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Design & Engineering Services
FIELD TEST OF HYBRID ROOFTOP UNIT PHASE 1
HT.11.SCE.008 Report
Prepared by:
Design & Engineering Services
Customer Service Business Unit
Southern California Edison
December 2012
Field Test of Hybrid Rooftop Unit HT.11.SCE.008
Southern California Edison Page ii Design & Engineering Services December 2012
Acknowledgements
Southern California Edison’s Design & Engineering Services (DES) group is responsible for
this project. It was developed as part of Southern California Edison’s HVAC Technologies
and System Diagnostics Advocacy (HTSDA) Program under internal project number
HT.11.SCE.008. Jay Madden, P.E., conducted this technology evaluation with overall
guidance and management from Jerine Ahmed. For more information on this project,
contact jay.madden@sce.com.
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.
Field Test of Hybrid Rooftop Unit HT.11.SCE.008
Southern California Edison Page iii Design & Engineering Services December 2012
Executive Summary According to the California Energy Commission’s 2003 Commercial End-Use Survey [1],
packaged air conditioning (AC) units cool 65% of commercial spaces and account for a large
amount of peak electrical demand loads during hot weather. These units can be retrofitted
with systems that lower electrical energy usage and demand by evaporating water to cool
air passing over the condenser coils.
The purpose of this field assessment is to evaluate the performance of hybrid evaporative
cooling technology on package air conditioning units serving a big-box retail store. Electrical
consumption, electrical demand, water consumption, and maintenance issues associated
with water evaporation are evaluated.
Hybrid evaporative-cooling systems are retro-fitted onto 13 rooftop packaged air
conditioning units serving the sales floor of a big-box retail store in Palmdale, California.
These systems are designed to pre-cool both condenser air and incoming outside air. The
systems were operated from September 1 to September 28, 2012 while cooling capacity,
compressor run-time, electrical consumption, and water consumption were measured. These measurements were taken at outdoor air temperatures ranging from 80 to 100°F. The
original air conditioning systems are also operated, from September 28 to October 26,
2012, and the same performance measurements were recorded.
Monitoring system problems prevent measurement of AC unit electrical and water
consumption. Phase 2 of this study addresses these monitoring issues. The following
performance improvements are observed in Phase 1, when evaporative pre-cooling is
applied:
Overall store electric meter load reduces 3%, but this value is not statistically
significant, given the many factors that affect this load.
Refrigerant saturated condensing temperature drops 20%.
Condensing pressure drops 25%.
After four weeks of operation, dissolved solids accumulate on the evaporative media
surfaces of three of the systems, and algae grows in the basins of four of the systems.
The operational and monitoring system issues on this project will be corrected and further
monitoring will be conducted during the summer of 2013. This added testing will provide
more detailed AC unit electrical and water consumption data. The longer term water
treatment issues will also be observed.
We recommend developing an eQuest computer model of electrical savings provided by this
measure, serving different commercial building types in various climate zones. This model
would be calibrated with actual measured results of this field assessment.
The results of this study suggest that SCE’s EE program adopt this technology, but as
mentioned earlier, continued monitoring at this site will provide a more definitive
recommendation. This study will occur during the summer of 2013.
Field Test of Hybrid Rooftop Unit HT.11.SCE.008
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Acronyms AC Air Conditioning
ASHRAE American Society of, heating, Refrigeration, and Air Conditioning Engineers
Btu British Thermal Unit
Btu/hr British Thermal Unit/hour
CFM Cubic Feet per Minute
COP Coefficient of Performance
DB Dry Bulb
DX Direct eXpansion
EE Energy Efficiency
EMCS Energy Monitoring and Control System
HTSDA HVAC Technologies and System Diagnostics Advocacy
HVAC Heating, Ventilating, and Air Conditioning
lb pound
OSA Outside Air
RA Return Air
RTU Rooftop Unit
SA Supply Air
SCE Southern California Edison
SCT Saturated Condensing Temperature
SST Saturated Suction Temperature
TxV Thermostatic Expansion Valve
WB Wet Bulb
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Contents
EXECUTIVE SUMMARY _______________________________________________ III
INTRODUCTION ____________________________________________________ 1
Evaporative Condensing ........................................................... 1
Hybrid Evaporative Cooling Technology ...................................... 1
Some systems being introduced on the market combine
condenser air evaporative pre-cooling with OSA sensible
pre-cooling. This technology circulates cooled water
through a finned coil, located in the OSA intake stream.
OSA cools while rejecting heat into the water. The
warmed water then drips through the condenser pre-
cooling media and evaporates.Background ....................... 1
Evaporative Pre-Cooling Condenser Air ...................................... 2
Dual-Cooling Technology .......................................................... 6
Market Barriers .................................................................. 7
ASSESSMENT OBJECTIVES ____________________________________________ 8
TECHNOLOGY/PRODUCT EVALUATION __________________________________ 9
TECHNICAL APPROACH/TEST METHODOLOGY ___________________________ 11
Field Testing of Technology .................................................... 11
Test Plan ......................................................................... 12 Instrumentation Plan ........................................................ 14
RESULTS_________________________________________________________ 15
Data Analysis ........................................................................ 15
Condenser Air Pre-cooling ................................................. 15 Overall Energy Impact ...................................................... 18 Water Usage .................................................................... 18 Water Treatment .............................................................. 19
DISCUSSION _____________________________________________________ 20
System Installation and Operation ........................................... 20
Energy Savings ..................................................................... 20
Water Treatment ................................................................... 21
RECOMMENDATIONS ______________________________________________ 23
CONCLUSIONS ___________________________________________________ 24
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APPENDIX A – SAMPLE REFRIGERATION CALCULATIONS ____________________ 25
APPENDIX B – MONITORING EQUIPMENT _______________________________ 30
REFERENCES _____________________________________________________ 33
Figures Figure 1. Direct Expansion Refrigerant Cycle ...................................................... 2
Figure 2. Air-Cooled Condenser ......................................................................... 3
Figure 3. Psychrometric Chart – Air-Cooled Condenser, Ontario, CA 1%
Design Dry-Bulb Temperature ............................................................. 4
Figure 4. Evaporative Pre-Cooling Condenser Air ................................................. 5
Figure 5. Psychrometric Chart – Evaporative Condenser, Ontario, CA 1%
