desert‐like. As the humidity increases, the ability for these

systems to cool the air decreases.

The dependency of the outside climate and humidity was one of the reasons that conventional air

conditioning systems, basically all vapor compression systems, became popular since their ability to cool

the air was not impeded by the outside humidity. The advantages of these conventional mechanical AC‐

systems were readily apparent as they were able to lower the temperature to a thermostatically

controlled temperature, even on humid days. The disadvantage of conventional vapor compression AC‐

systems is their significant electricity demand, as much as four times the electricity needs of evaporative

coolers. In addition, conventional AC‐systems are more expensive to install and maintain than

evaporative coolers. And, they require ozone‐damaging refrigerants (Green Cooling Initiative, 2016).

Palmer (2002) points out that in dry desert areas, evaporative coolers work the best, but they also

perform reasonably well fine most of the time in California's more humid climates. Sacramento, for

example, averages about 45 percent humidity on a typical hot summer afternoon, still dry enough for

evaporative cooling to work effectively. Despite the potential, however, fewer than five percent of

California homes and businesses use evaporative cooling.




July 9, 2016 Page 91 of 171

Several authors such as Kozubal et all (2011), Watt and Brown (1997), and Sahlot and Riffat (2016)

summarize the benefits of evaporative cooling as follows:

Evaporative coolers use significantly less electric energy use than energy intensive vapor

compressor based cooling. The only electricity demands of evaporative chillers are supply fans

and a sump pump. As a consequence, evaporative cooling requires 1/5 to 1/2 as much electricity

as mechanical cooling to operate.

Maintenance requirements are simpler for evaporative cooling systems than for refrigerated air

conditioning equipment. Refrigeration compressors, evaporators and condensers must operate

under high pressures, which require specialized tools and certified maintenance personnel.

Evaporative cooler users can maintain their peak cooling effectiveness without the need for

costly and sometimes unavailable specialized maintenance contracts.

The compounded (life‐cycle cost) of using evaporative cooling is less than a comparable

refrigerated air unit. This includes all dollar values such as first cost, energy, water, time value of

money and maintenance costs.

Evaporative cooling saves water at the power plant. During a typical summer in New Mexico, a

coal fired power plant using evaporative cooling towers will typically need about 0.95 gallons

per kWh. This quantity does not include the water needed to mine, process and deliver the coal

used to generate the electricity. The amount of water used by an evaporative cooler is stated in

terms of tons of cooling per gallon, which is on average one gallon of water per 0.6 ton of

refrigeration. Therefore, the water requirement for evaporative cooling is about 50 to 60% of a

typical water requirement for a thermal power plant.

Evaporative cooling does not directly use any working fluids which are detrimental to the earth’s

ozone layer. This is unlike most of the pre‐2000 commercial refrigerants whose use is regulated

in order to reduce their harmful impact on the environment. Evaporative coolers do not operate

under high pressure conditions and do not require any expensive controlled substances for their

operation.

An evaporative cooled building will always require less power to operate than a mechanically

cooled building. As a rule of thumb 0.5 to 5 kW and 3 to 10 kW of electrical power is required by

evaporative and mechanical cooling, respectively. The reduced power demand does not only

lower operational costs through smaller demand charges, but also lowers first costs for electrical

components will cost less.

The energy savings of evaporative cooling translates directly into reduced carbon dioxide and

other emissions from power plants, and decrease the peak electricity demand load that typically

occurs during peak summer cooling hours.

There are approximately 4 million evaporative air cooling units in operation in the United States

which provide an estimated annual energy savings equivalent to 12 million barrels of oil, and

annual reduction of 5.4 billion pounds of carbon dioxide emissions. They also avoid the need for

24 million pounds of refrigerant traditionally used in residential VAC (vapor‐compression air

conditioning or refrigerated air) systems".

Since evaporative cooling typically uses 100% outside air rather than recirculated air indoor air

quality is improved. The outside air and humidity added to the room air by an evaporative




July 9, 2016 Page 92 of 171

cooler can improve comfort conditions, flush out contaminants which are generated in the

building and reduce the incidence of static electric shock which can be detrimental to micro‐

electronics.

The working principle of evaporative cooling systems is water evaporation. This change of phase

requires latent heat to be absorbed from the surrounding air and the remaining liquid water. As a result,

the air temperature decreases and the relative humidity of the air increases. The maximum cooling that

can be achieved is a reduction in air temperature to the wet‐bulb temperature (WBT) at which point the

air would be completely saturated, but most evaporative cooling devices attain less than the technically

feasible temperature drop.

When air below saturation moves over a surface of water evaporates. This evaporation results in a

reduced temperature and an increased vapor content in the air. The bigger the area of contact between

the air and water the more evaporation occurs, resulting in more cooling and the addition of moisture.

In order for water to evaporate, heat is required. In order to evaporate one gallon of water

approximately 8,700 BTU’s of heat are required. For evaporative cooling, this heat is taken from the air,

thereby cooling it.

There are two basic types of evaporative air cooling (EAC) processes. Direct evaporative cooling occurs

when process air has direct contact with the evaporating water. Indirect evaporative coolers utilize a

heat exchanger; therefore, the process air never comes into direct contact with the cooling water. Direct

EAC is commonly used for residential cooling. Developments in the evaporative cooling industry have

reliably increased the efficiency or effectiveness of the cooling media. All direct EAC’s use 100% outside

air. Indirect evaporative air coolers (IEAC's) can be used in conjunction with direct EAC's and/or with

refrigerated air coolers. IEAC systems use electricity for the supply fan motor, a sump pump, and a

smaller secondary fan motor used for the heat exchanger’s airflow. The combination evaporative and

refrigerated system has a higher first cost, but offers a good mix of energy conservation and comfort.

Additionally, these redundant cooling systems are more reliable.

IEAC are rated from 60% to 78% thermal efficiency, depending on the configuration and the air speed

past the heat exchanger. IEAC systems can be used in combination with direct evaporative cooling, in

combination with refrigerated air systems or as a stand‐alone system. When combined with direct EAC

systems the effectiveness is additive. It is essential that indirect evaporative chillers do not add moisture

to the room air stream and therefore do not increase the room humidity level. Therefore, indirect

evaporative coolers can recirculate the room air.

The possible energy savings of evaporative coolers are significant when compared with conventional

mechanical refrigeration cycles. Table 7.2.1 shows a comparison of average energy savings through the

use of evaporative cooling systems Palmer (2002).




July 9, 2016 Page 93 of 171

Table 7.2.1: comparison of evaporative and conventional split system AC; (Palmer, 2002)

7.3 Direct Evaporative Cooler – Working Process

The typical direct evaporative cooler unit consists of a metal, plastic or fiberglass housing and frame, a

supply fan, water holding sump, water circulation pump, water distribution tubing, electric connections

and a wetted pad. These pads are the surface from which the water evaporates, and are usually made of

aspen shavings, paper or plastic media. Typical manufacturers stated evaporative effectiveness for this

type of wetted media is 65 to 78%. All EACs use a small fractional horsepower pump to raise the water

over the pads, then gravity and capillary action wet the entire area of the evaporative media.

Advantages of direct evaporative cooling are its low life‐cycle cost, improved indoor air quality, reduced

peak electrical demand, simple controls and low‐tech maintenance. Disadvantages include relatively

short service life of aspen media aspen pad coolers (1 to 3 years, depending on media type and

maintenance), the need for seasonal maintenance and reduced cooling performance during the wet

season.

There are many types of wetted media used in various configurations. Rigid media coolers are usually

more effective than flexible aspen pad coolers because they have more surface area per cubic volume of

media. A value of 100 to 130 square feet of surface area per cubic foot is common for rigid wetted

media (Palmer, 2002). As an added advantage rigid wetted media does not sag and reduce cooling

performance. It is available in various thicknesses between 2 and 24 inches, but 12‐inch thick media is

common for school and commercial building air handling units. As shown in the section on evaporative

media, the manufacturers ratings of effectiveness of rigid media is 75% to 95%, depending on the

thickness and air velocity through the media. Rigid media is washable, and with good regular

maintenance will last 7 to 10 years. One significant factor is that direct EAC systems cannot recirculate

the room air. All the air must be exhausted or otherwise relieved from the building.




July 9, 2016 Page 94 of 171

Figure 7.3.1 and 7.3.2 show a basic process diagram for a direct evaporative cooler and a typical Wetted

Aspen Pad Cooler, respectively.

Figure 7.3.1: Direct

evaporative cooling AC

(Palmer, 2002)

Figure 7.3.2: Typical

Wetted Aspen Pad Cooler

(Palmer, 2002)

7.4 Indirect Evaporative Cooler – Working Process

Indirect EAC's (IEAC's) have been in used not as long as direct EACs, but they have gained increased

acceptance because of more cost‐effective manufacturing and improved performance. Since IEAC’s use




July 9, 2016 Page 95 of 171

heat exchangers they do not add moisture to the room air stream, thus they only provide sensible

cooling.

Figure 7.4.1 shows a schematic view of an indirect evaporative cooling section located upstream of a

direct rigid media evaporative cooler. IEAC heat exchangers transfer heat across sheets of plastic or

metal configured to keep the two air streams from mixing. A smaller fan pulls the wetter air through the

secondary side of the heat exchanger. This type of air‐to‐air heat exchanger is used to transfer heat from

the primary space supply air stream to the evaporative cooled secondary air stream.

Figure 7.4.1: Schematic of

indirect evaporative

cooler; (Palmer, 2002)

Palmer (2002) suggested that following system configurations of IEACs:

Air‐to‐air heat exchangers (see Figure 7.4.2): Direct and indirect evaporative air cooling components

are commonly packaged together in the same cooling unit by the manufacturer. The main

advantages of using both types of evaporative cooling systems are the lower supply air temperature,

increased reliability and more available comfort hours. The direct EAC section will usually cool the

air more than the indirect section due to the losses of the IEAC heat exchanger

Combination IEAC and refrigerated systems (Figure 7.4.3): IEAC's are used in combination with

refrigerated air systems, because they do not add moisture to the air stream. The advantage of this

combination is that the reduction in air temperature due to the indirect evaporative section comes

at a significant cost savings, since cooling capacity by an IEAC is much cheaper than provided by a

conventional mechanical refrigerated system alone. The return air from the room can be

recirculated past both the IEAC and the refrigerated coil, or the IEAC can be located to pre‐cool only

the required outside air.




July 9, 2016 Page 96 of 171

Cooling tower "free cooling" (Figure 7.4.4): Indirect evaporative cooling system can be combined

with a waterside economizer. This is often the case with a chiller system that uses evaporative

cooling tower(s) for condenser heat rejection. When the outside air humidity level is sufficiently low,

the evaporative cooling tower system works so well that the chillers can be kept off much of the

time. The indirect system uses a water‐to‐air cooling coil in the room air stream. This evaporative

cooled water in the cooling tower sump can get very cool, especially when the air is dry. Depending

on the design parameters of the system, this water can be within 3 to 6 degrees of the wet‐bulb

temperature.

Figure 7.4.2: Air‐to‐Air

Heat Exchanger Indirect

+ Direct EAC; (Palmer ,

2002)

The figure shows a

section through an IEAC

combined with a direct

EAC.

Figure 7.4.3:

Combination

Evaporative and

Refrigerated Cooling

System Schematic;

(Palmer, 2002)




July 9, 2016 Page 97 of 171

Figure 7.4.4: “Free Cooling”

Evaporative and Refrigerated

Combination System; (Palmer,

2002)

7.5 Maisotsenko (M) ‐Cycle Enhanced Indirect Evaporative Cooling

A recent development in evaporative cooling is M‐cycle evaporative cooling, which can be utilized as a

special and highly effective process of indirect evaporative cooling. The M‐cycle operates on the

principles of sub‐wet bulb cooling. These properties make the M ‐ Cycle ideal for many evaporative

cooling applications, including cooling towers, where steady cooled water temperatures can be

delivered as the higher inlet temperatures increase the cooling capacity.

In conventional direct evaporative cooling technology, the temperature of the cold water produced is

limited to the outside air wet bulb temperature. As a result, the typical cooling towers are most suitable

for extremely dry regions. Unlike conventional evaporative cooling technologies, heat transfer within

the M ‐ Cycle is driven by a dew point temperature gradient, not wet bulb temperature. As such, only

the absolute outside ambient humidity is important, not the temperature of incoming air or water.

