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NIST Technical Note 1848

Sensitivity Analysis of Installation 

Faults on Heat Pump Performance

Piotr A. Domanski

Hugh I. Henderson

W. Vance Payne

http://dx.doi.org/10.6028/NIST.TN.1848
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NIST Technical Note 1848

Sensitivity Analysis of Installation 

Faults on Heat Pump Performance

Piotr A. Domanski

Energy and Environment Division

Engineering Laboratory

Hugh I. Henderson

CDH Energy Corporation

Cazenovia, NY

W. Vance Payne

Energy and Environment Division

Engineering Laboratory

This publication is available free of charge from:

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September 2014

U.S. Department of Commerce 

Penny Pritzker, Secretary

National Institute of Standards and Technology 

Willie E. May, Acting Under Secretary of Commerce for Standards and Technology and Acting Director 
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Certain commercial entities, equipment, or materials may be identified in this

document in order to describe an experimental procedure or concept adequately.

Such identification is not intended to imply recommendation or endorsement by the

National Institute of Standards and Technology, nor is it intended to imply that the

entities, materials, or equipment are necessarily the best available for the purpose. 

National Institute of Standards and Technology Technical Note 1848

Natl. Inst. Stand. Technol. Tech. Note 1848, 104 pages (September 2014)

 CODEN: NTNOEF

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iii

Sensitivity Analysis of Installation Faults on Heat Pump Performance

Piotr A. Domanski(a), Hugh I. Henderson(b), W. Vance Payne(a)

(a) National Institute of Standards and Technology, Gaithersburg, MD 20899-8631

(b) CDH Energy Corporation, Cazenovia, NY 13035-0641

ABSTRACT

Numerous studies and surveys indicate that typically-installed HVAC equipment operate inefficiently and 

waste considerable energy due to different installation errors (faults) such as improper refrigerant charge, 

incorrect airflow, oversized equipment, leaky ducts. This study seeks to develop an understanding of the

impact of different faults on heat pump performance installed in a single-family residential house. It 

combines building effects, equipment effects, and climate effects in a comprehensive evaluation of the 

impact of installation faults on a heat pump’s seasonal energy consumption through simulations of the 

house/heat pump system.

The study found that duct leakage, refrigerant undercharge, oversized heat pump with nominal ductwork, 

low indoor airflow due to undersized ductwork, and refrigerant overcharge have the most potential for 

causing significant performance degradation and increased annual energy consumption. The effect of 

simultaneous faults was found to be additive (e.g., duct leakage and non-condensable gases), little changed 

relative to the single fault condition (e.g., low indoor airflow and refrigerant undercharge), or well-beyond 

additive (duct leakage and refrigerant undercharge). A significant increase in annual energy use can be 

caused by lowering the thermostat in the cooling mode to improve indoor comfort in cases of excessive 

indoor humidity levels due to installation faults.

The goal of this study was to assess the impacts that HVAC system installation faults had on equipment 

electricity consumption. The effect of the installation faults on occupant comfort was not the main focus 

of the study, and this research did not seek to quantify any impacts on indoor air quality or noise 

generation (e.g., airflow noise from air moving through restricted ducts). Additionally, the study does not 

address the effects that installation faults have on equipment reliability/robustness (number of starts/stops, 

etc.), maintainability (e.g., access issues), or costs of initial installation and ongoing maintenance.
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TABLE OF CONTENTS

ABSTRACT ........................................................................................................................................... iii

TABLE OF CONTENTS .................................................................................................................... iv

LIST OF FIGURES ............................................................................................................................. vi

LIST OF TABLES ................................................................................................................................ viii

1. INTRODUCTION........................................................................................................................ 1

2. LITERATURE SURVEY........................................................................................................... 3

2.1 Field Surveys, Installation and Maintenance Issues .............................................................. 3

2.2 Heat Pump Oversizing, Undersizing and Part-load Losses.................................................... 5

2.3 Laboratory Studies of Performance Degradation of Heat Pumps Due to Faults.................... 6

3. HEAT PUMP PERFORMANCE DEGRADATION DUE TO FAULTS...................... 8 

3.1. Laboratory Measurements ...................................................................................................... 8

3.1.1 Experimental Apparatus and Test Conditions .............................................................. 8

3.1.2 Studied Faults and Their Implementation ..................................................................... 9

3.2 Fault Effects on Cooling Mode Performance ......................................................................... 11

3.2.1 Cooling Mode Normalized Performance Parameters and Correlation ......................... 11

3.2.2 Cooling Mode Charts with Normalized Performance Parameters ............................... 14

3.3 Fault Effects on Heating Mode Performance ......................................................................... 23

3.3.1 Heating Mode Normalized Performance Parameters and Correlation ......................... 23

3.3.2 Heating Mode Charts with Normalized Performance Parameters ................................ 23

4. BUILDING/HEAT PUMP MODELING APPROACH...................................................... 32

4.1. Building/Heat Pump Systems Simulation Models................................................................. 32

4.2 Building and Weather City Definitions ................................................................................. 34

4.3 Building and Enclosure Thermal Details ............................................................................... 35

4.3.1 Building Enclosure Air Leakage ................................................................................... 40

4.3.2 Duct Leakage and Thermal Losses................................................................................ 40

4.3.3 Moisture and Thermal Gains......................................................................................... 40

4.3.4 Moisture and Thermal Capacitance............................................................................... 40

4.3.5 Window Performance.................................................................................................... 41

4.4 Mechanical Ventilation ........................................................................................................... 41

4.5 Airflow Imbalance .................................................................................................................. 42

4.6 Heat Pump Specifications and Modeling................................................................................ 42

4.7 Cost of Electricity ................................................................................................................... 44

 

5. SIMULATIONS OF BUILDING/HEAT PUMP SYSTEM WITH INSTALLATION

 FAULTS ........................................................................................................................................ 45

5.1 Annual Energy Consumption in Baseline Houses ..........................................................................45

5.2 Simulations with Single Faults........................................................................................................46

5.2.1 Studied Faults .. .....................................................................................................................46 

5.2.2 Effect of Heat Pump Sizing ..................................................................................................46 

5.2.3 Effect of Duct Leakage .........................................................................................................54 

5.2.4 Effect of Indoor Coil Airflow ...............................................................................................60 

5.2.5 Effect of Refrigerant Undercharge ........................................................................................64 

5.2.6 Effect of Refrigerant Overcharge ..........................................................................................66 

5.2.7 Effect of Excessive Refrigerant Subcooling .........................................................................67 

5.2.8 Effect of Non-Condensable Gases ........................................................................................68 
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5.2.9 Effect of Voltage ...................................................................................................................69

5.2.10 Effect of TXV Sizing............................................................................................................71

 5.2.11 Discussion of the Effects of Single Faults ...........................................................................72

 5.3 Simulations with Dual Faults ........................................................................................................74

 5.3.1 Studied Fault Combinations...................................................................................................74

5.3.2 Effects of Dual Faults ............................................................................................................75 

5.3.3 Discussion of the Effects of Dual Faults ...............................................................................81 

 5.4 Effects of Triple Faults .................................................................................................................82

6. CONCLUDING REMARKS .................................................................................................... 83

7. NOMENCLATURE.................................................................................................................... 84

8. REFERENCES ............................................................................................................................ 85

ACKNOWLEGEMENTS ................................................................................................................... 92

APPENDIX A: DUCT LOSSES ....................................................................................................... 93
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LIST OF FIGURES

3.1 Schematic diagram of experimental apparatus (Kim et al. (2006)) ..................................................8 

3.2 Normalized performance parameters for the cooling mode TXV undersizing fault: 

 (a) capacity, (b) COP ......................................................................................................................14 

