Design & Engineering Services - ETCC ca

Design & Engineering Services

DEVELOPMENT OF A FAULT DETECTION AND

DIAGNOSTICS LABORATORY TEST METHOD FOR

A RESIDENTIAL SPLIT SYSTEM

HT.11.SCE.003 Report

Prepared by:

Design & Engineering Services

Customer Service Business Unit

Southern California Edison

July 2012

Development of a FDD Laboratory Test Method for a Residential Split System HT.11.SCE.003

Acknowledgements

Southern California Edison’s (SCE’s) Design & Engineering Services (DES) group is

responsible for this project. It was developed as part of Southern California Edison’s HVAC

Technologies and System Diagnostics Advocacy (HTSDA) program under internal project

number HT.11.SCE.003. DES project manager Sean Gouw conducted this test method

development with overall guidance and management from line manager Ramin Faramarzi,

and HTSDA program manager Jerine Ahmed. For more information on this project, contact

[email protected].

Disclaimer

This report was prepared by Southern California Edison (SCE) and funded by California

utility customers under the auspices of the California Public Utilities Commission.

Reproduction or distribution of the whole or any part of the contents of this document

without the express written permission of SCE is prohibited. This work was performed with

reasonable care and in accordance with professional standards. However, neither SCE nor

any entity performing the work pursuant to SCE’s authority make any warranty or

representation, expressed or implied, with regard to this report, the merchantability or

fitness for a particular purpose of the results of the work, or any analyses, or conclusions

contained in this report. The results reflected in the work are generally representative of

operating conditions; however, the results in any other situation may vary depending upon

particular operating conditions.


Southern California Edison Page iii Design & Engineering Services July 2012

EXECUTIVE SUMMARY This project intends to break new ground in the world of fault detection and diagnostics

(FDD), and improved heating, ventilating, and air conditioning (HVAC) performance through

enhanced maintenance. The project’s goal is to develop a laboratory test method for FDD

technologies for a residential split system air conditioner. The test method involves

imposing single and multiple cooling-mode faults under steady-state conditions. The test

method will evaluate an FDD technology in project HT.11.SCE.005 (“Laboratory Assessment

of a Retrofit Fault Detection and Diagnostics Tool on a Residential Split System”), and

evaluate fault impacts in project HT.11.SCE.007 (“Evaluating the Effects of Common Faults

on a Residential Split System”). The test method presented in this report is a first step, and

is a part of many ongoing efforts needed to explore solutions to the complex issues inherent

with FDD and HVAC performance and maintenance.

The approach to test method development was to leverage as much industry knowledge as

possible. This project sought feedback through engagement of a Technical Advisory Group

(TAG) and the Western HVAC Performance Alliance (WHPA) FDD subcommittee. AHRI

210/240 and methods from a previous FDD/HVAC maintenance study conducted at

Southern California Edison’s (SCEs) Technology Test Centers (TTC) provided the foundation

for this test method. Through use of the developed test method in conducting various fault

test scenarios, several technical challenges were encountered. The test method was

modified and enhanced to address the varying challenges.

This project successfully developed a steady-state test method suitable for simulating HVAC

faults in a laboratory environment in a repeatable manner. The best available information

was leveraged, but the resulting test method is not intended to be the final and universal

solution. Transient impacts of faults, as well as fault severity and prevalence actually

experienced in the field, are not captured with this test method.

To promote acceptance of a test method in the HVAC industry, its results must be

reasonably reflective of actual field conditions. The following areas are highlighted as

opportunities to improve and enhance the test method:

- Quantify fault severity and prevalence with field studies for laboratory test result

calibration and prioritization of future laboratory tests

- Investigate and simulate transient impacts of faults associated with cyclic laboratory

testing to ultimately determine if the additional test burden is necessary

In addition, it is important to determine what modifications are necessary to apply the test

method to meet the needs of the industry. Certain factors determine the plausibility of

testing for different entities. For example, it may be unreasonable to expect a manufacturer

to test high volumes of equipment with the same methods an academic facility might

employ to test a handful of systems. In order for the test method to be ultimately

successful, the following activities are recommended for future investigation:

1. FDD evaluation purposes for California utility incentive programs for HVAC

maintenance

2. Test requirements for future standards such as (but not limited to) voluntary

standards for the American Society of Heating, Refrigerating, and Air Conditioning

Engineers, Energy Star, Department of Energy Advanced Rooftop Unit Specifications,

federal rulemakings for residential or commercial equipment, and California Title-24

Investigation into these activities is paramount in promoting industry acceptance of an FDD

laboratory test method. Industry acceptance of an FDD laboratory test method is essential

in allowing broader adoption of FDD technologies and enhanced HVAC maintenance and

performance, as well as standardization of terminology and practices in the industry.


Southern California Edison Page iv

Design & Engineering Services July 2012

ABBREVIATIONS AND ACRONYMS

AFDD Automated Fault Detection and Diagnostics

AHRI The Air Conditioning, Heating and Refrigeration Institute

AMB Ambient

ANSI American National Standards Institute

ASHRAE The American Society of Heating, Refrigerating and Air Conditioning Engineers

BACnet Building Automation and Control Networks (Communications Protocol)

Btu British Thermal Unit

CASE Codes and Standards Enhancement

CI Capacity Index

COA Condensing (temperature) Over Ambient

CT Condensing Temperature

CZ Climate Zone

DB Dry-Bulb Temperature

DES Design and Engineering Services

DP Dew Point

EE Energy Efficiency

EER Energy Efficiency Ratio

EI Efficiency Index

ET Evaporator Temperature (Saturated)

ETO Education, Training, and Outreach

°F Degrees Fahrenheit

FDD Fault Detection and Diagnostics

hr Hour


Southern California Edison Page v


HTSDA HVAC Technologies and System Diagnostics Advocacy

HVAC Heating, Ventilating, and Air Conditioning

ITD Indoor Temperature Drop

kW Kilowatt

kWh Kilowatt-hour(s)

lbs. Pounds

ozs Ounces

LP Liquid Pressure

LT Liquid Temperature

min Minute

PDA Personal Digital Assistant

PIER Public Interest Energy Research

psi Pounds per square inch

°R Degrees Rankine

RA Return Air

RH Relative Humidity

RTU Rooftop Unit (Packaged)

RWB Return Wet-Bulb

SA Supply Air

SC Sub-cooling

SCE Southern California Edison

SCFM Standard Cubic Feet per Minute

SH Superheat

SME Subject Matter Expert

SP Suction Pressure


Southern California Edison Page vi Design & Engineering Services July 2012

ST Suction Temperature

SWB Supply Wet-Bulb

TAG Technical Advisory Group

TR Ton of Refrigeration

TTC Technology Test Center

TxV Thermostatic Expansion Valve

T/C Thermocouple

W Watt

WB Wet-Bulb Temperature

WHPA Western HVAC Performance Alliance


Southern California Edison Page vii Design & Engineering Services July 2012

CONTENTS

EXECUTIVE SUMMARY _______________________________________________ III

INTRODUCTION ____________________________________________________ 1

The FDD Project Series ............................................................ 1

Industry Input ........................................................................ 1

The Technical Advisory Group ................................................... 2

Problem Definition ................................................................... 2

Fault Detection and Diagnostics ................................................ 3

OBJECTIVE _______________________________________________________ 5

TEST METHOD DEVELOPMENT STRATEGY _________________________________ 6

THE HVAC TEST UNIT _______________________________________________ 7

THE FDD TECHNOLOGY _____________________________________________ 8

TEST EQUIPMENT INSTALLATION, INSTRUMENTATION, AND DATA ACQUISITION ____ 10

THE TEST METHOD _________________________________________________ 20

Control Parameters and Test Intervals ..................................... 20

Calculations .......................................................................... 21

Test Scenarios ...................................................................... 32

FDD Performance Testing ....................................................... 35

Baseline Testing .................................................................... 39

Single Fault Testing ............................................................... 44

Multiple Fault Testing ............................................................. 54

TEST METHOD CONCLUSIONS AND RECOMMENDATIONS ____________________ 59

APPENDIX _______________________________________________________ 61

REFERENCES _____________________________________________________ 64


Southern California Edison Page viii Design & Engineering Services July 2012

FIGURES Figure 1. Age of Central Air Conditioners in California ..................... 3

Figure 2. HVAC Test Unit - Indoor Unit (Left), Outdoor

Condensing Unit (Right) ................................................ 7

Figure 3. The FDD Technology ..................................................... 8

Figure 4. Test Setup - Comprehensive Diagram ........................... 11

Figure 5. Refrigerant-Side State Points ....................................... 12

Figure 6. Air-Side State Points ................................................... 13

Figure 7. Indoor HVAC Unit Sensor Placement ............................. 14

Figure 8. Condensing Unit Sensor Placement ............................... 15

Figure 9. Comparing Baseline Air-side Cooling Capacity

Calculations – Gross, Sensible, and Psychrometric-

based Latent ............................................................. 30

Figure 10. Comparing Baseline Air-side Cooling Capacity

Calculations – Gross, Sensible, and Scale-based Latent ... 30

Figure 11. Comparing Baseline Heat Rejection Calculation

Averages – Refrigerant Enthalpy Method and the Sum

of Air-side Gross Cooling Capacity and Compressor Heat

of Compression .......................................................... 31

Figure 12. P-H Diagram: Baseline Refrigeration Processes at

Varying Operating Conditions....................................... 40

Figure 13. P-H Diagram: Repeating the AHRI Test Condition ........... 41

Figure 14. Line Restriction Valve ................................................. 47

Figure 15. Non-Condensables - Nitrogen ..................................... 48

Figure 16. Evaporator Airflow Reduction ....................................... 50

Figure 17. Condenser Airflow Reduction ....................................... 52


Southern California Edison Page ix


TABLES Table 1. List of Measurement Points .............................................. 16

Table 2. Accuracy, Calibration Dates and Locations, and

Corresponding Key Monitoring Points for Sensors Used ... 19

Table 3. Control Parameters ........................................................ 20

Table 4. Calculation Methods........................................................ 21

Table 5. Baseline Gross Cooling Capacity: Refrigerant Enthalpy

Method vs. Compressor Regressions .............. 24

Table 6. Baseline Heat Rejection: Refrigerant Enthalpy Method vs.

Compressor Regressions ............................................. 24

Table 7. Baseline Compressor Power: Measured vs. Compressor

Regressions ............................................................... 25

Table 8. Baseline Refrigerant Mass Flow: Refrigerant Enthalpy

Method vs. Compressor Regressions .............. 25

Table 9. Comparing Evaporator Air Volumetric Flow Rate: Measured

SCFM vs. SCFM/Pressure Drop Equation Method ............ 28

Table 10. Baseline Gross Cooling Capacities: Refrigerant Enthalpy

Method vs. Air Enthalpy Method ................................... 29

Table 11. Baseline Test Scenarios ............................................... 32

Table 12. Single-Fault Test Scenarios ......................................... 32

Table 13. Multiple-Fault Test Scenarios ....................................... 34

Table 14. Tracking Test Timestamps ........................................... 37

Table 15. Measurement Averages for The Repeated Instances of

Test 2 ....................................................................... 41

Table 16. Test 2 Data Comparison: 8 lbs. 3 ozs. vs. 8 lbs. 9 ozs. ... 43

Table 17. Comparing a Redundant Outdoor Test Chamber

Temperature Sensor with Average Condenser Inlet

Temperatures for Baseline Tests .................................. 53

Table 18. Tests 43-45: Summary of Airflow and Compressor

Discharge Pressures ................................................... 56

Table 19. Tests 47-55: Summary of Refrigerant Charge, Airflow,

and Compressor Discharge Pressure ............................. 58

Table 20. Summary: Applicable Calculation Methods Per Test

Scenario ................................................................... 61


Southern California Edison Page x


EQUATIONS Equation 1. Energy Efficiency Ratio................................................ 21

Equation 2. Refrigerant-Side Gross Cooling Capacity ....................... 22

Equation 3. Refrigerant-Side Condenser Heat Rejection ................... 22

Equation 4. Refrigerant-Regression Condenser Heat Rejection .......... 23

Equation 5. Calculating Percent Variation ....................................... 23

Equation 6. Calculating Percent Difference...................................... 23

Equation 7. Air-Side Gross Cooling Capacity ................................... 26

Equation 8. Air-Side Gross Sensible Cooling Capacity ....................... 27

Equation 9. Air-Side Gross Latent Cooling Capacity ......................... 27

Equation 10. Air-Side Air Flow Rate ............................................ 27

Equation 11. Heat Rejection ...................................................... 31


Southern California Edison Page 1


INTRODUCTION

THE FDD PROJECT SERIES Southern California Edison (SCE) initiated a series of six projects under the Heating,

Ventilating, and Air Conditioning (HVAC) Technologies and System Diagnostics

Advocacy (HTSDA) program. These projects seek to explore several key items

regarding Fault Detection and Diagnostics (FDD) technologies.

- HT.11.SCE.002 - Development of a Fault Detection and Diagnostics

Laboratory Test Method for a Commercial Packaged Unit

- HT.11.SCE.003 - Development of a Fault Detection and Diagnostics

Laboratory Test Method for a Residential Split System (this report)

- HT.11.SCE.004 - Laboratory Assessment of Retrofit Fault Detection and

Diagnostics Tools on a Packaged Unit

- HT.11.SCE.005 - Laboratory Assessment of Retrofit Fault Detection and

Diagnostics Tools on a Residential Split System

- HT.11.SCE.006 - Evaluating the Effects of Common Faults on a Commercial

Packaged Unit

- HT.11.SCE.007 - Evaluating the Effects of Common Faults on a Residential

Split System

Projects HT.11.SCE.003, HT.11.SCE.005, and HT.11.SCE.007 focus on a residential

split system air conditioner. Projects HT.11.SCE.002, HT.11.SCE.004, and

HT.11.SCE.006 focus on a small commercial packaged rooftop unit (RTU) air

conditioner. The residential and commercial projects work together cohesively to:

- Develop a working laboratory test method

- Apply the working test method in a laboratory assessment project

- Update the working test method, as concurrent with lessons learned in the

laboratory assessment

- Using the data from the laboratory assessment

o Report on FDD performance

o Report on observed effects of faults

INDUSTRY INPUT Industry input was important during development and scoping of the residential FDD

project series. Channels such as the Western HVAC Performance Alliance (WHPA)

provided the means to do so. In particular, the WHPA’s Automated Fault Detection

and Diagnostics (AFDD) subcommittee played an important role in the realization of

the FDD project series.

Involvement with the AFDD subcommittee included frequent updates of concurrent

FDD related efforts. One such effort was a Codes and Standards Enhancement

(CASE) study AFDD proposal for Title-24, Part 6 (2010 California Energy Code). Part

of this effort included listing of the “highest priority” faults for the CASE proposal to

explore. This list, presented and vetted through the AFDD subcommittee, became

the basis of the scope of faults the FDD project series would explore.