Film‐type packing has gained prominence in the evaporative cooling process vessel design because of its

ability to expose greater water surface within a given packed volume. Its efficient heat transfer can give

greater cooling capacity, requiring less fan energy. Using the conventional evaporative cooling process

with any existing type of fill for cooling towers, it is possible to cool down fluid (air or water) to near the

wet bulb temperature. However, with the M‐cycle, using the existing film fill and a change of air and

water distributions cooled water temperatures approaching the dew point temperature of the working

gas is possible. Gas Technology Institute (GTI) (201) presented a comparison of a conventional and M‐

cycle evaporative cooling towers. Figure 7.5.1 and 7.5.2 compares the performance of a conventional

and M‐cycle evaporative cooling tower, respectively.




July 9, 2016 Page 98 of 171

Flow diagram packing Psychrometric chart

Subscripts: DB= dry bulb, WB= wet bulb, DP= dew point

Figure 7.5.1: Forced draft conventional cooling tower, (GTI, 2010) annotated

packing Psychrometric chart

Subscripts: DB= dry bulb, WB= wet bulb, DP= dew point

Figure 7.5.2: M‐cycle open cooling tower, (GTI, 2010) annotated




July 9, 2016 Page 99 of 171

Table 7.5.1 presents the results of the GTI investigation of the conventional and M‐cycle evaporative

cooling tower performance. The result presented in Table 7.5.1 highlights the significant process

improvement of the M‐cycle of providing cooling water for indirect evaporative cooling systems.

Table 7.5.1: Comparison of process effectiveness of conventional versus M‐cycle evaporative cooling tower

Process flow unit conventional cycle M‐cycle

Water to be cooled [F] 90 90

process air intake DB [F] 86 86

process air intake WB [F] 66 66

process air intake DP [F] 54.7 54.7

Cooled water temperature [F] 72 55.6

Dew point evaporative cooling processes, such as the M‐cycle, can be used to remove undesired heat

from a wide range of thermal applications, including building air conditioning systems. Dew point

evaporative cooling processes take advantage of the fact that when a parcel of air is sensibly cooled, the

saturated water vapor pressure decreases, reducing its wet bulb temperature, thus increasing its

evaporative cooling potential. Consequently, as the working fluid is humidified, the temperature of the

evaporative cooling liquid that it is in contact with is also cooled to theoretically as low as the incoming

air dew point temperature

GTI (2010) emphasized that the operating efficiency of the M‐cycle cooling tower can be enhanced by

reducing the dew point temperature of the incoming air. For example, reduced dew point temperatures

can be achieved by desiccants, which can take place before and/or after the air actually enters the dry

channel. Additionally, heat sources can be utilized to increase the system cooling capacity, in preheating

the incoming working gas (e.g. ambient air) which improves the thermal performance of the M‐Cycle.

This process characteristics is especially advantageous in high absolute humidity climates where the dew

point temperature of ambient air is in the range of 60 °F to 70 F. For even higher dew point

temperatures, the effect of changing absolute humidity, with changing temperature, will become even

larger. Similarly, use of a warmer evaporative liquid (e.g. hot water) will have the same effect as pre ‐

heating the incoming gas (e.g. air).

7.6 Absorption Chillers

The process of Absorption refrigeration is not a new process. In 1858 the absorption cooling process was

invented by the French scientist Ferdinand Carré where a mixture of water and sulphuric acid was used




July 9, 2016 Page 100 of 171

as the absorbent (Arsenault, 2016). The low coefficient of performance (COP) of the absorption cooling

cycle was the development of the alternative means to provide cooling capacity and refrigeration, the

vapor compression cycle, which was invented by Wills Carrier in 1902. The vapor compression

refrigeration cycle became the cooling process of choice, whereas the absorption chiller remains as an

alternative where suitable quality and quantity of process waste heat is available. The chillers can be

classified by the type of thermal energy source, whether they are single‐ or double‐effect processes and

whether they are directly or indirectly fired.

There are a number of absorption chillers available, including single‐effect indirect‐fired (steam, hot

water); double‐effect indirect‐fired (external heat source); and double‐effect direct‐fired (gas and/or oil

burner). Single‐effect absorption chillers have single set of process vessels. Double‐effect absorption

chillers have two sets which make them more energy efficient since process heat is not rejected to the

environment, as in the case of the single effect absorption chiller, but is recycled internally.

Absorption chillers have substantially reduced internal pressures to take advantage of the lower water

boiling temperatures. Absorption chiller internal pressures can range from 0.1 atmosphere (atm) to

below 0.01 atm (Sakraida, 2016). Figure 7.6.1 show a basic absorption chiller process diagram.

Figure 7.6.1.: Basic flow diagram of a single effect absorption chiller

The diagram indicates the mass flow of the absorption and regeneration process. Pressures and energy flows (e.g. heat rejection) are not all indicated. (EERE, 2016)

The absorption cooling process works on the basis of a solution of lithium bromide and water, which

absorbs water vapor and drives the process. As illustrated in Figure 7.6.1 water in tank A evaporates and

water vapor flows towards the absorber (tank B) which contains the absorbing solution. Evaporation of




July 9, 2016 Page 101 of 171

the water in in the evaporator (tank A) causes the temperature to decrease. The achievable

temperature in the evaporator depends on the evaporator pressure and the effectiveness of the

evaporative process. Since mass enters the absorber (tank B) and dilutes the solution, fluid from the

absorber is continuously pumped from tank B to the generator (tank C). In the generator, the solution is

heated either directly by a natural gas burner, or indirectly by means of a steam coil. While the solution

is heated, water evaporates and passes into the condenser (tank D). The regenerated and concentrated

solution accumulates in the sump of tank C, and is then returned to the absorber (tank B). Here, it again

absorbs water vapor that comes from the evaporator. Water vapor in the condenser (tank D) is cooled

by a separate coil of pipe through which cooling water passes. The condensed water is returned to the

evaporator (A).

Figure 7.6.2 shows the absorption cooling cycle by adding cooling and chilled water loops as well as

pumps.

Figure 7.6.2: Process diagram of a single effect absorption chiller with major process components added, (EERE, 2016)




July 9, 2016 Page 102 of 171

Figure 7.5.3: Single‐effect

absorption refrigeration cycle;

(Sakraida, 2016); annotated

A more detailed process illustration is provided in Figure 7.5.3. The following describes the main

components of the process shown in Figure 7.5.3.

Evaporator vessel (1): Liquid refrigerant (chilled water loop water) is pumped through the chilled water

tube bundle. The chilled water cycle provides cooling capacity to conditioned spaces by transporting

chilled water out of the evaporator. Water which cycled inside the evaporator is sprayed on the tube

bundle and evaporates. At the low evaporator pressure (as low as less than 0.01 atm), the water

evaporates at approximately 38 F, removing heat energy from the chilled water. Most lithium bromide

absorption chillers can only produce chilled‐water supply temperatures down to about 40 F (Sakraida

,2016). Process water which does not evaporate drops down into a well and is recirculated. The vapor

generated by evaporation flows to the absorber through a low pressure loss conduit.

Absorber vessel (2): The water vapor entering the absorber from the evaporator passes through a falling

film or spray of liquid lithium‐bromide solution. The lithium bromide solution absorbs the vaporized

refrigerant and becomes diluted. Cooling water absorbs the heat of vapor absorption and rejects it the

heat to the outside. A portion of the liquid lithium‐bromide solution in the well is recirculated to the

falling film or spray by the absorber spray pump (A). The remaining portion of the liquid lithium‐bromide

solution is transported by pump (B) to the generator/concentrator vessel. On its way to the generator/

concentrator vessel the solution passes through a heat exchanger (E) where process heat is recovered.




July 9, 2016 Page 103 of 171

Generator/concentrator vessel (3): In the generator/concentrator vessel the dilute lithium‐bromide

solution is heated by steam or hot water. The temperature of the lithium bromide solution increases to

a point where the water which is absorbed in the solution evaporates. The water vapor travels to the

condenser. The concentrated lithium bromide solution flows down to the absorber and passes through

the heat exchanger (C).

Condenser vessel (4): In the condenser, the water vapor is condensing on a tube bundle through which

cooling water flows, which rejects the heat of condensation to the outside. The condensate collects in a

well from where it overflows to the evaporator.

Expansion piping (5): The condensate (water) flows from the condenser through expansion piping to the

well in the evaporator vessel. Since the pressure in the condenser (about 0.1 atm) is significantly higher

than in the evaporator (less than 0.01 atm) the water flow experiences a drop in pressure and

temperature. Excess vapor, which is compromising the low process pressure in the evaporator is

expelled from the system by the purge pump. This completes the refrigerant cycle.

Figure 7.6.4 and 7.6.5 illustrate the difference between single cycle and multiple cycle absorption

refrigeration cycles. The energy transfer, e.g. the process flows inside the system and across the systems

boundaries is illustrated in Figure 7.6.4. Double‐effect absorption cycles capture some internal heat to

provide part of the energy required in the generator to create the high‐pressure refrigerant vapor. Using

the heat of absorption reduces the steam or natural gas requirements and boosts system efficiency.

Figure 7.6.5. shows the reuse of internal process heat energy by the following effect in the double‐effect

absorption cycle.




July 9, 2016 Page 104 of 171

Figure 7.6.4: Indirect heated single‐effect absorption refrigeration cycle; (Thermax, 2016)

The double effect chiller requires significantly more system components and controls, but the efficiency

of a double‐effect absorption cycle can be significantly increased of the single‐cycle process. Sakraida

(2016) reports that the coefficient of performance (COP) of the double‐effect is near double that of the

single‐effect system, with:

Hot water or steam single‐effect chiller . . . . . . . 0.60 to 0.75 COP

Hot water or steam double‐effect chiller . . . . . . 1.19 to 1.35 COP




July 9, 2016 Page 105 of 171

Figure 7.6.5: Indirect heated double ‐effect absorption refrigeration cycle; (Thermax, 2016)

SOLAIR (2009), an European Solar Cooling Initiative of 13 countries, suggests that many absorption

chiller products are available in the market, however typical chilling capacities of absorption chillers are

several hundred kW. For several years, the smallest machine available was a Japanese product with 35

kW capacity. The absorption chillers are typically supplied with district heat, waste heat or heat from co‐

generation. The required heat source temperature is usually above 80 °C for single‐effect machines and

the coefficient of performance (COP) is in the range from 0.6 to 0.8. Double‐effect machines with two

generator stages require driving temperature of above 140 °C, but the COP’s may achieve values up to

1.2.




July 9, 2016 Page 106 of 171

In a recent move to facilitate thermal solar energy conversion, several single‐effect absorption chillers

with capacities below 50 kW are now available. Therefore, solar air‐conditioning (SAC) systems can now

manufacture smaller units and be more flexible. A newly developed model for small capacity chiller

enables part‐load operation with reduced chilling power at a heat source temperature of 65 °C and a

COP of still approximately 0.7, which is quite promising in combination with a solar heat source. This

shows that there is a high potential for performance improvements of absorption chillers. The new

medium‐size and small‐size absorption chiller will be suitable to cover the cooling loads for building

areas from 200 m² to 500 m². The European manufacturers are located in Germany, Austria, Spain,

Sweden, Italy and Portugal. Some of the new absorption chiller models are still being tested in pilot

installations.

7.7 Adsorption Chillers

When compared to absorption refrigeration, adsorption chillers are a relatively recent thermal

technology development. Adsorption refrigeration builds on the older absorption chiller process, yet has

important operational differences, especially in regard to usability in solar energy conversion (SOLAIR,

2009). Adsorption chillers use solid sorption materials instead of liquid solutions to drive the

refrigeration process. Commercially available adsorption systems use water as refrigerant and silica gel

or zeolite as the sorbent material.

The currently available adsorption chillers consist of two process vessel compartments, one evaporator

and one condenser, which are used batch‐wise. Figure 7.7.1 illustrates the basic process scheme of

adsorption cooling. While regeneration of the sorbent occurs in the compartment 1, using hot water

from the external heat source, such as solar generated hot water, the sorbent in compartment 2

adsorbs the water vapor which is generated in the evaporator. Compartment 2 has to be cooled in order

to enable a continuous adsorption of water vapor. Due to the low pressure conditions in the evaporator

vessel, the refrigerant (water) vaporized by taking up heat of from the chilled water loop and thereby

producing cooling capacity. When sorption material in the adsorption compartment is saturated with

water vapor to a certain degree, the chambers are switched over in their function.

The available adsorption chiller technology can work with external heat sources at temperature of 80 °C,

where a coefficient of performance (COP) of about 0.6 is obtained. It is possible, however, to operate

the systems with heat source temperatures as low of approx. 60 °C. The capacities of the commercially

available chillers range from 5.5 kW to 500 kW chilling power. Greenchiller (2016) reports significant

benefits of using adsorption chillers, such as long life of the system and very low operating costs. The

chiller’s electrical load of just 0.4 KW, for the 1.8‐ton model, is for the controls and two small pumps.

One is a vacuum pump for non‐condensable gases and the other is a refrigerant liquid (water) pump

that runs only while unloading. Other operational costs such as maintenance are also very low.




