SERIITR-252-2718 

UC Category: 59a 

D E86004448 

SERl Desiccant Cooling Test 

Facility: Status Report 

Preliminary Data on the Performance of a 

Rotary Parallel-Passage Silica-Gel 

Dehumidifier 

Kenneth J. Schultz 

April 1986 

Prepared under Task No. 3023.21 

FTP NO. 01-548-85 

Solar Energy Research Institute 

A Division of Midwest Research Institute 

1617 Cole Boulevard 

Golden, Colorado 80401-3393 

Prepared for the 

U.S. Department of Energy 

Contract No. DE-AC02-83CH 10093 
NOTICE 

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PBEFACE 

In keeping with the national energy policy goal of fostering an adequate 

supply of energy at a reasonable cost, the United States Department of Energy 

(DOE) supports a variety of programs to promote a balanced and mixed energy 

resource system. The mission of the DOE Solar Buildings Research and 

Development Program is to support this goal by providing for the development 

of solar technology alternatives for the buildings sector. It is the goal of 

the program to establish a proven technology base to allow industry to develop 

solar products and designs for buildings that are economically competitive and 

can contribute significantly to the nation's building energy supplies. Toward 

this end, the program sponsors research activities related to increasing the 

efficiency, reducing the cost, and improving the long-term durability of 

passive and active solar systems for building water and space heating, 

cooling, and daylighting applications. These activities are conducted in four 

major areas: Advanced Passive Solar Materials Research, Collector Technology 

Research, Cooling Systems Research, and Systems Analysis and Applications 

Research. 

Advanced Passive Solar Materials Research - This activity area includes work 

on new aperture materials for controlling solar heat gains, and for enhancing 

the use of daylight for building interior lighting purposes. It also 

encompasses work on low-cost thermal storage materials that have high thermal 

storage capacity and can be integrated with conventional building elements, 

and work on materials and methods to transport thermal energy efficiently 

between any building exterior surface and the building interior by 

nonmechanical means. 

Collector Technology Research - This activity area encompasses work on 

advanced low- to medium-temperature (up to 180°F useful operating temperature) 

flat-plate collectors for water and space heating applications, and medium- to 

high-temperature (up to 400°F useful operating temperature) evacuated 

tube/concentrating collectors for space heating and cooling applications. The 

focus is on design innovations using new materials and fabrication techniques. 

Cooling Systems Research - This activity area involves research on highperformance dehumidifiers and chillers that can operate efficiently with the 

variable thermal outputs and delivery temperatures associated with solar 

collectors. It also includes work on advanced passive cooling techniques. 

Systems Analysis and Applications Research - This activity area encompasses 

experimental testing, analysis, and evaluation of solar heating, cooling, and 

daylighting systems for residential and nonresidential buildings. This 

involves system integration studies, the development of design and analysis 

tools, and the establishment of overall cost, performance, and durability 

targets for various technology or system options. 

iii 
This report documents the current status of the SERI Desiccant Cooling Test 

Facility and includes preliminary data on the performance of a spirally wound, 

parallel-passage, rotary dehumidifier using silica gel as the desiccant. 

This work was performed under Task 3023.21, Desiccant Cooling Research, during 

FY 1985. Testing of other rotary dehumidifiers will continue. The author 

thanks Terry Penney, among others, at SERI, and John Mitchell of the 

University of Wisconsin for their help and encouragement during this work. 

Kenneth J. Schultz 

1. u 

Approved for 

SOLAR ENERGY RESEARCH INSTITUTE 

Bdt]~~-- 

David Johnson, nager 

Thermal science$ Research Branch 

Z~uLrw. 

annon, Director 

at Research Division 
Objective 

This report describes the SERI Desiccant Cooling Test Facility as it now 

stands and gives preliminary data on a prototype spirally wound, parallelpassage rotary dehumidifier using silica gel as the desiccant. 

Discussion 

Much analytical and experimental work has been put into rotary desiccant dehumidifiers and desiccant cooling systems. However, there has been little coordination between these two areas. Desiccant dehumidifier models have been 

used extensively in systems analysis with little experimental verification. 

The SERI Desiccant Cooling Test Facility was constructed to remedy this 

situation. Data on various rotary dehumidifiers that have been obtained from 

this facility will be used to verify the accuracy of dehumidifier models, providing a new level of confidence in dehumidifier design and desiccant cooling 

systems analysis. 

Conclusions and Recommendations 

The SERI Desiccant Cooling Test Facility is operational. ,Several minor difficulties, including humidity control and inlet velocities, were discussed and 

improvements were suggested. 

We obtained satisfactory data on a prototype parallel-passage silica-gel 

rotary dehumidifier. Preliminary analysis of the data indicates an effective 

Lewis number near unity. However, nonuniformities in the passage spacings 

have reduced the effective number of transfer units by a factor of 3-4. 

Construction and testing of a dehumidifier with a matrix of uniform passage 

spacings are recommended. 
TABLE OF CONTENTS 

Page 

Nomenclature ............................................................. ix 

1.0 Introduction ........................................................ 1 

1.1 Background ..................................................... 1 

1.2 Purpose of Test Facility ....................................... 1 

2.0 Description of Test Facility ........................................ 5 

2.1 Overview ...................................................... 5 

2.2 General Layout ................................................. 5 

2.2.1 Flow Loop Design .................................. 5 

2.2.2 Test Unit Design ........................................ 5 

3.0 Data Acquisition and Control System ................................. 12 

4.0 Controls ............................................................ 15 

4.1 Flow Rates ..................................................... 15 

4.2 Temperatures ................................................ 17 

4.3 Humidity ........................................................ 17 

4.4 Wheel Speed .................................................... 18 

5.0 Instrumentation ..................................................... 19 

5.1 Temperature .................................................... 19 

5.2 Humidity ....................................................... 19 

5.3 Pressure ....................................................... 23 

5.4 Flow Rate .................................................. 25 

5.5 Wheel Speed .................................................... 25 

6.0 Preliminary Data .................................................... 26 

6.1 Description of Dehumidifier Wheel .............................. 26 

6.2 Seal Leakage Rates ......................................... 26 

6.3 Overall Facility Operation ..................................... 28 

6.4 Initial Tests .................................................. 32 

6.4.1 Low Speed Tests ......................................... 32 

6.4.2 High Speed Tests ....................................... 37 

7.0 Conclusions ......................................................... 41 

7.1 Facility Status and Preliminary Test Results ................... 41 

7.2 Future Work .................................................... 41 

8.0 References .......................................................... 42 

Selected Distribution List ............................................... 45 
LIST OF FIGURES 

Page 

The Ventilation Cycle Desiccant Cooling System ...................... 2 

The Recirculation Cycle Desiccant Cooling System .................... 3 

Test Facility Layout ................................................ 6 

Desiccant Cyclic Test Facility Looking South to North ............... 7 

Desiccant Cyclic Test Facility Looking North to South ............... 8 

System Capabilities ................................................. 9 

Rotary Dehumidifier in Test Facility with One Transition 

Section Removed ..................................................... 11 

Instrumentation and Data Acquisition Center ......................... 13 

Data Acquisition Program Flowchart ................................. 14 

Schematic of Control Functions and Required Ranges .................. 15 

Fan Performance Curve ............................................ 16 

Measurement Locations .......................................... 20 

Thermocouple Design ................................................. 23 

Sampling Points to Obtain Outlet Profile ............................ 24 

Coating of the Tape Film with Silica Gel ............................ 27 

First Test Dehumidifier from Rotary Heat Exchanger. Australia ....... 29 

Seal Designs ..................................................... 30 

Dehumidifier Seal Leakage Paths ..................................... 31 

Low Rotation Speed Outlet States .................................... 34 

F2 Effectiveness versus C ........................................ 36 2.j~ 

Fit to Very Low Speed Data .......................................... 36 

F2 Effectiveness versus C2ja ........................................ 37 

High Speed Test Results ............................................. 40 
LIST OF TABLES 

Page 

Projected Performance of Advanced Dehumidifier Systems at ARI 

Design Conditions ................................................... 3 

Specifications for Flow Loop Air States ............................. 6 

Data Acquisition and Control Unit Channel Allocation ................ 12 

Summary of Required Measurements ................................... 21 

Parameters of the First Test Dehumidifier ........................... 28 

Typical Seal Leakage Rates .......................................... 31 

Silica Gel and Matrix Properties .................................... 34 

Low Rotation Speed Tests ....................................... 35 

High Rotation Speed Tests ........................................... 39 
NOMENCLATURE 