Design Dry-Bulb Temperature ............................................................. 6
Figure 6. Dual-Cooling Technology .................................................................... 7
Figure 7. Assessment Site RTU Schedule ............................................................ 9
Figure 8. Assessment Site Roof Plan ................................................................ 10
Figure 9. Lennox AC Unit Condenser Configuration ............................................. 11
Figure 10. Monitoring Points ............................................................................ 13
Figure 11. Scale Formation on Evaporative Media ............................................... 19
Figure 12. Algae Growth in Sump ..................................................................... 19
Figure 13. Pressure-Enthalpy Chart for R-22 Refrigerant .................................... 25
Tables Table 1. Refrigerant Liquid Temperature, AC-7 and AC-8 ..................................... 15
Table 2. Refrigerant Compressor Lift, AC-7 ....................................................... 16
Table 3. Cooling Stages, Baseline, 3 Compressor AC Unit .................................... 16
Table 4. Cooling Stages, Evaporative Pre-Cooling, 3 Compressor AC Unit .............. 17
Table 5. Cooling Stages, Baseline, 4 Compressor AC Unit .................................... 17
Table 6. Cooling Stages, Evaporative Pre-Cooling, 4 Compressor AC Unit .............. 17
Table 7. Dual-Cooling System Pump Power Consumption .................................... 18
Table 8. Site Overall Energy Demand, Using Dry Bulb Temperature ..................... 18
Table 9. Monitoring Equipment ........................................................................ 30
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Equations Equation 1. Evaporator Heat Rejection ............................................................. 26
Equation 2. Evaporator Heat Rejection, R-22 Refrigerant, 40°F SST, 120°F SCT .... 26
Equation 3 .Evaporator Heat Rejection, R-22 Refrigerant, 40°F SST, 100°F SCT ..... 26
Equation 4. Cooling Capacity Increase, R-22 Refrigerant, 120°F to 100°F SCT ....... 27
Equation 5. Compressor Work ......................................................................... 27
Equation 6. Compressor Work, R-22 Refrigerant, 40°F SST, 120°F SCT ................ 27
Equation 7. Compressor Work, R-22 Refrigerant, 40°F SST, 100°F SCT ................ 28
Equation 8. Cooling Work Reduction, R-22 Refrigerant, 120°F to 100°F SCT ......... 28
Equation 9. Refrigeration Cycle Coefficient of Performance ................................. 28
Equation 10. COP, R-22 Refrigerant, 40°F SST, 120°F SCT ................................. 29
Equation 11. COP, R-22 Refrigerant, 40°F SST, 100°F SCT ................................. 29
Equation 12. COP Increase, R-22 Refrigerant, 120°F to 100°F SCT ...................... 29
Field Test of Hybrid Rooftop Unit HT.11.SCE.008
Southern California Edison Page 1
Design & Engineering Services December 2012
Introduction According to the California Energy Commission’s 2003 Commercial End-Use Survey [1],
approximately 65% of commercial floor area is conditioned by packaged air conditioning (AC)
units. These units consist of supply fans, Direct eXpansion (DX) cooling systems, heating, and
air filters. These systems are inexpensive to install and are prevalent in schools, smaller office
buildings, retail buildings, and other light commercial applications. Because of prevalence and
energy performance, commercial AC units are a large part of the electrical demand when
outdoor temperatures are high.
EVAPORATIVE CONDENSING DX systems provide cooling by rejecting heat from the indoor conditioned space to the
outdoor air. The greater the temperature difference between the rooftop unit’s (RTU’s)
supply air (SA) and outdoor air (OSA) temperatures, the more mechanical energy is
required to remove the heat from the conditioned space. By contrast, air conditioning
systems serving larger commercial applications take advantage of the dry climate of
the western United States by evaporating water to reject heat into the atmosphere.
This evaporative process reduces both energy consumption and peak electrical
demand by lowering the DX system’s condensing temperature.
Systems are being introduced to the market that bring the advantages of this
evaporative cooling process to the packaged AC system. These systems are designed
to be retrofitted onto existing RTUs. They operate by evaporating water over a media
in the condenser air stream, cooling the incoming condenser air.
For this field test, we study electrical consumption, peak electrical demand, water
consumption, AC system performance, and maintenance issues associated with
evaporating water.
HYBRID EVAPORATIVE COOLING TECHNOLOGY Some systems being introduced on the market combine condenser air evaporative pre-
cooling with OSA sensible pre-cooling. This technology circulates cooled water through
a finned coil located in the OSA intake stream. OSA cools while rejecting heat into the
water. The warmed water then drips through the condenser pre-cooling media and
evaporates.
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BACKGROUND
EVAPORATIVE PRE-COOLING CONDENSER AIR Commercial AC and refrigeration systems use mechanical energy to move heat from a
lower temperature space to a higher temperature heat sink. This heat sink is typically
the outdoor environment. The refrigerant DX system works as follows:
1. The low-pressure refrigerant boils in the evaporator, which removes heat
from the airstream. This colder air is then delivered to the conditioned
space.
2. The compressor, driven by an electric motor or other means, raises the
pressure of the refrigerant gas.
3. In the condenser, heat is transferred from the refrigerant gas to the heat
sink. This heat sink can be the outdoor air, a water stream, or evaporating
water.
4. The thermostatic expansion valve (TxV) reduces the pressure of the
refrigerant liquid and repeats the cycle.
Figure 1 shows the DX refrigerant cycle.
TxV
Compressor
Evaporator
Condenser
Inside Air
Outside Air
Work2
3
4
1
Heat
Heat
FIGURE 1. DIRECT EXPANSION REFRIGERANT CYCLE
Field Test of Hybrid Rooftop Unit HT.11.SCE.008
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The amount of energy required by the compressor in step 2 depends upon the
temperature at which the refrigerant gas condenses. In RTUs and air-cooled chillers,
condensers reject heat from the refrigerant directly into the outside air stream. In
these systems, higher outside air temperatures result in higher energy usage by the
compressor.
Evaporative cooling takes advantage of the outside air’s ability to absorb moisture and
the heat of vaporization. As water evaporates, heat is absorbed from the surrounding
air, refrigerant or the remaining water stream. Cooling towers, evaporative
condensers, and closed-circuit fluid coolers use this process. Evaporative condensers
operate at a lower temperature than air-cooled condensers, which lowers the energy
required by the compressor.
Figure 2 shows a typical air-cooled condenser for an AC unit. In this example, we used
1% American Society of Heating, Refrigeration, and Air Conditioning Engineers
(ASHRAE) design cooling conditions for Ontario, California (98° DB, 70°WB). Air is
drawn through a condenser coil, where heat from the refrigerant is rejected into the
airstream. The condenser fan then discharges this heated air into the atmosphere.
OUTSIDE AIR
CONDENSER COIL
CONDENSER FAN
DISCHARGE98° F
FIGURE 2. AIR-COOLED CONDENSER
Field Test of Hybrid Rooftop Unit HT.11.SCE.008
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Figure 3 shows the psychrometric process where air is sensibly heated in the
condenser.
