Evaporative and AirCooled Report.pdf

7/29/2019 Evaporative and AirCooled Report.pdf

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PERFORMANCE

EVALUATION

OFAIRCOOLED ANDEVAPORATIVECONDENSERS

Southern California EdisonRefrigeration Technology and Test Center

Energy Efficiency Division

March 4, 1998

Refrigeration Technology and Test Center


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DISCLAIMER

THIS WORK (WORK) WAS PERFORMED WITH REASONABLE CARE AND INACCORDANCE WITH PROFESSIONAL STANDARDS. HOWEVER, NEITHER SCE NOR ANYENTITY PERFORMING THE WORK PURSUANT TO SCES AUTHORITY MAKE ANYWARRANTY OR REPRESENTATION, EXPRESSED OR IMPLIED, WITH REGARD TO THISREPORT, THE MERCHANDABILITY OR FITNESS FOR A PARTICULAR PURPOSE OF THERESULTS OF THE WORK, OR ANY ANALYSES, OR CONCLUSIONS CONTAINED IN THISREPORT.

THE RESULTS REFLECTED IN THE WORK ARE BASED GENERALLY REPRESENTATIVE OFOPERATING CONDITIONS; HOWEVER, THE RESULTS IN ANY OTHER SITUATION MAYVARY DEPENDING UPON PARTICULAR OPERATING CONDITIONS.


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ACKNOWLEDGMENT

We would like to thank VaCom Technologies for assisting Southern California Edison (SCE) inconducting this test. Additionally, we would like to extend our thanks to Recold Division ofMarley Cooling Towers, Bohn Refrigeration Division of Heatcraft, Inc., and ABB for theirgenerous contributions in this project.

The entire data reduction, engineering analysis and report production in this project wasconducted by the staff of the Energy Efficiency Division of SCE.


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REFRIGERATION TECHNOLOGY AND TEST CENTERPERFORMANCE EVALUATION OF AIR COOLED AND EVAPORATIVE CONDENSERS

TABLE OF CONTENTS

Executive Summary................................................................................................................................... ii

List of Tables............................................................................................................................................ iv

List of Figures........................................................................................................................................... iv

Nomenclature..........................................................................................................................................viii

Introduction.................................................................................................................................................1

Report Organization....................................................................................................................................4

(1) Test Procedure......................................................................................................................................6

(2) Analysis................................................................................................................................................8

(3) Heat Rejection by Air Cooled Condenser..........................................................................................13

(4) Air Cooled Condenser Tests...............................................................................................................16

(5) Heat Rejection by Evaporative Condenser.........................................................................................25

(6) Evaporative Condenser Tests.............................................................................................................28

(7) Discussion of Results.........................................................................................................................39

(8) Conclusions........................................................................................................................................49


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ii

EXECUTIVE SUMMARY

The purpose of this project was to test and evaluate the impact of several common control strategies onthe heat rejection capacity and power consumption of air cooled and evaporative condensers. The testalso evaluated the effect of these control strategies on the condenser and refrigeration system efficiencies.

The test was conducted under actual climate conditions of Irwindale, CA. Throughout the entire thirteen

days of the test, the refrigeration load remained constant at seven tons while the main focus was on effectsof ambient conditions and control strategies on the condenser performance and power use.

Southern California Edison (SCE) conducted this test at its state-of-the-art Refrigeration Technology andTest Center (RTTC), located in Irwindale, CA. The RTTCs sophisticated instrumentation and dataacquisition system provided detailed tracking of the refrigeration systems critical temperature andpressure points during the test period. These readings were then utilized to quantify various heat transferand power related parameters within the refrigeration cycle.

In supermarkets refrigeration systems, compressors and condensers contribute a large portion of thesites electric energy use. Compressors can consume up to eight times more power than condensers incommercial supermarkets. In this test, however, the condenser comprised a larger than typical portion of

the power use. Hence, using the system and compressor power and efficiency results of this test in anabsolute sense will be misleading.

Optimizing the trade off between a compressors reduced power consumption and additional condenserfan energy consumption at lower head pressures has always been a challenge. This test attempted todemonstrate this trade off through the implementation of different fan control strategies. The mostcommon condenser fan control strategies for maintaining a fixed or floating compressor head are:

Two speed fans Staging of various size fans (fan cycling) Variable Speed Drive (VSD)

As a result, to address a wide range of fan and head pressure control combinations, this test evaluated theperformance of evaporative and air cooled condensers under the following scenarios:

1. Fixed condensing temperature at 90F, using VSD control on fan motor


3. Fixed condensing temperature at 90F, using two speed fan motors / fan cycling

4. Fixed condensing temperature at 65F, using two speed fan motors / fan cycling5. Variable condensing temperature set point, using two speed fan motors / fan cycling6. Variable condensing temperature set point, using VSD control on fan motor

In addition to the above scenarios, the evaporative condenser was tested with its fan operating at 80% ofits capacity at constant speed. The purpose of this test was to capture the effects of smaller sized fansoperating at maximum fixed speed on the condensing temperature of the system.


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Based on the tested control scenarios; the results indicate that except for one control scenario theevaporative condenser achieves better Energy Efficiency Ratios (EER) under various control modes(Figure 1). The EER simply reflects the ratio of the heat of refrigeration rejected by the condenser

(Btu/hr) to the condenser power input (Watts). The variable speed drive control with a 90F target SCTwas the only mode of control under which the air cooled condensers EER was higher than that of the

evaporative condenser. For the air cooled condenser, the variable speed drive fan control at 90F targetSCT required the lowest average condenser power use while yielding the highest SCT. The variablespeed fan control with variable set point achieved the next highest EER.

For the evaporative condenser, the constant speed operation of the fan at 80% of its rated capacity yielded

the most favorable results. Except for the 90F SCT tests, this mode of test produced the lowest SCTs,while the ambient wet bulb was the highest and the least amount of average condenser power wasrequired. The variable speed fan control with variable set point achieved the next highest EER.

The main difference between the two condensers, however, remained in the yielded SCTs and condenserpower use. Overall, the evaporative condenser achieved lower SCTs and consumed less power than theair cooled condenser.

Average Condenser Energy Efficiency Ratings

23.51

179.48

83.92

29.6623.3726.5928.1442.88

87.48

29.8634.7129.64

0

50

100

150

200

2 65 2 90 2 VSP VS 65 VS 90 VS VSP

EER(Btu/h/W)

Air Cooled Evaporative

Figure 1 - Average Energy Efficiency Ratings for Air Cooled and Evaporative Condensers


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iv

LIST OF TABLES

Table 1 - Average Condenser EER Values (Btu/h/W) for Various Test Modes.........................................................3

Table 2 - Various Condenser Performance Parameters per Amount of Heat Rejected..............................................3

Table 3 - Test Runs for the Evaporative Condenser...................................................................................................6

Table 4 - Test Runs for the Air Cooled Condenser ....................................................................................................7

LIST OF FIGURES

Figure 1 - Average Energy Efficiency Ratings for Air Cooled and Evaporative Condensers...................................iii

Figure 2 - Power Use Breakdown for Typical Supermarket.......................................................................................1

Figure 3 - Ambient Temperature Fluctuation for Entire Test Period..........................................................................2

Figure 4 - Schematic of an Air Cooled Condenser...................................................................................................13

Figure 5 - Variation of Condensing Coefficient With Distance Along Condenser Tube.........................................14

Figure 6 - Air Cooled Condenser Used in the Test...................................................................................................16

Figure 7 - Important Refrigeration System Parameters vs Hours of Test for the Air Cooled, Two

Stage Fans, 65F Target SCT ..................................................................................................................17

Figure 8 - SCT and Condenser Power Consumption vs Dry Bulb Temperature for the Air Cooled, ..........................

