ARTI-2 1 CR/605-200-50-01 

POTENTIAL BENEFITS OF SMART REFRIGERANT 

DISTRIBUTORS 

Final Report 

Date Published - December 2002 

W. Vance Payne and Piotr A. Domanski 

National Institute of Standards and Technology 

Building Environment Division 

MS 8631, Building 226, Room B114 

Gaithersburg, Maryland USA 20899 

Prepared for the 

4100 North Fairfax Drive, Suite 200, Arlington, Virginia 22203 

AIR-CONDITIONING AND REFRIGERATION TECHNOLOGY INSTlTl ITE 

Distribution A - Approved for public release; further dissemination unlimited. 

NISTIR 6946
DISCLAIMER 

This report was prepared as an account of work sponsored by the Air-conditioning and 

Refrigeration Technology Institute (ARTI) under its “HVAC&R Research for the 2 Is’ Century” 

(2 1 -CR) program. Neither ARTI, the financial supporters of the 2 1 -CR program, or any agency 

thereof, nor any of their employees, contractors, subcontractors or employees thereof - makes 

any warranty, expressed or implied; assumes any legal liability or responsibility for the accuracy, 

completeness, any third party’s use of, or the results of such use of any information, apparatus, 

product, or process disclosed in this report; or represents that its use would not infringe privately 

owned rights. Reference herein to any specific commercial product, process, or service by trade 

name, trademark, manufacturer, or otherwise, does not necessarily constitute nor imply its 

endorsement, recommendation, or favoring by ARTI, its sponsors, or any agency thereof or their 

contractors or subcontractors. The views and opinions of authors expressed herein do not 

necessarily state or reflect those of ARTI, the 2 1 -CR program sponsors, or any agency thereof. 

Funding for the 2 1 -CR program provided by (listed in order of support magnitude): 

- U.S. Department of Energy (DOE Cooperative Agreement No. DE-FC05-990R22674) 

- Air-conditioning & Refngeration Institute (ARI) 

- Copper Development Association (CDA) 

- New York State Energy Research and Development Authority (NYSERDA) 

- Refrigeration Service Engineers Society (RSES) 

- Heating, Refrigeration Air-conditioning Institute of Canada (HRAI) 

Available to the public from 

US. Department of Commerce 

National Technical Information Service 

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Springfield, VA 22 16 1 

(703) 487-4650 

Available to U.S. Department of Energy and its contractors in paper from 

U.S. Department of Energy 

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P.O. Box 62 

Oak Ridge, TN 3783 1 

(423) 576-8401 
ARTI-2 1 CW605-200-50-01 

POTENTIAL BENEFITS OF SMART REFRIGERANT DISTRIBUTORS 

Final Report 

Date Published - December 2002 

W. Vance Payne 

Piotr A. Domanski 

Prepared for the 

Under ART1 2 1 -CR Program Contract Number 605-20050 

AIR-CONDITIONING AND REFRIGERATION TECHNOLOGY INSTITUTE 
Use of Non-SI Units in a Non-NIST Publication 

It is the policy of the National Institute of Standards and Technology to use the International 

System of Units (metric units) in all of its publications. However, in North America in the 

HVAC&R industry, certain non-SI units are so widely used instead of SI units that it is more 

practical and less confusing to use measurement values for customary units only in figures and 

tables describing system performance. 
EXECUTIVE SUMMARY 

The main goal of this study was to investigate the benefits possible for finned tube refrigerant 

evaporators when refrigerant distribution was precisely controlled to produce a desired equal 

superheat in each circuit. This goal was accomplished by examining three different finned tube 

evaporators; a wavy fin, wavy-lanced fin, and a wavy-lanced fin evaporator with tube sheets 

separated. The effects of non-uniform airflow on capacity were also examined while superheat 

was controlled in each evaporator circuit. In parallel with the experimental effort, a modeling 

program was implemented and validated with the experimental results and then used to 

determine the savings in evaporator core volume possible if refrigerant distribution was 

controlled by a smart distributor. In extreme cases, the savings in core volume could be as much 

as 40 YO. 

Within the experimental part of this study, all three evaporators could avoid significant 

performance degradation using the ability to control superheat within each of the three finned 

tube circuits. As an example, with cross-counter flow configuration, uniform airflow, and exit 

manifold superheat fixed at 5.6 "C (10.0 OF), the wavy fin and wavy-lanced fin evaporator's 

capacity dropped by as much as 41 YO and 32 %, respectively, as the superheat was allowed to 

vary between the circuits. Control of superheat was shown to be even more important during 

cross-parallel refrigerant flow due to the rapid pinching of the refrigerant and air temperatures. 

For the wavy and lanced finned evaporators in cross-parallel flow, capacity dropped by 85 YO and 

78 YO as superheat changed from 5.6 "C (10.0 OF) to 16.7 "C (30.0 OF). As the coil faces were 

blocked to produce a non-uniform airflow, pressure drop through the coils increased 

1 
substantially and control of superheat was shown to restore performance. The non-uniform 

airflow tests showed that when airflow rate was held constant, the losses in capacity due to low 

airflow over a portion of the coil could be recovered to within 2% of the original uniform airflow 

capacity by controlling superheat. The more non-uniform the airflow over the coil, the more 

capacity was improved by controlling superheat. 

A combination of results obtained from laboratory tcsting and simulations indicate the influence 

of tube-to-tube heat transfer on capacity degradation. Thc impact of tube-to-tube heat transfer 

was negligible in tests with a uniform 5.6 "C (10 OF) supcrheat. but it was significant in tests 

involving 16.7 "C (30 OF) superheat. Between the two possible conduction mechanisms of heat 

transfer that may occur, longitudinal fin conduction is rcsponsible for degraded performance 

rather than longitudinal tube conduction, which has insignificant impact. The upgraded versi n 

of the EVAP5 evaporator model, which accounts for tubc-to-tube heat transfer based on tube 

temperatures, was able to predict key return bend temperatures which indicated the occurrence of 

tube-to-tube heat transfer. However, the study also confirmed that longitudinal heat conduction 

is affected by the fin design, air-side heat transfer coefficient, and moisture removal process. 

t 

.. 

11 

-~ 
ACKNOWLEDGEMENT 

This work was sponsored by the Air-conditioning and Refrigeration Technology Institute (ARTI) 

under ARTI 21 -CR Program Contract Number 605-20050. We acknowledge the comments and 

support from people associated with the sponsoring organizations and all members of the ARTI 

2 1 -CR Equipment Energy Efficiency Subcommittee. John Wamsley provided technician support 

for all of the experimental efforts, and Jong Min Choi assisted grcatly with his technical 

comments and skillful support. David Yashar contributed the analysis on longitudinal tube 

conduction and provided many helpful comments for the final draft of this rcport. 

... 

111 
TABLE OF CONTENTS 

EXECUTIVE SUMMARY ......................................................................................... i 

ACKNOWLEDGMENTS ........................................................................................... 111 

LIST OF TABLES ....................................................................................................... vi 

LIST OF FIGURES ..................................................................................................... viii 

NOMENCLATURE ..................................................................................................... xii 

CHAPTER 1 . SCOPE OF THE STUDY .................................................................. 1 

CHAPTER 2 . BACKGROUND AND LITERATURE REVIEW .......................... 3 

CHAPTER 3 . LABORATORY EXPERIMENT ..................................................... 7 

3.1 Experimental Setup .......................................................................................... 7 

3.2 Evaporators Selected for Testing ...................................................................... 11 

3.3 Test Conditions and Experimental Procedure .................................................. 14 

3.4 Experimental Results ....................................................................................... 19 

3.4.1 Cross-Counter AidRefrigerant Flow Configuration Tests .................... 19 

... 

... 

3.4.1 . 1 Non-Uniform Superheat Tests .................................................. 19 

3.4.1.2 Effects of Airflow Rate on Coil Capacity ................................. 29 

3.4.1.3 Effects of Non-Uniform Airflow on Coil Capacity .................. 31 

3.4.2 Cross-Parallel AidRefrigerant Flow Configuration Tests ..................... 48 

CHAPTER 4 . EVAPORATOR MODELING AND SIMULATIONS ................... 52 

4.1 Background on Evaporator Model EVAP5 ..................................................... 52 

4.2 Description of EVAP5 ..................................................................................... 53 

4.2.1 Modeling Approach ............................................................................... 53 . 

4.2.2 Heat Transfer and Pressure Drop Correlations ...................................... 55 
4.3 Modeling Issues ............................................................................................... 5 8 

.. 4.3.1 Refrigerant Distribution ......................................................................... 58 

4.3.2 Air-Side Heat Transfer Correlations ..................................................... 62 

4.3.3 Internal Heat Transfer in a Finned Tube Evaporator ............................. 65 

4.3.3.1 Longitudinal Tube Conduction ................................................. 66 

4.3.3.2 Tube-to-Tube Heat Transfer via Fins ....................................... 71 

4.4 Validation and Simulations with EVAP5 ........................................................ 77 

4.4.1 Validation of EVAPS ............................................................................ 77 

4.4.2 Possible Savings in Heat Transfer Area Due to 

Optimized Superheat ............................................................................. 93 

4.4.2.1 Cross-Counter Flow Configuratin with Uniform .. Air Flow Distribution ................................................................ 93 

4.4.2.2 Cross-Parallel Flow Configuration 

with Uniform Air Flow Distribution ........................................ 95 

4.4.2.3 Cross-Counter Flow Arrangement 

With Non-Uniform Air Flow Distribution ................................ 97 

CHAPTER 5 . CONCLUSIONS ................................................................................. 102 

APPENDIX A . SUMMARY OF TEST RESULTS .................................................. 104 

APPENDIX B . CAPACITY UNCERTAINTY ........................................................ 166 

APPENDIX C . USER’S INSTRUCTION FOR THE EVAP-COND 

VERSION USED IN THIS STUDY ............................................... 167 

BIBLIOGRAPHY ........................................................................................................ 178 

V 
LIST OF TABLES 

Table 3.1.1 : Essential Control Parameters .................................................................... 10 

Table 3.3.1 Experimental Test Conditions ................................................................... 14 

Table 3.3.2 Evaporator Tests ........................................................................................ 16 

Table 3.3.3 Test Number and Conditions for Each Evaporator Test ............................ 16 

Table 3.4.1.1 . 1. Non-Uniform Superheat Test Data for COIL-W and COIL-E ........... 20 

Table 3.4.1.2.1 : Capacity of the Test Evaporators at Varying Airflow Rates .............. 30 

Table 3.4.1.3.1 : COIL-W Performance with Non-Uniform Airflow ............................ 32 

Table 3.4.1.3.2: COIL-E Performance with Non-Uniform Airflow ............................. 33 

Table 3.4.2.1: Counter and Parallel Flow Performance of COIL-W and COIL-E ....... 49 

Table 4.4.1.1 : Measured and Simulated Capacities for 

COIL-W in Cross-Counter Flow Configuration .................................... 81 

Table 4.4.1.2: Measured and Simulated Capacities for 

COIL-E in Cross-Counter Flow Configuration ..................................... 82 

Table 4.4.1.3: Measured and Simulated Capacities for 

COIL-EC in Cross-Counter Flow Configuration ................................... 83 

Table 4.4.1.4a: EVAPS Validations for COIL.W, SI Units ......................................... 84 

Table 4.4.1.4b: EVAPS Validations for COIL-W, IP Units ......................................... 85 

Table 4.4.1 Sa: EVAPS Validations for CO1L.E. SI Units ........................................... 86 

Table 4.4.1.5b: EVAPS Validations for COIL.E, IP Units .......................................... 87 

Table 4.4.1.6a: EVAPS Validations for COIL.EC. SI Units ........................................ 88 

Table 4.4.1.6a: EVAPS Validations for COIL.EC, IP Units ........................................ 88 

Table 4.4.2.3.1 : Simulated Capacities and Refrigerant Distributions 

for Non-Uniform Inlet Air Velocity Profile ........................................ 99 

Table 4.4.2.3.2: Savings in Coil Volume Relative to COIL-W due to 

Optimized Refrigerant for 1 :2.5 and 1:4 Air Velocity Ratios ........... 101 

vi 

.. -__ ... . ..___
Table A.l.l: Wavy Fin Evaporators in Cross-Counter Flow ........................................ 104 

Table A.2.1: Wavy Fin Evaporator in Cross-Parallel Flow ........................................... 114 

Table A.3.1: Enhanced Fin (Wavy Lanced) Evaporators in Cross-Counter Flow ....... 119 

Table A.4.1: Enhanced Fin (Wavy-Lanced) Evaporators in Cross-Parallel Flow ........ 132 

Table A.5.1: Enhanced-Cut Fin (Wavy-Lanced) Evaporator in Cross-Counter Flow . 137 

Table A.6.1: Wavy Fin Evaporator with Non-Uniform Airflow .................................. 146 

Table A.7.1: Enhanced Fin (Wavy-Lanced) 

Evaporator with Non-Uniform Airflow ................................................... 153 

Table B . 1 : Relative Uncertainties of Two Evaporator Tests ........................................ 166 

vi i 
LIST OF FIGURES 

Figure 3.1.1 : Schematic diagram of the experimental setup ......................................... 7 

Figure 3.1.2: Condensing unit used to precisely control refrigerant conditions ........... 8 

Figure 3.2.1: COIL-EC showing separated tube depth rows ........................................ 12 

Figure 3.2.2: A schematic side view of refrigerant circuitry ........................................ 12 

Figure 3.2.3: A schematic side view of refrigerant circuitry for COIL-EC .................. 13 

Figure 3.2.4: Circuiting of all three evaporators ........................................................... 13 

Figure 3.3.1 : Capacity ratio for different shape fins relative 

to the capacity at test 9 for the wavy coil ................................................ 18 

Figure 3.4.1.1.1 : Capacity of evaporators for tests 5 through 12 .................................. 21 

Figure 3.4.1.1.2: Capacity ratio at different superheat conditions 

relative to test 9 for COIL-W and COIL-E ........................................ 21 

Figure 3.4.1.1.3: Capacity ratio for COIL-W relative 

to test 9 for wet and test 5 for dry conditions .................................... 23 

Figure 3.4.1.1.4: Capacity ratio for COIL-E relative 

to test 9 for wet and test 5 for dry conditions ................................... 23 

Figure 3.4.1.1.5: COIL.W, test 9. circuit bend 

temperatures for cross-counter flow ................................................. 26 

Figure 3.4.1.1.6: COIL.W, test 12. circuit bend 

temperatures for cross-counter flow ................................................. 26 

Figure 3.4.1 . 1 . 7. Capacity of COIL-E and COIL-EC relative 

to COIL.E, test 9 at two different airflow rates ................................ 27 

Figure 3.4.1 . 1 . 8. COIL-EC and COIL-E bend temperatures for test 9 ......................... 27 

Figure 3.4.1.1.9: COIL-EC and COIL-E bend temperatures for test 10 ....................... 28 

Figure 3.4.1.1.10: COIL-EC and COIL-E bend temperatures for test 13 ..................... 28 

Figure 3.4.1.1.1 1 : COIL-EC and COIL-E bend temperatures for test 14 ..................... 29 

Figure 3.4.1.2.1: Capacity as a function of air flowxate for wet coil conditions ......... 31 

... 