3.3 Normalized cooling performance parameters for improper indoor airflow ....................................17 

3.4 Normalized cooling performance parameters for refrigerant undercharge ....................................18

3.5 Normalized cooling performance parameters for refrigerant overcharge .......................................19 

3.6 Normalized cooling performance parameters for liquid line refrigerant subcooling .....................20 

3.7 Normalized cooling performance parameters for the presence of non-condensable gas ................21 

3.8 Normalized cooling performance parameters for improper electric line voltage ...........................22

3.9 Normalized heating performance parameters for improper indoor airflow ....................................26

3.10 Normalized heating performance parameters for refrigerant undercharge .....................................27

3.11 Normalized heating performance parameters for refrigerant overcharge .......................................28

3.12 Normalized heating performance parameters for improper refrigerant subcooling .......................29

3.13 Normalized heating performance parameters for the presence of non-condensable gas ................30

3.14 Normalized heating performance parameters for improper line voltage ........................................31

4.1 Screen shot of TRNBuild used to define the building envelope details .........................................34

4.2 IECC climate zone map ..................................................................................................................35

4.3 Schematic of a slab-on-grade house ................................................................................................37

4.4 Schematic of a house with basement ..............................................................................................38 

4.5 Schematic of a mechanical exhaust system.....................................................................................41

4.6 Capacity degradation due to defrost as a function of outdoor temperature.....................................44

5.1 Annual energy use for slab-on-grade houses for different heat pump sizings, scenario (2) ...........53

5.2 Annual energy use for houses with basement for different heat pump sizings, scenario (2) ..........54

5.3 Number of hours above 55 % relative humidity for a slab-on-grade house in Houston with duct 

 leak rates from 10 % to 50 % at three thermostat set point temperatures........................................57

5.4 Energy use for a slab-on-grade house in Houston with duct leak rates from 10 % to 50 % at 

 three thermostat set point temperatures related to energy use for the house at the default set 

 point and 10 % leak rate ..................................................................................................................58

5.5 Annual energy use for slab-on-grade houses for different indoor coil airflows relative to energy 

 use for the house in the same location with nominal airflow rate ...................................................60

5.6 Annual energy use for slab-on-grade houses at different level of refrigerant undercharge

 relative to the annual energy use for the house in the same location when the heat pump 

operates with the nominal refrigerant charge .................................................................................65

5.7 Annual energy use for slab-on-grade houses at different level of refrigerant overcharge 

 relative to the annual energy use for the house in the same location when the heat pump 

operates with the nominal refrigerant charge ..................................................................................67

5.8 Annual energy use for slab-on-grade houses at different level of refrigerant subcooling relative

 to the annual energy use for the house in the same location with the heat pump operating with 

the nominal refrigerant charge and subcooling ..............................................................................68

5.9 Annual energy use for slab-on-grade houses at different levels of input voltages relative to

The energy use for the house in the same location when the heat pump operates with nominal

voltage .............................................................................................................................................70

5.10 Annual energy use for slab-on-grade houses at different levels of TXV undersizing relative to

 the annual energy use for the house when the heat pump operates with a properly sized TXV.....72

5.11 Annual energy use by a heat pump in a slab-on-grade house resulting from a single-fault 

 installation relative to a fault-free installation. ...............................................................................72

5.12 Annual energy use for slab-on-grade houses with 14 dual-faults referenced to the energy use for 

 the house with fault-free installation ..............................................................................................81
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5.13. Annual energy use for houses with basement with 8 dual-fault installations referenced to energy

 use for the house with fault-free installation....................................................................................82

A1. Schematic representation of duct leakage in a home with attic ducts .............................................93
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viii

LIST OF TABLES

2.1 Selected studies on faults detection and diagnosis ...............................................................................6

3.1 Cooling and heating test temperatures .................................................................................................9

3.2 Measurement uncertainties ..................................................................................................................9

3.3 Definition and range of studied faults.................................................................................................10

3.4 Correlations for non-dimensional performance parameters in the cooling mode ..............................12

3.5 Example uncertainty propagation with normalized correlation (Y) uncertainty of 3 %

 for faulty COP and cooling capacity at AHRI Standard 210/240 B-test condition ............................12

3.6 Normalized capacity and COP correlation coefficients for a TXV undersizing fault.........................13

3.7 Correlations for non-dimensional performance parameters in the heating mode ..............................24

4.1 Comparison of residential building simulation software tools............................................................32

4.2 Comparison of general building calculation models ..........................................................................33

4.3 Climates, locations and structures considered ....................................................................................35

4.4 Specifications for simulated houses (HERS Index ≈100)...................................................................36

4.5 Calculation of R-values for basement walls and floor........................................................................39

4.6 Calculation of R-values for slab-on-grade floor .................................................................................39

4.7 Heat pump cooling characteristics......................................................................................................42

4.8 Thermostat cooling and heating set points..........................................................................................44

4.9 Cost of electricity................................................................................................................................44

5.1 Energy consumption and cost in baseline houses...............................................................................46

5.2 Studied faults in the cooling and heating mode .................................................................................46

5.3 Indoor airflow information for heat pump sizing scenario (1) and scenario (2) ................................48

5.4 Effect of 100 % unit oversizing on annual energy use for a slab-on-grade house for scenario (1)

 and scenario (2). .................................................................................................................................49

5.5 Effect of heat pump sizing on annual energy use for a slab-on-grade house with duct sized to 

 match heat pump size (scenario (1)) .................................................................................................50

5.6 Effect of heat pump sizing on annual energy use for a house with basement with duct sized to 

 match heat pump size (scenario (1)) .................................................................................................51 

5.7 Effect of heat pump sizing on annual energy use for a slab-on-grade house with fixed duct 

 size (scenario (2)) ..............................................................................................................................52

5.8 Effect of heat pump sizing on annual energy use for a house with basement with fixed duct 

 size (scenario (2)) ..............................................................................................................................53

5.9 Effect of duct leakage on annual energy use for a slab-on-grade house at default cooling set point 55

5.10 Effect of duct leakage on annual energy use for a slab-on-grade house at lowered cooling

 set point by 1.1 °C (2.0 °F) ..............................................................................................................56

5.11 Effect of duct leakage on annual energy use for a slab-on-grade house in Houston at lowered 

 cooling set point by 2.2 °C (4.0 °F) ..................................................................................................57

5.12 Effect of lowering cooling set point by 1.1 °C (2.0 °F) on annual energy use of a baseline

 slab-on-grade house and a house with basement .............................................................................59

5.13 Effect of indoor coil airflow on annual energy use for a slab-on-grade house when operating at 

 the default cooling set point..............................................................................................................61 

5.14 Effect of indoor coil airflow on annual energy use for a house with basement when operating 

 at the default cooling set point..........................................................................................................62

5.15 Effect of indoor coil airflow on annual energy use for a slab-on-grade house when operating 

 at a cooling set point that is 1.1 °C (2.0 °F) lower than the default value .......................................63 

5.16 Effect of indoor coil airflow on annual energy use for a house with basement when operating at 

 cooling set point that is 1.1 °C (2.0 °F) lower than the default value .............................................64 

5.17 Effect of refrigerant undercharge on annual energy use for a slab-on-grade house..........................65

5.18 Effect of refrigerant undercharge on annual energy use for a house with basement ........................65

5.19 Effect of refrigerant overcharge on annual energy use for a slab-on-grade house ...........................66
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5.20 Effect of refrigerant overcharge on annual energy use for a house with basement ..........................66

5.21 Effect of excessive refrigerant subcooling on annual energy use for a slab-on-grade house ...........67

5.22 Effect of excessive refrigerant subcooling on annual energy use for a house with basement .........68

5.23 Effect of non-condensable gases on annual energy use for a slab-on-grade house ..........................69

5.24 Effect of non-condensable gases on annual energy use for a house with basement .........................69

5.25 Effect of voltage on annual energy use for a slab-on-grade house ...................................................70