Two Stage Fans, 65F Target SCT..........................................................................................................17

Figure 9 - Subcooling Achieved by the Condenser vs Dry Bulb Temperature for the Air Cooled,

Two Stage Fans, 65F Target SCT...........................................................................................................18

Figure 10 - Important Refrigeration System Parameters vs Hours of Test for the Air Cooled,

Variable Speed Fan Drive, 65F Target SCT ........................................................................................18

Figure 11 - SCT and Condenser Power Consumption vs Dry Bulb Temperature for the Air

Cooled, Variable Speed Fan Drive, 65F Target SCT...........................................................................19

Figure 12 - Important Refrigeration System Parameters vs Hours of Test for the Air Cooled, Two

Stage Fans, 90F Target SCT ................................................................................................................20


Cooled, Two Stage Fans, 90F Target SCT...........................................................................................20


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Figure 14 - Important Refrigeration System Parameters vs Hours of Test for the Air Cooled,



Cooled, Variable Speed Fan Drive, 90F Target SCT...........................................................................22

Figure 16 - Important Refrigeration System Parameters vs Hours of Test for the Air Cooled,Two Stage Fans, Variable Target SCT ...................................................................................................22

Figure 17 - SCT and Condenser Power Consumption vs Dry Bulb Temperature for the AirCooled, Two Stage Fans, Variable Target SCT......................................................................................23

Figure 18 - Important Refrigeration System Parameters vs Hours of Test for the Air Cooled,Variable Speed Fan Drive, Variable Target SCT....................................................................................24

Figure 19 - SCT and Condenser Power Consumption vs Dry Bulb Temperature for the AirCooled, Variable Speed Fan Drive, Variable Target SCT ......................................................................24

Figure 20 - Schematic of an Evaporative Condenser................................................................................................25

Figure 21 - Evaporative Condenser Used for the Test..............................................................................................27

Figure 22 - Important Refrigeration System Parameters vs Hours of Test for the Evaporative,

Two Speed Fan, 65F Target SCT.........................................................................................................28

Figure 23 - SCT and Condenser Power Consumption vs Ambient Temperatures for the

Evaporative, Two Speed Fan, 65F Target SCT....................................................................................29




Evaporative, Variable Speed Fan Drive, 65F Target SCT ..................................................................30


Two Speed Fan, 90F Target SCT.........................................................................................................31


Evaporative, Two Speed Fan, 90F Target SCT....................................................................................31




Evaporative, Variable Speed Fan Drive, 90F Target SCT ...................................................................33

Figure 30 - Important Refrigeration System Parameters vs Hours of Test for the Evaporative,Two Speed Fan, Variable Target SCT....................................................................................................34


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Figure 31 - SCT and Condenser Power Consumption vs Ambient Temperatures for theEvaporative, Two Speed Fan, Variable Target SCT...............................................................................34

Figure 32 - Important Refrigeration System Parameters vs Hours of Test for the Evaporative,Variable Speed Fan Drive, Variable Target SCT....................................................................................35

Figure 33 - SCT and Condenser Power Consumption vs Ambient Temperatures for theEvaporative, Variable Speed Fan Drive, Variable Target SCT ..............................................................36

Figure 34 - Important Refrigeration System Parameters vs Hours of Test for the Evaporative,80% Fan Speed.......................................................................................................................................36

Figure 35 - SCT and Condenser Power Consumption vs Ambient Temperatures for theEvaporative, 80% Fan Speed..................................................................................................................37

Figure 36 - Water Consumption per MBtu of Rejected Heat for all Evaporative Condenser TestRuns........................................................................................................................................................38

Figure 37 - Important Heat Rejection and Evaporation Rate Values for all EvaporativeCondenser Test Runs..............................................................................................................................38

Figure 38 - Condenser Power Use vs Ambient Dry Bulb Temperature for Air Cooled and

Evaporative Condensers Operating with Fan Cycling and 65F Target SCT........................................39

Figure 39 - System and Condenser Energy Efficiency Ratings for Air Cooled and Evaporative

Condensers Operating with Fan Cycling and 65F Target SCT ............................................................40


Evaporative Condensers Operating with Fan Cycling and 90F Target SCT........................................41


Condensers Operating with Fan Cycling and 90F Target SCT ............................................................42

Figure 42 - Condenser Power Use vs Ambient Dry Bulb Temperature for Air Cooled andEvaporative Condensers Operating with Fan Cycling and Variable Target SCT ...................................43

Figure 43 - System and Condenser Energy Efficiency Ratings for Air Cooled and EvaporativeCondensers Operating with Fan Cycling and Variable Target SCT .......................................................43


Evaporative Condensers Operating with Variable Speed Fan Drive and 65F TargetSCT.........................................................................................................................................................44


Condensers Operating with Variable Speed Fan Drive and 65F Target SCT ......................................45

Figure 46 - Condenser Power Use vs Ambient Dry Bulb Temperature for Air Cooled andEvaporative Condensers Operating with Variable Speed Fan Drive and 90F TargetSCT.........................................................................................................................................................45


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Condensers Operating with Variable Speed Fan Drive and 90F Target SCT ......................................46

Figure 48 - Difference in Enthalpy of Air Entering the Condenser and Saturated Air at the SCTvs Ambient Wet Bulb Temperature for the Evaporative, Variable Speed Fan Drive,

Variable Target SCT Test Run...............................................................................................................47

Figure 49 - Condenser Power Use vs Ambient Dry Bulb Temperature for Air Cooled andEvaporative Condensers Operating with Variable Speed Fan Drive and Variable

Target SCT .............................................................................................................................................47

Figure 50 - Temperature Difference Between the Ambient Dry Bulb Temperature and the SCT vsAmbient Dry Bulb Temperature for the Air Cooled Condenser Operating withVariable Speed Fan Control and Variable Target SCT...........................................................................48

Figure 51 - System and Condenser Energy Efficiency Ratings for Air Cooled and EvaporativeCondensers Operating with Variable Speed Fan Drive and Variable Target SCT .................................49

Figure 52 - Average Condenser Energy Efficiency Ratings for the Evaporative and Air Cooled

Condensers for each Test Run................................................................................................................49

Figure 53 - Average Power Consumption, Saturated Condensing Temperatures, and AmbientTemperatures for all of the Air Cooled Test Runs..................................................................................50

Figure 54 - Average Power Consumption, Saturated Condensing Temperatures, and AmbientTemperatures for all of the Evaporative Condenser Test Runs..............................................................51

Figure 55 - Performance Parameters for Air Cooled and Evaporative Condensers for Full LoadOperation................................................................................................................................................51


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NOMENCLATURE

DBTamb Ambient dry bulb temperature (F)WBTamb Ambient wet bulb temperature (F)

Tsump Temperature of the evaporative condenser sump (F)Areacond Total surface area of condenser coil (ft2)

Coillength Total length of condenser coil (ft)CFM Volumetric flow rate of air through the condenser (ft3/min)kWfan Power consumed by condenser fan motor (kW)kWpump Power consumed by evaporative condenser water pump (kW)MFref Mass flow rate of refrigerant (lb/min)Pcond Pressure of refrigerant at condenser inlet (psig)(psia)SCT Saturated condensing temperature (with respect to condenser inlet conditions) (F)

Tcondin Temperature of refrigerant at condenser inlet (F)Tcondout Temperature of refrigerant at condenser outlet (F)Hcondin Enthalpy of superheated refrigerant at inlet of condenser (Btu/lb)Hsatliq Enthalpy of saturated liquid refrigerant in condenser (Btu/lb)

Tsccond Amount of subcooling achieved in condenser (F)

Cpsubcool Specific heat of subcooled liquid refrigerant at condenser exit (Btu/lb F)Hsubcool Enthalpy change between saturated liquid in condenser and subcooled liquid leaving

condenser (Btu/lb)Hcondout Enthalpy of subcooled refrigerant leaving condenser (Btu/lb)Hcond Enthalpy change between superheated and subcooled refrigerant at condenser inlet and exit

(Btu/lb)TDcond Temperature difference between SCT and WBT for evaporative condenser, between SCT

and DBT for air cooled condenser (F)ST Suction temperature of refrigerant (F)SP Suction pressure of refrigerant (psig)(psia)Hcompin Enthalpy of superheated refrigerant entering compressor (Btu/lb)DT Discharge temperature of refrigerant (F)

DP Discharge pressure of refrigerant (psig)(psia)Hcompout Enthalpy of superheated refrigerant leaving compressor (Btu/lb)

Tscin Temperature of refrigerant entering sub-cooler (F)Pscin Pressure of refrigerant entering sub-cooler (psig)(psia)Hscin Enthalpy of refrigerant entering sub-cooler (Btu/lb)

Tscout Temperature of refrigerant leaving sub-cooler (F)Pscout Pressure of refrigerant leaving sub-cooler (psig)(psia)Hscout Enthalpy of refrigerant leaving sub-cooler (Btu/lb)Mech SC Amount of mechanical subcooling (F)kWcond Power consumed by condenser (kW)kWcomp Power consumed by compressor (kW)Load Refrigeration load on the system (Btu/min)


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INTRODUCTION

The purpose of this project was to test and evaluate the impact of several common control strategieson the power use and performance of evaporative and air cooled condensers used in refrigerationsystems. The effects of these control strategies on the condenser and refrigeration systemefficiencies were also evaluated. Southern California Edison (Edison) conducted this test at itsstate-of-the-art Refrigeration Technology and Test Center (RTTC), located in Irwindale, California.