The following scope of faults was established for the FDD project series:

1. Low Refrigerant Charge

2. High Refrigerant Charge

3. Refrigerant Liquid Line Restrictions

4. Refrigerant Non-condensables

5. Evaporator Airflow Reduction

6. Condenser Airflow Reduction

THE TECHNICAL ADVISORY GROUP A Technical Advisory Group (TAG) was established to provide support with

specialized HVAC and FDD industry expertise. Specifically, feedback was sought

regarding the test method and the scope of test scenarios to explore. When

establishing the TAG, efforts were made to include as wide a range of participants as

possible. This included outreach to industry members from California utilities,

academia, and FDD and HVAC manufacturers. Included were: The University of

California Davis’ Western Cooling Efficiency Center (WCEC), New Buildings Institute

(NBI), Portland Energy Conservation Inc. (PECI), National Institute of Standards and

Technology (NIST), Climacheck, Field Diagnostics, Pacific Gas and Electric Company

(PG&E), Carrier Corporation, Purdue University, Pacific Northwest National

Laboratory (PNNL), Sempra, Taylor Engineering, and the University of Nebraska.

Several TAG members were also active attendees and participants of the WHPA

AFDD subcommittee meetings. TAG communication occurred through e-mail, phone

calls, discussion in WHPA AFDD subcommittee meetings, and through webinars

conducted on August 22, 2011 and July 11, 2012. Through these means, TAG

feedback was obtained prior to conducting the laboratory assessment and prior to

finalization of the project reports.

PROBLEM DEFINITION California homes consume approximately 85 billion kilowatt-hours (kWh) of

electricity annually.1 Of this, air conditioning equipment accounts for 9 billion kWh, or

around 10%.1 At least 10% of energy consumed by HVAC is wasted from excessive

run time and equipment and controls problems.2 Residential HVAC units also account

for approximately 24% of the peak demand in California.3 For all central air

conditioners in California, nearly half are over 10 years old.1 Figure 1 illustrates the

age of residential central air conditioners in California.

1 EIA Residential Energy Consumption Survey (RECS) 2005.

http://www.eia.gov/consumption/residential/data/2005/ 2 Advanced Automated HVAC Fault Detection and Diagnostics Commercialization Program.

http://www.archenergy.com/pier-fdd/ 3 HVAC Energy Efficiency Maintenance Study.

http://performancealliance.org/LinkClick.aspx?fileticket=FI-4e7Z8kLQ%3D&tabid=220




FIGURE 1. AGE OF CENTRAL AIR CONDITIONERS IN CALIFORNIA

Current HVAC maintenance practices face many hurdles and opportunities for

enhancement. Traditionally, these practices are open to varying interpretations and

are reactive in nature. Homeowners typically do not have maintenance contracts

established for regular servicing of their HVAC equipment. Homeowners usually call

in for maintenance after their equipment fails. In this manner, repair and

maintenance is not necessarily aimed at emphasizing optimization of equipment

efficiency.

FAULT DETECTION AND DIAGNOSTICS FDD technologies interpret parameters to detect symptoms of a faulty operating

state, and diagnose their root cause(s). FDD technologies may be built-in systems

for HVAC units or retrofit devices. Retrofit devices may be intended for long- or

short-term installation on HVAC units. FDD technologies may report their findings

through a means such as a display or automated/remote means.

FDD technologies have enormous potential to enhance the future of energy

efficiency. FDD can provide the information necessary to accurately and reliably

understand HVAC equipment performance, and improve HVAC maintenance through

preventative strategies. Ideally, FDD technologies would be implemented in an

automated fashion, outfitted for long-term use with a means for providing remote

connectivity. This would enable these technologies to actively inform building

operators, homeowners or service contractors and solicit corrective actions before

faults become severe, or before critical failures occur.

It is important to make a distinction between faults and failures. An HVAC unit may

still operate under a fault condition, albeit with reduced efficiency and/or

performance. Conversely, a failure mode is one that prohibits an HVAC unit from

operating at all. It is anticipated that most benefits of FDD are realized through

remediation of fault modes rather than failure modes. Failure modes are typically

reacted to and resolved regardless of the presence of FDD technologies.

0.7 0.7

1.0

1.6

0.9

0.4

0

0.2

0.4

0.6

0.8

1

1.2

1.4

1.6

1.8

( < 2 Years ) ( 2 - 4 Years ) ( 5 - 9 Years ) ( 10 - 19 Years ) ( ≥ 20 Years ) ( Don't Know )

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ANTICIPATED BARRIERS TO ADOPTION OF FDD

Cost Effectiveness – Cost effectiveness is dependent upon the difference between the

cost of the FDD technology, and the realized HVAC operating cost reductions.

Realizing operating cost reductions is not as straightforward with FDD as it is with

other “widget-based” technologies. Savings are dependent on:

1. Which faults occur in the HVAC system

2. Which faults are detected and diagnosed

3. Which faults are actively corrected

4. The financial impacts unique to the HVAC owner and application

Product Availability and Performance - The range of commercially available onboard

FDD products is still very limited and only a handful of in-field tools are available.

Currently, industry lacks standardized methodologies for both evaluating FDD

products and for simulating common maintenance faults. As a result, there is a

limited understanding regarding how well fault detection and diagnostics devices

perform. In addition, without a standardized method, it is challenging to make

comparisons between existing studies exploring the impacts of common HVAC faults.

As a result, the impacts of HVAC faults are not well understood, especially in

scenarios that consist of multiple simultaneous faults.

FDD and the “Human Element” – One potential benefit of FDD technologies is the

removal of uncertainties regarding varying human interpretation/diagnostics.

However, one must consider that there potentially may not be suitable technological

replacements for the creative/critical thinking abilities inherent with manual analysis

of complex problems. The level of human involvement appropriate for HVAC FDD in a

given application remains to be explored through continuing evaluations of FDD

technologies and the impacts of common faults.




OBJECTIVE The objective of this project is to develop a laboratory test method for FDD technologies.

The test method is to detail procedures to generate faults, and evaluate the response of

FDD to those faults. This report presents the final updated test methodology used in

HT.11.SCE.005, and examines the specific issues and lessons learned in the laboratory

assessment.

This test method is developed with the intention of informing SCE’s Energy Efficiency

Programs, as well as other developing FDD-related efforts such as Codes and Standards

Enhancement (CASE) studies for the California Code of Regulations, or the American Society

of Heating, Refrigerating and Air Conditioning Engineers (ASHRAE).




TEST METHOD DEVELOPMENT STRATEGY Consistency with current applicable HVAC testing methodologies is important to the industry

acceptance and success of an FDD test method. For this reason, the intention is to leverage

as much existing knowledge as possible. The Air Conditioning, Heating and Refrigeration

Institute (AHRI) establishes standards for HVAC equipment testing. The AHRI is widely

recognized and represents more than 300 heating, water heating, ventilation, air

conditioning, and commercial refrigeration manufacturers within the global HVAC industry.

AHRI 210/240-20081 and its incumbent referenced standards (such as ASHRAE Standard

37) were chosen as a basis for the FDD test method to build upon. The FDD test method is

used for FDD technologies suitable for unitary air-conditioners and air-source unitary heat

pumps with nominal capacity under 65,000 British thermal units per hour (Btu/h). The FDD

test method will leverage steady state wet-coil cooling mode testing, analogous to tests

outlined in AHRI 210/240.

In addition, a previous evaluation of FDD and HVAC faults was conducted on a packaged

RTU at SCE’s Technology Test Centers (TTC). This evaluation fed into a Public Interest

Energy Research (PIER) project2 as well as HVAC maintenance projects3 conducted under

SCE’s Education, Training, and Outreach (ETO) program. The procedures used for that

evaluation directly fed into the development of this test method. The resultant draft test

method was screened both through various subject matter experts (SMEs) at SCE’s TTC, as

well as through a TAG, composed of various key industry members.




THE HVAC TEST UNIT The HVAC test unit is a 3-ton (nominal) residential split system air conditioner,

manufactured by Trane. This air conditioner setup consists of one indoor unit (cooling coil:

“4TXC B042BC3HCAA,” and furnace “TUD1B080A9361A”), paired to an outdoor condensing

unit (XR80, 4TTB3 036D1000AA). It was not necessary to use the furnace contained within

the indoor unit, as the laboratory assessment focused on cooling mode operation. The test

unit is a fixed capacity setup (fixed-speed fans and compressor) that uses R-410a

refrigerant and features a thermostatic expansion valve (TxV). Figure 2 shows both the

indoor and outdoor units.

FIGURE 2. HVAC TEST UNIT - INDOOR UNIT (LEFT), OUTDOOR CONDENSING UNIT (RIGHT)

Various residential HVAC units exist in the field, comprising a number of different possible

physical configurations. This unit is just one possible configuration. It represents a standard

efficiency unit, relevant to the current generation of products that will be aging. Other

options to explore may include (but are not limited to) those that feature R-22 refrigerant,

fixed orifice expansion devices, or higher efficiency units (larger heat exchangers, more

efficient compressors, fans, etc.). Ultimately, field studies are needed to best characterize

the various equipment types, and inform about what is most prevalent in the field.




THE FDD TECHNOLOGY The FDD technology tested is a package of items, intended for use as a service technician’s

tool. A significant amount of HVAC maintenance-related information is also available

through reference literature and training provided by the FDD manufacturer. The package

includes:

- A personal digital assistant (PDA) mobile device

- (2) Air-side probes (supply air and return air) with each measuring both dry-bulb

(DB) and wet-bulb (WB) temperatures

- (1) Air-side sensor that measures DB temperature (condenser inlet air)

- (2) Clamp-on thermocouple (T/C) sensors (suction and liquid line refrigerant

temperatures)

- (3) Refrigerant pressure hoses (high and low side system pressures, and for

general charging/recovery/evacuation purposes)

- (1) Digital refrigerant manifold

For the purposes of this evaluation, this unit’s hoses and manifold were not used for

charging/recovery/evacuation. Figure 3 depicts a diagram of the FDD technology.

Manifold PDA

SA Sensor

RA Sensor

Ambient

Clamp-On T/C’sRefrigerant Pressure Hoses

FIGURE 3. THE FDD TECHNOLOGY

The PDA displays several screens of measurements and calculations. The tool steps through

its internal algorithms and displays its diagnosis in real-time fashion. The tool has

approximately 50 different diagnostic messages (not all were encountered during testing).

Measurements, calculations, FDD messages were observed to be simultaneously populated

about once every three seconds or so. This tool has no/limited logging capability: it is able

to log one set of readings, which may be uploaded to an online server for reporting. The

technology was provided as new, as calibrated from the manufacturer.




The following faults were applicable to the FDD technology:

- Low Refrigerant Charge

- High Refrigerant Charge

- Refrigerant Liquid Line Restrictions

- Refrigerant Non-condensables

- Evaporator Airflow Reduction

- Condenser Airflow Reduction

The PDA displays the following 19 measurements and calculations:

1. Suction pressure, pounds per square inch-gauge-pressure (psig)

2. Liquid pressure, psig

3. Suction temperature, degrees Fahrenheit (°F)

4. Liquid temperature, °F

5. Ambient air temperature, °F

6. Return air DB temperature, °F

7. Return air WB temperature, °F

8. Supply air DB temperature, °F

9. Supply air WB temperature, °F

10. Evaporator temperature, °F

11. Superheat, °F

12. Condensing temperature over ambient, °F

13. Sub-cooling, °F

14. Indoor temperature drop, °F

15. Efficiency Index

16. Capacity Index

17. Power, kilowatts (kW)

18. Runtime, hours

19. Dollar ($) Savings

Measurement and calculation items 1 through 14 were used for testing. For items 10 – 14,

marked in bold, the tool additionally has pre-established ranges to detect whether the

reported parameter is considered “Low”, “Ok (-)”, “Ok”, “Ok (+)” or “High”.




TEST EQUIPMENT INSTALLATION, INSTRUMENTATION,

AND DATA ACQUISITION The ducted HVAC system was installed with guidance from manufacturer-provided

literature, and the specifications of AHRI 210/240-2008 and its incumbent referenced

standards (ASHRAE 37-2009 – “Methods of Testing for Rating Electrically Driven Unitary Air-

Conditioning and Heat Pump Equipment”, ASHRAE Standard 41.2-1987 – “Standard

Methods for Laboratory Airflow Measurement”, etc.), with the exception of the gas hook-up

for heating. Forty-six feet of liquid line was needed for the test setup. A refrigerant mass

flow meter was installed on the liquid line. A ball valve was installed on the liquid line, after

the mass flow meter, for liquid line restriction testing. Sight-glasses were also installed to

identify the presence of mixed-phase refrigerant flow in the liquid line. Figure 4 provides a

comprehensive diagram of the HVAC unit setup and associated TTC instrumentation. In

addition, Figure 5 and Figure 6 are presented to highlight the key measured state points on

the refrigerant-side and the air-side.

FDD manufacturer specifications and standard practice dictated the installation of the FDD

technology and placement of its corresponding sensors. The FDD installation was

documented and verified by the FDD manufacturer. Table 1, Figure 7, and Figure 8 detail

the placement of all FDD sensors. In addition, Table 1 details the laboratory instrumentation

used by the TTC, along with the corresponding ASHRAE 37 measurement designations. In

this manner, key measurements and similarly located sensors are tracked and presented.

The FDD technology did not have the capability to log its readings. Accordingly, “spot

measurements” of FDD outputs were recorded manually onto a spreadsheet, once per

minute, for 10 entries per test scenario. For data logging of TTC instrumentation

measurements, the National Instruments’ SCXI data acquisition system was used. The data

acquisition system was set up to scan and log 95 data channels in one-minute intervals. As

part of the TTC’s quality control protocol, the data acquisition system was designed to be

completely independent of the supervisory control computer. This approach was taken to

ensure the data collection was not compromised by the control sequence’s priority over data

acquisition.

Data collected from TTC’s instrumentation were screened continuously to ensure key control

parameters were maintained within acceptable ranges. In the event that any of the control

parameters fell outside acceptable limits, the problem was flagged and a series of diagnostic

investigations occurred. Corrections were then made and tests were repeated, as necessary.

After the data passed the initial screening process, they were imported to a customized

refrigeration analysis model where detailed calculations were performed.

Table 2 provides the accuracy and calibration dates for all the sensors used in the lab

assessment. All instruments were calibrated before testing was conducted. Careful attention

was paid to the design of the monitoring system, with the objective of minimizing

instrument error and maintaining a high level of repeatability and accuracy in the data.




