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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
sean.gouw@sce.com.
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.
Development of a FDD Laboratory Test Method for a Residential Split System HT.11.SCE.003
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.
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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
Development of a FDD Laboratory Test Method for a Residential Split System HT.11.SCE.003
Southern California Edison Page v
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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
Development of a FDD Laboratory Test Method for a Residential Split System HT.11.SCE.003
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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
Development of a FDD Laboratory Test Method for a Residential Split System HT.11.SCE.003
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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
Development of a FDD Laboratory Test Method for a Residential Split System HT.11.SCE.003
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
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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
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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
Development of a FDD Laboratory Test Method for a Residential Split System HT.11.SCE.003
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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.
Development of a FDD Laboratory Test Method for a Residential Split System HT.11.SCE.003
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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
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Design & Engineering Services July 2012
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
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( < 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.
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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).
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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.
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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.
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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.
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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”.
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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.
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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
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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.
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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.
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FIGURE 7. INDOOR HVAC UNIT SENSOR PLACEMENT
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FIGURE 8. CONDENSING UNIT SENSOR PLACEMENT
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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
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ASHRAE 37-2009 Laboratory Measurements FDD Sensors
Measurement Units Name/Description Notes Name/Description Notes
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 - -
HD Cond Air OUT Temp 2
HD Cond Air OUT Temp 3
HD Cond Air OUT Temp 4
HD Cond Air OUT Temp 5
HD Cond Air OUT Temp 6
HD Cond Air OUT Temp 7
HD Cond Air OUT Temp 8
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) - -
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ASHRAE 37-2009 Laboratory Measurements FDD Sensors
Measurement Units Name/Description Notes Name/Description Notes
- 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 - -
- Rm 4 on Pole at 8ft - -
- Rm 4 on Pole at 7ft - -
- Rm 4 on Pole at 6ft - -
- Rm4 Pole at 5ft - -
- Rm 4 on Pole at 4ft - -
- Rm 4 on Pole at 3ft - -
- Rm 4 on Pole at 2ft - -
- Rm 4 on Pole at 1ft - -
- Rm 4 on Pole at Floor - -
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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
(SCE’s Metrology Lab)
Evap fan inlet
Relative Humidity (RH)
Vaisala, HMP 247 ± (0.5 + 2.5% of
reading)% RH
5-9-2011
(SCE’s Metrology Lab)
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
± 0.06% of full scale
4-14-2011
(Tektronix Calibration Lab)
Across indoor unit
Power Ohio Semitronics,
GW5-002C
± 0.2% of reading
± 0.04% of full scale
(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
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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
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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
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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
= Enthalpy at refrigerant-side state point R8, Btu/lb
EQUATION 3. REFRIGERANT-SIDE CONDENSER HEAT REJECTION
Where:
= Refrigerant-side heat rejection, Btu/hr
= Refrigerant mass flow rate, lbs/hr
= Enthalpy at refrigerant-side state point R4, Btu/lb
= Enthalpy at refrigerant-side state point R5, Btu/lb
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.
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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%.
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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
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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.
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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
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EQUATION 8. AIR-SIDE GROSS SENSIBLE COOLING CAPACITY
Where
= Air-side gross sensible cooling capacity, Btu/hr
= Air flow rate (measured or calculated), lbs/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
= Air flow rate (measured or calculated), lbs/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:
= Air flow rate (measured or calculated), lbs/hr
= 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
Development of a FDD Laboratory Test Method for a Residential Split System HT.11.SCE.003
Southern California Edison Page 28
Design & Engineering Services July 2012
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%.
Development of a FDD Laboratory Test Method for a Residential Split System HT.11.SCE.003
Southern California Edison Page 29
Design & Engineering Services July 2012
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%.