5.26 Effect of voltage on annual energy use for a house with basement ..................................................70

5.27 Effect of TXV sizing on annual energy use for a slab-on-grade houses..........................................71

5.28 Effect of TXV sizing on annual energy use for a house with basement...........................................71

5.29 Levels of individual faults used in Figure 5.11.................................................................................73

5.30 Combinations of studied faults ........................................................................................................74

5.31 Dual fault sets considered in simulations (heating and cooling) and their approximate collective 

 effect of energy use..........................................................................................................................74

5.32 Dual fault sets considered in simulations (heating and cooling) and their approximate collective 

 effect on annul energy use; TXV fault existing in cooling only......................................................75

5.33 Relative energy use for dual fault sets 1 to 5 for the slab-on-grade house in Houston ....................75

5.34 Relative energy use for dual fault sets 6 to 8 for the slab-on-grade house in Houston.....................76

5.35 Relative energy use for dual fault sets 9 to 11 for the slab-on-grade house in Houston...................76

5.36 Relative energy use for dual fault sets 12 to 14 involving cooling mode TXV

 for the slab-on-grade house in Houston ............................................................................................76

5.37 Relative energy use for dual fault sets 1 to 5 for the slab-on-grade house in Washington, DC .......77

5.38 Relative energy use for dual fault sets 6 to 8 for the slab-on-grade house in Washington, DC........77

5.39 Relative energy use for dual fault sets 9 to 11 for the slab-on-grade house in Washington, DC......77

5.40 Relative energy use for dual fault sets 12 to 14 involving cooling mode TXV

 for the slab-on-grade house in Washington, DC...............................................................................78

5.41 Relative energy use for dual fault sets 1 to 5 for the slab-on-grade house in Minneapolis .............78

5.42 Relative energy use for dual fault sets 6 to 8 for the slab-on-grade house in Minneapolis ..............78

5.43 Relative energy use for dual fault sets 9 to 11 for the slab-on-grade house in Minneapolis ............79

5.44 Relative energy use for dual fault sets 12 to 14 involving cooling mode TXV

 for the slab-on-grade house in Minneapolis......................................................................................79

5.45 Relative energy use for dual fault sets 6 to 8 for the basement house in Washington, DC ..............79

5.46 Relative energy use for dual fault sets 9 to 11 for the basement house in Washington, DC ............80

5.47 Relative energy use for dual fault sets 13 to 14 involving cooling mode TXV

 for the basement house in Washington, DC......................................................................................80
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1. INTRODUCTION

Space cooling is responsible for the largest share (at 21.3 %) of the electrical energy consumption in the 

U.S. residential sector (DOE, 2011). Space heating, for which a significant portion is provided by heat 

pumps, accounts for an additional 8.7 % electricity use. Consequently, there are increasing requirements 

that space-conditioning equipment be highly efficient to improve building energy efficiency as well as

address environmental concerns. To this end, state and municipal governments and utility partners have 

implemented various initiatives that promote sales of high-efficiency air conditioners (ACs) and heat 

pumps (HPs). However, there is a growing recognition that merely increasing equipment’s laboratorymeasured efficiency without ensuring that the equipment is installed and operated correctly in the field is 

ineffective. A key component for maximizing field equipment performance is to ensure that such 

equipment is sized, selected, and installed following industry recognized procedures. Consistent with this 

goal, the Air Conditioning Contractors of America (ACCA) released in 2007 a quality installation (QI) 

standard for heating, ventilating and air-conditioning (HVAC) equipment, which has been updated since 

then and achieved widespread recognition by various entities in the U.S. concerned with reducing energy 

consumption by buildings (ACCA, 2010). A companion standard (ACCA, 2011b) defines the verification 

protocols to ensure that HVAC systems have been installed according to the QI Standard. A related 

ACCA standard (ACCA, 2013) addresses residential maintenance issues.

Numerous studies and surveys indicate that typically-installed HVAC equipment operate inefficiently and 

waste considerable energy due to different installation errors (faults) such as improper refrigerant charge, 

incorrect airflow, oversized equipment, leaky ducts. However, it is unclear whether the effects of such 

faults are additive, whether small variances within a given fault type are significant, and which faults (in 

various applications and geographical locations) have a larger impact than others. If this information is 

known, better attention, resources, and effort can be focused on the most important design, installation, 

and maintenance parameters. 

This project seeks to develop an understanding of the impact of different commissioning parameters on 

heat pump performance for a single-family residential house application. It combines building effects, 

equipment effects, and climate effects in a comprehensive evaluation of the impact of installation faults 

on seasonal energy consumption of a heat pump through simulations of the house/heat pump system. The 

evaluated commissioning parameters include:

• Building subsystem

- Duct leakage (unconditioned space)

• Residential split, air-to-air heat pump equipped with a thermostatic expansion valve (TXV)

- Equipment sizing

- Indoor coil airflow

- Refrigerant charge

- Presence of non-condensable gases

- Electrical voltage

- TXV undersizing

• Climates (cooling and heating)

- Hot and humid

- Hot and dry

- Mixed

- Heating dominated

- Cold

• Single-family houses (the structures representative for the climate)

- House on a slab

- House with a basement.
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The goal of this study is to assess the impacts that HVAC system installation faults have on equipment 

electricity consumption. The effect of the installation faults on occupant comfort is not the main focus of 

the study, and this research did not seek to quantify any impacts on indoor air quality or noise generation 

(e.g., airflow noise from air moving through restricted ducts). Additionally, the study does not address 

the effects that installation faults have on equipment reliability/robustness (number of starts/stops, etc.), 

maintainability (e.g., access issues), or costs of initial installation and ongoing maintenance.
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2. LITERATURE SURVEY

The literature survey is presented in three sections. Section 2.1 presents selected publications related to air 

conditioner and heat pump installation and maintenance issues, Section 2.2 focuses on heat pump 

oversizing/undersizing and cycling loses, and Section 2.3 presents relevant studies on heat pump fault 

detection and diagnostics (FDD).

2.1 Field Surveys, Installation and Maintenance Issues

Numerous field studies have documented degraded performance and increased energy usage for typical

air conditioners and heat pumps installed in the United States. Commonly, system efficiency, peak 

electrical demand, and comfort are compromised. This degraded performance has been linked to several 

problems, which include:

- improperly designed, insulated, or balanced air distribution systems in the house

- improperly selected heat pump, either by the fact of overall performance characteristics due to mixmatched components or improper capacity (too large or too small) in relation to the building load 

- heat pump operating with a fault.

The first two problem categories are a result of negligent or incompetent work prior to the heat pump 

installation. The third problem category, a heat pump operating with a fault, can be a result of improper 

installation or improper maintenance. Field study reports describing observations and measurements on 

new installations are less common than publications on existing installations. For this reason, in this 

literature review we also include reports on maintenance practices, in particular those covering large 

numbers of systems. 

While discussing heat pump performance measurements taken in the field, we have to recognize that 

these field measurements offer significant challenges and are burdened by a substantial measurement 

uncertainty, much greater than the uncertainty of measurements in environmental chambers, which are in 

the order of 5 % at the 95 % confidence level. Typically, field study reports do not estimate the 

measurement uncertainty of the reported values; however, the number of installations covered by some of 

these studies provides an informative picture about the scope and extent of field installation problems. We 

may also note that most of the articles on field surveys are not published in indexed journals. 

Consequently, they are not searchable by publication search engines, and many of them are not readily 

available. In this literature review, we gave a preference to citing publications which can be readily 

obtained by a reader if desired.