Outdoor Section

Indoor Section

Return Duct

Evaporator Coil

Separation Wall

Mass Flow Meter

TxV

Compressor

Suction Line

Liquid Line

Discharge Line

Condenser Coil

Furn

ace

Evaporator Fan

Discharge Temp

Discharge Pressure

Suction Pressure

Liquid Line Temp

Liquid Line Pressure

Temp before & after Mass Flow Meter

Pressure before & after Mass Flow Meter

Temp before TxV

Pressure before TxV

Evap. Outlet Temp

Evap. Outlet Pressure

Supply Duct

Air Temp Grid

Inlet Air Temp Grid (4 sensors) on Each of the Four Sides of the Condensing Unit

Indoor Air:1. DB Temperature2. WB Temperature3. Relative Humidity

Entering Outdoor Air:1. DB Temperature2. Relative Humidity

Temp 1

Temp 2

Temp 3

Temp 4

Temp 6

Temp 5VoltsAmpsPower

Condenser Fan: VoltsAmpsPower

VoltsAmpsPower

Condensing Unit Total Power

T1 T2

T3 T4

Suction Temp

Scale: None

DB & RH

Indoor Test Chamber Outdoor Test Chamber

Drain Piping

Scale for Condensate Mass

Air Temp Grid

T1 T2 T3

T4 T5 T6

Air Temp Grid

T1 T2 T3

T4 T5 T6

DP

ASHRAE Airflow Measurement Apparatus

Volumetric Airflow Rate

(cfm)

Sight Glass

Duct Inlet Fan

WB

DP

T1 T2

T3 T4

Line Restriction

Valve

FIGURE 4. TEST SETUP - COMPREHENSIVE DIAGRAM




TxV

Refrigerant Mass Flow

Meter

Liquid Line Restriction

Valve

Compressor

R5R6

R7

Evaporator

R1

R2

R3

R4

(R8)

Condenser

Condensing Unit

FIGURE 5. REFRIGERANT-SIDE STATE POINTS

The following measurements are available at each refrigerant-side state point:

R1, Evaporator Outlet – Pressure, Temperature

R2, Condensing Unit Inlet – Pressure, Temperature

R3, Compressor Inlet – Temperature

R4, Compressor Outlet – Pressure, Temperature

R5, Condenser Outlet – Pressure, Temperature

R6, Mass Flow Meter Inlet – Pressure, Temperature

R7, TxV Inlet – Pressure, Temperature

Enthalpies are calculated at R1, R3, R4, R5, and R7. No measurements exist at state point

R8, but this state point is assumed to have the same enthalpy as state point R7. Refrigerant

mass flow is measured near state point R6.




Evaporator Coil

Furn

ace

Evaporator FanSupply

Duct

Scale: None

ASHRAE Airflow Measurement Apparatus

Duct Inlet Fan

A1

A2A3

A4

Return Duct

FIGURE 6. AIR-SIDE STATE POINTS

The following measurements are available at each air-side state point:

A1, Evaporator Fan Inlet – DB Temps (1-6), WB/relative humidity (RH)

A2, Evaporator Coil Inlet – DB Temps (1-4), dew point (DP)

A3, Evaporator Coil Outlet – DB Temps (1-6), DP

A4, Supply Duct – DB and RH

Enthalpies are calculated at all air-side state points for calculation and redundancy

purposes. Air volumetric flow is measured using the ASHRAE Airflow Measurement

Apparatus, located upstream of the indoor unit. Airflow is measured in units of Standard

Cubic Feet per Minute (SCFM). In addition, condensate from the evaporator is plumbed to a

separate tank outside of the test chamber. A scale continuously weighs this tank, and the

data is logged to the data acquisition system.




FIGURE 7. INDOOR HVAC UNIT SENSOR PLACEMENT




FIGURE 8. CONDENSING UNIT SENSOR PLACEMENT




TABLE 1. LIST OF MEASUREMENT POINTS

ASHRAE 37-2009 Laboratory Measurements FDD Sensors

Measurement Units Name/Description Notes Name/Description Notes

Barometric Pressure in Hg Barometric Pressure -

- -

Time 0:00:00 Date, Time -

Total Power W

- Total Power is Calculated

HD Cond Unit Watts -

HD Condenser Fan Watts -

HD Compressor Watts -

Indoor Unit Power W HD Evap Unit Watts -

Voltages V HD Evap Unit Volts -

HD Cond Unit Volts -

Frequencies Hz HD Evap Unit Frequency -

HD Cond Unit Frequency -

External Resistance to Airflow in H2O HD Static Press Evap Fan -

HD Static Press Evap Coil -

Fan Speed, Setting rpm Evap Fan RPM -

DB - Air Entering Indoor Unit °F

HD IDF Air T Ent Mid-T (State point A1) Mid - Middle

Lt - Left Rt - Right T - Top

B – Bottom

(1) RA Probe 100789 Measures DB & WB, Located at evap fan inlet,

near "HD Evap Fan Inlet Tw/Rh" sensor (State point A2)

HD IDF Air T Ent Lt-B

HD IDF Air T Ent Mid-B

HD IDF Air T Ent Rt-T

HD IDF Air T Ent Lt-T

HD IDF Air T Ent Rt-B

HD Evap Air T Ent Rt-B (State point A2) Lt - Left

Rt - Right T - Top

B – Bottom

HD Evap Air T Ent Lt-T

HD Evap Air T Ent Lt-B

HD Evap Air T Ent Rt-T

WB - Air Entering Indoor Unit °F

HD Evap Fan Inlet Tw <- Single sensor which measures air moisture content, calculates WB & RH

(State point A1) HD Evap Fan Inlet Rh

HD Evap Coil In Dewpoint (State point A2)

DB - Air Leaving Indoor Unit °F

HD Evap Air T Exit Rt-B

(State point A3)

(1) SA Probe 100780 Measures DB & WB, Located near "HD EvapCoil

Out Temp/RH" sensor (State point A3)

HD Evap Air T Exit Mid-B

HD Evap Air T Exit Lt-B

HD Evap Air T Exit Rt-T

HD Evap Air T Exit Mid-T

HD Evap Air T Exit Lt-T

HD Evap Coil Out Temp <- Two sensors, same location (State point A4)

WB - Air Leaving Indoor Unit °F HD Evap Coil Out Rh

HD Evap Coil Out Dewpoint (State point A3)

DB - Air Entering Outdoor Unit °F

HD Cond Air In N Lt-T

N - North Face S - South Face E - East Face

W - West Face Lt - Left

Rt - Right T - Top

B - Bottom

(1) Temp Probe Located at North Face of condensing unit, near sensor "5-3-0", this sensor is closest to the calc

avg of all sensors

HD Cond Air In N Lt-B

HD Cond Air In N Rt-T

HD Cond Air In N Rt-B

HD Cond Air In S Lt-T

HD Cond Air In S Lt-B

HD Cond Air In S Rt-T

HD Cond Air In S Rt-B

HD Cond Air In E Lt-T

HD Cond Air In E Lt-B

HD Cond Air In E Rt-T

HD Cond Air In E Rt-B

HD Cond Air In W Lt-T

HD Cond Air In W Lt-B

HD Cond Air In W Rt-T

HD Cond Air In W Rt-B






WB - Air Entering Outdoor Unit °F - Not used, air cooled condensing unit - -

DB - Air Leaving Outdoor Unit °F

HD Cond Air OUT Temp 1

Use Calc Avg - -








WB - Air Leaving Outdoor Unit °F - Not used, air cooled condensing unit - -

Velocity pressure at nozzle throat or static pressure diff across nozzle

in H2O

Current_10 Static press diff across nozzle, note:

LabView channel name was not re-named - -

Current_11 Static press diff across nozzle, note:

LabView channel name was not re-named - -

Temp at nozzle throat °F ASHRAE Box Air Temp 1 - - -

ASHRAE Box Air Temp 2 -

Press at nozzle throat in Hg

Voltage 6 Press at nozzle throat

note: LabView channel name was not re-named

- -

Voltage 7 - -

Voltage 8 - -

Voltage 9 - -

Condensing Pressure or Temperature psig or °F HD Cond exit Ref Psig (State point R5) (1) Refrigerant Pressure Hoses &

Elec Refg Manifold

Installed on extra Liquid Line Service Port at Condensing Unit. Extra service port installed to

allow TTC press transducer (State point R5)

Evaporator Pressure or Temperature psig or °F HD Evap Ref Exit Psig (State point R1) (1) Refrigerant Pressure Hoses &

Elec Refg Manifold

Installed on extra Compressor Suction Service Port at Condensing Unit. Extra service port

installed to allow TTC press transducer (State point R1)

Refrigerant Vapor Temperature Entering Compressor (10 in from shell)

°F HD Comp Suction Temp (State point R3) - -

Refrigerant Vapor Temperature Leaving Compressor (10 in from shell)

°F HD Comp Discharge Temp (State point R4) - -

Refrigerant Oil Flow Rate lbs./hr - - - -

Refrigerant Volume in Refg-Oil Mix ft^3/ft^3 - - - -

Condensate collection lbs./hr Scale in Pounds calculated per min - -

Refrigerant Liquid Temperature, Indoor Side °F HD Ref T Ent TXV (State point R7) - -

Refrigerant Liquid Temperature, Outdoor Side °F HD Cond exit Ref Temp (State point R5) (1) Clamp-On Thermocouple

100792 Near Liquid Line Service Port at Condensing Unit

(State point R5)

Refrigerant Vapor Temperature, Indoor Side °F HD Evap Ref Exit Temp (State point R1) - -

Refrigerant Vapor Temperature, Outdoor Side °F HD Cond Unit Suct Temp (State point R2) (1) Clamp-On Thermocouple

100791

Near Compressor Suction Service Port at Condensing Unit (State point R2)

Refrigerant Vapor Pressure, Indoor Side psig HD Cond exit Ref Psig (State point R5)

*Used to calculate evaporator temp - -

Other sensors, not listed in ASHRAE table

- HD Comp Discharge Psig (State point R4) - -

- HD Comp case Temp - - -

- HD Cond Unit Suct Psig (State point R2) - -

- HD Ref Psig Ent TXV (State point R7) - -

- HD Ref Psig Ent MFM (State point R6) - -

- HD Ref T Ent MFM (State point R6) - -






- Refrigerant MFM lbs. per min (State point R6) - -

- ASHRAE Box Rh Vaisala - - -

- ASHRAE Box Temp Vaisala - - -

- Rm 4 on Pole at Ceiling

Other TTC Sensors, not listed in ASHRAE table

- -

- Rm 4 on Pole at 9ft - -




- Rm4 Pole at 5ft - -





- Rm 4 on Pole at Floor - -




TABLE 2. ACCURACY, CALIBRATION DATES AND LOCATIONS, AND CORRESPONDING KEY MONITORING POINTS FOR SENSORS

USED

Sensor Type Make/Model

Accuracy

(NIST Traceable) Calibration Date

(location) Monitoring Points

Description

Temperature (type-T thermocouples)

Masy Systems,

Ultra-Premium Probe

± 0.18°C [at 0°C] (± 0.32°F)

5-4-2011

(In-house)

Evap fan inlet DB

Evap coil inlet DB

Evap coil outlet DB

Outdoor chamber DB

Cond inlet DB

Cond outlet DB

All refrigerant

temps

Relative Humidity (RH)

Vaisala, HMP 233

± 1% (0-90% RH)

± 2% (90-100% RH)

5-5-2011

(SCE’s Metrology Lab)

Evap outlet

Wet Bulb Vaisala, HMP 247 ± 0.013% of

reading

5-9-2011


Evap fan inlet

Relative Humidity (RH)

Vaisala, HMP 247 ± (0.5 + 2.5% of

reading)% RH

5-9-2011


Supply duct

Dew Point Edgetech, Dew

Prime DF Dew Point Hygrometer

± 0.2°C (± 0.36°F)

5-5-2011

(SCE’s Metrology

Lab)

Evap inlet

Evap outlet

Pressure

(0-1000 psig) Setra, C207

± 0.13% of full scale

4-14-2011

(In-house)

Comp discharge

Cond outlet

MFM inlet

TXV inlet

Pressure

(0-500 psig) Setra, C207

± 0.13% of full

scale

4-14-2011

(In-house)

Comp suction

Evap outlet

Pressure (0-10 inches of water, in-

wg)

Ashcroft, AQS-28304


4-14-2011

(Tektronix Calibration Lab)

Across indoor unit

Power Ohio Semitronics,

GW5-002C

± 0.2% of reading


(cond: 1,000W FS) (comp: 5,000W FS)

5-11-2011

(In-house)

Condensing unit

Compressor

Condenser fan

Power HIOKI 3169-21 ± 0.5% of reading 5-10-11

(In-house)

Indoor unit

Evap fan

Refrigerant Mass Flow Meter

Endress-Hauser, (Coriolis meter)

80F08-AFTSAAACB4AA

For liquids, ± 0.15% of reading

For gases, ± 0.35% of reading

7-22-2010

(Homer R. Dulin Co.)

Refrigerant flow rate

Scale HP-30K ± 0.1 gram

(± 0.0035 ounces)

11-29-2010

(In-house) Mass of condensate




THE TEST METHOD Updates to the test method were continually incorporated, as deemed necessary through

laboratory testing experience. In this section, the test method is presented, along with

various discussions on specific key lessons learned by conducting the laboratory

assessment.

CONTROL PARAMETERS AND TEST INTERVALS All testing is conducted similarly to the steady-state wet coil tests outlined in

AHRI – 210/240-2008. All test scenarios encompass a 1-hour span of data. This hour comprises a 30-minute pre-test interval, followed by a 30-minute data collection interval. Reported parameters are straight averages across

the 30-minute data collection interval. Table 3 lists the targeted control parameters used for testing.

TABLE 3. CONTROL PARAMETERS

Control Parameter

Test Operating Tolerance

Test Condition Tolerance Target Units

Outdoor Test Chamber

DB: Cond inlet DB 2.0 0.5

95, 75, or 115

(85 for select tests) °F

Indoor Test Chamber DB: Evaporator fan inlet

DB 2.0 0.5 80, 75, or 70 °F

Evaporator outlet DB 2.0 N/A N/A °F

Indoor Test Chamber WB: Evaporator fan inlet

WB 1.0 0.3

67, 63, or 59 (63.3, 67.9, or 72.4 for

select tests) °F

Evaporator outlet DP (calc’d equivalent)

~2.8 N/A N/A °F

Supply duct RH (calc’d equivalent)

~8 N/A N/A %

Electrical Voltage 2.0 1.5 208, 115

% of reading

(4, 2) (3, 1.7) (208, 115) V

Nozzle Press Drop 2.0 N/A N/A % of

reading




CALCULATIONS Various calculation methods are available for laboratory testing. Table 4 lists the

calculation methods used in this project.

TABLE 4. CALCULATION METHODS

# Calculation Methods Calculated Parameters

1 Refrigerant-side measurements and calculations

Gross cooling capacity, heat rejection

2 Refrigerant-side measurements and calculations -> compressor regression

Gross cooling capacity, refrigerant mass flow, compressor power

3 Air-side measurements and

calculations

Gross cooling capacity, sensible cooling

capacity, latent cooling capacity

4

Evaporator airflow equation: manufacturer literature of evaporator pressure drop vs. airflow

Evaporator air mass flow rate

5 Evaporator condensate scale Latent cooling capacity

A comprehensive summary of calculation methods applicable to a given test scenario

may be found in the appendix, in Table 20.

Energy Efficiency Ratio (EER) calculations are performed as follows:

EQUATION 1. ENERGY EFFICIENCY RATIO

Or

Where:

= Energy Efficiency Ratio (refrigerant-side-gross-

cooling-based), Btu/hr/Watt (W)

= Energy efficiency ratio (air-side-gross-cooling-based),

Btu/hr/W

= Refrigerant-side gross cooling capacity, Btu/hr

= Air-side gross cooling capacity, Btu/hr

= Total power (compressor + fans + misc.), W




Refrigerant-side calculations for gross cooling capacity and heat rejection are

performed as follows:

EQUATION 2. REFRIGERANT-SIDE GROSS COOLING CAPACITY

Where

= Refrigerant-side gross cooling capacity, Btu/hr

= Refrigerant mass flow rate, lbs /hr

= Enthalpy at refrigerant-side state point R1, Btu/lb


EQUATION 3. REFRIGERANT-SIDE CONDENSER HEAT REJECTION

Where:

= Refrigerant-side heat rejection, Btu/hr

= Refrigerant mass flow rate, lbs/hr



In addition, the HVAC unit’s manufacturer provided the compressor regression

curves. The regressions were used as a “sanity check” for the baseline scenarios, to

establish confidence in the test results. Given a saturated condensing temperature

and a saturated evaporator temperature, these curves are able to output cooling

capacity, refrigerant mass flow rate, and compressor power. Heat rejection is not a

direct output, but is estimated using cooling capacity and power (approximate heat

of compression) outputs.