The RTTCs sophisticated instrumentation and data acquisition system provided detailed tracking ofthe refrigeration systems critical temperature and pressure points during the test. These readingswere then utilized to quantify various heat transfer and power related parameters of the refrigerationcycle.

The overall capacity, or power use, of the compressor used in this test was only twice the capacityof the condenser, which does not reflect a realistic scenario. In actual supermarket refrigerationsystems, the compressors can use up to about 8 times more power than the condensers. Figure 2represents the relationship of compressors and condensers power use in an actual supermarketoperating in the coastal climate of Southern California from March to October. The unrealisticcompressor to condenser power use ratio in this test has caused the condenser to impose a heavierinfluence on the system power use profiles, whereas, in actual supermarkets, compressor power use

plays a much more significant role in overall system power use.

Typical Grocery Store Energy Use Percentages

0

20

40

60

80

100

March April May J une J uly August September OctoberPercen

to

fTotalEnergy

Use

dB

Compressoran

dCon

denser

Compressor Condenser

Figure 2 - Power Use Breakdown for Typical Supermarket


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The trade off between the compressors operation at lower discharge pressures and additional condenser fan energyconsumption and implementing an optimum condenser sizing and control strategy has always been a challenge.Most common condenser fan control strategies for maintaining a fixed or floating compressor head are:

Two speed fans

Staging of various size fans (fan cycling)

Variable Speed Drive (VSD)

As a result, this test evaluated the performance of evaporative and air cooled condensers under the followingscenarios:



3. Fixed condensing temperature at 90F, using two speed fan motors / fan cycling

4. Fixed condensing temperature at 65F, using two speed fan motors / fan cycling5. Variable condensing temperature set point, using two speed fan motors / fan cycling6. Variable condensing temperature set point, using VSD control on fan motor

In addition to above scenarios, the evaporative condenser was tested with its fan operating at 80% of its capacity at aconstant speed. The purpose of this test was to capture the effects of smaller sized fans operating at a maximumfixed speed on the condensing temperature of the system.

Figure 3 displays Irwindales dry bulb and wet bulb profiles over the thirteen test days. Throughoutthe entire thirteen days of the test, the refrigeration load remained constant at seven tons, while themain focus was on effects of ambient conditions and control strategies on the condenserperformance and power use.

Ambient Temperature Fluctuation and Refrigeration Load for Entire Test Period

hourly data

50

60

70

80

90

100

110

Day 1 Day 2 Day 3 Day 4 Day 5 Day 6 Day 7 Day 8 Day 9 Day 10 Day 11 Day 12 Day 13

Tempera

ture

(oF)

5

6

7

8

9

10

Re

frigera

tionLoa

d(tons

)

Dry Bulb Wet Bulb Refrigeration Load

Figure 3 - Ambient Temperature Fluctuation for Entire Test Period


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Based on the tested control scenarios, the results indicate that except for one control scenario theevaporative condenser achieved a better energy efficiency ratio (EER) than the air cooled condenser

(Table 1). The variable speed drive control with 90F target SCT was the only mode of controlunder which the air cooled EER was higher than the evaporative condenser. For the air cooled

condenser, the variable speed drive fan control at 90F target SCT required the lowest averagecondenser power use while yielding the highest SCT.

For the evaporative condenser, the constant speed operation of the fan at 80% of its rated capacity yieldedthe most favorable results. Except for the 90F SCT tests, this results of this test display the lowestSCTs, while the ambient wet bulb was the highest and required the least average condenser power.

The main difference between the two condensers, however, remained in the yielded SCTs and condenserpower use. Overall, the evaporative condenser achieved lower SCTs and consumed less power than theair cooled condenser.

TwoSpeed/Stage

Fans at

65F SCT

TwoSpeed/Stage

Fans at

90F SCT

TwoSpeed/Stage

Fans at

Variable SCT

VariableSpeed FanControl at

65F SCT


90F SCT


VariableSCT

Air Cooled 23.51 28.14 26.59 23.37 179.48 29.66

Evaporative 29.64 83.92 34.71 29.86 87.48 42.88

Table 1 - Average Condenser EER Values (Btu/h/W) for Various Test Modes

Table 2 compares the three critical performance parameters of the tested condensers. The fan air flowrates (CFM) andCoil Areadata were obtained from the condenser manufacturers catalogs based on peakdesign operation. As shown in Table 2, the evaporative condenser requires less coil area, less fan CFMand thereby less power than the air cooled condenser per unit of rejected heat.

Condenser Coil Area/HeatRejected,(ft

2/MBH)

CFM/HeatRejected,

(CFM/Btuh)

Condenser Power/Heat Rejected, (kW/MBH)

Air Cooled 0.429 0.279 0.043

Evaporative 0.076 0.069 0.034

Table 2 - Various Condenser Performance Parameters per Amount of Heat Rejected


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REPORT ORGANIZATION

Section (1) Test Procedure and Data Acquisition

This section contains detailed descriptions of the tests that were performed, including the date andtime that the tests were performed, the specific procedures followed for data acquisition, anddescriptions of the tools and hardware used during the stages of testing and analysis.

Section (2) Analysis

The overall approach used to reduce and analyze the collected data is discussed in this section. Itcontains sub-sections on:

2.1) Data collection procedure and initial reduction and screening2.2) Screening procedure2.3) Calculations used to evaluate refrigeration system

Section (3) Heat Rejection by Air Cooled Condenser

A brief discussion of the fundamentals of heat rejection by air cooled condensers is provided inthis section. It includes:

3.1) General description and theory of air cooled heat rejection3.2) Specifications of the air cooled condenser used in this test

Section (4) Air Cooled Condenser Tests - Observations on Individual Air Cooled Test Runsand the Heat Rejection Characteristics Displayed

In this section, explicit observations made during each air cooled condenser test is discussed.

4.1) Test 1 - Fan Cycling at 65F Target Saturated Condensing Temperature

4.2) Test 2 - Variable Speed Drive at 65F Target Saturated Condensing Temperature4.3) Test 3 - Fan Cycling at 90F Target Saturated Condensing Temperature

4.4) Test 4 - Variable Speed Drive at 90F Target Saturated Condensing Temperature4.5) Test 5 - Fan Cycling at Variable Target Saturated Condensing Temperature4.6) Test 6 - Variable Speed Drive at Variable Target Saturated Condensing Temperature

Section (5) Heat Rejection by Evaporative Condenser

A brief discussion of the fundamentals of heat rejection by evaporative condensers is provided inthis section. It includes:

5.1) General description and theory of heat rejection by evaporative condensers

5.2) Specifications of evaporative condenser used in this test


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Section (6) Evaporative Condenser Tests - Observations on Individual Evaporative Test Runsand the Heat Rejection Characteristics Displayed

In this section, explicit observations made during each evaporative condenser test is discussed.

6.1) Test 7 - Two Speed Fan at 65F Target Saturated Condensing Temperature

6.2) Test 8 - Variable Speed Drive at 65F Target Saturated Condensing Temperature

6.3) Test 9 - Two Speed Fan at 90F Target Saturated Condensing Temperature6.4) Test 10 - Variable Speed Drive at 90F Target Saturated Condensing Temperature6.5) Test 11 - Two Speed Fan at Variable Target Saturated Condensing Temperature6.6) Test 12 - Variable Speed Drive at Variable Target Saturated Condensing Temperature6.7) Test 13 - Fixed Fan Speed at 80%6.8) Water Evaporation Rate in Evaporative Condenser

Section (7) Discussion of Results - A Comparative Evaluation of the Two Condensing Types

The air cooled and evaporative condenser test results are compared and discussed in detail in thissection.

