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Technical Note 1667

Development and Verification of a Linear Fit Mixed

System Rating Method for Unitary Two-Speed and

Variable-Speed Air Conditioners

W. Vance Payne

U.S. DEPARTMENT OF COMMERCE

National Institute of Standards and Teclinology

Building Environment Division

Building and Fire Research Laboratory

Gaithersburg, Maryland 20899-8631

Nisr

National Institute of Standards and Technology

Technology Administration

United States Department of Commerce

QC

100

.U5753

#1667

2010

Technical Note 1667

Development and Verification of a Linear Fit Mixed

System Rating Method for Unitary Two-Speed and

Variable-Speed Air Conditioners

W. Vance Payne

U.S. DEPARTMENT OF COMMERCE

National Institute of Standards and Technology

Building Environment Division

Building and Fire Research Laboratory

Gaithersburg, Maryland 20899-8631

June 2010

*^«TE6 0*

U.S. DEPARTMENT OF COMMERCE

Gary Locke, Secretary

National Institute of Standards and Technology

Patrick D. Gallagher, Director

TABLE OF CONTENTS

Abstract 1

Acknowledgement 2

1: Introduction 3

2: Description of Evaporators and Matched System Condensing Unit 7

3: Experimental Method 7

3.1: Experimental setup 7

3.2: Data acquisition and measurement uncertainty 8

4: Two-Speed Matched and Mixed System Tests 10

5: Matched Coil Tests 11

5.1: Matched coil test results at A2, Ai, B2, Bi, Ev, and f^ conditions 11

5.2: Matched coil airflow specific cooling capacity 15

6: Mixed Coil #1 Tests 18

6.1: Mixed coil #1 test results at A2, Ai, B2, Bi, Ev, and Fi conditions 19

6.2: Mixed coil #1 airflow specific cooling capacity 23

7: Mixed Coil #2 Tests 25

7.1: Mixed coil #2 test results at A2, Ai, B2, Bi, Ev, and Fi conditions 26

7.2: Mixed coil #2 airflow specific cooling capacity 28

8: Matched Condensing Unit Tests 30

8.1: Matched condensing unit linear fits 31

8.2: Matched CD unit refrigerant massflow specific capacity 42

9: Comparison of Measured System Performance to Linear Fit Predictions 45

9.1: Calculation of capacity and EER 45

9.2: Calculation of SEER 49

10: A Discussion on Blower Efficiency 53

11: Concluding Remarks 55

References 57

Appendix A: Evaporator Coils and Condensing Unit Descriptions 58

Appendix B: Water-Cooled Condensing Unit 66

Appendix C: Water-Heated Evaporator Unit 76

Appendix D: Obtaining Data Used in this Report 84
List of Tables

1.1: Required steady-state, wet coil tests for single-,

two-, and variable-speed compressor systems 5

1 .2: Data submitted by OEMs to AHRI for two-speed

and variable-speed compressor condensing units 6

2.1: Evaporator descriptions 7

2.2: Matched and mixed system AHRI directory ratings 7

3.2.1: Measurement uncertainty 9

4.1: Matched system tests 10

4.2: Mixed system #1 tests 10

4.3: Mixed system #2 tests 10

5.1: Matched coil performance summary 11

5.1 .1 : Linear fits of matched coil-only capacity as a

function of evaporator exit refrigerant saturation temperature 12

6.1: Mixed coil #1 performance at various evaporator temperatures 18

6.1 .1 : Linear fits of mixed coil #1 , coil-only capacity as a

function of evaporator exit refrigerant saturation temperature 19

7.1: Mixed coil #2 performance at various evaporator temperatures 25

7.1 .1 : Linear fits of mixed coil #2 coil-only capacity as a

function of evaporator exit refrigerant saturation temperature 26

8.1: Matched condensing unit capacity 30

8.1 .1 : Linear fits of matched CD unit refrigerant-side capacity as a

function of OD service valve vapor refrigerant saturation temperature 31

8.1.2: Linear fits of matched CD power as a function of

OD service valve vapor refrigerant saturation temperature 32

9.1.1

9.1.2

9.2.1

9.2.2

9.2.3

9.2.4

9.2.5

System capacities, total power, and EER from the linear fit method 47

Condensing unit capacity without and with corrected Tevap 48

Bin method SEER calculated using linear fits at 10 °F superheat 50

Scaled SEER calculated using linear fits at 10 °F superheat 51

System classifications for cyclic degradation coefficient analysis 52

Categorized cyclic degradation coefficient values 52

Fexp for various mixed and matched system combinations 52

10.1: Fan efficiency for matched and mixed air handlers 53
List of Figures

1.1: Graphical illustration of the single-speed linear fit rating procedure 4

1 .2: Graphical illustration of the two-speed linear fit rating method 5

3.1.1: Experimental setup for evaporator and system testing 8

5.1.1: Matched coil total cooling capacity 13

5.1.2: Matched coil capacity at two different superheats 14

5.1.3: Coil capacity at A2 conditions for three refrigerant superheats 15

5.2.1: Linear fits to total, sensible, and latent cooling capacity

per unit airflow rate for all tests combined at a superheat of 10.0°F 16

5.2.2: Matched coil total cooling capacity per unit airflow rate at different superheats 17

6.1.1: Mixed coil #1 coil-only capacity for all liquid

refrigerant temperatures and constant superheat of 10 °F 20

6.1.2: Mixed coil #1 A2 coil-only capacity at three different refrigerant liquid temperatures 20

6.1 .3: Mixed coil #1 Ai coil-only capacity at different refrigerant liquid temperatures 21

6.1 .4: Mixed coil #1 B2 coil-only capacity at different refrigerant liquid temperatures 21

6.1.5: Mixed coil #1 Bi and Ev coil-only capacity for different refrigerant liquid temperatures .22

6.1 .6: Mixed coil #1 Fi coil-only capacity for different refrigerant liquid temperatures 23

6.2.1 : Mixed coil #1 coil-only capacity per unit airflow

rate for all liquid temperatures and a superheat of 10 °F 24

7.1.1: Mixed coil #2 coil-only capacity, high airflow, A2 and B2 conditions 27

7.1.2: Mixed coil #2 coil-only capacity, low airflow, Ai, Bi, Ev, and Fi conditions 27

7.2.1 : Mixed #2 coil-only capacity per unit airflow rate

for all liquid temperatures and a superheat of 10.0 °F 28

7.2.2: Mixed #2 coil-only capacity per unit airflow rate at different superheats 29

8.1.1: Matched CD unit A2 refrigerant-side capacity as a function of CD service

valve vapor refrigerant saturation temperature at several superheats 32

8.1 .2: Matched CD unit A2 power as a function of CD service valve

vapor refrigerant saturation temperature at several superheats 33

8.1.3: Matched CD unit Ai refrigerant-side capacity as a function of OD service

valve vapor refrigerant saturation temperature at several superheats 34

8.1 .4: Matched CD unit Ai power as a function of OD service valve

vapor refrigerant saturation temperature at several superheats 35

8.1.5: Matched CD unit B2 refrigerant-side capacity as a function of OD service

valve vapor refrigerant saturation temperature at several superheats 36

8.1 .6: Matched CD unit B2 power as a function of OD service valve

vapor refrigerant saturation temperature at several superheats 37

8.1 .7: Matched CD unit Bi refrigerant-side capacity as a function of OD service

valve vapor refrigerant saturation temperature at several superheats 38

8.1 .8: Matched CD unit Bi power as a function of OD service valve

vapor refrigerant saturation temperature at several superheats 39

8.1.9: Matched CD unit capacity for all conditions at a superheat of 10.0 °F 40
8.1.10: Matched CD unit power for all conditions at a superheat of 10.0 °F 41

8.2.1 : Matched CD unit capacity per unit of refrigerant mass flow rate at different outdoor air temperatures and constant superheat of 10.0 °F 42

8.2.2: Matched CD unit capacity per unit of refrigerant mass flow rate at different outdoor air temperatures and constant superheat of 20.0 °F 43

8.2.3: Matched CD unit refrigerant mass flow rate at high and low compressor speeds 44

9.2.1: Bin method SEER calculated with varied cyclic degradation coefficient 49

9.2.2: SEER for the two-speed systems calculated

with varied cyclic degradation coefficients 50

10.1: Matched system, high speed, blower power per unit of airflow rate as a function of external static pressure at constant airflow rate 54
Nomenclature

A EVAP-COND air-side heat transfer coefficient correction factor

A#-Test refers to AHRI Standard 210/240 test conditions of 35.0 °C (95.0 °F) outdoor air and

16.7 °C (80 °F) dry-bulb/ 19.4 °C (67 °F) wet-bulb indoor air conditions with #=^ for

low speed compressor, low speed indoor fan and #=2 for high speed compressor,

rated speed indoor fan

B#-Test refers to AHRI Standard 210/240 steady-state test conditions of 27.8 °C (82 °F)

outdoor air and 16.7 °C (80 °F) dry-bulb/ 19.4 °C (67 °F) wet-bulb indoor air conditions with #=1 for low speed compressor, low speed indoor fan and #=2 for high

speed compressor, rated speed indoor fan

Cd cyclic degradation coefficient as defined in AHRI Standard 210/240-2003

CD Unit condensing unit, the outdoor section of the split air-conditioner

CLF Cooling Load Factor as defined in AHRI Standard 210/240-2003

Diff abbreviation for difference

DOF degrees of freedom

EER Energy Efficiency Ratio as calculated in AHRI Standard 210/240-2003, (Btu/Wh)

ICM Indoor (independent) coil manufacturer

matched refers to a split air-conditioning system, an indoor section/condensing unit

combination, which rated performance is determined by laboratory testing; also may

refer to the evaporator which is used in the matched system.

mixed refers to a split air-conditioning system, an indoor section/condensing unit

combination, which rated performance is not determined by laboratory testing; also

may refer to the evaporator which is used in the mixed system.

n number of tests or number of data points

P electrical power, W

p#(82) condensing unit power at B#-Test condition (indoor fan power not included), W

P#(82) total power of air conditioner at B#-Test condition (condensing unit power plus indoor

fan power), W

AP EVAP-COND refrigerant-side pressure drop correction factor

Q Cooling capacity, W (Btu/h)
qf#(82) cooling capacity at B#-Test condition without accounting for indoor fan heat input, W

(Btu/h)

Q#(82) cooling capacity at BrTest conditions with the indoor fan heat input accounted for

,

W (Btu/h)

Q#(95) cooling capacity at A#-Test conditions without accounting for indoor fan heat input, W

(Btu/h)

0#(95) cooling capacity at A#-Test conditions with the indoor fan heat input accounted for, W

(Btu/h)

R EVAP-COND refrigerant-side heat transfer coefficient correction factor

scfm standard cubic feet per minute, equivalent to the volumetric flow rate of air with a

density of 0.075 Ib/ft^ SEER Seasonal Energy Efficiency Ratio as defined in AHRI Standard 210/240-2008,

Btu/(Wh)

SHR sensible heat ratio; the ratio of sensible capacity to total capacity

T temperature

& data standard deviation or fit standard error

SSE sum of squares of the error

ton cooling or heating capacity equal to 12 000 Btu/h or 3.517 kW

Subscripts

CD condensing unit of the split system air conditioner

coil refers to the indoor heat exchanger

eye cyclic testing

diff difference

dry dry-coil testing

evap refers to the indoor coil or evaporator at saturated refrigerant conditions

fan refers to the indoor coil blower

ID indoor

liq liquid refrigerant

mixed refers to the evaporator coil alone with respect to a system

OD outdoor

ref refrigerant

SB steady-state

suph superheat
NOTE

Use of Non-SI Units in a NIST Publication: The policy of the National Institute of Standards and

Technology is to use the International System of Units (metric units) in all of its publications.

However, in North America in the heating, ventilation and air-conditioning industry and in the

U. S. Department of Energy Test procedure referenced by this document, non-SI units are used

instead of SI units; therefore, it is more practical and less confusing to include values in customary units only.

NIST does not approve, recommend, or endorse any product or proprietary material. No

reference shall be made to NIST or to reports or results furnished by NIST in any advertising or

sales promotion which would indicate or imply that NIST approves, recommends, or endorses

any product or proprietary material, or which has as its purpose an intent to cause directly or

indirectly the advertised product to be used or purchased because of NIST test reports or

results.

Development and Verification of a Linear Fit Mixed System Rating

IVIethod for Unitary Two-Speed and Variable-Speed Air Conditioners

W. Vance Payne

National Institute of Standards and Technology

Abstract

A linear fit method of rating residential-type air conditioning systems was evaluated based on

performance predictions and laboratory testing of one two-speed matched system and two

mixed systems (matched two-speed condensing unit, matched indoor coil blower, and two

mixed coil blowers). The individual evaporators and the condensing unit were seperately tested

using water heated/cooled condensing/evaporating units over a range of evaporator refhgerant

saturation temperatures, evaporator superheats, and liquid refrigerant temperatures. Capacity

predictions were within ±1.0% of the tested values for the mixed systems, and the EER

predictions were within ±1.5 % of the measured EERs. The methods used for system rating on

the two-speed system can also be applied to a variable-speed system.