EQUATION 4. REFRIGERANT-REGRESSION CONDENSER HEAT REJECTION

Where

= Refrigerant-side heat rejection (regression-based), Btu/hr

= Refrigerant-side gross cooling capacity (regression), Btu/hr

= Compressor power (regression output), W

= Conversion factor = 3.41214163, Btu/hr/W

Percent variation is defined as the variation between two values, divided by the

average of the data set. This data set may comprise the two values, or it may

comprise several other values. For the purposes of this project, it is used when:

a. Comparing different methods of calculations of a certain parameter

b. Comparing values of a certain parameter from several repeat tests

Percent variation is given by the following equation:

EQUATION 5. CALCULATING PERCENT VARIATION

Percent difference is defined as the relative shift in a parameter, or the difference of

two values divided by one original value. Percent difference is used when comparing

a parameter from one fault test scenario, to its baseline scenario (shift in a

parameter due to a fault). The following equation provides the percent difference.

EQUATION 6. CALCULATING PERCENT DIFFERENCE

Compressor regression outputs and test measurements/calculations, along with the

associated percent variations are presented in Table 5, Table 6, Table 7, and Table 8.

Percent variations for gross cooling capacity ranged from -2% to -7%. Percent

variations for heat rejection ranged from -2% to -8%. Percent variations for

compressor power ranged from 7% to 11%. Percent variations for refrigerant mass

flow ranged from 0% to 4%.




TABLE 5. BASELINE GROSS COOLING CAPACITY: REFRIGERANT ENTHALPY METHOD VS. COMPRESSOR

REGRESSIONS

Test #

% Variation

Gross Cooling Capacity:

Refrigerant Enthalpy Method (Btu/hr)

Gross Cooling Capacity:

Compressor Regression (Btu/hr)

1 -7% 30,487 32,558

2 -7% 35,187 37,749

3 -8% 37,757 40,721

4 -4% 28,549 29,804

5 -4% 32,947 34,352

6 -4% 36,664 38,132

7 -2% 26,845 27,408

8 -2% 30,863 31,362

9 -2% 34,929 35,607

TABLE 6. BASELINE HEAT REJECTION: REFRIGERANT ENTHALPY METHOD VS. COMPRESSOR REGRESSIONS

Test # % Variation Heat Rejection:

Refrigerant Enthalpy Method (Btu/hr)

Heat Rejection: Compressor Regression

(Btu/hr)

1 -8% 39,450 42,569

2 -7% 42,966 45,994

3 -6% 44,701 47,673

4 -6% 37,252 39,558

5 -5% 40,542 42,510

6 -3% 43,146 44,654

7 -5% 35,221 36,898

8 -3% 38,236 39,385

9 -2% 41,358 42,106




TABLE 7. BASELINE COMPRESSOR POWER: MEASURED VS. COMPRESSOR REGRESSIONS

Test # % Variation Compressor Power:

Measured (W)

Compressor Power: Compressor Regression

(W)

1 8% 3,186 2,934

2 9% 2,649 2,416

3 11% 2,270 2,037

4 8% 3,089 2,859

5 9% 2,612 2,391

6 10% 2,118 1,912

7 7% 2,995 2,781

8 8% 2,546 2,351

9 10% 2,095 1,905

TABLE 8. BASELINE REFRIGERANT MASS FLOW: REFRIGERANT ENTHALPY METHOD VS. COMPRESSOR REGRESSIONS

Test # %

Variation Refrigerant Mass Flow:

Measured (lbs/min)

Refrigerant Mass Flow: Compressor Regression

(lbs/min)

1 0% 8.0 8.0

2 0% 8.2 8.2

3 0% 8.2 8.2

4 1% 7.4 7.3

5 1% 7.6 7.5

6 2% 7.7 7.5

7 2% 6.9 6.7

8 3% 7.1 6.9

9 4% 7.4 7.0

It is important to note that refrigerant-side calculation issues exist for tests featuring

non-condensables or mixed-phase refrigerant flow. Mixed phase refrigerant flow

occurred in the liquid line for all low charge tests (Tests 10-14, 47-55), and for all

liquid line restriction tests (Tests 25-27). With mixed phase liquid line refrigerant

flow, refrigerant properties look-ups become inaccurate at state points R5 through

R8. Without knowing refrigerant quality of the mixed-phase flow (percent vapor

composition of total mass), state point properties such as enthalpy cannot be

determined. In addition, refrigerant mass flow meter measurements are

compromised with mixed phase flow. As a result, refrigerant enthalpy method

calculations become questionable with the presence of mixed phase refrigerant flow.




In addition, with mixed-phase liquid line refrigerant flow, while the regression model

may still be suitable for predicting refrigerant mass flow and compressor power, any

gross cooling capacity outputs are suspect. The model cannot account for any effects

at the heat exchangers, and it is likely that heat transfer is compromised with mixed

phase flow. Forced convection heat transfer coefficients for gases typically range from approximately 4.4 to 44 Btu/hr-ft2-°R; forced convection heat transfer

coefficients for liquids typically range from approximately 8.8 to 3,500 Btu/hr-ft2-°R.iv

Furthermore, with a mixture of refrigerant and non-condensables (Tests 28-32),

refrigerant mass flow measurements are compromised, and refrigerant properties

look-ups for all refrigerant-side state points are no longer applicable. The regression

model also becomes inaccurate: the relationships between system pressures and

properties changes when pure R-410a is not present. For all tests featuring non-

condensables or mixed phase refrigerant flow, refrigerant-side based calculations for

gross cooling capacity (enthalpy method or regression method) are not used. The

air-enthalpy method must be relied upon in these cases.

Air-side calculations are performed as follows:

EQUATION 7. AIR-SIDE GROSS COOLING CAPACITY

or

Where

= Air-side gross cooling capacity, Btu/hr

= Air flow rate (measured or calculated), lbs/hr

= Enthalpy at air-side state point A2, Btu/lb

= Enthalpy at air-side state point A3, Btu/lb

= Air-side gross sensible cooling capacity, Btu/hr

= Air-side gross latent cooling capacity, Btu/hr




EQUATION 8. AIR-SIDE GROSS SENSIBLE COOLING CAPACITY

Where

= Air-side gross sensible cooling capacity, Btu/hr


= Specific heat of air (constant pressure), Btu/lb.-°F

= DB temperature at air-side state point A2, °F

= DB temperature at air-side state point A3, °F

EQUATION 9. AIR-SIDE GROSS LATENT COOLING CAPACITY

Or

Where

= Air-side gross latent cooling capacity, Btu/hr


= Humidity ratio at air-side state point A2, lbmoisture/lbdry air

= Humidity ratio at air-side state point A3, lbmoisture/lbdry air

= Heat of vaporization water at one atmosphere (atm), Btu/lb

= Condensate water flow rate (measured with scale), lb/hr

EQUATION 10. AIR-SIDE AIR FLOW RATE

Or

Where:


= Volumetric air flow rate, air at “standard” conditions, ft3/min

= Density of air at “standard” conditions = 0.075, lbs/ft3

= Volumetric airflow rate (equation output), ft3/min

= Conversion factor = 60, min/hr




Evaporator airflow was directly measured using calibrated instrumentation. In

addition, pressure drop measurements were measured, and used as a redundant

means to verify measured airflow rates. A second order polynomial was fitted to

eight manufacturer-published data points for evaporator air volumetric flow rate in

Standard Cubic Feet per Minute (SCFM) vs. evaporator pressure drop in psig. Data

points ranged from 0.05 in water (H2O) to 0.4 in H2O. After SCFM was calculated, it

was compared to SCFM measurements. Table 9 details the SCFM’s from both

methods and the percent variations. Percent variations ranged from -1% to 2%.

TABLE 9. COMPARING EVAPORATOR AIR VOLUMETRIC FLOW RATE: MEASURED SCFM VS. SCFM/PRESSURE DROP

EQUATION METHOD

Test # % Variation

Evaporator Air Volumetric Flow Rate

Measurements

(SCFM)

Evaporator Air Volumetric Flow Rate: SCFM/Pressure

Drop Equation Method

(SCFM)

1 5% 1,155 1,211

2 5% 1,147 1,208

3 5% 1,160 1,214

4 4% 1,165 1,208

5 3% 1,171 1,211

6 3% 1,177 1,213

7 3% 1,177 1,213

8 3% 1,184 1,215

9 2% 1,190 1,216

Table 10 presents the gross cooling capacity calculations from the refrigerant

enthalpy and air enthalpy methods, including percent variations. Percent variations

range from 3% to 11%.




TABLE 10. BASELINE GROSS COOLING CAPACITIES: REFRIGERANT ENTHALPY METHOD VS. AIR ENTHALPY METHOD

Test # % Variation

Gross Cooling Capacity: Refrigerant

Enthalpy Method Gross Cooling Capacity:

Air Enthalpy Method

1 4% 30,487 29,280

2 5% 35,187 33,319

3 10% 37,757 34,270

4 7% 28,549 26,613

5 7% 32,947 30,795

6 11% 36,664 32,826

7 3% 26,845 26,031

8 5% 30,863 29,349

9 5% 34,929 33,104

Figure 9 illustrates the air-side gross, sensible, and psychrometric-based latent

cooling capacity calculations for the baseline Tests 1-9. Percent variations are

compared between gross cooling capacity, and the sum of sensible and latent cooling

capacities. Percent variations were 2% across all tested scenarios.

Figure 10 illustrates the air-side gross, sensible, and condensate-scale-based latent

cooling capacity calculations for the baseline tests 1-9. Percent variations are

compared between gross cooling capacity, and the sum of sensible and latent cooling

capacities. Percent variations range from -2% to 5%.




FIGURE 9. COMPARING BASELINE AIR-SIDE COOLING CAPACITY CALCULATIONS – GROSS, SENSIBLE, AND

PSYCHROMETRIC-BASED LATENT

FIGURE 10. COMPARING BASELINE AIR-SIDE COOLING CAPACITY CALCULATIONS – GROSS, SENSIBLE, AND SCALE-BASED LATENT

29

,28

0

33

,31

9

34

,27

0

26

,61

3

30

,79

5

32

,82

6

26

,03

1

29

,34

9

33

,10

4

22

,56

2

24

,90

1

25

,36

1

21

,04

8

23

,76

0

24

,78

5

21

,08

2

23

,24

9

25

,88

4

6,0

41

7,6

33

8,1

05

5,0

21

6,4

02

7,3

59

4,4

90

5,5

75

6,6

27

0

5,000

10,000

15,000

20,000

25,000

30,000

35,000

40,000

0

5,000

10,000

15,000

20,000

25,000

30,000

35,000

40,000

1 2 3 4 5 6 7 8 9

Co

olin

g C

apac

ity

(Btu

/h)

Test Scenarios

Gross Sensible Latent - Psychr

29

,28

0

33

,31

9

34

,27

0

26

,61

3

30

,79

5

32

,82

6

26

,03

1

29

,34

9

33

,10

4

22

,56

2

24

,90

1

25

,36

1

21

,04

8

23

,76

0

24

,78

5

21

,08

2

23

,24

9

25

,88

4

5,5

51

8,1

61

9,3

85

5,0

94

7,0

21

8,5

42

3,6

59

5,6

38

7,2

86

0

5,000

10,000

15,000

20,000

25,000

30,000

35,000

40,000

0

5,000

10,000

15,000

20,000

25,000

30,000

35,000

40,000

1 2 3 4 5 6 7 8 9

Co

olin

g C

apac

ity

(Btu

/h)

Test Scenarios

Gross Sensible Latent - Scale

4%

1% -1%

2%

0%

-2%

5%

2%

0%

2% Variation for all scenarios:

(Gross) vs. (Sensible + Latent)




Airflow measurements are not done across the condenser, so a direct air-enthalpy

calculation cannot be used for air-side-based heat rejection calculations. Instead,

heat rejection is calculated through the sum of the air-side calculation for gross

cooling capacity and compressor power measurements (approximate heat of

compression).

EQUATION 11. HEAT REJECTION

Where

= Heat rejection (air-side based), Btu/hr

= Gross cooling capacity (air-side based), Btu/hr

= Compressor power (electrical measurement), W

= Conversion factor = 3.41214163, Btu/hr/W

Figure 11 illustrates the heat rejection calculations performed with the refrigerant

enthalpy method, and with the sum of air-enthalpy method gross cooling capacity

and power-measurement-based compressor heat of compression. Percent variations

range from -7% to 3%.

FIGURE 11. COMPARING BASELINE HEAT REJECTION CALCULATION AVERAGES – REFRIGERANT ENTHALPY METHOD

AND THE SUM OF AIR-SIDE GROSS COOLING CAPACITY AND COMPRESSOR HEAT OF COMPRESSION

39

,45

0

42

,96

6

44

,70

1

37

,25

2

40

,54

2

43

,14

6

35

,22

1

38

,23

6

41

,35

8

29

,28

0

33

,31

9

34

,27

0

26

,61

3

30

,79

5

32

,82

6

26

,03

1

29

,34

9

33

,10

4

10

,87

2

9,0

39

7,7

45

10

,54

1

8,9

11

7,2

28

10

,22

0

8,6

89

7,1

47

0.00

5000.00

10000.00

15000.00

20000.00

25000.00

30000.00

35000.00

40000.00

45000.00

50000.00

0

5,000

10,000

15,000

20,000

25,000

30,000

35,000

40,000

45,000

50,000

1 2 3 4 5 6 7 8 9

He

at T

ran

sfe

r (B

tu/h

)

Test Scenarios

Heat Rejection - Refg Enthalpy Method Gross Cooling Cap - Air Enthalpy Method Heat of Compression

2%

-1% -6%

0%

-2%

-7%

3% -1%

-3%




TEST SCENARIOS Table 11 lists all baseline tests, Table 12 lists all single-fault tests, and Table 13 lists

all multiple-fault tests.