7.1) Two Speed/Stage at 65F Target Saturated Condensing Temperature7.2) Two Speed/Stage at 90F Target Saturated Condensing Temperature7.3) Two Speed/Stage at Variable Target Saturated Condensing Temperature

7.4) Variable Speed Drive at 65F Target Saturated Condensing Temperature

7.5) Variable Speed Drive at 90F Target Saturated Condensing Temperature7.6) Variable Speed Drive at Variable Target Saturated Condensing Temperature

Section (8) Conclusions

This section contains a summary of test findings.


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1. TEST PROCEDURE

The refrigeration load for the heat rejection test was created by chilling city water using a plate heatexchanger. The heat absorbed by the refrigerant in the chiller was held constant by the use of a Btumeter and a water control valve. The flow rate of city water through the heat exchanger wasmodulated in order to hold constant the heat absorbed by the refrigerant.

Temperature measurements considered to be critical to the process were recorded from a group ofthree sensors; one resistance temperature device (RTD) and two thermocouples. These data pointswere extracted from the daily files, and the readings from the thermocouples were compared to eachother, and to the reading from the associated RTD. The difference between sensor readings wascompared to criteria established during the commissioning of the data acquisition system. Any datawhere the maximum difference fell outside the allowable standard deviation was flagged for furtherreview.

The operation of the compressor and condensers was controlled by the Danfoss NC-25 andComTrol MCS 4000 microprocessor controllers. The microprocessor controller was equipped witha stand-alone modem for remote access to the control parameters. An interface with themicroprocessor controller was made through a serial connection to a PC located in the computer

room. Through this interface, all parameters of the microprocessor control could be modified andinspected.

The Variable Frequency Drive (VFD) modulated the compressor speed (and thereby its capacity andthe refrigerant mass flow rate) according to inputs from the microprocessor controller. Themicroprocessor controller changed the VFD output to the compressor according to variations insuction pressure.

An evaporator pressure regulator (EPR) valve was utilized downstream from the evaporator (chiller)to maintain a desired pressure at the chiller and prevent the temperature from falling below aminimum set point. Without the EPR valve, the city water could begin to freeze within the heat

exchanger once the temperature fell below the minimum set point of 32F.

Tests were conducted over thirteen twenty-four hour periods under various control strategies(Tables 3 and 4). The evaporative condenser test runs were conducted first, followed by the aircooled condenser test runs. Seven control strategies were applied to the evaporative condenser, andsix were applied to the air cooled condenser.

Day 1Sept 7th

Day 2Sept 8th

Day 3Sept 9th

Day 4Sept 11th

Day 5Sept 12th

Day 6Sept 13th

Day 7Sept 22nd

VariableSpeedDrive,

Floating

HeadPressure

VariableSpeed Drive,

65FSaturated

CondensingTemperature

VariableSpeed Drive,

90FSaturated


Two SpeedFan Cycling,

90FSaturated


Two SpeedFan Cycling,

65FSaturated


Two SpeedFan

Cycling,Floating

HeadPressure

Fan MotorOperatingat 80% ofMaximum

Speed

Table 3 - Test Runs for the Evaporative Condenser


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Day 8Sept 16th

Day 9Sept 17th

Day 10Sept 20th

Day 11Sept 30th

Day 12Oct 2nd

Day 13Oct 3rd

Two StageFan Cycling,

90FSaturated


Two StageFan Cycling,

65FSaturated


Two Stage FanCycling,

Floating HeadPressure

Variable SpeedDrive,

90F SaturatedCondensing

Temperature


65F SaturatedCondensing

Temperature


Floating HeadPressure

Table 4 - Test Runs for the Air Cooled Condenser

With the objective of minimizing instrument error and maintaining a high level of repeatability andaccuracy in the data, careful attention was paid to the system design. With this end in mind, thefollowing steps were taken:

1. Use of sensors with the highest accuracy available.2. Minimization of sensor drift errors by use of redundant sensors.

3. Utilization of calibration standard instruments of the highest accuracy.4. Elimination of interference from power conductors and high frequency signals by double-

shielding sensor leads.

A Kaye Instruments Digi-4 Model #X1520S Data Scanner was used to log the data. Kayes Digi-4has a special emphasis on temperature measurement, with excellent thermocouple signal processing.

The data scanner can process 94 data channels. The scanner was calibrated at the factory, and istraceable to the National Institute of Standards and Technologys (NIST) standards. The system has

57 special grade type-T thermocouples accurate to 0.03C, 14 precision 100 platinum

Resistance Temperature Device (RTD) inputs accurate to 0.01C, and a combination of 23 analoginputs from pressure transducers, dew point sensors, flow meters, and CT-transducers. An RS-232communication link sent one data report that included instantaneous values of all data points every

ten seconds. To ensure that the data collection was not compromised by the control sequencespriority over data acquisition, the data acquisition system for the project was designed to becompletely independent of the supervisory control computer.

Every 10 seconds the data acquisition system sampled the scanned data and created time-stampedtwo-minute averages. The two-minute data was then saved to a file which was closed at the end ofeach 24-hour period. Every 24 hours, the data collected from the previous 24 hours wasdownloaded remotely to Southern California Edisons San Dimas office for screening andprocessing. The raw data were analyzed daily for consistency and accuracy. In the event that any ofthe test parameters fell outside acceptable limits, the problem was flagged. In such cases, test runswere repeated until correction of the problem.


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2. ANALYSIS

2.1 Data Collection/Reduction

The test facility is equipped with a sophisticated data acquisition system that can scan 94 sensorsand logs their outputs in two-minute intervals. Data was collected and stored for each sensor forthirteen days. Every 24 hours during the test, the data was downloaded and checked for consistency

and accuracy. Operating parameters were checked and deemed to be within acceptable limits beforethe next run was started.

The collected data points from the two-minute intervals were averaged into ten-minute intervals and usedfor further screening of the test data. The advantage of using ten-minute averages is that the data trendscan still be displayed with an acceptable resolution without an overwhelming number of data points. Thedata was then averaged into one-hour blocks for each 24-hour period. The hourly averages were used forrefrigeration system calculations. After the hourly data was developed, they were imported into SouthernCalifornia Edisons customized refrigeration analysis tool.

2.2 Screening Procedure

Once the data was compiled into hourly averages within the spreadsheet, tabular and graphicalrepresentations of various correlations and calculated parameters were produced. Several graphswere created to initially screen the data. The initial screening plots are located in Appendix C.Included in this group were the fundamental data points provided by the data acquisition system.

These plots were used to determine the validity of each test. After careful examination of the initialscreening plots, the informational plots were produced. This set provided relationships betweencalculated quantities. With more data points to analyze, the trends in the data were more easilydetermined by using the two-minute data as opposed to the hourly data.

2.3 Refrigeration Calculations

A series of calculations were performed to obtain the key refrigeration parameters including heat

rejection at the condenser, condenser efficiency, heat of compression, system efficiency, andcompressor power per ton. After the data was downloaded from the data logger and the data ofinterest was extracted, some preliminary calculations were performed. These calculations includedaveraging temperature data that were read by more than one sensor. Such data includedtemperatures at the condenser inlet and outlet, suction temperature, and temperature at the exit ofthe subcooler.

The heat rejection at the condenser, heat of compression, and compressor power all depend on therefrigerant enthalpies at different locations of the refrigeration cycle. Enthalpies can be eitherobtained from the refrigerant manufacturers data at various temperatures and pressures, orcalculated with respect to specific heat capacities and temperatures. In this analysis, some of theenthalpies were obtained from refrigerant property software and some by calculation.

Once the temperatures and pressures were determined, the enthalpies were obtained. DuPonts SuvaRefrigerant Expert Program, version 2.0 was used to determine the saturated and superheated vapor


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enthalpies. The data logger provided all pressures in gage units, and after conversion to absoluteunits, the Refrigerant Expert program was used to look up the enthalpies.

The enthalpies in the saturated phase were calculated using temperature-dependent expressionsprovided by DuPont, as well as using basic thermodynamic relationships. Equation 1, provided byDuPont, determined the saturated enthalpy in kJ /kg of refrigerant 404A for a temperature range of -20 C to 40 C. The temperatures of the saturated liquid were first converted to Celsius, then

inserted into equation 1 to produce the corresponding saturated enthalpy. The enthalpy of saturatedliquid in the condenser was found by use of equations 1 and 2.