Keywords: air conditioner, cooling capacity, mixed system, rating procedure, SEER, two-speed

system, variable-speed system
ACKNOWLEDGEMENT

This study was sponsored by the United States Department of Energy, Building Technologies

Program according to interagency agreement DE-EE0001097 under project managers Wes

Anderson and Michael Raymond. Mr. David Wilmering, Mr. Thomas Montgomery, and Mr. John

Wamsley provided assistance with test setup, operating the chambers, and analyzing data.
1: INTRODUCTION

A given condensing unit (outdoor section consisting of a condenser, compressor, and

associated tubing) is typically offered on the market in several air conditioner models, which

differ by the indoor sections they employ. For all models, the manufacturers must provide

performance information, which consists of the Seasonal Energy Efficiency Ratio (SEER) and

capacity at the 95 °F rating point, Q(95). Federal regulations require that only the highest sales

volume indoor-section/outdoor-section combination, referred to as the matched system, be

tested in a laboratory to obtain the ratings (CFR 2009a). For other combinations of indoor and

outdoor sections, so called mixed systems, the federal regulations allow the use of simplified

analytical methodologies upon approval by the U.S. Department of Energy (CFR 2009b).

The most commonly used simplified methodologies for rating mixed systems are those based

upon publicly available Q(95) and SEER of the matched systems (e.g., Domanski 1989). The

application of these methods requires predicting the capacity of the matched evaporator, which

is a major shortcoming because the rater is often not familiar with the matched system product

line. Since an inaccurate prediction of the matched evaporator performance leads directly to inaccurate mixed system ratings, a different rating method that excludes this step has the

inherent potential to be a better rating approach than the one currently used. Thus NIST

developed a single-speed cooling mode linear fit rating method to allow the prediction of the

SEER and Q(95) for mixed systems (Payne and Domanski 2006). This method was shown to be able to predict SEER and Q(95) for the tested mixed systems within ±5 % (Payne and

Domanski 2005).

Figure 1.1 shows the application of the single-speed linear fit method in a graphical form. This

method uses linear fits to the cooling capacity for the mixed coil, and cooling capacities, q(82)

and q(95), and power, p(82), for the condensing unit (CD unit). The lines are presented as a

function of the evaporator exit saturation temperature. Overlapping of the evaporator and CD

unit capacities provides mixed system capacities at 82 °F and 95 °F ambient temperatures.

Projecting the saturation temperature corresponding with operation at the 82 ""F ambient

temperature on the CD unit power chart provides the power requirement for the CD unit at the

82 °F rating point.

In practice the procedure illustrated in Figure 1.1 is performed mathematically. Power and

capacity linear fits for the outdoor section are determined by the OEM who also provides

EER(95) for their matched system. The ICM then generates linear fits for cooling capacity as a

function of evaporator exit refrigerant saturation temperature, Tevap, to overlay on the OEM

provided outdoor section linear fits. Using the matched system EER(95) and the ICM calculated

mixed system EER(95) as shown in Equation 1.1, a mixed system SEER can be calculated

using Equation 1.2, where Fexp is an expansion device correction factor.

EER(95) = ^5i?5) 1.1 P(95)

CppR _ qppR '-^'^W'^/ mixed p -i 9

^'-'-'^mixed ~ '-"^'^^matched -,-r-,/oc\ exp ' -^ EER(95)^3,,,3,

The use of EER(95), instead of EER(82), to determine mixed system SEER is a simplification

necessitated by the lack of direct knowledge of the matched system's cyclic degradation

coefficient, Cq.
u

n

Q.

<z O

*•» 'E13

U)c

'35 c

0) D

C

oo

o

Q.

Saturation Temperature

Figure 1.1: Graphical illustration of the single-speed linear fit rating procedure

Two-speed and variable-speed compressor outdoor units may be paired with different indoor

units just as single-speed outdoor units are paired with various indoor units. Calculating SEER

for modulating equipment is based on a temperature bin method and requires a larger number

of test points. Table 1.1 shows the required wet coil, steady-state tests for single-speed, twospeed, and variable-speed equipment as stated in the DOE test requirements (AHRI 2008).

The linear fit method for two-speed mixed systems is graphically depicted in Figure 1.2. Linear

fits are required to describe the matched system's cooling capacity and outdoor unit power. For

the purposes of explaining the proposed procedure for modulating equipment, this section will focus on two-speed systems, but the discussed concepts still apply to variable-speed

equipment. The proposed procedure requires that the rater have the matched system ratings

along with the linear fits for power and capacity for the quantities in Table 1.2. Table 1.2 also

shows what should be included in all submittals to the AHRI database; these quantities are necessary to bring consistency with single-speed submittals and to allow linear fit comparisons

by ICMs.
Table 1.1: Required steady-state

compressor systems

, wet coil tests for single-, two-, and variable-speed

Indoor/Outdoor

Dry Bulb

Temperature

(°F)

Indoor Air

Volume Rate

Compressor

Speed

Letter Designations

SingleSpeed

Compressor

Two-Speed

Compressor

VariableSpeed

Compressor

80/95 Certified Max Max A A2 A2

80/95 Min Min

80/82 Certified Max Max B Bs B2

80/82 Min Min Bi Bi

80/87 Intermediate Intermediate Ev

80/67 Min Min Fi Fi

o

n:Q.

n: O

*^c3

O

c

«

c

0) •acoo

0)o

Q.

Matched System

q{A2) •=

q(Fi) s

Mixed Evaporator

q(B2)

Saturation Temperature

~#:

TT

§

Condenser CONDENSING UNIT

%1

Compressor

Expansion

Device

INDOOR UNIT ^ Evaporator

o

'%.

Saturation Temperature '-'

Figure 1.2: Graphical illustration of the two-speed linear fit rating method (determination of p(B2)

is shown as an example)
Table 1 .2: Data submitted {and data that should be added to submittals) by OEMs to AHRI for two-speed and variable-speed compressor condensing units

Two-speed

Capacity,

Power

Capacity,

Power

Slope Intercept Tsuoh refrigerant Indoor airflow Tliq refrigerant

B, Slope Intercept Tsuoh refrigerant Indoor airflow Tliq refrigerant

Bi Slope Intercept Tsuoh refrigerant Indoor airflow Tliq refrigerant

Fi Slope Intercept Tsuoh refrigerant Indoor airflow Tliq refrigerant

Variable-speed

Capacity,

Power

Capacity,

Power

Slope Intercept Tsuoh refrigerant Indoor airflow Tliq refrigerant

B, Slope Intercept Tsuoh refrigerant Indoor airflow Tliq refrigerant

Bi Slope Intercept Tsuoh refrigerant Indoor airflow Tliq refrigerant

Slope Intercept Tsuoh refrigerant Indoor airflow Tliq refrigerant

Fi Slope Intercept \ Tsuoh refrigerant Indoor airflow Tliq refrigerant

1 ) Ita icized and blue-font

need to be included to

entries are not currently included in submittals to AHRI, but they

bring consistency to the AHRI database

When developing a rating for a two-speed mixed air conditioner, the ICM needs evaporator

capacity as a function of evaporator exit refrigerant saturation temperature at a fixed superheat

and a fixed refrigerant liquid inlet temperature at the expansion device. Or in other words, the

ICM needs an evaporator capacity linear fit as described in Equation 1.3.

q(A2, B2, ....Fi) = Slope Tevap + Intercept 1.3

Calculating SEER involves taking the ratio of the sum of the building loads, BLQ), divided by the

sum of input energy, E(j), for/ bins as shown in Equation 1 .4, where E(j) terms include the effect

of cycling losses for those temperature bins where the system capacity exceeds the building

load. Since the information on the cyclic degradation coefficient is not available. Equation 1.4

cannot be used. Instead we may use an approach similar to that used for single-speed systems

where SEER^ixed is derived from SEERmatched by scaling it with the ratio of corresponding EERs

and multiplying by Fexp (see Equation 1.2). For multi-speed mixed systems the rating equation

for SEER will take the form of Equation 1 .5.

SEER

ZBL(j)

ZE(j)

1.4

SEEf^mixed ~ SEEf^ matched EEERj, mixed ZEER,,

exp 1.5

matched

The EERs of the mixed and matched systems would be calculated from the linear fits for capacity and power. The calculation of SEER for mixed variable-speed equipment is more cumbersome than for mixed two-speed, but it will still follow the concept given by Equations 1.4

and 1.5.
For mixed systems with a variable-speed compressor, an additional complication and effort will be required over that for two-speed systems because of the intermediate test point. The OEMs

provide linear fits for capacity and power at all mandatory test conditions as shown in AHRI

210/240-2008.

The goal of this study was to evaluate the practicality and accuracy of the linear fit method

through its application to two-speed mixed systems. In this effort, NIST assumed the role of an

evaporator manufacturer and developed cooling capacity lines for two mixed evaporator coil blowers by testing them with a water-cooled condensing unit. To remove any doubt in the CD

unit linear fits, NIST also determined linear fits for the CD unit capacity and power by testing the

condensing unit with a water-heated evaporator. NIST then developed mixed system ratings

based upon the linear fits of the various components and compared the bin-method SEER

calculation to a matched system scaled SEER calculation method (illustrated by Equation 1.5).

2: DESCRIPTION OF EVAPORATORS AND MATCHED SYSTEM

CONDENSING UNIT

Table 2.1 shows basic information on the tested evaporators. All evaporators were of the

finned-tube design. Appendix A presents detailed design data, circuitry configuration, and

pictures of the coil blowers and condensing unit. All of the evaporators tested were equipped

with a fan and required indoor fan power measurement.

Table 2.1: Evaporator descriptions

Coil

Designation

Coil

Configuration

AHRI Type

Airflow

Direction

Tube Outside

Diameter

Expansion

Device

Refrigerant

Matched A RCU-A-CB Upflow 9.5 mm (0.375 in) TXV R410A

Mixed #1 A RCU-A-CB Upflow 9.5 mm (0.375 in) TXV R410A

Mixed #2 Inclined Slab

SDHV-RCU-ACB Horizontal 9.5 mm (0.375 in) TXV R410A

The air-cooled, matched system condensing unit had a two-speed compressor and vahablespeed fan. Ratings for this condensing unit with its matched indoor air handler and the first mixed air handler (mixed #1 indoor air handler) are given in Table 2.2. The mixed system

ratings for the second air handler (mixed #2 indoor air handler) with this condensing unit were

not available, but ratings for this evaporator with a single-speed condensing unit are given.

Table 2.2: Matched and mixed system AHRI directory ratings

System

Designation

Matched

Mixed #1

Mixed #2

AHRI Type

RCU-A-CB

RCU-A-CB

SDHV-RCU-A-CB

Capacity, kW

(Btu/h)

11.13(38000)

10.60(36200)

10.43(35600)

EER(A2)

(Btu/Wh)

14.60

13.00

9.65

SEER

20.0

17.5

11.0

3: EXPERIMENTAL METHOD

3.1: Experimental setup

Figure 3.1.1 shows the experimental setup, and Appendix B shows detailed pictures of the

water cooled condensing unit. The evaporator being tested was installed in the indoor
environmental chamber. Air was pulled through the evaporator by a centrifugal fan located at the outlet of the nozzle chamber ductwork. The adjacent outdoor chamber housed the air- cooled condensing unit for system tests or the water-cooled condensing unit and the laboratory

water-chiller for evaporator tests.

For evaporator tests with the water-cooled condensing unit, the water chiller control system

manipulated the temperature and mass flow rate of the water delivered to the condensing unit. The chiller rejected heat to the in-house chilled water loop. Heat rejection was to water and did

not require maintaining the outdoor chamber conditions.

The installation of the evaporator and test instrumentation conformed to ASHRAE Standard 37

and AHRI Standard 210/240. We used the air enthalpy method for the primary measurement of

the evaporator capacity with the refrigerant enthalpy method providing the secondary

measurement. Air dew-point temperature was measured at the inlet of the evaporator ductwork

and in the ductwork after the evaporator and several mixers. Twenty-five node thermocouple

grids, located on each side of the evaporator, were used to verify that the air was well mixed at each point. A 25-junction thermopile measured the air temperature change across the

evaporator. Barometric pressure, evaporator air pressure drop, air temperature and pressure

drop in the nozzle, and nozzle temperature were used along with the dew-point measurements

to establish the thermodynamic state of the air. The refrigerant enthalpy method was the

secondary measurement of evaporator capacity and required measurement of the evaporator

inlet and exit refrigerant temperatures and pressures in addition to mass flow rate. The

agreement between the air-side and refrigerant-side methods was always within 6 %.

i

TEST EVAPORATOR

BOOSTER FAN NOZZLE CHAMBER

INDOOR CHAMBER I

<=

OUTDOOR CHAMBER

CONCENSEB WATER (.es

J •••

HOUSE WATER Po<table

Chiller

D-Jyj

WATER-COOLED

CONDENSING UNIT

v> «»

AIR-COOLED CONDENSING UNIT

Figure 3.1.1: Experimental setup for evaporator and system testing

3.2: Data acquisition and measurement uncertainty

The measurements consisted of temperature, pressure, pressure difference, temperature

difference, dew-point temperature, fan amps, fan volts, and fan power. Table 3.2.1 lists the

8
measured quantities and their uncertainties for a 95 % confidence limit (two sigma on the mean

value) (Taylor and Kuyatt 1994).