A moist air 

c specific heat of air (k~/kg OC) P 

specific heat of moist air 

Cp~~ specific heat of dry desiccant 

C~~ 

specific heat of tape 

C~w specific heat of water 

C discharge coefficient (dimensionless) 

'i j reciprocal of the dimensionless wave speed for potential i in 

period j 

'max maximum value of fluid capacity rate in matrix (J/ks or kg/s) 

'min minimum value of fluid capacity rate in matrix (J/ks or kg/s) 

r matrix capacity rate (J/ks or kg/s) 

c 9: Cmin/Cmax 

C= Cr/Cmin 

COPth thermal coefficient of performance 

COPe electric coefficient of performance 

d nozzle diameter (m) 

D duct diameter (m) 

D A dry air 

DD dry desiccant 

I? i potential 1 or potential 2 

f gain factor, currently set to 1.0 (dimensionless) 

Gt gain factor, currently set to 1.0 

h specific enthalpy of moist air kg dry air) 

hs heat of adsorption 

hv latent heat of vaporization of water 
NOMENCLATURE (Continued) 

period 1 (process) or period 2 (regeneration) 

flow equation coefficient (kg-K/S-unit) 

temperature equation coefficient (k~/unit) 

wheel speed control equation coefficient (rphlunit ) 

coefficient for the Fincor controller (rpm/mA) 

coefficient for the HP 34978 (mA/uni t ) 

3 flow system constant (m /s/rpm) 

coefficient for the SCR controller (k~/mA) 

effective Lewis number 

mass flow rate (kg/s) 

set point mass flow rate (kg/s) 

flow rate of dry air (kg/s) 

mass flow rate of dry air in period j (kg/s) 

mass of dry gel (kg) 

active value of mDD 

physical value of mDD 

mass of tape 

wheel rotation speed (rph) 

set point wheel speed (rph) 

overall number of transfer units 

NTUo for sensible heat transfer 

NTUo for moisture transfer 

indicates calculation is based on physical measurements 

ambient pressure (Pa) 

saturation vapor pressure (Pa) 
NOMENCLATURE (Continued) 

amb 

t~ 

t~~ 

total absolute pressure (Pa) 

atmospheric pressure (Pa) 

equilibrium vapor pressure (atm) 

energy input rate (kW) 

gas constant for air (289.6 kg K) 

Reynolds number at the nozzle throat 

integer signal from IBM PC to HP 34978 (unit) 

offset signal (unit) 

temperature (OC) 

ambient temperature (OC) 

set point temperature ('12) 

dew point temperature (OC) 

absolute temperature (K) 

tape 

dew point temperature (K) 

millivolts referenced to OOC 

3 volume flow rate (m 1s) 

humidity ratio (kg/kg) 

desiccant moisture content (kg/kg) 

water 

dimensionless length 

Greeks 

B ratio of nozzle diameter to duct diameter (dimensionless) 

Y property ratio analogous to the ratio of heat capacities 
NOMENCLATURE (Concluded) 

mass flow ratio of dry desiccant to air (dimensionless) 

predicted l' j 

differential pressure across nozzle (Pa) 

temperature or moisture effectiveness 

E for heat exchanger 

dimensionless time 

number of mass transfer units for period j 

apparent A j 

predicted A j 
1.0 INTRODUCTION 

1.1 Background 

Solid desiccant cooling systems have 

received considerable interest as a 

mechanically simple, solar-fired 

option to conventional vapor compression air-conditioning systems for 

HVAC applications. These systems 

offer potentially lower cooling costs 

to the consumer. They would provide 

the natural gas utilities with a new 

summer market and may help reduce 

peak summer air-conditioning loads on 

electric utilities. 

Two system configurations (Figures 1-1 and 1-2) using rotary dehumidifiers are being considered for 

residential applications. Several 

studies have indicated that these 

systems can be competitive with vapor 

compression air-conditioning if thermal coefficients of performance (COP) 

greater than 1.2 can be obtained 

(Booz-Allen and Hamilton 1981; 

Jurinak 1982; Scholten and Morehouse 

1983 1. Commercial and industrial 

applications of solid desiccant dehumidification combined with vapor compression sensible cooling are also 

being considered (Cohen, Levine, and 

Arora 1983; Howe, Beckman, and 

Mitchell 1983). 

To make these systems economically 

viable researchers are trying to 

develop high-performance, solid desiccant dehumidifiers with low pressure drops for air-conditioning 

applications. The Institute of Gas 

Technology (IGT) investigated a corrugated passage, laminar flow matrix 

(Macriss and Zawacki 1982). SERI 

Table 1-1 shows the ~rojected performance of these systems. Further information is available in Schlepp and 

Schultz (1984). In addition, Kang 

(1985) has proposed new desiccant 

cycle arrangements with COPS exceeding 2.5. 

Commercialization of solid desiccant 

technologies is just beginning. 

American Solar King (1984) and Sharp 

(1982) are readying ventilation cycle 

systems for the residential market. 

~hermo~lectron/~~~ and Cargocaire are 

actively testing in the field systems 

that combine desiccant dehumidification with vapor compression sensible 

cooling (Cohen, Levine, and Arora 

1983). 

Developing solid desiccant dehumidifiers includes much analytical work 

in systems analysis and component 

modeling. This work has simulated 

seasonal system performance, estimated energy and cost savings, investigated the effects of various 

desiccant materials, and provided 

information on how to optimize performance through dehumidifier design 

and system operation. However, many 

of the models used, especially those 

available to the academic community 

and those not involved with hardware 

development, were not adequately 

verified against experimental data. 

The acceptance of solid desiccant 

cooling by the research, HVAC, and 

consumer communities depends on 

having adequate design tools for this 

fundamental heat and mass exchange 

process. 

further developed a parallel-passage, 1.2 Purpose of Test Facility 

laminar flow matrix (Schlepp and 

Barlow 1984). Exxon, with -funding The Solar Energy Research Institute's 

from the Gas Research Institute (SERI) Desiccant Cooling Cyclic Test 

(GRI), tested a similar design that Facility obtains test data on rotary 

contained additional heat capacity solid desiccant dehumidifiers opermaterial (Huskey et al. 1982). ated under conditions typical of 
I I I I I I I I I Lo.00 

10 40 70 100 

Dry bulb temperature (O C) 

1 4 

Figure 1-1. The Ventilation Cycle Desiccant Cooling System 

Outside -+ 

air 

Exhaust - 

air 

those encountered in solar desiccant 

cooling applications. Several dehumidifier designs will be tested. 

Dehumidifier models will be checked 

against this data, and modifications 

will be made, if necessary, to ensure 

the models match the data. These 

models will then allow us to accurately simulate dehumidifier performance, optimize system operation, 

accurately estimate seasonal energy 

and cost savings, and provide useful 

feedback for materials and hardware 

development. 

residence 

Two dehumidifier computer models will 

be compared with the data obtained: 

a finite difference program (MOSHMX) 

developed by Maclaine-cross (1974) 

and a simplified model (DESSIM) developed by Barlow (1982) at SERI. 

MOSHMX was compared with data from a 

rotary, packed-bed dehumidifier but 

with unsatisfactory results because 

Rotary 

heat 

exchanger 

3 ' 4 In (0

Rotary 

desiccant 

bed 

+ 

2 

N Evap. 

cooler 

t 

9 

-+Supply X 

air to 

residence 

8 7 

Solar 

heater f- - 

6 air from 

Evap. 

cooler Return 
Dry bulb temperature ("C) 

Figure 1-2. The Recirculation Cycle Desiccant Cooling System 

1 2 

Table 1-1. Projected Performance of Advanced Dehumidifier Systems 

at AEI Design Conditions 

,- 

- - 

4 Supply E 

air to 

3 4 - 

-- 

Treg Capacity Type Dehumidifier Cycle eHX COPth COP, o f 

(OC) kW k.J/kg Evaluation 

Evap. 

cooler 

IGT HCOP Vent 0.95 - - 8.4 18.7 0.95 6.2 Comput er 

SERI Vent 0.95 8 0 - - 16.7 1.07 7.2 Computer 

EXXON/GRI Vent 0.93 77 2.3 12.9 1.05 5.4 Laboratory 

Vent 0.93 7 7 - - 15.1 1.3 8.7 Computer 

+ 

Rotary 

Return + 

air from 

heat 

exchanger 

residence 

Exhaust tair 

residence 

Rotary 

* 

desiccant 

bed 

C- t-- Outside 

6 5 air 

Evap. - - cooler 

8 7 

Solar 

heater 
of difficulties in adequately modeling the solid-side moisture transfer 

resistance present in the wheel (van 

Leer sum and Close 1982) ; however, we 

expect it to work much better for 

high-performance designs. MOSHMX was 

used extensively for modeling dehumidifiers (Jurinak 1982; Brandemuehl 

1982; van den Bulck 1583). DESSIM 

was compared with data from a singleblow experiment for both a packed bed 

(Kutscher and Barlow 1982) and a 

parallel-passage dehumidifier (Barlow 

1982). It was not compared with 

rotary dehumidifier data, however. 