In a study of new installations, Proctor (1997) performed measurements on a sample of 28 air 

conditioners installed in 22 residential homes in a hot and dry climate (Phoenix, AR, USA). Indoor heat 

exchanger airflow averaged 14 % below specifications, and only 18 % of the systems had a correct 

amount of refrigerant. The supply duct leakage averaged 9 % of the air handler airflow, and the return 

leakage amounted to 5 %. The author cites several prior publications, which reported similar problems. 

Davis and Robison (2008) monitored seven new high efficiency residential heat pumps. They diagnosed 

several installation errors, which included a malfunctioning TXV, non-heat pump thermostat installed, 

incorrect indoor unit installed, and incorrect control wiring preventing proper system staging. The 

authors reported that once the problems were repaired, the systems performed at the expected levels. 

Parker et al. (1997) investigated the impact of indoor airflow on residential air conditioners in 27 

installations in Florida. They measured airflows ranging from 62.8 m3

∙h-1∙kW-1 to 246.4 m3∙h-1∙kW-1

(130 cfm/ton to 510 cfm/ton) while a typical manufacturer’s recommendation calls for 193.2 m3

∙h-1∙kW-1

(400 cfm/ton). Undersized return ducts and grills, improper fan speed settings, and fouled filters were the 

causes of improper airflow along with duct runs that were long, circuitous, pinched or constricted. 

Additional flow resistance can result from the homeowner tendency to increase air filtration via higher 
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efficiency filters during replacement; the measurements showed that substitution of high-efficiency filters 

typically reduces the airflow by 5 %. Low airflow has system energy-efficiency implications; reduction of 

airflow by 25 % from 193.2 m3

∙h-1∙kW-1 to 144.9 m3∙h-1∙kW-1 (400 cfm/ton to 300 cfm/ton) can reduce the 

efficiency of the air conditioner by 4 %. The authors commented that airflows below 169.1 m3

∙h-1∙kW-1

(350cfm/ton) render invalid most field methods for determining refrigerant charge and can lead to 

improper charging by a service technician who often does not check the evaporator airflow. 

Downey and Proctor (2002) reported on the field survey of 13 000 air conditioners installed on residential 

and commercial buildings. The measurements were collected during routine installation, repair, and 

maintenance visits. Of the 8873 residential systems tested, 5776 (65 %) required repairs, and of the 4384 

light commercial systems tested, 3100 (71 %) required repairs. Improper refrigerant charge was found in 

57 % of all systems. The authors noted that the simple temperature split method for identifying units with 

low airflow is flawed because it does not account for the system operating condition.

Proctor (2004) presented results from a survey study involving 55000 units. He reported that 60 % of 

commercial air conditioners and 62 % of residential air conditioners had incorrect refrigerant charge. In 

all, 95 % of residential units failed the diagnostic test because of duct leakages, poor duct insulation or 

excessive airflow restriction, improper refrigerant charge, low evaporator airflow, non-condensables in 

the refrigerant or an improperly sized unit.

Rossi (2004) presented measured performance data and statistics on unitary air conditioners. The data 

were gathered using commercially available portable data acquisition systems during normal maintenance 

and service visits. Out of 1468 systems considered in this study, 67 % needed service. Of those 15 % 

required major repairs (e.g., compressor or expansion device replacement), and 85 % required a tune-up 

type service (e.g., coil cleaning or refrigerant charge adjustment). Approximately 50 % of all units 

operated with efficiencies of 80 % or less, and 20 % of all units had efficiencies of 70 % or less of their 

design efficiency. 

Mowris et al. (2004) reported on field measurements of refrigerant charge and airflow, commonly 

referred to as RCA. Over a three-year period, 4168 new and existing split, package, and heat pumps were 

tested. The measurements showed that 72 % of the tested units had improper refrigerant charge, and 44 % 

had improper airflow. Approximately a 20 % efficiency gain was measured after refrigerant charge and 

airflow were corrected.

Neme et al. (1999) considered four installation issues − equipment sizing, refrigerant charging, adequate 

airflow, and sealing ducts − and assessed the potential benefits from improved installation practices. The 

authors relied on an extensive list of publications to determine the range of intensity of the four 

installation faults and the probable air conditioner efficiency gain resulting from a corrective action. The 

cited literature indicated the maximum efficiency improvement of 12 % for corrected airflow, 21 % for 

corrected refrigerant charge, and 26 % for eliminated duct leakage. The authors concluded that improved 

HVAC installation practices could save an average of 25 % of energy in existing homes and 35 % in new 

construction. They also pointed out that air conditioner oversizing has the potential of masking a number 

of other installation problems, particularly improper refrigerant charge and significant duct leakage, while 

a correctly sized air conditioner makes other installation problems more apparent, particularly at severe 

operating conditions.

Neal (1998) presented a methodology for calculating a field-adjusted seasonal energy efficiency ratio, 

which he referred to as SEERFA, with the goal to account for four installation errors and better represent 

the seasonal performance of the air conditioner installed in the field than the seasonal energy efficiency 

ratio (SEER) derived from tests in environmental chambers. He used correcting factors of value 1 or 

smaller, one for each installation fault, which act as multipliers on the SEER. He provided an example 
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indicating that, on average, a homeowner’s cooling cost is approximately 70 % higher than it could be 

with quality air conditioner installation. It should be noted that the proposed algorithm assumes no 

interaction between different faults, which seems to be an improper assumption. 

While the scope and specific findings presented in the above publications may differ, they uniformly 

document the prevalence of air conditioner and heat pump faults in the field and a significant performance 

degradation of this equipment.

2.2 Heat Pump Oversizing, Undersizing and Part-load Losses

It is generally accepted that equipment over-sizing will lead to significant part load losses due to cycling. 

Unit cycling increases energy use due to efficiency losses (Parken et al., 1985) and also can degrade the 

moisture removal capacity of the unit which leads to higher space humidity levels (Shirey et al., 2006). 

For nearly 50 years, proper sizing for residential air conditioners and heat pumps has typically been

defined using the ACCA Manual J (ACCA, 2011a).

The energy efficiency of a cycling system is governed by how quickly after startup the capacity and 

efficiency of the air conditioning unit reaches steady-state conditions. Parken et al. (1977) defined the 

‘Cyclic Degradation’ parameter (CD) as a simplified metric to predict part load losses. This parameter 

was integrated into the calculation procedure to determine the seasonal energy efficiency ratio (SEER) for 

air conditioners and heat pumps. That procedure has been incorporated into federal energy efficiency 

standards (Federal Register, 1979) and into AHRI Standard 210/240 (AHRI, 2008). The default value for 

CD in these calculation procedures is 0.25.

Many researchers have demonstrated the sensible and latent capacity of the air conditioner at startup is a 

complicated process (Henderson, 1990; O’Neal and Katipamula, 1991). The response includes the delays 

associated with pumping refrigerant from the low-side to the high-side of the system to establish the 

steady-state operating pressures as well as the first order delays due to heat exchanger capacitance. 

Several models have been proposed that represent the overall response as some combination of first order 

(time-constant) response, delay times, and other non-linear effects. Henderson (1992) compared all these 

and showed they generally could be represented as an equivalent time constant. 

As part of developing a model for latent degradation, Henderson and Rengarajan (1996) showed that the 

parameter CD can be directly related to equivalent time constant for capacity at startup while assuming a 

thermostat cycling rate parameter (Nmax) of 3.1 cycles per hour. O’Neal and Katipamula (1991) and 

Parken et al. (1977) also indirectly showed a similar relationship. The default value of 0.25 for CD is 

equivalent to an overall time constant of 1.27 minutes. 

Over the years since the SEER test and rating procedure has been developed, manufacturers have had a 

strong incentive to improve the cyclic performance of their systems. Dougherty (2003) demonstrated that 

the typical value of CD is now in the range 0.05 to 0.10 for most systems. So cyclic degradation and the 

part load efficiency losses may be of less consequence than was previously thought.