VI11 
Figure 3.4.1.3.1 : Idealized velocity profile over 

evaporator with upper half partially blocked ..................................... 32 

Figure 3.4.1.3.2: Uniform airflow velocity (fi/min) contour map for COIL-W ........... 35 

Figure 3.4.1.3.3: Non-uniform velocity (1 :1.5) (ft/min) contour map for COIL-W ..... 36 

Figure 3.4.1.3.4: Non-uniform velocity (1 :2) (ft/min) contour map for COIL-W ........ 37 

Figure 3.4.1.3.5: Relative capacities for COIL-W non-uniform airflow ...................... 38 

Figure 3.4.1 -3.6: Uniform velocity (ft/min) contour map for COIL-E ......................... 40 

Figure 3.4.1.3.7: Non-uniform velocity (1 :I . 26) (fi/min) contour map for COIL-E .... 41 

Figure 3.4.1.3.8: Non-uniform velocity (I : 1.36) (ft/min) contour map for COIL-E .... 42 

Figure 3.4.1.3.9: Non-uniform velocity (1 : 1.62) (fi/min) contour map for COIL-E .... 43 

Figure 3.4.1.3.10: Non-uniform velocity (1 : 1.75) (fi/min) contour map for COIL-E .. 44 

Figure 3.4.1.3.1 1 : Non-uniform velocity (1 :2.59) (fvmin) contour map for COIL-E .. 45 

Figure 3.4.1.3.12: Capacity relative to uniform airflow for COIL-E ........................... 47 

Figure 3.4.2.1 : Cross-parallel flow capacity comparison for 

COIL-W and COIL-E relative to their performance 

at test 9 with cross-counter flow .......................................................... 50 

Figure 4.2.1.1 : Representation of air distribution 

and refrigerant circuitry in EVAP5 ...................................................... 54 

Figure 4.3.1.1 : Possible refrigerant pressure profiles in 

a three-circuit evaporator fed by a refrigerant distributor ..................... 59 

Figure 4.3.1.2: EVAP-COND refrigerant input data 

options for evaporator simulations ....................................................... 60 

Figure 4.3.1.3: EVAP-COND input data window for 

simulations involving a refrigerant distributor .................................... 60 

Figure 4.3.1.4: EVAP-COND input data window for simulations 

with specified overall evaporator exit saturation 

temperature and superheat .................................................................. 61 

ix 
Figure 4.3.2.1: Comparison of air-side heat transfer correlations ................................ 64 

Figure 4.3.3.1 : Tested and predicted capacities for COIL-E 

and COIL-EC at 5.6 "C (10 OF) and 16.7 "C (30 OF) superheats ......... 66 

Figure 4.3.3.1.1 : Refrigerant circuitry configuration 

for the analyzed R-22 evaporator ...................................................... 68 

Figure 4.3.3.1.2: EVAP-COND window with evaporator design information ............ 69 

Figure 4.3.3.1.3: EVAP-COND window with evaporator operating conditions .......... 69 

Figure 4.3.3.2.1 : Schematic graph for longitudinal fin 

conduction between two adjacent tubes ............................................. 74 

Figure 4.3.3.2.2: Schematic of refrigerant circuitry for COIL.W, 

COIL-E and COIL-EC in cross-counter flow configuration ............. 75 

Figure 4.3.3.2.3: COIL-E measured and simulated capacities for 

tests with 5.6 "C (10 OF) and 16.7 OC (30 OF) superheats ................. 76 

Figure 4.4.1.1 : Design parameters for COIL-W ........................................................... 78 

Figure 4.4.1.2: Correction parameters for COIL-W and COIL-E ................................ 79 

Figure 4.4.1.3: Difference between simulated and measured capacities 

for all wet coil tests for COIL.W, CO1L.E. and COIL-EC ................. 89 

Figure 4.4.1.4 Difference between simulated and measured capacities 

for all dry coil tests for C0IL.W. CO1L.E. and COIL-EC ................... 89 

Figure 4.4.1.5: Tube numbering and refrigerant exit temperatures 

for individual tubes for test 9 and test 12 for COIL-W ........................ 91 

Figure 4.4.1.6: Refrigerant exit qualities for individual 

tubes for COIL-E and COIL-EC test 10 .............................................. 92 

Figure 4.4.2.1.1 : Simulation results for alternative coil designs to 

COIL-W in cross-counter flow configuration .................................... 95 

Figure 4.4.2.2.1 : Simulation results for alternative coil designs 

to COIL-W in cross-parallel configuration ....................................... 97 

Figure 4.4.2.3.1 : Air velocity profile representation for 

the 1 :3 top-to-bottom velocity ratio .................................................. 99 
Figure 4.4.2.3.2: COIL-W capacities at different air velocity ratios referenced 

to capacity at test 9 in cross-counter flow configuration .................. 100 

Figure 4.4.2.3.3 Two evaporators with a reduced number of tubes .............................. 101 

Xi 
NOMENCLATURE 

Af 

A0 

A,, 

A,, 

A, 

= finned surface area, m2 (ft2) 

= outside tube and fin area, m2 (ft2) 

= pipe mean surface area, m2 (ft2> 

= pipe outside surface area, m2 (ft2) 

= cross sectional area of the tube available for axial heat conduction, m2 (ft2) 

C,,, = minimum of mass flow rate times heat capacity for either fluid, W/K 

(Btu/(h."F) 

C, = maximum of the mass flow rate times the heat capacity for either fluid, W/K 

(B tu/( h*"F) 

Cpa 

cfm 

COP = coefficient of performance 

fpm 

hi 

hl 

hpf 

h, 

if, = latent heat of evaporation, kJ/(kg.K) (Btu/(lb."F) 

in WG = inches of water in gage pressure 

IP 

K 

L = length, m2 (ft2) 

ma = air mass flow rate, kg/s (lb/h) 

NTU = number of transfer units for the heat exchanger, dimensionless 

Q = capacity or heat transferred, W (Btu/h) 

scfm = airflow in standard cubic feet per minute where flowrate is taken at the air standard 

= specific heat at constant pressure for air, kJ/(kg.K) (Btu/(lb."F)) 

= airflow in cubic feet per minute 

= velocity in feet per minute 

= inside-tube heat-transfer coefficient, W/(m2.K), (Btu/(ft2-h."F)) 

= heat-transfer coefficient for condensate layer, W/(m2-K) (Btu/(ft2-h."F)) 

= heat-transfer coefficient for tube/fin contact, W/(m2.K) (Btu/(ft2.h-"F)) 

= air-side heat transfer coefficient, W/(m2.K) (Btu/(ft2.h."F)) 

= inch-pound or English system of units 

= material thermal conductivity, W/(m-K) (Btu/(ft-h."F)) 

density of 0.075 lbdft3 (ANSVASHRAE 5 1-1 985) 

SI = international system of units or metric system of units 
T 

t 

TXV 

U 

w 

XP 

a 

& 

@ 

h 

z 

643 

0, 

= temperature, K (OF) 

= thickness, m (ft) 

= thermostatic expansion valve 

= overall heat-transfer coefficient, W/(m2-K) 

=width, m (ft) 

= thickness of the tube wall, m (ft) 

(Btu/(ft2-h*oF)) 

= ifgw(@ - G)/(Cpa(Ta - Tw)) 

= heat-transfer effectiveness, fraction 

= fin efficiency, fraction 

= longitudinal heat conduction parameter, dimensionless 

= tube longitudinal conduction effect factor, fraction 

= humidity ratio of air at tube inlet, kgW/kg,d, (IbJlba,dy) 

= humidity ratio of saturated air at temperature of condensate wetting the tube, 

kgw&&,dry (lbdlba,dry) 

Subscripts 

a 

f 

i 

j 

nc 

wc 

r 

sim 

W 

= air 

= fin 

= inlet, inside, or tube numbering index 

=tube numbering index 

= no longitudinal conduction effects 

= considering longitudinal conduction effects 

= refrigerant 

= simulation 

= tube wall or water 

... Xlll 

1. SCOPE OF THE STUDY 

Typically, finned tube evaporators employ parallel refrigerant circuits to obtain an 

optimal refrigerant mass flux, which affects refrigerant heat transfer coefficient and 

pressure drop. Each circuit performs optimally when the superheat at its exit matches the 

desired overall superheat in the exit manifold. Circuit superheat is affected by the 

refrigerant mass flowrate and the air flowrate associated with each tube. 

Most evaporators use an inlet expansion valve with a flow distributor to control the bulk 

superheat at the evaporator exit manifold. The current practice does not embody means 

for adjusting refrigerant distribution between different circuits as needed. This means 

that non-uniform airflow or unintended pressure drops could cause some circuits to have 

excessive superheat while others may remain two-phase at the evaporator exit. In such 

situations, some circuits are inefficiently using coil area by transferring heat with 

superheated vapor instead of two-phase refrigerant. There are also mixing losses 

associated with reaching the final superheat as two-phase and superheated refrigerant mix 

in the evaporator exit manifold. 

The advances in micro electro-mechanical systems (MEMS) offers the opportunity to 

develop and place inexpensive flow control valves on each circuit of an evaporator. This 

would allow control of the refrigerant superheat at the exit of each circuit. This 

experimental investigation examines the benefits of controlling superheat by placing 

individual needle valves on each circuit of the evaporators. The study involves three 

evaporators containing three parallel refrigerant circuits in identical configurations. Two 

1 
of these evaporators equipped with enhanced (wavy-lanced) fins, respectively, were 

tested to examine the benefits of maintaining even superheat when compared to three 

different scenarios of excessive superheat. The third evaporator with wavy-lanced fins 

and separated (cut) depth rows facilitates documenting the impact of tube-to-tube heat 

conduction. Non-uniform superheats are imposed and compared to uniform superheats of 

5.6 "C (10.0 OF) and 16.7 "C (30.0 OF). Non-uniform airflow is also imposed while 

superheats are allowed to adjust naturally, and while superheats are controlled by the 

individual expansion valves. The NIST tube-by-tube evaporator model, EVAPS, is also 

modified and used to simulate the experimental results. 

The modeling part of the study discusses longitudinal tube and fin conduction and 

presents a scheme for including tube-to-tube heat transfer into a tube-by-tube simulation 

model. Validated results for an upgraded evaporator model are presented. 

2 

. ___~ 
2. BACKGROUND AND LITERATURE REVIEW 

Refrigerant incurs a phase change from the two-phase to the superheat zone in the 

evaporator. Large refrigerant mass flux not only increases the heat transfer rate, but also 

increases the pressure drop. Therefore, most refrigerant evaporators employ parallel 

circuits to provide a balanced effect on the evaporator capacity between the negative 

effect of refrigerant pressure drop and the positive effect of improved insidc-tube heat 

transfer coefficient. 

Even though all refrigerant circuits have the same inlet and outlct conditions. the 

refrigerant distribution is not uniform; the staggered tube arrangement can causc dt ffcrent 

heat transfer rates. Non-uniform refrigerant distribution is also due to thc thcrmal 

resistance in the superheat region increasing more rapidly than in the two-phase region. 

Superheated vapor at the evaporator exit is necessary to prevent liquid comprcssion and 

subsequent damage to the compressor, even though superheat reduces the pcrformance of 

the system. 

Refrigerant superheat in a given circuit is affected by the refrigerant mass flow mtc and 

the airflow rate over the coil area associated with that circuit. For a givcn air distribution 

there is one refrigerant flow rate that results in a desired superheat at thc individual 

circuit exit. When circuits are not well balanced, the target overall superheat is 3 result of 

mixing a highly superheated refrigerant and two-phase refrigerant leaving di ffcrent 

circuits. This causes significant degradation in evaporator capacity because the circuit 

with superheated refrigerant transfers less heat. 
Liang et aL(2001) conducted a numerical study of the refrigerant circuit. The governing 

equations and control volumes were presented with the simulation procedure for 

branches, tubes, and control volumes of a coil. Using the model, the heat transfer and 

fluid flow characteristics of the coils were studied. Compared to a common coil, the 

researchers found that using a complex refrigerant circuit arrangement where the 

refrigerant circuits are properly branched or joined may reduce the heat transfer area by 

around 5 % while maintaining constant capacity. The investigators experimentally 

validated 6 different refrigerant circuiting arrangements while maintaining the 

evaporators inlet and exit states. They used an R134a expansion valve inlet condition of 

40 "C (104 OF) saturated liquid with 5.0 "C (2.8 OF) of subcooling and an evaporator exit 

saturation temperature of 10 "C (50 OF) with 5.0 "C (2.8 OF) of superheat. Liang et al. 

noted that for a given evaporator load, designers must design the refhgerant circuitry to 

produce a refrigerant mass velocity that produces a maximum heat flux. Maximum heat 

fluxes vary with refrigerant circuiting due to varying levels of refrigerant pressure drop. 

Their model was able to predict evaporator capacity within 5 YO on four of the six coils 

while predicting refrigerant pressure drop to within 25 %. 

Kirby et al. (1998) experimentally investigated the performance of a 5275 W (18000 

Btuh) window air conditioner under wet and dry coil conditions with non-uniform 

airflow over the evaporator. The velocity variation over the evaporator varied by as 

much as a factor of 3, but upon correcting the non-uniformity of airflow, the investigators 

saw only a minor improvement in performance. This was a system study with no attempt 

4 
to maintain constant refrigerant states at the inlet and exit of the evaporator. A round 

disk was used to block 16 % of the central area of the evaporator while maintaining the 

original airflow. Tests were conducted with this blockage against the evaporator face and 

then moved in steps in the upstream direction. They found no capacity degradation 

greater than 2 YO. The authors noted that the blockage caused more of the evaporating 

rehgerant to exist in the two-phase state; thereby reducing the superheated area of the 

evaporator and offsetting the inability of the air velocity to compensate for the loss of 

heat transfer area. Wet-coil tests also showed very little difference in the sensible heat 

ratio with non-uniform airflow. The authors noted that any non-uniformity in airflow 

must be noted by designers and used to intelligently circuit the refrigerant to equalize 

exposure of the evaporating refkigerant to airflow. 

Chwalowski et al. (1 989) examined computer models and performed experiments using 

three different evaporators; two V-shaped evaporators with upflow and one vertical slab 

evaporator with horizontal and angled flow with respect to the approaching airstream. 

These were evaporator tests with fixed evaporator saturation pressures; therefore, these 

tests parallel the technique used in the current experimental investigation. The 

investigators determined coil face velocity for several of the configurations and noted 

non-uniformity of the airflow. Generally, the evaporator capacity varied with the 

configuration mainly as a function of the exit superheats at the two evaporator exits. As 

the coil angle with respect to the approaching airstream was varied, capacity degradations 

on the order of 20 YO were noted. Non-uniformities in superheat were produced by nonuniform airflow that produced differences in heat transfer and pressure drops within the 

5 
heat exchangers. The design of the circuitry and various splitting points within the 

evaporator produced differing capacities as a function of the coil orientation with the 

airstream. The investigators noted that none of the three evaporator models they 

considered could accurately predict capacity without some knowledge of the air velocity 

profile and refrigerant maldistribution. 

6 
3. LABORATORY EXPERIMENT 

3.1 Experimental Setup 

Figure 3.1.1 shows a schematic diagram of the experimental setup. The test rig consisted 

of three major flow loops: (1) a refrigerant flow loop containing a detachable test 

section, (2) a water flow loop used for the condensation heat exchanger and (3) an air 

flow loop used for the evaporation heat exchanger. The design of the rig allowed easy 

control of operating parameters such as condensing pressure and subcooling at the inlet 

of the expansion valve (evaporator inlet enthalpy), evaporating pressure at the exit of the 

Loop 

evaporator, and superheat at the exit of the evaporator. 

Liquid-cooled 

Mass flow meter 

(1) Refrigerant 

Flow Loop 

Filter drier 

Variable speed 

compressor 

Needle valve Turbine flow meter 

II 

Blower Air Flow 

Thermocouple Dew point Thermocouple Sensor Differential 

transducer Dew point grid pressure grid 

I sensor 

II I 

Figure 3.1.1 : Schematic diagram of the experimental setup 

7 
Figure 3.1.2 is a photo of the specially designed and constructed R22 condensing unit. 

The design of this condensing unit allowed complete control of the subcooled R22 liquid 

conditions. 

Figure 3.1.2: Condensing unit used to precisely control refrigerant conditions 

An open-drive compressor with a variable speed motor provided refrigerant mass flow, 

set the enthalpy entering the test section, set the condensing pressure, and set the 

subcooling at the inlet of the expansion valves. We controlled condensing pressure by 

adjusting cooling water flow using a hand-operated needle valve. To provide additional 

pressure control for the condenser, we also controlled the entering temperature of the 

cooling water. Water flow rate and temperature through the subcooler plate type heat 

8 
exchanger controlled the refrigerant subcooling at the inlet of the expansion valve. A 

loop supplying water to the condensing heat exchanger and subcooler consisted of a 

refrigerator, water storage tank, and a pump. The controls combined to produce 

evaporator inlet quality of 25 YO f 1 YO. 

The test section has three parallel refrigerant paths. The exit pressure of the evaporator 

was adjusted by a pressure regulating valve which was installed in the refrigerant line. 

The superheats of the three circuits of the evaporator were adjusted by three manual 

expansion valves. The design of the test section allowed easy installation and 

replacement of the evaporators. The pressure and temperature were measured upstream 

and downstream of the test rig. Flow conditions were also monitored using a sight glass 

at the inlet of the compressor and expansion valves. 

The total refrigerant flow rate was measured by a Coriolis-type mass flow meter in the 

liquid line between the subcooler and the expansion valves. Two turbine flow meters 

were installed to measure flow rate in two of the circuits. Flow through the third circuit 

was calculated by subtracting the flow through two of the circuits from the total mass 

flow. 

Air flow rate was measured in the air flow chamber according to ANSVASHRAE 5 1- 

1985. Evaporator capacity was calculated using the air enthalpy method and refrigerant 

enthalpy method following procedures specified in ASHRAE Standard 37 (1998). In the 

present experiments, the maximum difference between the air and rehgerant side 

capacity was less than 5 YO. Air velocity at the face of the evaporator was measured with 

a hot wire anemometer. 
Table 3.1.1 lists the parameters controlled to produce a successful evaporator test. The 

parameters are listed in the order they were set to produce a controlled test. 