10 15 20 25
30
35
40
45
50
55
55
60
60
ENTHALPY - BTU PER POUND OF DRY AIR
15
20
25
30
35
40
45
50
ENTH
ALPY -
BTU P
ER P
OUND O
F DRY A
IR
SATURATIO
N T
EMPER
ATURE -
°F
35
40
45
50
55
60
65
70
75
80
85
90
95
10
0
10
5
11
0
11
5
12
0
DR
Y B
UL
B T
EM
PE
RA
TU
RE
- °
F
.002
.004
.006
.008
.010
.012
.014
.016
.018
.020
.022
.024
.026
.028
10% RELATIVE HUMIDITY
20%
30%
40%
50%
60%
70%
80%
90%
35
3540
4045
45 50
50 55
55 60
6065
65
70
70
75
75
80
80
85 WET BULB TEMPERATURE - °F
85
90
12.5
13.0
13.5
14.0 VO
LUM
E- C
U.F
T. P
ER
LB. D
RY
AIR
14.5
15.0
HU
MID
ITY
RA
TIO
- P
OU
ND
S M
OIS
TU
RE
PE
R P
OU
ND
DR
Y A
IR
OSA, Ontario, CA Condenser Discharge
R R
ASHRAE PSYCHROMETRIC CHART NO.1
NORMAL TEMPERATURE
BAROMETRIC PRESSURE: 29.921 INCHES OF MERCURY
Copyright 1992
AMERICAN SOCIETY OF HEATING, REFRIGERATING AND AIR-CONDITIONING ENGINEERS, INC.
SEA LEVEL
0
1.0 1.0
-
2.0
4.08.0
-8.0-4.0-2.0
-1.0
-0.5-0.4-0.3-0.2-0.1
0.1
0.2
0.3
0.4
0.5
0.6
0.8
-2000
-1000
0
500
1000
1500
2000
3000
5000
-
SENSIBLE HEAT Qs
TOTAL HEAT Qt
ENTHALPY
HUMIDITY RATIO
h
W
FIGURE 3. PSYCHROMETRIC CHART – AIR-COOLED CONDENSER, ONTARIO, CA 1% DESIGN DRY-BULB TEMPERATURE
Field Test of Hybrid Rooftop Unit HT.11.SCE.008
Southern California Edison Page 5
Design & Engineering Services December 2012
Figure 4 shows the process when an evaporative pre-cooler is added to the condenser
as follows:
Water is introduced to the evaporative media, through which the incoming
condenser air flows.
The air comes in contact with the media, evaporating some of the water.
Sensible heat is removed from this airstream, lowering its dry bulb temperature
and carrying the evaporated water away in the airstream.
The outdoor air temperature is lowered by 20 °F before entering the condenser.
OUTSIDE AIREVAPORATIVE MEDIA
CONDENSER COIL
CONDENSER FAN
WATER SUMP
SPRAY PUMP
PRE-COOLED AIR
SPRAY NOZZLES
DISCHARGE98° F
78° F
FIGURE 4. EVAPORATIVE PRE-COOLING CONDENSER AIR
Field Test of Hybrid Rooftop Unit HT.11.SCE.008
Southern California Edison Page 6
Design & Engineering Services December 2012
Figure 5 is a psychrometric chart that shows the evaporative cooling process, followed
by the heat rejection at the condenser.
10 15 20 25
30
35
40
45
50
55
55
60
60
ENTHALPY - BTU PER POUND OF DRY AIR
15
20
25
30
35
40
45
50
ENTH
ALP
Y -
BTU
PER P
OUND O
F DRY A
IR
SATU
RATIO
N T
EM
PERATU
RE -
°F
35
40
45
50
55
60
65
70
75
80
85
90
95
10
0
10
5
11
0
11
5
12
0
DR
Y B
UL
B T
EM
PE
RA
TU
RE
- °
F
.002
.004
.006
.008
.010
.012
.014
.016
.018
.020
.022
.024
.026
.028
10% RELATIVE HUMIDITY
20%
30%
40%
50%
60%
70%
80%
90%
35
3540
4045
45 50
50 55
55 60
6065
65
70
70
75
75
80
80
85 WET BULB TEMPERATURE - °F
85
90
12.5
13.0
13.5
14.0
VO
LU
ME
- CU
.FT
. PE
R L
B. D
RY
AIR
14.5
15.0
HU
MID
ITY
RA
TIO
- P
OU
ND
S M
OIS
TU
RE
PE
R P
OU
ND
DR
Y A
IR
OSA, Ontario, CA
Pre-Cool Discharge Condenser Disch
R R
ASHRAE PSYCHROMETRIC CHART NO.1
NORMAL TEMPERATURE
BAROMETRIC PRESSURE: 29.921 INCHES OF MERCURY
Copyright 1992
AMERICAN SOCIETY OF HEATING, REFRIGERATING AND AIR-CONDITIONING ENGINEERS, INC.
SEA LEVEL
0
1.0 1.0
-
2.0
4.08.0
-8.0-4.0-2.0
-1.0
-0.5-0.4-0.3-0.2
-0.1
0.1
0.2
0.3
0.4
0.5
0.6
0.8
-2000
-1000
0
500
1000
1500
2000
3000
5000
-
SENSIBLE HEAT Qs
TOTAL HEAT Qt
ENTHALPY
HUMIDITY RATIO
h
W
FIGURE 5. PSYCHROMETRIC CHART – EVAPORATIVE CONDENSER, ONTARIO, CA 1% DESIGN DRY-BULB
TEMPERATURE
Lowering the condensing temperature of an AC system’s refrigerant affects efficiency
in two ways. The lower condensing pressure corresponding to a lower temperature
reduces the work done by the compressor. Additionally, a lower condensing pressure
allows a greater proportion of the refrigerant to condense.
Appendix A provides detailed sample calculations of a refrigeration cycle using R-22,
which demonstrates a 39% improvement in COP attainable by reducing the condenser
air temperature from 100°F to 80°F.
DUAL-COOLING TECHNOLOGY The following process describes the dual-cooling technology used in this study:
Incoming condenser air passes through wetted media. The resulting evaporative
process cools both the incoming condenser air and the remaining water in the
media.
Field Test of Hybrid Rooftop Unit HT.11.SCE.008
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Design & Engineering Services December 2012
The cooled water drains into a sump, and is then pumped through a finned coil in
the AC unit’s OSA intake.
Sensible heat is rejected from the OSA into the water loop, lowering the OSA
temperature.
The warmed water is distributed through the evaporative media, and the cycle
starts over.
Figure 6 shows a simplified diagram of the process using representative temperatures.
OUTSIDE AIR EVAPORATIVE MEDIA
CONDENSER COIL
CONDENSER FAN
WATER SUMP
SPRAY PUMP
PRE-COOLED AIR
SPRAY NOZZLES
DISCHARGE98° F 80° F
OUTSIDE AIR98° FPRE-COOLED OUTSIDE AIR
88° F
OUTSIDE AIR PRE-COOLING COIL
FIGURE 6. DUAL-COOLING TECHNOLOGY
MARKET BARRIERS
Ongoing maintenance is a potential market barrier because evaporative processes
increase the amount of regular RTU monitoring and maintenance. RTUs are usually
located on the building roofs so they do not receive regular maintenance. Regular
maintenance may include water treatment measures to prevent scale and algae build-
up in the equipment, such as bleeding a portion of the system water to maintain the
level of dissolved solids low, chemical treatment, or other non-chemical measures.
Water usage is also a potential market barrier. In addition to the cost of the water
consumed in the evaporation process and the bleed process, there is an electrical
penalty associated with delivering water to the site and treating the bleed water in the
sewer system. This penalty becomes part of the site’s water cost.