July 9, 2016 Page 107 of 171

Figure 7.7.1: Basic process

scheme of adsorption

cooling; (SOLAIR, 2009),

annotated

The German company InvenSor

GMBH, which is one the leading European manufacturer of adsorption chillers, offers systems with a

range of chilling capacities between 10 and 30 kW. The typical specification of InvenSor adsorption

chillers lists a thermal coefficient of performance (COP) of 0.72, as the efficiency of the use of heat for

cooling, and an outstanding maximum energy efficiency ratio (EER) of 33, which included the use of

electric energy.

Figure 7.7.2 illustrates the working principle of the InvenSor adsorption chiller. The compartments (1)

and (2) interconnect the evaporator and condenser vessels inside the machine. An illustrated step‐wise

description is illustrated in Figure 7.7.3. Figure 7.7.4. depicts the Adsorption Chiller InvenSor Model LTC

10, with a rated capacity of 10 kW. The machine operates with external heat sources as low as 65°C.




July 9, 2016 Page 108 of 171

Process Step A:

Compartment (1) serves as the receiving process vessel for the water vapor that is generated in the evaporator.

Compartment (2) serves as the regeneration of the sorbent.

(InvenSor, 2016), annotated

Process Step B:

Compartment (1) and (2) switch their functions.

Compartment (1) now serves as the regeneration of the sorbent

Compartment (2) now serves as the receiving process vessel for the water vapor that is generated in the evaporator.

(InvenSor, 2016), annotated

Definition of working fluid flow

Figure 7.7.2: Two‐phased working process of InvenSor adsorption cooling, (InvenSor, 2016), annotated




July 9, 2016 Page 109 of 171

Process step 1: Evaporation and adsorption

in one compartment Water in the evaporator vaporizes and creates chilling capacity that is transferred to the chilled water flow.

Water vapor flow to compartment (2) where vapor is adopted by the solid sorbent.

The heat of adsorption is rejected to the cooling water

Process step 2: Evaporation and adsorption

in two compartments When the adsorption capacity of the solid sorbent is diminishing compartment (2), the regenerated sorbent in compartment (1) is operating parallel to Compartment (2) as the absorber vessel.




July 9, 2016 Page 110 of 171

Process step 3: Evaporation and

Regeneration in one compartment

As the sorbent in compartment (2)

becomes saturated with water vapor, the

process function of compartment (2)

changes to desorption, where the solid

sorbent is regenerated.

External heat is supplied to compartment

(2) and water vapor is liberated from the

solid sorbent.

The liberated water vapor flows from

compartment (2) to the condenser where

the vapor condenses on cooling coils.

Compartment (1) is still in adsorption

mode.

Process step 3: Evaporation and

Regeneration in two compartments

As the sorbent in compartment (1)

becomes saturated with water vapor, the

process function of compartment (1)

changes to desorption, where the solid

sorbent is regenerated.

Both compartments are now in desorption,

e.g. regeneration mode,

The water vapor condensate flows back to the evaporator

Figure 7.7.3: Process steps 1 through 4 of the InvenSor adsorption cooling process, source (InvenSor, 2016),

graphics created by the author




July 9, 2016 Page 111 of 171

Figure 7.7.4: Adsorption Chiller InvenSor LTC 10, rated

capacity of 10 kW, (InvenSor, 2016)

The machine operates with external heat sources as

low as 65°C.

7.8 Comparison between Absorption and Adsorption Cooling Technology

Absorption refrigeration process has been used for a long time, while adsorption refrigeration is a

relatively recent development of thermal cooling technology. Bargman (2016) discusses the main

differences of the two thermal cooling processes and points out main design and operational

characteristics.

Bargman (2016) suggested that since there is no use of CFC's or ammonia, the thermally driven cooling

process is environment friendly. Adsorption chillers are known for their robust construction, ease of

installation, and in many cases, considered more advantageous than absorption chillers. Adsorption

chillers are not prone to problems such as crystallization, corrosion, hazardous leaks, and the electricity

consumption is very low. The typical cooling capacity of absorption and adsorption chiller are ranges

between 4.5 KW to 5 MW and 5.5 to 500 KW, respectively.

Perhaps the most significant difference between absorption and adsorption chillers is the expected

useful life of the equipment. Absorption chillers have an average lifetime of 7 to 9 years due to corrosion

problems, and the systems are high in complexity and maintenance time. On the other hand, adsorption

chillers have a high chiller life expectancy of about 30 years The main advantages of absorption and

adsorption chillers are indicated in Table 7.8.1.




July 9, 2016 Page 112 of 171

Table 7.8.1: Comparison of Adsorption Chillers vs Absorption Chillers; Bargman (2016)

FACTORS ADSORPTION CHILLERS ABSORPTION CHILLERS

Corrosion Protection Not Required High Corrosion Protection

Required

Crystallization No Crystallization Very High

Inhibitor Not Required Heavy Metal Inhibitors

Life Expectancy Greater than 30 Years 7 to 9 Years

Complexity Simple, Easy Mechanical

Operation

Complex, Chemical Operation

Replacement Requirements Not Required Heat Exchangers, Boilers,

Absorbent Replacement

Required

Temperature Down to 122°F Shut down at 180°F, Needs

Back‐up Heater

Chilled Water Output 40°‐ 55°F 48°F or More

7.9 Magnetic Refrigeration as an Alternative to Vapor Compression Refrigeration

While the man objective of this chapter is to present thermally operated air‐conditioning technology,

the recent cooling technology development of magnetic refrigeration is included in this literature and

technology review. Magnetic refrigeration has the potential of being a “disruptive cooling technology”

since it offers several significant benefits over the conventional vapor compression refrigeration

technology. Early magnetic cooling products suggest that these cooling systems are more energy

efficient, smaller and less noisy than vapor compression cooling and operate with much less

maintenance. In regard to energy efficient cooling and desiccant dehumidification systems, magnetic

refrigeration could be viable candidate to provide sensible cooling and waste heat for desiccant

regeneration.




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The magnetic refrigeration process has been known for over seventy years to achieve near 0 K

temperatures for small gas volumes by a process known as “adiabatic demagnetization refrigeration”.

Magnetic refrigeration is based on the magnetocaloric effect which was first observed in the early 20th

century by various researchers. Magnetic refrigeration has long been considered one of the simplest

ways to retrieve extremely low temperatures. However, the application of focus in current research has

shifted from providing extremely cold temperature to more practical uses, including consumer‐based

refrigerators. In the 1950's, several magnetic refrigerators operated at temperatures between 1 and 30

K. However, these were too inefficient to be used commercially and could only run for a couple of days

(O’Handley, 2004). Magnetic refrigeration has only very recently been developed to commercial

product status. The main advantage of magnetic refrigeration is the avoidance of refrigerant agents in

providing chilling capacity, which is eliminating several environmentally harmful effects of refrigerants,

including global warming and the depletion of the ozone layer.

When a magnetic field is applied to a magnetic material which has a magnetocaloric effect. A spin of the

magnets lowers the entropy of the system since disorder has decreased. In order to compensate for the

aligned spins, the atoms of the material begin to vibrate, and lower the entropy of the system again. In

doing so, the temperature of the magnetocaloric material increases. The warming and cooling process

can be compared, in principle, to a standard refrigerator which implements compressing and expanding

gases for variations in heat exchange and surrounding temperature. Figure 7.9.1 illustrates the basic

working scheme of the magnetic refrigeration cycle which is composed of the process of magnetization

and demagnetization, in which heat is discharged or absorbed in four steps.

While there have been early prototypes of magnetic refrigeration in ranges that can be used in

consumer products, high costs have been a main barrier to gaining market share from conventional

vapor compression refrigeration. One significant breakthrough on the way to making magnetic

refrigeration competitive with conventional systems is the discovery of the so‐called "giant" Magnetic

cooling effect which made previously considered implausible advancements possible. Some materials

were observed to have unusually large magnetocaloric effects, and this discovery made it much more

feasible to develop a machine at lower costs (Kuba and Ota, 2006).

The French company Cooltech is offering the first commercial magnetic refrigeration product (Cooltech,

2016). Figure 7.9.2 shows Cooltech’s proposed cooling application. Figure 7.9.3 shows a detail view of

the core of Cooltech’s magnetic cooling systems.




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Figure 7.9.1: Basic working

scheme of magnetic

refrigeration

(Cooltech, 2016)

Figure 7.9.2: Cooltech’s proposed cooling application for magnetic refrigeration system (Cooltech, 2016)




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Figure 7.9.3: The Magnetic Refrigeration System (M.R.S)

2015, (Cooltech, 2016)

This device is the first commercially available refrigeration

device based entirely on magnetocaloric effect.

Benefits of magnetic cooling are reported as flows (Cooltech, 2016):

Energy savings: Magnets are permanent and do not require an energy source to operate.

Achievable energy savings could be as high as 50%.

Cost reduction: Magnetic cooling provides lower operational costs due to optimized level of

maintenance (low rotational speed, low pressure, no leaks and no hazardous chemicals…) and

lower energy consumption.

No Refrigerant Gas: The magnetic refrigeration system does not use any refrigerant gas and is not

concerned with the regulations of harmful refrigerants.

Safety: Magnetic refrigeration provides safety due to the absence of refrigerant gas. This protects

the users from hazardous leaks. Measurements were performed on all Cooltech machines and

showed magnetic emissions that surrounded Cooltech’s devices were in safe ranges.

Reduced noise and vibrations: The magnetic refrigeration system operates with a low noise level

(less than 35db) and reduced vibrations.


SECTION 8 ‐ SOLID DESICCANT COOLING SYSTEMS


July 9, 2016 Page 116 of 171


This section presents the review of literature that discusses the integrated use of desiccant

dehumidification and cooling systems. Kolewar, A. et al (2014) suggested that desiccant cooling can be

either a substitute to the traditional vapor compression air conditioning technology or can expand

conventional AC systems and make them more cost efficient, accessible, and cleaner. The authors

pointed out that when powered by free energy sources (renewable energy sources) such as waste heat

and solar energy, desiccant cooling can considerably reduce the operating costs and increase noticeably

the user‐friendliness to the air conditioning for the populations in remote areas.

The emphasis of this literature and technology review is on alternative and energy efficient cooling

systems which are augmented by desiccant dehumidification. This section presents several system

combinations of solid desiccant and various cooling technologies.

8.1 Solid Desiccant Cooling Systems using Different Cooling Devices

Daou et al (2005) discussed the use of different cooling units for desiccant cooling systems. A desiccant

cooling system comprises three main components, namely the regeneration heat source, the

dehumidifier (desiccant material), and the cooling unit. The author pointed out that the cooling unit can

be the evaporator of a conventional air conditioner, an evaporative cooler or a cold coil. The function of

the cooling unit is handling of the sensible load while the desiccant removes the latent load. When a

desiccant wheel system is implemented, a heat exchanger is generally used in tandem with it to

preliminarily cool the dry and warm air stream before its further cooling by an evaporative cooler or a

cold coil, etc. In this case, the heat exchanger together with the evaporator cooler or the cold coil

constitutes the cooling unit. Fig. 8.1.1. shows a psychrometric representation of the different cooling

process combinations, with the following main process steps:

Desiccant dehumidification lowers humidity of the air and at the same time increases the temperature of the air through the heat of adsorption of water vapor on the solid desiccants. (state 1 to state 2)

A heat wheel lowers the air temperature by rejecting sensible heat (state 2 to state 3), while no latent load is removed.

A cooling coil can act in concert with the heat wheel and further lower the air dry bulb temperature (state 3 to state 4’).

Alternatively, an evaporative cooler can adiabatically lower the air dry bulb temperature and simultaneously increasing the humidity. (state 3 to state 4).

The evaporative cooling process depicted in Figure 8.1.1 is direct evaporative cooling where water is

sprayed directly into the air stream. The drawback of the direct evaporative cooling process is the added




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humidity in the process air, which can be avoided with indirect evaporative cooling. Figure 8.1.2 depicts

the psychrometric properties of both direct and indirect evaporative cooling.

Figure 8.1.1: Psychrometric chart illustrating the principle of desiccant cooling, (Daou, 2005) annotated

Indirect evaporative cooling uses another air stream, which is cooled evaporatively (called second air

stream), as the heat sink to cool the primary air stream, which is the process air. Typically, a plate heat

exchanger is used to transfer heat between the primary and secondary air stream. Direct evaporative

cooling works best in climates with low sensible heat ratio (SHR), since increasing the indoor humidity

does not compromise thermal comfort due to high relative humidity levels. The indirect evaporative

cooling is the better choice in humid climate since it provides sensible heat removal without adding

moisture into the process air. In regard to cooling efficiency, direct evaporative cooling is more efficient

because of the thermal losses of the secondary air cooling the primary air through the heat exchangers.