Evaporator Circuit Superheats 

Evaporator Exit Saturation Temperature 

Table 3.1.1 : Essential Control Parameters 

Parameter I Setpoint 

5.6 "C or 16.7 "C (10.0 "F or 30.0 OF): 

controlled by expansion/nccdlc valve 

opening and evaporator exit pressure 

7.2 "C (45.0 OF): controllcd by evaporator 

pressure regulator valve and compressor 

40.6 "C (1 05.0 OF): controlled by condenser 

water flowrate and compressor speed Upstream Liquid Saturation Temperature 

8.3 "C (1 5.0 OF): controlled by upstream 

pressure and refrigerant charge Liquid Line Subcooling 

I speed 

Evaporator Inlet Liquid Enthalpy 

Corresponds to the saturated liquid 

temperature of 40.6 "C ( 105.0 OF): when 

inlet pressures were increased. the inlet 

enthalpy was always monitored to produce 

an enthalpy equal to the saturated liquid 

enthalpy at 40.6 "C (1 05.0 OF) k 1.4 "C 

(2.5 "F) 
3.2 Evaporators Selected for Testing 

We used three finned tube heat exchangers of the same outside dimensions, tube spacing, 

and circuitry as the test evaporators: (1) COIL-W with wavy fins, (2) COIL-E with wavylanced (enhanced) fins, and (3) COIL-EC (Figure 3.2.1 ) with wavy-lanced (enhanced) 

fins and the tube rows separated to inhibit tube-to-tube heat transfer (enhanced-cut). 

Figures 3.2.2, 3.2.3, and 3.2.4 show the side views of the refrigerant circuits. The 

following are the main design parameters: 

(a) 3 depth rows with 25.4 mm (1 in) face spacing and 22.0 mm (0.866 in) row 

spacing 

(b) 3 refrigerant circuits as shown in Figures 3.2.1 and 3.2.2 

(c) 9.53 mm (0.375 in) diameter round copper tubes, smooth walls, 0.254 mm 

(0.0 10 in) wall thickness 

(d) 0.1 143 mm (0.0045 in) thick aluminum fins; wavy fins for COIL-W and 

louvered or slit fins for COIL-E and COIL-EC 

11 
Figure 3.2.1: COIL-EC showing separated tube depth rows 

Figure 3.2.2: A schematic side view of refrigerant circuitry 

12 
1234567 8910llI2l314151617L8 

Cuts in fins separating different depth rows. r 

Figure 3.2.3: A schematic side view of refrigerant circuitry for COIL-EC 

Ink :ts 

13 
3.3 Test Conditions and Experimental Procedure 

Table 3.3.1 lists the test parameters and environmental chamber conditions for tests on 

the three evaporators. All tests were conducted at the same 26.7 "C (80.0 OF) indoor drybulb with dew-point varying for wet-coil tests and dry-coil tests. Refrigerant R22 

conditions at the inlet to the expansion valves were controlled to maintain an enthalpy 

equivalent to a 48.9 "C (120.0 OF) saturation temperature with a subcooling of 8.3 "C 

(1 5.0 OF) k 1.4 "C (2.5 OF). Some tests required increasing the inlet pressure to produce 

the required superheats at the evaporator circuit exits. When the pressure was increased, 

the enthalpy and subcooling were adjusted to keep a constant enthalpy at the evaporator 

inlet. 

Table 3.3.1 : ExDerimental Test Conditions 

Variable I Value I Tolerance ~~ 

Indoor Dry-Bulb I 26.7 "C (80.0 OF) ~ I 0.28 "C (0.5 OF) 

Indoor Dew-Point for WetCoil Tests 

15.8 "C (60.4 OF) I 0.28 "C (0.5 OF) 

Evaporator Exit Saturation 1 7.2 "C (45.0 OF) ~ I 0.28 "C (0.5 OF) 

Temperature 

EvaporatorExpansion 48.9 "C (120.0 OF) 1.4 "C (2.5 OF) 

Valve Inlet Saturation 

Temperature 

Evaporator/Expansion 8.3 "C (15.0 OF) 1.4 "C (2.5 OF) 

Valve Inlet Subcooling 

Once an evaporator coil was mounted in the airflow chamber, indoor dry-bulb and dewpoint were stabilized for at least one hour. While indoor psychrometric conditions 

stabilized, the evaporator inlet R22 pressure and temperature were set by adjusting the 
flow control valves on the condensing unit. Water flow to the condenser and subcooler 

plate heat exchangers was adjusted to establish the evaporator expansion valves inlet 

pressure and temperature. The evaporator exit saturation temperature was set by 

adjusting the evaporator pressure regulating valve at the exit of the evaporator. Superheat 

conditions in the individual circuits were set by adjusting R22 mass flow through each 

circuit. Airflow rate over the evaporator was adjusted using the variable speed drive on 

the airflow chamber's pull-thru fan. 

Table 3.3.2 and 3.3.3 list the tests performed for the evaporators. Capacity specific 

airflow rate was initially established at 193 m'/kWh (400 scfm/ton) for test 9. Test 9 

required manipulating superheats and airflow rate to obtain the desircd airflow to 

capacity ratio. Test 9 capacity was then used to calculate the airflow mtc for the 

169 m3/kWh (300 scfm/ton) and 242 m3/kWh (500 scfm/ton) tests. Tests with ;I crossparallel flow configuration were performed by switching the refrigerant flow. 

Non-uniform airflow tests were performed with COIL-W and COIL-E. We established 

non-uniform air distribution by attaching a series of metal mesh plates to the upper half 

of the coil. 

A hot wire anemometer was used to measure airflow rate by traversing the coil at a 

minimum of 25 equally spaced points at the face of the coil. This measurcmcrit apeed 

with the chamber airflow within 2 %. 

15 
Evaporator 

COIL-W (wavy fins), cross-counter flow 

COIL-W (wavy fins), cross-parallel flow 

COIL-E (lanced fins), cross-counter flow 

COIL-E (lanced fins), cross-parallel flow 

COIL-EC, cross-counter flow 

Tests Performed 

1, 5-13 (10 tests) 

9- 12 (4 tests) 

1, 2, 5-14 (12 tests) 

9-12 (4 tests) 

1,2, 5,6,9, 10, 13, 14 (8 tests) 

Table 3.3.3: Test Numb( 

Ir I 

COIL-W, cross-counter flow, non-uniform 

airflow 

COIL-E, cross-counter flow, non-uniform 

airflow 

* 

9 (1/2 profile, no superheat adjustment), 9 (1/2 

profile, superheat adjusted), 9 (1 /3 profile, no 

superheat adjustment), 9 (1 /3 profile, superheat 

adjusted) [4 tests] 

9 (1/2 profile, no superheat adjustment), 9 (1/2 

profile, superheat adjusted), 9 (1/3 profile, no 

superheat adjustment), 9 ( 1 /3 profile, superheat 

adjusted) [4 tests] 

and Conditions for Each Evaporator Test 

Overall Superheat 

Superheats in Individual Circuits 

1 

2 

3 

5 I 

6 

7 

I 

9 

T++= 

14 

ired overall level of suDerheat is obtainec 

5.6 "C 

(1 0.0 OF) 

*/16.7/16.7 

(*/30/30) 

4 

8 

12 

16 

1) SI units of m3/kWh multiplied by capacity (Q) in kW to determke airflow, (IP units of cfdton 

multiplied by capacity (Q) in tons). 

16 
In total we performed 54 tests with uniform airflow and 28 tests with imposed nonuniform distribution of air. We conducted a total of 90 tests including repeats and tests 

that were excluded due to unsteady or non-standard conditions. 

The capacity characteristics of the three evaporators are shown below. Figure 3.3.1 

shows the capacity ratio at different test conditions to the capacity at test 9 for the wavy 

coil. Tests 1 , 9, and 13 are wet coil tests, and test 5 is a dry coil test. COIL-E and COILEC evaporators represented higher capacity than that of the COIL-W evaporator for the 

wet coil tests. Even though the air-side sensible heat transfer coefficient is much lower 

than the refrigerant side, the air-side thermal resistance is reduced due to the 

enhancement of moisture condensation and large finned area. For wet coil tests (tests 1, 

9, 13), the capacity of COIL-E and COIL-EC was larger than that of COIL-W. The 

COIL-EC produced higher capacity than COIL-E possibly because of the added fin 

leading edges agitating the boundary layer and increasing the air-side heat transfer 

coefficient even further in addition to eliminating some tube-to-tube heat transfer. 

17 
Q, 

v) 

0) 

- 1.40 

2 -- 1.20 

s 

0 1.00 c, 

0 Wavy 

rn Enhanced 

0 Enhanced-Cut 

c, 

0 

fn 

0.80 .. 

L 

c 0.60 

'Z 

c, 0.40 

2 

c, 

.- E 0.20 

0 tu 

P 

0.00 

1 

__ 

I 

5 9 13 

Figure 3.3.1: Capacity ratio for different shape fins relative to the capacity at test 9 for 

the wavy coil (Test 1 : low airflow,wet coil, Test 5: median airflow, dry 

coil, Test 9: median airflow, wet coil, Test 13: high airflow, wet coil). 

All tests have h 5.6 "C (10.0 OF) uniform superheat. 

18 
3.4 Experimental Results 

3.4.1 

3.4. I. I Non- Uniform Superheat Tests 

Figure 3.4.1.1.1 shows capacity at different superheat test conditions. These data are 

shown in Table 3.4.1.1.1. All of the coils showed a rapid decrease in capacity when 

Cross-Counter AidRefrigerant Flow Configuration Tests 

individual circuit superheat was increased with the overall superheat maintained at 5.6 "C 

(10 OF). Figure 3.4.1.1.2 shows the relative capacity of COIL-W tests 10 and 12 with 

respect to test 9. Even though the overall superheat was held at 5.6 "C (10 OF), the nonuniformity of superheat in cases 1 1 and 12 produced a 4 1 YO loss in capacity. This was 

almost as severe as the 43 % loss in capacity seen when overall superheat was held at 

16.7 "C (30.0 OF). 

Test 12 showed similar capacity as test 10 even though overall superheat was lower. At 

test condition 12, the mass flow rate through the top circuit was much higher than that at 

test 10. Therefore, the inlet refrigerant temperature of the evaporator for test 10 was 

higher than test 12 because exit pressure was the same. This means that the temperature 

difference between air and refrigerant for test 12 was higher than for test 10. This 

allowed test 12 to have a higher capacity than test 10. 

19 
P 

Table 

W020225B 

W020228A 

W02022 1 A 

W020225 A 

W020207B 

W020530A 

W02053 1 A 

W0202 15B 

W020322A 

W0203218 

W020322C 

W 0203228 

E020607A 

W0203 18A 

W0203 188 

W0203 19A 

Superheat tl 

Volumetric Flowrate of 

Cross-Counter AidRefriperant Flow 

0.6C 

0.77 

e controlled such that the desired overall level of superheat is obtain 

I a) SI units of m3/kWh multiplied by capacity (Q) in kW to determine airflow, (IP units of cfdton 

1 b) Capacity relative to the 5.6 "C (1 0.0 OF) tests noted by a rqtio of 1 in the row above. 

multiplied by capacity (Q) in tons). 
27000 I 

4 COIL-w 

COIL-E 

A COIL-EC 

M COIL-W Dry 

X COIL-E Dry 

__ 

L 

x 11000 

g 

W 

7000 9g 400 sdm/ton Dry 

3000 5

1000 

c 

r 

4 5 6 7 0 

0 

: 

16293 

4 2293 

400 sdrnlton Wet 

t 1293 

1293 

9 10 11 12 

Test Number 

Figure 3.4.1.1.1 

I 

: Capacity of evaporators for tests 5 through 12 

Q, .I 

.~ 

0 

m 

P 

9 

c, 

u) 

Q, 

IO 

U 

1 

0.8 

0.6 

.- 3 0.4 

0.2 

0 

Test 9 

~~ 

COIL-E 

Test 10 Test 12 

Figure 3.4. I. 1.2: Capacity ratio at different superheat conditions relative to test 9 for 

COIL-W and COIL-E (superheat cases follow Table 3.4.1.1.1). 

21 
For COIL-E, the reduction in capacity due to an increase in superheat was lower than 

COIL-W (Figure 3.4.1.1.1 and 3.4.1.1.2). COIL-E seemed to show a preference as to 

which flooded circuit (test 11 or 12) produced the higher capacity. When the middle 

circuit was flooded, the capacity decreased by 21 % compared to 32 % when the top 

circuit was flooded. 

Figure 3.4.1.1.3 shows the relative capacity of COIL-W for tests 10 and 12 with respect 

to test 9 and for tests 6 and 8 with respect to test 5. The capacity of the dry coil 

decreased by 28 % with non-uniform circuit superheats with overall superheat fixed at 

5.6 "C (10 OF). COIL-E capacity (Figure 3.4.1.1.4) dropped by 16 % with non-uniform 

superheat under dry conditions; again COIL-E showed that flooding the middle circuit 

produced a smaller capacity drop than flooding the top circuit while holding overall 

superheat constant at 5.6 "C (10 OF). 
1 .I 

1 

0.4 

% 0.3 

0 

0.2 

0.1 

0 

m 

Test 9 and 5 Test 10 and 6 Test 12 and 8 

Figure 3.4.1.1.3: Capacity ratio for COIL-W relative to test 9 for wet and test 5 for dry 

conditions 

1 

0.9 

0.8 

s 0.7 

.L 0.6 

2 0.5 

2 0.4 

0.3 

0.2 

0.1 

0 

Q, * 

Ial 

(P 

C 

- 

0 

m 

0 

Test 9 and 5 Test 10 and 6 Test 12 and 8 

Figure 3.4.1.1.4: Capacity ratio for COIL-E relative to test 9 for wet and test 5 for dry 

conditions 

The experimental investigation was designed to reveal some of the effects of tube-to-tube 

heat transfer by heat conduction through the fin material. The comparison was done by 

23 
examining COIL-E and COIL-EC and comparing their capacity at different levels of 

superheat. In addition to the capacity comparison between COIL-E and COIL-EC, direct 

evidence of conduction between tubes was noted from the thermocouple bend 

temperature data for COIL-W. 

Figures 3.4.1.1.5 and 3.4.1.1.6 show bend temperature data for test 9 (6.7 "C (10 OF) 

superheat on all circuits) and test 12 (flooded top circuit with 16.7 "C (30.0 OF) superheat 

on the other two circuits). The figure for test 9 shows a uniform temperature distribution 

between comparable bends in the three refrigerant circuits. The noticeable difference 

occurs for test 12 where the low temperature, flooded top circuit shows some thermal 

communication with the final tube passes of the middle circuit. The top circuit is 

showing an average surface temperature of approximately 9.4 "C (49 OF) throughout all 

of its tubes, while the middle circuit shows definite superheat at the third and fourth final 

tube bend with a temperature of 24.0 "C (75.2 OF). This is where the conduction between 

circuits was obvious; the surface temperature on the final two tubes bend was 22.3 "C 

(72.2 OF). This is a decrease in temperature due to conduction between the top circuit's 

tube and the middle circuit's tubes. 

The conduction effects were quantified in the tests conducted with COIL-E and COILEC; by separating the tube sheets in COIL-EC and thereby removing a majority of the 

conduction path between tubes. Figure 3.4.1.1.7 shows the capacity of COIL-E and 

COIL-EC relative to test 9 for COIL-E during cross-counter flow for tests 9, 10, 13, and 

14. These two coils used identical fin material and fin type; the only difference was the 

24 
tube sheets of COIL-EC were separated. Figure 3.4.1.1.8 shows that for test 9, COIL-E 

and COIL-EC have very similar bend temperatures. Figure 3.4.1.1.9 shows the same 

coils with the superheat increased to 16.7 "C (30.0 OF) at the coil exit (test 10). COIL-EC 

shows lower inlet temperatures than COIL-E even though expansion valve inlet and coil 

exit conditions are almost identical for both tests. The differences in temperatures seen 

with tests 9 and 10 are more pronounced for tests 13 and 14 at the higher airflow rate. 

These differences in temperatures could have produced the differences in capacity seen 

between COIL-E and COIL-EC. As noted above, the inlet and exit conditions for these 

coils were nearly identical, but COIL-EC always showed lower bend temperatures than 

COIL-E. This would mean that COIL-EC was operating at a higher average temperature 

difference with respect to the air than COIL-E. The greater average temperature 

difference for COIL-EC could translate to higher capacity than COIL-E. The test results 

showed that both coils produced very similar capacities when the overall superheat was at 

5.6 "C (10.0 OF). As the superheat was increased to 16.7 "C (30.0 OF), COIL-EC capacity 

was 10 % higher than COIL-E. As the airflow increased for tests 13 and 14, COIL-E and 

COIL-EC still produced nearly equal capacities at 5.6 "C (10.0 OF) superheat, but when 

superheat was increased to 16.7 "C (30.0 OF), COIL-EC had a 23 % higher capacity than 

COIL-E (Figures 3.4.1.1.10 and 3.4.1.1.1 1). This tends to lend more evidence to 

conduction effects between the tube sheets; eliminating some conduction paths improved 

the performance of the enhanced fin coil. 