Tab e 3.2.1: Measurement uncertainty

Quantity Range Uncertainty

Pressure

kPa to 3447 kPa

(0 psia to 500 psia)

±3.4 kPa

(±0.5 psi)

Temperature

-26.1 °Cto93.3°C

(-15°Fto200°F)

±0.3 °C

(±0.5 °F)

Temperature Difference °C to 27.8 °C

(0 °F to 50 °F)

±0.3 °C

(±0.5 °F)

Barometric Pressure

mm Hgto 1270 mm Hg

(0 in Hg to 50 in Hg)

±0.34 mm Hg

(±0.0135 in Hg)

Dew-point Temperature °C to 50 °C

(32°Fto 122 °F)

±0.2 °C

(±0.4 °F)

Pressure Difference

Pa to 374 Pa

(0inH2Oto1.5inH2O)

±3.8 Pa

(±0.02 in H2O)

Mass Flow kg/h to 544.3 kg/h

(Olb/h to 1200.0 Ib/h)

±1 %

Evaporator Capacity

5.56 kW to 14.4 kW

(19kBtu/hto49kBtu/h) ±3 % to ±7 %
4: TWO-SPEED MATCHED AND MIXED SYSTEM TESTS

Tables 4.1 through 4.3 list the test conditions and test results for the matched and mixed

systems.

Table 4.1: Ma1 ched system tes1 s

Test

Designation

Indoor

Airflow,

scfm

Average Evaporator

Exit Refrigerant

Saturation

Temperature, °F

Average

Refrigerant

Liquid Line

Temperature,

°F

Total

Capacity,

Btu/h

Sensible

Heat

Ratio

EER

Btu/kWh

A2 1240 52.7 95.8 36081 0.74 12.96

Ai 942 54.1 95.1 25307 0.77 14.85

B2 1240 51.4 82.9 38897 0.71 15.73

Bi 940 52.9 82.9 27484 0.74 19.48

Table 4.2: Mixed system #1 tests

Test

Designation

Indoor

Airflow,

scfm

Average Evaporator

Exit Refrigerant

Saturation

Temperature, °F

Average

Refrigerant

Liquid Line

Temperature,

°F

Total

Capacity,

Btu/h

Sensible

Heat

Ratio

EER

Btu/kWh

A2 1215 47.1 96.5 31967 0.78 11.27

Ai 965 49.1 96.2 22300 0.82 12.11

B2 1222 46.4 83.9 34905 0.75 13.78

Bi 971 48.5 83.7 24307 0.80 15.46

Table 4.3: Mixed system #2 tests

Test

Designation

Indoor

Airflow

, scfm

Average Evaporator

Exit Refrigerant

Saturation

Temperature, °F

Average

Refrigerant

Liquid Line

Temperature,

°F

Total

Capacity,

Btu/h

Sensible

Heat

Ratio

EER

Btu/kWh

A2 750 46.3 96.6 29981 0.63 9.87

Ai 753 51.4 96.7 22263 0.72 10.23

B2 760 45.2 84.1 32652 0.62 11.99

Bi 753 50.1 84.2 23864 0.70 12.53

10
5: MATCHED COIL TESTS

The matched system's air handler was attached to a water-cooled condensing unit and tested over a range of evaporator exit saturation temperatures, evaporator exit superheats and

refngerant liquid inlet temperatures as shown in Table 5.1. These tests allowed linear fits to be

developed for cooling capacity as a function of evaporator exit refrigerant saturation

temperature at a constant superheat for the various liquid temperatures corresponding to the

standard test conditions. Since the matched system was a two-speed system, the E^ test was

not required, but data was taken for an inlet refrigerant liquid temperature near 87 °F to explore

the effects of liquid refrigerant temperature on cooling capacity and to illustrate the usefullness

of the linear fit method for variable-speed equipment.

Table 5.1: Matched coil performance summary

Test

Evaporator Exit Saturation

Temperature w/

Superheat, Low - High,

op1

Coil Only Cooling

Capacity, High -

Low, Btu/h

^

Refrigerant

Liquid

Temperature,

Low -High. °F

Range of Coil

Sensible Heat

Ratio

(47.8, 19.8)-(55.6, 5.8) 25921 -21641 94.7-95.4 0.93-0.78

Ai (46.0, 20.0) -(53.5, 9.6) 30188-21386 94.5-95.5 0.85-0.71

B. (47.9, 20.6) -(54.0, 10.2) 34455 - 24871 81.9-82.5 0.89-0.74

Bi (47.0, 9.8) -(52.5, 10.1) 33712-23526 81.9-82.4 0.80-0.69

(47.2, 10.7) -(52.7, 10.2) 32079 - 22239 87.2-87.5 0.83-0.70

Fi (45.9, 10.5) -(52.2, 10.4) 36673 - 25697 67.1 -67.4

^ - Evaporator exit refrigerant saturation temperature and superheat (Tevap- Tsuph)

^ - Capacity at the temperature conditions listed in column 2

0.78-0.67

5.1: Matched coil test results at A2, Ai, B2, Bi, Ev, and Fi conditions

Table 5.1.1 shows the linear fits for the matched coil capacity as a function of evaporator exit

saturation temperature at various superheats. It would have been better to have more than

three points to do a linear fit, but time was a limiting factor and thus a visual inspection of the

trends seen in the linear fits with varying superheats is more indicative of the relationships than

a purely numerical analysis would indicate. The uncertainty in the capacity measurement at a

given evaporator temperature is at least ±3 %; this is neglecting the uncertainty of measuring

the evaporator temperature and superheat which also adds uncertainty to the capacity

determination and the linear fit. Visually inserting these error bars onto the data points and

extending all possible lines through the resulting range of points is one technique for seeing

similarities in the linear fits that may be confounded by comparing a strict linear fit to the

available data. For example, the Bi and Ai tests in Figure 5.1.1 at a superheat of 10 °F appear

11
to have very similar slopes, but they are not numerically equal in Table 5.1.1, yet visually they

appear equal and certainly the uncertainty in their slopes would overlap due to only three points

being used in the linear fit (see Payne and Domanski 2006 for a more detailed uncertainty

analysis of linear fits).

It is interesting to note the similarities in slopes between all tests at high airflow rates and low

airflow rates; 82 and A2 have very similar slopes just as Bi, Ai, and Ev show similar slopes.

Even Fi has a similar slope to the other low airflow tests, but it appears that the negative

approach temperature (67 °F - 80 °F = -13 °F) differentiates the Fi test from the other low airflow rate tests with positive approach temperatures for this coil.

Table 5.1 .1 : Linear fits of matched coil-only capacity as a function of evaporator exit refrigerant

saturation temperature (does not include fan heat)

Test

Number

of Points

in linear fit

Slope,

Btu/(h°F)

Intercept,

Btu/h

Pearson's

Correlation

Coefficient, R^

Airflow,

scfm

Average

Blower

Power

(W)^

A2(Tsuph=10°F) 4 -2703.7 171680 0.995 1225 274

A2 (Tsuph=5 °F) 3 -3428.8 212735 0.976 1231 280

A2 (Tsuph=20 T) 3 -2634.2 160962 0.999 1232 284

Ai (Tsuph=10°F) 3 -1814.5 118362 0.999 948 121

Ai (Tsuph=20 T) 3 -1835.2 114263 0.988 945 121

B2(Tsuph=10°F) 3 -2774.5 174701 0.999 1226 278

B2 (Tsuph=20 °F) 3 -2119.2 135725 0.999 1229 279

Bi (Tsuph=10°F) 6 -1839.4 120219 0.996 946 121

E,(Tsuph=10°F) 3 -1769.0 115665 0.998 947 119

Fi (Tsuph=10T) 4 -1725.9 116067 0.997 950 121

- Total static pressure drop seen across air handler was 60 Pa (0.24 inches of water gage)

12
38000.0

36000.0

34000.0

-? 32000.0 mX 30000.0

>. 28000.0 cO

O 26000.0

24000.0

22000.0

20000.0 45 46 47 48 49 50 51 52 53 54 55 Evap Exit Tsat (°F) Figure 5.1.1: Matched coil total cooling capacity (fan heat not included, Ev is low compressor

speed)

Fi, 950 scfm, 67.3 °F Liquid

B2, 1226 scfm. 82.1 °F Liquid

—^A2, Tsuph=10F

-•-A1,Tsuph=10F

-^82, Tsuph=10F - ->^B1,Tsuph=10F

-»-F1,Tsuph=10F \ —•— Ev, Tsuph=10F -

B,. 946scfm. 82.1 °F Liquid \

X

\^\ A2. 1224 scfm, 35.4 °F Lieluid

E„, 947 scfm, 87.3 °F LiquRK ,

-v -- \ \ X-., vX ^X\ \ ^

\X \V\

"^ \

A

' nX -x \

A,, 947 scfm . 95.0^1= Uquid

56

13
Figure 5.1.2 adds a higher superheat to the data presented in Figure 5.1.1. Visually, the effect

of the higher superheat is a lowering of the linear intercept with almost constant slopes for a

given test condition; tests at comparable liquid temperatures but different superheats appear to

differ in capacity by a constant offset. This trend is illustrated again in Figure 5.1.3 for the high

airflow case with near 95 °F liquid refrigerant temperatures; three superheats are shown

indicating almost equal slopes with a constant offset in cooling capacity.

38000.0

36000.0

20000

45 46 47 48 49 50 51 52 53 54 55 56 Evap Exit Tsat (°F) Figure 5.1 .2: Matched coil capacity at two different superheats

14
36000

34000

32000

S 30000

ffi

g- 28000 o

>.co

= 26000 oo

24000

22000

20000

A2, 1230 scfm, 95.1 "F Liquid

-•-A2, Tsuph=5F

• -A2 Tsuph=10F -

-fe-A2 Tsuph=20F

A2, 1224 scfm, 95 4 °F Liquid «^

A,, 1232 scfm, 94.9 °F Liquid \^^ \ \... \ \\

A,. X \\

\ \ \\ \ \\

X

48 49 50 51 52 53 54 55 Evap Exit Tsat (T) Figure 5.1.3: Coil capacity at A2 conditions for three refrigerant superheats56

5.2: Matched coil airflow specific cooling capacity

Figure 5.2.1 shows the coil-only cooling capacity divided by the standard airflow rate as a

function of the evaporator exit refrigerant saturation temperature. All refrigerant liquid

temperatures and airflow rates are presented in this figure for a constant evaporator exit

refrigerant superheat. Sensible and latent capacity lines are broken out of the total capacity to show dehumidification performance as a function of saturation temperature.

Figure 5.2.2 shows the airflow specific capacity for the matched coil at different superheats. As

seen in the previous linear fits, the Fi test points do not strictly group with the other total capacity line, but the three points are within +5.5 % of the line. All liquid temperatures are represented and seem to produce a weak effect on capacity.

15
40.00

35.00

g 30.00

m 25.00

O 20.00

5 1500

o<

O 10.00

5.00

0.00

^ Constant Superheat of 10 °F Matched Coil Total (Btu/h)/scfm

Sensible CoB (Btu/h)/scfm

A Latent Coil (Btu/h)/scfm

^"^^.^ ^

y = -2.0512x + 132.25 ^ ^~'^'^^^* R^ = 0.9708

y = -0.e771x + 65.828 ' ** ~ -^ ^ ^ ^\

R^ = 0.9589 ~ \

A.,

A '

"

" A "'"•^^. .. A y = -1.1741x +66 427 ^ A- - - R^ = 0.9761 A -

. . A

^ A AA ..A.

All liquid temperatures (67°Fto96''F) and airflow rates (940scfmto1235 scfrn) " " -.

45 46 47 48 49 50 51 52 53 54 55 56 Evap Exit Tsat (°F) Figure 5.2.1: Linear fits to total, sensible, and latent cooling capacity per unit airflow rate for all tests combined at a superheat of 10.0 °F (coil-only capacity with no accounting

for fan heat)

16
40

35

CD> 30

<» 25

ot

<

20

15

.

Matched Coil Total (Btu/hyscfm,Tsuph = 10F

Tsupti = 5F

ATsupti = 15 F »Tsuph = 20 F

"N ^S. -4 —-_ All liquid refrigerant temperatures

'^i-Tests Tsuph=10 =F, Tliq(66 - 96) °F » ^ / \ Tsupti=20°F,Tliq(81 -96)°F

* \

A \..