Field Test of Hybrid Rooftop Unit HT.11.SCE.008
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Assessment Objectives The objective of this field assessment is as follows:
Assess the feasibility of retrofitting a dual-cooling system on existing air conditioning
systems at a big-box retail store.
Measure the electrical energy and electric demand savings provided by evaporatively
pre-cooling the condenser air intake of an RTU.
Measure the AC units’ increased cooling capacity when dual-cool technology is used.
Observe the OSA pre-cooling of the dual-cool technology.
Measure the water usage of an evaporatively pre-cooled condenser air system.
Measure the effects of operating the dual-cooling technology on the entire sales area
of a big box retail store.
Observe the water conditions in the dual-cooling system, including but not limited to,
scale build-up and algae growth.
Field Test of Hybrid Rooftop Unit HT.11.SCE.008
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Design & Engineering Services December 2012
Technology/Product Evaluation In this assessment, dual-cooling systems were installed on 13 existing RTUs serving a big box
retail store in Palmdale, California. The RTUs served the sales floor of this building. They were
put into service in 2008. We chose this site because the local climate provides a wide range of
summer outside air temperatures, which allows us to measure and apply results over a wide
range of climate zones in SCE territory.
The dual-cool system’s manufacturer furnished and installed the dual-cooling technology, the
water and waste piping systems, associated electrical work, and system commissioning.
Western Cooling Efficiency Center of Davis, California furnished, installed, and operated the
monitoring system used to measure the performance of the dual-cooling technology and the
baseline AC system.
Adding an OSA coil to the existing RTUs affects the OSA air flow. An air balance contractor
reset the AC units’ OSA back to their original flows.
Figure 7 shows the existing RTUs that are retrofitted with a dual-cooling system.
UNIT # MFGR MODEL #
NOMINAL
CAPACITY
(TONS)
COMPRESSOR
QTY.
OUTSIDE
AIR
(CFM)*
RTU-5 LENNOX LGC180H2BL 15 3 1185
RTU-7 LENNOX LGC180H2BL 15 3 1580
RTU-8 LENNOX LGC156H2BL 13 3 1220
RTU-9 LENNOX LGC240H2BL 20 4 1910
RTU-10 LENNOX LGC240H2BL 20 4 2065
RTU-11 LENNOX LGC210H2BL 17.5 4 1905
RTU-12 LENNOX LGC240H2BL 20 4 1885
RTU-13 LENNOX LGC210H2BL 17.5 4 1765
RTU-14 LENNOX LGC210H2BL 17.5 4 1835
RTU-15 LENNOX LGC210H2BL 17.5 4 1700
RTU-16 LENNOX LGC210H2BL 17.5 4 1760
RTU-17 LENNOX LGC210H2BL 17.5 4 **
RTU-18 LENNOX LGC210H2BL 17.5 4 1720
AIR CONDITIONING UNIT SCHEDULE
* Rounded to the nearest 5 CFM.
** OSA damper broken
FIGURE 7. ASSESSMENT SITE RTU SCHEDULE
Field Test of Hybrid Rooftop Unit HT.11.SCE.008
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Design & Engineering Services December 2012
Figure 8 shows the locations of the AC units on the roof.
FIGURE 8. ASSESSMENT SITE ROOF PLAN
Field Test of Hybrid Rooftop Unit HT.11.SCE.008
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Design & Engineering Services December 2012
Technical Approach/Test Methodology
FIELD TESTING OF TECHNOLOGY Ten RTUs, RTU-9 through RTU-19, serve the main sales floor, and two additional units.
RTU-7 and RTU-8, serve the check-out area. The grocery area is served by RTU-23,
which is excluded from this assessment. Each RTU is an air-cooled packaged AC unit,
with a constant speed supply fan, DX cooling coil, gas-fired furnace, filter section, and
OSA economizer. The DX cooling system contains three to four independent refrigerant
circuits. Each air conditioning unit discharged supply air into the store through one
four-way ceiling diffuser and returned air from one ceiling register. AC unit cooling and
heating is controlled by thermostats mounted on columns in the sales area.
The store operating hours are 8:00 AM to 10:00 PM, Monday through Saturday, and
8:00 AM to 9:00 PM on Sunday. AC unit on and off scheduling is provided by the
facility’s Energy Monitoring and Control System (EMCS).
The existing AC units are Lennox model #LGC with two cooling stages. When stage
one cooling is requested, refrigerant circuits 1 and 2 are energized. Second stage
cooling energizes the remaining one or two refrigerant circuits, depending upon unit
size.
The AC unit condensers are constructed in a V-configuration. As a result, a condenser
air pre-cooler could not be configured to serve condenser coils 3 and 4. These two coils
are used on a call for second stage cooling and receive uncooled incoming air. Figure 9
illustrates the condenser configuration of a typical Lennox AC unit serving the project
site and the pre-cooler side panel. CO
ND
ENSER
CIRCUIT #2
CON
DEN
SER
CIRCUIT #1
CON
DEN
SER
CIRC
UIT
#4
CON
DEN
SER
CIRC
UIT
#3
AIR FLOW, COOLING STAGE #2
(BOTH SIDES)
CO
ND
ENSER
AIR
P
RE-C
OO
LERAIR FLOW, COOLING STAGE #1
PRE-COOLER SIDE PANEL
(BOTH SIDES)
CONDENSER FANS
RTU CABINET
OSA PRE-COOLING COIL
FIGURE 9. LENNOX AC UNIT CONDENSER CONFIGURATION
A one-row pre-cooling coil is installed on the OSA intakes of each tested AC unit. This
coil increases the pressure drop across the OSA intake, so the OSA damper minimum
position is reset to provide the original designed CFM.
Field Test of Hybrid Rooftop Unit HT.11.SCE.008
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The dual-cooler has a circulating pump, which is controlled to circulate water through
the OSA pre-cooling coil and the evaporative media whenever the OSA temperature is
above the 75°F set point (adjustable). The pump operates even if the AC unit is off or
if there is no call for cooling. A float valve, located in the pre-cooler sump, opens when
the water level decreases. A drain valve was manually adjusted to bleed a set amount
of water from the system whenever the circulating pump is running.
TEST PLAN
The field test evaluates baseline operation and dual-cooling technology as follows:
Comparing electrical energy, refrigerant liquid temperature, refrigerant
compressor pressure lift, and water usage of units RTU-7 and RTU-8.
Compressor pressure lift is the difference between the suction and discharge
pressures of the refrigerant compressors; the higher the lift, the harder the
compressor works.
Comparing the electrical energy of units RTU-10 and RTU-11.
Observing the electrical consumption of the entire site.
For both baseline and technology performance, the AC systems were monitored during
the store’s operating hours, when OSA temperature is between 80°F and 100°F.