25 
______- -Du 

Exit Refrigerant C0IL.W. Test 12, Bend Temperatures+) 

-b 416 

- - -. - - - -bU ........ 

........ -D722 

e; 

_______ ........ 

-D 736 

........ 

______. 

Figure 3.4.1.1.6: COIL-W, test 12, circuit bend temperatures for cross-counter flow (Top 

circuit flooding with 16.7 "C (30.0 OF) superheat on bottom two circuits 

to yield overall exit superheat of 5.6 "C (10.0 OF); refrigerant exit 

manifold saturation temperature set to 7.2 "C (45.0 OF)) 

- 55.0 
y 500*Q scfmhon 

400 scfrdton 1 .I 

-= 0.8 m 

a, 

p? 

0.7 

0.6 

0.5 

9 10 

Test Number 

13 14 

Figure 3.4.1.1.7: Capacity of COIL-E and COIL-EC relative to COIL-E, test 9 at two 

different airflow rates 

Coil-EC. Test 9. Bend Temperatures (OF) 

I Exit I , Exit Relrigerant 

Coil-E Test 9. Bend Temperatures (OF) 

.... IT....... 81 ....... b ... 48.9 

I u :I 

IQ ______ . 50.6. 51.4 .......... Q 118.3, 48.5 

< >.. 

56 0 

55 8 

Figure 3.4.1.1.8: COIL-EC and COIL-E bend temperatures for test 9 (cross-counter 

flow, wet coil, 5.6 "C (10.0 OF) superheat on all circuits) 

27 
Exit Refrigerant Coil-EC Test 10, Bend Temperalures ( OF) Coil-E Test 10. Bend Temperalures ( OF) 

- - - - - b 76.1 ....... ........ 

48.8 ---.-. - -D ..... -.----- -b ...... 

* 75.8 ....... --.--- - b 75.9 

...... 

72.1 

-------- ........ -b ..... 

- - b 75.6 -. - - - - -b 74.8 

....... 

- ....... -b ..... -D ... --. --- .)------ 

Figure 3.4.1.1.9: COIL-EC and COIL-E bend temperatures for test 10 (cross-counter 

flow, wet coil, 16.7 "C (30.0 OF) superheat on all circuits) 

56.7 

- e - - - - - - - 

56.5 

.. U - - e . -- -D I I 

I 1 Exil Refrigerant 

Coil-EC Test 13, Bend Temperalures (OF) 

1 ........ * y--- b 56.1 

<-- ; 8 8 153.5 

1 ........ 8 8. ....... * 

1 1 153.0 

g,,;---- - 55.6 

155.1 .... __.. 1 

Coil-E Test 13. Bend Temperatures (OF) 

Exil Refrigerant 

....... 8 

8 

28 
Coil-E Test 14, Bend Temperalures ( OF) Exit Relrigeranl Exit Refrigerant Coil-EC Test 14, Bend Temperalures [ OF) 

- .. - - - .b 74.2 -I - - - - -D 75.9 

....... 

76.1 

-..---- -D ........ -b ........ 

- - --- - - b 74.2 - 75.1 ........ ........ 

49.9 55.3 

------- -D ..... -b ..... I.. ....... ------ - b 74.5 ........ - 75.3 

-------- -b ... 

Figure 3.4.1.1.1 1 : COIL-EC and COIL-E bend temperatures for test 14 (cross-counter 

flow, wet coil, 16.7 "C (30.0 OF) superheat on all circuits) 

3.4.1.2 Effects ofAirftow Rate on Coil Capacity 

Air flow rate through the evaporator plays an important role in the capacity. When air 

flow rate is higher than the optimum quantity, the COP of the system decreases due to an 

increase in the air pressure drop and accompanying fan power consumption. But higher 

airflow enhances the heat transfer rate of the evaporator on the air-side due to the higher 

Reynolds number. Table 3.4.1.2.1 and figure 3.4.1.2.1 show the capacity variation of the 

tested evaporators as a function of air flow rate. As the air flow rate increased, total 

capacity increased due to the increased air mass flow rate. The latent heat transfer rate 

changed a little due to the constant evaporator pressure set by the evaporator pressure 

regulating valve. Other reasons that could play, a small part in the constant latent 

29 
capacity may be the condensed water on the surface of tube and fin mixing with the air 

which is unsaturated before latent heat transfer takes place at the range of these air flow 

rates. Secondly, the temperature difference between the air and the surface of 

condensing water decreases because thermal resistance increases due to the condensed 

water layer. As a result, the dominant increase in total capacity was caused by the 

increase in the sensible heat transfer rate. 

- T 

Test 

Name 

V020226P 

V020207E 

ir020301 P 

V020320E 

u'020607P 

u'0203 19E 

102041 7A 

102041 SA 

io204 1 6B 

Y 

V 

c 

Volumetric Flowrate 5 

of Air m3/h (scfm) - .- 

Superheat to 

1) SI units of m3/kWh multiplied by capacity (Q) in kW to determine airflow, (IP units of cfm/ton 

multiplied by capacity (Q) in tons). 

30 

- 
30000 I 8790 

E 20000 E 

-D- Wavy 

4 Enhanced 

a 

m' 

W 

.- 3 15000 ~ 

-6- Enhanced-Cut 0 

m Latent 

0 e 10000 1 

n - 

5860 s v 

)r 

tu 

Q 

c, 

-4395 '5 

-2930 5 

-1465 

o! I 

598.10 791.02 1002.55 

Air flow rate(CFM) 

Figure 3.4.1.2.1 : Capacity as a function of air flow rate for wet coil conditions 

3.4.1.3 Effects of Non-Uniform Airflow on Coil Capacity 

The combined effects of non-uniform airflow and evaporator superheat were examined 

by blocking the upper portion of the test evaporator on COIL-W and COIL-E during 

cross-counter flow operation for wet coil conditions. Figure 3.4.1.3.1 shows an idealized 

non-uniform velocity profile for a test coil with the upper half of the coil partially 

blocked. The velocity ratio was calculated by taking the average of the 15 velocity points 

on the top half divided by the average of the 15 velocity points on the lower half of the 

test evaporator. Test 9 conditions of 5.6 "C (10.0 OF) superheat were first performed, 

then the blockage was applied with no expansion valve adjustment (test 9A), and finally 

the expansion valves were adjusted to yield 5.6 "C (10.0 OF) superheat on all circuits (test 

9B). During these tests the standard airflow rate was held constant; the airflow was not 

31 
allowed to drop when the blockage was added regardless of the significantly higher 

pressure drop. Table 3.4.1.3.1 shows the performance of the coils with varying degrees 

of airflow blockage. 

Velocity Ratio Test name Test Airflow, 

(see Figure #) type1 3 m3/h(scitn) 

w020522 9 W 1244(732) 1:l 

1:1.5 

(Fig. 3.4.1.3.2) A 

(Fig. 3.4.1.3.3) A 

w020523 9A W 1252(737) 

w020524 9B W 1249(735) A 

B 

1:1.5 

1:l 

1:2 

w020528 9 W 1245(733) 

w020528 9A W 1239(729) 

w020529 9B W 1237(728) 

(Fig. 3.4.1.3.4) C 

A 1:2 

1) Test 9: uniform airtlow with superheat on all circuits of 5.6 "C (10.0 OF), 9A: 

Figure 3.4.1.3.1 : Idealized velocity profile over evaporator with upper half partially 

blocked 

Air-side Capacity 

Capacity, W Ratio, 

(Btu/h) Q/QTest 9 

6598 (22515) 1 

6351 (21670) 0.96 

6535 (22298) 0.99 

6636(22644) 1 

'6179(21085) 0.93 

6307(21521) 0.95 

same expansion valve setting as test 9 

Coil Air 

Pressure 

Drop, Pa 

(in WG) 

37.9 

(0.152) 

49.6 

49.8 

38.1 

(0.153) 

53.3 

(0.214) 

58.8 

(0.23 6) 

(0.199) 

(0.200) 

ut with 

non-uniform airflow and no superheat adjustment, 9B: expansion valves adjusted to yield 5.6 "C (10.0 "F) superheat on all 

circuits with non-uniform airflow. 

32 
Table 3.4.1.3.2: COIL-E Performance with Non-Uniform Airflow 

1) Test 9: uniform airflow with superheat on all circuits of 5.6 OC (10.0 OF), 9A: same expansion valve setling 4, IC\I ‘j hut ulih 

non-uniform airflow and no superheat adjustment, 9B: expansion valves adjusted to yield 5.6 “C (10.0 ‘I ) rupcrhcd Im JII 

circuits with non-uniform airflow. 

Figure 3.4.1.3.2 shows the air velocity contour map for COIL-W with no obstructions 

present. The volumetric flowrate for this test was 1244 m3h (732 scfm) with an avcrage 

velocity of 6437 m/h (352 fpm) and standard deviation of 512 m/h (28 fpm). Any nonuniformity in the unobstructed evaporator’s entrance region airflow was due to the 

dewpoint sampling tree, thermocouple grid, and fin angles. Figure 3.4.1.3.3 shows the 

33 
non-uniform velocity contour map for COIL-W when the flow was obstructed to the 

upper half of the coil. An average of the top half air velocity was compared to the bottom 

half average air velocity to yield the velocity ratio of 1 to 1.5. The volumetric flowrate 

for this test was 1252m3/h (737scfm) with a 4097mh (224fpm) and 6163m/h 

(337 fpm) average velocity on the upper and lower halves of the coil, respectively. 

Further obstruction was added to produce the velocity contours seen in Figure 3.4.1.3.4 at 

a velocity ratio of 1 to 2. The average velocities in this case were 4005 m/h (219 fpm) 

and 821 1 m/h (449 fpm) over the upper and lower halves of the coil, respectively. 

The imposed airflow blockage in the case of the 1 to 1.5 velocity ratio would have 

increase fan power by more then 30 YO. For the 1 to 2 velocity ratio case, the fan power 

would have increased by at least 54 YO relative to the uniform airflow case. 
16 

15 

14 

13 

12 

11 

0 

9 

a 

7 

6 

5 

4 

3 

2 

+ 

X Location 

Figure 3.4.1.3.2: Uniform airflow velocity (fvmin) contour map for COIL-W 

35 
Figure 3.4.1.3.3: Non-uniform velocity (1 : 1.5) (Wmin) contour map for C'OI L-W 

36 
I I 

5 14 15 16 

X Location 

Figure 3.4.1.3.4: Non-uniform velocity (1 :2) (Wmin) contour map for COIL-W 

Figure 3.4.1.3.5 shows the non-uniform air velocity evaporator capacity relative to the 

uniform air velocity capacity. The first case shows the effects of an imposed obstruction 

with a fixed area expansion device. The expansion device would not be able to correct 

superheat and capacity would drop by 4 YO in the 1 : 1.5 velocity ratio case and by 6.5 Yo in 

the 1:2 velocity ratio case. If the expansion device were able to correct each circuit, then 

the performance would improve closer to pre-obstruction, but for both cases the capacity 

37 
still decreased by 1.2 % and 4.5 %, respectively. The COIL-W non-uniform velocity 

results show that a more non-uniform velocity ratio produced a higher drop in capacity; 

and superheat correction still could not alleviate the entire penalty. 

1 .oo 

Velocity Ratio 

0.99 

0.98 

g 

€ 

g 

E 

Z 0.97 

5 0.96 

Oi 

I! 0.95 

s 

$ -0.94 

b 

B 0.93 n 

- 

0 

2 

0 

0.92 

0.91 

0.90 

Test 9A Test 9B Test 9A Test 9B 

Figure 3.4.1.3.5: Relative capacities for COIL-W non-uniform airflow (refer to Table 

3.4.1.3.1 for a description of the actual tests) 

Figure 3.4.1.3.6 shows the air velocity contour map for COIL-E with no obstructions 

present. The volumetric flowrate for this test was 1276 m3/h (75 1 scfm) with an average 

velocity of 6108 m/h (334 fpm) and standard deviation of 512 m/h (28 fpm). Any nonuniformity in the unobstructed evaporator’s entrance region airflow was due to the 

dewpoint sampling tree, thermocouple grid, and fin angles. Figure 3.4.1.3.7 shows the 

non-uniform velocity contour map for COIL-E when the flow was obstructed to the upper 

half of the coil. An average of the top half air velocity was compared to the bottom half 
average air velocity to yield the velocity ratio of 1 to 1.26. The volumetric flowrate for 

this test was 1281 m'/h (754 scfm) with a 5340 m/h (292 fpm) and 6748 m/h (369 fpm) 

average velocity on the upper and lower halves of the coil, respectively. Further 

obstruction was added to produce the velocity contours seen in Figure 3.4.1.3.8 at a 

velocity ratio of 1 to 1.36. The average velocities in this case were 5340 m/h (292 fpm) 

and 7279 m/h (398 fpm) over the upper and lower halves of the coil, respectively. More 

obstruction was added to produce a 1 to 1.62 average velocity ratio seen in Figure 

3.4.1.3.9. Average velocities in this case were 4609mh (252 fpm) and 7498 m/h 

(410 fpm) over the upper and lower halves of the coil, respectively. Again, obstruction 

was added to produce a 1 to 1.75 average velocity ratio seen in Figure 3.4.1.3.10. 

Average velocities in this case were 3621 m/h (198 fpm) and 6346 m/h (347 $m) over 

the upper and lower halves of the coil, respectively. The final obstruction was then added 

to produce a velocity ratio of 1 to 2.59. Average velocities in this case were 3219 m/h 

(176 fpm) and 8358 m/h (457 fpm) over the upper and lower halves of the coil, 

respectively. 

39 
3 i 5 6 9 10 11 12 13 14 15 16 

X Location 

Figure 3.4.1.3.6: Uniform velocity (Wmin) contour map for COIL-E 

40 
16 

15 

14 

13 

12 

11 

IC c 

0 

g gs 

J * 

E 

7 

E 

5 

4 

2 

c 

i 

X Location 

Figure 3.4.1.3.7: Non-uniform velocity (1 : 1.26) (ft/min) contour map for COIL-E 

41 
16 

15 

14 

13 

12 

11 

10 - 

2 

6 59 

-J - 

8 

7 

6 

5 

4 

3 

2 

34 5 

X Location 

Figure 3.4.1.3.8: Non-uniform velocity (1 :1.36) (Wmin) contour map for COIL-E 

42 
I 

- 

16 

15 

14 

13 

12 

11 

10 

- 

5 

j9 

-I - 

8 

7 

6 

5 

4 

3 

2 

X Location 

Figure 3.4.1.3.9: Non-uniform velocity (1 : 1.62) (Wmin) contour map for COIL-E 

43 
X Loca 

Figure 3.4.1.3.10: Non-uniform velocity (I : 1 .75) (Wmin) contour map for COIL-E 

44 
345 6 i 8 9 IO 11 12 

I 

I 

1 

I 

I 

I 14 l! 16 

X Location 

Figure 3.4.1.3.1 1: Non-uniform velocity (1:2.59) (Wmin) contour map for COIL-E 

Figure 3.4.1.3.12 shows the non-uniform air velocity evaporator capacity rc1atii.c to the 

uniform air velocity capacity for all COIL-E tests represented above and shmvn in Table 

3.4.1.3.2. The cases labeled “Test 9A” show the effects of an imposed obstruction with a 

fixed area expansion device. Cases labeled “Test 9B” have the superheat adjusted back 

to 5.6 “C (10.0 OF) on each circuit. The first case had a velocity ratio of 1 to 1.26 and a 

capacity specific airflow of 183 m3/kWh (378 scfmlton). Less than 1 % in capacity was 

45 
lost due to the obstruction, and this lost capacity was recovered when superheat was 

corrected (capacity was corrected within the uncertainty of our measurement at 

approximately 3.5 %). At a higher absolute airflow of 186 m3/kWh (385 scfdton), the 

losses in capacity with the obstruction were greater. 