Daou (2005) suggests that the effectiveness of direct and indirect evaporative cooling can be as high as

90% and 70 to 80%, respectively. Both processes are effective cooling technologies and can reach a COP

of 5 in dry climates. In humid climates their effectiveness decreases as the maximum temperature drop

is the difference between dry bulb and wet bulb temperatures. Therefore, in order to use evaporative

cooling in humid climates, dehumidification of the supply air is required upstream of the evaporative

cooler, which removes a portion of the humidity in the process air and thus creates the conditions of an

effective evaporative cooling process. Sultan et al (2015) presented the basic working principles of

desiccant cooling in Figure 8.1.3, where the desiccant dehumidifier is removing latent heat and a host of

cooling technologies is rejecting sensible load. In Figure 8.1.3 the desiccant dehumidifier is situated

upstream of the cooling device. This is only necessary, however, for evaporative cooling since it requires




July 9, 2016 Page 118 of 171

dry air to function efficiently. Therefore, especially in humid climate, evaporative cooling process has to

lower the humidity upstream of the evaporative cooling device.

Figure 8.1.2: Psychometric

of direct and indirect

evaporative cooling,

(Daou, 2005) annotated

Figure 8.1.3.

Basic working principle of

desiccant cooling

(Sultan et al, 2015)

Sultan et al (2015) presented a generic desiccant cooling process diagram (Figure 8.1.4) using solid

desiccant dehumidification and a range of cooling technologies. The numbered process steps in Figure

8.1.4 are as follows:




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Figure 8.1.4:

Generic schematic of desiccant cooling systems

The generic system can accommodate various cooling devices and regeneration heat sources

(Sultan et al, 2015), annotated

Process step 1 to 2: The outside air is dehumidified be the desiccant wheel (e.g. rotor) and the air

temperature is increased due to the heat of adsorption.

Process step 2 to 3: The dehumidified air is cooled at constant humidity using sensible heat exchanger

(heat wheel); the heat sink for the heat of adsorption is the return air from the conditioned spaces.

Process step 3 to 4: The air is cooled by one of the following cooling techniques; the air attains the

required temperature for suitable thermal comfort zone conditions. The candidate cooling technologies

include:

(A) direct or indirect evaporative cooling (B) cold coil, e.g. chilled water, (C) M‐cycle cooling,

(D) Evaporator of the VAC.




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Notes:

In case of DEAC cooling the water vapor is added as an isenthalpic process as shown in process 3–4

whereas the rest of cooling is performed at constant humidity that can be noticed from process 3–5.

The M‐cycle, which is an extremely efficient evaporative cooling process can be applied for cooling

towers, especially in humid climate. While an M‐Cycle based cooling tower design is not yet

commercially available, the principles of sub‐wet bulb cooling are getting considerable attention.

These favorable properties make the M ‐cycle attractive for many evaporative cooling applications.

Unlike many evaporative cooling technologies, heat transfer within the M‐cycle is driven by the dew

point temperature gradient, not wet bulb temperature. As such, only the absolute outside ambient

humidity is important, not the temperature of incoming air or water. (Gas Technology Institute, 2010)

Process step 6 to 7: Sensible heat is transferred from the supply to the exhaust air stream. Therefore,

the temperature of the airstream of the exhaust air stream is increased.

Process step 7 to 8: Heat is added for the regeneration of the solid desiccant.

Process step 8 to 9: Desiccant is regenerated (e.g. activated) by heating the desiccant.

Sultan (2013) suggested that the available solid desiccant cooling systems can be categorized as follows:

The type of the solid desiccant system ‐ adsorbent bed or rotating wheel type,

Number of stages ‐ single‐stage or multi‐stage dehumidification type,

Type of rotor ‐ single‐rotor or multi‐rotor type,

Integration into conventional HVAC systems ‐ hybrid system assisted by VAC, and

(Heat source used for regeneration ‐ electricity/gas/solar/waste energy.

TRANE (2005) discussed different system configurations of desiccant wheel and conventional cooling

devices. The three system configurations were wheel upstream of cooling coil, wheel downstream of

cooling coil and series regeneration. Refer to Figure 8.1.4.




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Figure 8.1.4: Typical

desiccant isotherms

(TRANE, 2005)

The case “wheel upstream of cooling coil” is often used in parallel arrangements of desiccant

dehumidification wheels with Type I or Type II desiccants. The desiccant wheel rotates between two

discrete air streams. The regeneration air stream may be the building exhaust or a second outdoor air

stream that is used exclusively to regenerate the desiccant material, where a suitable heat source

provides the heat for regeneration. As a result, water vapor transfers from the higher‐RH process air

(OA) to the lower‐RH regeneration air (RG’).

Figure 8.1.5 and 8.1.6 illustrates this system configuration and the psychrometric state points,

respectively. In the illustration, a second, dedicated, outdoor air stream regenerates the desiccant.

Figure 8.1.5 shows the desiccant wheel removing moisture from the process air stream—but for every

Btu of latent heat (moisture) removed, it adds more than one Btu of sensible heat. Therefore, air leaving

the process side of the wheel (OA’) is dry (at a low dew point) but hot (145 F DB in the case depicted).

Due to the costs of regeneration and re‐cooling, traditional desiccant wheels typically are used only

when the required process‐air dew point can’t be achieved with standard mechanical equipment. The

cost penalty becomes larger if the price of energy for wheel regeneration rises.)

The case “wheel downstream of cooling coil” is depicted in Figure 8.1.7. This wheel configuration applies

more favorably to the operating principles of cooling coils and desiccants. In this wheel configuration,

the process air (OA) first passes through a DX or chilled water cooling coil, where it’s cooled and

dehumidified. The cool, saturated air (CA) passes through the desiccant wheel, which adsorbs moisture

from the high‐RH air thereby lowering the dew point but raising the dry‐bulb temperature. The resulting

conditioned air (CA’) is dry and warm— but not as hot as in the “wheel upstream” configuration (Figure

8.1.5) described earlier. Water vapor transfers from the desiccant to the regeneration air (RG’) as the

wheel rotates into the regeneration air stream.




July 9, 2016 Page 122 of 171

Compared with the “wheel upstream” arrangement, the “wheel downstream” configuration can

dehumidify the process air to an equally low dew point, but requires less or perhaps non re‐cooling

because the resulting dry‐bulb temperature isn’t as hot. This system configuration still requires a

separate regeneration air stream, and that air typically must be heated to dry out the desiccant.

Figure 8.1.5: Desiccant dehumidification

wheel upstream of cooling coil, parallel

regeneration, (TRANE, 2005)

Figure 8.1.6: Desiccant

dehumidification wheel

upstream of cooling coil,

parallel regeneration,

(TRANE ,2005)

The case “desiccant wheel in series with the cooling coil” is depicted in Figure 8.1.9. This system

configuration has the regeneration side of the wheel upstream of the cooling coil and the process side

down stream of the coil. Moisture transfer occurs within a single air stream, where the desiccant wheel

adsorbs water vapor from the process air downstream of the cooling coil and then releases the collected

moisture upstream of that coil, allowing the cooling coil to remove it through condensation. A separate,

regeneration air stream is not needed. This wheel placement requires a Type 3 desiccant selected

specifically for this application, since desiccant must adsorb water vapor when the relative humidity of




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the air is high. When the RH is below 80 %, moisture adsorption ability drops significantly. Therefore,

adsorption is more driven by the Type 3 desiccant’s ability to regenerate at low temperatures than by

hot regeneration air, often without supplemental heat. A preheat coil can be added upstream of the

regeneration side of the wheel for applications that require even drier air.



downstream of cooling coil, parallel

regeneration, (TRANE, 2005)



downstream of cooling coil,

parallel regeneration,

(TRANE, 2005)

Adding the series configuration of the desiccant wheel changes the dehumidification performance of the

traditional cooling coil, trading sensible capacity for more latent capacity. The latent (dehumidification)

capacity of the cooling coil increases while the total cooling capacity (enthalpy change across the coil)

remains the same.




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Figure 8.1.9: Desiccant dehumidification wheel desiccant wheel in series with the cooling coil, (TRANE

,2005)

8.2 Solid Desiccant Cooling with Evaporative Cooling

In an early study Collier et al (1981) presented three basic process modes or cycles. These basic modes

and cycles have been used in many desiccant cooling applications:

Ventilation mode

Recirculation mode

Dunkle cycle

The ventilation mode is illustrated in Figure 8.2.1. In this mode, room air is used to regenerate the

dehumidifier bed and outdoor air is cooled. The air leaving the room (6) is evaporatively cooled (7) and

used as the cold sink for the dried room‐return air. The room‐exit air is heated during the heat

exchange. (8). The return air is then further heated by an external source of energy (Q) for desiccant

regeneration (9). Drying the desiccant cools and humidifies the air (10). Ambient air (1) is dried by the

desiccant (2) to supply room make‐up air and cooled by heat exchanger (3) with room‐exit air. The dried

and cooled ambient air is further chilled by evaporative cooling (4). Just before point (5), remix air is

introduced. This scheme mixes evaporative cooled room dried room make‐up air in order to control the

sensible heat factor.




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Figure 8.2.1: Ventilation mode (Collier et al, 1981), annotated The locations in the block diagram (a) correspond with state points in the psychrometric chart (b)

The recirculation mode shown in Figure 8.8.2. This mode uses the same components as the ventilation

mode except that room air is constantly reconditioned in a closed loop and outdoor air is used for

regeneration. Thermodynamically, this cycle has the advantage of processing air with greater cooling

capacity. The disadvantage is having a warmer cold‐sink temperature than the ventilation mode.

Selecting either mode is a trade‐off which depends upon room and ambient conditions. This trade‐off

will be discussed in greater detail later in this literature review. Another important difference between

this mode and the ventilation mode is that there is no direct fresh‐air supply, whereas the ventilation

mode uses all fresh air. For the recirculation mode, as well as for most vapor‐compression cycles, fresh

air to the building space is supplied by normal infiltration. In an era of tighter buildings, this fact is a

disadvantage, since air intake has to be controlled.

The Dunkle cycle shown in Figure 8.2.3. This process cycle combines the thermodynamic advantages of

both the ventilation and recirculation modes. The cycle uses the advantage of processing the higher

cooling availability room air as in the recirculation mode, while retaining the lower cold‐sink

temperature of the ventilation mode. This advantage in performance comes at the cost of increased

complexity and an additional sensible heat exchanger. As with the recirculation mode, the lack of

controlled fresh air to the building space is a disadvantage with tighter building envelopes.




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Figure 8.2.2: Recirculation mode; (Collier et al, 1981), annotated

The locations in the block diagram (a) correspond with state points in the psychrometric chart (b)

Pesaran and Hoo (1993) discussed desiccant cooling system where air is passed through a desiccant

dehumidifier in order to increase the performance of the downstream evaporative cooling. Figure 8.2.4

shows the desiccant cooling ventilation cycle which used a rotary desiccant dehumidifier, a heat

exchanger, two direct evaporative coolers, a desiccant regeneration heater, and ancillary equipment

such as fans and pumps. In this cycle outside air is dried in the dehumidifier and then cooled by

evaporative coolers and supplied to the conditioned space.




July 9, 2016 Page 127 of 171

Figure 8.2.3: Dunkle cycle;

(Collier et al, 1981), annotated

The locations in the block

diagram (a) correspond with

state points in the psychrometric

chart (b)

Figure 8.2.4: Evaporative cooling system, (Pesaran and Hoo, 1993), annotated




July 9, 2016 Page 128 of 171

The working process of the system depicted in Figure 8.2.4 is described in the following with

annotations referring to important process steps.

Drying of the outside air (1): Outside air is dried by a desiccant wheel which is reactivated by the

reheated exhaust air (process step (8)). Downstream of the desiccant wheel the air is lower in

humidity but warmer than the outside air, due to the heat of adsorption of the solid desiccants.

Cooling of the outside air (2): The conditioned outside air flows through an air‐to‐air heat

exchanger, such as a heat wheel, where heat is transferred from the outside air flow to the

return air.

First stage of evaporative cooling (3): The outside air passes through a direct evaporative cooler

where sensible heat is adiabatically removed from the air by the cooling effect of water droplets

vaporizing in the air stream. The resulting colder and moister supply air is provided to the

conditioned spaces.

Conditioned space (4): In the conditioned space sensible and latent load is added to the air,

which is vented out of the conditioned space.

Second stage of evaporative cooling (5): The return air passes through a direct evaporative

cooler where sensible heat is adiabatically removed from the air. This process results in a colder

and moister air downstream of the second stage evaporative cooler.