TABLE 11. BASELINE TEST SCENARIOS

Test # Description Indoor Chamber

Air Condition

Outdoor Chamber Air

Condition

1

Baseline

80◦F /67◦F /51% (DB/WB/RH)

115◦F DB

2 95◦F DB

3 80◦F DB

4

75◦F /63◦F /52% (DB/WB/RH)

115◦F DB

5 95◦F DB

6 75◦F DB

7

70◦F /59◦F /52%

(DB/WB/RH)

115◦F DB

8 95◦F DB

9 75◦F DB

TABLE 12. SINGLE-FAULT TEST SCENARIOS


Air Condition

Outdoor Chamber

Air Condition

10 Low Refrigerant Charge – 13%

80◦F /67

◦F /51%

(DB/WB/RH) 95

◦F DB 11 Low Refrigerant Charge – 27%

12 Low Refrigerant Charge – 40%

13 Low Refrigerant Charge – 40% 75

◦F /63

◦F /52%

(DB/WB/RH) 115

◦F DB

14 Low Refrigerant Charge – 40% 70

◦F /59

◦F /52%

(DB/WB/RH) 75

◦F DB

15 High Refrigerant Charge - 10%

80◦F /67

◦F /51%

(DB/WB/RH) 95

◦F DB 16 High Refrigerant Charge – 20%

17 High Refrigerant Charge – 30%

18 High Refrigerant Charge – 30% 75

◦F /63

◦F /52%

(DB/WB/RH) 115

◦F DB


◦F /59

◦F /52%

(DB/WB/RH) 75

◦F DB


◦F /63.3

◦F /70%

(DB/WB/RH) 85

◦F DB





Air Condition

Outdoor Chamber

Air Condition


◦F /67.9

◦F /70%

(DB/WB/RH) 85

◦F DB


◦F /67.9

◦F /70%

(DB/WB/RH) 95

◦F DB

23 Refrigerant Line Restrictions - 32 psi drop

80◦F /67

◦F /51%

(DB/WB/RH) 95

◦F DB 24 Refrigerant Line Restrictions - 66 psi drop

25 Refrigerant Line Restrictions - 98 psi drop

26* Refrigerant Line Restrictions – 88 psi drop 75

◦F /63

◦F /52%

(DB/WB/RH) 115

◦F DB

27* Refrigerant Line Restrictions – 96 psi drop 70

◦F /59

◦F /52%

(DB/WB/RH) 75

◦F DB

28 (see multiple faults)

80◦F /67

◦F /51%

(DB/WB/RH) 95

◦F DB

29 Non-Condensables – 0.3 ounces (ozs.) N2

30 Non-Condensables – 0.8 ozs. N2

30a (see multiple faults)

31 Non-Condensables – 0.8 ozs. N2 75

◦F /63

◦F /52%

(DB/WB/RH) 115

◦F DB

32 Non-Condensables – 0.8 ozs. N2 70

◦F /59

◦F /52%

(DB/WB/RH) 75

◦F DB

33 Evaporator Airflow Reduction – 36%

80◦F /67

◦F /51%

(DB/WB/RH) 95

◦F DB 34 Evaporator Airflow Reduction – 51%

35 Evaporator Airflow Reduction – 59%

36 Evaporator Airflow Reduction – 63% 75

◦F /63

◦F /52%

(DB/WB/RH) 115

◦F DB

37 Evaporator Airflow Reduction – 32% 70

◦F /59

◦F /52%

(DB/WB/RH) 75

◦F DB

38 Condenser Airflow Reduction – 467 psig Compressor Discharge Pressure

80◦F /67

◦F /51%

(DB/WB/RH) 95

◦F DB 39

Condenser Airflow Reduction – 575 psig

Compressor Discharge Pressure



75◦F /63

◦F /52%

(DB/WB/RH) 115

◦F DB


70◦F /59

◦F /52%

(DB/WB/RH) 75

◦F DB

*Note: The restriction imposed in Test 25 is the same restriction imposed in Tests 26 and 27. However, return and outdoor chamber air condition variations cause fluctuations in the measured pressure drop.




TABLE 13. MULTIPLE-FAULT TEST SCENARIOS


Air Condition

Outdoor Chamber

Air Condition

28 Fault 1: Non-Condensables – 0.3 ozs. N2

Fault 2: Low Charge – 32% 80◦F /67◦F /51% (DB/WB/RH)

95◦F DB

30a Fault 1: Non-Condensables – 0.8 ozs. N2

Fault 2: Low Charge – 76%

43*

Fault 1: Evaporator Airflow Reduction – 36%

Fault 2: Condenser Airflow Reduction – 438 (467) psig Compressor Discharge Pressure

80◦F /67◦F /51% (DB/WB/RH)

95◦F DB 44*



45*



46 Fault 1: High Refrigerant Charge – 30%


80◦F /72.4◦F /70% (DB/WB/RH)

95◦F DB

47 Fault 1: Low Refg Charge - 13%


80◦F /67◦F /51% (DB/WB/RH)

95◦F DB

48

Fault 1: Low Refg Charge - 13%

Fault 2: Condenser Airflow Reduction – 624 psig Compressor Discharge Pressure

49

Fault 1: Low Refg Charge - 13%



50 Fault 1: Low Refg Charge – 27%


80◦F /67◦F /51% (DB/WB/RH)

95◦F DB

51

Fault 1: Low Refg Charge – 27%


52




53 Fault 1: Low Refg Charge – 40%


80◦F /67◦F /51% (DB/WB/RH)

95◦F DB

54



55




*Note: Compressor discharge pressure is presented in the form, P’ (P)




Where

P’ = resultant compressor discharge pressure, after evaporator airflow reduction is imposed

P = the originally imposed pressure, prior to evaporator airflow reduction

FDD PERFORMANCE TESTING For this investigation, the following is adopted: A symptom is a deviation in an

operating parameter from what may be considered typical or expected from normal

operation. Examples of operating parameters in an HVAC system are high-side and

low-side pressures, superheat, sub-cooling, and airflow rate. On the other hand, a

fault is a root cause of one or more symptoms. For example, low sub-cooling and low

low-side pressure are possible symptoms of a low charge fault.

FDD performance is evaluated qualitatively on its functionality. Fault diagnosis is

considered the primary function. The ability to detect symptoms caused by faults is

considered a secondary function. As such, the following elements are considered for

discussion:

- How accurately are the imposed faults diagnosed by FDD?

- Do misdiagnoses occur?

- What fault thresholds does the FDD technology demonstrate sensitivity?

- How accurately are fault symptoms detected?

The FDD technology’s reported diagnosis of the HVAC system was recorded on a

spreadsheet, once every minute. This was repeated until 10 entries were captured.

In addition, “spot measurements” were recorded for the 19 key parameters that

were displayed. However, it was impossible to capture all 19 reported parameters in

unison with the FDD technology’s refresh rate of about three seconds.

COMMENT

Diagnostic messages instantaneously take all measurements and calculations

into account and represent the “bottom-line” interpretation of system

performance. It is of great interest to evaluate FDD technologies based on

what diagnostic is reported.

Significant FDD output variance was observed with real-time operation. This,

in combination with a 3-second display refresh rate, presented logistical

challenges in recording the outputs of the FDD technology. It becomes

difficult to tie “spot readings” of measurements and calculations back to the

overall diagnosis message, when they all cannot be recorded simultaneously.

Access to all 19 reported parameters requires navigation through several

different screens. The action of moving from one screen to another requires

the user to wait until the next refresh of the display screen. Each screen

displays a distinct portion of the 19 parameters and the overall diagnostic

message.




Table 14 details the timestamps for the test window and the recorded FDD readings.

The test window includes a “start,” “mid,” and “end” timestamp: the steady-state

period spans from “start” to “mid” times, while the measurement period spans from

“mid” to “end.” Timestamps for FDD readings are marked with a “start” and “end”

timestamp.

COMMENT

An appropriate window of time is needed to allow a fault to settle into an

HVAC system to achieve “steady state”. Initially, the time allotted was chosen

based on literature made available by the FDD manufacturer regarding

“steady state” HVAC unit operation. The language is as follows:

“The unit must be operating in steady state, which means there must be no

condenser fans cycling and no thermostatic expansion valve (TxV) hunting

and the compressor should be running continuously for 10 to 15 minutes prior

to testing. A good indication of steady state operation is when the liquid

temperature stops changing.”

Generally, FDD readings were conducted in a window of time that falls within

or after the steady state and measurement test windows. For Tests 11, 29,

30, and 38, the FDD reading times occurred before the steady-state period of

the test window, but were still conducted after allowing a faulted system

runtime within compliance of the above definition of steady state. Complete

initiation of faults occurred as detailed in the “Notes” column.




TABLE 14. TRACKING TEST TIMESTAMPS

TEST

# DATE

TEST WINDOW WINDOW OF FDD

READINGS NOTES

START MID END START END

2 10/3/201

1 9:49:40

AM 10:19:30

AM 10:49:20

AM 10:14:00

PM 10:24:00 PM

10 9/16/201

1 3:52:13

PM 4:22:04

PM 4:51:54 PM

5:10:00 PM

5:19:00 PM

11 9/20/201

1 9:27:19

AM 9:57:09

AM 10:26:59

AM 8:55:00

AM 9:05:00 AM

*Fault initiated by 8:36 AM

(19 min runtime)

12 9/21/201

1 5:20:02

PM 5:49:52

PM 6:19:42 PM

6:19:00 PM

6:28:00 PM

13 9/21/201

1 1:20:02

PM 1:49:52

PM 2:19:42 PM

1:41:00 PM

1:50:00 PM

14 9/22/201

1 9:44:03

AM 10:13:53

AM 10:43:43

AM 10:03:00

PM 10:12:00 PM

15 10/3/201

1 1:08:00

PM 1:37:50

PM 2:07:40 PM

1:41:00 PM

1:50:00 PM

16 10/3/201

1 2:46:40

PM 3:16:30

PM 3:46:20 PM

3:17:00 PM

3:26:00 PM

17 10/3/201

1 5:00:00

PM 5:29:50

PM 5:59:40 PM

5:38:00 PM

5:47:00 PM

18 10/4/201

1 9:05:19

AM 9:35:09

AM 10:04:59

AM 9:45:00

AM 9:54:00 AM

19 10/4/201

1 12:22:19

PM 12:52:09

PM 1:21:59 PM

12:47:00 PM

12:56:00 PM

20 10/4/201

1 3:55:19

PM 4:25:09

PM 4:54:59 PM

4:31:00 PM

4:40:00 PM

21 10/5/201

1 10:26:57

AM 10:56:47

AM 11:26:37

AM 11:06:00

AM 11:15:00

AM

22 10/5/201

1 2:59:57

PM 3:29:47

PM 3:59:37 PM

3:43:00 PM

3:52:00 PM

23 10/6/201

1 12:30:18

PM 1:00:08

PM 1:29:58 PM

1:37:00 PM

1:46:00 PM

24 10/6/201

1 2:55:17

PM 3:25:07

PM 3:54:57 PM

3:19:00 PM

3:28:00 PM

25 10/6/201

1 4:30:17

PM 5:00:07

PM 5:29:57 PM

5:02:00 PM

5:11:00 PM

26 10/7/201

1 9:29:13

AM 9:59:03

AM 10:28:53

AM 10:40:00

AM 10:49:00

AM

27 10/7/201

1 1:22:13

PM 1:52:03

PM 2:21:53 PM

1:50:00 PM

1:59:00 PM

28 11/7/201

1 3:30:37

PM 4:00:07

PM 4:29:37 PM

4:59:00 PM

5:08:00 PM

29 11/8/201

1 9:30:17

AM 10:00:07

AM 10:29:57

AM 9:17:00

AM 9:27:00 AM

* Fault initiated by 8:43 AM (34 min runtime)

30 11/8/201

1 3:30:17

PM 4:00:08

PM 4:29:58 PM

3:19:00 PM

3:29:00 PM

* Fault initiated by 1:51 PM

(1 hr + 28 min. runtime)

30a 11/17/20

11 10:00:14

AM 10:30:05

AM 10:59:55

AM 10:15:00

AM 10:24:00

AM

31 11/17/20

11 7:00:15

PM 7:30:05

PM 7:59:56 PM

8:27:00 PM

8:36:00 PM

32 11/17/20

11 4:00:15

PM 4:30:05

PM 4:59:55 PM

4:51:00 PM

5:00:00 PM

33 10/12/20 11:30:32 12:00:02 12:29:32 1:05:00 1:14:00 PM




TEST

# DATE

TEST WINDOW WINDOW OF FDD

READINGS NOTES

START MID END START END

11 AM PM PM PM

34 10/12/20

11 2:09:00

PM 2:38:50

PM 3:08:40 PM

3:30:00 PM

3:39:00 PM

35 10/11/20

11 2:29:54

PM 2:59:24

PM 3:28:55 PM

3:26:00 PM

3:35:00 PM

36 10/13/20

11 11:05:00

AM 11:34:50

AM 12:04:40

PM 12:45:00

PM 12:54:00 PM

37 10/14/20

11 9:20:18

AM 9:50:08

AM 10:19:58

AM 2:30:00

PM 2:39:00 PM

38 10/19/20

11 12:02:42

PM 12:32:32

PM 1:02:22 PM

10:40:00 AM

10:50:00 AM

* Fault initiated by 9:30

AM (1 hr + 10 min. runtime)

39 10/18/20

11 5:15:03

PM 5:44:53

PM 6:14:43 PM

5:55:00 PM

6:04:00 PM

40 10/21/20

11 2:30:17

PM 3:00:07

PM 3:29:57 PM

2:58:00 PM

3:07:00 PM

41 10/21/20

11 10:06:57

AM 10:36:47

AM 11:06:37

AM 10:46:00

AM 10:55:00

AM

42 10/20/20

11 2:29:58

PM 2:59:48

PM 3:29:38 PM

3:28:00 PM

3:37:00 PM

43 10/25/20

11 10:10:04

AM 10:39:54

AM 11:09:44

AM 10:50:00

AM 10:59:00

AM

44 10/25/20

11 2:30:04

PM 2:59:54

PM 3:29:44 PM

3:11:00 PM

3:20:00 PM

45 10/26/20

11 11:50:13

AM 12:20:03

PM 12:49:53

PM 3:26:00

PM 3:35:00 PM

46 10/27/20

11 12:30:12

PM 1:00:02

PM 1:29:52 PM

2:15:00 PM

2:24:00 PM

47 10/27/20

11

4:00:12

PM

4:30:02

PM 4:59:52 PM

4:40:00

PM 4:50:00 PM

48 10/31/20

11 12:00:10

PM 12:30:00

PM 12:59:50

PM 2:17:00

PM 2:27:00 PM

49 10/31/20

11 3:30:10

PM 4:00:00

PM 4:29:50 PM

4:32:00 PM

4:42:00 PM

50 11/1/201

1 11:00:07

AM 11:29:57

AM 11:59:47

AM 1:10:00

PM 1:20:00 PM

51 11/1/201

1 3:00:09

PM 3:29:59

PM 3:59:49 PM

3:19:00 PM

3:29:00 PM

52 11/2/201

1 9:30:12

AM 10:00:02

AM 10:29:52

AM 1:32:00

PM 1:42:00 PM

53 11/2/201

1 2:40:12

PM 3:10:02

PM 3:39:52 PM

3:25:00 PM

3:35:00 PM

54 11/3/201

1 11:00:15

AM 11:30:05

AM 11:59:54

AM 3:35:00

PM 3:45:00 PM

55 11/3/201

1 2:00:15

PM 2:30:05

PM 2:59:55 PM

2:10:00 PM

2:20:00 PM

Generally, testing experience showed more erratic FDD outputs in Tests 11

and 29. It is recommended to conduct FDD readings after or during the

measurement window (mid-to-end portion of test window).




BASELINE TESTING Limited data typically exist for equipment operating at lower indoor and outdoor test chamber air conditions. Indoor chamber air DB temperatures ranging from 70°F to

80°F were chosen as recommended by the TAG. Indoor chamber air humidity was

chosen to maintain a similar RH as that of the 80°F/67°F (DB/WB) AHRI 210/240

condition. Outdoor chamber air conditions were chosen from TAG input, and from

consideration of SCE service territories.