The data obtained from the test facility will be made available through 

later reports so other models can be 

checked. 

With minor modifications the test 

facility can also be used for other 

projects, such as testing heat exchangers, evaporative coolers, and 

liquid desiccant system components. 
2.0 DESCRIPTION OF TEST FACILITY 

2.1 Overview 

The test facility was designed to be 

flexible in characterizing the performance of dehumidifier components. 

The design was guided by past experience at SERI and elsewhere in building and operating similar facilities. 

Sources of this information outside 

SERI include the University of 

California at Los Angeles lark 

1979), IGT (Wurm et al. 19791, 

Garrett AiResearch (~ousseau 19821, 

Illinois Institute of Technology 

(~onier, Worek, and Lavan 19821, 

Monash ~niversit~/~niversity of New 

South Wales (Australia) (Ambrose, 

Maclaine-cross, Robson 19791, Commonwealth Scientific and Industrial 

Research Organization (van Leersum 

and Close 1982; Strahm and Wilson 

1982), and Exxon Corporation Energy 

Venture Development Group (Huskey et 

al. 1982). Past SERI work is documented in Kutscher and Barlow (1982) 

and more recently by Penney and 

Maclaine-cross (1985). 

The initial design of the test facility is reported in Schlepp, Schultz, 

and Zangrando (1984). Justification 

for the layout, control equipment, 

and instrumentation chosen is given 

there. Several minor changes were 

made as the loop was put together. 

This report presents the current capabilities and specifications of the 

test loop. Use of equipment, controls, or instruments by a particular 

manufacturer does not constitute an 

endorsement of that product ; any 

equivalent product may be used in its 

place. 

General Layout 

consists of two sections: the flow 

loop and the test unit. The flow 

loop contains all equipment and controls necessary to supply two airstreams of a given flow rate, temperature, and humidity to the test unit. 

The test unit consists of the test 

article and equipment and instrumentation to allow full testing of component performance. Figures 2-2 and 

2-3 show the test facility. 

2.2.1 Flow Loop Design 

The flow loop has two independent 

streams for supplying air to the test 

unit at a given flow rate, temperature, and humidity. Specifications 

for the flow loop air states are 

given in Table 2-1. Figure 2-4 shows 

the capabilities of the test facility 

in relation to conditions under which 

a typical desiccant cooling system 

would operate. 

At maximum flow rate a dehumidifier 

for an approximately 7-kW (2-ton) 

cooling system could be tested, 

assuming a cooling capacity of 

15.5 k~/kg (400 scfm/ton) could be 

generated (typical design values). 

Spiral duct of 0.3-m (12-in.) diameter and 22-gauge (0.75-mm) sheet 

metal was used for the flow loop. 

The airflow is induced by blowers 

powered by variable-speed motors, and 

electrical resistance heaters warm 

the air. Humidity is produced by 

injecting steam generated by an electric boiler. Subsequent sections of 

this report give details of these 

subsystems and their associated 

controls. 

2.2.2 Test Unit Design 

Figure 2-1 is a schematic of the test The test unit forms the framework for 

facility as installed in the west supporting the components to be 

high bay of SERI'S Field Test Labora- tested, holds the instrumentation to 

tory Building (FTLB). The facility measure the performance, and manages 
-- I I Future I I Experi- , 

u 

D.A.S. 

Fan I 

I I 

I HUM HTR Boiler HTR HUM 

I 

I 

I 

Figure 2-1. Test Facility Layout 

Table 2-1. Specifications for Flow Loop Air States 

Parameters Adsorption Regeneration 

Flow rate (kg/s) 0.04-0.43 0.04-0.43 

Temperature (OC) 22-40 60-90 

Humidity (kg/kg) 0.008-0.020 0.008-0.020 

the airflow to the components. For 

the current tests transition sections 

were fabricated to go from the round 

spiral duct to an approximately 0.8-m 

(2.5-ft) diameter half circle that 

matches the first test article. 

These sections, shown in Figure 2-5, 

are approximately 0.9 m (3 ft) long 

and were designed to expand the 

airflow with as little disruption as 

possible. However, the flow nozzle 

just upstream creates a rather 

turbulent jet of air that does not 

expand smoothly in the transition 

sect ion. The use of screens or 

longer D-sections may be needed to 

produce uniform flow at the test 

article. A straight 0.3-m (1-ft), 

D-shaped section connects the transition section with the test article 

and contains the thermocouple and 

humidity sampling arrays for measuring the dehumidifier performance. 

The sections of the duct between the 

inlet and outlet thermocouple locations were insulated to prevent heat 

losses. The regeneration stream 

ducting has been covered with two 

layers of 9-cm (3.5-in.) fiberglass 

bats backed with kraft paper. The 

process stream, because of its lower 

operating temperatures, was covered 

with only one layer. 





3.0 DATA ACQUISITION AND CONTROL SYSTEM 

A photograph of the data aquisition 

and control system is shown in 

Figure 3-1. The center of the data 

aquisition and control system is a 

Hewlett Packard HP 34978 Data 

~c~uisition/~ontrol Unit. This unit 

contains a digital voltmeter with a 

5-112 digit resolution and five slots 

for data acquisition and control 

cards. An extender unit (HP 34988) 

provides slots for ten additional 

cards. Table 3-1 summarizes the 

cards used and the channel allocations. Voltage measurement rates are 

25 readings per second with 5-112 

digit DVM resolution and auto-zero 

on. This can be increased to 300 

readings per second without auto-zero 

and use of 3-112 digit DVM resolution 

when acceptable. 

An IBM personal computer (PC) with 

512 kilobytes of memory and two, 

360-kilobyte disk drives converts 

data and controls storage and the 

system. The IBM PC communicates with 

the HP 3497A over an RS-232 (serial) 

interface at a rate of 4800 baud. An 

enquire/acknowledge (ENQIACK) so£ tware handshake is used to coordinate 

the communications. Data are stored 

directly on diskettes and sent to a 

line printer. The programming language is BASIC, either interpreted or 

compiled. A flowchart of the program 

structure is shown in Figure 3-2. 

The transient data and control loop 

of the program takes approximately 

nine seconds when running interpreted 

BASIC and five seconds when running 

the compiled version. 

Table 3-1. Data Acquisition and Control Unit channel Allocation 

Card Type Slots Channel s 

(signal ) Available Available Uses 

Inputs (instrumentation) 

Voltage (with OOC ref) 3 6 0 Type T thermocouples 

Voltage 2 4 0 Humidity (5 channels) 

Pressure (7) 

Wheel speed (1) 

Pulse counter 2 2 Wheel speed 

outputs (controls 

Current 3 6 Fans (variable speed motors) 

Heaters (silicon-controlled 

rectifier controllers) 

Humidifiers 

(electro-pneumatic valves) 

Voltage 1 2 Wheel speed 

(variable-speed motor) 


4.0 CONTROLS 

This section describes the equipment 60-Hz input. The power input is conand equations used to control the verted to a DC voltage proportional 

operation of the test facility. The to the input signal from the HP 3497A 

layout of the controls is shown sche- (4-20 mA, set by the computer) and 

matically in Figure 4-1. then inverted into an AC voltage 

(3-step approximation of a sine wave) 

at a frequency proportional to the 

4.1 Flow Rates 

Flow for each stream is produced by a 

blower sized to provide 0.43 kg/s at 

1000 Pa, 25O~, and 0.8 atm (standard 

pressure at SERI's elevation, approximately 6000 ft). A performance 

curve is shown in Figure 4-2. The 

blowers are powered by 2-hp, 230-V, 

3-phase motors that receive a variable frequency signal from a Fincor 

5100 variable frequency speed controller QINCOM International). The 

controller requires a 230-V, 1-phase, 

input reference signal. The output 

voltage and frequency are regulated 

to maintain a constant voltage-tofrequency ratio so the motor can 

operate at rated torque over the 

speed range. 

