Design & Engineering Services
Performance Evaluation of an EvaporativelyCooled Split-System Air Conditioner
ET 08.08 Report
Prepared by:
Design & Engineering Services
Customer Service Business Unit
Southern California Edison
November 20, 2009
Performance Evaluation of an Evaporatively-Cooled Split-System Air Conditioner
ET 08.08
Acknowledgements
Southern California Edison’s Design & Engineering Services (D&ES) group is responsible for
this project. It was developed as part of Southern California Edison’s Emerging Technology
program under internal project number ET 08.08. This technology evaluation was conducted
at the Technology Test Centers (TTC) within the D&ES group with overall guidance and
management from Paul Delaney. Project manager: John Lutton, Senior Lab Technician:
Bruce Coburn. For more information on this project please contact Paul.Delaney@sce.com.
Southern California Edison wishes to thank the Beutler Corporation for the contribution of
their AquaChill evaporatively-cooled condensing unit residential split-system air conditioner
for purposes of evaluation and testing at our laboratories.
Disclaimer
This report was prepared by Southern California Edison (SCE) and funded by California
utility customers under the auspices of the California Public Utilities Commission.
Reproduction or distribution of the whole or any part of the contents of this document
without the express written permission of SCE is prohibited. This work was performed with
reasonable care and in accordance with professional standards. However, neither SCE nor
any entity performing the work pursuant to SCE’s authority make any warranty or
representation, expressed or implied, with regard to this report, the merchantability or
fitness for a particular purpose of the results of the work, or any analyses, or conclusions
contained in this report. The results reflected in the work are generally representative of
operating conditions; however, the results in any other situation may vary depending upon
particular operating conditions.
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Performance Evaluation of an Evaporatively-Cooled Split-System Air Conditioner
ET 08.08
ABBREVIATIONS AND ACRONYMS
A/C
Air Conditioner
AHU
Air Handler Unit
AHRI
Air Conditioning, Heating and Refrigeration Institute (formerly ARI)
ARI
Air Conditioning and Refrigeration Institute
ASHRAE
American Society of Heating, Refrigeration, and Air Conditioning Engineers
BTU
British Thermal Unit
CFM
Cubic Feet per Minute, ft3/min
CPU
Central Processing Unit
CTAC
Customer Technology Application Center
CTZ
California Thermal Zone (also: Climate Zone)
DAT
Discharge Air Temperature, °F
DB
Dry-bulb temperature, °F
DX
Direct Expansion
ECCU
Evaporatively-cooled Condensing Unit
EER
Energy Efficiency Ratio
EMS
Energy Management System
fpm
Feet per minute
HDAC
Hot and Dry (region) Air Conditioner
hg
Inches of Water Gauge
HP
Horse Power
HVAC
Heating, Ventilation, and Air Conditioning
NIST
National Institute of Standards and Technology
PSI
Pounds per Square Inch (gauge)
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PSIA
Pounds per Square Inch (Absolute)
RH
Relative Humidity, %Rh
RMS
Root Mean Square
RTD
Resistive Thermal Device
RTTC
Refrigeration and Thermal Test Center
SCE
Southern California Edison
SCFM
Standard Cubic Feet per Minute
SH
Superheat
SHR
Sensible Heat Ratio
TC
Thermocouple
TXV
Thermostatic Expansion Valve
TTC
Technology Test Centers
UA
Heat Transfer Coefficient
VSD
Variable Speed Drive
WB
Wet-bulb temperature, °F
WCEC
Western Cooling Efficiency Center
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FIGURES
Figure 1. Entire Test Facility Layout..............................................8
Figure 2. Scroll Compressor Rack (a) and EMS Panels (b) ................9
Figure 3. Supply Air Plenum (a) and Return Air Opening with
Rack-Mounted Filters (b) ..............................................9
Figure 4. Variable Speed Drives of AHUs Serving Indoor (a) and
Outdoor (b) Controlled Environment Rooms ....................9
Figure 5. Outdoor (a) and Indoor (b) Controlled Environment
Rooms with Supply and Return Ducts ........................... 10
Figure 6. Psychrometric Plot of All Design Condition Climate Zones
Tested ..................................................................... 13
Figure 7. Graphical Data Acquisition Environment......................... 14
Figure 8. Modules for SCXI High Performance Signal Conditioning
and Switching Platform .............................................. 14
Figure 9. PCI 6052E Analog Input Card ....................................... 14
Figure 10. Four Externally Manifolded Pressure Taps Before Inlet of
the Test Unit............................................................. 16
Figure 11. Instruments for Measuring Electrical Input Power to Fan
Motors, Compressor Motor, and the Test Unit ................ 17
Figure 12. Digital Scale with Condensate Collection System ............. 17
Figure 13.
Heater/Chiller Loop (a) Condenser Water Inlet (b)
and Water Mass Flow Meter (c) ................................... 18
Figure 14. ASHRAE Air Flow Measurement Station......................... 19
Figure 15. Coriolis Refrigerant Mass Flow Meter ............................ 29
Figure 16. Performance Comparison of Manufacturer's Test Data to
TTC’s Test Data at AHRI 210/240 Baseline Conditions..... 32
Figure 17. Test Results of Evaporative Condenser A/C unit in
Various Outdoor Climate Zones ................................... 33
Figure 18. Water Consumed by Purge and Evaporation at Various
Climate Zone Conditions............................................. 35
Figure 19. Psychrometric Chart Showing Δω versus Δh for the
ECCU in Climate Zone 7 ............................................. 36
Figure 20. Psychrometric Chart Showing Δω versus Δh for the
ECCU in Climate Zone HDAC ....................................... 37
Figure 21. Total Water Consumption Rate versus Change in
Humidity Ratio of Outdoor Conditions ........................... 38
Figure 22. Total Water Consumption as a Function of Relative
Humidity .................................................................. 39
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Figure 23. Net Capacity, EER and Water Consumption as a
Function of Increasing Dry bulb Across Climate Zones .... 40
Figure 24. Air-Cooled Unit Performance at Various Dry bulb
Temperatures (Reference 1) ....................................... 41
Figure 25. Net Cooling Capacity Normalized per Ton for the
Evaporatively-cooled versus Air-Cooled Type A/C
Systems as a Function of Outdoor Dry bulb
Temperature............................................................. 42
Figure 26. Power Consumed by the Evaporatively-cooled versus
Air-Cooled Type A/C Systems as a Function of Outdoor
Dry bulb Temperature ................................................ 44
Figure 27. ECCU Power Consumption versus Increasing Wet Bulb
Temperature............................................................. 45
Figure 28. ECCU Discharge and Suction Pressure versus Increasing
Wet Bulb Temperature ............................................... 45
Figure 29. ECCU Saturated Condensing Temperature versus
Increasing Wet Bulb Temperature ................................ 46
Figure 30. EER of Evaporatively-cooled versus Air Cooled Type A/C
Systems as a Function of Outdoor Dry bulb
Temperature............................................................. 47
TABLES
Table 1.
Operating Conditions for Standard Rating Using AHRI
210/240................................................................... 11
Table 2.
Design Condition Climate Zones with Representative
Temperatures ........................................................... 12
Table 3.
Specifications of Sensors Used ..................................... 20
Table 4.
Minimum Requirements for Data Collection During
Cooling Capacity Tests as Specified by AHRI.................. 21
Table 5.
Summary Table of ECCU Performance Parameters Across
the different Climate Zone Conditions Tested................. 57
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EQUATIONS
Equation 1.Volumetric Airflow Rate .............................................. 25
Equation 2.Standard Volumetric Airflow Rate................................. 25
Equation 3.Air Mass Flow Rate .................................................... 25
Equation 4.Gross Cooling Capacity- Air-Enthalpy ........................... 26
Equation 5.Converting Btu/Hr to Tons .......................................... 26
Equation 6.Net Cooling Capacity.................................................. 26
Equation 7.Mass of Collected Condensate- Psychrometric................ 27
Equation 8.Latent Cooling Capacity- Psychrometric ........................ 27
Equation 9.Latent Cooling Capacity- Condensate ........................... 27
Equation 10.
Sensible Cooling Capacity ...................................... 28
Equation 11.
Volumetric Flow Rate per Gross Cooling Capacity ...... 28
Equation 12.
EER – Energy Efficiency Ratio ................................. 28
Equation 13.
Sensible Heat Ratio .............................................. 29
Equation 14.
Refrigeration Effect............................................... 30
Equation 15.
Gross Cooling Capacity- Refrigerant-Enthalpy ........... 30
Equation 16.
Compressor Efficiency ........................................... 31
Equation 17.
Moist Air Specific Enthalpy ..................................... 36
Equation 18.
Temperature Differential Across Evaporator.............. 52
Equation 19.
Evaporator Temperature Difference......................... 52
Equation 20.
Evaporator Coil Effectiveness ................................. 53
Equation 21.
Temperature Differential Across Condenser Coil ....... 53
Equation 22.
Condenser Temperature Difference......................... 53
Equation 23.
Condenser Arithmetic Mean Temperature Difference.. 54
Equation 24.
Condenser Heat of Rejection.................................. 54
Equation 25. Condenser Overall Heat Transfer Coefficient .............. 54
Equation 26.
Evaporator Coil Superheat..................................... 55
Equation 27.
Condenser Coil Subcooling .................................... 55
Equation 28.
Total System Subcooling....................................... 55
Equation 29.
Heat of Compression ............................................ 56
Equation 30.
Work of Compression ........................................... 56
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CONTENTS
EXECUTIVE SUMMARY _______________________________________________ 1
INTRODUCTION ____________________________________________________ 3
BACKGROUND ____________________________________________________ 4
ASSESSMENT OBJECTIVES ____________________________________________ 5
PRODUCT EVALUATED _______________________________________________ 6
Air Cooled versus Evaporatively-cooled Condensers.....................6
TECHNICAL APPROACH/TEST METHODOLOGY ____________________________ 8
Test Facility...........................................................................8
TEST METHODOLOGY ______________________________________________ 11
Test Protocol ....................................................................... 11
Test Method ........................................................................ 11
Instrumentation and Data Acquisition...................................... 13
Temperature Measurements .................................................. 15
Thermocouples .................................................................... 15
Refrigerant Temperature Measurements .................................. 15
Air Temperature & Humidity Measurements ............................. 15
Pressure Measurements ........................................................ 16
Electrical Measurements........................................................ 16
Condensate and Purge Measurements ..................................... 17
Condenser Sump and Spray Water Measurements .................... 18
Condenser Supply Water Measurements .................................. 18
Airflow Measurements........................................................... 19
Specifications of Sensors ....................................................... 20
Data to be Recorded ............................................................. 21
EVALUATIONS ____________________________________________________ 24
Approach ............................................................................ 24
Cooling Capacity .................................................................. 24
Air-Enthalpy Method ........................................................ 24
Refrigerant-Enthalpy Method............................................. 29
RESULTS ________________________________________________________ 32
Capacity and Performance of the Unit...................................... 32
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Water Consumption .............................................................. 34
Purge ............................................................................ 34
Evaporation .................................................................... 34
Performance and Water Consumption...................................... 39
Performance Comparison To Air-Cooled Air Conditioning Systems 40
Net Cooling Capacity........................................................ 41
Power ............................................................................ 43
Energy Efficiency Ratio (EER) ............................................ 46
CONCLUSIONS ___________________________________________________ 48
Performance........................................................................ 48
Water Consumption .............................................................. 48
Water Quality and Maintenance .............................................. 49
RECOMMENDATIONS ______________________________________________ 50
REFERENCES _____________________________________________________ 51
APPENDICES _____________________________________________________ 52
Appendix A.......................................................................... 52
Supplemental Performance Evaluations .............................. 52
Evaporator Coil Characteristic Performance .............................. 52
Condenser Coil Characteristic Performance .............................. 53
System Characteristic Performance......................................... 55
Heat and Work of Compression .............................................. 56
Appendix B.......................................................................... 57
Summary table of ECCU performance parameters ..................... 57
Notes: ................................................................................ 58
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EXECUTIVE SUMMARY
It has long been known that evaporatively-cooled condensers in air-conditioning systems
provide increased efficiency over air cooled condenser technology. The increased efficiency
is especially effective during peak demand periods that correspond with the hottest part of
the day. Some of the unknowns are the performance of these units over a wide range of
ambient conditions, the comparative performance to air cooled units, and information about
water consumption due to purging and evaporation.
The primary goal of this project is to evaluate the performance of an evaporatively-cooled
condensing unit as part of a residential split-system air conditioner in some of the climate
zone conditions found in Southern California Edison’s (SCEs) service territory. Additional
goals are to determine the performance degradation of the unit with increasingly harsh
climate conditions, look at general water consumption of the unit in these different climate
conditions, and to compare normalized performance data to existing information on aircooled condenser type air conditioning units from previous lab research conducted at SCEs
Technology Test Centers (TTC).
The evaporatively-cooled condensing unit of the residential split-system was installed in a
controlled environment room of the TTC and every major component was instrumented for
data collection. Likewise, the indoor unit of the split-system was installed in a separate
controlled environment room and every major component instrumented for data collection.
Baseline tests were then conducted where a minimum of two hours of test data was
collected under established indoor and outdoor conditions. These respective temperatures
and humidities derive from the rating conditions required by the Air-Conditioning, Heating &
Refrigeration Institute’s Standard 210/240-2003 for Unitary Air-Conditioning and Air-Source
Heat Pump Equipment.
The results of the Technology Test Center’s baseline data were then compared to the results
of the baseline tests previously conducted by the manufacturer at identical test conditions.
The closely matching results of this comparison established confidence in the results of
these independently conducted tests.
After the results were confirmed to be in close agreement, further tests were conducted at
various outdoor climate zone conditions. Each different climate zone test was conducted
under steady-state operation with the appropriate constant temperature and humidity in
each of the indoor and outdoor test chambers.
Data was collected for a minimum of two hours under several outdoor conditions
representative of climate zones within SCE service territory. Measured parameters were
later compared for performance variations across the different climate zone conditions.
The findings show that evaporatively-cooled condenser technology is able to produce the
same cooling capacity at lower energy consumption than current air-cooled condenser
technology. The efficiency of the evaporatively-cooled condenser decreased slightly
(approximately 10%) at the extreme climate zone conditions of high dry bulb and low wet
bulb but was largely unchanged across less extreme climate zone conditions. By contrast,
the efficiency of previously tested air cooled condensing units decreased severely (a drop of
34%) while increasing energy consumption at similar high dry bulb conditions.
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As expected, the evaporatively-cooled condenser technology showed increased water
consumption at hotter and dryer climate zone conditions. While the amount of water
consumed was not excessive, it should be weighed against the electrical energy savings
achieved in climate zone conditions where water consumption is an issue.
The manufacturer of the unit was able to down-size the compressor by taking advantage of
the heat transfer effectiveness of water and the evaporative cooling process to lower the
refrigerant saturated condensing temperature thereby increase the refrigeration effect. This
smaller compressor consumes less power for the same cooling capacity as an air-cooled
condensing unit with a larger compressor. The ability to obtain the same cooling capacity
with a smaller compressor results in an overall energy savings.
The performance of this evaporatively-cooled condensing unit shows the potential for
significant energy savings over similar air-cooled condensing technology of the same
capacity. The total energy savings could be substantial given significant market
penetration. This technology merits investigation into the creation of a customer rebate
program that will encourage market penetration.
Water consumption, due to evaporation and purge cycles, may be an issue to consider in
regions of water scarcity or in hot and dry climate zones where water consumption will
naturally be higher.
For purposes of determining energy savings, the amount of energy saved should account for
the amount of water consumed by evaporation and purge cycles, especially in regions where
water usage is of concern. Whether to use an energy cost versus water cost basis or other
basis is beyond the scope of this project.
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INTRODUCTION
Southern California Edison’s (SCEs) significant peak electrical demand is a result of the vast
number of air conditioners operating during the hot summer days. The efficiency of
conventional air-cooled air conditioning systems equipped with air cooled condensers
decreases as the outdoor temperature rises (see Reference 1). Because electricity cannot be
economically stored on a large scale, new power plants must be constructed to meet the
increasing peak demands of the state. A wide-scale reduction in energy consumption of air
conditioners on hot days would dramatically reduce peak demand.
It has long been known that evaporatively-cooled condensers in air conditioning systems
provide increased efficiency over air-cooled condenser technology. The increased efficiency
is especially effective during peak demand periods that correspond with the hottest summer
days.
A commercially available air conditioning system was developed using an evaporativecooled condensing unit (ECCU), which operates at much higher efficiency in hot weather
than conventional units with air cooled condensers.
Some of the unknowns are the performance of these units over a wide range of ambient
conditions, the comparative performance to air-cooled units, and information about water
consumption due to purging and evaporation.
If the manufacturer’s projections are validated, this technology could provide peak demand
reduction benefits to SCE and potentially provide significant long-term air conditioning
energy savings to customers in the hotter climate zones where air conditioning energy
savings are needed most.
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BACKGROUND
A large portion of SCE's peak power demand during the summertime is due to operation of
air conditioning systems. Broad implementation of a more efficient air conditioning
technology can reduce customer electrical energy usage and result in a significant reduction
in utility system peak demand requirements.
An evaporatively-cooled condensing unit was developed for a residential split-system air
conditioner in which water is sprayed on the condenser coils during operation to cool the
refrigerant in the coils. This provides a condensing temperature much lower than a
conventional air-cooled condenser system, resulting in greater operating efficiency and
reduced electrical demand.
The manufacturer claims efficiency improvements of up to 40% over air-cooled condenser
air conditioning (A/C) units on the hottest days in areas with high outdoor temperatures and
low humidity.
Water is consumed by evaporation in the cooling section of the condenser and by a periodic
flushing of the sump to avoid excessive mineral build-up in the cooling water. The water
usage can be a concern as there are many desert climate zones within SCE service territory.
The evaporative condenser air conditioner system is currently in limited production.
Investigating manufacturer performance claims and documenting water use characteristics
can help overcome market barriers and provide a basis for more widespread acceptance of
this emerging technology.
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ASSESSMENT OBJECTIVES
The primary objectives of this assessment are to quantify the cooling capacity, energy
efficiency and the water consumption characteristics of the evaporative condenser A/C
system in different SCE service territory climate zones.
In order to meet these primary objectives, several intermediate steps have to first be
satisfied. The overall sequence is detailed as follows:
1. Install, instrument and put into operation the unit to be tested.
2. Perform baseline ASHRAE/AHRI tests to quantify baseline performance.
3. Compare baseline performance to manufacturer’s baseline results for
agreement/confidence of the independent results.
4. Perform ASHRAE/AHRI tests at pre-selected SCE service territory climate zone
conditions.
5. Compare cooling capacity performance and energy efficiency of the unit across the
climate zones tested.
6. Quantify water consumption of the unit across the climate zones tested.
7. Compare cooling capacity performance and energy efficiency to water consumption
across the climate zones tested.
In the interest of reducing peak summer electrical demand and bringing more energy
efficient technology into use through market penetration, it is expected that this technology
may be considered a candidate for rebates or other incentive programs. To facilitate this
expectation, secondary assessment objectives were pursued to provide information about
how this technology compares to existing air conditioning technology.
Secondary assessment objectives:
1. Compare the cooling capacity performance of the unit to previously tested air-cooled
condenser A/C units for different climate zones.
2. Compare energy efficiency of the unit to previously tested air-cooled condenser A/C
units for different climate zones.
In conclusion, this project provides the assessments that are necessary to confirm the
electrical energy savings and peak demand reduction potential to SCE as required for
reducing future utility system peak power demand and long-term energy consumption. It
also provides validation of manufacturer’s performance claims to help reduce market
barriers that will ultimately result in a more widespread use of this promising energy
efficient technology.
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PRODUCT EVALUATED
The product evaluated was a commercially available residential split-system air conditioner
with an evaporatively-cooled condenser. The unit tested was rated at a capacity of three
tons (36,000 Btu/hr) and used the R-410a blend of environmentally friendly refrigerant.
The indoor portion of the A/C system was a standard 3-ton ‘A’ coil evaporator controlled by
a thermal expansion valve. The indoor blower fan was integral to the standard gas pack
heating unit - no heating components were tested. The supply side of the indoor unit was
fitted with a sufficient length of flexible duct to create the static pressure across the
evaporator required for testing. The key importance here is that the indoor unit was the
same as used in the manufacturer’s tests where results were later used for comparison with
Technology Testing Centers (TTC) tests.
AIR COOLED VERSUS EVAPORATIVELY-COOLED CONDENSERS
Conventional air-cooled condensing systems use outdoor air to cool and condense
the refrigerant gas in the condensing unit of the air conditioning system. The
efficiency of an air-cooled system is related to the outdoor dry bulb air temperature
at the condensing unit, and the efficiency decreases as this outside air temperature
increases.
In the draw-through type of evaporatively-cooled condenser system, water is
sprayed down onto the refrigerant lines while the counter flowing condenser inlet air
passes up through the water covered tubes of refrigerant. Condensing of the
refrigerant occurs as a result of a two-stage process of sensible and then latent heat
transfer. First, the heat of the refrigerant is conducted to the surface of the tubes
containing it where a thin film of sprayed water resides. The water conducts heat
away from the tube surface which increases the caloric heat content of the water
film. As the continuous flow of water sprays down on the refrigerant lines, the
upward flowing condenser inlet air absorbs moisture from the water spray and
becomes cooler.
Then, in the second stage of the condensing process, the cooled inlet air continues to
absorb moisture as it evaporates the water film on the refrigerant tube surfaces. It
is this latent mass transfer of moisture into the air (evaporation) that results in
further cooling of the refrigerant lines as the latent heat of vaporization of the water
is given up to the air that carries away the absorbed mass of water. This process
also describes the water consumption as more water continually replenishes that
which is evaporated.
The refrigerant condensing temperature is thus limited by the ambient air water
content or wet bulb temperature (for any given dry bulb temperature), that is
normally 14°F to 25°F lower than the condensing temperature limitation of ambient
dry bulb for air cooled condensers (Reference 5). The lower condensing temperature
allows for more efficient compressor operation and increased capacity. This
characteristic promises a performance advantage over air-cooled systems in climate
zones with high outdoor temperatures.
Achieving the benefit of increased efficiency using the evaporative condenser technology
requires the availability of sufficient water to meet the consumption requirements of the
system. Water is consumed by evaporation in the cooling section of the condenser and by a
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periodic purging of the sump to avoid excessive mineral build-up in the cooling water. The
purge is scheduled to take place once per hour of run time and releases about two gallons
of water each time. The manufacturer claims the water consumption is offset, to a small
extent, by a reduction in water used during electricity generation due to the reduced
electricity use of the higher efficiency evaporative condenser unit. However, there is also a
slight offset in energy savings due to the pumping energy required to deliver water to the
evaporative-condenser unit.
Performance parameters including capacity, power demand, and EER were evaluated for
various climate zones within the SCE service territory. Water consumption associated with
the operation of the system for the various climate zones was also quantified.
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TECHNICAL APPROACH/TEST METHODOLOGY
TEST FACILITY
The TTC is a testing facility of approximately 7,600 square-feet located in SCE’s Customer
Technology Application Center (CTAC) complex in Irwindale, California. The TTC is
comprised of lighting, refrigeration, and air-conditioning test centers. Figure 1 is a
schematic depicting the layout of the Refrigeration and Thermal Test Center (RTTC) portion
of the TTC. Within the RTTC, there are several test chambers:

Supermarket test chamber

Refrigerated walk-in test chambers

HVAC test chambers
The supermarket test area is comprised of a controlled environment room equipped with an
independent dehumidification, humidification, heating, and cooling system. Three
refrigeration racks and a variety of heat rejection equipment can serve various display case
systems that are housed in a controlled environment room.
The refrigerated walk-in testing areas consist of a cooler, a freezer and a loading dock
chamber. A multiplex compressor rack system equipped with sophisticated controls and
advanced features, serves the fan coil systems of these three chambers.
FIGURE 1. ENTIRE TEST FACILITY LAYOUT
The HVAC testing area is comprised of both an indoor and an outdoor controlled
environment chamber. Both chambers are served by their own individual heating, cooling,
dehumidification, and humidification systems. Two ultrasonic humidifiers, controlled by the
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facility’s sophisticated CPU-based Energy Management System (EMS), Figure 2 (b), inject
precise moisture quantities into both chambers. A centrifugal humidification system
augments the capacity of the ultrasonic units. Each room is served by an Air Handler Unit
(AHU), equipped with split Direct Expansion (DX) coils, Variable Speed Drive (VSD)
controlled variable air volume fans, variable supply air temperature control, and an electric
heating system controlled by pulse modulation circulates air through supply and return air
plenums, shown in Figure 3. A multiplex rack system, shown in Figure 2 (a), consisting of
two 15 HP scroll compressors, equipped with VSDs, shown in Figure 4, and variable suction
control serves the two AHUs.
(a)
(b)
FIGURE 2. SCROLL COMPRESSOR RACK (A) AND EMS PANELS (B)
(a)
(b)
FIGURE 3. SUPPLY AIR PLENUM (A) AND RETURN AIR OPENING WITH RACK-MOUNTED FILTERS (B)
(a)
(b)
FIGURE 4. VARIABLE SPEED DRIVES OF AHUS SERVING INDOOR (A) AND OUTDOOR (B) CONTROLLED ENVIRONMENT ROOMS
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The unit to be tested was installed in the test rooms in accordance with the manufacturer’s
installation instructions. Air velocity in the vicinity of the installed unit was monitored
closely and maintained below 500 feet per minute (fpm). Figure 5 (a) shows the outdoor
ambient controlled environment room with the evaporative condensing unit. The outdoor
controlled environment room’s 10-foot high ceiling provided sufficient clearance (more than
six feet) from condenser discharge. A distance of at least three feet was provided between
the test room’s walls and the equipment side surfaces. Figure 5 (b) shows the indoor unit in
the indoor ambient controlled environment room with the supply duct connected from the
outlet of the airflow measurement station.
(a)
(b)
FIGURE 5. OUTDOOR (A) AND INDOOR (B) CONTROLLED ENVIRONMENT ROOMS WITH SUPPLY AND RETURN DUCTS
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TEST METHODOLOGY
TEST PROTOCOL
The test protocol used in this performance evaluation was ASRHAE 37-2005 Methods of
Testing for Rating Electrically Driven Unitary Air-Conditioning and Heat Pump Equipment
(Reference 6). Instrumentation procedures, test setup and measurement methodologies
were derived from this protocol that were then applied to the relevant calculations used.
The test method used in this evaluation was the 2003 ARI Standard 210/240 Unitary AirConditioning and Air-Source Heat Pump Equipment (Reference 3). Test procedures, indoor
and outdoor baseline climate conditions and tolerances were derived from this test method
and applied to the test procedure.
TEST METHOD
Determination of the capacity and performance characteristics of the unit under different
climate conditions followed methods of testing specified by ASHRAE Standard 37 -2005 and
AHRI 210/240-2003 (formerly ARI). All tests were performed at steady-state conditions for
a period of at least two hours. As prescribed by the test standards, air-side and refrigerantside measurements are used to determine performance, particularly cooling capacity and
Energy Efficiency Ratio (EER).
Typically, the performance of a unit under test is to be validated by both air-side and
refrigerant-side enthalpy analyses. These independent determinations provide confidence in
the test results under ASHRAE guidelines.
It should be noted however, that ASHRAE does not require two simultaneous test methods,
only that when two methods are used, the total cooling capacity shall be the evaporator side
capacity of two simultaneously conducted methods of test which shall agree to within 6% to
establish confidence in the results of the tests.
The AHRI rating standard requires the unit under test be operated at the baseline conditions
for a minimum of one hour of steady state performance to establish stability in data
measurements and assure consistency of performance results. During this one hour
minimum, the indoor climate conditions are to be maintained at 80°F DB and 67°F WB.
Table 1 identifies the conditions specified by AHRI 210/240 that were used in the testing of
this unit.
TABLE 1.
OPERATING CONDITIONS FOR STANDARD RATING USING AHRI 210/240
TEST
(for Cooling Capacity)
INDOOR UNIT
Air Entering Evaporator
DB (°F)/ WB (°F)
OUTDOOR UNIT
Air Entering Air Cooled
Condenser
DB (°F)/ WB (°F)
Standard Rating Conditions
“A” Cooling Steady State
80°F/67°F
95°F/75°F
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In addition to the baseline test condition requirements of the performance rating
methodology, SCE has interest in the performance variations of the test unit across different
climate conditions within SCE service territory. Climate zones are defined by ambient dry
bulb and wet bulb temperatures specified for each zone.
Table 2 shows design conditions for several different climate zones within SCE service
territory in which the performance of this test unit may be of interest. Representative cities
and a description of the climate conditions were taken from the ASHRAE reference: Climatic
Data For Region X (see Reference 4 of appendix). These values represent the 0.5% mean
temperature conditions as given by the ASHRAE Region X (10) weather data (Reference 4).
The capacity, power demand, EER, and water consumption of the unit were determined
through the TTC tests for baseline conditions as well as for climate zones 6, 7 1 , 8, 9, 10, 13,
14, and 15. An additional climate zone specified as Hot and Dry Air Conditioner (HDAC) was
included because it was previously used to identify an extreme hot, dry climate condition
that is somewhat more severe than Climate Zone 15. See Table 2.
TABLE 2.
DESIGN CONDITION CLIMATE ZONES WITH REPRESENTATIVE TEMPERATURES
Temperature
Climate Zone
WB (°F)
Representative City
Description
TTC Baseline
95
75
ASHRAE 37-2005
ASHRAE Baseline
CTZ 6
84
67
Los Angeles
South Coast
1
1
DB (°F)
1
CTZ 7
83
69
San Diego
South Coast
CTZ 8
89
69
El Toro
South Coast
CTZ 9
94
68
Pasadena
South Coast
CTZ 10
100
69
Riverside
South Coast
CTZ 13
101
71
Fresno
Central Valley
CTZ 14
108
69
China Lake
Desert
CTZ 15
111
73
El Centro
Desert
CTZ HDAC
115
74
(Similar to) Needles @ 114/72
Desert
Climate Zone 7 is not is SCE service territory.
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Figure 6 below shows the psychrometric plot of all the design condition climate zones tested
as well as the baseline and the HDAC conditions.
ASHRAE PSYCHROMETRIC CHART NO.1
55
NORMAL TEMPERATURE
BAROMETRIC PRESSURE: 29.921 INCHES OF MERCURY
Copyright 1992
R
60
R
90
AMERICAN SOCIETY OF HEATING, REFRIGERATING AND AIR-CONDITIONING ENGINEERS, INC.
.028
SEA LEVEL
60
85
.026
85
WE
TB
80
.024
UL
BT
EM
PE
RA
TU
55
RE
-
.022
°F
80
.020
75
50
.018
75
70
.016
Baseline
70
.014
HUMIDITY RATIO - POUNDS MOISTURE PER POUND DRY AIR
65
CTZ 7
%
90
%
CTZ 8
CTZ 13
60
CTZ 10
45
50
40
50
45
35
%
60
CTZ 14
%
DRY BULB TEMPERATURE - °F
55
50
40%
40
30%
35
20%
15
20
110
105
100
95
90
85
80
75
70
65
60
55
50
45
40
35
Y
VE HUMIDIT
10% RELATI
10
.012
.010
CTZ 15 CTZ HDAC
CTZ 9
%
70
40
.008
.006
35
.004
.002
115
80
55
CTZ 6
120
65
60
45
30
25
ENTHALPY - BTU PER POUND OF DRY AIR
FIGURE 6. PSYCHROMETRIC PLOT OF ALL DESIGN CONDITION CLIMATE ZONES TESTED
INSTRUMENTATION AND DATA ACQUISITION
This section addresses instrumentation requirements for temperature, pressure, electrical
equipment, refrigerant flow, and condensate collection. With the objective of minimizing
random and systematic uncertainties, careful attention was paid to the design of the data
acquisition system. With this in mind, the following steps were taken:

Use of sensors with very high accuracy

Minimization of random errors by use of multiple sensors

Use of calibration standard instruments of very high accuracy.
The test facility is equipped with a sophisticated computer-based data acquisition system.
The National Instruments SCXI high-performance signal conditioning and instrumentation
system for PC-based data acquisition and control was used to acquire and log test data. The
data acquisition system was programmed to process and average 100 reads from 110 data
channels every 20 seconds. This system was calibrated at the factory, and is traceable to
the National Institute of Standards and Technology’s (NIST) standards. The collected and
stored data for each sensor was then checked for consistency and accuracy at the end of
each test scenario. Consistently, the operating parameters were checked and deemed to be
within acceptable limits before the next run began.
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The Data Acquisition System consists of LabVIEW 8.6 software and SCXI hardware. The
software includes a graphical data acquisition environment, and a data logging and
supervisory control add-on, see Figure 7.
FIGURE 7. GRAPHICAL DATA ACQUISITION ENVIRONMENT
The hardware includes three high-performance signal conditioning and switching platforms
(SCXI-1000) with 4-slot chassis where each platform can house four modules, see Figure 8.
FIGURE 8. MODULES FOR SCXI HIGH PERFORMANCE SIGNAL CONDITIONING AND SWITCHING PLATFORM
There are eight 32-channel SCXI-1100 analog input modules, and two 16-channel SCXI1122 isolated sensor modules. In addition, a PCI-6052E 333kS/s, 16-bit, and 16-channel
analog input card was used to convert analog data to digital, see Figure 9.
FIGURE 9. PCI 6052E ANALOG INPUT CARD
The collected data points from the 20-second intervals were averaged into one-minute
intervals and used for further screening of the test data. The advantage of using one-minute
averages is that the data trends can still be displayed with an acceptable resolution while
enabling the engineering model to generate relevant calculated hourly results such as
cooling loads. The primary data points used for comparative analysis are based on
refrigerant enthalpy results.
After the data was compiled into one-minute averages within the engineering model, tabular
and graphical representations of various correlations and calculated parameters were
produced. Several graphs were created to initially screen the calculated data. Various critical
raw data was continuously screened for validation prior to importing the data into the TTC’s
engineering model. After careful examination and upon validation of the initial screening
plots, the informational plots were produced. This set of plots provided relationships
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between the calculated quantities. In cases where data flaws were detected, a series of
diagnostic investigations were carried out, and through this process, corrections were made
and tests were repeated.
Under steady-state test conditions, every 20 seconds, the data acquisition system sampled
and averaged 100-points per channel of scanned data, which was then saved to a file at the
end of each test. SCE engineers reviewed the initial data at the TTC to ensure that the
control parameters were within range. In the event that any of the control parameters fell
outside acceptable limits, the problem was flagged. In such cases, test runs were repeated
until the problem was corrected.
TEMPERATURE MEASUREMENTS
All temperature measurements followed ASHRAE Standard 41.1-1986.
THERMOCOUPLES
All thermocouples (TCs) were Type-T (copper-constantan) and had their junctions
secured with soft solder and electrically isolated with heat shrink tubing. Upon
application to the contact surface to be measured, the TCs were attached with
thermally conductive paste for optimum heat transfer. Each TC was individually
calibrated within 0.18oF accuracy, exceeding ASHRAE’s requirement of 0.2oF.
REFRIGERANT TEMPERATURE MEASUREMENTS
Refrigerant vapor and liquid line temperatures were determined by attaching
thermocouples on the suction, discharge, and liquid lines as described previously.
For stability of readings, an eight inch length of the TC wire was wrapped around the
location of pipe to be measured after the tip of the TC was properly affixed with
thermal paste. Three separate layers of insulation and heat shield material were
then wrapped around the TC element and pipe.
AIR TEMPERATURE & HUMIDITY MEASUREMENTS
Air temperature measurements were taken within the 2’x2’ air duct at four locations
throughout the air distribution ducts. For critical measurements, including air
entering and leaving the evaporator, a temperature grid and a dew point sampling
grid were constructed and mounted before and after each side of the ‘A’ shaped
evaporator. The air inlet and outlet temperature grids each consisted of four TCs.
The two air inlet dew point sampling grids were plumbed together and routed to a
DewPrime® dew point temperature measurement instrument. Likewise, the two air
outlet dew point sampling grids were plumbed and routed to a separate DewPrime
instrument. All probe connections were insulated to minimize air leakage.
Similarly, for the air temperature at the inlet of the condenser unit, a grid of six TCs
was installed on each of the three air inlet sides of the condenser. Each of these TC
channels was recorded separately and their average was calculated as the average
inlet air temperature to the condenser. A similar grid of 8 TCs was used at the outlet
of the condenser to record the condenser outlet air temperature.
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The inlet and outlet dew point temperatures of the condensing unit were each
measured using a grid of three sampling points for inlet and outlet dew point
temperature, respectively.
PRESSURE MEASUREMENTS
The accuracy of all the pressure measuring instruments was ± 0.13% of full scale.
The smallest scale division of the pressure sensors used for these tests never
exceeded 2.5 times their specified accuracy. All duct static pressures were measured
with manometers having an accuracy of + 1% full scale. In each designated
location, one side of the manometer was connected to four externally manifolded
pressure taps in the supply duct. The other side of the manometer was connected to
four externally manifolded pressure taps centered in each face of the return duct
before inlet to the unit, see Figure 10.
To determine the external static pressure, ASHRAE 37 requires that pressure taps be
centered in each discharge duct face at a distance of twice the square root of the
cross sectional dimension from the test unit’s outlet. Due to the space restrictions in
the test facility, taps were installed at a distance shorter than those specified by the
ASHRAE standard.
FIGURE 10. FOUR EXTERNALLY MANIFOLDED PRESSURE TAPS BEFORE INLET OF THE TEST UNIT
ELECTRICAL MEASUREMENTS
The total power supplied to the test unit was delivered through a STACO® voltage
stabilization device that provides consistent stable voltage at 208 volts AC with an
accuracy of ±1.2 volts.
To measure electrical component parameters, a power transducer was used to
sample current and voltage for each electrical component of the system. The
compressor, water circulation pump, condenser fan, water purge pump, indoor
blower fan, entire indoor unit, and entire condensing unit were each measured
separately. Each output of these transducers was recorded on a separate channel by
the data acquisition system. Additionally, the total power, voltage, current, phase
angle and frequency supplied to the unit were measured by an independent Nexus®
power analysis unit. The output data from the Nexus was also recorded by the data
acquisition system. The results of these two separate sets of data were compared to
establish confidence in the readings by ensuring the sum of the individual
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measurements was in close agreement with the totalizing unit. See Figure 11. All
tests were performed at 208 volts and 60 hertz.
FIGURE 11. INSTRUMENTS FOR MEASURING ELECTRICAL INPUT POWER TO FAN MOTORS, COMPRESSOR MOTOR,
AND THE TEST UNIT
CONDENSATE AND PURGE MEASUREMENTS
A special piping assembly was constructed to transfer condensate from the
evaporator pan to a separate container placed on a digital scale through a gravity fed
system, see Figure 12. Water removed from the condenser cooling water sump
during hourly purge cycles was also routed to the scale. Purge cycles were
recognized in the data as step increases on the graph of water accumulated over
time while the evaporator condensation rate was determined from the slope of the
line of the rate of water accumulation over time. The collective weight of the
condensate and the purge water accumulated was measured every 20 seconds by a
31,000 gram capacity digital scale with a standard deviation of 0.1 g, linearity of 
0.2 g, and minimum deviation of 0.01%. The scale system automatically purged
when necessary.
FIGURE 12. DIGITAL SCALE WITH CONDENSATE COLLECTION SYSTEM
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CONDENSER SUMP AND SPRAY WATER MEASUREMENTS
The average water temperature in the sump at the bottom of the condensing unit
was measured at the inlet to the circulation pump using three TCs. The data
recorded from each of these three channels was averaged together in the data
analysis.
The temperature of the water spray onto the refrigerant condensing coils was
measured using an array of eight TCs attached to the water distribution lines within
the body of the condensing unit. The data recorded from each of these eight
channels was also averaged together in the data analysis.
CONDENSER SUPPLY WATER MEASUREMENTS
Cooling water input to the condenser was provided through a constant temperature
and pressure supply line to the condenser water sump. Water was supplied to
replace evaporation and to replace water removed during the hourly purge of the
sump. The hourly purge is designed to prevent mineral buildup in the cooling water.
Water was supplied to the condenser sump by drawing, as needed, from a
continuously circulating constant temperature water supply loop where the
condenser inlet water was controlled to a constant 85oF by means of a heater and
chiller arrangement, Figure 13 (a). Condenser inlet water temperature and pressure
were also measured at the condenser inlet point of the circulation loop, Figure 13
(b). To measure total water consumption, a Coriolis mass flow meter was installed
on the supply side of the constant temperature circulation loop as show in Figure 13
(c).
(a)
(b)
FIGURE 13. HEATER/CHILLER LOOP (A) CONDENSER WATER INLET (B)
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(c)
AND WATER MASS FLOW METER (C)
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AIRFLOW MEASUREMENTS
To enable precise measurements of airflow rates for Heating, Ventilation, and Air
Conditioning (HVAC) equipment, TTC staff designed and supervised construction of
an ASHRAE airflow measurement station, as shown in Figure 14. The airflow station
was built according to ASHRAE Standard 41.2-1987 and contains four calibrated flow
nozzles with flow straighteners located both up- and down-stream of the nozzles.
This measurement station was located in the indoor room and incorporated a
supplemental blower fan with a variable frequency drive. Adjustments to this
variable frequency drive fan allowed for a wide range of air velocity and static
pressure conditions. Adjustments to the fan speed of the airflow station allowed
matching the static pressure and cubit feet per minute (cfm) across the evaporator
coil of the unit under test to the requirements of the ARI 210/240-2003 Standard.
FIGURE 14. ASHRAE AIR FLOW MEASUREMENT STATION
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SPECIFICATIONS OF SENSORS
The list of sensors used and their respective NIST traceable accuracies are shown in
Table 3, below.
TABLE 3.
SPECIFICATIONS OF SENSORS USED
SENSOR TYPE
MAKE/MODEL
ACCURACY [NIST
TRACEABLE]
Temperature (TC)
Type T in Teflon Jacket
Temperature (RTD)
Hy-Cal Engineering Model RTS-37-A100
±0.06°C (±0.108°F)
Dew Point
EDGETECH Model 2000 Dew Prime DF
Dew Point Hygrometer- S2 Sensor
±0.2°C (±0.36°F)
Relative Humidity
General Eastern Humiscan
Pressure (static)
MAMAC Systems Model PR-274/275
Pressure (barometric)
PTB 100A Barometric Pressure
Transducer
Pressure (differential)
Dresser Industries Inc. ASHCROFT
IxLdp
0.25% F.S.
Pressure
Setra Model 207, 100-500 psig
pressure ranges
±0.13%F.S.
Power
Ohio Semitronics Model GW-5
E.I.L. AC Watt Transducer
Electro Industries Nexus 1250
0.2% of reading
0.5% F.S.
0.06 of reading
Refrigerant Mass Flow
Endress & Hauser Promass 80 M15
Coriolis Mass Flow Measuring System
±0.25% (gas)
±0.05% (liquid)
Condenser Cooling
Water Mass Flow
Endress & Hauser Promass 80 F08
Coriolis Mass Flow Measuring System
±0.25% (gas)
±0.05% (liquid)
Velometer
TSI Inc. Model #8455 Air Velocity
Transducer
±0.5% F.S.
±2.0% of reading
Velometer
TSI Inc. Model #8475 Air Velocity
Transducer
±1.0% F.S.
±3.0% of reading
Scale
HP-30K
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±0.1°C (±0.18°F)
±1.0%RH (for 0.5-90%
RH range)
±1.0%F.S.
±0.3 hPa
±0.1 Gram
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DATA TO BE RECORDED
The minimum amount of data that is required by AHRI 210/240 to be recorded
during the cooling capacity test, is shown in Table 4.
TABLE 4.
MINIMUM REQUIREMENTS FOR DATA COLLECTION DURING COOLING CAPACITY TESTS AS SPECIFIED
BY AHRI
AIR ENTHALPY METHOD
REFRIGERANT ENTHALPY
METHOD
Barometric pressure
(in. Hg)
X
X
Power input to
equipment (W or Wh)
X
X
X
X
X
X
External resistance to
airflow (in. Hg)
X
-----
Fan speed, if
adjustable (rpm)
X
-----
Dry-bulb temp of air
entering equipment
(°F)
X
-----
Wet-bulb temp of air
entering equipment
(°F)
X
-----
Dry-bulb temp of air
leaving equipment
(°F)
X
-----
Wet-bulb temp of air
leaving equipment
(°F)
X
-----
Condensing pressure
or temp (psig/°F)
-----
X
Evaporator pressure
or temp (psig/°F)
-----
X
Refrigerant-oil flow
rate (ft3)
-----
X
Volume of refrigerant
in refrigerant-oil
mixture (ft3/ft3)
-----
X**
Refrigerant liquid
temp, indoor side
(°F)
-----
X
ITEM
Applied voltage(s)
Frequency (Hz)
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AIR ENTHALPY METHOD
REFRIGERANT ENTHALPY
METHOD
Refrigerant liquid
temp, outdoor side
(°F)- Required only
for line loss
adjustment
-----
X
Refrigerant vapor
temp, indoor side
(°F)
-----
X
Refrigerant vapor
temp, outdoor side
(°F)- Required only
for line loss
adjustment
-----
X
Refrigerant vapor
pressure, indoor side
(psig)
-----
X
ITEM
**Volume of refrigerant in refrigerant-oil mixture was not captured in this project
The AHRI Standard requires the cooling capacity tests to yield the following results:
 Total cooling capacity, Btu/hr
 Sensible cooling capacity, Btu/hr
 Latent cooling capacity, Btu/hr
 Indoor side airflow rate, CFM standard air
 External resistance to indoor airflow, in-wg
 Total power input to equipment or all equipment components, watts
 Air temperatures:

Outdoor and indoor dry-bulb

Outdoor and indoor wet-bulb
 EER, Btu/hr/watts
Test results generated under this project, however, provided information beyond the
AHRI requirements. The following lists additional results that were obtained.
 Coil sensible heat ratio
 Supply CFM/ton
 Coil superheat
 Condenser and total system sub-cooling
 Supply air temperature profile
 Refrigeration effect
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 Condenser coil’s arithmetic mean temperature difference
 Evaporator and condenser coils heat transfer coefficient (UA)
 System power factor
 Work of compression
 Heat of compression
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EVALUATIONS
APPROACH
The cooling capacity of the air-conditioning equipment was determined based on the
ASHRAE 37 -2005 Standard. According to this standard, the cooling capacity can be
determined based on indoor air-enthalpy and refrigerant enthalpy methods. The total
cooling capacity of two simultaneously conducted methods of tests should agree
within 6.0%. In addition to cooling capacity of the unit under consideration, the
condenser performance, system EER, and water consumption were also evaluated.
The thermodynamic properties of air and airflow inside ducts were determined
according to the ASHRAE Handbook of Fundamentals. Thermal Analysis Partners,
LLC’s refrigerant property program, xProps® version 1.3, was used to determine
refrigerant properties. The saturation temperatures were also determined using the
xProps software.
COOLING CAPACITY
The gross cooling capacity is the rate of cooling or heat removal (in Btu/hr) that
takes place at the evaporator coil of the unit. Cooling capacity was determined based
on two methods; the air-enthalpy and refrigerant-enthalpy methods. In the airenthalpy method, cooling capacity was determined based on properties of entering
and leaving air, and the associated air mass flow rate. The cooling capacity using the
refrigerant-enthalpy method was determined based on mass flow rate of refrigerant,
and the refrigerant properties at the inlet and outlet of the evaporator coil.
According to AHRI 210/240, the cooling capacity rating should include the effects of
blower fan heat, but not include supplementary heat.
AIR-ENTHALPY METHOD
The air-enthalpy method employs the measured psychrometric properties of air
flowing across the evaporator coil of the unit. These measurements included dry-bulb
and dew-point temperatures of the air up-stream and down-stream of the coil,
volumetric airflow rates through the coil as well as static pressure drop across the
coil. The enthalpy and mass flow rates of the air were then determined through
calculation using these measured values. The resulting gross and net cooling
capacities were then determined.
Prior to determining the cooling capacity of the test unit, the volumetric airflow in the
duct was determined based on velocity pressure readings and air density. The
equation for determining volumetric airflow from measured velocity pressure and
obtained air density is shown in Equation 1.
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EQUATION 1. VOLUMETRIC AIRFLOW RATE

cfm  C *

2 * P v * gc 
 *A
ρ

where
cfm
= volumetric airflow rate, ft3/min
C
= unit conversion factor, (136.8)
Pv
= velocity pressure, in-wg
gc
= gravitational constant, (32.174 lbm-ft/lbf-s2)

= density of air, lb/ft3
A
= duct cross sectional area, ft2
The standard volumetric airflow rate was determined based on measured and
standard air specific volume, which corresponds to about 60oF at saturation and 69oF
dry air at 14.7 psia. The obtained volumetric airflow rate and the ratio of the
measured and standard specific volume of air were used to obtain the standard
volumetric airflow rate, Equation 2:
EQUATION 2. STANDARD VOLUMETRIC AIRFLOW RATE
 υstd 

 υair 
SCFM  cfm * 
where
SCFM
= standard volumetric airflow rate, ft3/min
std
= specific volume of air at standard conditions, (13.33 ft3/lb)
air
= measured specific volume of air, ft3/lb
After the standard volumetric airflow rate is determined, the air mass flow rate is
determined by simply multiplying the volumetric airflow rate by the air density using
Equation 3:
EQUATION 3. AIR MASS FLOW RATE

m air  SCFM * ρ * k
where

m air
= airflow rate inside the duct, lb/hr
k
= conversion factor, (60 min/hr)
SCFM
= standard volumetric airflow rate, ft3/min

= density of air, lb/ft3
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The measured air properties, specifically dew point and dry-bulb temperatures, were
used to determine air enthalpies at the inlet and outlet of the evaporator coil, based
on the ASHRAE Handbook of Fundamentals. After the air enthalpies and airflow rate
are determined, Equation 4 is used to obtain the gross cooling capacity.
EQUATION 4. GROSS COOLING CAPACITY- AIR-ENTHALPY


Q  m air * hair - in - hair - out 
where

Q
= gross cooling capacity, Btu/hr
hair - in
= entering air enthalpy, Btu/lb
hair - out
= leaving air enthalpy, Btu/lb
It is sometimes useful to determine the cooling capacity in tons. Thus, the cooling
capacity of the unitary air conditioning equipment was divided by 12,000, a
conversion factor for Btu/hr to tons. See Equation 5.
EQUATION 5. CONVERTING BTU/HR TO TONS


Q (tons) 
Q
12,000
where

Q (tons)
= gross cooling capacity of air, tons
The net cooling capacity can be determined by simply subtracting the heat gain due
to the evaporator fan from the gross cooling capacity, Equation 6. This methodology
was used to exclude the heat input from the evaporator fan motor.
EQUATION 6. NET COOLING CAPACITY


Q net  Q  kWevap - fan * k 
where

Q net
= net cooling capacity of air, Btu/hr
kWevap - fan
= evaporator fan power, kW
k
= conversion factor, (3,413 Btu/hr/kW)
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Two methodologies were used to determine the latent indoor cooling capacity,
psychrometric data and values taken directly from the condensate scale reading. The
mass of collected condensate was determined psychrometrically by using airflow rate
and absolute humidity of air at the inlet and outlet of the evaporator coil, Equation 7.
EQUATION 7. MASS OF COLLECTED CONDENSATE- PSYCHROMETRIC


m wp  m air * ωair - in - ωair - out 
where

m wp
= mass of collected condensate based on psychrometric data, lb/hr
ωair - in
= absolute humidity of air at the evaporator inlet, lbw/lba
ωair - out
= absolute humidity of air at the evaporator outlet, lbw/lba
After the condensate weight was determined, the latent cooling capacity was
obtained by simply multiplying the condensate mass by the heat of vaporization of
water. Equation 8 and Equation 9 show the latent cooling capacity calculations using
psychrometric and scale reading methodologies, respectively.
EQUATION 8. LATENT COOLING CAPACITY- PSYCHROMETRIC


Q lp  m wp * hfg
where

Q lp
= latent indoor cooling capacity using psychrometric data, Btu/hr
hfg
= heat of vaporization of water, (1,060 Btu/lb)

= mass of collected condensate based on psychrometric data, lb/hr
m wp
EQUATION 9. LATENT COOLING CAPACITY- CONDENSATE


Q ls  m ws * hfg
where

Q ls

= latent indoor cooling capacity based on scale reading, Btu/hr
m ws
= mass of collected condensate from scale reading, lb/hr
hfg
= heat of vaporization of water, (1,060 Btu/lb)
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After the gross cooling capacity and latent cooling capacity were determined, the
sensible cooling capacity was obtained by using Equation 10.
EQUATION 10.