1) H =A +BT +CT2

H =Enthalpy (kJ/kg)A =200B =1.438333C =0.003916667

T =Temperature (C)

where A, B, and C were constants determined by DuPont from the relationship between saturated

enthalpy and temperature. Next, equation 2 was used to convert the enthalpy in kJ/kg to Btu/lb. Becauseof a change in reference states from SI to English units, a reference conversion, H (ref), was included inEquation 2.

2) H (Btu/lb) =[H (kJ/kg) - H (ref)] 0.43021(Btu/lb / kJ /kg)

H (ref) =145.6 kJ/kg for R404A

The Refrigerant Expert program does not provide subcooled enthalpies, therefore, in order to findthe enthalpies for subcooled refrigerant, the thermodynamic relationship between enthalpy andtemperature was incorporated. For this relationship, however, the correct liquid specific heatcapacity was needed. Equation 3, provided by DuPont, calculated the liquid specific heat capacity

of refrigerant 404A for a temperature range of -40 F to 140 F. The enthalpy at the exit of thecondenser was found using this equation and the subcooled temperature at this location.

3) Cp =0.306 +4.083E-4 T - 1.194E-6 T2 +8.056E-8 T3

Cp = Liquid Heat Capacity (Btu/lb F).T = Average temperature of the subcooled liquid (F) for a range of -40 F to 140 F.

The temperature difference between the saturated liquid in the condenser and the subcooled liquidleaving the condenser was needed in order to find the corresponding enthalpy change due tosubcooling within the condenser.

4) DTsccond =SCT - Tcondout


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DTsccond = Temperature difference between saturated liquid in the condenser and subcooledliquid leaving the condenser (F).

SCT = Temperature of saturated liquid in the condenser (F). This value wasdetermined by look-up using DuPonts Suva Refrigeration Expert Program andcondenser inlet pressure data from the data acquisition system.

Tcondout = Average temperature of refrigerant at condenser exit (F). This value was readdirectly using the data acquisition system.

Next, the enthalpy change between the subcooled and saturated liquid was calculated by utilizingthe following thermodynamic relationship.

5) DHsubcool =Cp DTsccond

DHsubcool = The enthalpy change between the subcooled liquid leaving the condenser and saturatedliquid in the condenser (Btu/lb).

Finally, the enthalpy of the subcooled liquid was computed. In order to accomplish this, the enthalpychange between the subcooled and saturated liquid was subtracted from the enthalpy of saturated liquid.

6) Hsubcool =Hsatliq - DHsubcool

Hsubcool = The subcooled liquid enthalpy leaving the condenser (Btu/lb).Hsatliq = Saturated liquid enthalpy (Btu/lb). This value was determined using equations 1 and 2.

After determination of all enthalpy values, calculations were made to determine parameters ofinterest such as heat of compression, heat rejection at the condenser, and energy efficiency rating.

The refrigeration load is the amount of cooling or heat removal that takes place at the evaporator. Thisparameter was obtained directly from the Btu meter readings in Btu/min, but a unit conversion wasneeded to express this value in its usual units, Btu/hr (equation 7).

7) Qload =Load k

Qload = Refrigeration load imposed on the system (Btu/hr).Load = Refrigeration load imposed on the system (Btu/min).k = Conversion factor, 60 (min/hr).

It is also important to determine the refrigeration load in tons. Thus, the refrigeration load was divided by12,000, a conversion factor from Btu/hr to tons (equation 8).

8) Qload (tons) =Qload / 12000

Qload (tons) = Refrigeration load (tons).


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Obtaining the heat of compression is also of interest because it can be used to calculate the amountof heat rejection at the condenser and the theoretical compressor power. Because the heat ofcompression is the difference between the enthalpies at the suction and discharge sides of thecompressor, this value is affected by variations in saturated condensing temperature. The saturated

suction temperature remained unchanged at about 32 F throughout the test. In order to obtain theheat of compression, the mass flow rate of refrigerant was multiplied by the difference between theenthalpies of the compressor inlet and outlet (equation 9).

9) Qcomp =MFref k (Hcompout - Hcompin)

Qcomp = Heat of compression (Btu/hr).k = Conversion factor, 60 (min/hr).Hcompout = Superheated enthalpy at the outlet of the compressor (value determined by look-up

using DuPonts Suva Refrigeration Expert Program)(Btu/lb).Hcompin = Superheated enthalpy at the inlet to the compressor (value determined by look-up

using DuPonts Suva Refrigeration Expert Program)(Btu/lb).

The total heat rejected at the condenser was obtained by the product of mass flow rate and change inrefrigerant enthalpies between the inlet and outlet of the condenser, which includes de-superheating,

latent (or phase change), and subcooling heat removals within the condenser (equation 10).

10) Qcond =MFref k (Hcondin - Hcondout)

Qcond = Heat Rejection at the Condenser (Btu/hr).k = Conversion factor, 60 (min/hr).

To prevent flashing in the liquid line, additional subcooling is accomplished with a glycol chiller.Determination of the mechanical subcooling accomplished here is important to perform a proper heatbalance of the system. The amount of heat rejected at the mechanical subcooler is determined by theproduct of mass flow rate of refrigerant and change in refrigerant enthalpies between the condenser outletand subcooler outlet, assuming an adiabatic transport of liquid refrigerant (equation 11).

11) Qsubcooler =MFref k (Hcondout - Hsc out)

Qsubcooler = Heat rejected in the subcooler (Btu/hr).k = Conversion factor, 60 (min/hr).Hsc out = Liquid enthalpy at the outlet of the subcooler (Btu/lb).

Another important parameter was the total amount of heat rejected by the system. This is computed bysumming the heat rejected in the condenser and the heat rejected in the subcooler (equation 12).

12) Qtot rej =Qcond +Qsubcooler

Qtot rej = Total amount of heat rejected by the system (Btu/hr).


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The capacity of a condenser can also be expressed per lineal foot of coil. The heat rejection in thecondenser was divided by the total lineal footage of the condenser coil from the manufacturers data toobtain this value (equation 13).

13) Qcondft=Qcond / Coillength

Qcondft = Heat rejected by the condenser per lineal foot of condenser coil (Btu/hr).

Coillength = Total length of coil in the condenser (ft).

The Energy Efficiency Rating (EER) of the refrigeration system is defined as the ratio of the amount ofcooling achieved (refrigeration load, Btu/hr) to the total power used by the system (compressor powerplus condenser power, kW) (equation 14). A large EER indicates a more efficient system.

14) EERsystem=Qload/ (kWcomp+kWcond) k

EERsystem = Energy efficiency rating based on total system power (Btu/hr/W).kWcomp = Power consumed by the compressor (kW).kWcond = Total power consumed by the condenser including fans and pump (kW).k = Conversion factor, 1000 (W/kW).


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3. HEAT REJ ECTION BY AIR COOLED CONDENSER

The air cooled condenser (Figure 4) rejects the total heat of refrigeration through a heat exchangerto the ambient air. A fan(s) blows ambient air across the finned coils of the heat exchanger whichcarries high pressure superheated refrigerant vapor. The superheated refrigerant vapor de-superheats and undergoes a phase change converting into liquid (typically sub-cooled) within the

heat exchanger by rejecting its heat to the ambient air.

fan

air out

air in

condensercoil

highpressure,

superheatedrefrigerant

vapor

saturatedor

subcooledliquid

refrigerant

Figure 4 - Schematic of an Air Cooled Condenser

3.1 Theory of Heat Rejection by Air Cooled Condenser

The mechanism of condensation is complex. The ability of the refrigerant to condense depends onseveral variables including density, viscosity and conductivity of condensate, latent heat of vaporizationof refrigerant, temperature difference between refrigerant vapor and coil surface, and the number of tubesand their diameter. At the entrance to the condenser tube, superheated refrigerant vapor has lowcondensing ability. This is an inherent thermal characteristic of gases in convection heat transfer. Oncede-superheating ends and the condensing of the vapor begins, the ability of the refrigerant to transfer heatand condense increases. On the other hand, liquid travels slower than gas, and filling more of the heattransfer surface with liquid decreases the exposure of the vapor to the coil surface, resulting in a suddendecrease in rate of condensation (see Figure 5).