Warming of the return air (6): The air‐to‐air heat exchanger is cooled by the return air, thereby

adding sensible heat to the outside air.

Regeneration heat supply (7: Thermal heat is provided to a portion of the return air thereby

heating up the air flow that passes through the desiccant wheel. A remaining portion of the

return air bypasses the desiccant wheel which provides for a precise heating of the desiccant

wheel and avoids over heating of return air and the desiccants.

Drying of the desiccant wheel (8): The heated air flow through the desiccant wheel dries the

desiccant matric of the wheel and provides adsorption capacity for the outside air (1) as the

reactivated portion of the desiccant wheel is available for the outside air passing through the

wheel matrix.

The advantage of the desiccant cooling cycle depicted in Figure 8.2.4 is that the system works in climate

zones with high humidity levels. In humid higher climate zones evaporative cooling becomes less

effective as a higher wet point of the outside air limits the removal of sensible heat. In addition, in

humid climate zones humidity has to be removed from the outside air as it is supplied to the

conditioned spaces. The drying of the outside air flow upstream of the evaporative cooler lowers the

humidity level and therefore enables more effective evaporative cooling.




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The level of dehumidification upstream of the evaporative cooling has to be higher in the case of

indirect evaporative cooling than for direct evaporative cooling systems. The indirect evaporative

cooling process does not add humidity to the air, which is an advantage. However, since the

effectiveness of indirect evaporative cooling is lower than the direct process only detailed design

analysis can ascertain whether direct or indirect evaporative cooling cycle is the best match for the

desiccant system.

8.3 Solid Desiccant Cooling to Increase Conventional AC System

Pesaran (1993) presented an application of an indirect evaporative coolers working in conjunction with

two types of solid desiccant systems. The entire system was implemented to augment the cooling

capacity of a conventional VAV system. The author presented a design strategy to add cooling capacity

and enhance ventilation rates to the conditioned space. Rather than adding approximately 500 tons of

conventional chiller capacity to the existing building. the author proposed to add a desiccant cooling

system for pre‐conditioning the supply air. The proposed desiccant cooling pre‐conditioning process is

described hereafter.

Figure 8.3.1 depicts the existing AC system which was undersized for the actual load.

Figure 8.3.1. Schematic of an Existing All‐Air VAV System of an Office Building; (Pesaran, 1993) The figure shows the existing AC system which was undersized to provide 15.0 and 2.5 BTUH/sqft of sensible and latent load. In addition, the existing 0.1 CFM/sqft ventilation rate was deemed as not sufficient.




July 9, 2016 Page 130 of 171

Figure 8.3.2. shows the proposed addition of a pre‐conditioning outside air supply to the existing AC

system. The added desiccant cooling system included two solid desiccant systems and an indirect

evaporative cooler.

Outside air enters the system and the pressure is increased in order to overcome pressure

losses of the air as it passes through two desiccant wheels and one indirect evaporative cooler.

The pressurized outside air supply passes through an enthalpy wheel where humidity and heat is

transferred from the outside air to the return air (e.g. exhaust air).

Downstream of the desiccant wheel the air flow is separated into two flow paths. A portion of

the air flow passes through desiccant dehumidifier (e.g. desiccant wheel) while the remaining

portion bypasses the desiccant wheel. The humidity of the air flowing through the desiccant

wheel is transfers to the solid desiccants of the wheel.

The solid desiccant material of the desiccant wheel is activated (e.g. regenerated) by a separate

outside air flow that is heated by a natural gas furnace. An air‐to‐air heat exchanger pre‐heats

the outside air upstream of the natural gas furnace.

Downstream of the desiccant wheel an indirect evaporative cooler provide sensible heat

removal to the air flow which serves as the preconditioned outdoor air supply.

The system is depicted in Figure 8.3.3. It provides an example of how desiccant cooling can enhance

existing conventional AC‐systems in retrofit applications. In the early 1990s a desiccant cooling system

that enhances conventional HVAC installations was first introduced and sold by ICC Technologies, Inc.

The system used a desiccant drier, heat exchanger, and evaporative coolers in series to dry and then

cool the air. The systems were marketed as alternatives or supplements to more traditional commercial

air conditioning units, largely in light commercial and supermarket systems. The type of desiccant

material used in the system was a so‐called Engelhard Titanium Silicate (ETS), which is suitable for low‐

temperature applications. One form of ETS desiccant could be regenerated at less than 140°F, so waste

heat from condensing units of electric chillers could be used. Figure 8.3.3 shows ICC's DESI/AIR System

which with a desiccant wheel, heat‐exchanger wheel, and evaporative cooling pads, and can be used as

a preconditioner for ventilation air. (ICC Systems 1993).




July 9, 2016 Page 131 of 171

Figure 8.3.2: Schematic of an Existing All‐Air VAV System of an Office Building; (Pesaran, 1993), annotated




July 9, 2016 Page 132 of 171

Figure 8.3.3: ICC Technologies DESI/Air System (ICC Technology, 1993)




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Daou et al (2005) suggested that desiccant cooling in conjunction with conventional cooling coils can be

successfully used in radiant cooling applications. The author pointed out several generic benefits of

radiant systems. Such radiant cooling systems include metal ceiling panels, chilled beams, and tube

embedded ceilings, wall and floors (e.g. also called core cooling). Benefits suggested by the author

included energy savings in the range of 27 to 37%, due to a significantly reduced ventilation rate, the

lower energy demand of pumping chiller water in lieu of air and the decoupling of ventilation to sensible

heat rejection. The author also stressed benefits of radiant cooling on occupant comfort.

The block diagram and configuration of the proposed system is depicted in Figure 8.3.4 (a) and (b),

respectively. The system included desiccant a wheel and a heat wheel which dehumidify and provide

sensible cooling to the incoming air. The supply air had to be dried in order to prevent condensation on

the radiant surfaces and a resulting reduction in comfort. Chilled water discharge from the cold coil was

circulated through the chilled‐ceiling system. The incoming air was dehumidified by a desiccant wheel

and pre‐cooled by heat wheel before been cooled further by the cold coil to the supply temperature.

The author suggested that sensible load is entirely handled by the chilled‐ceiling radiant cooler while the

latent load is extracted by the desiccant. The use of the desiccant wheel is of very importance for

comfort point of view if this system is to be used in a hot and humid climatic zone.

Daou (2005) stated several advantages of chilled‐ceiling itself and the integration of the decoupled

desiccant assisted ventilation and sensible heat rejection as follows:

The sensible and latent loads are handled independently, realizing the so‐called decoupled

cooling approach.

The absence of reheating allows great energy saving, the author suggested about 44% of

primary energy savings in comparison with the vapor compression

A single water chiller is used since water is circulated in series in the cold coil and the cooling

panel. The temperature of chilled water can be as high as 17 °C, which increases the COP of the

chiller

The system can be driven by low grade energy, especially the free energy such as solar energy

and waste heat.

The system delivers a draft and noise free conditioned air.




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(a) Block diagram of the desiccant cooling system The desiccant cooling system provides latent heat removal and ventilation

(b) Sensible heat rejection from the conditioned space through a chilled ceiling The discharge of the cooling coil is cold enough to serve as the chilled water supply of the chilled ceiling

Figure 8.3.4: Desiccant cooling integrated into radiant cooling systems (Daou, 2005)




July 9, 2016 Page 135 of 171

8.4 Solid Desiccant Colling with Solar Heat

Parmar and Hindoliya (2011) suggested that solar energy is a suitable option for regeneration of

desiccant cooling system and thus saves the regeneration energy input. The authors suggest the use of

evaporative cooling techniques in humid climate, with a solid desiccant as the dehumidifier. Figure 8.4.1.

illustrates the basic approach of using solar heat for the activation of the desiccant material.

Figure 8.4.1: Basic approach of using solar heat for the activation of the desiccant material. (Parmar and Hindoliya, 2011)

This literature search revealed an increasing interest in the use of solar heat for desiccant cooling. Three

examples of solar supported desiccant cooling applications are presented hereafter.

Sultan et al (2015) proposed hybrid solar thermal and electric driven desiccant cooling system that is

shown in Figure 8.4.2. During night time the system uses stored thermal energy that was harvested

during the day. This electric back‐up system made the system cost effective by using relatively lower

price off‐peak electricity and prevented uncertainty by preventing total reliance on solar energy. The

system was tested and the performance was satisfactory.

The outdoor and indoor conditions were set to 30°C and 60% RH and 26°C and 55% RH, respectively.

Effect of desiccant regeneration temperature on the system performance, measured in COP is shown in

Table 8.4.1. The results suggest that the dehumidification performance by the system was improved

with the increase in regeneration temperature, but this came with the penalty of increased supply air

temperatures. The system COP decreased from 0.44 to 0.35 with the increase in regeneration

temperature from 60 to 75°C. The authors concluded that the thermal heat supplied to the desiccant AC

system is not exactly proportional to the actual cooling effect and that an optimum regeneration air

temperature is required to obtain efficient performance of a desiccant AC system.




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Figure 8.4.2: Schematic diagram of solar thermal and electric driven desiccant AC‐system (Sultan, 2015)

Table 8.4.1: Response of COP to regeneration temperature, for system depicted in Figure 8.4.2; (Sultan, 2015)




July 9, 2016 Page 137 of 171

Archibald (2001) presented a solar hybrid desiccant cooling system which is depicted in Figure 8.4.3. The

system includes a solar thermal system, a desiccant evaporative cooler, one additional direct

evaporative cooler, one indirect evaporative cooler and one hot water heating system using the waste

heat from the desiccant regeneration. The psychometrics of the integrated cycle are shown in Figure

8.4.4. The cycle shown does not use recovered heat from the heat exchanger for desiccant regeneration.

Instead, the excess solar heated air from the solar thermal tile roof is used as the primary heating source

for desiccant regeneration.

8.4.3: Solar roof integrated desiccant cooling system schematic,

(Archibald, 2001)

8.4.4: Solar roof integrated desiccant cooling system psychometrics

(Archibald, 2001)




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Sultan (2015) discussed a solar driven two‐stage desiccant cooling system (TSDC), assisted by a VAC

system to condition an office building (see Figure 8.4.5). The desiccant cooling unit had a design cooling

capacity of 10kW, a VAC unit with nominal cooling capacity of 20 kW and a flat plate solar collector array

of 90 m2. Internal heat exchangers were used to reduce the adsorption heat in the process air flow. The

system proved that it can provide sufficient cooling and dehumidification even at high temperatures and

high humidity level, such as outside conditions of 35°C and a humidity ratio of 21.54 g/kg.

The two‐stage dehumidification provided higher air flow rate as compared to single‐stage at a

comparable dehumidification level. Another benefit of the multi‐stage dehumidification is that the

desiccant regeneration temperature can be decreased, which increases the overall COP. The fact that

lower regeneration temperatures are used makes the system more suitable for utilizing low grade waste

heat and solar energy for desiccant regeneration.

Figure 8.4.5: Schematic diagram of solar driven two‐stage hybrid DAC system, (Sultan, 2015)


SECTION 9 ‐ LIQUID DESICCANT COOLING SYSTEMS


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Liquid desiccants have been in use in various industrial applications. Their use in conjunction with

building cooling applications, however, has been very limited, since the solid desiccant dehumidifiers,

such as the rotary desiccant devices, have become HVAC industry standards. ES Magazine (2010)

suggested that various companies had proposed liquid desiccant dehumidification for HVAC applications

in the past. Although these systems had produced very impressive EER and COP, these systems were

small and few were installed, therefore these systems never overcame the critical mass for larger scale

adoption. The authors stated, however, that there is a renewed interest in liquid desiccant technology

for HVAC systems and these new liquids desiccant systems are showing very encouraging performances

characteristics. This section presents several of recent liquid desiccant cooling applications.

9.1 General System Configuration of Liquid Desiccant Cooling Systems

Sahlot and Riffat (2016) suggested advantages and disadvantages of liquid desiccant cooling systems as

follows:

Advantages

(1) Low‐pressure drop across the liquid desiccant system makes them suitable to use with low

regeneration temperatures.

(2) The ability to pump liquid desiccants makes the entire unit small and compact.

(3) Liquid desiccants can be stored and used when heat source is not available. This is advantageous

when heat source is only intermittently available for desiccant regeneration.

Disadvantages

(1) Liquid desiccants like lithium chloride, lithium bromide and all other salts are corrosive and can

damage the desiccant system.

(2) Any carry‐over of liquid desiccant along with supply air stream can cause harm to the health of

the occupants.

(3) In order to handle large volume of liquid desiccant, large pumps are required, which draws a

large amount of power.

(4) Desiccants of aqueous salts also face the problem of crystallization.