Air conditioning unit system energy usage is compared for the following OSA dry bulb
temperature bins:
80º – 85º F
85º – 90º F
90º – 95º F
95º – 100º F
For each test, the following measurements were recorded:
OSA temperature, dry bulb and wet bulb
System voltage
System kW and kVa
Compressor kW, cooling stages 1 and 2
Condenser fan kW
Return air temperature, dry bulb and wet bulb
Supply air temperature, dry bulb and wet bulb
Pre-cooler circulating pump kW
Pre-cooler water temperature, inlet
Pre-cooler water temperature, outlet
Make-up water consumption*
Refrigerant suction pressures and temperatures*
Refrigerant compressor pressures and temperatures*
Refrigerant condensing temperature*
* - RTU-7 and RTU-8
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Figure 10 shows the general location of the points monitored for each air conditioning
unit. OSA temperatures were measured at one point, for all air conditioning units.
Return Duct
Compressor 4
Evaporator Coil
Supply Fan
Condenser FansSupply Duct
OSA
Supply Air Dry
Bulb Temp
Return Air Dry
Bulb Temp
kW
Compressor 3
Compressor 2Compressor 1
kW
OSA PRE-COOLER
OSA Dry Bulb
Temp
OSA Wet Bulb
Temp
Discharge Air Dry
Bulb Temp
Discharge Air Wet
Bulb Temp
Return Air Wet
Bulb Temp
Supply Air Wet
Bulb Temp
Water Temp Out
Water Temp In
F Make-up Water
Gallons
Unit Voltage
EVAPORATIVE PRE-COOLER
kW
CIRCULATING PUMP
kWsystem
kW
#1 #2 #3 #4
Condenser Fans
kVasystem
PSUCTION1
PSUCTION2
PSUCTION3
PSUCTION4
TSUCTION1
TSUCTION2
TSUCTION3
TSUCTION4
PDISCH1
PDISCH2
PDISCH3
PDISCH4
TCOND1
TCOND2
TCOND3
TCOND4
TDISCH1
TDISCH2
TDISCH3
TDISCH4
FIGURE 10. MONITORING POINTS
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The effects of the evaporative pre-cooling system on the RTU cooling performance
were measured in the following ways:
Saturated condensing temperature.
Pressure lift or the difference in suction and discharge pressure of the
compressors.
Overall RTU power, in kW.
Compressor run-time, 1st and 2nd stages of cooling.
Compressor amperage, 1st and 2nd stages of cooling.
Dual-cooling water usage is determined by metering the main water supply to all dual-
cooling units, as well as separate sub-meters to four individual units, for the duration
of the test.
The electrical distribution of the test site did not allow our monitoring equipment to
measure the electrical usage of all of the AC units, so the overall effect on the store of
the dual-cooling system was observed by reviewing 15 minute building meter data.
Water conditions, including algae growth and water scale formation, are visually
observed and recorded during the test.
INSTRUMENTATION PLAN
Appendix B provides the list of monitored points, sensors, and sensor accuracy. Data is
measured and recorded once per minute.
A control module is installed in each RTU to read the monitored points and record
data. These data are uploaded to WCEC’s project engineer.
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Results SCE operated and monitored the dual-cooling technology on the assessment site from
September 1 to September 28, 2012. During this period, outside air dry bulb temperatures
reached 100°F. On the morning of September 28, 2012, the dual-cooling system was disabled
as follows:
Turned off the circulating pumps.
Turned off the water supplies.
Drained the water sumps.
Removed the evaporative media. This step was necessary to prevent an airflow
restriction that results in lower condenser air flow and/or higher condenser fan energy
compared to baseline conditions.
The baseline AC system is operated and monitored from September 28 through October 26, 2012. During this period, outside air dry bulb temperatures also reached 100°F.
DATA ANALYSIS
CONDENSER AIR PRE-COOLING
The amount of condenser air pre-cooling provided by the technology is indirectly
measured by measuring the DX system SCT and condensing pressure.
Table 1 summarizes the observed and measured Saturated Condensing Temperature
(SCT) of the RTU-7 and RTU-8 refrigerant circuits.
TABLE 1. REFRIGERANT LIQUID TEMPERATURE, AC-7 AND AC-8
INCOMING OSA
TEMPERATURE
(°F)
AC-7 AC-8
OUTLET TEMP.
(°F)
OUTLET TEMP.
(°F)
BASELINE MEASURE ∆TEMP BASELINE MEASURE ∆TEMP
80°F - 85°F 101 86 15 101 97 4
85°F - 90°F 106 87 19 112 103 9
90°F - 95°F 111 91 20 118 100 18
95°F - 100°F 115 95 20 120 100 20
Several observations are made from the data collected as follows:
The second stage of cooling for AC-7 and AC-8 never energized.
The first stage of cooling of AC-8 did not energize for more than a few hours during
the baseline test.
Monitoring of AC-8 was interrupted on October 14, 2012.
SCE calculated compressor lift for the active refrigeration systems by subtracting the
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measured discharge pressure from the measured suction pressure. Table 2 summarizes the
compressor lifts for baseline and post measure operation.
TABLE 2. REFRIGERANT COMPRESSOR LIFT, AC-7
OSA TEMPERATURE
COMPRESSOR LIFT
(PSI)
BASELINE MEASURE % DIFF.
80°F - 85°F 181 147 19
85°F - 90°F 195 146 25
90°F - 95°F 210 156 26
95°F - 100°F 220 166 24
For a constant cooling load, increased compressor capacity is reflected by reduced
compressor run times. Table 3 and Table 4 show the compressor run times for the three
compressor AC units, before and after evaporative cooling.
TABLE 3. COOLING STAGES, BASELINE, 3 COMPRESSOR AC UNIT
OSA TEMPERATURE, DRY BULB
COOLING STAGES RUN TIME, %
AC-7 AC-8
NO
COOLING
STAGE
#1
STAGE
#2
NO
COOLING
STAGE
#1
STAGE
#2
80 - 85°F 50 50 0 97 3 0
85 - 90°F 0 100 0 95 5 0
90 - 95°F 0 100 0 98 2 0
95 - 100°F 0 100 0 89 11 0
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TABLE 4. COOLING STAGES, EVAPORATIVE PRE-COOLING, 3 COMPRESSOR AC UNIT
OSA DRY BULB TEMPERATURE
COOLING STAGES RUN TIME, %
AC-7 AC-8
NO
COOLING
STAGE
#1
STAGE
#2 NO
COOLING
STAGE
#1
STAGE
#2
80 - 85°F 0 100 0 94 5 1
85 - 90°F 0 100 0 86 14 0
90 - 95°F 0 100 0 74 26 0
95 - 100°F 0 100 0 46 54 0
Table 5 and Table 6 show the compressor run times for the four compressor AC units before
and after evaporative cooling.