>^ S..,^^ y = -2.0512x + 132.25 *s ^ ^S^^^^ R^= 9708

y=-1.9286x + 120.31 \

R^ = 0.9856 * v^ A *

45 46 47 49 50 52 53 54 55 56 57 Evap Exit Tsat (°F) Figure 5.2.2: Matched coil total cooling capacity per unit airflow rate at different superheats

(coil-only capacity with no accounting for fan heat)

17
6: MIXED COIL #1 TESTS

The mixed #1 system's air handler was attached to a water-cooled condensing unit and tested over a range of evaporator exit saturation temperatures, evaporator exit superheats and

refrigerant liquid inlet temperatures as shown in Table 6.1. These tests allowed linear fits to be

developed for cooling capacity as a function of evaporator exit refrigerant saturation

temperature at a constant superheat at the various liquid temperatures corresponding to the

standard test conditions. Since the matched system was a two-speed system, the Ev test was

not required, but data was taken for inlet refrigerant liquid temperature near 87 °F to explore the

effects of liquid refrigerant temperature on cooling capacity and to illustrate the applicability of

the linear fit method for variable-speed equipment.

Table 6.1 : Mixed coil #1 performance at various evaporator temperatures

Test

Evaporator Exit

Saturation Temperature

and Superheat, Low -

High, °F^

Coil Only Cooling

Capacity, Low -

High, Btu/h ^

Refrigerant Liquid

Temperature, Low

-High, °F

Range of

Coil

Sensible

Heat Ratio

(45.0, 10.0) -(53.2, 9.8) 34755-17130 94.8-105.4 0.99-0.76

Ai (45.8, 10.2) -(54.0, 10.0) 27682-13501 94.8-105.4 0.99-0.76

B, (46.1, 9.8) -(52.8, 10.4) 33144-16914 82.1 -88.7 0.99-0.77

Bi (43.8, 10.2) -(51.4, 10.4) 31791 -16432 82.0-88.6 0.94-0.71

(43.8, 10.2) -(51.4, 10.4) 31791 -16432 84.5-88.6 0.94-0.71

, Fi I (42.6, 10.1) -(48.0, 10.2) I 34114-22999 | 66.6-73. 3

^ - Evaporator exit refrigerant satuation temperature and superheat (Tevap, Tguph)

^ - Capacity at the temperature conditions listed in column 2

0.83-0.69

18
6.1: Mixed coil #1 linear fits at A2, Ai, B2, Bi, Ev, and Fi conditions

Table 6.1 .1 shows the linear fits for the mixed #1 coil at the various standard test conditions and

a constant superheat. Liquid refrigerant temperature entering the expansion valve was varied

around the outdoor air temperature corresponding to the given test condition; if the coil had

been connected to an air-cooled condensing unit, the refrigerant liquid temperature would be

close to or higher than the outdoor air temperature.

As seen with the matched coil, capacity slopes at high airflow rates and low airflow rates are

similar (Figure 6.1.1). Figures 6.1.2 through 6.1.6 show the weak effects of different liquid

refrigerant inlet temperatures on coil capacity.

Table 6.1 .1 : Linear fits of mixed coil #1 , coil-only capacity as a function of evaporator exit

refrigerant saturation temperature (does not include fan heat)

Test

Number of

Points in linear fit''

Slope,

Btu/(h°F)

Intercept,

Btu/h

Pearson's

Correlation

Coefficient,

R2

Airflow,

scfm

Average

Blower

Power,

A2

(Tsuph=10°F)

12 -2072.6 127404 0.994 1210 378

Ai

(Tsuph=10°F)

11 -1771.6 108723 0.995 964 225

B2

(Tsuph=10°F)

9 -2353.5 141020 0.989 1213 374

Bi

(Tsuph=10°F)

9 -1971.9 118075 0.998 967 228

Ev

(Tsuph=10°F)

6 -2001.3 119593 0.999 967 232

Fi

(Tsuph=10°F)

10 -2183.1 127818 0.998 967 226

^- Includes all refri gerant liquid 1.emperatures1 ^- External static pressure drop seen across air handler was (0.22 to 0.24) inches water gage

19
36000

34000

12000

Evap Tsat (°F) Figure 6.1.1: Mixed coil #1 coil-only capacity for all test conditions for all liquid refrigerant

temperatures and constant superheat of 10 °F

37000

32000

r 27000

2"

>

E

= 22000

17000

12000

1

y=-2144.4x + 131082

R' = 0.9996 95 °F

High Airflow, 95F Liquid

High Airflow, 99F Liquid AHigh Airflow, 105F Liquid

y = -2084.8x + 128113 *-»^

R^ = 0.998 ,05 op

1-

, , 1

1

1 r

y = -1795.7x + 113359

r2 = 0.9992 99 "F

44 45 46 47 48 51 52 53 54 55 49 50 Evap Tsat fF) Figure 6.1.2: Mixed coil #1 A2 coil-only capacity at three different refrigerant liquid temperatures

and superheat of 10 °F

20
30000.0

28000.0

26000.0

__ 24000.0 £

3st 22000.0 u

a

a.

(S O^ 20000.0 e9

e

18000.0

16000.0

14000.0

12000.0

Low Airflow, 95 F Liquid

Low Airflow. 99 F Liquid

A Low Airflow. 105 F Liquid A

* *

^ y = - 1 693x + 1 0497

-Xa^ R= = 0.9956 95 ^F

y --i745.3x+ 107742

y = -1892.4x. 114313 ^^ R^ = 0.9985 .q, c. r2 = 0.9999 99 Op - ^

105 F

A

45 46 47 48 49 50 51 Evap Tsat (°F)52 53 54 55

Figure 6.1.3: Mixed coil #1 Ai coil-only capacity at different refrigerant liquid temperatures

35000

33000

31000

29000

S 27000

g- 25000 o

c9 23000

21000

19000

17000

15000

Higti Airflow, 82 F Liquid

High Airflow 85 F Liquid

AHigh Airflow, 88 F Liquid

^ *V^ y= -2216 2x + 134415 "*« R2= 9858 ^^'^

V \,

S.' N^

y=-23644x. 141650 ^ >V V

= -25(M.9x ^ 148237 R^= 9991 88 °F ^>sV R^ = 993 85 F

Ta

45 46 47 48 49 50 51 Evap Tsat fF)

52 53 54 55

Figure 6.1.4: Mixed coil #1 B2 coil-only capacity at different refrigerant liquid temperatures

21
35000

33000

31000

29000

£ 27000

S- 25000 O

9 23000

21000

19000

17000

15000

-

Low Airflow, 82 F Liquid

Low Airflow. 85 F Liquid

A Low Airflow. 88 F Liquid

^

.V y = -1919.6x + 115437 " , R2 = 9997 82 °F %

\

'V* y = -1989.7x + 119070 „ V^ R^ = 0.9997 "^ "^

y = -201 1.8X + 120090 ^ ^*V

R2 =0.9978 ^^°^ N^\

42 43 44 45 46 47 48 49 50 51 52 53 54 55 Evap Tsat (°F) Figure 6.1.5: Mixed coil #1 Bi and Ev coil-only capacity for different refrigerant liquid

temperatures

22
35000

33000

_ 31000 m

>. ^ 29000

(OQ.nU

>. 9 27000

25000

23000

-2240.2X + 130295

R^ = 0.9985

67 °F

Low Airflow, 67 F Liquid

Low Airflow, 70 F Liquid ALow Airflow, 73 F Liquid

y = -2109.5x + 124569

F^ = 0.9989 70 °F

y = -2189.1x + 128068

R^ = 0.9998 73 °F

21000

42 43 44 45 46 47 48 49 Evap Tsat {°F)

Figure 6.1.6: Mixed coil #1 Fi coil-only capacity for different refrigerant liquid temperatures

50

6.2: Mixed coil #1 airflow specific cooling capacity

As seen with the matched coil, mixed #1 coil airflow specific capacity was very linear with

evaporator exit refrigerant saturation temperature. Figure 6.2.1 shows total, sensible and latent

airflow specific capacity for all liquid temperatures and airflow rates at a constant superheat.

The Fi test does not stand out for mixed #1 coil as it did for the matched coil. This difference

may be due to a coil circuiting or coil geometry effect (face velocity, etc.).

23
38

3634

3230

2826

E 24

>. 22 3 20 m

> 18 I

16

oo 14

12

10

42

A -

Mixed #1 Latent (Btu/ti)/scfm

Mixed #1 Sens (Btu/h)/scfm

AMixed #1 Total Coil Capacity/scfm

A y=-2 1012x + 124 73

A R^= 0.9673 ^ AA A " AA Constant Superheat of 1 °F A^

^ -*^ _ /^ ^ ^ A ^ ~- --b-^bL A ^^^i A A

y = -0 9685X + 66.045 ^»-~.

~- -1- ^ A R' = 9588 -^ *- --. ^ A A

''"'*•-->

^**^--s^ y

= -1 1327X + 58 684 -^ "" "*

-1l» *

R^ = 9612 ^ ^

^" •^J

43 44 45 46 47 48

EvapExitTsatC^F)

49 50 51 52 53

Figure 6.2.1 : Mixed coil #1 coil-only capacity per unit airflow rate for all liquid temperatures and

a superheat of 10 °F (SHR=1.0 @ 51.8 °F)

24
7: MIXED COIL #2 TESTS

Table 7.1 shows the range of evaporator temperatures tested and the resulting cooling

capacities and sensible heat ratios for all of the tests performed. The mixed #2 coil was part of a small duct, high velocity air handler. In addition to operating at external static pressures

greater than 1 .2 inH20, the sensible heat ratios for this air handler were lower than the matched

and mixed #1 coils at comparable evaporator saturation temperatures.

Table 7.1 : Mixed coil #2 performance at various evaporator temperatures

Test

Evaporator Exit Saturation

Temperature w/ Superheat,

Low -High, °F

'

Coil Only Cooling

Capacity, Low -

High, Btu/h ^

Refrigerant Liquid

Temperature, Low

-High, °F

Range of

Coil

Sensible

Heat Ratio

(47.1, 10.2) -(53.7, 10.0) 31121 -20214 94.6-105.4 0.80-0.66

Ai (44.9, 9.8) -(52.1, 10.4) 28753-19487 94.6-105.6 0.73-0.63

Bp (46.0, 10.3) -(52.5, 11.6) 33117-22755 81.7-88.3 0.77-0.66

Bi (44.2, 10.0) -(50.6, 10.5) 30216-21453 88.0-88.1 0.72-0.63

(44.2, 10.0) -(50.6, 10.5) 30216-21453 88.0-88.1 0.72-0.63

Fi I (41.4, 10.3) -(49.0, 10.5) | 33244-23758 66.7-73.2 0.69-0.61

- Evaporator exit refrigerant saturation temperature and superheat (Tevap, Tguph)

^ - Capacity at the temperature conditions listed in column 2

25
7.1: Mixed coil #2 linear fits at A2, Ai, B2, Bi, Ev, and Fi conditions

Table 7.1.1 shows the linear fits for the mixed #2 coil at airflow rates and liquid refrigerant

temperatures corresponding to the standard test conditions with a constant evaporator exit

refrigerant superheat. As shown in Figure 7.1.1 and 7.1.2, liquid refrigerant temperature had a weak effect on coil capacity even for the low temperature liquid refrigerant tests.

Table 7.1 .1 : Linear fits of mixed coil #2 coil-only capacity as a function of evaporator exit

refrigerant saturation temperature (does not include fan heat)

Test

Number of

Points in linear fit^

Slope,

Btu/(h°F)

Intercept,

Btu/h

Pearson's

Correlation

Coefficient,

R^

Airflow,

scfm

Average

Blower

Power,

A2(Tsuph=10T) 11 -1631.3 107936 0.99 761 561

Ai (Tsuph=10T) 9 -1394.7 91813 0.99 607 483

B2(Tsuph=10°F) 11 -1717.5 112592 0.99 763 566

Bi(Tsuph=10°F) 2 -1366.4 90651 1.0 617 491

Ev(Tsuph=10°F)

Same as

Bi

Fi (Tsuph=10°F) 7 -1336.0 88883 0.99 607 484

Includes all refrigerant liquid temperatures near the test condition's outdoor air temperature

Total external static pressure drop seen across air handler was 1 .8 inches of water gage or greater

26
36000

34000

32000

? 30000 S

Q. 28000 u

26000

24000

22000

20000

Mixed Coil #2, High Airflow (754scfm to 772 scfm)

Total Coll Capacity at different liquid

^. refrigerant inlet temperatures

72 Fto 89 F LIq Temps

A94 Fto 106 F Liq Temps

OAII Liquid Temps —

\

^N y=-1697.6x + 111524 All Points ^N R2 = 0.9896

V.