Henderson and Rengarajan (1996) developed a similar part load model to consider the degradation of air

conditioner latent or moisture removal capacity at cyclic conditions. This model focused on situations 

when the fan operated continuously, but the compressor cycled. A more comprehensive study was 

completed by Shirey et al. (2006) and a more detailed model was developed with physically-based model 

parameters. The resulting model and the more comprehensive understanding of parametric conditions for 

a wide variety of systems and conditions allowed them to develop a refined model for latent degradation 

that could also consider the case when the fan cycles on and off with the compressor (Auto Fan Mode) –

the practice most commonly used with residential systems. 
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Field testing and simulation analysis have been used to assess the impact of over-sizing on energy use and 

space humidity levels. Sonne et al. (2006) changed out oversized air conditioner units in four Florida 

houses and replaced them with units sized according to ACCA Manual J (ACCA, 2011a). Detailed 

performance data was collected both before and after the right-sized unit was installed. Their study found 

mixed results in terms of seasonal energy use and space humidity levels. In some houses energy use was 

higher, in some it was lower, and in others the results were inconclusive. Similarly, relative humidity 

(RH) appears to be either slightly higher and or unchanged after the right-sized unit was installed. They 

also speculated that duct leakage impacts were greater for the right-sized unit since longer periods of 

system operation were required to meet the same load. More duct leakage increases the thermal losses to 

the attic (supply ducts are colder for longer ‘on’ periods) and brings in more fresh air into the system. 

Both these effects increase the sensible and latent loads imposed on the system.

A simulation study by Henderson et al., (2007) also confirmed the modest and somewhat unexpected 

impact of oversizing. They found that, when 20 % duct leakage was factored into the simulations, both 

energy use and space humidity levels were only slightly affected, even when both latent degradation 

effects and part load cyclic efficiency losses were considered. For example, oversizing by 30 % in Miami 

for the HERS Reference house increased energy use by only 2 % and actually resulted in slightly lower 

space humidity levels.

2.3 Laboratory Studies of Performance Degradation of Heat Pumps Due to Faults

Several studies on degradation of the air conditioner and heat pump performance due to different faults 

are documented in the literature. While in most cases the main interest of these studies was the fault 

detection and diagnosis (FDD), some of the findings can be used in the analysis of effects of faulty 

installation. Reports of major studies on FDD for HVAC systems started to appear in the literature in the 

nineties, and the number of publications noticeably increased in the last fifteen years. 

Table 2.1 lists a few examples of studies published since 2001. The reports by Kim et al. (2006) and 

Payne et al. (2009) present detailed literature reviews up to the dates these reports were published and 

include laboratory data for the cooling and heating mode, respectively. These laboratory data are used in 

our report; however, they had to be extended through tests in environmental chambers to provide 

complete coverage of the whole range of installation faults of interest in this study (see chapter 3 of this 

report).

Table 2.1. Selected studies on faults detection and diagnosis

Investigators System Type Study Focus

Comstock and Braun (2001) Centrifugal chiller Experiment, eight single faults

Kim et al. (2006, 2009) Split residential heat pump Experiment for cooling mode,

single-faults 

Chen and Braun (2001) Rooftop air conditioner Simplified rule-based chart method

Navarro-Esbri et al. (2007) General vapor compression system Dynamic model based FDD for 

real-time application

Payne et al. (2009) Single-speed, split residential heat pump Experiment for heating model, 

single-faults

Wang et al. (2010) HVAC system for new commercial 

buildings

System-level FDD involving 

sensor faults

Cho et al. (2005) Air-handling unit for buildings Multiple faults

Li and Braun (2007) Direct expansion vapor compression system Multiple faults

Du and Jin (2008) Air handling unit Multiple faults

Southern California Edison 

Design and Engineering 

Services (SCE 2012)

Single-speed, split residential air 

conditioner

Single faults, dual faults, and triple 

faults
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A large number of laboratory cooling mode tests were performed by Southern California Edison (SCE 

2012) to determine the effects of common faults on air conditioner performance. These faults included 

indoor airflow, outdoor airflow, refrigerant charge, non-condensables, and liquid line restrictions. 

SCE single-fault tests at a low refrigerant charge showed similar degradations in cooling capacity and 

total power as Kim et al. (2006); SCE reported -3 % and 0 % change in cooling capacity and total power, 

respectively, at 13 % undercharge while Kim et al. (2006) reported -5 % and -2 % change at 10 % 

refrigerant undercharge. However, at higher fault levels, SCE measured much higher performance 

degradation than Kim et al.; cooling capacity and total power changed by -54 % and -5 %, respectively, at 

27 % undercharge (SCE) compared to -17 % and -3 % at 30 % undercharge (Kim et al., 2006). These 

large differences in cooling capacity change for a similar fault level exemplify differences in the effect a 

given fault may have on different systems. In the case of refrigerant undercharge fault, it is possible that 

different internal volumes were a factor in the different system responses. 

SCE also performed several tests with dual and triple faults, which included reduction of the outdoor 

airflow by imposing different levels of airflow restriction. For the highest level of outdoor airflow 

blockage, 40 % refrigerant undercharge, and 56 % reduction in indoor airflow, the cooling capacity 

decreased by almost 70 %. The conducted multiple fault tests show the range of possible performance 

degradation, however, more tests are required to allow modeling of these faults within annual simulations 

of the house/heat pump system.
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3. HEAT PUMP PERFORMANCE DEGRADATION DUE TO FAULTS

A significant number of laboratory tests were taken by Kim et al. (2006) and Payne et al. (2009) to 

characterize heat pump performance degradation due to faults. For the purpose of this study, we 

conducted additional tests using the same heat pump and test apparatus to expand the ranges of previously 

studied faults and to include faults that were not covered earlier, specifically, improper electric line 

voltage and improper liquid line subcooling. The goal of this experimental effort was to enable the 

development of correlations that characterize the heat pump performance operating with these faults. 

These correlations are presented in a non-dimensional format with performance parameters expressed as a 

function of operating conditions and fault level. 

3.1 Laboratory Measurements

3.1.1 Experimental Apparatus and Test Conditions

The studied system was a single-speed, split heat pump with an 8.8 kW (2.5 ton) rated cooling capacity. 

The heat pump was equipped with a thermostatic expansion valve (TXV). Figure 3.1 shows a schematic 

diagram of the experimental setup with the locations of the main measurements. The air-side 

measurements included indoor dry-bulb and dew-point temperatures, outdoor dry-bulb temperature, 

barometric pressure, and pressure drop across the air tunnel (not shown on the schematic). Twenty-five 

node, T-type thermocouple grids and thermopiles measured air temperatures and temperature change, 

respectively. On the refrigerant side, pressure transducers and T-type thermocouple probes measured the 

inlet and exit parameters at every component of the system. 

Figure 3.1. Schematic diagram of experimental apparatus (Kim et al. (2006))

Tables 3.1 presents the cooling and heating test conditions (indoor dry bulb, indoor dew point, and 

outdoor dry bulb temperatures), and Table 3.2 presents the measurement uncertainties. For the uncertainty 

analysis and detailed description of the experimental setup the reader should refer to Kim et al. (2006). 