For a given flow system (i.e., pressure drop versus flow rate), the volume flow rate of air produced by a 

fan is proportional to the fan speed. 

The Fincor controller produces a fan 

speed proportional to the input signal. The HP 3497A linearly converts 

DAS/control 

ACT ACT 

Process stream (1 ) Regeneration stream (2) 

Additional 

Function Code Range Input (DAS) requirements 

Air flow CF-1,2 0-0.43 kg/s 4-20 mA 230 V, 14 

Air temp. CT-1 22-40" C 0-12 mA 

CT-2 60-90" C 0-12 mA 

Humidity CH-1,2 0.008-0.02 kg/kg 3-15 mA 3 atm 

compressed air 

Rotation CR-1 1-20 rph 0-10 V 0-5 V, TTL readout 

Figure 4-1. Schematic of Control Functions and Required Ranges 
0 0.2 0.4 0.6 0.8 1 .O 1.2 1.4 

I I 1 2.0 

Type: Compact GI 

Size: 124 

Conditions: 

1525 m (5000 ft) 

- 1.5 

elevation 

21 " C (70" F) 

2984 rpm - 1.0 

- 0.5 

I I 

0 5 10 15 20 25 30 

lo2 ft2/min 

Source: New York Blower Co 

Figure 4-2. Fan Performance Curve 

an integer input from the IBM PC to a So = offset signal (unit) 

current signal (0-10,000 from the PC Kf = flow equation coefficient 

results in a signal of 0-20 mA). (kg-K/s-unit) 

Assuming we can neglect ambient pres- t = temperature (OC). 

sure variations, the following equation can be written for mass flow Using differential proportional conrate: trol, the control equation for the 

P P mass flow rate becomes 

rh = - V = - - So) RT RT 

273 + t 

Snew = Sold + Gf Kf (hs- 15) , (4-1 

where 

where 

rh = mass flow rate (kg/s) 

P = ambient pressure (Pa) 

R = gas constant for air 

(289.6 .J/kg-K) 

T = absolute temperatur (K) 3 V = volume flow rate (m Is) 3 KS = flow system constant (m /s rpm) 

KF = coefficient for the Fincor controller (rpm/mA) 

KHP = coefficient for the HP 3497A 

(mA/uni t ) 

S = integer signal from IBM PC to 

HP 3497A (unit) 

lil = set point mass flow rate (kg/s) 

G; = gain factor, currently set to 

1.0 (dimensionless). 

The constant Kf is determined experimentally and must be redetermined 

each time the pressure drop characteristics of the system are changed 

(such as changing flow nozzles). 

Using the above equations, flow rates 

can be brought up to the set points 

in approximately 20-30 s (for a 10-s 
sample time) and typically can be 

held to within less than +0.5%. Measuring and calculating flow rates are 

discussed in Section 5.4. 

4.2 Temperatures 

Heat is added to the airstreams 

through a pair of 480-V, 3-phase 

electric resistance duct heaters. 

The process stream heater has a capacity of 6 kW, and the regeneration 

stream heater has a capacity of 

35 kW. Energy input to each heater 

is regulated by a silicon-controlled 

rectifier (scR) controller (Halmar 

Electronics, series LZF~) that receives a 0- to 5-mA signal from the 

HP 3497A. 

The temperature of the flow stream 

leaving the heater is given by 

where 

= ambient temperature (OC) 

= energy input rate (k~) 

c = specific heat of air 

(kJ/kg OC) 

KSCR = coefficient for the SCR controller (kW/mA) 

Kt = temperature equation coefficient (k~/unit) 

K~~ = coefficient for the HP 3497A 

(rn~/unit). 

Again, using differential proportional control, the control equation 

for temperature becomes 

where 

Gt = gain factor, currently set to 

1.0 

ts = set point temperature (OC). 

The constant Kt can be determined 

from the characteristics of the 

HP 34976 output card and the SCR controllers. Power input to the heaters 

is monitored by an Ohio Semitronics 

AC watt transducer (PC 5 series). 

Once we approach set point conditions, we can hold the temperature to 

within 20.2~~. Temperatures can 

overshoot at the beginning of experiments, especially for set points much 

higher than ambient, because of the 

slow response time of the heater element and lack of insulation on the 

ducting between the heaters and the 

test unit. This problem can be reduced by manually controlling the 

signal from the IBM PC. Large step 

changes in temperature can still take 

5-10 min. 

4.3 Humidity 

We maintained the humidity conditions 

by injecting steam into the airstreams. The steam is generated by 

an Electro-Steam, 50-kW electric 

boiler with on/off control through a 

Honeywell pressure regulator with a 

1-psi (7-kPa) differential. The 

steam is injected into the airstream 

through a Walton ST-100 steam humidifier. The steam flow to each stream 

is regulated by a Fisher Controls 

electro-pneumatic valve. 

Controlling the steam flow proved to 

be a difficult task. Maintaining a 

constant inlet humidity ratio is complicated by the steam pressure fluctuation in the boiler and fluctuations in the lab air humidity ratio 

caused by the cycling of the building 

air conditioning system. The control 

valves exhibit much hysteresis when 
changes in direction are called for; 

this has made computer control not 

feasible. 

Currently, a relative humidity sensor 

(General Eastern, model 450) placed a 

short distance downstream of the humidifier provides a voltage signal to 

a diff erentialfintegral controller 

(Leeds and Northrup, Electromax 111). 

The controller outputs a current signal to the control valve proportional 

to the difference between the sensor 

input voltage and the controller set 

point. With this setup the process 

stream humidity ratio can be maintained to within 20.3 gfkg. However, 

the signal from the relative humidity 

sensor is not very sensitive to 

changes in humidity ratio at the low 

relative humidities (<5%) of the regeneration stream. Therefore, the 

regeneration control valve position 

is set manually to provide the required steam flow. Variations in regeneration inlet humidity ratio of 

t0.7 g/kg are typical. This variation is caused by the boiler pressure 

fluctuations. Using a pressure regulator with a narrower band would help 

this problem. If the room humidity 

ratio varies significantly, the operator needs to constantly monitor the 

regeneration valve position. 

4.4 Wheel Speed 

The dehumidifier wheel is circumferentially friction-driven by a DC 

servomotor (Electro-craft Corp.) 

turning a rubber-rimmed drive wheel 

through a reduction gear box (Leedy 

Manufacturing Co., 220:l ratio). The 

DC servomotor is powered by a linear 

amplifier (also Electro-Craft) that 

outputs a voltage proportional to a 

210-V input signal from the HP 3497A. 

Tachometer feedback from the motor to 

the amplifier ensures a constant dehumidifier rotational speed to within 

20.03 rph. The ratio of the diameter 

of the first test article to that of 

the drive is 4.06:l. The maximum 

wheel speed for this test article is 

135 rph with the above setup. 

Wheel speed is highly repeatable 

(20.02 rph) for a given signal from 

the HP 3497A; however, wheel speed 

varies somewhat nonlinearly (amplif ier) with input signal and we could 

not get a reasonable regression fit. 

Therefore, we developed a proportional control equation to obtain accurate set point wheel speeds: 

where 

N = set point wheel speed (rph) 

8 = wheel rotation speed (rph) 

Kw = wheel speed control equation 

coefficient (rphfunit). 

The constant Kw contains the coef f icient that describes the HP 34978 

signal card; the amplifier and motor 

combination; and the speed ratios of 

the motor, gearbox and drive wheel, 

and dehumidifier wheel, and is determined experimentally by timing wheel 

revolutions or using a strobe light. 
5.0 INSTRUMENTATION 

Figure 5-1 shows schematically the 

measurement locations in the test 

loop. Table 5-1 summarizes these 

measurements and the instruments 

used. Each type of instrumentation 

is described further in the following 

sections. 

5.1 Temperature 

We measured temperatures using copper 

and constantan (type T) thermocouples. For bulk inlet and outlet 

temperatures we placed an array of 

four thermocouples connected in parallel in the duct to obtain an area 

weighted average. The junction design 

shown in Figure 5-2 provides sufficient bare wire to reduce conduction 

2nd radiation errors. The conduction 

error was calculated to be less than 

0.01~~; the radiation error less than 

0.05~~. 

The profile of temperatures versus 

rotation angle at each dehumidifier 

outlet face is measured by an array 

of thermocouples arranged as shown in 

Figure5-3. The grid is closely 

spaced in the region near where the 

dehumidifier rotates into the airstream to resolve the sharp gradient 

present. The more widely spaced grid 

over the rest of the face is sufficient for the more gentle gradients 

there. 