SENSIBLE COOLING CAPACITY

Q s  Q - Q ls
where

Q
= sensible indoor cooling capacity, Btu/hr
s
The standard volumetric airflow rate per gross cooling capacity of the airconditioning equipment was obtained using Equation 11.
EQUATION 11.
SCFM
Ton

VOLUMETRIC FLOW RATE PER GROSS COOLING CAPACITY
SCFM

Q (tons)
where
SCFM
= Standard volumetric airflow rate per ton of cooling capacity,
Ton
ft3/min/tons
The EER of the unit depends on the total power usage, as well as the net cooling
capacity of the unit. The total power usage includes compressor, condenser fan, and
evaporator fan. The EER of the unit can be determined by simply dividing the net
cooling capacity by the measured total input power to the unit, Equation 12.
EQUATION 12.
EER – ENERGY EFFICIENCY RATIO

EER 
Q net
Wtotal
where
EER
= energy efficiency ratio of the unit, Btu/hr/watts
Wtotal
= measured total input power to the unit, watts
The sensible heat ratio (SHR) was determined using Equation 13. It is used to
compare the amount of sensible cooling being done on the air to the gross cooling
capacity, which includes the amount of moisture removal.
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Performance Evaluation of an Evaporatively-Cooled Split-System Air Conditioner
EQUATION 13.
ET 08.08
SENSIBLE HEAT RATIO

SHR 
Qs

Q
where
SHR
= sensible heat ratio, unit-less
REFRIGERANT-ENTHALPY METHOD
Pressures and temperatures were measured across all components of the
refrigeration system to verify the enthalpies at all points in the refrigeration cycle.
These measurements were taken in the refrigerant lines approximately 10 inches
from the relevant components such as compressor, condenser coil, and evaporator.
A Coriolis mass flow meter was installed in the liquid line at the evaporator unit,
Figure 15. This assembly was positioned upstream of the refrigerant metering
device. Pressure drop across the flow meter was closely monitored so that liquid
refrigerant did not flash and undergo a temperature drop larger than 3°F.
FIGURE 15. CORIOLIS REFRIGERANT MASS FLOW METER
Additionally, temperature and pressure sensors as well as a sight glass were installed
immediately downstream of the mass flow meter. This provided confirmation that
the total sub-cooling of liquid refrigerant had not exceeded 3°F (exiting the flow
meter) and that no vapor bubbles passed through the flow meter.
The refrigerant enthalpy method is generally an easier and more reliable method for
determining cooling capacity than the air-enthalpy method as there are fewer
properties to be measured and fewer components in the cooling capacity equation.
Pressures, temperatures, and the mass flow of refrigerant can be measured directly
and the behavior of refrigerant is generally more consistent and stable over time
compared to properties of air. As an example, opening the door to the indoor test
chamber for a short period during a test would affect the air-side properties much
more than the refrigerant-side properties.
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Performance Evaluation of an Evaporatively-Cooled Split-System Air Conditioner
ET 08.08
The refrigerant enthalpy method requires refrigerant temperature and pressure
measurements entering and leaving the evaporator coil as well as the mass flow rate
of the refrigerant through the system. A change in enthalpy across the evaporator is
determined from the refrigerant pressure and temperature measurements at the
inlet to the evaporator (before the expansion device) minus the corresponding
enthalpy due to the pressure and temperature of the refrigerant exiting the
evaporator. This difference is referred to as the refrigeration effect (Equation 14
below) and is the quantity of heat that each unit of mass of refrigerant absorbs to
cool the indoor space. It simply represents the capacity of the evaporator per pound
of refrigerant. The xProps program was used to determine refrigerant vapor and
liquid refrigerant enthalpies from the temperature and pressure readings.
EQUATION 14.
REFRIGERATION EFFECT
RE  hrefrig  out  hrefrig  in
where
RE
= refrigeration effect, Btu/lb
hrefrig  out
= superheated refrigerant enthalpy at the evaporator outlet, Btu/lb
hrefrig  in
= subcooled liquid refrigerant enthalpy at the expansion device inlet,
Btu/lb
The gross cooling capacity of the unit was determined simply by multiplying the
mass flow rate of refrigerant by the refrigeration effect, Equation 15.
EQUATION 15.

GROSS COOLING CAPACITY- REFRIGERANT-ENTHALPY

Q  m refrig * RE
where

Q

= gross cooling capacity of refrigerant, Btu/hr
m refrig
= mass flow rate of refrigerant, lb/hr
RE
= refrigeration effect, Btu/lb
As was the case with the air-enthalpy gross cooling capacity, Equation 5 is used to
convert the refrigerant-enthalpy gross cooling capacity to tons of cooling.
The net cooling capacity can be determined by subtracting the heat gain in Btu/hr
due to the evaporator fan motor from the gross cooling capacity, as shown in
Equation 6.
Equation 9 was used to determine the latent cooling capacity. With the latent and
net cooling capacities calculated the sensible cooling capacity can be calculated
(Equation 10).
The standard volumetric airflow rate per gross cooling capacity of refrigerant was
obtained using Equation 11.
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Performance Evaluation of an Evaporatively-Cooled Split-System Air Conditioner
ET 08.08
The same equations were used to calculate the EER and SHR using the refrigerantenthalpy data as with the air-enthalpy. These are Equation 12 and Equation 13,
respectively.
The compressor efficiency was determined based on the gross cooling capacity and
the compressor input power, Equation 16. The compressor efficiency represents the
gross cooling capacity per power input to the compressor.
EQUATION 16.
COMPRESSOR EFFICIENCY