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Entrance (after

desuperheating)Exit

Distance Along CondenserTube

Figure 5 - Variation of Condensing Coefficient With Distance Along Condenser Tube

The heat rejection of an air cooled condenser is expressed by equation 16.

16) Qrej =U A LMTD

Qrej = Total heat rejection at the condenser (Btu/hr).U = Overall coefficient of heat transfer (Btu/hr-ft2).A = Surface area (ft2).LMTD = Log Mean Temperature Difference as a function of refrigerant condensing

temperature and air inlet and outlet temperatures (F).

The LMTD can be calculated by using the following equation, simplified to become equation 17.

LMTD =[(SCT - Ti) - (SCT - To)] / [ln (SCT - Ti) / (SCT - To)]

17) LMTD =(To - Ti) / [ln (SCT - Ti) / (SCT - To)]

To = Dry-bulb temperature of air leaving the condenser coil (F).

Ti = Dry-bulb temperature of air entering the condenser coil (F).SCT = Saturated Condensing Temperature of refrigerant corresponding to compressor

discharge pressure (F).

Typically, in most practical cases (To) is not measured and consequently LMTD is approximated andsubstituted by TD given by equation 18.

18) TD =SCT - Ti

Hence, the condenser heat rejection can be approximated by equation 19.

19) Qrej =U A TD


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A condensers overall coefficient of heat transfer can vary as the velocity of the cooling medium (in thiscase air) changes. Since the face area of the condenser heat exchanger is constant, any change in thevolume rate of flow of air across the condenser will result in a change in the air velocity. Increasing theair velocity will produce turbulent flow and increase the (U), but will decrease after a certain point.Improving the overall coefficient of heat transfer will provide lower condensing temperatures and resultin reduced compression work. The rise in the fans static pressure loss beyond a certain point and theincrease in air flow rate will, however, negate the benefits from operating compressors at lower discharge

pressures (or SCT). This is more evident following the fan laws. The increase in fan power as a cubicfunction of air flow rate (or rpm) is given by equation 20.

20) bhp1/ bhp2 =(cfm1 / cfm2)3

bhp1 = Initial fan motor brake horsepower.bhp2 = Increased fan motor brake horsepower.cfm1 = Initial air flow rate across the condenser coil (ft

3/min).cfm2 = Increased air flow rate across the condenser coil (ft

3/min).

From an operational stand point, close attention should be paid to effective sizing and fan speed control

strategies. The following sections discuss the effects of various fan control strategies with respect tospecific SCT targets. Again, the goal should be to obtain an optimum balance point between a desirableSCT (resulting in less compressor work) and a minimum possible fan power consumption.

3.2 Description of Test Air Cooled Condenser

The air cooled condenser used for this test had the following specifications. Figure 6 shows a photographof the unit.

Make: Bohn

Model No.: BRH027

Capacity: 195 MBtu/h (15 TD, R-404A)

CFM: 23,200

Coil: 14 tube circuits, 35.6 ft2 fin surface area

Fan Motors: 2 motors, each 1.5 HP, 1140 RPM


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Figure 6 - Air Cooled Condenser Used in the Test

4. AIR COOLED CONDENSER TESTS

4.1 Test 1 - Fan Cycling at 65F Target Saturated Condensing Temperature

The control setting was programmed to allow fan capacity modulations of 0%, 50%, and 100% with

respect to target discharge pressure. At condensing temperatures higher than 65F, both fans operate at

full load. With a design Temperature Difference (TD) of 15 F, and low ambient dry bulb temperatures,

condenser capacity could reduce to 50% by stepping down one of the fans. This condition, however,never occurred since the actual climatic conditions during this particular test never reached sufficientlylow dry bulb temperatures allowing a 50% condenser capacity reduction. Figure 7 depicts the dry bulbtemperature profile on 9/17-9/18/97 in conjunction with other test parameters. During this test interval,

the daily range (maximum peak - min) was 26.7 F, and the dry-bulb temperature had an average,

minimum, and maximum of 75.6 F, 64.3F and 91.0F, respectively.


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0

2

4

6

8

10

1 3 5 7 9 11 13 15 17 19 21 23

Hours of Test

Poweran

d

Re

frigera

tion

Loa

d

0

20

40

60

80

100

Dry

Bu

lb

Tempera

ture

(oF)

Compressor Power (kW) Condenser Power (kW) System Power (kW)Refrigeration Load (tons) Dry Bulb Temperature

Figure 7 - Important Refrigeration System Parameters vs Hours of Test for the Air Cooled, Two Stage Fans, 65F Target SCTTest Run (10 minute data)

Figure 7 correlates the key parameters of this test. As shown, the refrigeration load (similar to other testmodes) remained at a constant rate of 7 tons during the 24 hours of the test. With relatively high ambientdry bulb temperatures and low target SCT, the controls did not allow the fans to cycle off. As Figure 7

indicates, both fans operated continuously aiming unsuccessfully at a low but unattainable SCT of 65F.Consequently, condenser power follows a constant straight line with a zero slope as the dry bulbtemperature fluctuates. On the other hand, the increase in dry bulb temperature caused an increase insaturated discharge pressure and power use of the compressor. The effects of the compressor power useincrease is reflected in the system power profile (Figure 7). Figure 8 clearly shows the progressiveincrease of saturated condensing temperature with respect to ambient dry bulb temperature, whilecondenser power remained nearly constant.

50

60

70

80

90

100

110

64 64 65 65 67 68 69 70 70 71 73 75 76 77 77 79 81 81 82 85 86 89 89 91

Dry Bulb Temperature (oF)

Sa

tura

tedCon

dens

ing

Temperatu

re(oF)

0

1

2

3

4

Con

denserPo

wer

(kW)

Saturated Condensing Temperature Condenser Power

Figure 8 - SCT and Condenser Power Consumption vs Dry Bulb Temperature for the Air Cooled, Two Stage Fans, 65 F TargetSCT Test Run (hourly data)

The performance of the condenser under this mode of control clearly deteriorated as the ambienttemperature increased, resulting in higher discharge pressures and compressor power use. In addition, itwas observed that the condensers ability to sub-cool the liquid refrigerant at the condenser exit


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diminished with an increase in dry bulb temperature. Figure 9 indicates a 0.89F drop in condenser sub-

cooling per 10 F increase in dry bulb temperature under the specific conditions and mode of control ofthis test.

3

4

5

6

7

64 64 65 65 67 68 69 70 70 71 73 75 76 77 77 79 81 81 82 85 86 89 89 91


Con

denser

Su

bcoo

ling(

oF)

Condenser Subcooling

Figure 9 - Subcooling Achieved by the Condenser vs Dry Bulb Temperature for the Air Cooled, Two Stage Fans, 65F Target

SCT Test Run (hourly data)

4.2 Test 2 - Variable Speed Drive at 65F Target Saturated Condensing Temperature

This test took place on a milder day (10/02-10/03/97) than the previous test (AC 2 65), with a temperature

range of 11.5F. The average, minimum, and maximum dry bulb temperatures were 71.1F, 66.5F,

and 78.0F respectively (Figure 10).

0

2

4

6

8

10

1 3 5 7 9 11 13 15 17 19 21 23Hours of Test

Poweran

dRe

frigera

tion

Loa

d

0

10

20

30

40

50

60

70

80

Dry

Bu

lbTemperat

ure

(oF)

Compressor Power Condenser Power System PowerRefrigeration Load (tons) Dry Bulb Temperature

Figure 10 - Important Refrigeration System Parameters vs Hours of Test for the Air Cooled, Variable Speed Fan Drive, 65F

Target SCT Test Run (10 minute data)


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Similar to the AC 2 65 test, due to a low target SCT of 65F and relatively high dry bulb ambientconditions, both fans ran continuously at full load. As a result of the ambient and low SCT targetrestrictions, the fan control strategy did not play any role in this test.

In this test, as expected, the SCT increased linearly with respect to dry bulb temperature while thecondenser power remained constant at 3.6 kW (Figure 11). The lower peak SCT is due to a lowerambient dry bulb temperature than the previous test.