To obtain the temperatures we measured the thermocouple electromotive 

force (EMF) with the digital voltmeter in the HP 3497A through a voltage card that provides an electronic 

O'C (k0.2Oc) reference temperature. 

The thermocouple wire was calibrated 

in SERI'S ~nstrumentation/~etrolo~~ 

Recharge Center. The following 

third-order regression fit results in 

residuals of less than +0.04'~ over 

the range of 0~-100~~: 

where 

t = temperature (OC) 

v = millivolts referenced to O'C. 

Assuming the above noted errors combine randomly, the uncertainty in 

temperature measurements is less than 

k0.5'~. 

5.2 Humidity 

Humidity measurements are made using 

dew point hygrometers (General 

Eastern, model 1100DP/llll~). These 

instruments have optical sensors and 

use chilled mirrors to control condensation. The mirror temperature is 

monitored by a platinum resistance 

thermometer and puts out a voltage 

signal linearly proportional to the 

dew point. These sensors have been 

calibrated by General Eastern using 

instruments, equipment, and standards 

directly traceable to the National 

Bureau of Standards. The stated uncertainty of the sensors is i0.2'~. 

We connected the four instruments in 

series to simultaneously monitor the 

same air sample to confirm this. 

We wrapped electrical heat tape 

around the dew point hygrometers to 

keep the temperature above the dew 

point, which prevents condensation in 

the tubing and sensor cavity. 

Equation 5-2 converts dew point temperature to humidity ratio: 

where 

w = humidity ratio (kg/kg) 

'sat = saturation pressure (Pa) 



I Con 

Figure 5-2. Thermocouple Design 

tDP = dew point temperature (OC) 

Pt = total absolute pressure (Pa). 

Maclaine-cross (1974) developed the 

following correlation for saturation 

pressure: 

Psat = exp (23.28199 - 

3780.82 

TDP 

where 

TDP = dew point temperature (K). 

The uncertainty of this correlation 

is less than 20.1% (Maclaine-cross 

1974). The uncertainty in humidity 

ratio, calculated from a Taylor series expansion and an assumption of 

independent errors, is less than 

?2.5%. 

We obtained the bulk average inlet 

and outlet humidity ratios by mixing 

air sampled across the duct cross 

section. The outlet humidity profiles are obtained from a grid of air 

sample points, shown in Figure 5-3. 

An electrically driven switching 

valve (Valco Instruments, 16 port 

SC type) directs each sample, in 

turn, to a dew point hygrometer. 

During steady-state operation, the 

outlet profile remains constant, so 

one instrument is sufficient for monitoring each face. This arrangement 

cannot adequately monitor the profiles under transient conditions. 

5.3 Pressure 

Pressure measurements are required 

for (1) pressure drop characteristics 

of the wheel, (2) pressure differences across the dehumidifier seals 

that cause leakage, (3) pressure differences across flow nozzles to calculate flow rates, and (4) ambient 

absolute pressure for air density and 

humidity ratio calculations. 

Capacitance pressure sensors are used 

(MKS Instruments, Baratron type 221A) 

for differential pressure measurements and have a range of 0-10 in. of 

water (0-250 Pa). These sensors put 

out a voltage signal linearly proportional to a pressure difference. The 

test loop uses six sensors: two to 

monitor the inlet flow rates and four 

to monitor the pressure difference 

between each face of the dehumidifier 

wheel and the ambient air. The pressure drop through the wheel and 

across the radial seals separating 

the two flow streams are then easily 

calculated. 

We checked the calibration of these 

instruments by comparing them with a 

Hooke gauge manometer (Dwyer Instruments, Microtector) with a resolution 

of +0.0002 in. of water. The results 

showed a typical differential pressure error of less than +0.5%. 

Again, we use a capacitance sensor to 

monitor the ambient absolute pressure 

(MKS Instruments, Barat ron type 

220A). This sensor was checked 

against a calibrated aneroid barometer with a resolution of 0.2 torr in 

the SERI Instrumentation/~etrolog~ 

Recharge Center. Considering temperature effects on the pressure measurement the uncertainty in the 

MKS instrument is less than +0.3%. 

5.4 Flow Rate where 

We obtain the flow rates by measuring 

the pressure difference across ASMEstandard, long-radius flow nozzles. 

A set of nozzles will be used, ranging from 10-18 cm (4-7 in.) in diameter, to cover the range of flow 

rates (0.1-0.43 kgls). We installed 

ten diameters of straight duct with 

flow straighteners upstream of the 

nozzle to conform to ASME standards. 

Mass flow rate is calculated from 

where 

& = mass flow rate (kgls) 

D = duct diameter (m) 

C = discharge coefficient (dimensionless) 

B = ratio of nozzle diameter to duct 

diameter (dimensionless) 

R = gas constant for air 

(289.6 ~/kg-~) 

P = total absolute pressure (Pa) ~k = differential pressure across 

nozzle (Pa) 

t = temperature (OC). 

For a 0.305-m (12-in.) diameter duct 

the discharge coefficient (Bean 1971) 

is given by 

d = nozzle diameter (m) 

Red = Reynolds number at the nozzle 

throat. 

Since the uncertainty in Eq. 5-3 is 

larger than the variation with 

Reynolds number, the coefficient is 

calculated in the data aquisition and 

control program based on the set 

point conditions. This removes the 

need for an iterative solution technique. The uncertainty in discharge 

coefficient is given as *2% (Bean 

1971) together with the uncertainties 

in temperature and pressure; the 

uncertainty in flow rate, based on a 

Taylor series expansion and assumption of independent errors, is k3 .OX 

[see Schlepp, Schultz, and Zangrando 

(1984) for more details]. 

5.5 Wheel Speed 

As discussed in Section 4.4, wheel 

speed is measured by monitoring the 

output of the feedback tachometer on 

the wheel drive motor. We calibrated 

this output versus wheel speed by 

timing a number of revolutions of the 

drive wheel and comparing them with 

the output. In addition, we installed an optical encoder on the 

wheel drive motor. It outputs two 

signals: a 1 pulse per revolution 

and a 1000 pulse per revolution 

square wave train. By counting 

pulses we can accurately measure 

wheel speed and easily determine the 

control equation coefficient Kw (Section 4.4). Repeatability of average 

wheel speeds is less than k0.02 rph; 

variation in wheel speed with time is 

less than t0.03 rph. 
6.0 PRELIMINARY DATA 

6.1 D~sc~~D~~ou of Dehumidifier Wheel 

We chose as the desiccant a spirally 

wound, parallel passage design using 

silica gel as the first test article. 

The design is similar to prototypes 

tested at SERI in the past (Schlepp 

and Barlow 1984) under single blow 

conditions. This allows us to check 

the test system against previous work 

and provides a logical basis from 

which to proceed to cyclic testing. 

The parallel-passage design was also 

chosen to simplify the modeling efforts since the transfer coefficients 

for this geometry are well-known 

based on theoretical and experimen~al 

work and the solid-side moisture diffusion resistance is low (Schlepp and 

Barlow 1984). 

To construct the dehumidifier approximately 400 m of 25-um (0.001-in.) 

thick polyester film with an acrylic 

adhesive was coated with grade 11 

Davison, silica-gel particles ranging 

in size from 180-350 um. This coating process is shown in Figure 6-1. 

We used the Davison grade 11 gel because it contained the particle sizes 

we needed and has an isotherm very 

similar to the grade 40 previously 

used in Schlepp and Barlow (1984). 

The coated tape was then sent to 

Rotary Heat Exchangers Pty. Ltd. of 

Bayswater, Victoria, Australia, for 

winding. Rotary Heat Exchangers uses 

this winding technique to produce 

parallel-passage heat exchangers from 

a 75-pm (0.003-in.) thick polyester 

film. Table 6-1 summarizes the dehumidifer design. Figure 6-2 shows 

several views of the dehumidifier. 

Because of manufacturing difficulties, the windings are not as tight 

as desired; therefore, passage spacing is nonuniform. Although flow 

through the dehumidifier is fairly 

uniform overall, there are places 

where little air passes through and a 

lot of air passes through. This may 

be because of the compressibility of 

the desiccant-coated sheet between 

the spacers. The force exerted by 

the outer windings on the inner windings down along the spokes causes the 

desiccant particles to deform the 

tape film between the spacers as the 

particles from one side of the sheet 

fill the voids between the particles 

on the other side of the sheet. This 

compression releases the tension on 

the inner windings and causes 

sagging. 