As per the current 2008 Title-24 standards, outdoor design conditions are selected

from reference Joint Appendix JA2. For general comfort cooling applications, outdoor

conditions are based on the 0.5% cooling DB and Mean Coincident WB values. SCE

service territories contain climate zones (CZ’s) 6, 8, 9, 10, 13, 14, 15, and 16. CZ 15, El Centro, had the most extreme 0.5% cooling DB design condition of 111°F. In

addition, California research evaluating the performance of air conditioners optimized for “hot and dry” climates, defines the “hot and dry” DB at 115°F.v Insight on the

lower bound of outdoor test chamber design conditions was drawn from the 2009

ASHRAE Fundamentals. CZ6 had the lowest cooling design DB value of the SCE CZ’s:

The Los Angeles International Airports’ (CZ6) 2% cooling design DB condition was 77.8°F.

Ultimately, 115°F was chosen as the upper limit, 95°F was chosen as the

intermediate, and 75°F was chosen as the lower limit for outdoor chamber test

conditions. The current standard condition in AHRI 210/240 is 95°F. These tests were

conducted with no directly imposed faults, maintaining control parameters at several

selections of return and outdoor chamber air conditions. Testing was conducted with

the FDD technology installed and functional. Baseline testing required that the

installed liquid line restriction valve be set to the wide open position. Diagnostic

outputs were recorded for Test 2.

COMMENT

One might assume that ideally, baseline testing should be done without

installing the FDD technology, so as to quantify unbiased, independent HVAC

performance. However, given that the evaluated FDD generally demonstrates

itself as a “non-invasive” technology, this is not considered particularly

critical. Installation of this particular FDD was not anticipated to have any

significant impact on HVAC operation. The FDD’s instrumentation consists of

items like air sensors, clamp-on thermocouples, and refrigerant pressure

hoses. The FDD setup did not introduce anything significantly different from

what is used in typical laboratory instrumentation. More exploration may be

done to quantify impacts, if any, of FDD technology instrumentation. This

work could lead to more clarity regarding what could be considered

“invasive.”

Factory-shipped refrigerant charge for the HVAC system’s condensing unit is 6 lbs. 9 ozs. Manufacturer literature specifies 10°F condenser sub-cooling to achieve proper

charge levels, after a refrigerant line-set and indoor unit have been added/installed.

Baseline tests were conducted with a total refrigerant charge of 8 lbs. 3 ozs.




Figure 12 maps the refrigeration cycle process on a pressure-enthalpy diagram,

using measurements/calculations from baseline Tests 1-9. The tested variations in

outdoor test chamber conditions are generally observed to have a more pronounced

impact than the explored indoor chamber air condition variations. Test 2 was

repeated on three separate days, to establish confidence in laboratory test results.

Figure 13 maps the refrigeration cycle process on a pressure enthalpy diagram, for

the three repeated instances of Test 2. Close grouping of the processes shows good

agreement. Table 15 also illustrates the maximum percent variations of key

parameters between the original Test 2 and its three repeated instances.

FIGURE 12. P-H DIAGRAM: BASELINE REFRIGERATION PROCESSES AT VARYING OPERATING CONDITIONS

0

100

200

300

400

500

600

700

50 70 90 110 130 150 170 190 210

Pre

ssu

re (

psi

g)

Enthalpy (Btu/lb)

Saturation Dome - IIR Ref Test 1 Test 2

Test 3 Test 4 Test 5

Test 6 Test 7 Test 8

Test 9




FIGURE 13. P-H DIAGRAM: REPEATING THE AHRI TEST CONDITION

TABLE 15. MEASUREMENT AVERAGES FOR THE REPEATED INSTANCES OF TEST 2

# PARAMETER (UNITS) TEST 2 TEST 2

(REPEAT

1)

TEST 2

(REPEAT

2)

TEST 2

(REPEAT

3)

MAXIMUM

%

VARIATION

1 Indoor Chamber Air Condition DB (°F)

80 80 80 80 0%

2 Indoor Chamber Air Condition WB (°F)

67 67 67 67 0%

3 Outdoor Test Chamber DB (°F) 95 95 95 95 0%

4 Compressor Power (W) 2,649 2,647 2,669 2,670 1%

5 Evaporator Fan Power (W) 536 546 546 542 2%

6 Condenser Fan Power (W) 123 123 123 123 0%

7 Total Power (W) 3,325 3,328 3,352 3,344 1%

8 EER - Refg-side (Btu/hr/W) 10.6 10.4 10.6 10.5 2%

9 EER - Air-side (Btu/hr/W) 10.0 9.5 10.0 9.9 5%

10 Refrigerant-side Gross Cooling Capacity (Btu/hr)

35,187 34,639 35,509 35,245 2%

11 Air-side Gross Cooling Capacity (Btu/hr)

33,319 31,646 33,402 33,229 5%

12 Refrigerant-side Heat Rejection (Btu/hr)

42,966 42,439 43,149 42,997 1%

13 Heat Rejection: Air-side Gross Cooling Capacity + Compressor

Power (Btu/hr)

42,358 40,679 42,510 42,339 4%

14 Air-side Sensible Cooling Capacity (Btu/hr)

24,901 24,631 25,915 25,032 4%

R1 R2

R4 R5

R7

R8

0

100

200

300

400

500

600

700

50 70 90 110 130 150 170 190 210

Pre

ssu

re (

psi

g)

Enthalpy (Btu/lb)

Test 2 Saturation Dome - IIR Ref Test 2 (Repeat 1)

Test 2 (Repeat 2) Test 2 (Repeat 3)




# PARAMETER (UNITS) TEST 2 TEST 2

(REPEAT

1)

TEST 2

(REPEAT

2)

TEST 2

(REPEAT

3)

MAXIMUM

%

VARIATION

15 Air-side Latent (Psychrometric) Cooling Capacity (Btu/hr)

7,633 6,295 6,725 7,414 19%

16 Air-side Latent (Scale) Cooling Capacity (Btu/hr)

8,161 7,732 8,337 8,243 5%

17 Evaporator Air Flow Rate (SCFM) 1,147 1,164 1,163 1,157 1%

18 Refrigerant Mass Flow (lbs./min) 8.2 8.1 8.3 8.3 1%

19 Evaporator Superheat (°F) 10 9 11 9 12%

20 Total Superheat (°F) 15 14 15 14 10%

21 Condenser Sub-Cooling (°F) 6 6 6 5 12%

22 Total Sub-Cooling (°F) 10 10 10 9 5%

23 Saturated Condensing Temperature (°F)

112 112 112 112 0%

24 Saturated Evaporator Temperature (°F)

47 47 48 49 3%

25 Compressor Discharge Pressure

(psig) 391 393 393 392 1%

26 Compressor Suction Pressure (psig)

136 134 136 135 2%

COMMENT

After conducting tests 1-55, it was discovered that manufacturer literature for refrigerant charging called for 10°F at the condensing unit service port.

Laboratory testing and subsequent baseline charging procedures used total sub-cooling measured at the indoor unit: condenser sub-cooling was 6°F. In

order to ensure the integrity of the laboratory tests, Test 2 was repeated with 10°F condenser sub-cooling. This resulted in 8 lbs. 9 ozs. of refrigerant

charge. Table 16 depicts key performance parameters for the original Test 2,

and the repeated Test 2 with 8 lbs. 9 ozs. of refrigerant.




TABLE 16. TEST 2 DATA COMPARISON: 8 LBS. 3 OZS. VS. 8 LBS. 9 OZS.

# PARAMETER (UNITS)

TEST 2

(BASELINE

8 LBS. 9

OZS.)

TEST 2

(8 LBS. 3

OZS.)

%

DIFFERENCE

1 Indoor Chamber Air Condition DB (°F)

80 80 0%

2 Indoor Chamber Air Condition WB (°F)

67 67 0%

3 Outdoor Test Chamber DB (°F) 95 95 0%

4 Compressor Power (W) 2,675 2,649 -1%

5 Evaporator Fan Power (W) 539 536 -1%

6 Condenser Fan Power (W) 122 123 0%

7 Total Power (W) 3,348 3,325 -1%

8 EER - Refg-side (Btu/hr/W) 10.6 10.6 0%

9 EER - Air-side (Btu/hr/W) 9.6 10.0 4%

10 Refrigerant-side Gross Cooling

Capacity (Btu/hr) 35,344 35,187 0%

11 Air-side Gross Cooling Capacity (Btu/hr)

32,186 33,319 4%

12 Refrigerant-side Heat Rejection (Btu/hr)

43,146 42,966 0%

13 Heat Rejection: Air-side Gross

Cooling Capacity + Compressor Power (Btu/hr)

41,313 42,358 3%

14 Air-side Sensible Cooling Capacity (Btu/hr)

24,367 24,901 2%

15 Air-side Latent (Psychrometric) Cooling Capacity (Btu/hr)

7,070 7,633 8%

16 Air-side Latent (Scale) Cooling Capacity (Btu/hr)

8,142 8,161 0%

17 Evaporator Air Flow Rate (SCFM) 1,150 1,147 0%

18 Refrigerant Mass Flow (lbs./min) 8.1 8.2 1%

19 Evaporator Superheat (°F) 10 10 0%

20 Total Superheat (°F) 14 15 5%

21 Condenser Sub-Cooling (°F) 10 6 -39%

22 Total Sub-Cooling (°F) 14 10 -28%

23 Saturated Condensing Temperature (°F)

113 112 -1%

24 Saturated Evaporator Temperature

(°F) 48 47 -1%

25 Compressor Discharge Pressure (psig)

396 391 -1%

26 Compressor Suction Pressure (psig) 133 136 2%




COMMENT

It is also noteworthy to mention that California’s Title-24 (2008) reference appendices give Home Energy Rating System (HERS) raters a +/- 4°F

tolerance for determining acceptable charge with sub-cooling.vi

Given the marginal additional refrigerant needed and the results of repeat

testing relative to previous repeated tests, data integrity is intact.

For future reference, anyone performing tests of this nature should always

comprehensively verify that the HVAC setup is at the nominal/original

manufacturer specification: verifying and adhering to correct evacuation and

charging procedures, maintaining cleanliness of evaporator and condenser

surfaces, etc.

SINGLE FAULT TESTING For all single-fault testing scenarios, the strategy was to:

- Capture the effects of three incremental fault levels at a standardized condition of 80°F/67°F (DB/WB) indoor chamber, 95°F outdoor

chamber

- Capture the effects of the most pronounced fault level, at two extra

combinations of indoor and outdoor test chamber air conditions

Increments of faults are generally chosen based on the following criteria:

- Is the fault increment representative of what happens in the field?

- Does the fault increment induce a failure mode or otherwise prohibit the HVAC system from operating in a steady state fashion? Examples

include:

o A condenser airflow reduction may be severe enough to cause HVAC

system shutdown, by tripping the high-pressure switch.

o A liquid line restriction could be severe enough to drop low-side pressures

to a point that would cause HVAC system shutdown by tripping the low-

pressure switch.

o A liquid line restriction could be severe enough to drop the evaporator

temperature low enough to cause coil frosting.

When moving between test scenarios, proper measures were taken to reverse the

effects of each fault to bring the HVAC system back to baseline operating conditions,

as needed. Bringing the system back to baseline was not necessary after performing

incremental faults in the same family/category. For example, after testing 13% low

charge, it was not necessary to bring the system back to baseline before proceeding

with a 27% low charge test. Care was taken during refrigerant charging and recovery

procedures:

1. Self-locking hoses are used to eliminate the need to purge hoses to prevent

refrigerant contamination with air.

2. Refrigerant charging is slowly done through an evaporator outlet service port:

slow metering of refrigerant, coupled with the suction line run from the indoor




test chamber back to the condensing unit, ensures no liquid refrigerant is

introduced into the compressor.

Procedures for specific faults are detailed in the sections below. Tests were

conducted in a manner analogous to that of the cooling mode steady-state wet-coil

tests in AHRI 210/240.

A. (Tests 10-14) Low Refrigerant Charge

The low refrigerant charge fault describes a state where an HVAC system contains

refrigerant charge levels significantly below that which was intended by the

manufacturer. Low charge levels may occur because of improper charging or

servicing practices, or general system leakage. The HVAC system will have less

working fluid available to remove heat from the conditioned space(s) and may

operate with significant performance degradation.

Increments of low refrigerant charge are defined on a percent of nominal charge lost

basis (by mass). The 13% low charge scenario refers to an HVAC system that

contains 87% of its nominal charge. Defined, nominal charge is the amount of

refrigerant required to achieve compliance with manufacturer specifications.

Refrigerant was recovered from an HVAC system as it was running. The HVAC

system had a baseline charge of 8 lbs. 3 ozs. Refrigerant was recovered from the

high-side service port near the TxV inlet, into an appropriate reclaim tank. The

weight of the tank is continually measured and refrigerant is weighed into the

reclaim tank. Low charge levels were tested in 1 lb. 1.5 oz. increments of removal.

This corresponds to total charge levels of 7 lbs. 1.5 ozs., 6 lbs., and 4 lbs. 14.5 ozs.,

for Tests 10, 11 and 12, respectively. Tests 13 and 14 capture different combinations

of indoor and outdoor chamber air conditions with a total charge of 4 lbs. 14.5 ozs.

The intent of the test scenarios was to create a performance curve with varied low

charge fault levels. Initially, 10% increments were chosen, however, documentation

errors of the HVAC unit’s nominal charge lead to miscalculation of the incremental

refrigerant mass to be removed. This resulted in tests conducted with 1 lb. 1.5 oz.

increments. Regardless, the tested increments still achieved the desired goal of

creating a performance curve with varied low-charge fault levels.

COMMENT

For future reference, it is suggested that anyone performing tests of this

nature should always comprehensively verify that the HVAC setup is at the

original manufacturer specification: verifying correct evacuation and charging

procedures have been followed, maintaining cleanliness of evaporator and

condenser surfaces, etc.

Mixed-phase refrigerant flow was encountered in the liquid line at all tested

levels of low charge. Refrigerant-side calculations of gross cooling capacity

were compromised through errors with refrigerant mass flow measurement

and enthalpy calculations at state points R5 through R8. The refrigerant

regression model could still be used to report refrigerant mass flow and check

compressor power. Air-side measurements were solely relied upon for gross

cooling capacity calculations.




B. (Tests 15-22) High Refrigerant Charge

The high refrigerant charge fault describes a state where an HVAC system contains

refrigerant charge levels significantly above the original manufacturer specifications.

High levels of refrigerant charge may occur from improper charging/servicing. The

HVAC system will have excessive working fluid available to remove heat from the

conditioned space. As a result, the system may operate with increased high side

pressures, significant performance degradation, and may run the risk of introducing

liquid refrigerant into the compressor.

Increments of high refrigerant charge are defined on a percent of nominal charge

basis. For example, the 20% high charge scenario refers to an HVAC system that

contains 120% of its nominal charge. Nominal charge shall be defined as the amount

of refrigerant required to achieve manufacturer specifications.