TABLE 5. COOLING STAGES, BASELINE, 4 COMPRESSOR AC UNIT
OSA DRY BULB TEMPERATURE
COOLING STAGES RUN TIME, %
AC-10 AC-11
NO
COOLING
STAGE
#1
STAGE
#2 NO
COOLING
STAGE
#1
STAGE
#2
80 - 85°F 93 7 0 14 86 0
85 - 90°F 86 14 0 1 99 0
90 - 95°F 67 33 0 0 100 0
95 - 100°F 76 24 0 0 100 0
TABLE 6. COOLING STAGES, EVAPORATIVE PRE-COOLING, 4 COMPRESSOR AC UNIT
OSA DRY BULB TEMPERATURE
COOLING STAGES RUN TIME, %
AC-10 AC-11
NO
COOLING
STAGE
#1
STAGE
#2 NO
COOLING
STAGE
#1
STAGE
#2
80 - 85°F 92 8 0 15 85 0
85 - 90°F 86 14 0 1 99 0
90 - 95°F 86 14 0 0 100 0
95 - 100°F 69 31 0 0 100 0
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The dual-cooling system’s circulating pumps added an electrical load to the system. In the
application tested, the pumps ran continuously whenever the outdoor air temperature
exceeded 75°F, regardless of the call for cooling or if the RTU was operating. Each pump
consumed 1 amp power @ 115 volts, based on monitored data. These results are shown in
Table 7.
TABLE 7. DUAL-COOLING SYSTEM PUMP POWER CONSUMPTION
AC UNIT PUMP CURRENT (AMPS) PUMP POWER (KW) PUMP OPERATING HOURS
AC-7 1 .12 62%
AC-8 1 .12 70%
AC-10 1 .12 70%
AC-11 1 .12 59%
We are not able to compare RTU energy usage in kW for baseline and measure conditions
because of errors in data output. We are investigating these errors.
OVERALL ENERGY IMPACT
SCE completed a general review of 15-minute electrical utility data. We compared the
baseline air-cooled operation to dual-cooled operation at OSA dry bulb temperatures ranging from 80 to 100°F. Table 8 summarizes the overall performance, based on dry
bulb temperature.
The evaporative systems serving RTU-15 and RTU-18 did not operate during the test
period. If these systems were operating correctly, overall building energy usage may
have been less.
TABLE 8. SITE OVERALL ENERGY DEMAND, USING DRY BULB TEMPERATURE
SITE ELECTRIC METER DEMAND (KW)
OSA DRY BULB
TEMPERATURE (°F) BASELINE DUAL-COOL ∆ DECREASE
(%)
80 - 85 483 471 12 2.4
85 - 90 521 505 16 3.0
90 - 95 541 517 25 4.6
95 - 100 568 552 16 2.7
WATER USAGE
SCE placed water meters on the main water supply to all of the dual cool units. Water
meters are also placed on the individual water supplies to RTU-7, RTU-8, RTU-10, and
RTU-11. These meters send a signal to the on-site monitoring system and also have a
physical counter.
We did not monitor water consumption because of problems with the monitoring
system. Additionally, the installation team did not record the initial values on the flow
meters’ counters, so we were unable to use this method to measure water use.
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WATER TREATMENT
After one month of dual cooling operation, SCE removed the evaporative media and
visually observed scale formation and algae growth. The evaporative media surfaces
show minor amounts of precipitated solids on ¼ of the RTUs, but no precipitates on
evaporative media surfaces of the remaining units. Figure 11 shows the extent of scale
formation on the outside face and the top of the evaporative media.
FIGURE 11. SCALE FORMATION ON EVAPORATIVE MEDIA
Algae growth is observed in the sump of the dual-cooling system. Figure 12 indicates
the level of growth after one month of operation without water treatment. These
observations were shared with the dual-cool system’s manufacturer.
FIGURE 12. ALGAE GROWTH IN SUMP
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Discussion Observations are made regarding the ease of installation of the evaporative pre-cooling
systems, their operation, measured performance improvements, and water treatment.
SYSTEM INSTALLATION AND OPERATION The evaporative pre-cooling system was installed on the 13 RTUs at the project site in
4 days, using a qualified mechanical contractor. The contractor and manufacturer
surveyed the site for a day prior to this installation. This work included the installation
of an entire water distribution system on the roof to serve these units and drainage
connections to the existing condensate piping. There was one problem during the
installation process. Due to confusion between sub-contractors, the evaporative pre-
cooling systems serving RTU-15 and RTU-18 were not placed in operation.
The remaining systems operated during the test period with no problems. SCE
interviewed the facility manager several times, and he did not report any complaints.
We did notice that three units leaked a small amount of water onto the roof. These
leaks were reported to the manufacturer for repair.
A constant amount of water was bled from each system, to maintain the total
dissolved solids level of the system water below a point where solids would precipitate
onto the evaporative media. This bleed discharged into the AC unit’s condensate drain.
The condensate drains for all of the RTUs drained into a common piping system, which
discharged indirectly into a mop sink at the rear of the building. Because the weather
in Palmdale, CA is arid, the amount of condensate draining into the sink is normally
minimal. However, the addition of the bleeds from each evaporative system created a
constant flow of water into the mop sink. The mop sink was able to accommodate this
flow, even when debris from other sources partially blocked the sink’s strainer.
The circulating pumps for the evaporative systems operated continuously when the
outside air temperature was above 75° F. The pumps did not have capability to shut
off when the RTUs were not operating.
The evaporative pre-cooling system and the water piping system serving it are
vulnerable to freezing conditions. For example, outside air temperature sensors
recorded sub-zero temperatures for an eight hour period on the morning of November
12, 2012. For this reason, SCE decommissioned the system before the threat of
exposure to sub-zero temperatures. SCE drained the water from the systems’ sumps,
piping, and the exposed water piping on the roof.
ENERGY SAVINGS The energy savings are as follows:
Refrigeration saturated condensing temperatures are reduced by 20% when
evaporative pre-cooling is applied to RTU-7 and RTU-8, at outdoor dry bulb
temperatures from 90° - 100°. Smaller SCT reductions are measured on these units
@ OSA dry bulb temperatures between 80° - 90°.
DX system pressure lift is reduced by 25% in RTU-7, when evaporative pre-cooling
is applied. The RTU-8 DX system did not operate long enough during baseline
monitoring to provide results for comparison.
The evaporative pre-cooling system did not reduce compressor run-time for the
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four RTUs monitored in this test. The measured run-time in cooling stage 1 for
RTU-7, 8, 10, and 11 are almost unchanged from baseline to measure.
The cooling stage 1 for RTU-7 and RTU-11 operated almost continuously, while the
cooling stages of adjacent units RTU-8 and RTU-10 did not energize often.
RTU-7 and RTU-8 serve almost identical cooling loads in the front check-out area of
the store. RTU-8 is closer to the entry doors and theoretically subject to a higher
cooling load.
RTU-10 and RTU-11 served similar areas of the sales floor and are expected to
have similar cooling loads.
Errors in the data output do not allow us to compare the overall RTU energy usage,
between measure and baseline.
The system tested in this study included an OSA pre-cooling coil, which was designed
to lower the sensible cooling load imposed upon the RTUs. In this facility, the OSA
cooling load is a large proportion of the overall cooling load. Other components of the
cooling load are people, lighting, wall and roof loads. The retail store has minimal glass
area and is built to 2008 energy codes, so the exterior wall and roof do not contribute
much to the overall cooling load.