In the time available to them Rotary 

Heat Exchangers was unable to find an 

engineering solution to the compressibility problem. However, it is not 

the winding technique used by Rotary 

Heat Exchangers that is at fault but 

rather a limitation of the materials 

supplied by SERI with which they had 

to work with. SERI is investigating 

possible solutions. Basically, the 

recommendation is to use a smaller, 

more uniform desiccant particle and a 

stronger tape film. 

In spite of these shortcomings we decided to conduct experiments on this 

wheel to see just how much effect the 

nonuniform spacing would have and 

also to provide a complete shakedown 

of the test facility. 

6.2 Seal Leakage Rates 

The purpose of this facility is to 

obtain data on the performance of a 

dehumidifier without concern for the 

physical design of the housing, 

seals, supports, etc. These effects 

will be present in the raw experimental data. However, they can be eliminated if they are known. Therefore, 

we measured seal leakage rates and 

correlated them with the appropriate 

pressure differences to correct the 

Table 6-1. Parameters of the First Test Dehumidifier 

Outside diameter 

Hub diameter 

Passage depth 

Number of spokes 

Spacer dimensions - thickness 

- width 

Number of windings 

Gel particle size range 

Sheet thickness (with gel) 

Passage width, average 

Sheet-to-sheet spacing, average 

Coated tape density (dry) 

Dry silica gel fraction 

Mass of dry gellperiod 

Mass of tapelperiod 

Face arealperiod 

raw data. , This method is based on 

Strahm and Wilson (1982). 

A 76-ym (3-mil) polyester film attached to the spokes that contacts 

the center support structure, as 

shown in Figure 6-3a, acts as a radial seal to separate the two airstreams. The radial seal length is 

approximately 0.6 m on each face. 

The contact area is such that at 

least one seal is always functioning. 

This seal design has been used by 

Rotary Heat Exchangers in their 

rotary heat exchangers. Similar 

material acts as a circumferential 

seal, shown in Figure 6-3b, to prevent air from flowing around the dehumidifier between the rotor and the 

outer housing. The film is serrated 

to prevent the seal from buckling and 

to ensure contact with the rim. 

Figure 6-4 shows the various leakage 

paths in the dehumidifier, and 

Table 6-2 lists typical values for 

the leakage rates. The out leakage 

rate is shown for the fans upstream 

of the wheel. If they were downstream, ambient air would leak into 

the system; however, other leakages 

would be unaffected because the pressure differences across the seals 

would be similar in either case. .For 

an inlet flow rate of 0.23 kg/s, the 

pressure difference through the wheel 

and so also across the radial and 

circumferential seals was approximately 68-75 Pa. 

6.3 Overall Facility Operation 

The test facility as described previously performed very well. Operation and control of the loop were 

fairly simple, although maintaining 

constant inlet humidity ratios was 

tedious, especially during the first 

hour or so of operation. 

Closure of the dehumidifier water 

and energy balances, as defined by 

the following equations, was 

satisfactory: 

Water balance closure: 


Figure 6-4. Dehumidifier Seal Leakage Paths 

PP out 

Process stream 

__1_1__, - 

Table 6-2. Typical Seal Leakage Rates 

mp 

7 

!I-4 rad 

='R 

Percentage of Inlet Flow (%) 

Locat ion 

Process Stream Regeneration Stream 

c t2-3 rad 

7 

Out 3.0-3.5 

Circumferential 2.0-2.5 

Radial 3.8-4.3 

In 4 f2-4 c~rc \ rn2 

1 

Regeneration stream 

where where 

W = desiccant moisture content 

(kg/kg) 

~DA = flow rate of dry air (kg/s) 

w = humidity ratio (kg/kg). 

Energy balance closure: 

h = specific enthalpy of moist air 

(J/kg dry air). 

The state points are numbered as in 

Figure 5-1. Values for AW are typically less than +3% with maximums of 

+5%. Values for AH are somewhat 

higher because of small differences 

between large numbers since the dehumidification process results in small 

air enthalpy changes. Similar results hold for the overall test rig. 
This is satisfactory for the present 

test article; however, improvements 

will be needed for testing higher 

performance wheels. 

6.4 Initial Tests 

Initial tests included several 

single-blow tests in an attempt to 

match previously taken data by 

Schlepp and Barlow (1984). Because 

of slight differences in geometry, we 

could not exactly match the transfer 

units and desiccant-to-air ratio. 

However, the single-blow data obtained matched the previous data well 

enough to indicate that the test 

facility and the wheel were performing satisfactorily. Also, we ran 

several cyclic tests to check full 

operation of the rig and wheel, again 

obtaining satisfactory results. 

These data are not presented here 

because the rig was not insulated at 

the time and, therefore, the energy 

balances did not close. 

After installing needed insulation, 

we conducted two sets of tests to determine several important parameters 

of the wheel. A set of very low rotation speed tests was performed to 

determine how much of the silica gel 

in the wheel is actually active. A 

set of very high rotation speed tests 

was performed to determine the effective Lewis number (ratio of the heat 

transfer coefficient to the mass 

transfer coefficient) of the wheel. 

Two parameters that describe the dehumidifier and that are used in the 

following analyses are the capacity 

rate ratio for period j, r and the j ' number of mass transfer units for 

period j, A j These are defined 

further in Jurinak (1982); Maclainecross (1974); Maclaine-cross and 

Banks (1972); Schlepp, Schultz, and 

Zangrando (1984); and Schultz and 

Schlepp (1984). For a particular 

wheel these parameters become simple 

functions of air flow rate and wheel 

rotation speed. For the first test 

dehumidifier these parameters are 

calculated as 

(Nu = 7.5 parallel-passage, laminar) 

(6-3) 

where 

mDAj = mass flow rate of dry air in 

period j (kg/s) 

N = wheel rotation speed (rph) 

p = indicates calculation is based 

on physical measurements, 

and where the constants 4.63 and 

0,00408 were determined from the 

physical design of the wheel. 

6.4.1 Low-Speed Tests 

At very low rotation speeds the wheel 

comes to equilibrium with the inlet 

airstream before it rotates into the 

next period. Each differential slice 

of the wheel essentially goes through 

a complete single-blow test. The 

method of analysis used here transforms the coupled heat and mass 

transfer processes in the dehumidifier into approximately uncoupled 

processes, each analogous to heat 

transfer alone, which are then superimposed (Jurinak 1982; Maclaine-cross 

1974; Maclaine-cross and Banks 1972). 

With appropriate assumptions that 

hold fairly well for high-performance 

dehumidifiers the governing equations 

for heat and mass transfer in the 

dehumidifier can be transformed into 

a set of equations in terms of two 

potentials, Fl and F2. These potentials are simllar to air enthalpy and 

relative humidity lines on a psychrometric chart. The transformed equations, which have the same form as 

the equation for heat transfer alone, 

are as follows (Jurinak 1982; 
Maclaine-cross 1974; Maclaine-cross 

and Banks 19721, 

where 

Fa = potential 1 or potential 2 : = period 1 (process) or period 

2 (regeneration) 

0' = dimensionless time 

x' = dimensionless length 

and 

where 

Cij = reciprocal of the dimensionless 

wave speed for potential i in 

period j 

y = property ratio analogous to the 

ratio of heat capacities 

mDD = mass of dry gel (kg) 

%A = flow rate of dry air (kgls). 

Because of the values of y, 

wave moves through the wheel muc the Fi 

more slowly than the F wave does. 

At Low rotation speeds, a efined as 

wave moves completely through 

the w eel before it rotates into the 

next period. An F2 effectiveness q 

can be defined analogously to ti2 

heat transfer effectiveness and it 

can be shown that 

for very low rotation speeds 

(Maclaine-cross 1974; Maclaine-cross 

and Banks 1972). We can calculate 

 TI^^ from test data using the following equations: 

where the state points are as shown 

in Figure 6-5. Assuming that the 

properties of the gel/matrix combination are known, test data at very 

low rotation speeds would provide a 

check on the amount of active dry gel 

m~pa in the dehumidifier. The silica 

ge and matrix properties used are 

given in Table 6-3. 

To perform this check the five tests 

listed in Table 6-4 were run and analyzed with the above method. The 

outlet states of each test are shown 

in Figure 6-5. Note that as the 

rotation speed increases, the outlet 

states move toward the intersection 

point of the Fi characteristics. As 

the wheel speed increases, more air 

is processed by the wheel and less 

comes through unaffected. 