Development of a FDD Laboratory Test Method for a Residential Split System HT.11.SCE.003
Southern California Edison Page 30
Design & Engineering Services July 2012
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)
Development of a FDD Laboratory Test Method for a Residential Split System HT.11.SCE.003
Southern California Edison Page 31
Design & Engineering Services July 2012
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%
Development of a FDD Laboratory Test Method for a Residential Split System HT.11.SCE.003
Southern California Edison Page 32
Design & Engineering Services July 2012
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
Test # Description Indoor Chamber
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
19 High Refrigerant Charge – 30% 70
◦F /59
◦F /52%
(DB/WB/RH) 75
◦F DB
20 High Refrigerant Charge – 30% 70
◦F /63.3
◦F /70%
(DB/WB/RH) 85
◦F DB
Development of a FDD Laboratory Test Method for a Residential Split System HT.11.SCE.003
Southern California Edison Page 33
Design & Engineering Services July 2012
Test # Description Indoor Chamber
Air Condition
Outdoor Chamber
Air Condition
21 High Refrigerant Charge – 30% 75
◦F /67.9
◦F /70%
(DB/WB/RH) 85
◦F DB
22 High Refrigerant Charge – 30% 75
◦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
40 Condenser Airflow Reduction – 613 psig Compressor Discharge Pressure
41 Condenser Airflow Reduction – 622 psig Compressor Discharge Pressure
75◦F /63
◦F /52%
(DB/WB/RH) 115
◦F DB
42 Condenser Airflow Reduction – 612 psig Compressor Discharge Pressure
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.
Development of a FDD Laboratory Test Method for a Residential Split System HT.11.SCE.003
Southern California Edison Page 34
Design & Engineering Services July 2012
TABLE 13. MULTIPLE-FAULT TEST SCENARIOS
Test # Description Indoor Chamber
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*
Fault 1: Evaporator Airflow Reduction – 51%
Fault 2: Condenser Airflow Reduction – 489 (575) psig Compressor Discharge Pressure
45*
Fault 1: Evaporator Airflow Reduction – 59%
Fault 2: Condenser Airflow Reduction – 575 (613) psig Compressor Discharge Pressure
46 Fault 1: High Refrigerant Charge – 30%
Fault 2: Evaporator Airflow Reduction – 58%
80◦F /72.4◦F /70% (DB/WB/RH)
95◦F DB
47 Fault 1: Low Refg Charge - 13%
Fault 2: Evaporator Airflow Reduction – 35%
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%
Fault 2: Evaporator Airflow Reduction – 35%
Fault 3: Condenser Airflow Reduction – 607 (624) psig Compressor Discharge Pressure
50 Fault 1: Low Refg Charge – 27%
Fault 2: Evaporator Airflow Reduction – 51%
80◦F /67◦F /51% (DB/WB/RH)
95◦F DB
51
Fault 1: Low Refg Charge – 27%
Fault 2: Condenser Airflow Reduction – 618 psig Compressor Discharge Pressure
52
Fault 1: Low Refg Charge – 27%
Fault 2: Evaporator Airflow Reduction – 46%
Fault 3: Condenser Airflow Reduction – 602 (618) psig Compressor Discharge Pressure
53 Fault 1: Low Refg Charge – 40%
Fault 2: Evaporator Airflow Reduction – 55%
80◦F /67◦F /51% (DB/WB/RH)
95◦F DB
54
Fault 1: Low Refg Charge – 40%
Fault 2: Condenser Airflow Reduction – 615 psig Compressor Discharge Pressure
55
Fault 1: Low Refg Charge – 40%
Fault 2: Evaporator Airflow Reduction – 58%
Fault 3: Condenser Airflow Reduction – 601 (615) psig Compressor Discharge Pressure
*Note: Compressor discharge pressure is presented in the form, P’ (P)
Development of a FDD Laboratory Test Method for a Residential Split System HT.11.SCE.003
Southern California Edison Page 35
Design & Engineering Services July 2012
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.
Development of a FDD Laboratory Test Method for a Residential Split System HT.11.SCE.003
Southern California Edison Page 36
Design & Engineering Services July 2012
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.
Development of a FDD Laboratory Test Method for a Residential Split System HT.11.SCE.003
Southern California Edison Page 37
Design & Engineering Services July 2012
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
Development of a FDD Laboratory Test Method for a Residential Split System HT.11.SCE.003
Southern California Edison Page 38
Design & Engineering Services July 2012
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).
Development of a FDD Laboratory Test Method for a Residential Split System HT.11.SCE.003
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Design & Engineering Services July 2012
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.