P, T
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Table 3.1. Cooling and heating test temperatures 

Cooling Heating

TID o

C (oF)

TIDP o

C (oF)

 TOD o

C (oF)

TID o

C (oF)

TIDP o

C (oF)

TOD o

C (oF)

21.1 (70) 10.3 (50.5) 27.8 (82) 18.3 (65) dry -8.3 (17)

21.1 (70) 10.3 (50.5) 37.8 (100) 21.1 (70) dry -8.3 (17)

26.7 (80) 15.8 (60.4) 27.8 (82) 21.1 (70) dry 1.7 (35)

 26.7 (80) 15.8 (60.4) 35.0 (95) 21.1 (70) dry 8.3 (47)

 26.7 (80) 15.8 (60.4) 37.8 (100)

Note: The dew-point temperature in the cooling mode corresponds to a relative humidity of 50 %

Table 3.2. Measurement uncertainties

3.1.2 Studied Faults and Their Implementation 

Table 3.3 lists seven studied faults, including their definition and range. The first six faults were studied 

experimentally. The impact of the last listed fault, cooling-mode TXV undersizing, was determined 

based on a detailed analysis; the inherent variable-opening capability masks the TXV undersizing, and the 

performance penalty occurs only after the outdoor temperature is below a certain threshold temperature, 

referred to by us as the ‘departure temperature’, which is related to the level of this fault. We did not 

include the TXV mismatched fault in the heating mode because it is very unlikely to occur as the heating 

TXV is installed in the outdoor section at the factory at time of assembly. 

The indoor airflow fault was implemented by lowering the speed of the nozzle chamber booster fan to 

increase the external static pressure across the indoor air handler. The fault level was calculated as a ratio 

of the fault-imposed air mass flow rate to the no-fault air mass flow rate, with the -100 % fault level 

indicating a complete loss of airflow. 

The no-fault refrigerant charge was set in the cooling mode at the AHRI 210/240 Standard A-test 

condition (AHRI, 2008). The refrigerant undercharge and overcharge faults were implemented by adding 

or removing the refrigerant from a correctly charged system. The fault level was defined as the ratio of 

the refrigerant mass by which the system was overcharged or undercharged to the no-fault refrigerant 

charge, with 0 % indicating the correct, no-fault charge, -100 % indicating no refrigerant charge, and 

100 % indicating doubled charge.

Measurement Measurement Range Uncertainty at the 95% 

confidence level

Air dry-bulb temperature (-9 ~ 38) oC ((15 ~ 100) oF)) ±0.4 oC (±0.7 oF)

Air dew-point temperature (0 ~ 38) o

C (32 ~ 100) oF) ±0.4 o

C (±0.7 oF)

Air temperature difference (0 ~ 28) oC (0 ~ 50) oF) ±0.3 oC (±0.5 oF)

Air nozzle pressure (0 ~ 1245) Pa ((0 ~ 5) in H2O) ±1.0 Pa (0.004 in H2O)

Refrigerant temperature (-12 ~ 49) o

C ((10 ~ 120) oF) ±0.3 o

C (±0.5 oF)

Refrigerant mass flow rate (0 ~ 272) kg∙h-1 ((0 ~ 600) lb∙h-1

) ±1.0 % 

Cooling capacity (3 ~ 11) kW ((3 ~ 11) kW) ±4.0 %

Power (25 ~ 6000) W ((25 ~ 6000) W) ±2.0 %

COP 2.5 ~ 6.0 ±5.5 %
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Table 3.3. Definition and range of studied faults

Fault name Symbol Definition of fault level Fault range 

(%)

Improper indoor airflow rate AF % above or below correct airflow rate -50 ~ 20

Refrigerant undercharge UC % mass below correct (no-fault) charge -30 ~ 0

Refrigerant overcharge OC % mass above correct (no-fault) charge 0 ~ 30

Improper liquid line refrigerant 

subcooling (indication of 

improper refrigerant charge)

SC % above the no-fault subcooling value 0 ~ 200

Presence of non-condensable 

gases

NC % of pressure in evacuated indoor 

section and line set, due to noncondensable gas, with respect to 

atmospheric pressure

0 ~ 20

Improper electric line voltage VOL % above or below 208 V -8.7 ~ 25

TXV undersizing, cooling TX % below the nominal cooling capacity -60 ~ -20 

The amount of refrigerant in a TXV-equipped system can also be estimated by examining the refrigerant 

subcooling in the liquid line; this method is commonly used by field technicians installing or servicing a 

heat pump. Therefore, we also characterized the effect of refrigerant overcharge by noting the liquid line 

subcooling at increased charge levels. The ratio of fault-imposed subcooling to the no-fault subcooling 

indicated the fault level with the 0 % fault corresponding to the proper subcooling, and the 100 % fault 

indicating a doubled subcooling level.

The non-condensable gas fault is caused by incomplete evacuation of the system during installation or 

after a repair that required opening the system to the atmosphere. When a new heat pump is installed, the 

outdoor unit is typically pre-charged, and the installer needs to evacuate the indoor section and the 

connecting tubing before charging it with refrigerant. Industry practice (ACCA, 2010) is to evacuate the 

system to a vacuum of 500 μPa (29.9 in Hg vacuum). The non-condensable gas fault was implemented by 

adding dry nitrogen to the evacuated system before the charging process. This fault level is defined by the 

ratio of pressure in the evacuated indoor section due to non-condensable to the atmospheric pressure. The 

0 % fault level occurs when the refrigerant charging process starts with a vacuum, and the 100 % fault 

level would occur when the nitrogen filled refrigerant lines are at atmospheric pressure before the 

refrigerant is charged. 

The electrical line voltage fault was implemented by varying the supply voltage to the system from the 

nominal, no-fault value of 208 VAC. The fault level was defined by the percentage by which the line 

voltage was above or below the nominal level, with a positive fault indicating a voltage above 208 VAC. 

TXV mismatch results in the TXV being unable to adjust its opening to match the refrigerant mass flow 

rate pumped by the compressor. This fault level is defined as the ratio of the difference in the nominal 

system capacity and the TXV capacity with respect to the nominal system capacity. With this definition, it 

is assumed TXVs are rated at the midpoint of their opening range of ±40 %. 
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3.2 Fault Effects on Cooling Mode Performance

3.2.1 Cooling Mode Normalized Performance Parameters and Correlations

The cooling mode tests considered the effect of faults on six performance parameters: total cooling 

capacity (Qtot,, capacity includes the indoor fan heat), refrigerant-side cooling capacity (QR,, capacity does 

not include the indoor fan heat), coefficient of performance (COP), sensible heat ratio (SHR), outdoor 

unit power (WODU, includes the compressor, outdoor fan, and controls powers), and total power (Wtot, 

includes WODU and indoor fan power). These parameters are presented in a dimensionless, normalized 

format obtained by dividing the values as obtained for the heat pump operating under a selected fault to 

their value obtained for the heat pump operating fault free. We used Eq. (3.1) to correlate the 

dimensionless parameters as a function of the indoor dry-bulb temperature (TID), outdoor dry-bulb 

temperature (TOD), and fault level (F). 

Y= Xfault

Xno-fault

=1+(a1+a2TID+a3TOD+a4F)F (3.1) 

where a1

, a2, a3, and a4 are correlation coefficients, Xfault and Xno-fault are performance parameters for a 

faulty and fault-free heat pump, and Y is a dimensionless parameter representing the ratio of the faulty 

performance from that of the fault-free heat pump. 

Table 3.4 shows coefficients for a correlation using three input variables, TID, TOD, and F. The 

coefficients were determined by means of a multivariate polynomial regression method using the 

normalized values of performance parameters determined from heat pump test data. If the heat pump is 

fault free, values of all normalized parameters equal unity. The fit standard error of the normalized 

correlation dependent variable, Y, was a maximum of 3 % over the range of operating conditions listed in 

Table 3.1. Table 3.5 shows an example of propagation of uncertainty for the faulty COP and cooling 

capacity obtained from calculations using the measurement uncertainties of the corresponding fault-free 

values and the 3 % uncertainty in the dimensionless parameter Y.