The F2 effectivenesses are plotted 

against C2. calculated from physical 

measuremenJg in Figure 6-6. Note 

that the slope of the line deviates 

from unity even at very low C2 .p. A 

linear least squares fit througA zero 

and the results of tests LS-1A and 

LS-2A, shown in Figure 6-7, give a 

slope of 0.903. This indicates that 

the active amount of silica gel in 

the matrix is approximately 90% of 

that calculated to be there, or 

where the subscript a refers to the 

active value and the subscript p 

refers to the physical value. This 

could be a result of the tape adhesive blocking some of the silica gel 

pores and reducing the gel's water 

holding capacity. Future sorption 

equilibrium tests at SERI on the tape 

and 
A 

I - 30 

P:, = 0.80 atm 

41 P 

- 20 

- 10 

1 I I I I 1 I I .o 

10 20 30 40 50 60 70 80 90 100 

Temperature, t (O C) 

Figure 6-5. Low Rotation Speed Outlet States 

Table 6-3. Silica Gel and Matrix Properties - -- - - 

Equilibrium relationship (Brandemuehl 1982) 

Pv = --- I [(29.91)(2.112~)~~~~ lhs/hv , 29.91 

where 

Pv = equilibrium vapor pressure (atm) 

Psat = saturation vapor pressure (atm) 

W = desiccant moisture content (kg/kg). 

Heat of adsorption (Brandemuehl 1982) 

hs/hv = 1 + , 

where 

hs = heat of adsorption 

hv = latent heat of vaporization of water. 

Specific heat of silica gel (~chlepp and Barlow 1984) = 921 ba kg OC 

Specific heat of tape (Schlepp and Barlow 1984) = 1840 k kg OC 
Table 6-4. LorRotation Speed Tests - -- -- - - - -- 

Average inlet conditions 

Intersection points 

t3iP = 60.4'~ 

Average ys 

y21 = 23.7 

Outlet states 

Test No. t3 W3 

(OC) (g/kg) 

Effectivenesses 

Test No. "2 1 "22 C2 1~ C22~ '2 la '22a 

and gel combination may answer the A = 22 or NTUo = 11 , Latter question. j P 

Shown in Figure 6-8 is the expected where 

theoretical results for F2 effectiveness. The deviation appears to be be- A = predicted number of mass 

cause of a reduced number of transfer jp transfer units for period j 

units. Based on the physical dimensions of the wheel, the NTUs are cal- NTUo = overall number of transfer 

culated to be units. 

A Process period 

Regeneration period 

A Process (theory, NTU, = 11) 

0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1.0 1.1 1.2 

C,, (physical) j = 1: process; j = 2: regeneration 

Figure 6-8. P2 Effectivess versus C2je 

The apparent values from the data 

appear to be 

where 

A = apparent number of mass trans- ja fer units for period j. 

This result is discussed further in 

Section 7.0. 

6.4.2 High-Speed Tests 

At very high rotation speeds the dehumidifier acts as a total enthalpy 

exchanger. The outlet of one stream 

approaches the inlet of the opposite 

stream. In terms of the analogy 

method the fast moving El wave does 

not have time to move completely 

through the wheel, 

If the inlet conditions have the same 

humidity ratio, the wheel will then 

seem to act as a temperature exchanger. From an experimentally 

obtained temperature effectiveness 

we can determine the overall number 

of transfer units for heat transfer 

NTUto (Maclaine-cross and Banks 1972; 

Kays and London 1964). If the inlet 

conditions have the same temperature 

but different humidity ratios, the 

wheel acts as a mass exchanger. As 

with temperature, we can determine 

the overall number of transfer units 

for moisture transfer NTUw . The 

ratio of these values is tEe Lewis 

number, 

NTU+ n 

Using the notation of Kays and London 

(19641, 
where 

E = temperature or moisture effec- 

.#. t iveness c " - 

- Cm.n'Cmax c: = crfcmin 

and 

= &DAC~A I for heat transfer 

= ~DA for mass transfer 

I - 

- ['DD~(C~DD + WcpW) + mTcpT]N 

r for heat transfer (6-13) 

( = mDDaN for mass transfer, 

where 

A = moist air 

DA = dry air 

DD = dry desiccant 

T = tape 

W = water. 

Table 6-5 lists the tests we ran to 

check the Lewis number of the first 

test dehumidifier. The state points 

are shown on a psychrometric chart in 

Figure 6-9. The results show that 

the wheel exhibics a Lewis number 

near unity. Based on period 2 results, the Lewis number could be as 

high as 2. This variation is a result 

of uncertainties in the mass and energy balance closures. 

Again, as for the low speed tests, 

the experimentally determined NTUs 

are about a factor of 3-4 less than 

was expected based on the physical 

dimensions of the wheel. For set A 

it was 

and for set B, 

This is most likely a manifestation 

of the nonuniform spacing of the 

windings and subsequent nonuniform 

flow through the matrix. Further 

testing will be done to resolve this 

question. We are interested in 

whether or not the heat and mass 

transfer coefficients are affected 

similarly. 
Table 6-5. High Rotation Speed Tests 

Inlet States 

"'IDA tl " 1 m2~~ t2 2 N Test No. 

(kg/s) (OC > (g/kg) (kg/s) (OC) (g/kg) (r/h) 

- -- p~~ - 

Outlet States for Period 1 

Regenerator Parameters 

Heat Transfer Mass Transfer 

Test No. J. c3k ,yk c " 

r NTUt o r NTUwo c * 

HS-1A 0.99 5.4 3.55 0.97 2.4 3.40 1.04 

HS-2A - - - - - - 0.98 2.4 

HS-3A 0.97 5 .O 3 .OO - - - - 3*00\ - - 1.00 

7.0 CONCLUSIONS 

7.1 Facility Status and Preliminary 

Test Results 

SERI's Desiccant Cooling Cyclic Test 

Facility is assembled and operating. 

The facility can test bench-scale, 

rotary solid desiccant dehumidifiers 

over a wide range of controlled conditions representative of those encountered by solar desiccant cooling 

systems. We can monitor transient 

and steady-state operation. The cloInitial tests indicate that approximately 90% of the silica gel thought 

to be present is active. The overall 

Lewis number of the wheel is near 

unity. However, the nonuniform flow 

passage spacing seems to be reducing 

the effective number of trans£ er 

units from theory by a factor of 3-4 

We will perform further tests to tr) 

to understand this. 

sure of the overall mass and energy 7.2 Future Work 

balances on the rig - are satisfactory 

for experiments on the first test 

article; however, improvements will 

be needed to test higher performance 

wheels. We can acquire accurate data 

suitable for model validation. 

We can presently control inlet humidity ratios satisfactorily, although 

the process is somewhat tedious. 

Velocities at the wheel inlet faces 

are somewhat nonuniform and highly 

turbulent. The current dehumidifier 

appears to provide sufficient pressure drop to even out the flow. However, it may be desirable to redesign 

the entrance ducting for future 

tests. 

A spirally wound, parallel-passage, 

rotary dehumidifier using silica gel 

as the desiccant was constructed and 

installed in the test loop. Initial 

tests, single-blow and cyclic, were 

conducted with satisfactory results. 

The test matrix presented in Schultz 

and Schlepp (1984) will be carried 

out, This will fully characterize 

the performance of this first rotary, 

parallel-passage dehumidifier and 

allow us to validate the models for 

this design over a wide range of 

operating conditions. 

Complete validation of the analytical 

models requires a range of designs to 

be tested. Based on the results of 

the first dehumidifier, we will acquire and test other dehumidifier 

geometries. Several choices are 

available, including constructing a 

second parallel-passage dehumidifier 

with improved passage spacing uniformity (Maclaine-cross 1985). This 

could be done at SERI or by other 

commercial manufacturers. A choice 

will be made as soon as sufficient 

information is available from the 

first dehumidifier. 
8.0 REFERENCES 

Ambrose, C. W., I. L. Maclaine-cross, Howe, R. R., W. A. Beckman, and J. W. 

and E. B. Robson, 1979, The Use of Mitchell, 1983 (June), "~ornrnercial 

Rotary Regenerative Heat Exchangers Applications for Solar Hybrid 

for the Conversion of Energy in Desiccant Systems," Proceedings of 

Buildings, Clayton, Victoria, the 1983 Annual ASES Meeting, 

Australia: Monash University, Anaheim, CA: ASES. 

Dept. of Mech. Engr. 

Huskey, B., et al., 1982, Advanced 

Barlow, R. S., 1982 (Dec.), Analysis SolarIGas Desiccant Cooling System, 

of the Adsorption Process and of GRI 81/10064, Chicago, IL: Gas ReDesiccant Cooling Systems: A search Institute. 