The losses in capacity with no superheat correction ranged from slightly more than 2 % at 

the 1 to 1.36 velocity ratio to approximately 5.5 % at the 1 to 2.59 velocity ratio. When 

the superheat was corrected, much of the loss in capacity was recovered. Please note that 

all of these tests were performed at a constant airflow rate. As shown in Table 3.4.1.3.1 

and 3.4.1.3.2, the air pressure drop across the evaporator increased substantially when the 

blockage was imposed. This would translate into much higher fan power requirements 

and a subsequently lower COP. In the case of the 1 to 1.36 velocity profile, fan power 

would have increased by 21 % with the imposed blockage (power equals flowrate times 

the pressure drop). For the 1 : 1.62 velocity profile, fan power would have increased by 19 

%. For the 1:1.75 velocity ratio, fan power would have increased by 24 %. For the 

highest blockage test with a velocity ratio of 1 :2.59, the fan power would have increased 

by at least 20 YO. 
186 rn3/kWh (385 scfrn/ton) 183 rn3/kWh (378 scfrn/ton) 

1.01 - 

Velocity 

Ratio 

1:1.26 

1 .oo 

- B 0.99 

‘t: 

a E 0.98 

e 

5 0.97 

5 0.96 

a 

P 0.94 

0 

al 

0 

- 

- 

3 0.95 - 

0 m 

0 

0.93 

0.92 

0.91 

Test 9A Test 98 Test 9A Test 9B Test 9A Test 9B Test 9A Test 9B Test 9A Test 9B 

Figure 3.4.1.3.12: Capacity relative to uniform airflow for COIL-E (refer to Table 

3.4.1.3.2 for a description of the tests) 

47 
3.4.2 

Table 3.4.2.1 shows the performance of COIL-W and COIL-E with refrigerant circuited 

in cross-counter flow and cross-parallel flow with a capacity specific airflow rate of 

193 m3/h (400 scfdton). COIL-W airflow rates for parallel flow and counter flow were 

912 m3/h (537 sch) and 1240 m3/h (730 scfm), respectively. COIL-E airflow rates for 

parallel flow and counter flow were 895 m3/h (527 scfm) and 1288 m'/h (758 scfm), 

respectively. The main information to be gained from this comparison is the rapid 

reduction in capacity that follows any increase in superheat during parallcl flow (Figure 

3.4.2.1). This was evident for both coils tested under parallel flow conditions. 

Cross-Parallel AidRehnerant Flow Configuration Tests 

As the superheat increased from 5.6 "C (1 0.0 OF) to 16.7 "C (30.0 OF), the rcfrigcrant and 

air temperatures became pinched very quickly. During all tests the evaporator pressure 

was fixed to give an evaporating saturation temperature of 7.2 "C (45.0 OF). When 

superheat was increased to 16.7 "C (30.0 OF), the exiting refrigerant temperature 

approached 23.9 "C (75.0 OF); pinching of the two streams occurred rapidly. This is 

evident from examining the capacity of test 10 in parallel flow. 

Figure 3.4.2.1 shows that COIL-W capacity in parallel flow decreased by 84.8 90 as the 

superheat increased from 5.6 "C (10.0 OF) to 16.7 "C (30.0 OF); test 9 to test IO. Tests 11 

and 12 also produced very low capacity due to two of the three circuits having a 6.7 "C 

(30.0 OF) superheat. Although tests with the center circuit flooding produced higher 

capacity than the top circuit flooding for both COIL-W and COIL-E. 
Table 

W020207B 

W 020530 A 

W02053 1 A 

W020215B 

E020607 A 

W0203 18A 

W0203 18B 

4.2.1 : Counter and Parallel 

Volumetric Flowrate of 

5 Air m’/h (cfm) I .- * 

1 93.Qta 242-QJa 

(400.Q) (500.Q) 

;low Performance of COIL-W and COIL-E 

Cross-Counter AidRefrigerant Flow 

W0203 19A 

Cross-Parallel AirIRefrigerant Flow 

1 b : capacity at test 9 for the coil specified. 
1.10 

1 .oo 

0.90 

0.80 

I Cross-Counter FIOW /--GiGl 

0 

2 0.60 

0.40 

0.30 

0.20 

0.10 

0.00 

9 9 10 11 12 

Test Number 

Figure 3.4.2.1 : Cross-parallel flow capacity comparison for COIL-W and COIL-E 

relative to their performance at test 9 with cross-counter flow 

COIL-E capacity in parallel flow was 3.9 % less than COIL-W at the test 9 condition of 

193 m3/kWh (400 scfdton). This is the opposite of the counter flow capacity results 

where COIL-E had a 5.4 YO greater capacity than COIL-W. COIL-E produced a slightly 

lower capacity than COIL-W at the test 9 parallel flow conditions. This was mainly due 

to the 196 m3/kWh (406 scfin/ton) for COIL-W and the 193 m3/kWh (399 scfdton) for 

COIL-E. The ideal airflow would have produced 193 m3/kWh (400 scfdton) for both 

coils, but COIL-W airflow was high while COIL-E airflow was slightly low. These 

differences in airflow produced the accompanying difference in capacity. Also, a 

secondary factor in the lower capacity of COIL-E than COIL-W may be due to the more 

rapid pinching of COIL-E than COIL-W due to the higher air-side heat transfer of the 

enhanced fins. 
COIL-E capacity in parallel flow decreased by 77.6 % as the superheat increased from 

5.6 "C (1 0.0 OF) to 16.7 "C (30.0 OF); test 9 to test 10. Again, rapid pinching of the two 

fluid streams reduced capacity. 

51 
4 EVAPORATOR MODELING AND SIMULATIONS 

4.1 

Modeling of finned tube heat exchangers started at NIST with a tube-by-tube simulation 

model originally formulated by Chi (1979). Over the years, the model underwent 

significant upgrades, which are documented in Domanski and Didion (1 983), Domanski 

(1 99 l), Lee and Domanski (1 997), and Domanski (1 999b). These upgrades included the 

capability to account for air maldistribution and its interaction with refrigerant 

distribution, the extension to zeotropic mixtures, the extension to new refrigerant 

property representations, and implementations of new simulation correlations. The 1 999 

upgrade equipped the evaporator model with a graphical user interface (GUI). The GUI 

serves as a pre- and post-processor; it facilitates preparation of simulation input data, 

including the layout of refrigerant circuitry, and allows the user to display detailed 

performance results for individual tubes after a simulation run is completed. In 2002 

under a parallel ARTI-21CR/605-500100 project (Domanski and Payne, 2002), the 

EVAP-COND simulation package was developed that included the evaporator and 

condenser models, EVAP5 and COND5, working under the same GUI. Both models 

were validated with experimental data taken with R22 and R410A heat exchangers at 

NIST. Since the current project and ARTI-21CW605-500100 project partially 

overlapped in time, the EVAP-COND attached to the ARTI-2 1 CW605-500100 report 

included some of upgrades developed under this study, e.g. the option to simulate the 

evaporator together with a refrigerant distributor. Not included in the release version of 

EVAP-COND are the upgrades to the evaporator model that account for longitudinal fin 

Background on Evaporator Model EVAPS 
conduction, which were needed to perform simulations with controlled refrigerant 

superheats at individual refrigerant circuits. 

4.2 Description of EVAPS 

4.2.1 Modeling Approach 

Figure 4.2.1.1 presents the refrigerant circuitry and air velocity representation used by 

EVAPS. The tube-by-tube modeling approach recognizes each tube as a separate entity 

for which the model performs simulation calculations. These calculations are based on 

inlet refrigerant and air parameters, their properties, and mass flow rates. The simulation 

begins with the inlet refrigerant tubes and proceeds to successive tubes along the 

refrigerant path in the heat exchanger. At the outset of the simulation, the air temperature 

is only known for the tubes in the first row and has to be estimated for the remaining 

tubes. A successful simulation run requires several passes (iterations) through the 

refrigerant circuitry, each time updating inlet air and refrigerant parameters for each tube. 
\\ I1 

Refrigerant inlet tubes 

Figure 4.2.1.1 : Representation of air distribution and refrigerant circuitry in EVA 

L 

Coil height 

.P5 

Heat transfer calculations start by calculating the heat-transfer effectiveness, E, by one of 

the applicable relations (Kays and London, 1984). With the air temperature changing due 

to heat transfer, the selection of the appropriate relation for E depends on whether the 

refrigerant undergoes a temperature change during heat transfer. Once E is determined, 

heat transfer from air to refrigerant is obtained using equation 4.2.1.1. 

(4.2.1.1) 

The overall heat-transfer coefficient, U, is calculated by equation 4.2.1.2, which sums up 

the individual heat-transfer resistances between the refrigerant and the air. 
U= 1 AoXp +-1 + + A+ 

A 

-1 

(4.2.1.2) 

The first term of equation 4.2.1.2 represents the refrigerant-side convective resistance. 

The second term is the conduction heat-transfer resistance through the tube wall, and the 

third term accounts for the conduction resistance through the water layer on the fin and 

tube. The fourth term represents the contact resistance between the outside tube surface 

and the fin collar. The fifth term is the convective resistance on the air-side where the 

multiplier (l+a) in the denominator accounts for the latent heat transfer on the outside 

surface. For a dry tube a = 0.0 and l/h, = 0.0. Once the heat transfer rate from the air to 

the refrigerant is calculated, the tube wall and fin surface temperatures can be calculated 

directly using heat-transfer resistances. Then, the humidity ratios for the saturated air at 

the wall and fin temperatures are calculated, and mass transfer from the moist air to the 

tube and fin surfaces is determined. For more detailed information on heat transfer 

calculations refer to Domanski (1 991). 

4.2.2 

EVAPS uses the following correlations for calculating heat transfer and pressure drop. 

Heat Transfer and Pressure Drop Correlations 

Air Side 

heat-transfer coefficient for flat fins: Wang et al. (2000) 

heat-transfer coefficient for wavy fins: Wang et al. (1 999a) 

heat-transfer coefficient for slit fins: Wang et al. (2001) 

55 
heat-transfer coefficient for louver fins: Wang et al. ( 1999b) 

fin efficiency: Schmidt method, described in McQuiston et al. (1 982) 

A correlation for calculating the tube-collar junction resistance is not listed because all 

air-side heat transfer correlations authored by Wang include the heat transfer resistance 

between the tube and collar. 

Refrigerant Side 

single-phase heat-transfer coefficient, smooth tube: McAdams, described in 

ASHRAE (2001) 

evaporation heat-transfer coefficient up to 80% quality, smooth tube: Jung and 

Didion (1 989) 

evaporation heat-transfer coefficient up to 80% quality, rifled tube: Jung and 

Didion (1 989) correlation with a 1.9 enhancement multiplier suggested by 

Schlager et al. (1 989) 

mist flow, smooth and rifled tubes: linear interpolation between heat transfer 

coefficient values for 80 YO and 100 YO quality 

single-phase pressure drop, smooth tube: Petukhov ( 1970) 

two-phase pressure drop, smooth tube, lubricant-free refrigerant: Pierre ( 1 964) 

two-phase pressure drop, rifled tube: Pierre (1 964) correlation for smooth tube 

with a 1.4 multiplier suggested by Schlager et al. (1 989) 

single-phase pressure drop, return bend, smooth tube: White, described in 

Schlichting (1 968) 

two-phase pressure drop, return bend, smooth tube: Chisholm, described in 

Bergles et al. (1 98 1). The length of a return bend depends on the relative 

locations of the tubes connected by the bend. This length is accounted for in 

pressure drop calculations. 

56 
4.2.3 Representation of Refrigerant Properties 

Representation of thermodynamic and transport properties is based on REFPROP6 

property routines (McLinden et al., 1998). Because EVAP5 simulations are 

computationally intensive, using a refrigerant property look-up tables is a practical 

necessity if simulation runs are expected to take less than 60 seconds. This is particularly 

true in case of REFPROP6 for which property calculations are several times more CPU 

demanding than for REFPROPS. EVAP-COND employs a pressure-enthalpy-based 

system of look-up tables, which includes eight different routines that retrieve the desired 

state or transport property. The look-up scheme is applicable to single component 

refrigerants and refrigerant mixtures. If a given refrigerant state falls outside the range of 

the look-up table, then EVAP-COND calls a REFPROP6 refrigerant property routine 

directly. 

Since REFPROP6 property calculations may not converge occasionally , particularly 

during phase equilibrium calculations for refiigerant mixtures, EVAP-COND employs an 

error evasive scheme. Under this scheme, EVAP-COND attempts to obtain a given 

property even if REFPROP flash calculations do not converge, e.g., if a routine PHFLSH 

crashes, a routine that uses TPRHO is invoked to attempt to iteratively match TPRHO’s 

enthalpy value with the known (target) value. If both REFPROP flash calculations do not 

converge, then the data point in the refrigerant look-up table is flagged and look-up table 

routines iterate properties for this point using refrigerant properties in the neighboring 

nodes of the table. 

57 
4.3 Modeling Issues 

The following four sections present four major modeling issues that received special 

consideration during this study. 

4.3.1 Refrigerant Distribution 

Simulating refrigerant distribution is an important aspect of heat exchanger simulation 

because of its impact on the heat exchanger performance. It is also know that in some 

designs a non-uniform air distribution may affect refrigerant distribution. In a heat 

exchanger with multiple circuits, refrigerant distributes itself in appropriate proportions 

so that the refrigerant pressure drop from the inlet to the outlet for all circuits is the same. 

In the context of simulating refrigerant distribution, a refrigerant circuit starts at the point 

of the first split of refrigerant stream after leaving the condenser and ends at the final 

merging point before entering the suction line leading the refrigerant to the compressor. 

If the refhgerant enters the evaporator by a single tube, the first split, if any, will exist 

within the coil assembly. If the evaporator has several inlet tubes and a refrigerant 

distributor is used, the first refrigerant split typically occurs at the inlet to the distributor 

tubes just after the expansion process in a thermostatic expansion valve (TXV) or a short 

tube restrictor. Note, that in this design, refrigerant pressures and temperatures at 

different inlet tubes may be different, as graphically shown in Figure 4.3.1.1. Such 

different refrigerant pressure and temperature profiles also occurred during the tests with 

controlled uneven exit superheats (refrigerant distributions), namely tests 3, 4, 7, 8, 11, 

and 12. 

58 
I Evapmtor 

fl 

outlet 

Enthalpy 

Figure 4.3.1.1 : Possible refrigerant pressure profiles in a three-circuit evaporator fcd by a 

refrigerant distributor 

Under this project, two simulation methods were developed to simulatc ciaporator 

performance with controlled refhgerant superheats at the evaporator outlct tubes. The 

first scheme involving a general model for a refrigerant distributor was introduccd into 

EVAP-COND as one of the eight evaporator simulation options. When thc ciqwrator is 

simulated using this option, the refrigerant operating input data are the condcnscr cxit 

bubble-point temperature, condenser subcooling, evaporator exit dew-point tcmpcrature, 

and evaporator superheat (in the release version of EVAP-COND, condenser subcooling 

and evaporator superheat are imposed as 5.0 "C (9.0 OF)). For this option, as thc first task 

EVAP-COND runs preliminary simulations to establish dimensions of the rcfrigerant 

distributor tubes that would inflict a 70 kPa (10.2 psi) refrigerant pressure drop. Once the 

distributor tubes are sized, EVAP-COND proceeds to main simulations in which it 

59 
establishes refrigerant distribution between different circuits based on the total pressure 

drop. This total pressure drop includes the pressure drop in a given distributor tube and 

the refrigerant circuit in the coil assembly it feeds. In the test version used in this study, 

EVAP-COND additionally solicits a “restriction factor” for each distributor tube, which 

acts as a multiplier to the pressure drop calculated by the program. By inputting values 

different from 1 .O, the user can control refrigerant distribution and refhgerant superheat 

at different evaporator exit tubes. The program iterates the refrigerant mass flow rate 

until the overall superheat is reached at the evaporator exit. Figures 4.3.1.2 and 4.3.1.3 

present the eight refiigerant input data options and the input data window for EVAPCOND simulations involving a refrigerant distributor, respectively. 

Figure 4.3.1.2: EVAP-COND refrigerant input data options for evaporator simulations 

Figure 4.3.1.3: EVAP-COND input data window for simulations involving a refiigerant 

distributor 

60 
While the simulation option involving a refrigerant distributor is useful for a coil designer 

for typical simulations, this option proved to be somewhat impractical for controlling 

individual exit superheats when we tried to reproduce our test results. This was due to a 

non-linear dependence of individual exit superheats on the “restriction factors”. For this 

reasonanother simulation scheme was developed in which the user directly assigns 

refbgerant distribution between different circuits. The operating conditions are as shown 

in Figure 4.3.1.4 with the refrigerant inlet quality and distribution (in fractions) solicited 

by the follow-up DOS window. While holding the refrigerant distribution constant, the 

program iterates the overall refrigerant mass flow rate and inlet pressures at individual 

inlet tubes to converge on the target exit pressure (the same for each exit tube) and 

overall target superheat. Different individual superheats can be obtained by specifying 

different refrigerant distributions. Eventually, all simulation results for this study were 

obtained using the second scheme. 