In the basic liquid desiccant dehumidification process moisture from the process air is removed in the

dehumidification or absorber unit, where the desiccant absorbs the water vapor from the process air

due to a difference in vapor pressure between the desiccant solution and process air. While water vapor

is absorbed, the temperature of liquid desiccant rises due to heat of absorption and condensation being

liberated. The diluted desiccant is pumped back to the regenerator for regeneration. Downstream of the




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regeneration the diluted solution passes through a plate heat exchanger for heat recovery and a heating

coil, where its temperature of the desiccant solution is raised. In the regenerator, the hot diluted

solution is exposed to regenerative air stream, and moisture is transferred from the weak solution to air

due to the difference in vapor pressure.

Lowenstein (1998) discussed the evolution of industrial liquid desiccant dehumidifier and suggested that

liquid desiccant devices used in conjunction with HVAC would have the same basic working scheme as

industrial liquid desiccant systems. The author pointed out that the liquid desiccant would have to be

adapted to the level of the target dew point and other issues, such as carry over and desiccant

isotherms. Figure 9.1.1 shows a simple industrial desiccant system configuration.

Figure 9.1.1.: Simple industrial desiccant system configuration (Lowenstein, 1998)

Figure 9.1.2 shows a generic process flow diagram of a liquid desiccant system comprised of

dehumidification and regeneration units. Figure 9.1.3 depicts the psychrometric performance of the

liquid desiccant process, showing the state points (A) through (D) corresponding to the locations in

process diagram in Figure 9.1.2. Figure 9.1.3 indicates that the vapor pressure of the liquid desiccant

solution decreases with increasing concentration.




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Figure 9.1.2.: Generic liquid desiccant process diagram (Sahlot and Riffat ,2016)

Figure 9.1.3.: Psychrometric process (Sahlot and Riffat ,2016) The process states (A) through (D) correspond to the locations A through D in the process diagram

The process vessels for liquid desiccant dehumidification are classified with regard to internal geometry,

such as finned‐tube surface, spray and packed tower and their heat extraction process, adiabatic and

internally cooled dehumidifiers. Figure 9.1.4 (a) and (b) shows an adiabatic and internally cooled

dehumidifier, respectively. Sahlot and Riffat (2016) discuss advantages of multiple stage

dehumidification and suggest that the effectiveness of the entire system can be significantly improved

by adding successive effects. The authors suggest that the successive stages can be optimized not only

in size but also be using different liquid desiccant systems.




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(A) Adiabatic dehumidifier: Process heat removal occurs through air passing through the tower. Higher flow rates have disadvantages.

(B) Internally cooled dehumidifier: Process heat removal occurs through cooling device inside the tower.

Figure 9.1.4: Liquid desiccant dehumidification – type of heat removal (Sahlot and Riffat ,2016)

Kassem (2013) investigated three common flow patterns in an adiabatic dehumidifier: parallel flow,

cross‐flow and counter flow, (see Figure 9.1.5) through simulations and experimental verification. Flow

patterns of liquid desiccant and contact air determine the contact area and the process of interaction

between desiccant and airflow. The author concluded that the performance of counter flow is best

followed by cross‐flow, while the performance of parallel flow is not optimum.

Figure 9.1.5: Flow patterns of air and liquid desiccant in dehumidification unit, (Kassem, 2013)

Figure 9.1.6 illustrates basic process characteristic of liquid desiccant cooling. A liquid desiccant system

used in cooling applications has several main components namely the dehumidifier (also absorber or

conditioner), the regenerator (also desorber), the regeneration heat source and the cooling unit. Packed




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columns or beds are the most frequently used technology for these components. The dehumidification

process (sorption) is carried out by spraying the liquid desiccant into the process air to absorb the

moisture out of the air. The liquid desiccant falls to a sump and is pumped back to the top of the process

vessel where nozzles to be sprayed back to the air and onto the packing.

As the desiccant solution becomes diluted as a result of absorbing moisture out of the air stream the

solution must be regenerated. Liquid desiccant material continuously circulated through the regenerator

where an external heat source drives desorption. Also, the desiccant solution is heated before any

connection with air to raise the partial pressure of the desiccant above that of the of the air. Thus, the

moisture content of desiccant is transferred to the regeneration air (desorption), which leaves the

regenerator as hot and humid air. After the regeneration process, the liquid desiccant solution becomes

concentrated and free from moisture content and must be cooled to lower the vapor pressure. Before

being sprayed into the process air again to complete the cycle, the liquid desiccant solution is cooled by

a chiller or cooling tower to adjust the temperature to the desired level (cooling). Sensible cooling is

provided via employing traditional vapor compression and vapor absorption, direct or indirect

evaporative coolers.

(a) Schematic of a desiccant cooling air conditioning (b) Moisture removal process by desiccant

Figure 9.1.6: Basic process characteristics of liquid desiccant cooling, (Abdulrahman, 2013)




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9.2 Liquid Desiccant Systems to Enhance Conventional AC System

Unlike solid desiccants, liquid desiccant dehumidification systems have not become a standard design

component to augment conventional AC systems using vapor compression cooling. The literature

reviewed presented some liquid desiccant dehumidification applications, which will be presented in the

following.

Daou (2006) presented a vapor compression air conditioning which is aided by a liquid desiccant cooling

system. Figure 9.2.1 shows the system process diagram. Cool strong desiccant solution is sprayed onto

the top of the dehumidifier. Since desiccant solution vapor pressure is less than that of the air vapor

pressure, water vapor migrates from the air stream to the desiccant solution and condenses therein.

Consequently, the process heat liberated increase in the temperature of the solution. The process air

stream is slightly cooled down due to its contact with the cold desiccant solution. The dehumidified and

warm process air stream then passes successively through the evaporative cooler and the evaporator of

the traditional refrigerant vapor compression AC‐unit, before the air is delivered to the conditioned

space. The diluted desiccant solution, which exits the dehumidifier, is circulated through the regenerator

where it is heated. The moisture absorbed in the dehumidifier unit is now lost to the scavenger air

stream. The hot and strong desiccant solution is thereafter cooled down in the pre‐cooler and then

cooled further in the heat exchanger (HX) before being ready again to dehumidify the incoming process

air.

Because the latent load is handled independently by the desiccant dehumidifier, ventilation air does not

have to be cooled below its dew point. The temperature of evaporation can thus be lifted up to 15 °C

from its generally level of 5 °C for cooling‐based dehumidification in the traditional vapor compression

system. This saves energy since reheating is not required and the increase in cooling coil temperature

increases the system’s coefficient of performance (COP).

AIL (2016) presented an innovative and highly effective integration of liquid desiccant into a vapor

compression cycle. This system is the Liquid‐Desiccant Direct‐Expansion air conditioner (LDDX), which is

especially suited for hot and moist outdoor conditions. The system is described as an efficient, high

latent cooling system that can adjust its Sensible Heat Ratio (SHR) between 0.35 to 0.75 and still retains

a high overall efficiency with minimal losses in its rated total cooling capacity. The electric efficiency

stays close to its projected value of 11.4 EER. AIL stated that the LDDX system is the first embodiment of

a compressor‐based liquid‐desiccant air conditioner in which a solution of lithium chloride floods the

surfaces of the air conditioner’s evaporator and condenser providing direct contact between the

desiccant and the air flowing through these two coils.




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Fig. 9.2.1: Schematic of liquid desiccant aided vapor compression air conditioning, (Daou, 2006)

Figure 9.2.2 (a) and (b) shows a 3D depiction of the LDDX system and a liquid desiccant flow diagram,

respectively. Figure 9.2.2 (a) shows a strong desiccant applied to the top of a wicking‐fin evaporator and

weak desiccant to the top of a wicking‐fin condenser. The strong desiccant absorbs water vapor from

the process air flowing through the evaporator. The weak desiccant flowing off the evaporator is

warmed in the interchange heat exchanger before it is delivered to the top of the condenser. The heat

rejected in the condenser further warms the weak desiccant which then desorbs water to the cooling air

that flows through the condenser. The warm, strong desiccant that flows off the condenser is cooled in

the interchange heat exchanger before it is supplied to the evaporator.

The essential characteristic of the LDDX is its ability to supply cool, unsaturated air without reheat.

Whereas a conventional DX air conditioner without reheat might have a sensible heat ratio (SHR) of

0.75, the LDDX can have an SHR as low as 0.35. The SHR adjustment is achieved by varying the amount

of liquid desiccant that is recirculated over the evaporator. Figure 9.2.2 (b) illustrates a diverting valve

which controls the desiccant flow from the evaporator to the condenser.

The essential characteristic of the LDDX is its ability to supply cool, unsaturated air without reheat and

to avoid overcooling when the latent load is high, e.g. with low SHR. In applications where high dew

point air is processed, for example Dedicated Outdoor Air Systems operating in humid locales, the most

efficient cooling system would be a two stage system, with a first‐stage conventional DX air conditioner




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that delivers saturated, partly cooled air to a second‐stage LDDX, which dehumidifies the air. Such a two

stage system in shown in Figure 9.2.2.

(a) 3D‐rendering of the LDDX system

Showing the refrigeration and liquid desiccant flow

system

(b) Flow system of the liquid desiccant loop

shoeing the diverting valve for SHR adjustment

Figure 9.2.2: Description of the single stage LDDX system (AIL, 2016)

Figure 9.2.3: Description of a two‐stage LDDX system (AIL, 2016)




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Sahlot (2016) discussed a liquid desiccant system which was combined with a vapor compression and

demonstrates that such an integrated system can be highly efficient in space cooling with possible

savings in primary energy of up to 50%. The COP of the system tested was 3.8, with a total cooling

capacity of 6.15 kW, using 2.6 kW VCS. Lithium chloride was the liquid desiccant solution. Figure 9.2.4

shows the schematic diagram. Strong solution from the tank is pumped and sprayed uniformly over the

evaporator surface area. Process air is passed through the evaporator. The evaporator and desiccant are

engaged in simultaneously cooling and dehumidifying the process air while the diluted desiccant

solution is collected in the weak solution tank. The diluted solution is then pumped to absorb heat from

the heat exchanger that uses the waste heat rejected from the condenser of the vapor compression

system to preheat the diluted solution. A heating coil in the regenerator tank provides additional heat

for regeneration.

Figure 9.2.4: A schematic of hybrid desiccant‐assisted air conditioner, (Sahlot, 2016)

Kassel (2013) demonstrated a simple integration of a liquid desiccant dehumidifier into a vapor

compression air‐conditioning cycle. The hybrid system is depicted in Figure 9.2.5.




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Such a system can succeed in obtaining high thermal performance and energy savings since waste heat

is recovered internally and used for the desiccant regenerator to recover waste heat. Although the

performance of the regenerator is not as good as the regenerator using the heating desiccant solution,

this system uses free heat from waste hot air.

In order to overcome intrinsic disadvantages of using waste heat, the author discusses the importance

of identifying the optimum inlet parameters that drive the rate of evaporation of water vapor from the

regenerated liquid desiccant, a was solution of CaCl2. The main inlet parameters considered are the inlet

air temperature, solution flow rate, solution inlet concentration, airflow rate and humidity of inlet air.

The author reported that the regeneration process is highly dependent on the air inlet temperature,

humidity, and flow rate. An increases in air velocity from 0.5 to 1.5 m/s resulted in the overall mass‐

transfer coefficient in the structured packing dehumidifier and regenerator from 4.0 to 8.5 g/m2 s and

from 2.0 to 4.5 g/m2 s, respectively. Higher solution temperature resulted in lower overall mass‐transfer

coefficients.

Figure 9.2.5: An energy‐efficient air conditioner with a desiccant solution regenerator using exhausted hot air from the condenser (Kassel, 2013)

Jradi and Riffat (2014) demonstrated the feasibility of desiccant supported air‐conditioning in a tri‐

generation system powered by biomass gasification. The tri‐generation system used was a combined

heating, cooling and power system, which was operated using gas turbine, internal combustion unit

(Figure 9.2.6). Regeneration heat for the liquid desiccant was obtained from the waste heat of the gas

turbine that ran with gasified biomass.




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Figure 9.2.6: An internal combustion‐based tri‐generation system with a biomass gasification unit, (Jradi and Riffat, 2014)

9.3 Liquid Desiccant Cooling with Evaporative Cooling and Solar Energy

Liquid desiccants perform the same basic function as their solid desiccant counterpart in integrated

cooling applications with evaporative cooling. The main function is to regulate humidity levels of the

supply air to the conditioned spaces. In climates with higher humidity levels, the effectiveness of

evaporative cooling is impeded and, in the case of direct evaporative coolers, adding more moisture to

the already humid outdoor air would create uncomfortable, or even unhealthy indoor conditions. This

section presents several desiccant cooling systems where sensible load is removed by means of

evaporative cooling. The systems are examples of basic systems designs. A very innovative and ground

breaking approach to evaporative air‐conditioning is presented in the Chapter 9.4. The use of

evaporative cooling in conjunction with liquid desiccant dehumidification will be illustrated with

examples of solar and other heat sources for regeneration.