Compeff 
Q
Wcomp
where
Compeff
= compressor efficiency in terms of cooling capacity per input power,
Btu/hr/watts
Wcomp
= measured total input power to the compressor, watts
Additional supplemental performance evaluation information pertaining to evaporator
coil characteristics, condenser coil characteristics and related parameters is
presented in Appendix A.
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Performance Evaluation of an Evaporatively-Cooled Split-System Air Conditioner
ET 08.08
RESULTS
CAPACITY AND PERFORMANCE OF THE UNIT
48,000
16.0
42,000
14.0
36,000
12.0
30,000
10.0
24,000
8.0
18,000
6.0
12,000
4.0
6,000
2.0
0
0.0
Manufacturer's Tests
TTC Tests
Demand (kW) & EER (Btu/hr/Watt)
Net Cooling Capacity (Btu/hr)
The first test at the TTC of the evaporative condensing A/C unit system was
conducted to verify the manufacturer’s stated performance of the air conditioner and
to develop a baseline for comparing the performance of the unit in different SCE
climate zones. The TTC tests showed nearly identical capacity to the manufacturer’s
test data but at a 13% greater power demand resulting in a 12% reduction in EER
compared to the manufacturer’s test data, shown in Figure 16.
AHRI 210/240 Baseline Conditions (95°F DB, 75°F WB)
Net Capacity (Btu/hr)
Power (kW)
EER (Btu/hr/Watt)
FIGURE 16. PERFORMANCE COMPARISON OF MANUFACTURER'S TEST DATA TO TTC’S TEST DATA AT AHRI
210/240 BASELINE CONDITIONS
One of the reasons for the difference in power is that the manufacturer’s tests did
not actually measure indoor blower fan wattage but used the ASHRAE indoor blower
fan default value of 438 watts for their calculations. The SCE tests however, used
the actual measured values of indoor blower fan wattage in determining total unit
power. These measured values were consistently higher than the ASHRAE default
value but the difference does not entirely account for the difference in power
between the two independent tests.
The methodology of each party’s tests was based on the AHRI 210/240 baseline
conditions of 95°F dry bulb and 75°F wet bulb using the refrigerant enthalpy method
for determining performance capacity of the unit. This is the more reliable and
preferred method of performance verification as there are fewer measured data
values and thus fewer possibilities for introducing error to the results compared with
the air enthalpy method. Additionally, attaining closely matched results using the
same methodology in analyzing the independently conducted baseline tests provides
increased confidence in the results.
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Performance Evaluation of an Evaporatively-Cooled Split-System Air Conditioner
ET 08.08
The TTC baseline and manufacturer’s baseline tests were determined to be in
agreement based on the refrigerant enthalpy test methodology used and the
relatively close values attained by each testing group. Therefore, for the purpose of
comparison, the subsequent tests across different climate zones were conducted.
As mentioned previously in the Test Methodology section, ASHRAE requires that
when two simultaneous test methods are employed that the evaporator side cooling
capacity of the two simultaneous methods match within 6%. However, the airenthalpy and refrigerant enthalpy methods for the TTC tests did not match to within
6.0% as required when two simultaneous methods are used. The reason for the
greater than 6.0% difference between air- and refrigerant enthalpy method results
was suspected to be related to errors in the measurement instruments or incorrect
calibrations of the air-side instruments.
Investigation of these presumptions later revealed a problem with the cold-junction
calibrations of the TCs in the old data acquisition chassis compared to those of the
new chassis added shortly before the beginning of this project. However, the close
agreement of the independently conducted baseline tests at identical test conditions
provides confidence in the TTC varying climate zone test results based on the
analysis using only the refrigerant-enthalpy method.
48,000
16.0
42,000
14.0
36,000
12.0
30,000
10.0
24,000
8.0
18,000
6.0
12,000
4.0
6,000
2.0
0
Demand (kW) & EER (Btu/hr/Watt)
Net Cooling Capacity (Btu/hr)
The results of the different climate zone tests are shown in Figure 17 where each
climate zone is specified by a dry bulb (DB) temperature and a wet bulb (WB)
temperature to define a state point for that CTZ condition. The Capacity, Power, and
EER were fairly uniform over Climate Zones 6, 7, 8, 9, and 10, with small declines
(up to a maximum of 10%) in EER for Climate Zones 13, 14, 15, and the HDAC
condition. Capacity was maintained within 2% of the CTZ 6 value for Climate Zones
6, 7, 8, 9, and 10, with a maximum reduction of 5.7% for the HDAC condition.
0.0
Baseline CTZ 6
(95 DB,
(84 DB,
75 WB) 67 WB)
CTZ 7
CTZ 8
CTZ 9
(83 DB,
(89 DB,
(94 DB,
69 WB)
69 WB)
68 WB)
CTZ 10
CTZ 13
CTZ 14
CTZ 15
HDAC
(100 DB, (101 DB, (108 DB, (111 DB, (115 DB,
69 WB)
71 WB)
69 WB)
73 WB) 74 WB)
Climate Zones (CTZ) in DB°F and WB°F
Net Capacity (Btu/hr)
Power (kW)
EER (Btu/hr/Watt)
FIGURE 17. TEST RESULTS OF EVAPORATIVE CONDENSER A/C UNIT IN VARIOUS OUTDOOR CLIMATE ZONES
Southern California Edison
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Performance Evaluation of an Evaporatively-Cooled Split-System Air Conditioner
ET 08.08
These results show that the test unit experienced a relatively small degradation in
performance in the hotter climate zones. For the HDAC zone, the hottest condition
evaluated, the unit is able to deliver 94.3% of capacity and 90% of the EER it
achieved in Climate Zone 6, one of the cooler zones evaluated.
A summary table of the performance parameters across the different climate zones
tested is listed in Appendix B at the end of this report. To put these results in
perspective, they were compared to similar tests performed at the TTC in 2004 on
air conditioners with conventional air-cooled condensers. This comparison is
developed in a subsequent section of this report.
WATER CONSUMPTION
Total water consumption of the evaporatively-cooled condensing unit was due to the
amount of water consumed through evaporation and the amount of water purged
during the purge cycles. Each of these is explained in more detail below.
PURGE
Generally, supply water to an evaporative cooler inherently contains dissolved
calcium, lime and other minerals. During the evaporation process, only the water
component evaporates leaving an increased concentration of minerals in the sump
water as the supply continuously replenishes the water that is evaporated. The
purge activity is a method of eliminating the water that becomes concentrated with
minerals that might otherwise build up as scale and subsequently reduce the
effectiveness of heat transfer surfaces.
The interval of the ECCU purge cycle may be manually adjusted inside the unit to
accommodate different water qualities depending on the quality of the feed water for
any particular installation. The purge interval for all tests conducted in this
evaluation was set to one hour. When the water quality contains a high
concentration of dissolved minerals, it may be desirable to purge the water more
frequently to minimize scale buildup on internal heat transfer surfaces and pump
components.
The amount of water consumed by the purge activity is independent of climate zone
test conditions. The measured purge amount is a function of the purge frequency
interval (one hour), the pump on-time duration and the purge pump flow rate. The
amount of water purged at each purge interval is relatively constant at
approximately 1.8 gallons per one hour flush interval (see Figure 18 below).
Climate zone 7 test conditions had the lowest purge water amount at 1.6 gal/hr
while the highest purge water amounts were at CTZ 10 and CTZ 13 each with 2.1
gal/hr. This small variation from lowest to highest of approximately 0.5 gal/hr is
likely related to variations in purge pump operating parameters such as flow rate or
time to establish flow, such as in priming the pump.
EVAPORATION
The remainder and larger component of the water consumption is due to evaporation
during the cooling process and is directly related to the climate zone conditions as
explained below.
Water consumption of the ECCU due to evaporation tends to increase with hotter and
dryer climate zones. The increase is approximately 27% from the lowest usage of 5
Southern California Edison
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Performance Evaluation of an Evaporatively-Cooled Split-System Air Conditioner
ET 08.08
gal/hr for the mild condition of climate zone 7 to the highest usage of almost 7 gal/hr
for the extreme conditions of climate zones 14, 15, and the HDAC condition (see
Figure 18).
Total water usage for evaporation and the hourly purge of water in the sump over all
climate zones was in the 6.5 to 8.5 gal/hr range during continuous operation. This is
equivalent to about 2.1 to 2.8 gal/hr per ton of air conditioning capacity for this 3ton unit.
9.00
8.00
Water Usage (gal/hr)
7.00
6.00
5.00
4.00
3.00
2.00
1.00
0.00
HDAC
CTZ15
CTZ14
CTZ13
Baseline CTZ6 CTZ7 CTZ8 CTZ9 CTZ10
(95 DB, (84 DB, (83 DB, (89 DB, (94 DB, (100 DB, (101 DB, (108 DB, (111 DB, (115 DB, 75 WB) 67 WB) 69 WB) 69 WB) 68 WB) 69 WB) 71 WB) 69 WB) 73 WB) 74 WB)
Climate Zones (CTZ) in DB°F and WB°F
Water Cons umed by Purge
Water Cons umed by Evaporati on
FIGURE 18. WATER CONSUMED BY PURGE AND EVAPORATION AT VARIOUS CLIMATE ZONE CONDITIONS
The amount of water consumed due to evaporation in the process of cooling the
refrigerant is related to the climate zone conditions in which the unit operates.
Specifically, the hotter and dryer the ECCU inlet air, the more water is consumed by
the evaporative cooling process. Any given climate zone condition can be specified
by a dry bulb (DB) temperature and a wet bulb (WB) temperature to define a state
point for that CTZ condition. The water content of such a state point is referenced
by the humidity ratio (ω).
Figure 19 shows a psychrometric chart plot of the ECCU inlet and outlet air state
points under the relatively mild conditions of climate zone 7. Though this particular
climate condition is not in SCE service territory, it is useful to compare with more
extreme climate zone conditions. The vertical graduations are dry bulb
temperatures, the horizontal graduations are humidity ratios, and the diagonal lines
represent the enthalpy or total heat content of the air at any given state point. The
enthalpy at any given state point is described in the ASHRAE Fundamentals
Handbook by the following relationship:
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Performance Evaluation of an Evaporatively-Cooled Split-System Air Conditioner
EQUATION 17.
ET 08.08
MOIST AIR SPECIFIC ENTHALPY
h  0.240T   (1061  0.444T) Btu/lb
As the condenser inlet air passes through the ECCU, it absorbs moisture in the
evaporation process until the air becomes saturated with water vapor. This is shown
in Figure 19 as the Condenser Outlet state point on the saturation curve of the
psychrometric chart plot.
The difference or change in humidity ratio (Δω) of the condenser outlet to condenser
inlet air state points represents the amount of water absorbed by the air as it passes
through the condensing unit (again, Figure 19). This is the water consumption due
to evaporation. It is interesting to note that for this climate zone, the Δω is nearly
proportional to the change in enthalpy (Δh) or heat content of the air after passing
through the ECCU. The increase in the air enthalpy comes from the heat of the
refrigerant as it is transferred to the water film surface through conduction and then
to the air through the latent process of evaporation.
Figure 19 shows a small drop (~1.5°F) in dry bulb temperature along the x-axis and
an increase in humidity ratio along the right y-axis as the state point of the air
moves from the Condenser Inlet condition to the Condenser Outlet condition.
ASHRAE PSYCHROMETRIC CHART NO.1
55
NORMAL TEMPERATURE
BAROMETRIC PRESSURE: 29.921 INCHES OF MERCURY
R
60
R
Copyright 1992
90
AMERICAN SOCIETY OF HEATING, REFRIGERATING AND AIR-CONDITIONING ENGINEERS, INC.
.028
SEA LEVEL
60
85
.026
CTZ 7 Condenser Outlet
Δh CTZ 7
85
WE
TB
80
.024
UL
B
TE
MP
ER
A
55
TU
RE
- °F
.022
80
.020
75
50
.018
ΔωCTZ 7
75
70
.016
70
.014
HUMIDITY RATIO - POUNDS MOISTURE PER POUND DRY AIR
65
CTZ 7 Condenser Inlet
65
%
90
%
80
70
55
50
50
40
%
50
45
35
%
60
DRY BULB TEMPERATURE - °F
45
60
%
4 0%
40
30%
35
20%
10
15
20
110
105
100
95
90
85
80
75
70
65
60
55
50
45
40
35
E HUMIDITY
10% RELATIV
.010
40
.008
.006
35
.004
.002
115
55
.012
120
60
45
30
25
ENTHALPY - BTU PER POUND OF DRY AIR
FIGURE 19. PSYCHROMETRIC CHART SHOWING Δω VERSUS Δh FOR
THE ECCU IN CLIMATE ZONE 7
Figure 20, below, shows a psychrometric chart plot of the inlet and outlet air to the
condenser under the relatively extreme conditions of the hot and dry HDAC climate
zone. In this plot we see a large increase in the humidity ratio (Δω is large) as the
Southern California Edison
Design & Engineering Services
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Performance Evaluation of an Evaporatively-Cooled Split-System Air Conditioner
ET 08.08
condenser inlet air has a much greater ability to absorb moisture compared to the
milder CTZ 7 condenser inlet air conditions. In this extreme climate condition, the xaxis of Figure 20 also shows a large drop in the dry bulb temperature of the
condenser outlet air. This indicates that some of the water consumed during
evaporation in this extreme climate condition is inadvertently used to cool the
condenser outlet air in addition to that used in cooling the refrigerant. This is
inherent to the use of the evaporation process in a hot and dry climate zone.
The greater Δω of the extreme climate zone conditions results in a higher water
consumption at the higher dry bulb temperature CTZs. It is interesting to note that
the overall change in the enthalpy (Δh) of the air is approximately the same as that
for the mild climate zone condition, CTZ 7 (ΔhCTZ 7 = 12.3 versus ΔhCTZ HDAC = 11.9).