Pseudo-Steady-State Model for Coupled Heat and Mass Transfer, Jurinak, J. J., 1982 (Aug.), OpenSERI/TR-631-1330, Golden, CO: Cycle Desiccant Cooling: Component 

Solar Energy Research Institute. Models and System Simulations, 

Ph.D. thesis, Madison: University 

Bean, H. S., ed., 1971, Fluid Meters, of Wisconsin, Solar Energy LaborTheir Theory and Application, 6th atory. 

edition, New York: ASME Research 

Committee on Fluid Meters. 

Booz-Allen and Hamilton, 1981, Evaluation of Residential and Commercial SolarIGas Heatine and Cooling Technologies, GRI-7910105, 

Chicago, IL: Gas Research 

Institute. 

Brandemuehl, M. J., 1982, Analysis of 

Heat and Mass Transfer Regenerators 

with Time Varying or Spatially Nonuniform Inlet Conditions. Ph.D. 

thesis, Madison: University of 

Wisconsin, Dept. of Mech. Engr. 

Clark, J. E., 1979, Design and Construction of Thin, Adiabatic Desiccant Beds for Solar Air Conditioning Applications, University of 

California at Los Angeles. 

Kang, T. S., 1985, "~diabatic Desiccant Open Cooling Cycles," Master's 

thesis, Kensington, Australia: 

University of New South Wales, 

School of Mech. and Ind. Engr. 

Kays, W. M., and A. L. London, 1964, 

Compact Heat Exchangers, New 

York: McGraw-Hill. 

Kutscher, C. F., and R. S. Barlow, 

1982 (Aug.), Dynamic Performance of 

Packed-Bed Dehumidifiers: Ex~erimental Results from the SERI Desiccant Test Laboratory, SERI/TR-252- 

1429, Golden, CO: Solar Energy 

Research Institute. 

Maclaine-cross, I. L., 1974, A Theory 

of Combined Heat and Mass Transfer 

in Regenerators, Ph.D. thesis, 

Clayton, Victoria, Australia: 

Monash University, Dept. of Mech. 

Cohen, B., A. Levine, and R. Arora, Engr . 1983 (~eb.1, Field ~evelo~rnent of a - 

Desiccant-Based Space-Conditioning Maclaine-cross, I. L., 1985 (Apr. 1, 

System for Supermarket Appli- Potential Improvements to the SERI 

cations, TE4308-42-83, Chicago, Parallel-Passage Dehumidifier 

IL: Gas Research Institute. Matrix, Internal Branch Report, 
Thermal Research Branch 252, 

Golden, CO: Solar Energy Research 

Institute. 

Maclaine-cross, I. L., and P. J. 

Banks, 1972, "coupled Heat and Mass 

Transfer in Regenerators - Prediction Using an Analogy with Heat 

Transfer," lnttl J. of Heat and 

Mass Transfer. Val. 15. DD. 

Macriss, R. A., and T. S. Zawacki, 

1982 (Feb.), "~igh COP Rotating 

Wheel Solid Desiccant system," Proceedings of the 9th Energy Technology Conference, Washington, DC. 

Monier, J. B., W. M. Worek, and Z. 

Lavan, 1982 an.), Testing of a 

Cross-Cooled Solar-Powered Desiccant Cooling - System. DES-82-1, 

Chicago, IL: Illinois Institute of 

Penney, T., and I. Maclaine-cross, 

1985 (May), Promising Advances in 

Desiccant Cooling, SERIlTP-252- 

2683, Golden, CO: Solar Energy Research Institute. 

Rousseau, J., 1982 (Nov.), Development of a Solar Desiccant Dehumidifier: Final Summary Report, 82- 

18944, Torrence, CA: Garrett 

AiResearch Manufacturing Co. 

Schlepp, D., and R. Barlow, 1984 

1, Performance of the SERI 

Parallel-Passage Dehumidifier, 

SERI/TR-252-1951, Golden, CO: 

Solar Energy Research Institute. 

Schlepp, D., and K. Schultz, 1984 

(Sept.), High Performance Solar 

Desiccant Cooling Systems: Performance Evaluations and Research 

Recommendations, SERI~TR-252-2497, 

Golden, CO: Solar Energy Research 

Institute. 

Schlepp, D., K. Schultz, and F. 

~angrando, 1984 (Aug. 1, Facility 

Design for Cyclic Testing of Advanced Solid Desiccant Dehumidifiers. I SERI/TR-252-2464. Golden. 

CO: Solar Energy Research 

Institute. 

Scholten, W. B., and J. H. Morehouse, 

1983, Active Program Research Requirements: Final Report, ~c~ean, 

VA: Science Applications, Inc. 

Schultz, K., and D. Schlepp, 1984 

(Apr. j , ~Gerimental Test -Plan for 

Cvclic Testing of Parallel Passage 

Rotarv Desiccant Dehumidifiers. 

Draft, Golden, CO: Solar Energy 

Research Institute. 

Sharp Co., Product literature, 1982, 

Nora, Japan. 

Strahm, C. S., and S. G. Wilson, 

1982a (July), Dehumidifier Seal 

Development, DET. 21, Highett, 

Victoria, Australia: CSIRO Division of Energy Technology. 

Strahm, C. S., and S. G. Wilson, 

1982b (July), Dehumidifier Test 

Facility Rig Description, Operation. and Maintenance. DET. 22. 

Highett, Victoria, Australia: 

CSIRO Division of Energy Technology. 

van den Bulck, E., 1983, "Analysis of 

Solid Desiccant Rotary Dehumidifiers," M.S. thesis, Madison: 

University of Wisconsin, Dept. of 

Mech, Engr. 

van Leersum, J., et al., forthcoming, t t Comparison of the Results from 

Experiments on a Rotary Particulate 
Bed Dehumidifier with Those from a Venhuizen, D., 1984 (Oct.), "Solar 

Numerical Model," submitted to ASME ~ing's Cooling A ambit," Solar Age, 

J. of Solar Energy Engr. Vol. 9, No. 10, p. 25-27. 

van Leersum, J. G., and D. J. Close, 

1982 (June), Experimental Verification of Open Cycle Cooling 

System Component Models, Draft, 

Madison: University of Wisconsin. 

Wurm, J., et al., 1979 (NOV.), SolarMEC Development Program: Semiannual Progress Report, C00-4495- 

23, Chicago, IL: Institute of Gas 

Technology. 
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Page 1 SERTR-252-278 I - ...--> .. -* 

%EW d _Subt~tle Yeslccant Coolliig Tesc Tacility: Status Report 

Preliminary Data on the Performance of a Rotary 

Parallel-Passage Silica-Gel Dehumidifier 

I 

7. Author@) 18. Performing Organization Rept. No. 

Kenneth J. Schultz 

9. Perform~ng Organization Name and Address 

Solar Energy Research Institute 

1617 Cole Boulevard 

Golden, Colorado 80401 

I I Technical Report 

16. Project/Task/Work Unit No. 

3023.210 

11. Contract (C) or Grant (G) No. 

(C1 

12. Sponsoring Organization Name and Address 

I 

15. Supplementary Notes 

13. Type of Report & Period Covered 

16. Abstract (Limit: 200 words) 

This report describes the SERI Desiccant Cooling Test Facility. The facility can 

test bench-scale rotary dehumidifiers over a wide range of controlled 

conditions. We constructed and installed in the test loop a prototype parallelpassage rotary dehumidifer that has spirally wound polyester tape coated with 

silica gel. The initial tests gave satisfactory results indicating that 

approximately 90% of the silica gel was active and the overall Lewis number of 

the wheel was near unity. The facility has several minor difficulties including 

an inability to control humidity satisfactorily and nonuniform and highly 

turbulent inlet velocities. To completely validate the facility requires a range 

of dehumidifier designs. Several choices are available including constructing a 

second parallel-passage dehumidifier with the passage spacing more uniform. 

1 Document Control 11. SERI Report NO. 1 2. NTlS Accession No. 1 3. Recipient's Accession No. 1 

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17. Document Analysis 

a.Descriptors Cooling ; Dehumidifiers ; Desiccants ; Test Facilities ; Testing 

b. Identifiers/Open-Ended Terms 

c. UC Categories 

59a 

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National Technical Information Service 

U.S. Department of Commerce 

5285 Port Royal Road 

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20. Price 

I I 

Form No. 0069 (3-25-82)