Separately testing the effects of the evaporative pre-cooling and the OSA sensible pre-
cooling is difficult because it’s not possible to operate the condenser air pre-cooler
without the OSA pre-cooler, unless bypass piping and valves are added to the system.
The OSA pre-cooling does not affect the SCT or the discharge pressure of the
refrigerant system, so these two variables can be studied to isolate the effects of the
condenser air pre-cooler.
A study of overall energy usage derived from applying evaporative pre-cooling to all
RTUs serving the sales floor shows a 3% savings. This result is affected by two factors
as follows:
Space cooling represents approximately 10% of the overall electrical consumption
in a big box retail store that has supermarket refrigeration. While other major
components of the electrical consumption, like lighting, television displays, and in-
store product cooler remain fairly constant, there is enough variance in overall load
to question whether the observed 3% reduction is statistically significant.
The evaporative pre-coolers of RTU-15 and RTU-18, did not operate correctly. If
they had been operating correctly, overall energy savings may have been higher.
WATER TREATMENT Packaged air-cooled RTUs require a minimal amount of service to provide satisfactory
space temperature control. These systems may operate inefficiently for long time
periods because they are not serviced until the occupants complain that the space is
too hot or cold.
HVAC and refrigeration systems that use evaporative cooling have a higher level of
continuous maintenance. For example, water contains dissolved solids, which remain
after a portion of the water is evaporated. The concentration of these solids increases
and, if not addressed, reaches a level where they cannot remain dissolved. At that
point, the solids precipitate out as scale on solid surfaces. After four weeks of
operation, we saw solids on the evaporative media of four of the systems tested on
this site. Figure 11 shows an example of scale formation on the evaporative media.
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To prevent scale formation, we used a constant water bleed to allow a drain valve to
release a constant flow of water from the system sump. This water is replaced by fresh
incoming water, which dilutes the concentration of solids. The proportion of bleed
water required to prevent scaling depends on the concentration of dissolved solids in
the incoming water. In larger evaporative systems, like cooling towers, a conductivity
meter is used to measure the level of dissolved solids in water. This meter
automatically opens a bleed valve when the solids level reaches setpoint. That type of
system provides better control of water loss and dissolved solids level than a constant
bleed.
In this study, we observed scaling on three of the 11 units even though each unit
received the water of the same quality and operated under the same conditions. It is
possible that the three units had a smaller bleed rate than the others.
Biological growth forms in water basins. For evaporative systems, the basin water is
normally 80° to 100°F, which provides a hospitable environment for biological growth.
The average sump water temperature of the operating pre-cooling system is 72°F,
which is lower than what is typically observed in open cooling tower basins. We also
observed algae growth in the basins of two operating systems, with the largest
concentration in non-operating RTU-15. It is possible that RTU-15 provided a better
environment for algae growth because the evaporative system did not work, which
allowed the basin water to heat to a higher temperature.
Biological growth in the water of evaporative systems is controlled by several methods,
including injecting biocides into the water or imposing an electrical charge into the
water stream. The system we studied did not provide a way to control biological
growth.
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Recommendations The results of this study suggest that SCE’s EE program adopt this technology, but we
recommend that SCE monitor the test site during hot summer months to provide a conclusive
recommendation.
Recommended Phase 2 work includes the following:
Address the issues in the monitoring system, collect data through summer 2013 and
finalize the study.
Improve the measurement of water consumption, RTU electrical energy consumption,
and compressor electrical consumption to provide definitive results.
Calculate unit EER with dual-cooling technology.
Review the results of other lab testing and field assessments of this technology,
performed by both SCE and PG&E as follows:
SCEs recently completed laboratory test of a new Trane RTU.
PG&E reports of field assessments performed in the summer of 2012.
SCE’s installation of two new Trane RTUs in a shopping mall office in Ontario, CA.
Develop an eQuest computer model of electrical savings provided by evaporatively
pre-cooling condenser air for RTUs, serving different commercial building types in
various climate zones. Calibrate this model based on actual measured results obtained
from this field assessment.
Study reliable, inexpensive water treatment systems for the proposed evaporative
technology. Air-cooled RTUs are popular choices for commercial HVAC systems
because they require less continual maintenance. In the last few years, commercial
RTU manufacturers have introduced larger (60 ton capacity) RTUs into the market,
which have evaporative condensers instead of air-cooled condensers. These products
feature a water treatment system that controls biological growth and scale formation.
This technology can be considered for use with evaporative pre-coolers.
Study the effectiveness of OSA sensible pre-coolers. The challenge is providing a
sensing method for measuring the leaving air temperature from this pre-cooling coil.
This is a challenge because the return air and OSA paths of an RTU typically mix soon
after the OSA intake damper. This makes it difficult to isolate the OSA temperature
downstream of a pre-cooling coil for measurement.
Provide feedback to manufacturer regarding system controls, water treatment, and
winterizing.
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Conclusions Adding evaporative pre-coolers to the condenser air stream of RTUs provides a significant
20% to 25% reduction in DX system saturated condensing pressure and temperature. This
reduction occurs when the OSA temperatures are 80° to 100°F, which is when peak cooling
loads and energy consumption occur. Operating this system on most of the RTUs that serve a
large retail facility provide a measurable electrical energy savings.
Based on the results of this study, we should consider evaporative pre-cooling for air-cooled
RTUs in California. Installing this technology is relatively simple, but sites should consider the
following operational issues:
Increased maintenance costs for water treatment
The ability of the existing facility’s cold water system to serve the added equipment.
A suitable location to drain bleed water
The requirement to winterize the evaporative system to prevent freeze damage, in
some locations
Controls that prevent pump operation and water bleed when the RTU is not in service
Difficulty with the monitoring system and the data produced by it prevented an analysis of the
following aspects of this technology:
Water consumption.
Actual electrical energy savings for individual RTUs
Electrical energy reduction of the DX compressors
Sensible cooling provided by the OSA pre-cooling coil
Corrections to the monitoring system used in this project and a longer test period will provide
better data analysis for this technology.
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Appendix A – Sample Refrigeration
Calculations Lowering the condensing temperature of an AC system’s refrigerant affects efficiency as
follows:
The lower condensing pressure corresponds to a lower temperature which reduces the
work done by the compressor
The lower condensing pressure allows a greater proportion of the refrigerant to
condense.
Figure 13 diagrams the cycle for a theoretical R-22 refrigeration system. The red lines indicate the cycle @ SCT = 120°F, and the blue lines indicate the cycles @ SCT = 100°F. This
diagram graphically represents the increase in cooling capacity and decrease in compressor
work associated with reducing the system’s condensing temperature. The higher SCT corresponds to an air-cooled AC unit operating @ 100°F OSA, while the lower SCT
corresponds to an AC unit with pre-cooling lowering entering condenser air temperature to 80°F.
FIGURE 13. PRESSURE-ENTHALPY CHART FOR R-22 REFRIGERANT
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Equation 1 calculates the heat rejected in the process of evaporating refrigerant in an ideal
refrigeration cycle.