This is due to evaporation of the water contributing to the large drop in air dry bulb
temperature at the discharge of the ECCU during operation in hot and dry climate
zone conditions.
ASHRAE PSYCHROMETRIC CHART NO.1
55
NORMAL TEMPERATURE
BAROMETRIC PRESSURE: 29.921 INCHES OF MERCURY
R
60
R
Copyright 1992
90
AMERICAN SOCIETY OF HEATING, REFRIGERATING AND AIR-CONDITIONING ENGINEERS, INC.
.028
SEA LEVEL
85
60
CTZ HDAC Condenser Outlet
.026
Δh CTZ HDAC
85
WE
TB
80
.024
UL
B
TE
MP
ER
AT
UR
E
55
- °F
.022
80
.020
75
50
.018
ΔωCTZ HDAC
75
70
.016
70
%
90
%
80
50
40
35
%
40
30%
20 %
15
20
110
105
100
95
90
85
80
Y
75
70
65
60
55
50
45
40
35
E HUMIDIT
10% RELATIV
10
40
.008
Condenser Inlet
4 0%
35
.010
CTZ HDAC
%
60
50
45
.012
.006
35
.004
.002
120
55
50
60
%
115
70
DRY BULB TEMPERATURE - °F
55
45
HUMIDITY RATIO - POUNDS MOISTURE PER POUND DRY AIR
65
60
45
.014
65
30
25
ENTHALPY - BTU PER POUND OF DRY AIR
FIGURE 20. PSYCHROMETRIC CHART SHOWING Δω VERSUS Δh FOR
THE ECCU IN CLIMATE ZONE HDAC
Figure 21, below, shows the increasing total water consumption rate as a function of
the increasing change in humidity ratio (Δω) for the outdoor climate zone conditions.
Specifically, this change in humidity ratio (Δω) is the difference between what the
humidity ratio is at saturation and the actual humidity ratio for the specific dry bulb
of each climate zone condition. Thus, the dryer the entering condenser air, the
greater is the Δω or the ability of that air to absorb moisture, therefore the higher
the water consumption due to evaporation. The data shows that CTZ 7 has the
lowest evaporation rate that corresponds to the smallest change in humidity ratio
(Δω) or ability to absorb moisture. Likewise, the CTZ HDAC conditions have the
highest ability to absorb moisture (Δω) and the greatest water consumption due to
evaporation.
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Performance Evaluation of an Evaporatively-Cooled Split-System Air Conditioner
ET 08.08
Figure 21 charts the total water consumption rate. The rate due to evaporation only
is the difference between the total water consumption and the purge water
consumption, or approximately 2 gal/hr/ton less at each data point.
9
Water Consumption Rate (gal/hr)
8.5
CTZ 14
8
CTZ 15
CTZ HDAC
CTZ 13
Baseline
7.5
CTZ 9
CTZ 10
CTZ 6
7
6.5
CTZ 8
CTZ 7
6
y = 37.618x + 6.3742
2
R = 0.9509
5.5
5
0.01
0.02
0.03
0.04
0.05
0.06
Δω
Water Consumption Rate (gal/hr)
Linear (Water Consumption Rate (gal/hr))
FIGURE 21. TOTAL WATER CONSUMPTION RATE VERSUS CHANGE IN HUMIDITY RATIO OF OUTDOOR CONDITIONS
As relative humidity is a more general reference to the moisture content of the air
for any given dry bulb temperature, we would expect to see a correlation similar to
that of the change in humidity ratio (Δω). Figure 22 shows reference to the more
common metric of relative humidity. Though the correlation is not as precise as to
the change in humidity ratio (Δω), it may be easier to use as a reference. This graph
shows total water consumed in gallons per hour and references Table 4 of dry bulb
and wet bulb conditions for the different climate zones tested.
Water consumption in hot and dry areas can become an issue, but the test data
provides an accurate basis for estimating water consumption needs in various
climate zones corresponding to future installation projections for the units. The
trend line of Figure 22 indicates an increase of approximately 0.5 gal/hr total water
consumption per ton of net cooling capacity for each 10% decrease in relative
humidity.
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Performance Evaluation of an Evaporatively-Cooled Split-System Air Conditioner
ET 08.08
10
9.5
9
CTZ HDAC
CTZ 15
Water Usage (gal/hr)
8.5
CTZ 14
CTZ 13
8
CTZ 10
Baseline
7.5
CTZ 9
CTZ 6
CTZ 8
7
6.5
y = ‐0.0468x + 8.9062
CTZ 7
2
R = 0.9208
6
A 10% decrease in RH increases water use by 0.5 gal/hr.
5.5
5
50
45
40
35
30
25
20
15
10
Relative Humidity, (%)
Tota l Water Us e
Li nea r (Total Water Us e)
FIGURE 22. TOTAL WATER CONSUMPTION AS A FUNCTION OF RELATIVE HUMIDITY
PERFORMANCE AND WATER CONSUMPTION
Figure 23 shows a comparison of net cooling capacity, EER and water consumption
across the different climate zone conditions tested. Notice the general trend of
increased water consumption as the hot and dry climate conditions become more
extreme. Also, the increasing hot and dry conditions reflect a slight but general
trend in decreased cooling capacity performance.
Southern California Edison
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November 2009
ET 08.08
40,000
16.0
35,000
14.0
30,000
12.0
25,000
10.0
20,000
8.0
15,000
6.0
10,000
4.0
5,000
2.0
0
EER (Btu/hr/watt) & Water Consumed (gal/hr)
Net Capacity (Btu/hr)
Performance Evaluation of an Evaporatively-Cooled Split-System Air Conditioner
0.0
CTZ 7
CTZ 6
CTZ 8
CTZ 9
Baseline CTZ 10 CTZ 13 CTZ 14 CTZ 15
HDAC
(83 DB, (84 DB, (89 DB, (94 DB, (95 DB, (100 DB, (101 DB, (108 DB, (111 DB, (115 DB,
69 WB) 67 WB) 69 WB) 68 WB) 75 WB) 69 WB) 71 WB) 69 WB) 73 WB) 74 WB)
Climate Zones (CTZ) in DB°F and WB°F
Net Capacity
EER
Water Consumed
FIGURE 23. NET CAPACITY, EER AND WATER CONSUMPTION AS A FUNCTION OF INCREASING DRY BULB ACROSS
CLIMATE ZONES
Overall, it can be concluded that the evaporatively-cooled condenser system offers
the potential of more consistent performance in hot climate zones compared to
conventional air-cooled condenser systems that experience substantial performance
and efficiency degradation when outdoor temperatures are high.
PERFORMANCE COMPARISON TO AIR-COOLED AIR
CONDITIONING SYSTEMS
In 2004, tests were conducted at the TTC to evaluate the performance of six
conventional air cooled air conditioning systems at high ambient temperatures
(Reference 1). The 5-ton rooftop package units evaluated included standard and
high-efficiency models from three different manufacturers. The performance of each
unit was evaluated by conducting controlled environment tests of varying outdoor
dry bulb while maintaining constant indoor conditions (temperature and humidity).
The outdoor dry bulb temperature was incrementally increased in several stages
from 85°F to 130°F to capture the performance parameters at each stage. The
performance parameters evaluated included capacity, power demand, and EER.
For the purpose of a general comparison to the ECCU over the 85°F to 115°F dry
bulb temperature range, the results of all six air-cooled units (standard and highefficiency) are averaged together and presented in Figure 24 below. These results
were taken from Reference 1 and re-formatted for presentation here. Over the
tested temperature range of 85°F to 115°F, the average performance data of the six
air cooled units showed a 13% decrease in net cooling capacity with a simultaneous
26% increase in power (kW) consumption with increasing dry bulb temperature. The
resulting Energy Efficiency Ratio (EER) decreased by 34%.
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72,000
14
Net Cooling Capacity (Btu/hr)
10
48,000
8
36,000
6
24,000
4
12,000
EER (Btu/hr/Watt) & Demand (kW)
12
60,000
2
0
0
(DB 85°F)
(DB 95°F)
(DB 105°F)
(DB 115°F)
Air-Cooled Unit Performance at Various Drybulb Temperatures.
Net Capacity (Btu/hr)
EER (Btu/hr/Watt)
Power (kW)
FIGURE 24. AIR-COOLED UNIT PERFORMANCE AT VARIOUS DRY BULB TEMPERATURES (REFERENCE 1)
It should be noted that the tests on the air-cooled units were conducted over an
outdoor temperature range of 85°F to 130°F under no specified wet bulb
temperature as wet bulb does not affect the performance of an air-cooled condenser
unit. The results of these air cooled units were determined as a function of outdoor
dry bulb temperature only. The tests on the evaporatively-cooled unit however,
were performed based on a series of climate zone conditions defined by specified dry
bulb and wet bulb temperatures as the evaporative cooling process and thus the
evaporative condenser performance depend on the wet bulb temperature. The
results of the tests on the ECCU were determined as a function of both dry bulb and
wet bulb.
It should also be noted that the air-cooled units referenced in Figure 24 were 5-ton
roof-top package units while the ECCU tested was a 3-ton residential split-system.
All general performance comparisons, therefore, were normalized for a ‘per-ton of
net cooling capacity’ comparison between the two unit types.
NET COOLING CAPACITY
In lieu of a direct comparison between the two unit types, a general comparison can
be made by examining the performance at several outdoor dry bulb temperatures.
Figure 25 shows, for example, CTZ 6 can be represented by a dry bulb temperature
of 84°F, and the HDAC zone by 115°F. Comparing these points with the air-cooled
condenser test results for an 85°F to 115°F range, it can be seen that the capacity of
the ECCU is reduced by 5.7% while the capacity for the average of the air-cooled
units is reduced by 13% over this range (see Figure 25).
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12,800
12,600
Net Cooling Capacity per Ton (Btu/hr/ton)
12,400
12,200
CTZ 6
CTZ 10
12,000
CTZ 8
CTZ 7
CTZ 9
11,800
Baseline
CTZ 13
CTZ 15
11,600
CTZ 14
CTZ HDAC
11,400
11,200
11,000
10,800
80
85
90
95
100
105
110
115
120
Outdoor Dry Bulb Temperature (°F)
ECCU Net Capacity (Btu/hr)
Air-Cooled Net Capacity (Btu/hr)
FIGURE 25. NET COOLING CAPACITY NORMALIZED PER TON FOR THE EVAPORATIVELY-COOLED VERSUS AIRCOOLED TYPE A/C SYSTEMS AS A FUNCTION OF OUTDOOR DRY BULB TEMPERATURE
It is interesting to note the air cooled units have a high net cooling capacity per ton
at low dry bulb temperature but a low net cooling capacity per ton at the higher dry
bulb temperatures. This is because the condensing temperature of air-cooled units is
directly dependent on the outdoor ambient dry bulb temperature. As the outdoor
temperature gets cooler, the condenser is able to more efficiently reject heat to the
outside air and the cooling capacity of the unit increases. The ECCU condensing
temperature, however, is driven by the moisture content of the condenser inlet air as
previously discussed and so does not show as significant a change in cooling capacity
with lower dry bulb temperatures.
At the AHRI 210/240 Standard Rating Condition of 95°F for outdoor air-cooled units,
the net cooling capacity is slightly higher than 12,000 Btu/hr/ton. This is an
indication that the units are designed to be slightly oversized or over rated at low dry
bulb temperatures in order for the performance to be acceptable at higher dry bulb
temperatures. This over sizing is, apparently, necessary because the performance
decreases significantly with increasing dry bulb temperatures.
It is also interesting to note that the ECCU is slightly under sized at the same AHRI
Standard Rating Condition of 95°F. In fact, the net cooling capacity per ton is mostly
below 12,000 Btu/hr/ton over the same temperature range. The ECCU however,
does not show as significant a loss of net cooling capacity as the air-cooled unit over
the same temperature range. This is indicated by the slope of their respective trend
lines.
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The average net cooling capacity of the air-cooled units decreased 13% or 54.4
Btu/hr/ton per °F increase over this temperature range. Compare this to the less
severe decrease in net cooling capacity of only 5.7% or 22 Btu/hr/ton per degree °F
increase for the ECCU over the same dry bulb temperature range. This is an
absolute drop in net cooling capacity of 1,632 Btu/hr/ton for the air-cooled units and
only 690 Btu/hr/ton for the ECCU.
Figure 25 indicates that at approximately 101°F the net cooling capacity per ton of the
ECCU becomes greater than that of the air-cooled units. Also, neither the air-cooled units
nor the ECCU unit was able to produce their rated cooling capacity at the extreme
temperature of 115°F; though the net cooling capacity per ton of the ECCU was greater than
that of the air cooled units at this temperature.
POWER
Perhaps one of the strongest advantages to the evaporative condenser system is the
relatively flat or steady power consumption across a wide range of outdoor dry bulb
temperatures (see Figure 26). This is where the performance of air-cooled
condenser systems is greatly compromised. As outlined previously, when
temperatures get hotter the cooling capacity of air cooled units decreases while they
consume even more power to try to satisfy the need for cooling. This is an inherent
problem with air cooled condenser technology as the compressor has to work harder
to reject the heat of the refrigerant to the hotter outside air. The result is evident as
Figure 26 shows the increase in power consumption with increasing outdoor
temperatures for the air-cooled condenser units versus the relatively unchanged
power consumption for the ECCU.
Additionally, at the low dry bulb temperatures where the air-cooled units perform
best, the power demand per ton of cooling capacity for the ECCU was 0.3 kW/ton
more efficient. This is indicated by the upward offset in the power demand curve for
the average of the air-cooled units.
Another interesting point to note is that for the same cooling capacity per ton at
approximately 101°F dry bulb (refer to Figure 25), the trend line of the power
consumed per ton at 101°F (Figure 26) indicates the air-cooled unit consumes
significantly more power than the ECCU for the same output cooling capacity. This
implies for this dry bulb, that even when the net cooling capacities per ton are the
same, the air-cooled unit requires 0.379 kW/ton more power to produce that cooling
capacity than does the ECCU.
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1.6
1.4
Demand (kW/Ton).
1.2
1.0
0.8
Baseline
CTZ 8
CTZ 7
CTZ 9
CTZ 6
CTZ 15
CTZ 13
CTZ 10
CTZ HDAC
CTZ 14
0.6
0.4
0.2
0.0
80
85
90
95
100
105
110
115
120
Outdoor Dry Bulb Temperature (°F)
ECCU Demand (kW/Ton)
Air-Cooled Demand (kW/Ton)
FIGURE 26. POWER CONSUMED BY THE EVAPORATIVELY-COOLED VERSUS AIR-COOLED TYPE A/C SYSTEMS AS A
FUNCTION OF OUTDOOR DRY BULB TEMPERATURE
The change in power demand shown for the air-cooled units as temperature