Figure 4.3.1.4: EVAP-COND input data window for simulations with specified overall 

evaporator exit saturation temperature and superheat. 
4.3.2. Air-side Heat Transfer Correlations 

Often the most significant part of heat transfer resistance between the air and refrigerant 

is on the air-side of the heat exchanger. For this reason, a literature review on the latest 

air-side heat transfer correlations was performed at the outset of this study. Of our 

particular interest were correlations for wavy and lanced (slit) fins - the two fin types 

used in the evaporators tested under this project. 

Figure 4.3.2.1 compares the predictions of different correlations available in the 

literature. These predictions were calculated for typical fin designs for a three-depth-row 

heat exchanger. The layout of different prediction lines in the figure may serve as an 

explanation why predicting performance of a finned-tube heat exchanger may be 

difficult. For wavy fins, the correlation by Wang et al. (1 999a) and Kim et al. (1 997) are 

in a close agreement, while the correlation by Webb (1 990) calculates heat transfer 

coefficients up to 50 YO lower that the two first methods. In the air velocity range of 1.8 

m/s (5.9 Ws) to 2.1 m/s (6.9 Ws), the Webb correlation breaks sharply due to switching 

between two different algorithms with a changing air-side Reynolds number. At some 

point the Webb correlations provide a value for the wavy fin heat transfer coefficient that 

is lower than that for a flat fin, which is not a realistic prediction. 

For slit (lanced) fins, the correlations by Nakamura and Xu (1983) and Wang et al. 

(2001) may differ by more than 40 %, depending on air velocity. This spread may be 

indicative of the general fact that some correlations do not predict well outside the 

geometries for which they were developed. A measurement uncertainty in one or both 
experiments may also be a contributing factor to this large discrepancy. In addition, it 

should be noted that the Nakayama and Xu (1 983) predictions do not approach zero at air 

velocities below 2 m/s (6.5 ft/s), the trend exhibited by the other correlations. Regarding 

louver fins, the correlation by Wang et al. (1999b) shows a step change in the 1.5 m/s 

(4.9 Ws) to 1.8 m/s (5.9 Ws) range caused by using two different algorithms, similar to 

the Webb correlation for the wavy fins. 

The analysis of relative predictions of the air-side heat transfer coefficient provided the 

reason for replacing the existing correlations in EVAPS with correlations published by 

Wang and his co-workers for all types of fins, i.e., flat, wavy, louver, and slit fins. It was 

judged that a better degree of prediction consistency can be obtained with all correlations 

developed by the same author. Still, the reader should note a reservation regarding the 

louver fin correlation, which did not provide smooth predictions in Figure 4.3.2.1. 

In conclusion, we have to recognize the spread in performance between different 

enhanced fins, either realistic or perhaps, in some instances, overstated by correlations. 

To accommodate these differences and facilitate accurate evaporator model predictions, 

EVAP-COND provides an option that allows the user to “tune” evaporator simulated 

performance to the laboratory data by specifying a “correcting parameter” for the air-side 

heat transfer coefficient (such correcting parameters are also allowed for the refrigerantside heat transfer coefficient, and refhgerant pressure drop). 

63 
200 

-m- Flat Fin (Gray 8 Webb, 1986) 

-c- Flat Fin (Wang et al., 2000) 

- --b Wavy Fin (Wang et al., 1999a) Y E 160 - -0- Wavy Fin (Kim et al.. 1997) 

-f- Lanced Fin (Nakayama 8 Xu. 1983) - +.Slit Fin (Wang et al., 2001) - 140 - -A- Louver Fin (Wang et al., 1999b) 

a 

8o - -0- Wavy Fin (Webb, 1990) n 

cv 

'J 120- 

c 

0 I I I I I I I I I 

Velocity (mls) 

Figure 4.3.2.1 : Comparison of air-side heat transfer correlations 

64 
4.3.3. Internal Heat Transfer in a Finned-Tube Evaporator 

The current study stipulated evaporator tests with a superheat as high as 16.7 "C 

(30.0 OF) to assess capacity degradation at large and uneven superheats. As presented in 

previous sections, these tests for COIL-E and COIL-EC produced rather interesting data 

suggesting that internal heat transfer within the evaporator metal body may be the culprit 

for significant capacity degradation. Figure 4.3.3.1 presents measured capacities at 

193 m3/kWh (400 cfidton) and includes capacities predicted by EVAPS (internal heat 

transfer not considered). The figure shows that tested capacities at 5.6 "C (10 OF) overlap 

for both coils and are reasonably well predicted by EVAP5. However, at 16.7 "C (30 OF) 

superheat COIL-E capacity was tested to be 346 W (1 180 Btuh) less that COIL-EC 

capacity and 966 W (3295 Btuh) less than the simulated value. The lower capacity 

degradation for COIL-EC can be explained by the fin cuts, which physically separated 

different tube depth rows and disallowed heat transfer between neighboring rows. 

65 
24000 

26000 c 

18000 

--t Coil-E, simulation 

--m- Coil-E. test 

I' I 

0 10 20 30 40 

Superheat (OF) 

Figure 4.3.3.1: Tested and predicted capacities for COIL-E and COIL-EC at 5.6 "C (10 

OF) and 16.7 "C (30 OF) superheats (Test 9 and Test 10 opcrating 

conditions. Internal heat transfer within the evaporator metal body not 

considered.) 

The following two sections discuss the two modes of the internal hcat transfer, 

longitudinal tube conduction and longitudinal fin conduction (tube-to-tubc hcat transfer), 

and their impact on evaporator performance. 

4.3.3. I Longitudinal Tube Conduction 

The general theory states that if a temperature gradient exists in a wall of a heat 

exchanger, then conduction heat transfer will occur along that wall and may thcrcfore 

degrade the performance of the heat exchanger. Kays and London (1 984) idcntificd the 

major parameters affecting the magnitude of the performance degradation due to this 

phenomenon as follows: 
(4.3.3.1.1) 

The magnitude of the performance degradation becomes larger with increasing h, 

C,,,i,,/Cmax, and NTU. Kays and London stated that this reduction in performance is seen 

in heat exchangers designed for high effectiveness (~>0.9), however, they did not provide 

much of a quantitative analysis. Ranganayakulu et. al. ( 1996) carried out a series of finite 

element simulations to quantify the magnitude of the performance degradation in a heat 

exchanger due to longitudinal heat conduction.. The rcsults of their simulations are 

represented by the “conduction effect factor”, 2, in terms of the effectiveness with no 

longitudinal conduction effects, E,, , and the effectiveness considering longitudinal 

conduction effects, E,~. 

EN, - Ewc z= 

‘NC 

(4.3.3.1.2) 

The conduction effect factor can be read from the charts presented in the paper for given 

&,A, Cmin/Cmax, and NTU. Ranganayakulu et. al. suggested 0.8 as the effectiveness limit 

below which the impact of longitudinal conduction is negligible. 

With the theory and the results from the numerical simulations at hand, the impact of 

longitudinal tube conduction for a typical finned-tube evaporator was examined. Using 

EVAPS, a 10.6 kW (3 ton) evaporator was simulated to identify the tubes with two-phase 

R-22 (in which the longitudinal heat conduction does not occur) and the tubes with a 

67 
superheated refrigerant (in which longitudinal heat conduction does take place). Then the 

capacity penalty for the superheated tubes was calculated as a fraction of capacity of 

these tubes and as a fraction of the evaporator capacity. 

Figure 4.3.3.1.1 displays an evaporator side view with a schematic of the refhgerant 

circuitry. Figure 4..3.3.1.2 contains a Coil Design Data window from EVAP-COND with 

the coil design information, and Figure 4.3.3.1.3 presents the operating conditions of the 

evaporator. 

Figure 4.3.3.1.1 : Refrigerant circuitry configuration for the analyzed R-22 evaporator 

(inlet tube: tube # 24, outlet tubes: tube # 1 and tube # 16; ) 
Figure 4.3.3.1.2: EVAP-COND window with evaporator design information 

Figure 4.3.3.1.3: EVAP-COND window with evaporator operating conditions 

For the evaporator exit superheat of 8.0 "C (14.4 OF), the simulations showed that only 5 

of the 48 tubes in the heat exchanger have superheated refrigerant and experience a 

temperature change. This means that 43 of the tube passes in the heat exchanger will not 

experience any axial heat conduction because there will be no temperature difference for 

this to occur (neglecting the marginal drop of saturation temperature due to the pressure 

69 
drop). An example of a tube with superheated vapor, tube number 15, has the following 

values for the aforementioned parameters. 

= 0.00 1 1 

-=0.441 Cmi" 

cmax 

NTU=0.368 

(4.3.3.1.3) 

(4.3.3.1.4) 

(4.3.3.1.5) 

These parameters lie below the range of data given by Ranganayakulu. Using 

extrapolation, it was determined that the conduction effect factor would be approximately 

0.0005. This means that this particular tube in the heat exchanger would see a loss in 

capacity of one twentieth of one percent due to axial heat conduction. When this capacity 

degradation is summed over all of the tubes in the entire heat exchanger where this effect 

occurs, the capacity reduction totals 0.13 W (0.45 Btuh), which is insignificant when 

compared to the predicted performance of the evaporator being 49000 W (1 6800 Btu/h). 

It should be noted that the effectiveness of tube 15 is 0.29. Hence, our result agrees with 

the general statements by Kays and London and that of Ranganayakulu et. al. that the 

longitudinal heat transfer has an insignificant effect for heat exchangers with an 

effectiveness less than 0.8. The negligible impact of the longitudinal tube conduction on 

evaporator performance permits neglecting this heat transfer in modeling of a finned-tube 

heat exchanger. It may be further inferred that the same conclusion can be made for the 

R407C zeotropic mixture. Although a 7 "C (12 OF) glide associated with R407C phase 

change produces a temperature difference promoting longitudinal heat transfer, in the 

70 
analyzed evaporator this glide would be distributed over 10 m (33 ft) of tube passes and 

would result in a small longitudinal temperature gradient. 

4.3.3.2 Tube-to-Tube Heat Transfer via Fins 

If we recognize that longitudinal tube heat conduction has a negligible impact, then the 

difference in capacity degradation between COIL-E and COIL-EC at 16.7 "C (30 OF) 

superheat must be due to longitudinal fin conduction. In COIL-EC, the continuous cuts in 

the fins separating different depth rows prevent heat transfer between di ffcrent depth 

rows. However, the fins join the adjacent tubes in the same depth row. and some heat 

transfer between them occurs. This is why COIL-EC still experiences a declinc in 

capacity, but not as much as COIL-E. 

Our literature review located two publications that shed some light on the longitudinal fin 

conduction phenomenon. Heun and Crawford (1 994) performed analytical study of the 

effects of longitudinal fin conduction on multipass cross-counterflow single-depth-row 

heat exchanger. They considered the fins to have one-dimensional tcmpcrature 

distributions and solve them using a system of non-dimensional differentiah equations. 

Their results showed that longitudinal fin conduction always degrades heat exchanger 

performance and this effect is stronger for a low normalized fin resistancc and large 

values of the ratio of air-side conductance to air heat capacity rate. 

Romero-Mender et al. (1997) also studied tube-to-tube heat transfer in a single-row 

finned-tube heat exchanger. They assumed the fins to be continuously and uniformly 
distributed along the length of each tube. With his continuum assumption, they solved a 

system of ordinary differential equations for steady-state refrigerant and tube-wall 

temperatures. They identified four non-dimensional groups that effected the degradation 

of evaporator capacity. These group are: (1) the ratio of the thermal conductance for 

convective heat transfer between the refrigerant and the wall to the product of 

refrigerants heat capacity and mass flow rate, (2) the ratio of the thermal conductance for 

external heat transfer from the unfinned portion of the tube to the internal thermal 

conductance, (3) the ratio of the thermal conductance for convection from the fin to the 

thermal conductance for conduction along it, and (4) the ratio of the thermal conductance 

for heat conduction along the insulated fin to the thermal conductance between the 

refrigerant and the wall. Their study also indicated the number of tubes to be an 

influencing factor. The study by Romero-Mender et al. indicates that tube-to-tube heat 

transfer always degrades capacity and that the influencing parameters they identified 

have a non-linear impact on capacity degradation over the wide range of values studied. 

For some parametric values they found the degradation in a single-row heat exchanger to 

be as high as 20 %. 

Our literature review have not located a publication that would discuss longitudinal fin 

conduction in a multi-depth-row evaporator, but it may be safely expected that capacity 

degradation would be higher than that for a single-depth-row coil. The literature review 

identified a very recent paper which presents three simulation models for finned-tube 

single-phase dehumidifying heat exchangers, the most advanced of which accounts for 

tube-to-tube heat transfer (Oliet et al. (2002)). This model, referred to as 

72 
“advancedCHESS”, is based on a finite volume approach with discretization of the heat 

exchanger domain into a set of control volumes as fin-and-tube elements where both 

local thermophysical properties and empirical coefficients are computed. While 

“advancedCHESS” appears to be a research model, two other models are more practical 

for production simulations. The two less advanced models, called “basicCHESS” and 

“quickCHESS”, do not consider tube-to-tube heat transfer. 

While the above publications are very interesting and valuable, they do not offer a 

methodology for accounting for tube-to-tube heat transfer in a tube-by-tube simulation 

model. The number of influencing parameters identified for a dry fin by RomeroMender et al. (1997) suggests that a fully fundamental approach will be difficult to 

implement into a tube-by-tube evaporator model, which uses the adiabatic fin tip 

assumption and considers an individual tube as a separate entity for heat transfer 

calculations. Our attempts to apply a few algorithms derived from their paper, however, 

were unsuccesful because we were unable to resolve the “clashes” between the 

fundamental algorithms and the current simulation scheme used in EVAPS. It appears 

that merging a fundamental scheme into EVAPS amounts to a separate project that 

should be dedicated to this task. 

To reach the objectives of the project within a stipulated effort and time, a practical 

scheme was developed, which uses the temperature difference between neighboring tubes 

as the driving force for heat transfer. This scheme approaches the tube-to-tube heat 
transfer problem in a similar way Sheffield (1988) studied fin collar-tube heat transfer 

resistance as shown in Figure 4.3.3.2.1. 

-T 

Figure 4.3.3.2.1: Schematic graph for longitudinal fin conduction between two adjacent 

tubes 

To determine the heat transferred, the Fourier Law of Conduction is applied. The effects 

of the available width and configuration of the conducting material (fin) are represented 

by a “shape factor” S used in equation 4.9: 

(4.3.3.2.1) 

To show the impact of fin conduction, Lee and Domanski used this scheme considering 

up to six immediate neighboring tubes. For example, for the circuitry used in the 

evaporators tested under this project and shown in Figure 4.3.3.2.2 the intermediate 

neighbors for tube 25 are tubes 7, 8, 26,44,43, and 24. 
Figure 4.3.3.2.2: Schematic of refrigerant circuitry for COIL-W, COIL-E and COIL-EC 

in cross-counter flow configuration (inlet tubes: 42,48, 54; outlet 

tubes: 1,7, 13) 

Extensive experimenting with this scheme for the coils tested under this project indicated 

that it was important to add additional neighbors to the group of immediate neighbors 

considered so far. This need demonstrated itself not only in predicted capacity values 

but also in simulation runs, which did not yield gradually changing predictions at small 

changes in imposed refrigerant superheat at the outlet tubes. Based on these 

observations, depending on location up to six second-order neighbors were added. These 

are other tubes in the coil assembly that a given tube “can see”. Tube 25 has four 

second-order neighbors; they are tubes 6, 9, 45, and 42. For tube 9, the immediate 

neighbors are tube 8, 10, 27, and 26, and the second-order neighbors are tubes 25, 45, 

and 28. In a three-depth-row coil, the maximum number of second-order neighbors is 

four. A five-depth-row coil is need for a tube located in the middle depth row to have all 

six second-order neighbors. 

The value of the shape factor depends on a fin design. For flat and wavy fins the fin 

material is continuous. Lanced fins, however, have numerous cuts, which reduce the fin 

cross-section area that is available for heat transfer. Hence, the shape factor for flat and 

wavy fins should be expected to have higher values than for lanced or louver fins. Since 

75 
we are not aware of any publication that quantifies the fin shape factor, the values for 

COIL-W and COIL-E were assigned based on their respective results for test 10. Once 

these values for shape factors were assigned, they were left unchanged for the remaining 

simulations. COIL-EC used the same shape factor as COIL-E, however, any tube in 

COIL-EC could have only two neighbors, the closest two tubes located in the same depth 

row. 