44 45 46 47 48 49 50 51 Evap Exit Tsat (°F)52 53 54 55

Figure 7.1.1: Mixed coil #2 coil-only capacity, high airflow, A2 and B2 conditions

36000

34000

32000

30000

mt 28000

26000

24000

22000

20000

18000

Mixed Coil #2, Low Airflow (595 - 622 scfm)

Total Coil Capacity at different liquid refrigerant inlet temperatures. Constant Tsuph = 10 °F

67 F Liquid Refrig A 70 F to 73 F Liq Temps

9 88 F to 106 F Liq Temps o All Liquid Temps

Linear (All Liquid Temps)

•-^^ •

y = .1353.5x. 89798 ^„p^j^j^

R^ = 0.9917

40 42 44 46 48 50 52 Evap Exit Tsat (°F) Figure 7.1.2: Mixed coil #2 coil-only capacity, low airflow, Ai, Bi, Ev, and Fi conditions

54

27
7.2: Mixed coil #2 airflow specific cooling capacity

Figure 7.2.1 shows the airflow rate specific capacity for the mixed #2 coil as a function of

evaporator exit refrigerant saturation temperature. All approach temperatures are represented

well by this linear fit; there is no offset for the Fi tests as was seen in the matched coil tests.

Figure 7.2.2 shows the effect of different superheats on the airflow specific capacity. The

results are very linear at the various superheats; there is only an offset between the various

superheats.

65

60

55

50 c 45 3

ffi r 40

35

E 30 u

a>Q. w 25 o5 20

5 15

10

Constant Superheat of 10 °F, Tliq(66 °F - 106 °F), SCF"M(595-901) Mixed #2 Coil Only Total (Btu/h)/scfm

Mixed #2 Latent (Btu/hyscfm A AMixed#2 Sensible (Btu/h)/scfm "^ ^ ^

y = -2.3284x + 151.63

.s. r2= 0.9845 '^v.

^^ - ^ ^ '^'^^L^

^^' X^^

•-- ..^ ^^^''•^^^-A^.,^^^,.^J^

~**-^* y = -0.9932X + 74.598

p2 = n 978'> *^-^

y = -1.3352x + 77.034 » •»-»^

R^ = 0.98!Dt) ~ ^~ »•

40 41 42 43 44 45 46 47 48 49 50 51 52 53 54 55 56 EvapExitTsat(°F)

Figure 7.2.1 : Mixed #2 coil-only capacity per unit airflow rate for all liquid temperatures and a superheat of 10.0 °F (SHR=1.0 @ 57.7 °F)

28
—•— Predicted Total / scfm at Tsupti = 10 F -•— Predicted Sensible /scfm atTsupti=10 F -A— Predicted Latent / scfm at Tsupti=1 F -O— Meas Total at Tsupti=1 F -B— Meas Sens/scfm atTsuph=10 F A Meas Lat'scfm atTsuph=10 F -— Meas Total/scfm at Tsuph=15 F -MeasTotal/scfmatTsupti=20 F Meas Sens/scfm alTsuph=15 F

- Meas Sens/scfm at Tsuph=20 F -Meas Lafscfm atTsuph=15 F

- Meas Lat/scfm at T5uph=20 F

47 Evap Exit Tsat ( F) Figure 7.2.2: Mixed #2 coil-only capacity per unit airflow rate at different superheats

29
8: MATCHED CONDENSING UNIT TESTS

The matched system condensing unit was connected to a water-heated evaporator

arrangement as shown in Appendix A. The CD unit was located in the outdoor psychrometric

chamber and air conditions were established at the various standard test conditions. Table 8.1 shows the range of tests performed with the matched system condensing unit connected to the

water-heated evaporator. The Ev test was performed at low compressor speed and low outdoor

airflow as established by the CD unit's controls.

Table 8 1: Matched condensing unit capacity

Test

OD Vapor at Service Valve

Temperature w/

Superheat, Low

-High, °F '

Refrig. Side

Cooling Capacity,

Low - High,

Btu/h 2

Refrigerant

Liquid

Temperature,

Low -High, °f

Refrig.

Subcooling at OD Service

Valve, Low -

High, °F

OD Total

Power at Conditions

in Col. 2, W

A2

(Tsuph^ 10 °F)

(46.3, 9.7)

(57.6, 10.1

34330-41911 97.7-99.4 7.6-8.6 2428 - 2558

A2

(Tsuph = 15 °F)

(43.9, 14.9)

(53.5, 14.8

33012-39386 96.9-97.6 8.3-9.5 2404-2514

A2

(Tsuph = 20F)

(43.9, 20.0)

(53.3, 20.1

33108-39398 96.6-97.0 9.1 -10.4 2417-2518

Ai

(Tsuph = 10 °F)

(47.7, 10.0)

(53.6, 10.2

23600 - 26760 97.6-99.4 4.8-7.2 1594-1579

Ai

(Tsuph = 15 °F)

(45.5, 14.9)

(57.0, 15.1

22687 - 28788 96.6-98.4 5.3-8.8 1601 -1568

Ai

(Tsuph = 20 °F)

(41.5,20.1)

(58.5, 20.0

21105-29538 96.5-97.9 6.7-8.4 1620-1558

B2

(Tsuph ^ 10 °F)

(46.4, 10.1)

(57.7, 10.1

37430-46090 83.8-84.1 10.1 -12.1 2145-2270

B2

(Tsuph = 15 °F)

(45.1, 15.2)

(56.2, 15.3

36555 - 44848 83.5-83.7 10.2-12.2 2132-2256

B2

(Tsuph = 20 °F)

(41.6, 20.2)

(52.9, 20.3

34343 - 42350 83.1 -83.3 9.9-12.0 2098-2219

1322-1306 Bi

(Tsuph ^ 10 °F)

(49.9, 10.3)

(56.2, 10.3

27550-31017 83.6-84.1 8.6-9.1

Bi

(Tsuph ^ 15 °F)

(47.2, 15.5)

(54.2, 15.2

26083 - 29809 83.2-83.6 8.6-9.5 1330-1312

Bi

(Tsuph = 20 °F)

(45.0, 19.9)

(51.9, 19.8

25092 - 28798 82.9 8.5-9.5 1336-1308

(Tsuph = 10 °F)

(44.9, 9.8)-

(51.4,9.7)

23999 - 27276 88.6-89.2 7.2-7.9 1438-1418

Fi (Tsuph = 10 °F)(45.8, 10.1)-

(51.8, 10.2)

27493 - 30737 68.7-69.2 8.5 1081 -1064

"^ - Evaporator exit refrigerant saturation temperature and superheat (Tevap, Tsuph)

^ - Capacity at the temperature conditions listed in colunnn 2 (Col 2)

30
8.1 : Matched condensing unit linear fits Tables 8.1.1 and 8.1.2 list the linear fits for refrigerant-side capacity and CD unit total power for

all the data seen in the following figures. Figures 8.1 .1 through 8.1 .8 show the effects of varied

superheat on the refrigerant-side capacity and CD unit total power as a function of refrigerant

saturation temperature at the vapor service valve. Refrigerant-side capacity, in the odd

numbered Figures 8.1.1 through 8.1.7, was a weak, but visible, function of superheat over the

ranges tested (10 °F to 20 °F). CD unit total power, in the even numbered Figures 8.1.2 through

8.1.8, also showed dependence upon superheat but with less linearity than capacity.

Figures 8.1.9 and 8.1.10 shows refrigerant-side capacity and CD unit power at all standard test conditions at a constant superheat with the points connected by straight lines (these are not

linear fits overlayed onto the points).

Table 8.1.1 : Linear fits of matched CD unit refrigerant-side capacity as a function of CD service

valve vapor refrigerant saturation temperature

Test

Superheat,

°F

Number of

Points in linear fit

Slope,

Btu/(h °F)

Intercept,

Btu/h

Pearson's

Correlation

Coefficient, R^

Average

Subcooling,

°F

A2 10 3 672.97 3184.62 0.99 8.2

A2 15 5 659.13 4033.72 0.99 8.9

A2 20 4 663.68 3952.13 0.99 9.8

Ai 10 5 543.10 -2399.960 0.99 6.0

Ai 15 5 535.11 -1720.11 0.99 7.3

Ai 20 5 500.03 222.12 0.99 7.4

B2 10 3 766.41 1870.97 0.99 10.9

B2 15 3 746.07 2895.88 0.99 11.3

B2 20 3 709.55 4791.64 0.99 10.8

Bi 10 3 551.19 -12.458 0.99 8.8

Bi 15 3 535.51 806.41 0.99 9.0

Bi 20 3 535.71 924.27 0.99 9.0

Ev 10 4 511.43 997.40 0.99 7.7

Fi 10 3 535.15 2991.85 0.99 8.4

31
Table 8.1.2: Linear fits of matched CD power as a function of OD service valve vapor

refrigerant saturation temperature

Test

Superheat,

°F

Number of

Points in linear fit

Slope,

W/°F

Intercept, W

Pearson's

Correlation

Coefficient, R^

Average

Subcooling,

T

^2 10 3 1 1 .446 1899.41 0.99 8.2

A2 15 5 11.493 1902.11 0.99 8.9

A2 20 4 10.679 1947.53 0.99 9.8

Ai 10 5 -2.763 1726.11 0.80 6.0

Ai 15 5 -2.637 1719.74 0.84 7.3

Ai 20 5 -3.549 1767.62 0.99 7.4

B2 10 3 11.046 1632.55 0.99 10.9

B2 15 3 11.129 1630.56 0.99 11.3

B2 20 3 10.747 1649.07 0.99 10.8

Bi 10 3 -2.617 1452.90 0.99 8.8

Bi 15 3 -2.549 1449.37 0.99 9.0

Bi 20 3 -4.056 1520.21 0.97 9.0

Ev 10 4 -3.238 1584.62 0.99 7.7

Fi 10 3 -2.831 1210.27 0.99 8.4

2c

43000

42000

41000

40000

39000

38000

37000

36000

35000

34000

33000

32000

Tsuph = 10°F 15 °F 20 °F

40

y = 672.97x + 3184.6 y= 659.1 3x + 4033.7 y = 663.68x + 3952.1 R^ =0.9999 R^ = 0.9985 R^ = 0.9996

AH ^/

y = 659.41 X +4030.6 y/^

Tsuph = 10 F Tsuph = 15 F A Tsuph = 20 F X A2, All Tsuph

Linear (Tsuph = 10 F) Linear (Tsuph = 1 5 F) Linear (Tsuph = 20 F)

R^ = 0.9975 yT

y^

^yf'

J^'

r

y^ ^

42 44 46 54 56 58 60 48 50 52 OD Service Valve Vapor Tsat (°F) Figure 8.1.1: Matched CD unit A2 refrigerant-side capacity as a function of OD service valve

vapor refrigerant saturation temperature at several superheats

32
2580

2560

2540

2520

5 2500 «

o

Q.CO

o

2480

2460

2440

2420

2400

2380

Tsuph = 10 T 15^ 20 °F y=11.446x + 1899.4 y= 11.493x + 1902.1 R^ = 0.9993 R^ = 0.9966

y = 10.679X + 1947.5 R^ = 0.9994

All yfr

y= 11. 02x + 1926.1 y^

R^ = 0.9894 yy

Tsuph = IGF

Tsuph = 15

F

A Tsuph = 20 F

X A2, All Tsuph

Linear (Tsuph = 10 F)

Linear (Tsuph = 15 F) - - Linear (Tsuph = 20 F)

Linear (A2, All Tsuph)

yy^

v^'' >^

.0'

^.-^4'''

' "r - 1 —"•[•-- , [ , 1 1 1 40 42 44 46 48 50 52 54 56 58 OD Service Valve Vapor Tsat (°F) Figure 8.1 .2: Matched CD unit A2 power as a function of OD service valve vapor refrigerant

saturation temperature at several superheats

60

33
32000

30000

£ 28000

o

csQ.

ra u

Tsuph =

10 °F

26000

in

~ 24000 «

te.