increased was 27% or 10 watts/ton per °F over the range of temperatures tested.
The change in power demand for the ECCU unit over the same temperature range
was nearly zero at less than 0.1 % or less than 1 Watt/ton per °F. This is an
approximation based on the trend line of the experimental data. This result closely
matches the expected trend for power demand because the saturated condensing
temperature and thus the condenser discharge pressure are dependent on the
moisture content of the condenser inlet air. This means for a given wet bulb, as the
dry bulb temperature increases the amount of water evaporated within the unit
increases because more water can be evaporated from the air and the thin water film
covering the refrigerant tubes.
As confirmation of this, Figure 27 shows the ECCU power consumption as a function
of increasing ambient wet bulb temperature changes only slightly as the wet bulb
temperature increases.
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4
Power (kW/ton)
3.5
3
2.5
2
y = 0.012x + 2.4986
1.5
2
R = 0.838
1
CTZ 6
(84 DB,
67 WB)
CTZ 9
(94 DB,
68 WB)
CTZ 7
(83 DB,
69 WB)
CTZ 8
(89 DB,
69 WB)
CTZ 10 CTZ 14 CTZ 13 CTZ 15
HDAC Baseline
(100 DB, (108 DB, (101 DB, (111 DB, (115 DB, (95 DB,
69 WB) 69 WB) 71 WB) 73 WB) 74 WB) 75 WB)
Increasing Wet Bulb Temperature (°F)
Power (kW/ton)
FIGURE 27. ECCU POWER CONSUMPTION VERSUS INCREASING WET BULB TEMPERATURE
Similarly, Figure 28 shows that the ECCU discharge pressure that caused the change
in power consumption also follows the same trend of a slight increase with increasing
wet bulb temperature.
400
350
Pressure (psig)
300
250
200
150
100
50
0
CTZ 6
(84 DB,
67 WB)
CTZ 9
(94 DB,
68 WB)
CTZ 7
(83 DB,
69 WB)
CTZ 8
(89 DB,
69 WB)
CTZ 10
(100 DB,
69 WB)
CTZ 14
(108 DB,
69 WB)
CTZ 13
(101 DB,
71 WB)
CTZ 15
(111 DB,
73 WB)
CTZ HDAC BASELINE
(115 DB, (95 DB,
75 WB)
74 WB)
Increasing Wet Bulb (°F)
Discharge Pressure, psig
Suction Pressure, psig
FIGURE 28. ECCU DISCHARGE AND SUCTION PRESSURE VERSUS INCREASING WET BULB TEMPERATURE
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Saturated Condensing Temperature (°F)
Figure 29 shows that the ECCU saturated refrigerant condensing temperature that
caused the change in discharge pressure is also seen to follow the same slightly
increasing trend with increasing wet bulb temperature. This average saturated
condensing temperature of 84°F is determined by the average 84°F wet bulb
temperature of the water on the refrigerant tubes.
100
90
80
70
60
50
40
30
20
y = 0.6776x + 81.789
R2 = 0.83
10
0
CTZ6
(84 DB,
67 WB)
CTZ9
(94 DB,
68 WB)
CTZ7
(83 DB,
69 WB)
CTZ8
(89 DB,
69 WB)
CTZ10
CTZ14
CTZ13
CTZ15
HDAC Baseline
(100 DB, (108 DB, (101 DB, (111 DB, (115 DB, (95 DB,
69 WB) 69 WB) 71 WB) 73 WB) 74 WB) 75 WB)
Increasing Wet Bulb Temperature (°F)
SCT (°F)
FIGURE 29. ECCU SATURATED CONDENSING TEMPERATURE VERSUS INCREASING WET BULB TEMPERATURE
The trend of Figure 29 back to Figure 27 indicates that as the inlet condenser air is
more humid, there is less capacity of that air to perform the mass transfer process of
evaporating the water on the condenser refrigerant tubes so the compressor has to
work harder (more power) to reject the heat of the refrigerant in the condensing
unit. However, the ECCU does not severely increase power consumption with
increasing wet bulb temperatures nor with increasing dry bulb temperatures as does
air-cooled condenser technology.
ENERGY EFFICIENCY RATIO (EER)
The EER is defined as the ratio of net cooling capacity over the power required to
produce that cooling capacity. This ratio is defined at the AHRI 210/240 rating
temperature of 95°F DB for air cooled condensers and 95°F DB at 75°F WB for
evaporatively-cooled condensers. Although the EER is defined at a single point for
reporting purposes, Figure 30 shows that it changes as outdoor temperatures
change.
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15.0
14.0
13.0
EER (Btu/hr/watt)
12.0
11.0
10.0
9.0
8.0
7.0
6.0
80
85
90
95
100
105
110
115
120
Outdoor Dry Bulb Temperature (°F)
ECCU EER (Btu/hr/Watt)
Air-Cooled EER (Btu/hr/Watt)
FIGURE 30. EER OF EVAPORATIVELY-COOLED VERSUS AIR COOLED TYPE A/C SYSTEMS AS A FUNCTION OF
OUTDOOR DRY BULB TEMPERATURE
Similar to the decrease in net cooling capacity shown in Figure 25, the EER of the
evaporatively-cooled condenser also decreased (by approximately 10%) as a
function of increasing temperature conditions (Figure 30). However, the decrease in
EER is more severe (approximately 34%) as a function of increasing temperature
conditions for the air cooled condensing units. This is a direct result of the decrease
in net cooling capacity that is coincident with the increased power consumption as
the dry bulb temperatures increased.
Overall, the performance of the evaporatively-cooled condensing unit suffers less in
terms of net cooling capacity and energy consumption than does an air-cooled
condensing unit. However, this performance benefit involves consumption of water,
which can become an issue in arid areas with limited water supply.
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CONCLUSIONS
The tests demonstrated that the evaporatively-cooled system experiences substantially less
performance degradation under hot climate conditions than conventional air-cooled air
conditioner systems. The energy savings over air-cooled condenser technology increase as
the outdoor temperature increases where conventional air-cooled systems become less
efficient. Additionally, the performance advantage of the evaporative condenser system is
greatest in climate zones with high outdoor temperatures and low humidity.
PERFORMANCE
The humidity ratio of air passing through the evaporatively-cooled condenser drives
the saturated condensing temperature of the refrigerant toward the wet bulb
temperature of the air that is always lower than the dry bulb temperature.
Condensing temperature is thus limited by ambient wet bulb temperature that is
typically 14°F - 25°F below the ambient dry bulb temperature.
The TTC test results confirm the performance of evaporatively-cooled condenser
technology to be significantly better than that of air-cooled condenser technology in
hot and dry climate zones. In mild climate conditions of less than 101°F DB the aircooled units demonstrated higher cooling capacity. However, the test results also
showed the ECCU had up to 3.6% higher net cooling capacity per ton than the aircooled units in the extreme hot and dry climate zone conditions. At temperatures
above 101°F DB the performance of air-cooled condenser technology A/C degrades
dramatically. The measured test data also showed up to 5 points greater EER and
0.5 kW/ton less power consumption for the ECCU than for the air-cooled units across
the climate/temperature conditions tested.
Under ARI-Test A conditions of 95°F DB and 75°F WB, the ECCU produced a net
cooling capacity of 11,825 Btu/hr/ton. The unit consumed 0.9 kW/ton of power and
yielded an EER of 13.5 Btu/hr/watt.
In general, at higher DB temperatures the evaporatively-cooled condenser
technology allows a higher net cooling capacity per ton than that of the air-cooled
condenser technology without compromising the comfort conditions of supply air
temperature and humidity. Additionally, the ECCU technology does not severely
increase power consumption with increasing outdoor temperatures as does the aircooled condenser technology.
WATER CONSUMPTION
Taking advantage of the evaporative cooling process using water is beneficial to
energy efficiency. The lowered saturated refrigerant condensing temperature due to
wet bulb temperature results in a significant benefit of reduced energy consumption.
Water consumption versus water availability, however, may be an issue in areas
where evaporatively-cooled condensing units perform optimally.
Under ARI-Test A conditions of 95°F DB and 75°F WB, the ECCU purged 0.59
gal/hr/ton of water each hour and evaporated 1.86 gal/hr/ton for a total water
consumption of 2.45 gal/hr/ton.
The total ECCU water consumption increased with increasing changes in the humidity
ratio (Δω) of the condenser inlet air or about 0.5 gal/hr increase for every 10%
decrease in relative humidity.
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In regard to water consumption, it should be noted that electrical power generation
plants consume fresh water if their steam condensers operate with water from rivers
or lakes. In this case, on a grid-level scale some of the water consumption may be
offset to some extent by a reduction in water used in electricity generation
proportional to the reduction in electrical energy consumption by the more efficient
ECCU units. However, it takes energy to pump water to the site of an ECCU and
there are thus energy losses in the water distribution system that may offset some of
these gains.
WATER QUALITY AND MAINTENANCE
Study of the maintenance requirements of the ECCU system was outside the scope of
this performance evaluation. It should be noted however, that the ECCU system
includes additional components over an air-cooled system that are required to pump
and handle the water and these may introduce the potential for higher maintenance
costs. In contrast, however, it should be noted that in hot climate zones where the
ECCU system is most advantageous, the maintenance needs of conventional aircooled air conditioners will likely be higher than average because the system is
operating much of the time under extreme conditions.
Another concern is the long-term performance of the ECCU technology that may be
affected by buildup of hard water deposits on the refrigerant lines. This in turn
decreases the heat transfer effectiveness from the refrigerant lines to the surface
water film and influences the ability to cool the refrigerant. This buildup of hard
water deposits ultimately results in decreased performance of the unit. The degree
of decreased performance is a function of the water quality (hardness), period of
exposure, and evaporation rate or climate zone. While it is certain that water quality
is a long-term issue with this technology, the long-term effects of water quality are
beyond the scope of this evaluation.
The long-term effects of water quality on the performance of the ECCU will next be
evaluated by the Western Cooling Efficiency Center (WCEC) in Northern California.
The intent here is to take the unit to failure mode while evaluating the long-term
operational effects of hard water deposits on the refrigerant lines and other internal
components of the condensing unit.
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RECOMMENDATIONS
To further assess the potential for commercial viability and possible value of an incentive
program for evaporatively-cooled condenser technology on air conditioning systems, some
additional evaluations should be undertaken.
1. Conduct field testing to verify that laboratory test results can be achieved under
actual operating conditions in the field.
2. Investigate impact of water usage on water supply, environmental impact, and cost
issues in hot dry climate zones where the system is most effective. The water usage
may be partially offset by a reduction in power plant water usage proportional to the
reduction in electrical energy consumption by the unit.
3. Investigate potential for reducing the system water consumption. Options may
include potential recycling of condensate water for cooling, reduction in sump purge
volumes or cycles, and use of purge water for irrigation.
4. Investigate potential maintenance, reliability, corrosion, scaling, and other water
related operational issues.
5. Perform cost/benefit assessment of installation of systems in hotter SCE climate
zones. Include current and future cost projections, maintenance, water consumption,
lifetime, operating costs, and other related issues.
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REFERENCES
1. Performance Evaluation of Rooftop Air Conditioning Units at High Ambient Temperatures,
Faramarzi R., Coburn B., Sarhadian R., Mitchell S., Pierce R., 2004, American Council for
an Energy-Efficient Economy (ACEEE) Paper 202. Southern California Edison,
Refrigeration and Thermal Test Center, Design and Engineering Services.
2. 2005 ASHRAE Handbook Fundamentals, American Society of Heating, Refrigeration and
Air-Conditioning Engineers, Inch-Pound Edition. Copyright 2005 by the American
Society of Heating, Refrigeration and Air-Conditioning Engineers, Inc.
3. Air-Conditioning & Refrigeration Institute 2003 Standard for Unitary Air-Conditioning and
Air-Source Heat Pump Equipment – Standard 210/240. Copyright 2003 by AirConditioning and Refrigeration Institute.
4. Climatic Data for Region X: Arizona, California, Hawaii, Nevada. ASHRAE Publication
SPCDX. Golden Gate and Southern California Chapters. American Society of Heating,
Refrigeration & Air-Conditioning Engineers, Fifth Edition, May 1982, Copyright 1982.
5. 2004 ASHRAE Handbook HVAC Systems and Equipment, American Society of Heating,
Refrigeration and Air-Conditioning Engineers, Inch-Pound Edition. Copyright 2004 by
the American Society of Heating, Refrigeration and Air-Conditioning Engineers, Inc.
6. ASHRAE Standard 37-2005 Methods of Testing for Rating Electrically Driven Unitary AirConditioning and Heat Pump Equipment, American Society of Heating, Refrigeration and
Air-Conditioning Engineers, Inc. Copyright 2005 ASHRAE, Inc.
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APPENDICES
APPENDIX A
SUPPLEMENTAL PERFORMANCE EVALUATIONS
EVAPORATOR COIL CHARACTERISTIC PERFORMANCE
Due to temperature stratifications on the exit side of the evaporator coil, the air
temperature at this point was calculated using the measured temperature after the
evaporator fan and subtracting the change in temperature caused by the evaporator
fan motor. The temperature differential across the evaporator coil was determined
based on the measured air temperature at the inlet of the evaporator coil, and the
computed air temperature at the outlet of the evaporator coil, Equation 18.
EQUATION 18.
TEMPERATURE DIFFERENTIAL ACROSS EVAPORATOR
kWevap  fan * k 