Figure 4.3.3.2.3 shows tested and simulated capacities for COIL-E at conditions of test 1, 

2, 9, 10, 13, and 14. For the tests at 5.6 "C (10 OF) superheat (tests 1, 9 and 13), the 

model predicted measured capacities within 5.1 %. For the tests with 16.7 "C (30 OF) 

superheat, the differences in between tested and predicted and tested capacities were - 

6.7 %, 3.1 %, and - 2.9 %. Without accounting for tube-to-tube heat transfer, EVAP5 

would overpredict the capacities at 16.7 "C (30 OF) superheat by approximately 20 %. 

Section 4.4 presents validation results for all three evaporators. 

28000 

26000 

24000 

f 

2 22000 

!2 

c x 20000 

h 

.- ._ 

I v. 

18000 

16000 

0 

14000 

12000 

+Tests 1 & 2; test 

+Tests 1 & 2; simul. 

+Tests 9 & 10; tests 

--Tests 9 8 10; simul. 

+Tests 13 & 14; tests 

+Tests 13 & 14; simul. 

0 5 10 15 20 25 30 35 

Superheat (OF) 
Figure 4.3.3.2.3: COIL-E measured and simulated capacities for tests with 5.6 “C (10 OF) 

and 16.7 “C (30 OF) superheats (tests 1,2,9, 10, 13, 14) 

4.4 

4.4.1 Validation of EVAPS 

The majority of laboratory test data measured in this study are for a cross-counter flow 

configuration with non-disturbed (uniform) air velocity profile at the coil inlet. These 

measurements for COIL-W, COIL-E, and COIL-EC were used to validate EVAPS and 

explore the impact of tube-to-tube heat transfer. The imposed variety of rekgerant 

superheats at individual exit tubes combined with a wide range of air flow provided a 

unique set of challenging test data for validating any evaporator model. 

Validation and Simulations with EVAPS 

The refrigerant circuitry for the tested evaporators has already been presented in Figure 

4.3.3.2.2 as it is displayed by the EVAP-COND interface. Also copying from the 

respective window of EVAP-COND, Figure 4.4.1.1 shows the key design parameters of 

COIL-W. Except for a different fin design, these parameters were the same for COIL-E 

and COIL-EC. 

At the outset of simulations for each coil, EVAPS was “tuned” to predict the performance 

of a given evaporator at the conditions of test 9. This was accomplished by inputting 

“corrections parameters” for the refrigerant heat transfer coefficient, refrigerant pressure 

drop, and air-side heat transfer coefficient. (Section 4.3.2 discusses the reasons for using 

these parameters in the context of prediction discrepancies between different air-side 

correlations). Figure 4.4.1.2 presents the correction parameters for COIL-W and COIL-E 

77 
as they were input into the EVAP-COND window. The input for COIL-EC was different 

by the value for the air-side heat transfer coefficient, which was 0.62 instead of 0.65. The 

1.6 value for the refrigerant pressure drop parameter accounts for the impact of lubricant, 

which can be responsible for 35 % pressure drop underprediction. Since the parameter 

for refhgerant heat transfer coefficient was set to 1.0, the 0.65 or 0.62 value for the airside heat transfer coefficient accommodates the heat transfer adjustment on the air and 

refrigerant sides. These correction parameters were used in all simulations for the 

respective coils. 

Figure 4.4.1.1 : Design parameters for COIL-W 
Figure 4.4.1.2: Correction parameters for COIL-W and COIL-E 

Tables 4.4.1.1, 4.4.1.2, and 4.4.1.3 show tested and simulated total capacities for COILW, COIL-E, and COIL-EC, respectively, in the test matrix format. Tables 4.4.1.4a, 

4.4.1.4b, 4.4.1.5a, 4.4.1.5b, 4.4.1.6a, and 4.4.1.6b present total and sensible capacities, 

sensible heat ratios, and differences between simulated and measured results. It is 

convenient to screen the accuracy of capacity predictions by reviewing Figures 4.4.1.3 

and 4.4.1.4. Figure 4.4.1.3 shows that the maximum error in capacity prediction for wet 

coil tests was 5.5 % for all air velocities and refrigerant superheat scenarios. For dry coil 

tests shown in Figure 4.4.1.4, EVAP5 predicted the three capacities at uniform 5.6 "C 

(10 OF) superheat within 5.0 %. (They are represented by the first bar for each coil). For 

the remaining six tests at different superheat scenarios, the only capacity predicted within 

5.0 % was for COIL-EC (represented by the last bar on the right hand side). Capacities 

for COIL-W and COIL-E capacities were unpredicted by as much as 20.0 %. The 

inability of EVAP5 to account accurately for longitudinal fin conduction for dry coil tests 

can be explained by the fact that the used algorithm considers only temperature 

differences between neighboring tubes (the first order effect) and neglects variations in 

air-side heat transfer. This conclusion agrees with the observation by Romero-Mendez et 

79 
al. (1997) who indicated thermal conductance for convection from the fin as one of the 

parameters affecting tube-to-tube heat transfer. This and other effects identified by 

Romero-Mendez et al. (1 997) should receive detailed attention in a future study dedicated 

to this challenging modeling issue. 
Y 

; 

6 

5 

2 

c, 

c 

c 

f 

+ 

- 

X 

X 
n 

W 

Y 

; 

rn 

2 

s 

m 

0 

N 

_.I 

c 

s + 

rx 

.f= 

cdu- 5s 

u 

d oooommr-wm 

----00000 

2 9999????? 

Q 

Q 

cu 

N 

0 

cu 

0 

3 

z 

Y 

4: 

* 

L: 

7 

0 

rn m 

0 

33 

-N 
m - Y 

tn 1 

w & 

* 

ci 

b 

YW 

c 

0 

01 

-M 

.- Y 

32 

s 

hg 
c1 v1 

d 

a 

E 

Y 

% + 

I 

v 
8 '? 

\o 

m a! 

n 

a! - 

m 

y-! 

42 

u 

c 

s 

-ri 
go - 0 N 0 N Ob 

.e 

v 

I 

I 

I 

- Y v) 

2 

2 

Y m 

G 

c 

0 

c 

8 

-N -N 
5.0 

0.0 

-5.0 

-10.0 __ 

-15.0 

1 

-20.0 

Figure 4.4.1.3: Difference between simulated and measured capacities for all wct coil tests for 

COIL-W, COIL-E, and COIL-EC 

10.0 

5.0 

0.0 

-5.0 

-10.0 

-1 5.0 

-20.0 

COIL-E oCOIL-EC 

I 

, 

i 

Figure 4.4.1.4 Difference between simulated and measured capacities for all dry coil tests for 

COIL-W, COIL-E, and COIL-EC 

89 
Tube-to-tube heat transfer demonstrates itself in temperatures that can be measured on coils 

return bends, as it was shown for COIL-W in Figures 3.4.4.1 and 3.4.4.2. Figure 4.4.1.5 shows 

similar information (refrigerant temperature at tube exits) as it is displayed by EVAP-COND for 

the same tests. For test 9 with even refrigerant superheat, refrigerant temperatures are similar for 

each circuit; refrigerant temperatures reflect drop in refrigerant pressure until the last two tubes 

in each circuit (2 and 1, 8 and 7, and 14 and 13) in which the refrigerant is superheated. For test 

12, the first circuit is in two-phase flow until the exit tube 1, while the refrigerant leaving two 

other exit tubes (7 and 13) is highly superheated. Tubes 7 and 8 experience a drop in 

temperature compared to tube 9 because of their vicinity to the left-hand side circuit with twophase, low-temperature refrigerant. Tubes 13 and 14 also experience temperature drop, 

however, it is small because the adjacent tubes are also superheated. This simulation results 

agree in principle with the measured return bend temperature of Figures 3.4.4.1 and 3.4.4.2. 

90 
V air flow 

q----------- 0 q----------- Q Q - - - <- - - - - - - 0 48.7 48.9 49.1 49.1 49.3 48.6 48.7 48.9 48.0 49.1 49.3 

- - - e-2- - - - e-2- - - - p s---q 

*---++++--d e--- --- --- 0 - - - - - - * --- 47'7 4/47'3 

- - -€-I& - - - 9 59'9L - - -- L--- ->---$I 

e----------- 71'6 z7 

47.7 d47,3 48.4 48.2 48.0 47.8 

48'6L- 48.6 48.7 48.0 49.1 49.1 49.3 

\-.-c-$ ____ ----Q 

48.4 48.2 48.0 47.8 47.7 47.3 48.4 48.2 48.0 47.8 

57.9 52.3 46.0 46.4 46.8 47.1 57.4 52.0 46.0 46.4 46.8 47.1 57.0 51.3 46.0 46.4 46.8 47.1 

I 

Refrigerant exit temperatures (OF) for COIL-W, test 9 

I 

q-------<----o q---c--------Q Q-----------O 

58.8 46.3 45.3 46.5 45.5 50.9 46.5 46.7 45.7 46.7 

- - - 

e--- - - - ?Q-g - - - 

72.1 j3.4 66.9 68.7 69.8 70.5 

w'3L 48.9 49.1 40.1 49.3 49.5 49.8 

0 b ---- c)--s------- \ 

48.7 48.4 48.2 48.0 47.8 47.5 65.1 69.3 70.5 71.2 

47.5 45.3 46.9 46.4 46.8 47.1 68.4 72.1 73.2 73.0 74.8 76.1 74.5 74.8 75.0 75D 75.0 74.7 

Refrigerant exit temperatures (OF) for COIL-W, test 12 

Figure 4.4.1.5: Tube numbering and refrigerant exit temperatures for individual tubes for test 9 

and test 12 for COIL-W, (inlet tubes: 42,48, 54; outlet tubes: 1, 7, 13) 

Figure 4.4.1.6 presents refrigerant exit qualities for individual tubes for COIL-E and COIL-EC 

for test 10 (16.7 "C (30 OF) even superheat). The figure shows that refrigerant reaches superheat 

(quality 1) two tubes earlier in COIL-E than in COIL-EC in each refrigerant circuit. This is a 

result of heat transfer between different tube depth rows that was allowed in COIL-E and was 

inhibited in COIL-EC. Corresponding tube temperatures predicted by EVAPS agree with those 

measured during laboratory tests. 

91 
~ ~~~ 

Refrigerant exit qualitiesfraction) for COIL-EC, test 10 

Figure 4.4.1.6: Refrigerant exit qualities for individual tubes for COIL-E and COIL-EC test 10 

(16.7 "C (30 OF) even superheat) 

The last two columns in Tables 4.4.1.4, 4.4.1.5, and 4.4.1.6 compare simulatcd capacities that 

were obtained with and without accounting for tube-to-tube heat transfer. For the tests with a 

uniform superheat of 5.6 "C (1 0 OF), the difference in capacities is not greater than 2.7 Oh for any 

of the three coils. For the tests with uniform superheat of 16.7 "C (30 OF). the capacities differ 

by 63.1 YO and 56.3 YO for COIL-W, 33.6 YO, 18.8 YO, 19.2 YO for COIL-E, and 1.5 O/b. 2.5 %, 2.5 

%, and 4.9 YO for COIL-EC. These results demonstrate the impact of fin dcsign on tube-to-tube 

heat transfer. 

Thevalidation of EVAP5 used the full set of COIL-W, COIL-E, and COIL-EC measurements in 

cross-counter flow configuration, which constituted the majority of the measurcments taken in 

this study. The validation effort was not extended to the eight data points takcn in cross-parallel 

flow configuration because six of these tests that involved 16.7 "C (30 OF) superheat resulted in 

severe pinching within less than 1 "C (1.8 OF) in at least one of the circuits. Such a close 
approach of refrigerant and air causes a profound convergence problem for a tube-by-tube model 

in which air temperatures upstream of each tube have to be iterated around a target value. 

Furthermore, for evaluating the potential reduction in heat exchanger volume, only capacity 

predictions at test 9 with 5.6 "C (10 OF) superheat were needed, and these were attainable with 

EVAP5. 

Test 9 measured capacities for COIL-W and COIL-E were 4729 W (16146 Btu/h) and 4549 W 

(15522 Btuh), while EVAP5 predictions were 4796 W (16366 Btu/h) and 5482 W (18705 

Btu/h), respectively. This is a very good prediction for COIL-W, within 1.3 YO, while the 

discrepancy for COIL-E is 20.5 %. It should be noted that a higher capacity should be expected 

for COIL-E than for COIL-W, as it was predicted by EVAPS and always was obtained from 

laboratory measurements except this time. It is possible that some condensate holdup might 

have influenced the measured capacity for COIL-E. With this, it was concluded that EVAP5 

properly simulated coils in a cross-parallel flow set up. Consequently, COIL-W was applied in a 

later section to examine potential savings in evaporator core volume for the cross-parallel 

configuration. 

4.4.2 

4.4.2.1 

Considering similar performance degradations for different refiigerant superheat scenarios, 

possible savings in heat exchanger material are demonstrated using test 12 of COIL-W as an 

example. In our simulations, it was assumed that smart refrigerant distributors would optimize 

refrigerant distribution so the evaporator obtains maximum capacity. In these tests with a 

Possible Savings in Heat Transfer Area Due to Optimized Superheat 

Cross-Counter Flow Configuration with Uniform Air Flow Distribution 

93 
uniform air velocity profile, the evaporators reached maximum capacity when the refrigerant 

split between the three circuits resulted in uniform superheat at the individual outlet tubes. 

For these simulations, five alternative coils were coded with a smaller number of tubes than 

COIL-W. All simulation runs had the same inlet air condition, refrigerant inlet quality, and 

refrigerant outlet pressure and superheat at the evaporator exit. Also, inlet air velocity was the 

same for each coil as for COIL-W. A coil with a smaller face area had a lower volumetric flow 

rate than COIL-W, proportional to the percentage that its face area was reduced. 

Figure 4.4.2.1.1 shows coil designs and simulation results. Four out of five alternative coil 

designs offered both savings in the heat exchanger core volume and an increase in coil capacity. 

The coils with a lower volumetric flow of air would also provide savings in fan power. 

94 
-o--H 

14 I5 

Q = 16016 Btdh 

Air flow rate = 732.7ft3/min 

HX vol. core saved = 33.3 % 

QIQtest12 = 1.2 1 

Q = 13622 Btu/h 

Air flow rate = 65 1.4 ft3/min 

HX vol. core saved = 40.7 % 

QlQtest12 = 1-03 

~ ~~ 

Air flow rate = 61 0.6 ft3/min 

QIQtest12 = 0.93 

Q = 17387 Btdh 

Air flow rate = 61 0.6 ft3/min 

HX vol. core saved = 20.0 % 

Q/Qtest12 = 1.31 

Q = 14712 Btu/h 

Air flow rate = 488.5 ft3/min 

HX vol. core saved = 33.3 % 

Q/Qtest 12 = 1 .I3 

Figure 4.4.2.1.1 : Simulation results for alternative coil designs to COIL-W in cross-counter flow 

configuration. Performance is compared to COIL-W test 12 with simulated 

capacity of 13225 Btu/h. 

4.4.2.2 Cross-Parallel Flow Configuration with Uniform Air Flow Distribution 

Simulations were also performed to demonstrate possible savings in coil material (core volume 

of the heat exchanger) for COIL-W in cross-parallel configuration. Test 9 (W020304a) was used 

as a reference. Alternative coil designs with two depth rows were only examined because, in the 

cross-parallel configuration, more depth rows are not beneficial due to pinching. 

95 
All simulations were run using test 9 operating conditions, including 6 "C (1 0.8 OF) superheat, 

with the difference that the volumetric flow of air was adjusted so that each coil had the same 

inlet air velocity as COIL-W. This means that a coil with a smaller face area had a lower 

volumetric flow rate than COIL-W by the same percentage its face area was reduced. 

Figure 4.4.2.2.1 shows coil designs and simulation results. Each of the four presented two-depth 

row designs offered improved capacity and savings in coil core volume. The coils with a lower 

volumetric flow of air would also provide savings in fan power. The smallest coil with 12x2 

tube arrangement matched the capacity of test 12 with a 33.3% savings in coil material. 

96 
Air flow rate = 536.7 fi3/min 

Q/Qtest12 = 1.99 

Q = 1 I876 Btdh 

Air flow rate = 477.1 ft3/min 

HX vol. core saved = 40.7 % 

Q/Qtcst12 = 1.62 

I t- ct-8- 14 15 -2 I 

Air flow rate = 447.3 ft3/min 

QIQtert12 = 1.33 

Q = 7378.7 Btdh 

Air flow rate = 357.8 ft'/min 

HX vol. core saved = 33.3 % 

Q/Qtest12 = 1 .O 1 

Figure 4.4.2.2.1 : Simulation results for alternative coil designs to COIL-W in cross-parallel 

configuration. Comparisons are to COIL-W, test 12 with tested capacity of 

731 1 Btu/h. 