The basic concept of liquid desiccant based evaporative cooling, either direct or indirect, is illustrated in

Figure 9.3.1 and 9.3.2. Figure 9.3.1 illustrates the concept of liquid desiccant cooling, which is two‐stage

process of removing sensible and latent loads. A mixture of outside and return air first passes through




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first stage which is the liquid desiccant dehumidifier where humidity is reduced. Then, in the second

stage evaporative cooling reduces the air temperature, which means it reduces the outside air dry bulb

temperature and the heat of absorption from the first stage desiccant dehumidification.

Figure 9.3.1: Concept of liquid desiccant based evaporative cooling system), (Buker, 2015)

Legend:

Figure 9.3.2 illustrates the difference between direct and indirect evaporative cooling. The main

difference being that direct evaporation increases the humidity in the air whereas the indirect process

does not, since the heat transfer occurs through an impermeable membrane, such as a metal plate. For

both types of evaporative cooling, however, the importance of the first stage dehumidification is that

humidity levels are reduced so that evaporative cooling can be effective, especially in humid climate.

The use of solar heat for regeneration is of significant importance and many system developments of

liquid desiccants cooling use of solar or some other low grade waste heat source for desiccant

regeneration. Advantages of liquid desiccant applications using solar heat is pointed out by a number of

researches and HVAC practitioners. For example, Enteria and Mizutani (2011) discussed that liquid

desiccant regeneration can be accomplished with at a temperature as low as 40–50 °C, thus at

temperatures that can be easily obtained by solar thermal devices, such as flat plate collector. That fact

that regeneration temperatures can be below 80 °C is one of the most favorable features of the liquid

desiccant systems. The authors state that carry‐over of desiccant droplets to the supply air is one of the

main concerns of the liquid desiccant cooling systems. Liquid desiccant solutions are usually corrosive

and simple handling of the working media is difficult. These design problems, however, are currently

being solved thanks to fast technology advancement. The authors conclude that the potential of the

liquid desiccant is remarkable.




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(a) (b)

(a) Indirect evaporative cooler has two air streams: the primary air stream is cooled by heat transfer over metal

plates of an air‐to‐air heat exchanger and rejects heat to the secondary air stream, e.g. there is only heat

transfer and no mass transfer (b) direct evaporative cooling has only one air stream: the secondary air stream

has an adiabatically evaporation of water which results in significant temperature drop but also increase in

humidity in the air stream; (b) direct evaporative cooler is cooled by water vaporizing and lowering the

temperature, but the humidity level of the primary air stream increases.

Figure 9.3.2: Concept of liquid desiccant based evaporative cooling system and illustration of direct and indirect

evaporative cooling; (Buker, 2015)

Buker and Riffat (2015) discussed a solar liquid‐desiccant air conditioner (LDAC) coupled with a direct

evaporative cooler for the climate in Queensland, Australia. The system is shown in Figure 9.3.3. The

system used a high effectiveness cross‐flow polymer plate heat exchanger as an absorber. The liquid

desiccant agent was a lithium chloride solution which was regenerated by using hot water derived from

flat plate solar collectors. The process air was dehumidified by strong desiccant solution and then the

warm and dry air was cooled and humidified through the direct evaporative cooler. The supply air was

then provided to the conditioned space. The test results revealed that electrical COP of the system was

around 6 with 3.5 kW electrical energy consumption and the effectiveness of the conditioner with the

desiccant solution was about 82%.

Buker and Riffat (2015) discussed a prototype of a solar driven liquid desiccant systems which was

operated in the hot and humid tropical climate of Thailand. The system (Figure 9.3.4) used only solar

energy for the regeneration of liquid desiccant. The authors pointed out that the system was capable of

reducing the temperature by 1.2 °C and relative humidity by 11.1%. The reason for the limited

performance was the performance of the conventional cooling tower. Since the ambient air was humid,

the cooling of process water in the cooling tower was only limited.




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Figure 9.3.3:

Schematic diagram of

the solar LDAC system

in Queensland,

Australia, (Buker and

Riffat, 2013)

Figure 9.3.4: Schematic diagram of the solar regenerated liquid desiccant system, (Buker and Riffat, 2015)

Das et al (2013) presented a small capacity solar liquid desiccant cooling prototype system which was

tested in the laboratory (See Figure 9.3.5). One of the main differences to other liquid desiccant cooling




July 9, 2016 Page 153 of 171

systems was the intent to avoid carry‐over of liquid desiccant solutions to the supply air. The system

provided 100% outdoor air and consisted of a dehumidifier or absorber (b), a regenerator or desorber

(i), an indirect evaporative cooler (d), a cooling tower (f), two water‐solution heat exchangers (g, j), and

one solution to solution heat exchanger (h). Hot and humid ambient air (a) entered the unit and passed

through a dehumidifier (b). In the dehumidifier (also called an absorber), concentrated liquid desiccant

absorbed water vapor from the process air. The air and the liquid channels were separated by

semipermeable microporous membranes. Therefore, only the water vapor passed through the

membranes into the liquid desiccant when the vapor pressure of the air exceeds that of the liquid

desiccant. The authors emphasized that indirect contact between desiccant and air stream in the

dehumidifier eliminates carryover of the liquid desiccant into the supply air.

Downstream, the dehumidified air was first cooled by a fin‐tube air‐water heat exchanger and an

indirect evaporative cooler (d) and then supplied to the room via a fan (e). The indirect evaporative

cooler used part of the dehumidified process, while the recirculated air got humidified and exhausted to

the atmosphere. Moisture‐laden dilute desiccant from the dehumidifier was preheated in the solution

heat exchanger (h) by recovering heat from the hot concentrated desiccant leaving the regenerator and

stored in a tank (o). The preheated dilute desiccant was then pumped to the plate heat exchanger (j)

where it was further heated by hot water. Water was heated in solar collectors (k) (Evacuated tube

Collectors, Heat Pipe Collectors) and stored in a hot water storage tank (l). The hot dilute desiccant

flowed through the regenerator (i), where it released water vapor to the air stream owing to the vapor

pressure difference. The concentrated desiccant from the regenerator was stored in a tank (n) and then

precooled in the solution‐solution heat exchanger (h). The solution was then pumped to the plate heat

exchanger (g) where it was cooled by water from the cooling tower. Cooling increased the vapor

pressure difference between air and concentrated desiccant, resulting in higher dehumidification

efficiency. Das et al (2013) stated that the system performances need some further optimization to

lower thermal and pressure losses.

Since the liquid desiccant cooling is well suited for air at higher humidity level, air at lower humidity

levels does not necessarily need dehumidification upstream of the evaporative cooler to function

effectively. The level of humidity in the air can either by a characteristic of region or of season. For

instance, a region with constantly high humidity levels in the summer or during periods in the year could

have dryer air conditions for the rest of the year. These changes would require that the liquid desiccant

systems are suitable to adjust for different outdoor air conditions. Kim et al (2014) discussed an air

conditioning system which offered flexibility by engaging the different main systems components.




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Figure 9.3.5: Prototype of solar liquid desiccant cooling prototype system (Das, 2013)

Kim et al (2014) discussed the performance enhancement of an indirect and direct evaporative‐cooling‐

assisted 100% outdoor air system (IDECOAS) by a liquid desiccant dehumidification unit. The authors

called the liquid desiccant enhanced IDECOAS the LD‐IDECOAS. The authors state that prior studies had

indicated that an IDECOAS could save as much as 50% of the energy requirement for a conventional

variable air volume (VAV) system. However, during a hot and humid summer, the energy savings

potential of the IDECOAS was significantly reduced by the low process effectiveness of the evaporative

cooling units. A desiccant system to precondition the incoming outdoor air upstream of the IDECOAS

was added to mitigate high humidity effects on the evaporative coolers. Figures 9.3.6 through 9.3.8

show the system configurations that were used in comparing the annual performance of the

conventional VAV, the IDECOAS and the LD‐IDECOAS, respectively.




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Figure 9.3.6: Schematics of a conventional VAV system used in system comparison (Kim et al, 2014)

Figure 9.3.7: Schematics of IDECOAS used in system comparison (Kim et al, 2014)

Figure 9.3.8: Schematics of LD‐IDECOAS used in system comparison (Kim et al, 2014); colored system components refer to four regions in psychrometric chart shown in Figure 9.3.10




July 9, 2016 Page 156 of 171

Kim et al (2014) stated that both solid and liquid desiccant systems could be used to dehumidify the

outdoor air flow. The authors, however, suggested that liquid desiccant system can provide deeper

dehumidification with less regeneration energy consumption compared to solid desiccants. Additional

benefits of the liquid desiccant included less fan energy due to lower pressure drops in the airside of the

system, and the ability to use solar derived regeneration heat for the desiccant.

Figure 9.3.9: Liquid desiccant sub‐system of the LD‐IDECOAS (Kim et al, 2014)

The LD‐IDECOAS shown in Figure 9.3.8 consists of an indirect (IEC) and direct (DEC) evaporative cooler

and a liquid desiccant system for dehumidification of the process air. The liquid desiccant loop is

depicted in Figure 9.3.9. The heat of regeneration of the liquid desiccant units can be derived from a

range of heat sources, including solar heat or low grade waste heat. In order to achieve an enhanced

cooling effect in the IEC, the exhaust air can be supplied into the secondary side of the IEC when the

exhaust air has a lower wet bulb temperature than the outdoor air. The heating coil (HC) and sensible

heat exchanger (SHE) are located at the exhaust air side to maintain adequate temperature set points

during periods when no cooling was needed. The supply air flow rate was modulated on the basis of the

air‐conditioning load of the conditioned zones, in the same way as in a conventional VAV systems.

Figure 9.3.10 shows a psychrometric chart which the authors divided into four regions. These four

regions represented either different climate zones or different phases during the annual climate cycle at

a particular location. Within the four regions the components of the system would operate at different

levels or would be turned off or on or they would be bypassed. Table 9.3.1 summarizes how the main

system components of direct and indirect evaporative coolers and the liquid desiccant unit would

operate. The authors concluded from their system analysis that all different system operating settings

(e.g. Regions A through D) performed well with different components on or off and that the required

supply air temperature and humidity could be continuously maintained. The results indicated that on an

annual level the LD‐IDECOAS resulted in a 64% energy savings relative to the conventional VAV system.

By using solar heat for the desiccant regeneration would enhance the energy savings further. The




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authors point out that the use of the liquid desiccant systems resulted in a 23% annual energy reduction

over the IDECOAS system (e.g. without dehumidification) due to the increased effectiveness of the

evaporative cooling process. While the energy savings are higher in the LD‐ IDECOAS than in the

IDECOAS, in very humid conditions the LD‐ IDECOAS would be required to provide suitable indoor

humidity levels, since the IDECOAS would not be able to provide such humidity levels. Figure 9.3.11

suggest additional system improvements of the LD‐IDECOAS could include the use of solar heat and

cooling tower for the liquid desiccant system.

Figure 9.3.10: Four regions of performance of the LD‐IDECOAS (Kim et al, 2014)

Table 9.3.1: Modes of operation of LD, DEC and IEC in four regions of psychrometric chart, (Kim et al, 2014)




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Figure 9.3.11: LD‐IDECOAS with thermal components added (Kim et al, 2014)

9.4 Desiccant Enhanced Evaporative Air‐Conditioning

This section is dedicated to the discussion of a recent desiccant cooling systems development by

National Renewable Energy Laboratory (NREL) which AC‐system developers called nothing short of

“revolutionary AC‐technology” (MIT Technology Review, 2016).

The system is referred to as DEVAP, from Desiccant Enhanced Evaporative Air Conditioning. As discussed

in Section 9.3 the thermodynamic performance of evaporative cooling is impeded in humid weather and

adding desiccants to dehumidify the supply air to the evaporative coolers can rectify this shortcoming.

The DEVAP technology solves the combined desiccant dehumidification and evaporative cooling by using

membrane technology that separates desiccant from air travelling through the process vessel. The

polymer membrane used in the process has pores about 1 micrometer to 3 micrometers in diameter, big

enough that water vapor passes through easily while the salty liquid cannot pass through it. The

membrane material is also coated with a Teflon‐like substance to repel liquid water. The desiccant pulls

moisture from the airstream, leaving dry, warm air. This dry and warm air stream enters a second

channel, where water evaporates indirectly to cool a secondary airstream, which in turn cools the first

airstream. The result is cool and dry air.