The following is an explanation of the procedure used to calculate the dimensionless capacity and COP 

due to undersizing of the cooling mode TXV. This fault occurs if the expansion valve’s equivalent orifice 

area is too small to control refrigerant superheat during periods of low ambient temperature conditions at 

reduced condenser pressures. A properly sized TXV will regulate refrigerant flow rate and maintain 

proper superheat over a wide range of indoor and outdoor air temperatures. However, if the indoor TXV 

is undersized for the particular outdoor unit, the system performance is degraded due to a restricted mass 

flow of refrigerant at certain evaporator and condenser pressure differentials. The rated TXV capacity 

and nominal system capacity are used to determine the TXV undersizing fault level. For example, if a 

7.0 kW (2 ton) TXV is installed in a system with the nominal capacity of 8.8 kW (2.5 ton), the fault level 

is 20 %. (F = 1-7.0/8.8=0.20). 

Since the pressure difference between upstream and downstream becomes smaller with decreasing 

outdoor temperature, the TXV opens to increase refrigerant mass flow rate at low outdoor temperatures. 

The outdoor temperature at which the TXV reaches its maximum orifice size, referred to as the ‘departure 

temperature’, is determined from calculations and empirical fits to previous data. The resulting departure 

temperature below which the TXV cannot supply adequate mass flow rate is given by Eq. (3.2).

Tdep[°C]=80.326∙F+11.682 (3.2)
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Table 3.4. Correlations for non-dimensional performance parameters in the cooling mode

* All temperatures are in Celsius

** FSE (fit standard error) equals the square root of the sum of the squared errors divided by the degrees of freedom *** The applicable range of SHR for wet coil predictions: 0.7 to 0.85

Table 3.5. Example uncertainty propagation due to normalized correlation (Y) uncertainty of 3 % for 

faulty COP and cooling capacity at AHRI Standard 210/240 B-test condition (AHRI, 2008) 

Fault Parameter Parameter Value Uncertainty (%)

(95 % confidence level)

10 % reduced indoor 

airflow

COP 3.67 ± 6.4

Cooling capacity 9.4 kW ± 5.0

Fault Performance 

parameter Y

Y=1+(a

1

+a

2

TID+a

3

TOD+a

4

F)·F*

FSE**

a

1 a2 a3 a4

Improper indoor 

airflow rate (AF)

COP YCOP = YQtot/ YWtot 1.65E-02

Qtot 1.85E-01 1.77E-03 -6.40E-04 -2.77E-01 1.53E-02

QR 2.95E-01 -1.17E-03 -1.57E-03 6.92E-02 5.39E-03

SHR*** 5.93E-02 5.16E-03 1.81E-03 -2.89E-01 9.82E-03

WODU -1.03E-01 4.12E-03 2.38E-03 2.10E-01 6.91E-03

Wtot 1.35E-02 2.95E-03 -3.66E-04 -5.88E-02 5.68E-03

Refrigerant 

undercharge (UC))

COP YCOP = YQtot/ YWtot 1.17E-02

Qtot -5.45E-01 4.94E-02 -6.98E-03 -1.78E-01 1.02E-02

QR -9.46E-01 4.93E-02 -1.18E-03 -1.15E+00 1.44E-02

SHR*** 4.19E-01 -2.12E-02 1.26E-03 1.39E-01 8.56E-03

WODU -3.13E-01 1.15E-02 2.66E-03 -1.16E-01 5.14E-03

Wtot -2.54E-01 1.12E-02 2.06E-03 5.74E-03 5.29E-03

Refrigerant overcharge 

(OC)

COP YCOP = YQtot/ YWtot 2.00E-02

Qtot 4.72E-02 -1.41E-02 7.93E-03 3.47E-01 1.96E-02

QR -1.63E-01 1.14E-02 -2.10E-04 -1.40E-01 5.67E-03

SHR*** -7.75E-02 7.09E-03 -1.93E-04 -2.76E-01 7.34E-03

WODU 2.19E-01 -5.01E-03 9.89E-04 2.84E-01 5.17E-03

Wtot 1.46E-01 -4.56E-03 9.17E-04 3.37E-01 5.43E-03

Improper 

liquid line refrigerant 

subcooling (SC)

COP YCOP = YQtot/ YWtot 2.26E-02

Qtot 6.77E-02 0.00E+00 -1.22E-03 -1.91E-02 2.18E-02

QR 4.16E-02 0.00E+00 -3.51E-04 -1.55E-02 1.39E-03

SHR*** -9.04E-02 0.00E+00 2.13E-03 1.60E-02 3.06E-02

WODU 2.11E-02 0.00E+00 -4.18E-04 4.25E-02 4.34E-03

Wtot 1.06E-02 0.00E+00 -2.93E-04 3.88E-02 4.84E-03

Non-condensable gas 

(NC)

COP YCOP = YQtot/ YWtot 1.71E-02

Qtot 2.77E-01 -1.75E-02 1.78E-02 -1.96E+00 1.63E-02

QR -1.78E+00 4.04E-02 1.78E-02 9.98E-01 9.59E-03

SHR*** -4.67E-01 1.69E-02 9.89E-04 2.90E-01 5.59E-03

WODU -6.92E-01 2.01E-02 1.20E-02 6.62E-01 6.13E-03

Wtot -5.37E-01 1.52E-02 1.09E-02 4.36E-01 6.20E-03

Improper line voltage 

(VOL)

COP YCOP = YQtot/ YWtot 1.98E-02

Qtot 5.84E-01 -1.21E-02 -8.57E-03 -3.35E-01 1.80E-02

QR 1.03E-01 -6.10E-03 3.64E-03 -1.04E-01 6.41E-03

SHR*** -6.65E-02 5.21E-03 -2.10E-03 4.23E-02 2.95E-02

WODU 7.66E-01 -3.85E-03 -1.83E-02 1.14E+00 4.39E-03

Wtot 9.06E-01 -6.37E-03 -1.75E-02 1.10E+00 7.39E-03

TXV undesizing, 

cooling (TXV) Refer to Eqs. (3.6, 3.7) and Table 3.6 
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The cooling capacity and the gross COP of the undersized TXV-equipped system can be expressed as 

functions of outdoor temperature and fault level. To develop equations for the normalized capacity and 

COP, non-dimensional variables for outdoor temperature, cooling capacity, and gross COP are defined by 

Eqs. (3.3, 3.4, 3.5), respectively, where TOD has Celsius units. 

Tr= TOD

35 °C (3.3)

YQ= Qundersized

Qno−fault

(3.4) 

YCOP=

COPundersized

COPno-fault

(3.5)

The correlations for determining normalized cooling capacity and normalized gross COP are given by 

Eqs. (3.6) and (3.7) and are presented in a graphical form in Figure 3.2. The coefficients are listed in 

Table 3.6. 

YQ=a1+a2Tr+a3F+a4Tr

2

+a5TrF+a6F2 if TOD≤Tdep or YQ=1 if TOD>Tdep (3.6)

YCOP=b1+b2Tr+b3F+b4Tr

2

+b5TrF+b6F2 if TOD≤Tdep or YQ=1 if TOD>Tdep (3.7)

Table 3.6. Normalized capacity and COP correlation coefficients for a TXV undersizing fault 

Coefficients for YQ Coefficients for YCOP

a1 9.1440E-01 b1 8.4978E-01

a2 2.0903E-01 b2 4.0050 E-01

a3 -5.4122E-01 b3 -8.4120E-01

a4 1.2194E-01 b4 7.5740E-02

a5 -2.9428E-01 b5 -3.3105E-01

a6 -3.0833E-02 b6 2.0290E-01

A complete and detailed discussion of the TXV undersizing fault correlation development is beyond the

scope of this report and is presented by Payne and Kwon (2014). 
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Figure 3.2. Normalized performance parameters for the cooling mode TXV undersizing fault: 