4.4.2.3 Cross-Counter Flow Arrangement with Non-Uniform Air Flow Distribution 

The tests performed in the lab with non-uniform air distributions resulted in complicated velocity 

profiles that currently cannot be reproduced in EVAP5 simulations. For this reason, the 

simulations were performed with one-dimensional, non-uniform velocity profiles that were 

independent of the tests performed in the lab and represented a different application case 

scenario. For these simulations, COIL-W test 9 with a uniform air distribution was selected as a 

reference test, and additional simulations were performed for two-step velocity profiles where 

the top (left) to bottom (right) velocity ratios were 1 : 1.5, 1 :2.0, 1 :2.5, 1 :3, 1 :3.5, and 1 :4. 

97 
Two runs were performed for each velocity ratio: the first - with a uniform refrigerant 

distribution, and the second - with a refrigerant distribution optimized to obtain maximum 

capacity. During all these tests, the external run parameters (refrigerant inlet state, exit pressure 

and superheat, air flow rate, etc.) were the same. Figure 4.4.2.3.1 presents a velocity profile 

representation for the 1 :3 ratio as it was input into EVAP-COND. Since our velocity profiles 

have a near step change, they represent a more radical departure from uniformity than the 

profiles obtained in the laboratory. 

Table 4.4.2.3.1 summarizes simulation results, and Figure 4.4.2.3.2 presents simulated capacities 

as referenced to the capacity of test 9. The table and figure show that the capacity degrades 

linearly with degradation of the air velocity. For the 1:4 air velocity ratio and uniform 

refrigerant distribution, the obtained capacity was only 63 YO of the reference test 9 value. 

However, with optimized refrigerant distribution, as is the purpose of smart distributors, the 

obtained capacity was within 7 YO of the reference capacity. 
Location 

Figure 4.4.2.3.1 : Air velocity profile representation for the 1 :3 top-to-bottom velocity ratio 

(during tests the coil was positioned vertically, turned clockwise by 90 ") 

Table 4.4.2.3.1 : Simulated Capacities and Refrigerant Distributions for Non-Uniform Inlet Air 

99 
1.100 

1 .ooo 

0.900 

0.800 

0.700 

0.600 

0.500 

II I 

8 rn 8 8 

8 optimized refrigerant mass flow rate 

uniform refrigerant mass flow rate 

1 1.5 2 2.5 3 3.5 4 4.5 

Air velocity ratio (bottomltop) 

Figure 4.4.2.3.2: COIL-W capacities at different air velocity ratios referenced to capacity at test 

9 in cross-counter flow configuration 

To assess the savings in the heat exchanger material due to optimized control of refrigerant 

superheat, two evaporators with reduced number of tubes were coded, shown in Figure 4.4.2.3.3, 

and simulations were performed for 1:2.5 and 1:4 air velocity profiles. The two-depth row 

evaporator had the same face area as COIL-W and was simulated with the same volumetric flow 

rate of air. For the three-depth-row, the volumetric flow rate was reduced by 16.7 %, which 

corresponds to the reduction of the coil face area in relation to that of COIL-W. The results 

presented in Table 4.3.3.1.1 show that the benefit of optimizing refrigerant distribution increases 

with the level of non-uniformity in the air velocity profile. For the 1:4 air velocity ratio, 

optimizing refrigerant distribution allows a reduction in coil volume of 33.3 YO. For the 1 :2.5 air 

velocity ratio, the use of 15x3 coil with a slightly increased volumetric flow rate could produce a 

16.7 'KO reduction in the coil volume. 
Coil 15x3 

‘Oil 

Coil 18x2 

Figure 4.4.2.3.3: Two evaporators with a reduced number of tubes 

Air Savings 

watt flow rate Btu/h volume 

Capacity volumetric Capacity in coil 

fi3/min % 

Air velocity 

(top to bottom) 

Refkgerant ratio 

distribution 

Table 4.4.2.3.2: Savings in Coil Volume Relative to COIL-W due to Optimizcd Rcfrigerant for 

1 :2.5 and 1 :4 Air Velocity Ratios 

COIL-W 

COIL-W 

COIL-W 

uniform 1:l 20.7 6485 733 22 127 0 

optimized 1 : 2.5 20.7 6305 733 21515 0 

uniform 1 : 2.5 20.7 5256 733 17935 0 

- 18x2 

15x3 

I I 

optimized 1 : 2.5 20.7 4558 733 15553 33.3 

oDtimized 1 : 2.5 17.3 4890 61 1 16685 16.7 

COIL-W 

COIL-W 

uniform 1:l I 20.7 6485 73 3 22 I27 0 

outimized 1 :4 I 20.7 6059 733 20675 0 

COIL-W 

18x2 

15x3 

101 

uniform 1 :4 20.7 41 10 73 3 13024 0 

optimized 1 :4 20.7 4335 73 3 14700 33.3 

optimized 1 :4 17.3 4474 61 1 15266 16.7 
5 CONCLUSIONS 

This collection of experimental data for the three evaporators has revealed interesting results 

related to non-uniform refrigerant distribution and conduction between tubes through the fins. 

With cross-counter refrigerant flow, uniform airflow, and exit manifold superheat fixed at 5.6 "C 

(10.0 OF), the wavy fin and wavy-lanced fin evaporator's capacity dropped by as much as 41 % 

and 32 YO, respectively, as the superheat was allowed to vary between the circuits. Control of 

superheat was shown to be even more important during cross-parallel refrigerant flow due to the 

rapid pinching of the refrigerant and air temperatures. For the wavy and lanced finned 

evaporators in cross-parallel flow, capacity dropped by 85 % and 78 % as superheat changed 

from 5.6 "C (10.0 OF) to 16.7 "C (30.0 OF). 

As the coil's faces were blocked to produce a non-uniform airflow, control of superheat was 

shown to restore capacity if the volumetric flow of air was unchanged. The tests showed that 

when airflow rate was held constant, the losses in capacity due to non-uniform airflow could be 

recovered to within 2 % of the original uniform airflow capacity by controlling superheat. The 

more non-uniform the airflow over the coil, the greater was the benefit of controlling superheat. 

For the lanced fin coil, as the airflow ratio between the top half and lower half of the coil varied 

from 1 : 1.26 to 1 :2.59, superheat control improved capacity by 1.4 % and 4.6 YO, respectively. 

In parallel with the experimental effort, the NIST evaporator model EVAPS was upgraded to 

control refrigerant distribution and account for tube-to-tube heat transfer. The model was 

validated with the experimental results and then used to determine the possible savings in 
evaporator core volume if refrigerant distribution was controlled by a smart distributor. 

extreme cases, the savings in core volume could be as much as 40 %. 

In 

A combination of results obtained from laboratory testing and simulations indicated the influence 

of tube-to-tube heat transfer on capacity degradation. The impact of tube-to-tube heat transfer 

was negligible in tests with a uniform 5.6 "C (10 OF) superheat, but it was significant in tests 

involving 16.7 "C (30 OF) superheat. Between two possible conduction mechanisms by which 

such heat transfer may occur, longitudinal fin conduction was chiefly responsible for degraded 

performance while longitudinal tube conduction had insignificant effect. The upgraded version 

of the EVAPS evaporator model, which accounts for tube-to-tube heat transfer based on tube 

temperatures, was able to predict key return bend temperatures that indicated the occurrence of 

tube-to-tube heat transfer. However, the study also confirmed that longitudinal heat conduction 

is affected by the fin design, air-side heat transfer coefficient, and moisture removal process. 

Consequently, a more detailed modeling scheme needs to be developed to capture other effects 

influencing tube-to-tube heat transfer. Such a study would not only improve the modeling of 

evaporators but also of condensers and of gas coolers, where internal heat tansfer may be even 

more pronounced. 

103 
APPENDIX A. SUMMARY OF TEST RESULTS 

A. 1 Wavy fin evaporator in cross-counter flow 

Test names 

Table A. 1.1 : Wavv Fin EvaDorators in Cross-Counter Flow 

Test type 

W020225B 

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5 

6 

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W020225A 

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A.2 Wavy fin evaporator in cross-parallel flow 

Test names Test type 

W020304A 

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9 

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Table A.4.1: Enhanced Fin (Wavy-Lanced) Evaporators in Cross-Parallel Flow I 

132 
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7 Test names Test type 

E0204 17A 1 

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Test names Test type 

W020522A 9 

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A.7 Enhanced fin (wavy-lanced) evaporator with non-uniform airflow 

Test names 

E020604A 

Table A.7.1: Enhanced-Cut Fin (Wavy-Lanced) Evaporator with Non-Uniform Airflow 

Test type' Velocity ratio 

9 1:l 

E020604B 

E020605A 

E020607A 

9A 1:1.26 

9B 1 : 1.26 

9 1 :1 

E020607B 

E0206 1 OA 

E0206 1 1 A 

E0206 12A 

E0206 1 3A 

E020620A 

E02062 1 A 

E020624A 

airflow and no superheat adjustment, 9B: expansion valves adjusted to yield 5.6 OC (10.0 OF) superheat on all EIKUII\ with ncm-uniform airflow. 

..- 

9A I : I .36 

9B 1 : 1.36 

9A 1 : 1.62 

9B 1 : 1.62 

9A 1 : 1.75 

9B 1 : 1.75 

9A 1 :2.59 

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APPENDIX B. CAPACITY UNCERTAINTY 

Coil 

Designation Test Name 

Table B.l lists the relative uncertainty in the air-side capacity for two representative tests at low 

Percent Uncertainty 

at a 95 YO Confidence kW (Btu/h) Description kW (Btu/h) Limit on the Mean 

Capacity, Uncertainty Capacity, 

and high evaporator capacity. Two tests are shown below, with the first test being a typical test 

E0204 1 6B 

W0203 1 1 B 

at a capacity comparable to a majority of the other tests for all coils. The second test listed in 

Typical of 7.8 (26546) 4.2 

all tests 7.8 (26546) Enhancedcut 

Wavy 0.90 (3078) Worst case 0.90 (3078) 8.9 

Table B.l shows a worst case test for COIL-W in parallel flow with an extremely low capacity. 

For the majority of tests, the uncertainty in the evaporator capacity was at the 4 % to 5 'YO level. 

A complete description of the propagation of error technique used to calculate uncertainty is 

given in Payne and Domanski (2001). 

Table B. 1 : Relative Uncertainties of Two Evaporator Tests 

166 

-_ 
APPENDIX C. USER’S INSTRUCTION FOR THE EVAP-COND VERSION 

USED IN THIS STUDY 

A CD attached to this report contains a version of EVAP-COND that was specifically developed 

for this study. The following pages describe how to install and use the model. As needed for 

this study, this version of EVAP-COND simulates only evaporators with multiple inlets using 

the option that solicits refiigerant outlet saturation temperatures and global superheat. This 

option is identified in the figure below with the EVAPORATOR OPERATING CONDITIONS. 

The condenser, which normally is included in the EVAP-COND package is not provided here. 

The attached CD package contains the following two files: 

EV-CD.exe - self-extracting file with all files needed for 

executing EVAP-COND. 

EVAP-COND instructions.pdf - file with visual instructions on how to use EVAPCOND. (You need Version 5 of Adobe Reader to 

read this file.). The instructions are also included 

in this appendix. 

Installation of EVAP-COND on your PC 

Execute file EV-CD.exe to expand it on your hard drive. You will be prompted to select a 

directory where you want the program to reside. When the installation is completed, you should 

see EVAP-COND directory and two subdirectories called FLUIDS and MIXTURES. 
In the main directory (EVAP-COND), EVAP-COND.exe is the interface. EVAP5.exe is the 

evaporator. Files with the affix.dat are example cases of input data used in this study. Files with 

the affix.opc extension contain corresponding operating conditions. 

Next step 

Refer to the following pages or EVAP-COND instructions.pdf for further information about 

EVAP-COND capabilities and limitations. They will also assist you in your first evaporator 

simulation run. It is recommended for the user to follow the steps described there to familiarize 

yourself with the model. Because of constant upgrading of the model, thc simulation results you 

are going to obtain may not be the same as those presented in EVAP-COND instructions.pdf. 

Control of the option to simulate with or without longitudinalfin conduction 

The user can control the option of using longitudinal fin conduction in a simulation by accessing 

file TUNE.TXT selecting 0 or 1 for the flag, as it is explained in the filc. and saving the file. 

TUNE.TXT is located in the EVAP-COND directory. 

The following pages contain general visual instructions for using EVAP-COND as they are 

presented in the file EVAP-COND instructions.pdf located on the attached CD. The option 

developed for this project is marked and available in this package is marked. 

168 
I 'I NAPCOND is a software package 

that contains NIST's simulation models 

pages prwlde basic instructions on hov; 

to use this package. The instructions 

indude preparation of inwt data, 

Capabilities include: 

tube-by-tube simulation 

non-uniform air distribution 

simulation of refrigerant distribution 

- condenser model cawble of 

simulattims above the critical mint 1 I Piotr A Domanski I 

169 
. - 

I L...,,." 

1 70 
IRCUIT DESIGN 

171 
172 
173 
EVAPS OPENING WINDOW 

EVAPS REF 

. DISTRIBUTILQN 

174 
175 
LATION SUMMARY (cont.) 

(Complete printoul of file Jns) 

1 

NAPS SILmLITION SLlUllARV-- 

CDll ID MULPLE-5 

REFRIGERANT R22 

120 0 Ikwh] 3 735 [bU] 

1 b77 [kUl 5 413 [hU] 7 0 6 2 IC] 8 7 11 I [Cl b55 0 622 9 [LPsl 0 200 1 000 

1 0 989 76 00 

16 1 000 23 0 16 0 

Multiplier for mfras heat transtor -11 1 00 

Multinlia for airaide h-t trans -11 1 00 MultlDller for rufrlgsr.nt dmp 1 00 

MULATE EVAPORATOR ? 

Run Windows Explorer and go to the directory containing EVAP-COND exe 

Doubleclffik on EVAP-COND exe to start the program 

Open file EXAMPLE-5 dat to smulate the evaporator After the file is loaded you will see a schematic representing a 

slde VI%W of the evaporator The red cwcle(s) indates the inlet tube lo Ihe evaporator The blue cncles indlcate the 

outlet tubes The hortzontal line at the bottom of the screen indmtes the ar velocdy proliie at the evapMator mlel 

Revmw Input Data Click on the EdrVCoil Design menu dem to review the evaporator design nbrmation You may 

select elther the SI or Bntish system of units for your input data and sirnulaton resuns 

Clck on the EdMOperating Condifiono/Evaporamr/inlel pressure and quaiffy menu dern to review operatum 

condiilons Note that the loaded optwn has a mark on the left-hand side Since EVAPS simulates perfmance tubeby-tube from the mk4 to outlet the opmns thal speedy any outlet refrigeranl parameler involve fieratwe cans to the 

optlon that spec~Ms refrlgerant inlet pressure and quality until the larger outlet parameters are obtained (eg 

saturabn temperature and superheat) 

Clck on the €dl#e/ocify Profile menu dem to review the air velocity profile You may use the air mass flow rate 

specmsd earlmr in the Operating Conditions mndow or integrate the air velocdy profile In general the first option IS 

recommended unless very delalled and accurate local measurements of the veloclty profile were taken You may 

change the ab- veloclty prof& using a mouse by clcking the left button 

Run a simulation Clck on the Run Simulation menu item and select EVAPS An MS-DOS mndow wlll appear and 

wlll give you a message when a snnulatlon run is successfully completed 

Examine local and global Simulation results EVAPB bwMles global simulation resuns to file SI res (SI system of units) 

and BT res (Brdrh system of units) The same infwmatron IS provlded in the pulldown menu in the unts setected for 

data input 

176 
PREPARE OUR DATA F 

SekI IEdWOpemtmg Conditions menu dem to input operabng condlbons data 

Select Ediuverocity Protile lo change rhe veloclty profile using a mouse (I& button) 

Spec@ refngecanl crcuary 

If yw are coding evaporatw cwcurtry. start mth one of the inlet lubes and proceed dormstream If you are coding 

condenser circultry, start wkh one of the outlet lubes and proceed upslream. I e, in either case yx~ have 

to start frcin the sde that Is char to the saturated lquid lma 

To draw a return bend, pml the mouse on a tube, press the left button, drag the mouse to the next lube, and release 

If yuu wan1 lo moddy a crcuiey. you may delete a part of it starting from a gwen tube end ending by the exlt tube by 

Once a CKCUII IS coded. It can be used for both evaporator and condenser elmulalions based on specmed operating 

pomnbng the mouse on the gwen tube and double-clcking the left button 

wndltlons I I 

SUGGESTIONS 

QUESTIONS 

3 rn 

National Institute of Standards and Technology 

Building and Fire Research Laboratory 

Gaithenburg, MD, USA 

I 

177 
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