22000

20000

40

15 °F 20 °F

y = 543.1x-2400

R^ = 0.9932

y = 535.1 1x - 1720.1 V = 500.03x + 222.12 A1 Tests

r2 ^ Q gggg r2 = 0.9993

--

Tsuph = 10 F Tsuph = 15 F A Tsuph = 20 F X A1, All Tsuph

Linear (Tsuph = 10 F)

Linear (Tsuph = 15 F)

Linear (Tsuph = 20 F)

Linear (A1. All Tsuph)

All A

y = 508.38x- 381.23 Jii^'^'^^

R^ = 0.9912 J^^

^^

.^'

.J^^^'

^^y^

42 44 46 48 50 52 54 56 58 60 OD Service Valve Vapor Tsat (°F) Figure 8.1.3: Matched CD unit Ai refrigerant-side capacity as a function of OD service valve

vapor refrigerant saturation temperature at several superheats

34
a>o

Q.

ift'^n Tsuph=10°F 15 °F 20°F

y=-2.7626x + 1726.1 R^ = 0.8013 A

y = -2.6368x + 1719.7 y= -3.5485x + 1767.6 A1 Tests

R^ = 0.8429 R^ = 0.9948

iftin

N

All

Tsuph = 10 F a Tsuph= 15 F A Tsuph = 20 F

X AI.AIITsuph

Linear (Tsuph = 10 F) _

Linear (Tsuph = 1 5 F) - - - Linear (Tsuph = 20 F)

ifinn ^^N^

y=-3.2629x+ 1752.5

R' = 0.9363

i(^Qn -

"^ - '^W"

i*^fin - ^V^*

1*^7^

ic;fin - Removed Tsuph=10 °F w/ Tsat higher than 55 "P^^s^^

1550 -

A

1 1 1 1 1 1 1 1 1 40 42 44 46 48 50 52 OD Service Valve Vapor Tsat (°F)

54 56 58 60

Figure 8.1 .4: Matched CD unit Ai power as a function of OD service valve vapor refrigerant

saturation temperature at several superheats

35
48000

46000

44000 -

Tsuph= 10°F

3*^GO>.o

IDa

42000

40000 -

5 38000

O)

36000

34000

32000

48 50 52 OD Service Valve Vapor Tsat (°F) Figure 8.1 .5: Matched CD unit B2 refrigerant-side capacity as a function of OD service valve

vapor refrigerant saturation temperature at several superheats

36
Tsuph=10°F

2280

2260

2240

2220

2200

2180

2160

2140

2120

2100

2080

15°F 20 T

y=11.046«* 1632.6 R'.l

y = 11.129x + 1630.6

R= = 1 y= 10.747X + 1649.1 R^ = 0.9968 B2 Tests

Tsuph = 10 F

Tsuph = 15 F A Tsuph = 20 F

X B2. All Tsuph

Linear (Tsuph = 1 F)

Linear (Tsuph = 1 5 F)

Linear (Tsuph = 20 F)

All y= 10.939X + 1639.4 R^ = 0.9986 /x

y^

y^

y^

y^

y^

1 1 1 1 \ 1 1

1

'

1 40 42 44 46 48 50 52 54 56 58 OD Service Valve Vapor Tsat (°F) Figure 8.1 .6: Matched CD unit B2 power as a function of OD service valve vapor refrigerant

saturation temperature at several superheats

60

37
32000

31000

30000

29000 -

28000

u

<0a.

ra O

o

•am

**c

§ 27000

26000

25000

24000

Tsuph= 10 °F 15 T 20 "F

y = 551.19x- 12.458 y = 535.51x + 806.41 y = 535 71 x + 924.27 R^ = 0.9985 r2 = 1 R^ = 0.9993 B1 Tests

A. Tsuph = 10 F

Tsuph = 15 F A Tsuph = 20 F X BI.AIITsuph

Linear (Tsuph = 10 F)

Linear (Tsuph = 1 5 F)

Linear (Tsuph = 20 F)

All j/^

y = 530.69x + 1098.5 >"

R^ = 0.9985 >

Linear (Bi , All Tsupli) jTy

jr

yy^

1 1 1 1 1 1

1 1 1 40 42 44 46 48 50 52 54 56 58 60 OD Service Valve Vapor Tsat (°F) Figure 8.1 .7: Matched CD unit Bi refrigerant-side capacity as a function of OD service valve

vapor refrigerant saturation temperature at several superheats

38
1345

1340

1335

1330

I

1325 o

Q.§ 1320 Q

O

1315

1310

1305

Tsuph= 10 °F 15 °F 20 °F

1300

y = -2.6168X + 1452.9

R^ = 0.9982

y = -2. 5488x + 1 449. 4 y = -4. 0558x + 1 520.

2

R^ = 0.9903 r2 = 0.9748 B1 Tests

4\

Tsuph = 10 F

Tsuph = 15 F A Tsuph = 20 F

X BI.AIITsuph

Linear (Tsuph = 10 F)

Linear (Tsuph = 1 5 F)

Linear (Tsuph = 20 F)

Linear (B1. All Tsuph)

>>. \ y = -2 7834X + 1460.6

S, R'= 0.9213 "%v^

x^

'^^^..

^^

&

\. \

42 44 46 48 50 52 54 56 58 OD Service Valve Vapor Tsat (°F) Figure 8.1 .8: Matched CD unit Bi power as a function of OD service valve vapor refrigerant

saturation temperature at several superheats

60

39
50000

45000

2 40000

" 35000 a

«9

I* 30000

25000

20000

-•—A1, Tsuph = 10 F

-ii-A2. Tsuph = 10 F -e^-B1, Tsuph = 10 F --82. Tsuph = 10 F --Ev, Tsuph = 10 F -+-F1. Tsuph = 10 F

IVlatched System Condensing Unit

^-^-^^^"C--

^^-'--^^^Z--^^''^^

_^^

^^^Z^^^'C^^

40 42 44 46 48 50 52 54 56 58 OD Service Valve Vapor Tsat (°F)

Figure 8.1.9: Matched CD unit capacity for all conditions at a superheat of 10.0 °F

60

40
2800

2600

2400

2200

I

2000

a.5

ot 1800 ca

o

1600

1400

1200

-Al.Tsuph = 10 F -A2,Tsuph = 10 F -Bl.Tsuph = 10 F -B2,Tsuph = 10 F

-Ev.Tsuph = 10 F

-F1 Tsuph = 10 F

Matched System Condensing Unit

1000

40 42 44 46 48 50 52 54 56 58 OD Service Valve Vapor Tsat (°F) Figure 8.1.10: Matched CD unit power for all conditions at a superheat of 10.0 °F60

41
8.2: Matched CD unit refrigerant mass flow specific capacity (change in enthalpy)

As seen in Figure 8.2.1, condensing unit refrigerant-side capacity is a function of OD service

valve vapor refrigerant saturation temperature and superheat, outdoor air temperature and

compressor speed. A superheat increase from 10°F to 20 °F raised specific capacity by

approximately 2.2 Btu/lb. Mass flow rate was also affected by the resulting change in suction

density seen with the change in superheat.

q/mdot = flCTsat, Tsuph, Tod) see Figure 8.2.1

mdot = f2(Tsat, Tsuph, n) see Figure 8.2.3

q = (q/mdot)(mdot) = f1 • f2 = f3(7sat, T^uph, Tod, n)

8.2.1

8.2.2

8.2.3

Use of a compressor map to predict mass flow rate would allow refrigerant side capacity to be

predicted at high and low compressor speeds and associated outdoor airflow rates given the

linear fits to the data shown in Figures 8.2.1 and 8.2.2. Figure 8.2.3 shows the refrigerant mass

flow rate as a function of evaporator exit refrigerant saturation temperature at high and low

airflow rates corresponding to the standard test conditions with two different levels of evaporator

exit superheat. The addition of superheat produces a negative offset for refrigerant mass flow

rate.

84

82

£

ffi 80

6 78

76

a.

V)IcCA

(A

(0 i 74 3

Q

O

72

70

+ + 67 °F OD Air •" —

A2,Tsuph=10 F

AB2,Tsuph=10 F

xA1.Tsuph=10 F

XBI Tsuph=10 F

• Ev.Tsuph = 10 F

+ F1 Tsuph=10 F

& A

« 82''FODA]r * A A

• • • 87 °F OD Air •

95 TOD Air x x X

X

40 42 44 46 54 56 58 60 48 50 52 OD Service Valve Vapor Tsat (°F) Figure 8.2.1 : Matched CD unit capacity per unit of refrigerant mass flow rate at different outdoor

air temperatures and constant superheat of 1 0.0 °F

42
84

82

EI

80 >

'5t 78

a.g 76 o

-74

O

o

72

70

—

A2, Tsuph=20 F

AB2. Tsuph=20 F

A1,Tsuph=20 F

• B1.Tsuph=20 F

• •

82 °F OD Air • A

A

A

95°FODAir

40 42 44 46 48 50 52 54 56 58 60 OD Service Valve Vapor Tsat (°F) Figure 8.2.2: Matched CD unit capacity per unit of refrigerant mass flow rate at different outdoor

air temperatures and constant superlieat of 20.0 °F

43
750

650

E

I

550 o

t 450

350

250

A2, Tsuph=10 F

;tc A1,Tsuph=10 F A B2, Tsuph=10 F B1,Tsuph = 10 F • Ev, Tsuph=10 F X F1,Tsuph = 10 F + Low Airflow, Tsuph=10 F

- High Airflow, Tsuph = 10 F Low Airflow, Tsuph=20 F

High Airflow, Tsuph=20 F — -Linear (Low Airflow, Tsuph=10 F)

y = 9.3651x+ 48.623

r2 = 0.9923

High airflow ^^^^.^^--^"'''^^

- - - Linear (Low Airflow, Tsuph=20 F) ^^_-^--;:::::^

^_t^—

^^^-^-—"^ y = 6. 3365X + 35.847

R^ = 0.9363

y = 8.229x+ 87.252 ,

R =0.987 Low airflow „ —- -^ "^ . - - ' '

' K^ —^"^ . * -

-

^ ,.--'' ' y = 6.314X + 23.627

r2 = 0.9876

40 42 44 46 48 50 52 OD Service Valve Vapor Tsat (°F)

54 56 58 60

Figure 8.2.3: Matched CD unit refrigerant mass flow rate at liigh and low compressor speeds

44
9: COMPARISON OF MEASURED SYSTEM PERFORMANCE TO LINEAR FIT

PREDICTIONS

9.1 : Calculation of capacity and EER

With the coil capacity coefficients and CD unit capacity coefficients, the linear fit method can be

used to calculate cooling capacity and EER for the matched and mixed systems.

The calculation procedure can be implemented computationally by solving the set of two linear

equations for the evaporation temperature at which the cooling capacity of the coil equals the

cooling capacity of the CD unit:

QcD = °CD +''^CD'evap = ^coil - ^coil + ' evap^coil 9.1.1

J

^ (Bcoil -BCD

'''' (AcD-A

- 9.1.2

CD ~~'"'coil /

In the equations above, B represents the intercept and A represents the slope for the CD unit

(CD subscript) and evaporator coil (coil subscript), respectively. Applying the obtained value of

the saturation temperature into either capacity equation yields the capacity of the evaporator.

The rated cooling capacity of the system can be obtained by reducing the evaporator capacity

by the fan heat. For coils equipped with a fan, the fan heat was measured; for other coils it can

be calculated according to AHRI Standard 210/240 (AHRI 2008).

^otal = Qcoil ~ ^D fan 9.1.0

Similarly, the total power of the system can be obtained by applying the value of the evaporator

saturation temperature from Equation 9.1.2 into the condensing unit power Equation 9.1.4 and

making adjustment for the indoor fan power as shown in Equation 9. 1 .5.

PcD = ^CD + ^CD^evap 9. 1 .4

Motal - PcD + HDfan 9.1 .5

Table 9.1.1 compares the matched and mixed system tests to the linear fit calculated values.

Table 9.1.1 uses the linear fits at a constant superheat of 10.0°F. No correction was made for

pressure drop in the refrigerant vapor line; no adjustment of evaporator saturation temperature

was applied to the CD unit evaporator saturation temperature measured at the service valve.

Liquid refrigerant temperature determines the inlet enthalpy for the evaporator and thus will have some effect upon cooling capacity. This effect was simulated and empirically correlated

for R22 and R410A coils in the previous study by Payne and Domanski (2006). In that study,

the effects of liquid temperature (and superheat) were included by adjusting the apparent

evaporator temperature. In the previous single-speed linear fit method, this empirical correction

was applied to adjust the rated cooling capacity, 0(95), and to determine the CD unit power at the corrected Tgvap- In the case of two-speed and variable speed equipment, the adjustment for

liquid temperature and superheat differentce between the mixed coil and matched system

condensing unit would be applied for each standard test conditions to correct the Tgvap for each

case. The correction has the following form shown in Equations 9.1.6 through 9.1.10.

45
step 1 : Estimate the correction for the indoor section capacity equation, z^

^

N -0.123 /_ X -0.0879

'^^cor

' liq.CD

V^ liq.coil J

' suph.CD

\y suph.coil J

9.1.6

where: Tiiqco - refrigerant liquid temperature as listed for the outdoor section at the A Test

conditions (°F)

TsuphCD - refrigerant superheat at the evaporator exit as listed for the outdoor section at the A Test conditions (°F)

Tijq.coii - refrigerant liquid temperature used during the generation of the linear fit for the

indoor coil (°F)

Tsuphcoii - refrigerant superheat at the evaporator exit used during the generation of the

linear fit for the indoor coil (°F)

Step 2: Estimate the evaporator refrigerant saturation temperature at the standard test conditions, Tevap

-Icor ^coil ^CD

evap D. D.

9.1.7

'CD ^Icor '-'coil

Step 3: Improve the estimate of the correction for indoor section capacity equation, £2cor

_ ' liq.CD ' suph.CD ^2cor

\^ liq.coil J V

suph.coil J

9.1.8

where: b1 = -0.123

T

A

evap "so"

b2 = -0.0879

Tevap "50~

TevapO - evaporator refrigerant saturation temperature calculated from Equation 9.1.7,

converted to °F (if calculated in °C)

Step 4: Calculate evaporator refrigerant saturation temperature at the standard test conditions.

evap-

' evao ( / ^2cor ' ^coil ^CD

'-'cD ~ ^2cor '-'coil

9.1.9

Step 5: Calculate mixed system capacity at the standard test conditions, Qmixed

.

Qmixed ~ QCD " Ofan. mixed ~ Cqd "* DcD ' ' evap ~ Wfan. mixed 9.1.10

Table 9.1.2 applies the correction to some mixed coil linear fits determined at different

superheats. For the tests shown, the matched CD unit linear fit at 10 °F is used and the

corrected evaporator saturation temperature is calculated.