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Kozubal et al (2011), who are the inventor of the DEVAP process, discussed that simply combining

desiccant‐based dehumidification and indirect evaporative cooling technologies is feasible, but has not

shown promise because the equipment is too large and complex. The DEVAP technology provides a

single cooling core, evaporative and desiccant cooling, which makes the system easier to install and

operate and renders dehumidification many times more efficient. The authors claimed that due to the

DEVAP technology of using membranes, the system solves the important issues of eliminating desiccant

entrainment into the airstream. When an additional water containment is used wet surfaces are

eliminated, which prevents bacterial growth and mineral buildup, therefore avoiding cooling core

degradation.

DEVAP decouples cooling and dehumidification performance, which results in independent temperature

and humidity control. The energy input is largely switched away from electricity to low‐grade thermal

energy including waste heat, solar, or biofuels. Thermal energy consumption correlates directly to the

humidity level in the operating environment. The authors suggested that yearly combined source energy

savings can be as high as 90% for dry climate and about 30% for very humid climate. The authors further

pointed out that liquid desiccant technology is a new science with unpracticed technology

improvements that can reduce energy consumption an additional 50%.

The DEVAP technology builds on the previously developed liquid‐desiccant‐based A/C (LDAC) technology

by an industry partner of NREL. In this particular LDAC process liquid desiccant is absorbed into the

conditioner (absorber) where the inlet ambient air is dehumidified. The liquid desiccant is regenerated

in the regenerator (desorber) where the water vapor desorbs into the exhaust air stream. This process is

called low flow liquid desiccant A/C, because the desiccant flow is minimized in both heat and mass

exchangers to the flow rate needed to absorb the necessary moisture from the air stream. The heat and

mass exchangers must therefore have integral heating and cooling sources (55 F –85 F cooling tower

water is supplied to the conditioner). The regenerator uses hot water or hot steam at 160°–212 F.

The DEVAP enhancements to the LDAC system is illustrated in Figure 9.4.1. The figure depicts how the

DEVAP cooling core enhances the already developed LDAC technology and converts it from a dedicated

outdoor air system to an air conditioner that performs space temperature and humidity control and

provides all the necessary ventilation air. DEVAP can be configured to provide 30% ‐ 100% outdoor air.

As a further significant advantage the DEVAP system does not require a cooling tower, which reduces

maintenance requirements. Figure 9.4.1 illustrates the three basic ways to regenerate the desiccant

system with a thermal source: solar, water heater, and a double effect. The water heater or boiler can

be fueled by many sources, including natural gas, combined heat and power (CHP), or even biofuels.




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Figure 9.4.1: DEVAP enhancement for LDAC, (Kozubal et al, 2011)




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Figure 9.4.2 illustrates, and the following steps describe the physical DEVAP process:

1. Ventilation air [1] and warm indoor air [2] are mixed into a single air stream.

2. This mixed air stream (now the product air) is drawn through the top channel in the heat exchange pair.

3. The product air stream is brought into intimate contact with the drying potential of the liquid desiccant [d] through a vapor‐permeable membrane [e].

4. Dehumidification [ii] occurs as the desiccant absorbs water vapor from the product air.

5. The product air stream is cooled and dehumidified, then supplied to the building space [3].

6. A portion of the product air, which has had its dew point reduced (dehumidified), is drawn through the bottom channel of the heat exchange pair and acts as the secondary air stream.

7. The secondary air stream is brought into intimate contact with the water layer [c] through a vapor‐permeable membrane [b].

8. The two air streams are structurally separated by thin plastic sheets [a] through which thermal energy flows, including the heat of absorption [i].

9. Water evaporates through the membranes and is transferred to the air stream [iii].

10. The secondary air stream is exhausted [4] to the outside as hot humid air.

Figure 9.4.2: Physical DEVAP concept description, (Kozubal et al, 2011)

Figure 9.4.2 indicates the use of two types of membranes, the water‐side and the desiccant‐side

membranes. The desiccant‐side membrane is necessary to guarantee complete containment of the

desiccant droplets, which thus prevents desiccant leaks and entrainment into the air stream. The water‐

side membrane is part of the original DEVAP concept, but is not a necessary component. Its advantages

are complete water containment and dry surfaces, which are made completely of plastic and resist

biological growth.




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Kozubal (2013) provided an illustrated DEVAP heat and mass flow process which is depicted in Figure

9.4.3. From the psychrometric performance the process has the significant advantage that the

achievable dry bub temperature is not limited by the wet bulb temperature at saturation but rather by

the dry bulb temperature of the air. This characteristic is comparable with the M‐cycle evaporative

cooling which was discussed in Section 7.

(a) Physical heat and mass flow diagram for DEVAP process The locations in the schematics correspond to the state point in the psychrometric chart

(b) Psychrometric chart with state point for DEVAP process

Figure 9.4.3: Thermal performance of the DEVAP process, (Kozubal, 2013)




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Kozubal et al (2011) presented possible DEVAP installations in residential and commercial buildings.

Figure 9.4.4 illustrates proposed generic DEVAP implementations in residential and commercial building

implementations. In both cases solar heat is used for all or a major part of the desiccant regeneration.

In a residential applications DEVAP performs the following functions:

Air conditioner with independent temperature and

humidity control

Dedicated dehumidifier

Mechanical ventilator

In commercial applications DEVAP units would be installed as roof top units (RTU). There are several options desiccant regeneration heat sources, such as:

Natural gas only

CHP with or without natural gas backup

Solar heat with or without natural gas backup.

The figure above shows DEVAP system with a solar and CHP options.

Figure 9.4.4: Example of commercial and residential installation of DEVAP A/C, (Kozubal et al, 2011)

Kozubal et al (2011) performed a simplified cost analysis for a typical commercial DEVAP installation,

which assumed monthly electricity and natural gas costs, but did not include time‐of use electricity

costs, such peak power. Because electric power draw is a main concern of commercial peak

consumption, inclusion of time of use costing would increase electricity costs and therefore savings,

thus improve the economics of DEVAP applications. A 15‐year present value analysis suggested a 28%

and 39% internal rate of return (IRR), for the humid Houston and dry Phoenix climate, respectively.


SECTION 10 ‐ GENERAL FINDINGS OF THE LITERATURE REVIEW


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The main findings of this literature and technology review are as follows:

The basis for dehumidification is described by means of psychrometric properties which can be

readily used to describe the heat and mass transfer of dehumidification of air.

There are two major forms of dehumidification, cooling based and desiccant humidification,

which are used for most air drying applications. There are some specialized dehumidification

using pressure and electricity as the driving forces to remove water, but these applications are

not used on a greater scale and are not discussed in this literature and technology review.

Cooling‐based dehumidification cools the moist air to below dew point and induces

condensation of the water vapor phase, thereby removing the moisture from the air. Cooling‐

based dehumidification can only dry air to a dew point that corresponds to the freezing

temperature of water, since at temperatures below freezing, the cooling coils generate ice and

the process is no longer effective.

Desiccant dehumidification removes water vapor through absorption into liquid desiccants or

adsorption to a solid desiccant material. The desiccant material takes up water vapor due to

differences in vapor pressure of the desiccants and the air to be dehumidified. Desiccants can

reach lower dew point temperature than cooling‐based dehumidification.

Cooling‐based dehumidification in building conditioning has lower first costs than desiccant

dehumidification, and therefore cooling‐based dehumidification systems are mostly used. The

potential advantages of lower operating cost through desiccants have not been a strong enough

incentive to increase the market share of desiccants in building dehumidification.

Desiccant dehumidification has its roots in specialized industrial and commercial applications.

Desiccants are used in a wide range of applications to mitigate negative effects of humidity in

products and processes. Typically, in their usual industrial or commercial applications desiccants

are develop for specific applications and are sold for a dedicated market segment.

Desiccant dehumidification has been used in conjunction with building air‐conditioning systems

for several decades. While there were a limited number of liquid desiccant air‐condition

applications, the vast majority of working desiccant systems in buildings are using solid

desiccant material.

Desiccant dehumidification units have been added to conventional vapor compression air

condition systems in efforts to increase the energy efficiency of the overall system and utilize

heat and humidity sinks for energy recovery.




July 9, 2016 Page 165 of 171

Over the past 20 years there has been increasing interest to use desiccant dehumidification as a

way to achieve separation of the three basic air‐conditioning functions, which are sensible

cooling, latent heat removal and space ventilation. The typical cooling‐based air‐conditioning

combines these three functions, however, it is next to impossible to optimize all three in one

system setting, because building and climate conditions may require one function to dictate the

performance envelope of the overall system. For example, in humid climates, the requirement

to remove latent heat often results in overcooling of the spaces since sensible heat removal

would require less cooling capacity than has to be supplied for humidity control. In such cases,

separating the three basic functions of air‐conditioning systems would provide ample

opportunity for effective system optimization.

When desiccants are used for humidity control, the sensible heat removal can be carried out at

above dew point set‐points, which avoids overcooling and results in more energy effective use

of the space and building conditioning system. Providing sensible cooling at a higher

temperature than the below dew point temperature of a conventional vapor compression AC

system increases the coefficient of performance of the vapor compressor.

When desiccants take up moisture, they become saturated as the water vapor pressure in the

desiccant material approaches the water vapor pressure in the air stream. At that point in the

process, the water in the desiccant has to be removed by applying heat, in order to establish a

water vapor pressure differential that drives the desorption process. Depending on the type of

desiccant material there are several different of heat sources that can be used for desiccant

regeneration, such as high grade (steam, natural gas burners) and low grade (solar, waste

process heat) heat sources.

When desiccants are used in system integration with dedicated sensible heat removal, the term

“desiccant cooling” is often used. This term indicates that latent loads in the moist air are

removed by desiccants, while the sensible heat removal can be carried out by a range of cooling

technologies, including conventional and alternative cooling methods. Alternative cooling

methods can include open‐cycle evaporative cooling, closed‐cycle adsorption and absorption

and magnetic refrigeration.

Typically, conventional evaporative cooling works effectively only in low‐humidity climates. In

moist, high humidity climates, the high humidity level impedes the cooling effectiveness of

water vaporization. If desiccants are used upstream of evaporative cooling devices, evaporative

cooling can also be used in moist climates. This opens up a significant market for evaporative

cooling also in the tropics and sub‐tropics, with their hot and moist climates.

Evaporative cooling works either in a direct or indirect mode. In the direct mode water is

sprayed into the air stream adiabatically reducing the dry bulb temperature of the air stream. As

a side effect the humidity increases in the airstream since water vapor is generated through




July 9, 2016 Page 166 of 171

vaporization. In the indirect evaporative cooling mode, the air stream to the conditioned spaces

is cooled through air‐to‐air heat exchange with a secondary air stream which experiences

evaporative cooling. The advantage of indirect evaporative cooling is that no additional humidity

is added to the primary air stream. The disadvantage is additional heat transfer steps, with

unavoidable heat losses.

The literature and technology review reports on a recently developed enhanced indirect

evaporative cooling technology, called the M‐cycle, which is based on a so‐called dew‐point

reduction process. This innovative indirect evaporative cooling technology provides chilled

water at near to dew point rather than wet point temperatures, even when working with high

humidity air. This increased range of application of evaporative cooling effectiveness could open

up new and energy efficient thermal cooling applications.

While, generally, solid desiccants were used for the design of desiccant supported evaporative

cooling, there has been a significant interest in the use of liquid desiccant systems in the past 10

years. Liquid desiccant systems typically require lower regeneration temperatures than solid

desiccants. This makes them more suitable for the use with low grade heat sources, such as

solar and waste heat from processes (e.g. combined heat and power systems). In addition, liquid

desiccant systems can avoid significant energy for fans, and these systems are usually

significantly smaller and allow for a more space efficient installation arrangement of the

equipment. The significant disadvantage of liquid desiccants is that concentrated desiccant

solutions can be very corrosive and need specialized handling. Another operational difficulty is

the possibility of liquid desiccant solution entrainment into the process air. The literature and

technology review identified technology development efforts to overcome these drawbacks of

liquid desiccant systems.

A recently developed integrated liquid desiccant and evaporative cooling process is the DEVAP

technology, which has been called a “breakthrough cooling technology”. The DEVAP system

solves many performance problems of earlier liquid desiccant cooling systems and provides a

compact core cooling device that can be used in residential and commercial applications. The

system uses innovative membrane technology which contains the desiccant and avoids

entrainment into the process air stream. The thermal performance of the system offers a high

coefficient of performance. The system can be used in high humidity climate through a dew‐

point reduction process, similar to the M‐cycle.

The conclusions of this literature and technology review are used to identify candidate desiccant

cooling systems for which the feasibility for Hawaii’s climate is assessed in the second part of

this project.


REVIEWED LITERATURE


July 9, 2016 Page 167 of 171

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