EQUATION 1. EVAPORATOR HEAT REJECTION
Where: = change in enthalpy, evaporator,
= Enthalpy, leaving refrigerant
= Enthalpy, entering refrigerant
Equation 2 applies the formula in Equation 1 to a system with 40°F SST and 120°F SCT.
EQUATION 2. EVAPORATOR HEAT REJECTION, R-22 REFRIGERANT, 40°F SST, 120°F SCT
Equation 3 applies the formula in Equation 1 to a system with to 40°F SST and 100°F SCT.
EQUATION 3 .EVAPORATOR HEAT REJECTION, R-22 REFRIGERANT, 40°F SST, 100°F SCT
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Equation 4 calculates the increase in cooling capacity for a theoretical R-22 system, when
decreasing the SCT from 120°F to 100°F.
EQUATION 4. COOLING CAPACITY INCREASE, R-22 REFRIGERANT, 120°F TO 100°F SCT
Equation 5 calculates the work performed by the compressor in an ideal refrigeration cycle.
EQUATION 5. COMPRESSOR WORK
Where: = change in enthalpy, compressor,
= Enthalpy, leaving compressor
= Enthalpy, entering compressor
Equation 6 applies the formula in Equation 5 to a system with 40°F SST and 120°F SCT.
EQUATION 6. COMPRESSOR WORK, R-22 REFRIGERANT, 40°F SST, 120°F SCT
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Equation 7 applies the formula in Equation 5 to a system with 40°F SST and 100°F SCT.
EQUATION 7. COMPRESSOR WORK, R-22 REFRIGERANT, 40°F SST, 100°F SCT
Equation 8 calculates the cooling work reduction, R-22 refrigerant, 120°F to 100°F SCT.
EQUATION 8. COOLING WORK REDUCTION, R-22 REFRIGERANT, 120°F TO 100°F SCT
The coefficient of performance (COP) of a refrigeration system is the ratio of cooling provided
to work input. Equation 9 calculates the COP for a theoretical refrigeration system.
EQUATION 9. REFRIGERATION CYCLE COEFFICIENT OF PERFORMANCE
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Equation 10 calculates the COP for an R-22 refrigeration system @ 40°F SST and 120°F SCT.
EQUATION 10. COP, R-22 REFRIGERANT, 40°F SST, 120°F SCT
Equation 11 calculates the COP for the system in Equation 10 with 100°F SCT, and the
increase in performance.
EQUATION 11. COP, R-22 REFRIGERANT, 40°F SST, 100°F SCT
Equation 12 calculates the COP improvement when reducing SCT from 120°F to 100°F.
EQUATION 12. COP INCREASE, R-22 REFRIGERANT, 120°F TO 100°F SCT
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Appendix B – Monitoring Equipment Table 9 provides information about the monitoring equipment used in this study.
TABLE 9. MONITORING EQUIPMENT
SENSOR TYPE MAKE/MODEL ACCURACY RTUS
TOSA Vaisala HUMICAP HMP110 ±0.36°F 7,8,20
RHOSA Vaisala HUMICAP HMP110 ±1.7% RH 7,8,20
TRA Vaisala HUMICAP HMP110 ±0.36°F 7,8,10,11,20,21
RHRA Vaisala HUMICAP HMP110 ±1.7% RH 7,8,10,11,20,21
TSA Vaisala HUMICAP HMP110 ±0.36°F 7,8,10,11,20,21
RHSA Vaisala HUMICAP HMP110 ±1.7% RH 7,8,10,11,20,21
∆PSA Dwyer 0-1.0 “WC = 4-20 mA 7,8,10,11,20,21
∆POSA Dwyer 0-0.25 “WC = 4-20 mA 7,8,10,11
WATER OMEGA FTB 4105 A P ±2% 7,8,10,11
MAIN OMEGA FTB8010B PR ±2% 20
OSA Position RA/OSA Damper Actuator 0–10 Vdc NC 7,8,10,11,20,21
±0.06%CTC1 100 Ω SA1-RT-B ±0.06% 7,8,20
CTC2 100 Ω SA1-RT-B ±0.06% 7,8,20
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SENSOR TYPE MAKE/MODEL ACCURACY RTUS
CTC3 100 Ω SA1-RT-B ±0.06% 7,8,20
CTC4 100 Ω SA1-RT-B ±0.06% NA
TLOW, C1 100 Ω SA1-RT-B ±0.06% 7,8,20
TLOW, C2 100 Ω SA1-RT-B ±0.06% 7,8,20
TLOW, C3 100 Ω SA1-RT-B ±0.06% 7,8,20
TCD OUT 3 W 100 Ω SA1-RT-B ±0.06% 8.20
THI, C1 100 Ω SA1-RT-B ±0.06% 7,8,20
THI, C2 100 Ω SA1-RT-B ±0.06% 7,8,20
THI, C3 100 Ω SA1-RT-B ±0.06% 7,8,20
TCD OUT 3 WO 100 Ω SA1-RT-B ±0.06% 8,20
PLOW, C1 ClimaCheck 200200, 10bar ±1% 7,8,20
PLOW, C2 ClimaCheck 200200, 10bar ±1% 7,8,20
PLOW, C3 ClimaCheck 200200, 10bar ±1% 7,8,20
PLOW, C4 ClimaCheck 200200, 10bar ±1% NA
PHI, C1 ClimaCheck 200100, 35bar ±1% 7,8,20
PHI, C2 ClimaCheck 200100, 35bar ±1% 7,8,20
PHI, C3 ClimaCheck 200100, 35bar ±1% 7,8,20
Field Test of Hybrid Rooftop Unit HT.11.SCE.008
Southern California Edison Page 32
Design & Engineering Services December 2012
SENSOR TYPE MAKE/MODEL ACCURACY RTUS
PHI, C4 ClimaCheck 200100, 35bar ±1% NA
CTPUMP NK AT1-005-000-SP AC current transducer to 0-5
Vdc
7,8,10,11,
T CD OUT 1 100 Ω SA1-RT-B ±0.06% 7,8,20
T CD OUT 2 100 Ω SA1-RT-B ±0.06% 7,8,20
T CD OUT 3 100 Ω SA1-RT-B ±0.06% 7,8,20
T CD OUT 4 100 Ω SA1-RT-B ±0.06% NA
T SUMP Thermocouple Type T Place below low water level 7,8,10,11
T WC IN Thermocouple Type T Insulate 7,8,10,11
TWC OUT Thermocouple Type T Insulate 7,8,10,11
KWSYSTEM Dent Powerscout 3 RS485 connection to
dataTaker
7,8,10,11,20,21
CTCF 1&2 NK AT1-005-000-SP AC current transducer to 0-5
Vdc
NA
CTCF 3&4 NK AT1-005-000-SP AC current transducer to 0-5
Vdc
NA
Field Test of Hybrid Rooftop Unit HT.11.SCE.008
Southern California Edison Page 33
Design & Engineering Services December 2012
References [1] Itron, Inc., "California Commercial End-Use Survey," California Energy Commission,
Sacramento, 2006.
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