46
tU 0)

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p

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d

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p1 p1 o

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9.2: Calculation of SEER

Linear fit method SEER may be directly calculated using the bin method if the matched system

CD unit linear fits for capacity and power are provided. The rater only needs linear fits for the

mixed coil capacity at the standard conditions with corresponding indoor blower power. Table

9.2.1 shows the calculated SEER for the matched system and two mixed systems using the

linear fits at a superheat of 10 °F and providing the indoor fan power correction for the matched

system (Figure 9.2.1).

Figure 9.2.2 shows the effect of varying the cyclic degradation coefficient from a value of 0.05 to 0.25; there is a range of approximately 7.5 % with respect to the SEER values at a Cd=0.25.

375

350

325

I

300 o

Q.B 275

250

225

200

175

150

-

SCFM =1199.8

stdev = 15 7 Matched System Blower

-Linear (Matched System Blower)

- Poly (Matched System Blower) y = 352 47x + 116 29 R2 = 9812 y

yC- yf^

- yy

y = -255 13x2 + 567.53x + 77.082 ^ ^'^^^

r2 = 9935 ^y^

/>^

^y^

y^^/^

1 ———£-— , 1 1 1 1 1 1 11

1 000 050 100 150 200 250 300 350 400 450 500 560 600 650 700

Total External Static Pressure (In H2O)

Figure 9.2.1: Matched system, high speed, blower power as a function of external static

pressure at constant airflow rate

49
Table 9.2.1: Bin method SEER ca culated using linear fits at 10 ''F superheat

Type

' evap

.

°F

Indoor

Airflow, scfnn Pfe.W Ofan. Btu/h

PcD. w

q, Btu/h Q, Btu/h

Total

Power,

W

EER,

Btu/Wh

SEER,

Btu/Wh

Matched w/ Cd=0.25

A2 49.9 1241 170 580 2471 36766 36186 2641 13.70

17.64

B2 48.8 1234 170 580 2172 39279 38699 2342 16.53

Bi 50.3 943 70 239 1321 27709 27470 1391 19.74

Fi 50.0 942 70 239 1069 29754 29515 1139 25.92

Mixed #1 w/ Cd=0.25

A2 45.2 1209 391 ' 1334.1 2417 33632 32298 2808 11.50

14.26

B2 44.6 1222 392 1337.5 2125 36053 34715 2517 13.79

Bi 46.8 960 237 808.6 1330 25784 24976 1567 15.93

Fi 45.9 960 237 808.6 1080 27566 26758 1317 20.31

Mixed #2 w/ Cd=0.25

A2 45.5 750 587 2002.8 2420 33777 31774 3007 10.57

11.43

B2 44.6 760 593 2023.3 2125 36033 34010 2718 12.51

Bi 47.3 753 586 1999.4 1329 26048 24048 1915 12.56

Fi 45.9 753 586 1999.4 1080 27557 25558 1666 15.34

All data taken at 208 VAC, single-phase power.

No fan power credit given for Mixed #1 system.

0.00005 0.05 Cyclic Degradation Coefficient

Figure 9.2.2: SEER for the two-speed systems calculated with varied cyclic degradation

coefficients

50
Table 9.2.2 shows the results of using Equation 1.5 (shown below) to scale the matched system

SEER to determine the mixed systems' SEER. Fexp was determined using Tables 9.2.3 and

9.2.4 as shown in the single-speed linear fit method (Payne and Domanski 2006). The

manufacturer's rated SEER for the matched system is 20. No cyclic testing was done to determine the actual degradation coefficient, and testing voltage was 208 VAC. Once the fan

power credit is given to the matched system, and a low Cd value is used, the bin SEER

becomes 19.203 even at the lower voltage test conditions. This is within 5 % of the 20 SEER

claimed for the matched system. Mixed system #2 would have had a similar increase in SEER

if the fan power correction was performed, but the fan power curve at constant CFM was not

produced. The key point to note through all of this testing is that the performance of the

matched and mixed systems is very linear; this means the linear fit method of scaling the

matched system SEER to calculate the mixed system SEER will work.

SEER^i^gj - SEER^a,j;hed

EEERj, mixed Vffr '-p copy of 1.5

^'^'^'^j, matched Table 9.2.2: Scaled SEER calculated using linear fits at 10 °F superheat with Cd = 0.00005 and

Cd = 0.25

Type EER,

Btu/Wh Z^ERmixed ^EERmatched 'exp

X^E'^l.mixed 2^EERi rnatched

Scaled

SEER,

Btu/Wh

Bin

SEER,

Btu/Wh

% diff wrt Bin

Method

A2 13.704

75.894

Matched

75.894 1.0 100%

19.203 19.203 0.0 %

B2 16.526

Bi 19.744

16.973 16.973 0.0 %

Fi 25.920

A2 11.501

61.539

Mixed #1

75.894 1.0 81.1 %

15.571 15.577 -0.04 %

B2 13.791

Bi 15.934

13.763 14.26 -3.49 %

Fi 20.313

A2 10.568

50.976

Mixed #2

75.894 1.0 67.2 %

12.898 12.285 3.30 %

B2 12.513

Bi 12.557

1 1 .400 11.431 -0.27 %

Fi 15.338

Dougherty (2003), working with DOE and AHRI, performed a statistical analysis of

experimentally determined Cd values for a large sample of systems. He grouped the studied

systems into four basic categories shown in Table 9.2.3. The analysis of Cd values for these

four system categories produced the Cd percentiles shown in Table 9.2.4. Using the 95'*" percentile values for each system category in Table 9.2.4, in addition to Domanski's (1989)

empirical correction for time delay relays and different expansion devices, yields the Fexp values

in Table 9.2.5.

51
Table 9.2.3: System classifications 1 or cyclic degradation coefficient analysis (Dougherty 2003)

System

Category

Equalize During Off

Cycle

Indoor Fan Off Delay System

Components

A Yes No

Cap Tube

Orifice Bleed TXV

B1 No No

Non-Bleed TXV

Electronic Expansion Device

Liquid Line Solenoid

B2 Yes Yes

Cap Tube

Orifice Bleed TXV

C No Yes

Non-Bleed TXV

Electronic Expansion Device

Liquid Line Solenoid

Table 9.2.4: Categorized cyclic; degradation coefficient values (Douglierty 2003)

Percentile A B1 B2 C

99'" 0.24 0.16 0.22 0.15

95'" 0.22 0.14 0.14 0.12

90'" 0.16 0.14 0.12 0.10

85'" 0.14 0.12 0.11 0.09

80'" 0.12 0.12 0.10 0.08

75'" 0.12 0.11 0.10 0.07

70'" 0.11 0.11 0.09 0.06

60'" 0.10 0.9 0.08 0.05

50'" 0.09 0.07 0.07 0.04

Sample Size 77 58 109 78

Table 9.2.5: Fgx:, for various mixed and matched system combinations

Matched System

A B1 B2 C

Mixed

System

A 1.000 0.990 0.990 0.974

B1 1.010 1.000 1.000 0.985

B2 1.010 1.000 1.000 0.985

C 1.026 1.016 1.016 1.000

52
10: A DISCUSSION ON BLOWER EFFICIENCY

An interesting analysis of fan power and efficiency was performed by Messmer (2010). His

analysis examines the AHRI 210/240 default for fan power per unit airflow rate (0.365 W/scfm)

and how this can be related to fan static pressure rise, fan mechanical efficiency, and fan motor

efficiency. In his analysis he illustrates the some of the possible assumptions about the fan

performance that lead to a value of 0.365 W/scfm. He showed that the following three

assumptions produced the default fan power of 0.365 W/scfm:

1

)

Blower efficiency = 0.55

2) Motor efficiency = 0.55

3) External static pressure across blower = 0.94 inches of water.

Messmer pointed out that high efficiency products, with variable-speed blower motors, will easily

require lower than the default fan power. One reason for this is the higher motor efficiency (0.75

to 0.85) and "flat" nature of the fan curves resulting in a low speed mechanical efficiency of

approximately 0.50. Thus at low speed, Messmer's calculations showed a fan power per unit

airflow rate of approximately 0.157 W/scfm.

For the testing performed at NIST the matched system and mixed system #1 used a variable

speed, high efficiency blower motor, while mixed system #2 used a single-speed blower found

in small duct, high velocity systems. Table 10.1 shows high and low speed fan efficiency for the

airflow rates and fan powers seen in Table 9.2.1. As noted by Messmer, the high efficiency

blower in the matched system produced fan power per unit of airflow rate very close to his

calculations.

1fable 10.1: Fan efficiency for matched and mixed air handlers

Type

Indoor

Airflow,

scfm

Pfan.\N

External

Static

Pressure,

in H2O

Blower,

W/scfm

Blower,

scfm/W

Matched

A2 1241 279 0.24 0.22 4.45

B2 1234 277 0.24 0.22 4.45

Bi 943 113 0.24 0.12 8.35

Fi 942 112 0.24 0.12 8.41

Mixed #1

A2 1209 391 0.2 0.32 3.09

B2 1222 392 0.2 0.32 3.12

Bi 960 237 0.21 0.25 4.05

Fi 960 237 0.21 0.25 4.05

Mixed #2

A2 750 587 2.2 0.78 1.28

B2 760 593 2.3 0.78 1.28

Bi 753 586 2.2 0.78 1.28

Fi 753 586 2.2 0.78 1.28

In another attempt to look at the performance of the electronically commutated blower motor in the matched system, fan power was recorded for constant airflow rate at various external static

53
pressures. The resulting plot was shown in Figure 9.2.2 with power as the ordinate and in Figure 10.1 with power per unit airflow rate as the ordinate. The linear fit correlation coefficients

show that this blower setup is very linear over this range of external static pressures. Using the

linear fit of Figure 10.1 , the external static pressure at 0.365 W/scfm would equal 0.93 inches of

water gage. This result is as predicted by Messmer's analysis summarized above.

0.300

_ 280 --

> 260

a.I

0.240 t

<

? 0220

g 200

£ 180

0.160

S 0.140

120

100

/

Matched System Blower

-Linear (Matched System Blower)

- Poly. (Matched System Blower)

SCFM = 1199.8 stdev= 15 7

y = 0.2822x +01012 ^' R^= 9854 y' ^^

y = -0.1713x^ + 0.4266x + 0.0749 /j^^^

R^ = 9941 4 ^^/^

y^

y^

^X/ ^y X^^

0000 0050 0100 0150 0.200 0.250 300 350 400 450

External Static Pressure Drop (in H2O)

500 0550 0600 650 0.700

Figure 10.1 : Matched system, high speed, blower power per unit of airflow rate as a function of

external static pressure at constant airflow rate

54
11: CONCLUDING REMARKS

Coil cooling capacity was examined for the matched system coil and two mixed system coils. Liquid refhgerant temperature was varied and shown to have a weak effect on linear fit slopes,

but this effect may not be negligible. The previous correction method developed by Payne and

Domanski (2006) was used to correct the calculated mixed system evaporator temperature for

several example tests. For those coil linear fits determined at liquid refrigerant temperatures

and superheats different from the matched system CD unit, the corrections moved capacity in the right direction and corresponded to past trends seen with mixed system testing. Superheat

correction was applied for several examples and shown to be in the correct direction, but the

magnitude of this correction may not be sufficient in all cases. Coil manufacturers would need

to modify this superheat correction to produce better agreement in cases where superheat was

substantially different (more than 5 °F) from that used to generate the matched CD unit linear

fits. A thorough examination of subcooling should be performed for the CD unit. For the tests presented here, the CD unit charge was set at the A2 test conditions and then remained

unchanged. Subcooling will affect compressor power and thus EER, but different subcoolings were not investigated here.

An attempt was made to normalize the coil's cooling capacity by examining the ratio of coil capacity to standard airflow rate (Btu/(h scfm)). The airflow specific capacity trends, at a given

superheat for the matched and mixed coils, were extremely linear even when all liquid

temperatures were included in the figures. This type of coil capacity normalization may be

useful to determine whether a certain coil is being rated consistently as it is applied to different

manufacturer's CD units. A linear fit of airflow specific capacity at a fixed superheat may be

generated from two points (possibly taken from the AHRI database). If a particular mixed

system utilizing this coil produces airflow specific capacity that is outside an acceptable limit, then the mixed system ratings may be suspect.

A similar attempt was made to normalize condensing unit cooling capacity by dividing the

refrigerant-side capacity by the refrigerant mass flow rate. This quantity equals the change in enthalpy of the refrigerant as it passes through the CD unit [(Btu/h) / (Ib/h) = Btu/lb].

Condensing unit data showed that the mass flow rate specific capacity was approximately

constant at a fixed superheat for a given outdoor air temperature regardless of compressor

speed. Therefore, someone could easily use a compressor map to determine refrigerant mass

flow rate at a specific evaporator saturation temperature and apply this mass flow rate to the

specific capacity line at the appropriate outdoor air conditions to determine refrigerant-side

capacity. This kind of analysis was presented only as a possible means for CD unit

manufacturers to characterize their product's performance.