The extreme test scenario may be considered to be the state right before:

- Liquid refrigerant is introduced into the compressor

or

- The HVAC system shuts down on high head pressure

o The high pressure cutout limit for the HVAC system’s compressor was

determined to be 650 psig

(Tests 15-19)

High charge faults were imposed by weighing in additional refrigerant to a running

HVAC unit. This HVAC unit had a baseline charge of 8 lbs. 3 ozs. Only new R-410a

refrigerant was added, to eliminate the possibility of contaminants. Since R-410a is a

blend, it is essential to ensure liquid is pulled from the supply tank. If vapor is pulled

from the supply tank, the constituents of the blend will boil away from the supply

tank at different rates, thereby changing the ratio of the blend added to the HVAC

system. R-410a was pulled from the supply tank as a liquid, and throttled into the

system as a vapor without impacts to the blend’s ratio. Charging was done through a

low-side service port near the evaporator outlet. Care was taken to throttle

refrigerant in slowly, so as not to introduce any liquid refrigerant into the

compressor.

COMMENT

Caution: It is important to note that the evaporator outlet service port was

one of the available service ports not outfitted with a pressure transducer.

Previous attempts at adding refrigerant through ports outfitted with pressure

transducers resulted in “drift” of those sensors. Readings floated out of

agreement with other redundant pressure sensors. The affected pressure

transducer was replaced, and charging procedures were changed accordingly.

Testing was performed with 10%, 20%, and 30% high charge levels (13 oz.

increments). This corresponds to total charge levels of 9 lbs., 9 lbs. 13 ozs., and 10

lbs. 10 ozs., respectively. It was decided not to mimic the percent increments

analogous to the low charge tests: high charge tests run the risk of “slugging” the




compressor with liquid refrigerant. Tests 18 and 19 capture different combinations of

indoor and outdoor chamber air conditions with the 24% high charge fault.

(Tests 20-22)

These three test scenarios were specifically chosen to provide data to other parallel

efforts: fault modeling challenges exist regarding prediction of HVAC performance at

more pronounced levels of wet evaporator coils and low sub-cooling. Tests 20

through 22 were conducted with a total refrigerant charge of 10 lbs. 10 ozs. (24%

high charge) at additional permutations of indoor and outdoor chamber air

conditions.

C. (Tests 23-27) Refrigerant Line Restrictions

The refrigerant line restrictions fault describes a state in which refrigerant flow is

unintentionally restricted in a certain part of the liquid line. These cause unwanted

pressure drops at certain points in the system. These restrictions include sources

such as bent refrigerant lines, dirty liquid line filter-driers, or solder blockages at pipe

joints. Restricted/clogged expansion devices may also exhibit similar impacts to the

HVAC system. High levels of line restriction may result in system failure on low

suction pressure or evaporator frosting.

A refrigerant line restriction was simulated with a ball valve on the liquid line of the

HVAC system. The pressure differential across the restriction valve was measured.

Increments of line restrictions were defined in set pressure drops across the

restriction valve, measured in pounds per square inch (psi). Forty-six ft. of liquid line

was needed for the laboratory setup.

FIGURE 14. LINE RESTRICTION VALVE

The extreme test scenario may be considered to be the state right before:

- Evaporator frost forms (saturated evaporator temperatures below 32°F),

or

- The HVAC system shuts down on low suction pressure

o The low pressure cutout limit for the HVAC system’s compressor was


Incremental line restriction faults were initiated while the HVAC unit was activated

and running. Restrictions were initiated with the installed liquid line ball valve until

certain measured psi drops in the liquid line were achieved. Testing was performed

with 32 psi, 66 psi, and 98 psi liquid line restrictions. Two additional tests capturing




different combinations of indoor and outdoor chamber air conditions were also

conducted with the same restriction that created the 98-psi fault. However, the

variance in operating conditions attributed to changes in measured psi drops, even

though the restriction remained the same. As a result, the liquid line pressure drop

was 88 psi for Test 26 and 96 psi for Test 27.

COMMENT

Use of a ball valve is not recommended as fine-tuned adjustments proved

difficult. Excessive tuning, required to dial in target psi drops, adds to test

burden. Target pressure drops were originally 30 psi, 60 psi, and 90 psi.

Valve selection for liquid line restriction testing should consider those that

feature more precise control and low losses in the fully open position, such as

a needle valve.

D. (Test 29, 30, 31, and 32) Non-Condensables

The refrigerant line non-condensables fault describes a state in which contaminants

such as air, water vapor, or nitrogen mix with the refrigerant in an HVAC system.

The physical properties of these contaminants and their subjection to the HVAC

system’s working pressures mean they exist as gases throughout the system. These

contaminants impose their own properties on the overall working fluid, which

typically results in performance degradation. Non-condensables may be introduced

through faulty equipment servicing.

Refrigerant non-condensables were simulated with nitrogen gas. Increments were

defined on a mass-of-nitrogen basis. Figure 15 displays the nitrogen tank and scale

setup used for testing.

FIGURE 15. NON-CONDENSABLES - NITROGEN

A mass of nitrogen exists at a specific pressure, when introduced into an empty non-

running HVAC system by itself. This pressure is directly proportional to the mass of

nitrogen introduced into an empty system. The extreme scenario may be considered

as the mass of nitrogen that pressurizes to one atm, when introduced into an HVAC

system by itself (no refrigerant). This represents a scenario where an empty HVAC

system, exposed to atmospheric pressure, is not subjected to evacuation, and is

charged with refrigerant.




Appropriate amounts of nitrogen faults were determined through pretesting.

First, all refrigerant was recovered from the HVAC system. Refrigerant was

recovered from the service port at the TxV inlet, until the HVAC units’

operating pressures reached levels near those that would cause high or low

pressure cut-out. Accordingly, the unit was then shut off, and refrigerant was

completely recovered through both the service ports on the TxV inlet and the

evaporator outlet. Then, a 1,000-micron vacuum (approximately 99.9% of a

perfect vacuum) was pulled on the system. Lastly, nitrogen was weighed into

the HVAC system through both the TxV inlet and evaporator outlet ports, until

one atm was achieved. This corresponded to 0.8 ozs. of nitrogen for this

specific test setup.

COMMENT

It is important to note the use of a separate tool for measurement of

nitrogen-charged HVAC system pressures in the vicinity of 1 atm. Pressure

transducers used in laboratory instrumentation of the HVAC system were

suitable for higher typical operating pressures of HVAC systems, but not for

pressures of one atm or vacuum pressures. The Robinair 14830A Thermistor

Vacuum Gauge used for HVAC system evacuation, was suitable for vacuum

pressures in the range of 25,000 to 50 microns, but not for pressures around

1 atm (1 atm = 750,000 microns). A separate tool, the Fluke PV350

pressure/vacuum transducer module, was used to confirm 1 atm of nitrogen

charge. Pressure was measured at the TxV inlet service port.

There are two types of methods in which non-condensables may be introduced into

an HVAC system for testing:

1. The correct nominal charge has been weighed in with an additional specified

mass of nitrogen. That is to say, the lbs. of correct nominal charge of the

system is known, and simply weighed into a system that contains non-

condensables.

2. A specified mass of nitrogen has been added, and a non-nominal charge of

refrigerant has been added. That is to say, refrigerant is incrementally added

to an HVAC system containing non-condensables, until the design sub-cooling

value has been achieved (10°F).

Method 1 is considered representative of a singular, non-condensables fault. Method

2 is considered representative of a multiple simultaneous fault comprised of non-

condensables and low refrigerant charge. As such, Method 1 was used for singular-

fault testing.

For Tests 29 and 30, the HVAC unit underwent steady state testing with mixture

iterations consisting of 8 lbs. 3 ozs. of refrigerant and nitrogen levels of 0.8 ozs. and

0.3 ozs. (approximately one third of the one atm pretest amount). The mixture

containing 8 lbs. 3 ozs. of refrigerant and 0.8 ozs. of nitrogen was tested at two

additional combinations of indoor and outdoor chamber air conditions (Tests 31 and

32).




E. (Tests 33-37) Evaporator Airflow Reduction

The evaporator airflow reduction fault describes a state in which the HVAC system’s

evaporator is reduced due to factors such as: airflow obstructions, dirty/fouled

evaporator, dirty filters, or evaporator fan problems. Evaporator airflow reductions

result in lower evaporator temperatures/pressures and significant performance

degradation may result. High levels of evaporator airflow reduction may result in

system failure on low suction pressure or evaporator frosting.

Airflow was measured at the evaporator. The evaporator airflow reduction fault was

simulated by restricting evaporator airflow at the duct inlet fan. Evaporator airflow

reduction faults were imposed with a wooden-sheet obstruction. Fault increments

were based on the percent reduction in evaporator airflow. For example, given a

baseline airflow of 1,147 SCFM, an HVAC unit running with 772 SCFM would

represent a 33% evaporator airflow reduction fault scenario (where 1-[772/1,147] =

33%). Figure 16 depicts the evaporator airflow reduction.

FIGURE 16. EVAPORATOR AIRFLOW REDUCTION

The extreme test scenario may be considered as the state right before:

- Evaporator frost forms (saturated evaporator temperatures under 32°F),

or

- The HVAC system shuts down on low suction pressure

o The low pressure cutout limit for the HVAC system’s compressor is


To establish the fault increments, pretesting for the extreme condition was needed.

Airflow restriction was performed while monitoring the evaporator airflow,

compressor suction pressure, and Saturated Evaporator Temperature (SET). A SET of 35.2°F was chosen as the extreme condition, as it established a point before

evaporator frost would occur. Suction pressure at this point was 106 psig: low-

pressure switch trip would not likely occur before coil frosting would. Refrigerant-side

cooling capacity was found to be 80% of its nominal value. Evaporator airflow rates




of 772 SCFM, 584 SCFM, and 496 SCFM were used for Tests 33, 34, and 35,

respectively.

COMMENT

Evaporator airflow reduction faults were initially defined as “evaporator heat

transfer reduction” faults. The evaporator airflow rates of 772 SCFM, 584

SCFM, and 496 SCFM were established to give even increments of reduction

in evaporator heat transfer. The strategy of using heat transfer reduction

increments was dropped in favor of airflow reduction increments, for the

following reasons:

o Heat transfer reduction increments understate the severity of the fault

increment. At AHRI test chamber conditions, a 20% reduction in heat

transfer at the evaporator requires airflow to be reduced by 57%!

o Heat transfer reduction increments require iterations of airflow

adjustment that are dependent on real-time monitoring of evaporator

heat transfer rates. This significantly adds to test burden. In addition,

the calculation method chosen (air-side, refrigerant-side, etc.) impacts

the level of repeatability/uncertainty for the increments.

Two additional tests captured an extreme fault condition at two additional indoor and

outdoor chamber air conditions (Test 36 and 37). In these scenarios, restrictions

were initiated up to a point before evaporator coil frost. The SET was allowed to drift

down to 32.6°F in Test 36, and down to 32.5°F in Test 37. Evaporator airflow rates of

442 SCFM and 815 SCFM were used for Tests 36 and 37, respectively.

F. (Tests 38-42) Condenser Airflow Reduction

The condenser airflow reduction fault describes a state in which the HVAC system’s

condenser encounters reduced airflow because of factors such as: condenser

inlet/outlet obstructions/fouling, or condenser fouling. Condenser airflow reductions

result in higher refrigerant condensing temperatures/pressures and significant

performance degradation may result. High levels of condenser airflow reduction may

result in system failure on high head pressure.

The condenser airflow reduction faults were simulated by restricting airflow at the

condenser inlets. Increments are imposed with paper towel blockages at the inlet to

the condensing unit. Figure 17 depicts the simulated condenser airflow reduction.




FIGURE 17. CONDENSER AIRFLOW REDUCTION

The extreme scenario may be considered as the state before the HVAC system shuts

down on high head pressure. For the test setup, airflow was not measured at the

condenser. Condenser airflow reduction faults were tracked through increments of

compressor discharge pressure.

Airflow restriction was performed while monitoring the compressor discharge

pressure. The HVAC unit’s high-pressure switch is set to trip when compressor

discharge pressures reach 650 psig. At a discharge pressure of 613 psig, the

refrigerant-side heat rejection was recorded and established as the extreme fault

condition. Tests 38, 39, and 40 ran at compressor discharge pressures of 467 psig,

575 psig, and 613 psig, respectively.

COMMENT

Condenser airflow reduction faults were initially defined as “condenser heat

transfer reduction” faults. The compressor discharge pressures of 467 psig,

575 psig, and 613 psig were established to give even increments of reduction

in condenser heat rejection. The strategy of using heat transfer reduction

increments was dropped in favor of airflow reduction (compressor discharge

pressure) increments, for the following reasons:

o Heat transfer reduction increments understate the severity of the fault

increment. At AHRI test chamber conditions, an 11% reduction in heat

rejection at the condenser requires airflow to be reduced until

compressor discharge pressures rise to 613 psig (high pressure cutout

at 650 psig)!

o Heat transfer reduction increments require iterations of airflow

adjustment that are dependent on real-time monitoring of condenser

heat rejection calculations. This significantly adds to test burden. In

addition, the calculation method chosen (air-side, refrigerant-side,

etc.) impacts the level of repeatability/uncertainty for the increments.

Two more tests captured the extreme fault at two additional indoor and outdoor

chamber air conditions (Test 41 and 42). In these scenarios, restrictions were

initiated until a point before the high-pressure switch would trip. In Test 41, the




discharge pressure was allowed to float up to 622 psig. In Test 42, the discharge

pressure was allowed to float up to 612 psig.

COMMENT

The condenser uses an axial-flow fan (or “propeller” type). Unlike the

centrifugal fan used in the evaporator, the condenser’s axial flow fan is not

capable of overcoming large pressure drops. At Tests 40 and 42, air

restrictions caused backflow of air to occur across the condenser. Effectively,

this shifts where the condenser “inlet” and “outlet” are located for

measurement purposes. Accordingly, condenser inlet and outlet air DB

measurements were ignored. This presents an issue for controlling outdoor

chamber temperatures by controlling condenser inlet temperatures to the

operating and average deviation targets. Separate measurements were

needed. An array of thermocouples setup along a pole running from the floor

to the ceiling of the outdoor test chamber was available as an alternative.

Trending temperatures of the 11 available thermocouples were analyzed for

previous tests that did not feature condenser backflow issues. Percent

variations ranging from 0.5% to 1.9% were observed when comparing the

outputs of the thermocouple at a level of 5 ft. off the ground, with the

condenser inlet temperature. As such, the 5-ft level thermocouple was chosen

as a suitable replacement monitoring point for establishing outdoor chamber

air conditions at proper tolerances. Table 17 illustrates the percent variations

between the thermocouple at the 5-ft level and the condenser inlet

temperatures; the averages across the 30-minute measurement periods are

compared.

TABLE 17. COMPARING A REDUNDANT OUTDOOR TEST CHAMBER TEMPERATURE SENSOR WITH AVERAGE CONDENSER

INLET TEMPERATURES FOR BASELINE TESTS

TEST # TEMP - RM4 POLE AT 5FT (°F) CONDENSER INLET TEMP (°F) % VARIATION

1 114.4 115.0 0.5%

2 94.0 94.7 0.7%

3 79.8 80.3 0.6%

4 114.3 115.0 0.6%

5 94.3 95.0 0.7%

6 73.7 75.1 1.9%

7 114.3 115.0 0.6%

8 94.3 95.2 0.9%

9 73.7 75.0 1.7%




MULTIPLE FAULT TESTING Multiple-fault scenarios explored three levels of imposed fault combinations, and focused on operation at standard AHRI indoor and outdoor test chamber

air conditions.