50

60

70

80

90

100

110

67 67 67 67 67 68 68 68 68 69 69 69 69 70 72 73 74 74 75 76 76 78 78 78


Sa

tura

tedCon

dens

ing

Tempera

ture

(oF)

0

1

2

3

4

Con

denser

Power

(kW)


Figure 11 - SCT and Condenser Power Consumption vs Dry Bulb Temperature for the Air Cooled, Variable Speed Fan Drive, 65F Target SCT Test Run (hourly data)

4.3 Test 3 - Fan Cycling at 90F Target Saturated Condensing Temperature

Under this mode of test, condenser capacity can be reduced to 50% by cycling off one of the fans during

periods with sufficiently low ambient dry bulb temperatures, while maintaining the target 90 F SCT.The dry bulb fluctuation pattern shown in Figure 12 represents Irwindales weather on 9/16-9/17/97 with

daily range, minimum, maximum and average temperatures of 17.2 F, 70.1F, 87.3F and 77.0F,respectively.

The total system power use follows a close profile to that of ambient dry bulb (Figure 12). During low

ambient dry bulb temperatures with a 15F design TD, the condenser achieved the target SCT of 90F.From the 10th hour to 20th hour of the test, the ambient dry bulb temperature was low and compressorpower reached a minimum constant level indicating a non-fluctuating SCT, while the condenser fanscycled, resulting in a reduction in condenser power (Figure 12).


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0

2

4

6

8

10

1 3 5 7 9 11 13 15 17 19 21 23

Hours of Test

Poweran

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R

efrigera

tion

Loa

d

0

1020

30

40

50

60

70

80

90

Dr

yBu

lbTempera

ture

(oF)


Figure 12 - Important Refrigeration System Parameters vs Hours of Test for the Air Cooled, Two Stage Fans, 90 F Target SCTTest Run (10 minute data)

Figure 13 clearly shows that during this mode of test, the condenser fans cycled and the SCT stayed

constant at 90F until an ambient dry bulb temperature of 77F was reached. At the ambient dry bulb

temperature of 77F, the SCT exceeded the set point of 90F. As a result, both fans operated at full loadcontinuously attempting to lower the SCT to 90 F (Figure 13). Although the condenser ran at either 50%or 100% capacity, its power use may appear between the two limits (Figure13). The plotted condenserpower and SCTs are actually the 2-minutes average readings that were collapsed into hourly blocks. The

hourly SCT and condenser kW data for dry bulb temperatures 77F and less include some averaged two-minute intervals during which fans repeatedly cycled off.

70

75

80

85

90

95

100

70 70 70 70 71 71 71 72 73 73 74 75 75 76 77 80 81 82 84 85 85 86 87 87


SCT(oF)

0

1

2

3

4

5

Con

denser

Power

(kW)

Saturated Condensing TemperatureCondenser Power

Figure 13 - SCT and Condenser Power Consumption vs Dry Bulb Temperature for the Air Cooled, Two Stage Fans, 90F TargetSCT Test Run (hourly data)


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This test was conducted by modulating the fans speed using a VSD controller. The fans speed varied at

full load down to 20% Hz, targeting a fixed SCT of 90 F at low ambient dry bulb temperatures. The drybulb profile shown in Figure 14 represents Irwindales weather on 9/30-10/1/97 with daily range,

minimum, maximum and average temperatures of 25.6 F, 65.2F, 90.8F and 73.5F, respectively.

As shown in Figure 14, the VSD controller reduced the fan speed, thus the power from 3.5 to 0.3 kW (by

91.4% ) during periods when ambient dry bulb temperature was lower than about 75F. At temperatures

higher than 82F the SCT exceeded the 90F target. As a result, both fans operated at full capacityaiming unsuccessfully to lower the SCT. During these periods compressor and system power use reachedtheir highest kW (Figure 14).

0

2

4

6

8

10

1 3 5 7 9 11 13 15 17 19 21 23Hours of Test

Poweran

dRe

frigera

tion

Loa

d

0

10

20

30

40

50

60

70

80

90

100

Dry

Bu

lbTempe

rature

(oF)


Figure 14 - Important Refrigeration System Parameters vs Hours of Test for the Air Cooled, Variable Speed Fan Drive, 90FTarget SCT Test Run (10 minute data)

Figure 15 illustrates the relationship of SCT and condenser power with ambient dry bulb temperature.The VSD maintains a minimum condenser fan speed and hence the power at 0.3 kW at ambient

temperatures between 65F and 75F. From the dry bulb temperature of 75F and up, it gradually

ramps up the fan speed reaching full capacity at 82F. Figure 15 depicts this transition as reflected in the

increase of condenser power. Note that SCT exceeds the 82F target at ambient temperature 77F(Figure 15).


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50

60

70

80

90

100

110

65 65 66 66 66 66 66 67 67 67 68 69 70 71 73 74 77 78 81 82 87 88 90 91


Satura

tedCon

dens

ing

T

empera

ture

(oF)

0

1

2

3

4

Con

denser

Power

(kW)


Figure 15 - SCT and Condenser Power Consumption vs Dry Bulb Temperature for the Air Cooled, Variable Speed Fan Drive, 90F Target SCT Test Run (hourly data)

4.5 Test 5 - Fan Cycling at Variable Target Saturated Condensing Temperature

Under this mode of test, condenser capacity can be reduced to 50% by cycling off one of the fans duringperiods with sufficiently low ambient dry bulb temperatures, while maintaining a fixed 15 F TD abovethe fluctuating ambient dry bulb temperature. The dry bulb fluctuation pattern shown in Figure 16represents Irwindales weather on 9/20-9/21/97 with daily range, minimum, maximum and average

temperatures of 27.9F, 59.9F, 87.8F, and 71.4F, respectively.

0

2

4

6

8

10

1 3 5 7 9 11 13 15 17 19 21 23Hours of Test

Poweran

dRefrigera

tion

Loa

d

0

10

20

30

40

50

60

70

80

90

Dry

Bu

lbTem

pera

ture

(oF)


Figure 16 - Important Refrigeration System Parameters vs Hours of Test for the Air Cooled, Two Stage Fans, Variable TargetSCT Test Run (10 minute data)


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The condenser power reaches its maximum at the extreme ends of the power curve (Figure 16), whichcorresponds to the highest dry bulb temperatures during the test period. As the ambient dry bulbtemperature decreases, the fans cycle off more frequently to achieve the target TD above ambientcondition.

50

60

70

80

90

100

110

60 60 61 61 62 63 64 65 65 65 66 68 70 71 73 77 78 81 82 82 83 84 84 88


Sa

tura

tedCon

dens

ing

Tempera

ture

(oF)

0

1

2

3

4

Con

denser

Power

(kW)


Figure 17 - SCT and Condenser Power Consumption vs Dry Bulb Temperature for the Air Cooled, Two Stage Fans, VariableTarget SCT Test Run (hourly data)

The general relationship between the saturated condensing temperature and the condenser power useunder this mode of capacity control is shown in Figure 17. The hourly averages in these graphs providean overall trend and understanding of such relationships. With variable set point control the SCT varieswith respect to ambient dry bulb temperature. Fan power, however, remained at its lowest during lowerambient temperatures, while SCT continued increasing with dry bulb temperature. Beyond the ambient

temperature of 81F, both fans begin to operate continuously at full capacity (Figure 17).

4.6 Test 6 - Variable Speed Drive at Variable Target Saturated Condensing Temperature

Under this mode of test, condenser capacity can be reduced by reducing the fans speeds to 33% CFM

with the aid of the VSD controller, while maintaining a fixed 15F TD above the fluctuating ambient drybulb temperature. The dry bulb fluctuation pattern shown in Figure 18 represents Irwindales weather on

10/03-10/04/97, with daily range, minimum, maximum, and average temperatures of 23.2F, 62.8F,

86.0F, and 71.6F, respectively.

Despite the previous tests, an unexpected observation was made in this test. It was found that duringlower ambient dry bulb conditions, the condenser fan works harder to achieve the target TD than at highertemperatures (Figure 18). It seems the condenser can achieve the TD set-point much easier when theambient temperature is high, resulting in less fan power use. On the other hand, at higher dry bulb

temperatures it results in higher saturated condensing temperatures which causes the compressor toconsume more power (Figure 18).