Development of a FDD Laboratory Test Method for a Residential Split System HT.11.SCE.003
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Design & Engineering Services July 2012
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
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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)
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# 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.
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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%
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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
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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.
Development of a FDD Laboratory Test Method for a Residential Split System HT.11.SCE.003
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Design & Engineering Services July 2012
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
Development of a FDD Laboratory Test Method for a Residential Split System HT.11.SCE.003
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Design & Engineering Services July 2012
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
determined to be 50 psig
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
Development of a FDD Laboratory Test Method for a Residential Split System HT.11.SCE.003
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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.
Development of a FDD Laboratory Test Method for a Residential Split System HT.11.SCE.003
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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).
Development of a FDD Laboratory Test Method for a Residential Split System HT.11.SCE.003
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Design & Engineering Services July 2012
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
determined to be 50 psig
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
Development of a FDD Laboratory Test Method for a Residential Split System HT.11.SCE.003
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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.
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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
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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%
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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.
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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.
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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
*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
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.
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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.
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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)*
*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
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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.
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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
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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
3 Y Y Y Y Y Y Y Y Y Y Y Y
4 Y Y Y Y Y Y Y Y Y Y Y Y
5 Y Y Y Y Y Y Y Y Y Y Y Y
6 Y Y Y Y Y Y Y Y Y Y Y Y
7 Y Y Y Y Y Y Y Y Y Y Y Y
8 Y Y Y Y Y Y Y Y Y Y Y Y
9 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
12 N N Y N Y Y Y Y Y Y N Y
13 N N Y N Y Y Y Y Y Y N Y
14 N N Y N Y Y Y Y Y Y N Y
15 High Charge
Y Y Y Y Y Y Y Y Y Y Y Y
16 Y Y Y Y Y Y Y Y Y Y Y Y
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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
17 Y Y Y Y Y Y Y Y Y Y Y Y
18 Y Y Y Y Y Y Y Y Y Y Y Y
19 Y Y Y Y Y Y Y Y Y Y Y Y
20
High Charge
Y Y Y Y Y Y Y Y Y Y Y Y
21 Y Y Y Y Y Y Y Y Y Y Y Y
22 Y Y Y Y Y Y Y Y Y Y Y Y
23
Line Restrictions
N N Y N Y Y Y Y Y Y N Y
24 N N Y N Y Y Y Y Y Y N Y
25 N N Y N Y Y Y Y Y Y N Y
26 N N Y N Y Y Y Y Y Y N Y
27 N N Y N Y Y Y Y Y Y N Y
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
30 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
31 N N Y N Y Y Y Y Y Y N N
32 N N Y N Y Y Y Y Y Y N N
33
Evaporator Airflow Reduction
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
35 Y Y Y Y Y N Y Y Y Y Y Y
36 Y Y Y Y Y N Y Y Y Y Y Y
37 Y Y Y Y Y N Y Y Y Y Y Y
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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
38
Condenser Airflow Reduction
Y Y Y Y Y Y Y Y Y Y Y Y
39 Y Y Y Y Y Y Y Y Y Y Y Y
40 Y Y Y Y Y Y Y Y Y Y Y Y
41 Y Y Y Y Y Y Y Y Y Y Y Y
42 Y Y Y Y Y Y Y Y Y Y Y Y
43 Evaporator and
Condenser Airflow Reduction
Y Y Y Y Y N Y Y Y Y Y Y
44 Y Y Y Y Y N Y Y Y Y Y Y
45 Y Y Y Y Y N Y Y Y Y Y Y
46 High Charge and
Evaporator Airflow Reduction
Y Y Y Y Y N Y Y Y Y Y Y
47 Low Charge &
Evaporator Airflow Reduction
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
53 N N Y N Y N Y Y Y Y N Y
48 Low Charge &
Condenser Airflow Reduction
N N Y N Y Y Y Y Y Y N Y
51 N N Y N Y Y Y Y Y Y N Y
54 N N Y N Y Y Y Y Y Y N Y
49 Low Charge, Evaporator and Condenser Airflow
Reduction
N N Y N Y N Y Y Y Y N Y
52 N N Y N Y N Y Y Y Y N Y
55 N N Y N Y N Y Y Y Y N Y
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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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