For all multiple-fault scenarios tested, the strategy was to focus on capturing

the effects of three incremental fault levels at the standardized condition of 80°F/67°F (DB/WB) indoor, and 95°F outdoor. Increments of faults are

generally chosen based on the following criteria:

- Is the fault increment representative of what happens in the field?

- Does the fault increment induce a failure mode, or otherwise prohibit

the HVAC system from operating in a steady state fashion? Examples include:

o A condenser airflow reduction may be severe enough to cause HVAC

system shutdown, by tripping the high-pressure switch (failure).

o A liquid line restriction could be severe enough to drop low-side pressures

to a point that would cause HVAC system shutdown by tripping the low-

pressure switch (failure).

o A liquid line restriction could be severe enough to drop the evaporator

temperature low enough to cause coil frosting (transient impacts, will

build up).

When moving between test scenarios, proper measures were taken to reverse the

effects of each fault to bring the HVAC system back to baseline operating conditions,

as needed. Procedures for specific faults are detailed in the proceeding sections.

G. (Tests 28 and 30a, 2-faults) Non-condensables and Low Refrigerant

Charge

Non-condensables may be introduced into an HVAC system in two methods:

1. The correct nominal charge of the system is known, and simply weighed into

a system that contains non-condensables.

2. Refrigerant is incrementally added to an HVAC system containing non-condensables until the design sub-cooling value has been achieved (10°F).

Method 1 is considered representative of a singular, non-condensables fault. Method

2 is considered representative of a multiple simultaneous fault comprised of non-

condensables and low refrigerant charge. As such, Method 2 was used for these

multiple-fault tests. Testing was conducted with 0.8 oz. and 0.3 oz. levels of nitrogen

(see singular non-condensables fault pre-testing). Method 2 required a few iterations to determine the level of refrigerant needed in order to achieve the 10°F sub-cooling

that would falsely indicate proper charge.




COMMENT

This iterative method required the following considerations:

It was unknown how much refrigerant could be added with the set level of

nitrogen, to allow the HVAC system to operate. Too little refrigerant and the

system would have tripped off on low suction pressure. Too much refrigerant, and the tester would have overshot the 10°F sub-cooling goal. From that

point on, additional refrigerant only worked to further increase sub-cooling.

Attempts to recover refrigerant would result in pulling an unknown mixture of

nitrogen and R-410a. Overshooting the sub-cooling target means the tester

must recover the entire mixture and start all over again.

First, all refrigerant was recovered. Then, a 1,000-micron vacuum (approximately

99.9% of a perfect vacuum) was imposed on the empty HVAC system. Then, a

specified amount of nitrogen was weighed in. Then, the minimum amount of

refrigerant was added to allow the HVAC system to operate without tripping off on

low suction pressure. Pretests with this setup indicated that at 2 lbs. of R-410a (no

nitrogen), the unit was still able to operate without tripping on low suction pressure.

Two lbs. of R-410a represented the starting point.

Additional refrigerant was then incrementally weighed in slowly, until proper sub-

cooling was achieved. With 0.3 ozs. of nitrogen (Test 28), a total of 5 lbs. 9 ozs. of R-410a (32% low charge) was needed to achieve 10°F sub-cooling. With 0.8 ozs. of

nitrogen (Test 30a), 10°F sub-cooling was achieved with a total of 2 lbs. of R-410a

(76% low charge).

COMMENT

Severe performance penalties were realized in Test 30a: 0.8 ozs. N2 + 76%

low charge at AHRI conditions: -95% EER and -96% gross cooling capacity

penalties. This test scenario was originally established to mimic a scenario

where an HVAC system, vented to atmosphere, was not subjected to any

vacuum before being charged with refrigerant to a target sub-cooling value.

Due to the extreme nature of the performance impact, it may not be suitable

to assume this as a realistic scenario of maintenance mal-practice that would

go unnoticed in the field.

H. (Tests 43 – 45, two faults) Evaporator and Condenser Airflow Reduction

Airflow was measured at the evaporator, but not at the condenser. Evaporator

airflow reduction may be imposed at different airflow increments, but condenser

airflow reduction is imposed in compressor discharge pressure increments.

COMMENT

A reduction imposed at the evaporator results in a reduction of the total heat

rejected at the condenser, and reduced high-side pressures. This dynamic

must be considered because condenser airflow measurements are not

available to establish condenser reduction increments. Condenser reductions

were imposed first, to achieve compressor discharge pressures similar to the

single-fault scenarios. A condenser reduction paired with an evaporator

reduction will have lower resultant compressor discharge pressures.




Reductions were imposed at the condenser, prior to imposing any faults at the

evaporator. Condenser airflow reductions were imposed to achieve compressor

discharge pressures similar to those found in Tests 38 – 40. After the reductions at

the condenser had been established, evaporator airflow levels similar to those used

in Tests 33 – 35 were imposed. Table 18 summarizes the airflow levels and

compressor discharge pressure increments used in the evaporator and condenser

airflow reduction multiple fault tests 43-45. For comparison, details are also provided

for the single-fault evaporator airflow reduction and condenser airflow reduction fault

tests.

TABLE 18. TESTS 43-45: SUMMARY OF AIRFLOW AND COMPRESSOR DISCHARGE PRESSURES

Test #

Compressor Discharge Pressure (psig)

Evaporator Airflow (SCFM)

Evaporator Airflow Reduction

33 387 772

34 382 584

35 376 496

Condenser Airflow

Reduction

38 467 1,165

39 575 1,165

40 613 1,164

Evaporator and Condenser Airflow

Reduction

43 438 (467)* 774

44 489 (575)* 590

45 575 (613)* 494


Where

P’ = resultant compressor discharge pressure, after evaporator airflow

reduction is imposed


I. (Test 46, two faults) High Charge and Evaporator Airflow Reduction

This test scenario was specifically chosen to provide data to other parallel efforts:

fault modeling challenges exist regarding prediction of HVAC performance at more

pronounced levels of wet evaporator coils and low sub-cooling. Test 46 was

conducted with a total refrigerant charge of 10 lbs. 10 ozs. (30% high charge) and

an evaporator airflow rate of 505 SCFM (analogous to Test 35). Test 46 was

conducted at an indoor chamber air condition of 80°F/70% (DB/RH) and an outdoor

chamber air condition of 95°F.




J. (Tests 47 – 55)

The families of multiple-fault tests were categorized in the following manner:

1. Tests 47, 50, and 53 (2-fault): Low refrigerant charge and evaporator airflow

reduction increments

2. Tests 48, 51, and 54 (2-fault): Low refrigerant charge and condenser airflow

reduction increments

3. Tests 49, 52, and 55 (3-fault): Low refrigerant charge, evaporator airflow,

and condenser airflow reduction increments

However, tests were conducted in order from 47 – 55.

COMMENT

Testing in a sequence that linearly progresses through these families does not

make the most sense from a logistics standpoint (in the order: 47, 50, 53, 48,

51, 54, 49, 52, and 55). Specifically, this requires more iteration of

refrigerant recovery and addition. Taking steps to conduct tests with similar

refrigerant charge levels together avoids extra burden.

Low charge increments were imposed in the same 1 lb. 1.5 oz. increment conducted

in the single low charge fault tests: 13%, 27%, and 40% low charge. For multiple-

fault Tests 47 – 55, low charge faults were always imposed first. Evaporator airflow

reductions were always imposed last. Evaporator airflow rates were imposed at

levels similar to those of Tests 33 – 35. Condenser heat rejection was established by

floating compressor discharge pressure up to levels right before trip of the high-

pressure switch.

Table 19 summarizes the refrigerant charge, evaporator airflow rates, and

compressor discharge pressures associated with Tests 47 – 55. For Tests 49, 52, and

55, compressor discharge values change from the initially imposed values, due to the

dynamic between evaporator and condenser heat transfer. Compressor discharge

values are presented both in their initially imposed values and in their resultant

values. This is done to illustrate that the condenser airflow reductions in Tests 48,

51, and 54, are the same as those done in Tests 49, 52, and 55, respectively.




TABLE 19. TESTS 47-55: SUMMARY OF REFRIGERANT CHARGE, AIRFLOW, AND COMPRESSOR DISCHARGE PRESSURE

Test # Total Refrigerant

Charge Evaporator Airflow

(SCFM) Compressor Discharge

Pressure (psig)

47 7 lbs. 1.5 ozs.

(13% low charge) 782 382

50 6 lbs.

(27% low charge) 587 367

53 4 lbs. 14.5 ozs.

(40% low charge) 544 351

48 7 lbs. 1.5 ozs.

(13% low charge) 1161 624

51 6 lbs.

(27% low charge) 1159 618

54 4 lbs. 14.5 ozs.

(40% low charge) 1171 615

49 7 lbs. 1.5 ozs.

(13% low charge) 776 607 (624)*

52 6 lbs.

(27% low charge) 653 602 (618)*

55 4 lbs. 14.5 ozs.

(40% low charge) 507 601 (615)*


Where

P’ = resultant compressor discharge pressure, after evaporator airflow

reduction is imposed





TEST METHOD CONCLUSIONS AND

RECOMMENDATIONS Several lessons were learned in the course of developing and applying this test method. The

following highlights key findings and challenges:

FDD technologies should be fundamentally evaluated based on what diagnostic is

expected to be reported. While symptom detection is important, it should be

considered secondary.

Refrigerant-side analysis can exhibit challenges for test scenarios featuring non-

condensables or mixed phase refrigerant flow.

Future FDD tests should require FDD readings be conducted after or during the

“measurement window.”

Baseline testing should be allowed with an installed FDD, if it is not anticipated to

significantly impact HVAC performance.

o It may be prudent to explore the impacts, if any, of FDD

instrumentation/setup.

Tests of this nature should always comprehensively verify that the HVAC setup is at

the original manufacturer specification. This may include verifying that correct

evacuation and charging procedures have been followed, or maintaining cleanliness

of evaporator and condenser surfaces.

Testing should explore whether the setup encounters pressure transducer “drift,”

when adding refrigerant into a service port at/near the transducer.

Use of a ball valve to simulate liquid line restriction testing proved problematic. Valve

selection for restriction testing should consider those that feature precise control and

low losses in the fully open position, such as needle valves.

When performing evaporator or condenser airflow reduction fault testing through

restricted airflow, consider the applicability of evaporator or condenser fan power

measurements.

Severe airflow restrictions caused backflow of air to occur across the condenser,

affecting condenser inlet and outlet temperature readings. It is best to use suitable

replacements from available redundant sensors.

For non-condensables and low charge multiple fault testing, when adding refrigerant

to a pre-determined amount of nitrogen, once the sub-cooling target is exceeded,

additional refrigerant only serves to increase sub-cooling further. As a result, the

entire mixture must be recovered and set again.

For non-condensables and low charge multiple fault testing: severe performance

penalties were realized in Test 30a (0.8 ozs. N2 + 76% low charge at AHRI

conditions). This test scenario was originally established to mimic a scenario where

an HVAC system, vented to atmosphere, was not subjected to any vacuum before

being charged with refrigerant to a target sub-cooling value. Due to the extreme

nature of the performance impact, it may not be suitable to assume this as a realistic

scenario of HVAC maintenance mal-practice that would go unnoticed in the field.

When conducting multiple fault tests consisting of evaporator and condenser airflow

reductions, account for the dynamic interaction between evaporator and condenser

heat transfer. Conduct condenser airflow reductions first.

Taking steps to conduct tests with similar refrigerant charge levels together avoids

extra burden.




The resulting test method is not intended to be the final and universal solution. There is

ample opportunity to improve and enhance the test method. For example, future

investigation should be conducted to address transient impacts of faults associated with

cyclic laboratory and field testing. Additionally, it is also important to determine what

modifications are necessary to apply the method to various applications. Certain factors

determine the plausibility of testing for different entities. For example, it may be

unreasonable to expect a manufacturer to test high volumes of equipment with the same

methods an academic facility might employ to test a handful of systems. A few noteworthy

applications to explore may include:

1. A voluntary ASHRAE test standard

2. A mandatory Title-24 manufacturer test requirement

3. Evaluation purposes for utility incentive programs




APPENDIX

TABLE 20. SUMMARY: APPLICABLE CALCULATION METHODS PER TEST SCENARIO

Parameter -> Gross Cooling Capacity Heat Rejection Evaporator Airflow

Sensible Cooling

Latent Cooling Refrigerant Mass Flow

Measurement Type -> Refrigerant Refrigerant Air Refrigerant

Air & Electrical

Air Air Air Air Scale Refrigerant Refrigerant

Calculation Method ->

Enthalpy Method

Compressor Regression

Enthalpy Method

Enthalpy Method

Gross Cooling + Heat of

Compression

Measure Pressure

drop eqn

ṁ x Cp x ∆T

Psychrometric Condensate

Scale Measure Regression

1

Base

Y Y Y Y Y Y Y Y Y Y Y Y

2 Y Y Y Y Y Y Y Y Y Y Y Y








10

Low Charge

N N Y N Y Y Y Y Y Y N Y

11 N N Y N Y Y Y Y Y Y N Y




15 High Charge







Sensible Cooling



Air & Electrical



Enthalpy Method


Enthalpy Method

Enthalpy Method


Compression

Measure Pressure

drop eqn

ṁ x Cp x ∆T






20

High Charge




23

Line Restrictions






28

Non-condensables

N N Y N Y Y Y Y Y Y N N

29 N N Y N Y Y Y Y Y Y N N


30a N N Y N Y Y Y Y Y Y N N



33


Y Y Y Y Y N Y Y Y Y Y Y

34 Y Y Y Y Y N Y Y Y Y Y Y








Sensible Cooling



Air & Electrical



Enthalpy Method


Enthalpy Method

Enthalpy Method


Compression

Measure Pressure

drop eqn

ṁ x Cp x ∆T



38

Condenser Airflow Reduction






43 Evaporator and





46 High Charge and



47 Low Charge &


N N Y N Y N Y Y Y Y N Y

50 N N Y N Y N Y Y Y Y N Y


48 Low Charge &





49 Low Charge, Evaporator and Condenser Airflow

Reduction

N N Y N Y N Y Y Y Y N Y






REFERENCES

1 Air-Conditioning, Heating and Refrigeration Institute (2008), 2008 Standard for Performance Rating

of Unitary Air-Conditioning & Air-Source Heat Pump Equipment.

2 Architectural Energy Corporation (December 15 2007), Advanced Automated HVAC Fault Detection and Diagnostics Commercialization Program. California Energy Commission Contract # 500-03-030. Project 4: Advanced Packaged Rooftop Unit. Deliverable D4.6b. ARTU Performance Test

Report.

3 Design & Engineering Services, Customer Service Business Unit, Southern California Edison (January 13, 2009), HVAC Maintenance ETO 08.02,03,04,05,06,07,08, and 09.

iv Incropera, DeWitt, Bergman, Lavine, (2007), Introduction to Heat Transfer. 5th Edition, John Wiley & Sons.

v Southern California Edison, Proctor Engineering Group, Ltd., Bevilacqua-Knight, Inc. (July 2008),

Energy Performance of Hot, Dry Optimized Air-Conditioning Systems. http://www.energy.ca.gov/2008publications/CEC-500-2008-056/CEC-500-2008-056.PDF

vi California Energy Commission. Reference Appendices for the 2008 Building Energy Efficiency Standards for Residential and Non-Residential Buildings. http://www.energy.ca.gov/title24/2008standards/

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