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0

2

4

6

8

10

1 3 5 7 9 11 13 15 17 19 21 23

Hours of Test

Poweran

d

R

efrigera

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Loa

d

0

1020

30

40

50

60

70

80

90

DryB

ulbTempera

ture

(oF)

Compressor Power (kW) Condenser Power (kW) System Power (kW)

Refrigeration Load (tons) Dry Bulb Temperature

Figure 18 - Important Refrigeration System Parameters vs Hours of Test for the Air Cooled, Variable Speed Fan Drive, VariableTarget SCT Test Run (10 minute data)

Figure 19 correlates the saturated condensing temperature and condenser power variations with respect to

ambient dry bulb temperature. At 71

F ambient temperatures and lower, the condenser fans operate atalmost full speed to achieve the lowest SCT. As the ambient temperature rises above 64F, the SCT

increases while the condenser power remains unchanged. This trend continues until ambient dry bulb

reaches 71F. At temperatures higher than 71F, SCT continues to increase while condenser fan speed,hence power, drops down to its lowest value of 1.6 kW.

50

60

70

80

90

100

110

63 63 63 63 64 64 65 66 67 67 67 69 70 71 71 74 75 78 79 81 83 84 86 86


Sa

tura

tedCon

dens

ing

Tempera

ture

(oF)

0.0

0.5

1.0

1.52.0

2.5

3.0

3.5

4.0

Con

denser

Power

(kW)


Figure 19 - SCT and Condenser Power Consumption vs Dry Bulb Temperature for the Air Cooled, Variable Speed Fan Drive,Variable Target SCT Test Run (hourly data)


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5. HEAT REJ ECTION BY EVAPORATIVE CONDNESER

Similar to air cooled condensers, evaporative condensers change the refrigerant state of the high pressure,superheated vapor discharged from the compressor to liquid. The de-superheating and phase changetakes place within the condenser tubes. Figure 20 illustrates the basic components of an evaporativecondenser. Sprayed water over the condenser tubes and the passage of forced air flow through them arethe primary media to absorb the heat of refrigeration.

Figure 20 - Schematic of an Evaporative Condenser

5.1 Theory of Heat Rejection by Evaporative Condensers

The heat transfer in evaporative condensers takes place in two phases. The first phase of heat transfer isbetween condensing refrigerant and the water film covering the coil given by equation 21.

21) Qrej1 =A U (SCT - Tws)

Qrej1 = Heat transferred between refrigerant and water film on the coil surface (Btu/hr).


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A = Coil surface area (ft2).

U = Overall coefficient of heat transfer of coil.

SCT = Saturated condensing temperature of refrigerant at discharge pressure (F).

Tws = Temperature of water film surface (F).

The second phase of heat transfer takes place between the water film on the coil surface and the airpassing through the coil. This heat flow is dependent on the enthalpy of the air entering the condenser

and the enthalpy of the saturated air (adjacent to the water film) at the refrigerant condensing temperature.This phase of heat removal is expressed by equation 22.

22) Qrej2 =A Uws (hSCT - ha)

Qrej2 = Heat transfer between water film and air over the coil surface (Btu/hr).A = Coil surface area (ft2).

Uws = Overall coefficient of heat transfer between the water film/air interface and enteringair.

hSCT = Enthalpy of saturated air at SCT (Btu/lb).ha = Enthalpy of moist air entering the condenser (Btu/lb).

The evaporation of sprayed water into the air provides the predominant phase of heat rejection inevaporative condensers. Hence, the ambient wet-bulb temperature becomes one of the main drivingforces affecting the performance of these condensers. Apart from the physical characteristics of thecondenser (surface area and coil material) and the ambient conditions, the performance of the evaporativecondenser is greatly influenced by the air flow and spray water rates. Equation 23 correlates the overallheat transfer ability of an evaporatively cooled condenser as a function of air flow and water spray rates.

23) H =k G0.48 L0.22

H = Condenser capacity indicator.k = Constant.G = Rate of water spray.

L = Rate of air flow.

This expression suggests that the water spray rate has a greater influence on the condenser capacity thanthe does the air flow rate. In actual applications, however, water flow rate is rarely controlled. On theother hand, air flow rate is widely controlled to achieve optimum condensing temperatures and condenserpower use. The increase in air velocity affects the overall coefficient of heat transfer of the condenser(U). At higher air velocities, the Reynolds number along the coil surface increases which improves theNussalt number and hence the heat transfer coefficient. The improvement in heat transfer rate lowers theSCT, and hence the heat of compression. In actual applications, however, minimizing the condenser fanpower use and achieving the lowest SCT is the goal.

The following sections discuss the affects of various fan control strategies, with respect to specific target

saturated condensing temperatures, on the evaporative condenser and system performance. Again, the


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realistic objective should be to obtain an optimum balance point between a desirable SCT (resulting inless compressor work) and a minimum possible fan power consumption.

5.2 Description of Test Evaporative Condenser

The evaporative condenser used for this test had the following specifications. A photograph of the unit is

shown in Figure 21. The capacity of this unit exceeded the heat rejection requirements of this test.Consequently, to better match the refrigeration load, 1/3 of the condenser coil was valved off.

Make: Recold

Model No.: JC-30

Capacity: 441,000 Btu/hr (27o TD, R-12)

CFM: 5900

Coil: 21 tube circuits, 9.7 ft2coil surface area

Fan Motor: 3 HP, 796 RPM

Sump: 43 gallon, 0.5 HP pump, 40 GPM spray

Figure 21 - Evaporative Condenser Used for the Test


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6. EVAPORATIVE CONDENSER TESTS

6.1 Test 7 - Two Speed Fan at 65F Target Saturated Condensing Temperature

The control setting was programmed to allow the condenser fan to operate at a low or high speed withrespect to target discharge pressure and sump temperature. At condensing temperatures higher than

65F, the fan operated at high speed. With a design Temperature Difference (TD) of 15F, and a low

ambient wet-bulb temperature, the condenser capacity could be reduced by stepping down the fan speed.This condition, however, did not occur since the actual climatic conditions during this particular test

never reached sufficiently low wet bulb temperatures (roughly 50F) to allow such capacity reduction.Figure 22 depicts the wet-bulb temperature profile on 9/12-9/13/97 in conjunction with other test

parameters. During this test interval, the daily range was 14.4 F, the wet bulb temperature had an

average, minimum, and maximum of 64.7 F, 57.4F, and 71.5 F, respectively.

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Figure 22 - Important Refrigeration System Parameters vs Hours of Test for the Evaporative, Two Speed Fan, 65F Target SCTTest Run (10 minute data)

Due to the relatively high wet bulb temperatures, the condenser was incapable of meeting the target SCT.Figure 22, presenting 10 minute data, clearly indicates that the condenser power stayed constantregardless of the daily wet-bulb fluctuation. Compressor power, however, increased during the hours thatthe ambient wet-bulb temperature rose.

The relationship between SCT and condenser power with respect to ambient wet-bulb temperature isshown in Figure 23. The condenser fan runs continuously at a fixed high speed while the SCT graduallyincreases as the ambient wet bulb temperature goes up. A steep incline in SCT takes place between

57F and 61F wet bulbs. Beyond 61F wet bulb temperature, SCT increases with a smaller slope.


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Figure 23 - SCT and Condenser Power Consumption vs Ambient Temperatures for the Evaporative, Two Speed Fan, 65FTarget SCT Test Run (hourly data)


The control setting was programmed to allow the condenser fan speed to modulate from 100% to 33%CFM using a variable speed drive controller. The VSD controller slows down the fan speed with respectto target discharge pressure and sump temperature. With a design Temperature Difference (TD) of

15F and a sufficiently low ambient wet-bulb temperature, the condenser capacity could be reduced byslowing down the fan speed. Similar to the previous test, this condition never occurred since the actualclimatic conditions during this particular test did not reach sufficiently low wet-bulb temperatures

(roughly 50F), to allow such capacity reduction. Figure 24 depicts the wet-bulb temperature profile on9/8-9/9/97 in conjunction with other test parameters. During this test interval, the daily range was 11.2

F, the wet bulb temperature had an average, minimum, and maximum of 67.5 F, 62.2F, and 73.4 F,respectively.

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Compressor Power (kW) Condenser Power (kW) System Power (kW)Refri eration Load (tons) Wet Bulb Tem erature

Figure 24 - Important Refrigeration System Parameters vs Hours of Test for the Evaporative, Variable Speed Fan Drive, 65FTarget SCT Test Run (10 minute data)


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Condenser fan power remains constant at 2.8

Documents

Evaporative and AirCooled Report.pdf