Tevap  Tevap in  Tevap out  Tevap in   Tevapfan out 


.24 * m air 

where
ΔTevap
= temperature differential across the evaporator coil, oF
Tevap-in
= air temperature at the inlet of the evaporator coil, oF
Tevap-out
= air temperature at the outlet of the evaporator coil, oF
Tevapfan-out = air temperature at the outlet of the evaporator fan, oF
k
= conversion factor, (3,413 Btu/hr/kW)
Another indication of coil performance is the evaporator temperature difference (TD).
It is defined as the difference in temperature between the temperature of the air
leaving the evaporator and the saturation temperature of the refrigerant
corresponding to the pressure at the compressor inlet, Equation 19. The saturation
temperature at the inlet of the compressor is determined using the xProps program.
EQUATION 19.
EVAPORATOR TEMPERATURE DIFFERENCE
TDevap = Tevap-out – SST
where
TDevap
SST
Southern California Edison
Design & Engineering Services
= evaporator TD, oF
= saturated suction temperature based on compressor inlet pressure,
F
o
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Additionally, the evaporator coil effectiveness was determined. As shown in Equation
20, the effectiveness is defined as the ratio of actual to maximum possible heat
transfer rate.
EQUATION 20.
ε
EVAPORATOR COIL EFFECTIVENESS
ΔTevap
Tevap  in  SST
where

= evaporator coil effectiveness, unit-less
CONDENSER COIL CHARACTERISTIC PERFORMANCE
The temperature differential across the condenser coil was determined based on
measured air temperatures at the inlet and outlet of the condenser coil, Equation 21.
EQUATION 21. TEMPERATURE DIFFERENTIAL ACROSS CONDENSER COIL
ΔTcond = Tcond-in – Tcond-out
where
Tcond
= temperature differential across the condenser coil, oF
Tcond-in
= air temperature at the inlet of the condenser coil, oF
Tcond-out
= air temperature at the outlet of the condenser coil, oF
The condenser TD is defined as the difference in temperature between the
temperature of the air entering the condenser and the saturation temperature of the
refrigerant corresponding to the pressure at the compressor outlet, Equation 22. The
saturation temperature at the outlet of the compressor was determined using the
xProps program.
EQUATION 22. CONDENSER TEMPERATURE DIFFERENCE
TDcond = SCT – Tcond-in
where
TDcond
= condenser TD, oF
SCT
= saturated condensing temperature based on compressor outlet
pressure, oF
The heat exchange effectiveness of the condenser is also dependent on its LMTD and
its UA. However, utilizing the arithmetic mean temperature difference can provide an
approximation to the actual LMTD. The arithmetic mean temperature difference is
determined using the air temperatures at the inlet and outlet of the condenser and
the saturated condensing temperature based on compressor outlet pressure
according to Equation 23.
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EQUATION 23.
Tmean 
ET 08.08
CONDENSER ARITHMETIC MEAN TEMPERATURE DIFFERENCE
SCT  Tcond  in   SCT  Tcond  out 
2
where
Tmean
= arithmetic mean temperature difference, oF
Prior to determining the UA of the condenser coil, the heat of rejection at the
condenser needed to be obtained. According to Equation 24, the heat of rejection is
the sum of gross cooling capacity and the heat of compression, that can be directly
determined from the input power to the compressor.
EQUATION 24. CONDENSER HEAT OF REJECTION


Q rej = m refrig ( hrefrig  in - hrefrig  out )
where

= heat of rejection at the condenser, Btu/hr
Q rej

m refrig
= mass flow rate of refrigerant, lb/hr
hrefrig  out
= subcooled liquid refrigerant enthalpy at the condenser outlet, Btu/lb
hrefrig  in
= superheated refrigerant enthalpy at the condenser inlet, Btu/lb
After the heat of rejection and arithmetic mean temperature difference were
determined, Equation 25 was used to obtain the effective overall heat transfer
coefficient of the condenser coil (UA).
EQUATION 25. CONDENSER OVERALL HEAT TRANSFER COEFFICIENT

UAcond 
Q rej
Tmean
where
UAcond
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Design & Engineering Services
= effective overall heat transfer coefficient of the condenser coil,
Btu/hr-oF
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Performance Evaluation of an Evaporatively-Cooled Split-System Air Conditioner
ET 08.08
SYSTEM CHARACTERISTIC PERFORMANCE
One of the system parameters is the evaporator coil superheat. The evaporator coil
superheat was determined based on vapor refrigerant temperature at the evaporator
coil outlet and the saturation temperature of the refrigerant corresponding to the
pressure at the compressor inlet according to Equation 26.
EQUATION 26. EVAPORATOR COIL SUPERHEAT
TSH-evap = Tv – SST
where
TSH-evap
= evaporator coil superheat, oF
Tv
= vapor refrigerant temperature at the evaporator coil outlet, oF
Equation 27 was used to determine the condenser subcooling. Condenser subcooling
was obtained by subtracting the liquid refrigerant temperature at the condenser coil
outlet from the saturated condensing temperature based on compressor outlet
pressure.
EQUATION 27. CONDENSER COIL SUBCOOLING
TSC-cond = SCT – TL
where
TSC-cond
= condenser coil subcooling, oF
SCT
= saturated condensing temperature based on compressor outlet
pressure, oF
TL
= liquid refrigerant temperature at the condenser coil outlet, oF
The total system subcooling, that takes place after the liquid refrigerant leaves the
condenser coil until it reaches the refrigerant metering device of the unit, was
determined using Equation 28, i.e., subtracting the liquid refrigerant temperature at
the metering device inlet from the saturation temperature of refrigerant
corresponding to the pressure after the mass flow meter (metering device inlet). The
liquid refrigerant temperature at the metering device inlet excludes the subcooling
effect of the mass flow meter.
EQUATION 28. TOTAL SYSTEM SUBCOOLING
TSC-total = STmfm-out – TMD-in
where
TSC-total
= total system subcooling, oF
STmfm-out
= saturation temperature of refrigerant based on pressure after mass
flow meter, oF
TMD-in
= refrigerant temperature at the metering device inlet, less subcooling
effect, oF
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Performance Evaluation of an Evaporatively-Cooled Split-System Air Conditioner
ET 08.08
HEAT AND WORK OF COMPRESSION
The heat of compression, which is a function of refrigerant suction and discharge
enthalpies, is obtained using Equation 29.
EQUATION 29. HEAT OF COMPRESSION
h= hdischarge – hsuction
where
h
= Heat of compression, Btu/lb
hdischarge
= Enthalpy of superheat vapor discharged from compressor, Btu/lb
hsuction
= Enthalpy of low pressure vapor at compressor inlet, Btu/lb
After determining the heat of compression, the work of compression was obtained by
multiplying the heat of compression by the mass flow rate of refrigerant, Equation
30.
EQUATION 30. WORK OF COMPRESSION


W comp  m refrig * h
where

W comp

= Work of compression, Btu/hr
m refrig
= mass flow rate of refrigerant, lb/hr
h
= Heat of compression, Btu/lb
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Performance Evaluation of an Evaporatively-Cooled Split-System Air Conditioner
ET 08.08
APPENDIX B
SUMMARY TABLE OF ECCU PERFORMANCE PARAMETERS
TABLE 5. SUMMARY TABLE OF ECCU PERFORMANCE PARAMETERS ACROSS THE DIFFERENT CLIMATE ZONE CONDITIONS TESTED
Parameter
CTZ
Baseline
CTZ 6
CTZ 7
CTZ 8
CTZ 9
CTZ 10
CTZ 13
CTZ 14
CTZ 15
CTZ
HDAC
CTZ Dry Bulb Temp. (°F)
95
84
83
89
94
100
101
108
111
115
CTZ Wet Bulb Temp. (°F)
75
67
69
69
68
69
71
69
73
74
35,474
36,343
35,840
35,753
35,736
36,188
35,258
34,580
34,766
34,274
11,825
12,114
11,947
11,918
11,912
12,063
11,753
11,527
11,589
11,425
2.62
2.50
2.56
2.55
2.53
2.55
2.57
2.54
2.61
2.61
0.87273
0.83254
0.85245
0.84997
0.84425
0.85099
0.85813
0.84581
0.86993
0.87135
EER (Btu/hr/watt)
13.5
14.6
14.0
14.0
14.1
14.2
13.7
13.6
13.3
13.1
Water Evaporated
(gal/hr)
5.57
5.14
4.93
5.25
5.67
5.72
5.63
6.45
6.49
6.74
1.71
1.86
1.58
1.88
1.48
2.12
2.03
1.82
1.83
1.79
7.35
7.01
6.52
7.14
7.46
7.83
7.75
8.27
8.32
8.53
0.01414
0.0103
0.01203
0.01064
0.00874
0.00812
0.00943
0.00629
0.00875
0.00867
0.02261
0.01526
0.01269
0.01955
0.02683
0.03509
0.0352
0.04953
0.05264
0.061
39.81
41.24
49.59
36.3
25.58
19.79
22.28
12.13
15.42
13.62
Net Cooling Capacity
(Btu/hr)
Net Cooling Capacity per
ton (Btu/hr/ton)
Power (kW)
Power per ton (kW/ton)
Water Purged (gal/hr)
Total Water Consumption
(gal/hr)
Outdoor Humidity Ratio
(ω)
Change in Humidity Ratio
Across Condenser (Δω)
Outdoor Relative
Humidity (%Rh)
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Performance Evaluation of an Evaporatively-Cooled Split-System Air Conditioner
ET 08.08
NOTES:
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November 2009