A cursory examination of default fan power was presented to illustrate the assumptions

necessary to produce the default 0.365 W/scfm mandated by the AHRI 210/240 test procedure.

The analysis presented by Messmer (2010) was confirmed by blower power measurements

done during this testing; high efficiency, variable-speed, ECM blower performance may be

predicted with knowledge of the motor efficiency, blower wheel mechanical efficiency, and static

pressure drop. Testing of the matched system blower at constant airflow rate showed that

power was very linear over a wide range of external static pressures. The measurements also

indicated that the default of 0.365 W/scfm greatly exceeded the matched systems 0.143 W/scfm

(linearly extrapolated down to 0.15 in H20). This gross overestimate of fan power by the default

55
value necessitates that ICM's purchase the variable-speed air handlers with which they want to rate their coil in order to determine reasonably accurate SEER values for the mixed system.

This seems to be overly burdensome especially when a more reasonable default power could

be calculated.

In lieu of performing a calculation for the default fan power, a statistical analysis of the various

adjustable speed blowers found in the AHRI database could be performed. This type of

analysis would be similar to that performed by Dougherty (2003) when he presented an analysis

of the cooling mode, cyclic degradation coefficients determined in the AHRI test program;

blower efficiency (W/scfm) could be statistically analyzed and grouped by relevant blower

parameters and characteristics. Such an analysis could incorporate adjustable speed air handlers and furnace blowers. The ease of acquiring detailed fan power data is questionable,

but this type of analysis would be foolproof in that it would only look at measured results and not

attempt a calculation of default fan power.

The results of this investigation will be used to produce a detailed test procedure similar to Payne and Domanski (2006). This type of procedure is meant to guide raters in developing a

linear fit based Alternate Rating Method (ARM) for their mixed systems. The rater is free to modify and use parts of any procedure to create an ARM that specifically applies to their

products.

56
REFERENCES

ASHRAE 1988. ANSI/ASHRAE Standard 37. Methods of testing for rating unitary air

conditioning and fieat pump equipment. American Society of Heating, Refrigerating and

Air-Conditioning Engineers. 1791 Tullie Circle NE, Atlanta, GA, USA.

AHRI 2008. Standard 210/240, Standard for unitary air-conditioning and air-source tieat pump

equipment, Air-Conditioning, Heating, and Refrigeration Institute, 4100 North Fairfax

Drive, Suite 200, Arlington, VA 22203.

CFR, 2009a. Code of Federal Regulations, Part 430, Appendix M to Subpart B, Uniform test method for measuring the energy consumption of central air conditioners, Office of the

Federal Register, National Archives and Records Administration, Washington, DC.

http://www.gpoaccess.gov/cfr/retrieve.html

CFR, 2009b. Code of Federal Regulations, Part 430.24, Subsection ml. Units to be tested.

Federal Register, National Archives and Records Administration, Washington, DC.

Dougherty, B., 2003. New defaults for the cyclic degradation coefficient used in rating air

conditioners and heat pumps, Seminar 40, ASHRAE Annual Meeting in Kansas City, KS,

July 2.

Domanski, P.A., 1989. Rating procedure for mixed air-source unitary air conditioners and heat

pumps operating in the cooling mode - revision 1, NISTIR 89-4071, U. S. Dept of

Commerce, Natn'l Institute of Standards and Technology, Gaithersburg, Maryland USA

20899.

Messmer, C.S., 2010. Personal communications. Unico System, 326 Bluff View Circle, Saint

Louis, MO 63129-5060, USA.

Payne, W.V. and Domanski, P.A., 2005. A curve-based mixed system rating method for unitary

air conditioners, NISTIR 7225, U. S. Dept of Commerce, Natn'l Institute of Standards

and Technology, Gaithersburg, Maryland USA 20899.

http://www.bfrl.nist.gov/863/HVAC/pubs/index.htm

Payne, W.V. and Domanski, P. A., 2006. Linear fit-based rating procedure for mixed air-source

unitary air conditioners and heat pumps operating in the cooling mode, NISTIR 7325, U.

S. Dept of Commerce, Natn'l Institute of Standards and Technology, Gaithersburg,

Maryland USA 20899.

http://www.bfrl.nist.gov/863/HVAC/pubs/index.htm

Taylor, B. N., and Kuyatt, C. E., 1994. "Guidelines for evaluating and expressing the uncertainty

of NIST measurement results", NIST Technical Note 1297, 1994 edition, U.S.

Department of Commerce.

57
APPENDIX A: EVAPORATOR COILS AND CONDENSING UNIT

DESCRIPTIONS

Appendix A presents specifications for the evaporators and condensing unit tested at NIST. It includes

pictures, design data, and refrigerant circuitry representations in the input format of the EVAP-COND

simulation package.

Matched System Coil

Coil Design Data

- Data for a section ^

No. of tubes in depth row ttl

:

130

No. of tubes in depth row tt2: 30

No. of tubes in depth row tt3: 30

No. of tubes in depth row tt4: 10 No. of tubes in depth row tt5:

Matched System Coil

UnitsSl Units 15? British Units

Number of slabs W

Tube data

1 Tube length in 18.75

Inner diameter in 0.3125

1 Outer diameter in 0.375 Tube pitch i in 1

Depth row pitch in 0.75

Inner surface Rifled T hermal conductivity ] B tu/(ft. h. F) 21 6. 671

Fin data

Thickness in 1 0.004

Pilch in 0.0714285

Type Louver ^^J i

Thermal conductivity Btu/(ft.h.F) 1 132. 891

Volumetric flow rate

Fan power

Cancel

fl^/min 1200 W 277

OK

Figure A1 : Matched coil description

58
Figure A2: Matched coil refrigerant circuitry

Figure A3: IVIatched coil side view

59
Mixed System Coil #1

Coil Design Data

Data for a section

No. of tubes in depth row ttl; 14

No. of tubes in depth row U2: |14 No. of tubes in depth row tt3: 14

No. of tubes in depth row tt4:

; No. of tubes in depth row tt5:

1

Mixed System Coil 81

Units-r SI Units P? British Units

Number of slabs F

Tube data Tube length

1

in |17.8 1

Inner diameter in 0.3125

1 Outer diameter in jO.375

Tube pitch in 1

Depth row pitch in |0.75

Inner surface 1 Rifled _^

Thermal conductivity Btu/(ft.h.F) j21 6.6711 1

Fin data

Thickness

Pitch Type

Thermal conductivity

in |0.004

in 0.0714284 '

j Lanced \

Btu4ft.h.F) jl 32.891

Volumetric flow rate

Fan power

Cancel

ft^/min 1200

W 393

OK

Figure A4: Mixed coil #1 description

Figure A5: Mixed coil #1 refrigerant circuitry

60
Figure A6: Mixed coil #1 side view

61
Mixed System Coil #2

Coil Desien Data

Data for a section

No. of tubes in depth row ttl

No. of tubes in depth row tt2

No. of tubes in depth row tt3

No. of tubes in depth row tt4:

No. of tubes in depth row tt5 No. of tubes in depth row tt6

15 [ir

15

15

15

15

Tube data Tube length in 33.8

Inner diameter in 0.3125

Outer diameter in 0.375 Tube pitch

Depth row pitch

in 1

in 0.G25

Inner surface Rifled _^ , Thermal conductivity Btu/(ft.h.F) 216. 671

1

Mixed Coil tt2

Units

r SI Units R? British Units

Number of slabs

Fin data 1

Thickness '"

jo. 004

Pitch in Type

Thermal conductivity B tu/(ft. h. F)

0.0G6G667 Wavy _^ 1 132.891 j

Volumetric flow rate

Fan power

Cancel

fP/min 750

W 585

OK

Figure A7: Mixed coil #2 description

--0

77 7S

<> O79 -o oso -O Q

S1 82 S3s. 0-—O61 62 O —

—o— -o o- 63 64 65

1-0-

9

o- -O o86 87 88 89-O

90

-o—-o70 71 ^ 47 -o o48 49

-O -O-

-o <: 50

-C'

~^\ 67 6i\ \6

-q JbH—0- - -^H—o o o 51V 52V -3 54\ 55 5€ 57

-O O72 73 74 75 O -0- -O 'VA -0-- o—o31 32 33 34 35 36 / 3A 38 3?v 40 41 42 43 - —o16 17 -O— —O- -o o19 20 21 22 23 24 • <>-

1 2 -o o- -o o5 6

-o o7 8

:.S 59 60 -0---0- -O ^

44 45 O

17 28 29

O

12 13 1i 15

Airflow from the bottom

Figure A8: Mixed coil #2 refrigerant circuitry

62
Figure A9: Mixed coil #2 side views

63
Matched System Condenser

Coil Design Data

Data for a section 1

No. of tubes in depth row ttl

No. of tubes in depth row tt2

No. of tubes in depth row U3

No. of tubes in depth row tt4

No. of tubes in depth row tt5

48

48

Matched Condensing Unit

Units r SI Units 17 British Units

Tube dataTube length

Inner diameter

Outer diameter Tube pitch

Depth row pitch

Inner surface

Thermal conductivity

—

h |88

in 0.275591

|in

in

0.334B4G

0.85

' in |0.736

Rifled _^

Btu/(ft.h.F) j21 6.6711

Number of slabs

Fin data -

Thickness

Pitch Type

in 0.004

0.05

Lanced

Thermal conductivity ] Btu/(ft.h.F) j 132.891

Volumetric flow rate ft*/min

Fan power |W

Cancel

3200

100

OK

Figure A10: Matched condensing unit coil description

& '» SI a MS* »""» w' '» »*'"jS f^' '5 e5' •« « "ee er"'» 9»''n n"h ti i* 75 "w W'n ^ '« «* o » "Jm ^ » r w « « »i e n »* » yfe

1 2 3 4 « % 7 e 9 t« II t: 13 U 15 ie 1! 16 19 » >l a » i» » K r » M » 31 3; 3! >* H M !7 Jl » *a *i *i «3 AJ *! *« *7 ^ Airflow from bottom

Figure A1 1 : Matched condensing unit coil refrigerant circuitry

64
kT'^? ^^^^^^ ~^^l ir%

m H^lM

B^^^^S^

Figure A12:

^i

Matched condensing unit coil pictures

65
APPENDIX B: WATER-COOLED CONDENSING UNIT

Figure B1 : Water-cooled condensing unit with variable-speed, open-drive compressor

66
Figure B2: Oil separator arrangement

67
Figure B3: Rotameter/flowmeter used to adjust water flowrate to condenser heat exchanger

68
Figure B4: Rotameter/flowmeter used to adjust water flow to subcooler heat exchanger

69
Figure B5: Right side view of water cooled condensing unit showing power contactor box, hi/lo

pressure safety switch, and manual on/off switch

70
Figure B6: Brazed plate condenser heat exchanger

71
Figure B7: Brazed plate subcooler heat exchanger

72
Figure B8: Compressor and motor for water-cooled condensing unit (guard removed)

Figure B9: Open drive compressor name plate

73
Figure B10: Variable speed drive for water-cooled condensing unit

74
wmoG

SERIAINO H02050AOOA1

(HITPUT

ZOXI

awtf 3 m Hi tf

laWMMMMMMMIMMMMIMtiMMMM^^

i m.i i K jwi

Il l '" «iii««amwMW«HMMWw«iKiiM«Hiniii^^ awe m I-4M iz

t ., 11 1 I I

- 1,1, 11 11, i. .ii .. .i.>i >, ii.,j< iyiiiiiji i i ii M»Hi«»ii<iiiiWM»i>|gggMU[ |

ilil»iJiBl Wipiw^W^^ ni iiii nff ii suxDMo nmmu

:mm mm

iMpMlHMgm ««m-v ' - ^l am mf i rni i'l uriiiii ritri iri irr

I n

I m

, WI>i f

)wri ]WL/ii >.»•DH||»illt;, i--

Figure B11: Variable speed drive name plate

Figure B12: Portable water chiller connected to house water and used to control'water

temperature fed to the water-cooled condensing unit

75
APPENDIX C: WA TER-HEA TED EVAPORA TOR UNIT

Figure C1: Water-heated evaporator unit sJiowing two evaporators and one superheater

76
Figure C2: Left water-heated evaporator showing refrigerant expansion valves in parallel

11
Figure C3: Water-heated evaporator plate heat exchanger side-view

78
Figure C4: Water and refrigerant line connections

79
Figure C5: Superheater heat exchanger

80
Figure C6: Refrigerant expansion valve connnections

81
Figure C7: Right-side view of water-heated evaporator unit

82
Figure C8: Hot and cold house water mixed before going to water-heated evaporator unit

83
APPENDIX D: OBTAINING DATA USED IN THIS REPORT

Please contact Vance Payne for a copy of the data used to generate this report.

Vance Payne

National Institute of Standards and Technology

100 Bureau Drive, MS 8631

Gaithersburg, MD 20899

Email: vance.pavne(a)nist.qov

Phone: 301-975-6663

84