sce design and engineering services · the operational and monitoring system issues on this project...

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 [email protected].

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.


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.


Southern California Edison Page iv

Design & Engineering Services December 2012

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


Southern California Edison Page v


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


Southern California Edison Page vi Design & Engineering Services December 2012

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


Southern California Edison Page vii Design & Engineering Services December 2012

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


Southern California Edison Page 1


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.




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




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




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




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




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.




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.




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.




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-11 LENNOX LGC210H2BL 17.5 4 1905






RTU-17 LENNOX LGC210H2BL 17.5 4 **


AIR CONDITIONING UNIT SCHEDULE

* Rounded to the nearest 5 CFM.

** OSA damper broken

FIGURE 7. ASSESSMENT SITE RTU SCHEDULE




Figure 8 shows the locations of the AC units on the roof.

FIGURE 8. ASSESSMENT SITE ROOF PLAN




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.




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




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




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.




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




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




TABLE 4. COOLING STAGES, EVAPORATIVE PRE-COOLING, 3 COMPRESSOR AC UNIT

OSA DRY BULB TEMPERATURE


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



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



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




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.




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




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




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.




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.




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.




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.




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




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




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. COMPRESSOR WORK, R-22 REFRIGERANT, 40°F SST, 120°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




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




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





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, C4 ClimaCheck 200200, 10bar ±1% NA

PHI, C1 ClimaCheck 200100, 35bar ±1% 7,8,20







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




References [1] Itron, Inc., "California Commercial End-Use Survey," California Energy Commission,

Sacramento, 2006.

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