(a) capacity, (b) COP

3.2.2 Cooling Mode Charts with Normalized Performance Parameters

Figures 3.3 through 3.8 show variations of the normalized performance parameters with respect to fault 

levels at five operating conditions. The figures present the measured data points and correlations 

developed for COP, capacity, SHR, total power and, for some faults, the outdoor unit power. The outdoor 

unit power is included for improper indoor airflow (AF) and improper liquid line refrigerant subcooling 

(SC) faults where the trends of the total power and the outdoor unit power were not similar. In some of 

the figures there is a significant difference between the correlation fits and the actual data points. The 

correlations were developed for all indoor and outdoor test conditions, and thus the fit sum of squared 

deviations was minimized. In addition, the normalized value for the heat pump operating with no fault 

was calculated from the fault-free correlation as presented by Kim et al. (2010); therefore, no-fault tests 

may actually have normalized values somewhat different from unity due to the inability of the no-fault 

correlation to predict the no-fault parameter exactly. Scatter of normalized no-fault data around unity 

indicates measurement uncertainty, correlation uncertainty, and uncertainty caused by different system 
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15

installations. The data for Figures 3.6 and 3.8 were collected after the system was removed and reinstalled in the test chambers; therefore, one would expect more scatter in the normalized no-fault 

correlations due to this installation repeatability uncertainty. This installation repeatability uncertainty is 

also indicative of what could be seen in field installations when applying the same no-fault correlations 

from system to system. 

Figure 3.3 shows the normalized parameters at a reduced and increased indoor airflow. For the studied 

airflow range from -50 % to +20 % of the nominal value, the change in outdoor unit power ranged 

from -3 % to 0 %, respectively, with small variations between different operating conditions. Total power 

varied from -5 % to 2 % within the same range of airflow rate, which indicates the varied power of the 

indoor fan at this fault. COP and capacity were markedly degraded at a decreased airflow and somewhat 

improved at the increased airflow above the nominal level; however, these increases in COP and capacity 

were associated with a significant increase in SHR, which may not be a desirable change from the 

homeowner’s comfort point of view. The difference between total power and outdoor unit power is due to 

the power of the indoor blower, which was nominally 430 W. Outdoor unit power was relatively constant 

under this fault. As a result, COP slightly increased at the max fault level by the increased indoor airflow.

 

Figures 3.4 and 3.5 show the variation of the normalized values for refrigerant charge faults. The changes

in COP and total capacity for refrigerant undercharge are larger than those for refrigerant overcharge. A 

30 % undercharge reduced capacity by almost 15 % on average reducing COP by 12 % while a 30 % 

overcharge produced little reductions or small increases in capacity with 6 % greater total power and 3 % 

reduced COP on average because of the increased discharge pressure. In case of different outdoor 

temperature conditions, COP and capacity increased as the outdoor temperature increased for the 

undercharged condition. Farzad et al. (1990) also showed that higher refrigerant flow rate is one reason 

for the higher capacity at higher outdoor temperatures for the conditions of undercharge.

In this study, a subcooling temperature of 4.4 °C (8.0 °F) was regarded as the no-fault condition under the 

considered test conditions. Figure 3.6 shows the effects of increased subcooling at the TXV inlet. The 

departure of the normalized values of COP and cooling capacity from the correlations in the figure are 

mostly due to the TXV attempting to correct mass flow rate (reduce effective orifice size) as subcooling 

increases. If more data were available with subcooling being varied randomly from high to low values, 

hysteresis effects and TXV hunting effects would be better captured. COP and capacity normalized 

correlations for higher levels of subcooling still represent the general trends in system performance. 

Increased subcooling is a symptom of excessive refrigerant charge, and it has the same effect; higher 

subcooling leads to reduced condensing area and increased condensing pressure. In the studied heat 

pump, refrigerant overcharging by 30 % corresponded to approximately doubling of refrigerant 

subcooling. For this level of fault the COP degradation was within 4 %. For the highest subcooling fault 

of 181 % of the nominal value, the impact on the capacity was minor but the outdoor unit power increased 

by 15 %, which resulted in a similar decrease in the COP.

Figure 3.7 shows the variation of the normalized values for chosen performance parameters versus noncondensable gas (NC) fault level. Non-condensable gases increase the condensing pressure above that 

corresponding to the saturation pressure of the refrigerant at the same temperature due to the partial 

pressure of the NC components. As a result, increased total power consumption and decreased COP can 

be seen in the Figure 3.7. Maximum degradation of COP at the 20 % fault level was about 5 % for the 

condition of TID=26.7 °C (80.0 °F) and TOD=27.8 °C (82.0 °F).

Figure 3.8 shows the variation of the normalized values for chosen performance parameters for the line 

voltage variation fault conditions. A line voltage of 208 V was set as the no-fault condition. Total external 

static pressure for the indoor air handler was set at 125 Pa (0.5 in H2O) at the no-fault line voltage, which 

produced a nominal indoor fan power demand of 430 W. As voltage increased, fan speed and static 
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16

pressure increased thus producing increased fan power. Total power consumption increased almost 

linearly as the fault level increased. The fan power increased more than the compressor power when the 

voltage was increased. An average increase of 27 % for the fan power and 9 % for the compressor power 

occurred at the max fault level. At fault levels over 20%, the degradation of COP is greater than 10 %.

The presented measurements for the cooling mode indicate that the refrigerant undercharge fault has the 

highest potential for degrading air conditioner efficiency. For 30 % percent undercharge – a fault level 

commonly observed during field surveys – the system efficiency is decreased between 7 % and 15 %, 

depending on operating conditions. 

A reduction of the airflow rate by 30 % (also a commonly observed fault) can reduce the efficiency by 

6 %, and this level of degradation persists independently of operating conditions. Refrigerant 

overcharging by 30 % resulted in COP degradation on the order of 4 %. COP degradation within 3 % 

was measured for improper electric voltage and non-condensable gas faults. The non-condensable gas 

fault can be misdiagnosed in the field as refrigerant overcharge, which may prompt a serviceman to 

remove some of the refrigerant from the system, thus triggering an undercharge fault.
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Fault level (%)

Figure 3.3. Normalized cooling performance parameters for improper indoor airflow

(The numbers in the legend denote test conditions, TID (°C)/ TOD (°C))
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Figure 3.4. Normalized cooling performance parameters for refrigerant undercharge

(The numbers in the legend denote test conditions, TID (°C)/ TOD (°C))
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Figure 3.5. Normalized cooling performance parameters for refrigerant overcharge

(The numbers in the legend denote test conditions, TID (°C)/ TOD (°C))
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Figure 3.6. Normalized cooling performance parameters for improper liquid line refrigerant subcooling

(The numbers in the legend denote test conditions, TID (°C)/ TOD (°C))
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Figure 3.7. Normalized cooling performance parameters for the presence of non-condensable gas

(The numbers in the legend denote test conditions, TID (°C)/ TOD (°C))
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Figure 3.8. Normalized cooling performance parameters for improper electric line voltage

(The numbers in the legend denote test conditions, TID (°C)/ TOD (°C))
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3.3 Fault Effects on Heating Mode Performance

3.3.1 Heating Mode Normalized Performance Parameters and Correlation

The heating mode tests considered the effect of faults on five performance parameters: coefficient of 

performance (COP), total heating capacity (Qtot,, includes the indoor fan heat), refrigerant-side heating 

capacity (QR,, does not include the indoor fan heat), outdoor unit power (WODU, includes the compressor, 

outdoor fan, and controls powers), and total power (Wtot, includes WODU and indoor fan power). These 

parameters are presented in a dimensionless, normalized format obtained by dividing these parameter 

values as obtained for the heat pump operating under a selected fault by the no-fault value. The 

normalized parameters were correlated as a function of outdoor dry-bulb temperature (TOD) and fault level 

(F). These two parameters were the only values varied for the heating mode tests; indoor dry-bulb 

temperature did not vary enough to use in the heating mode correlations. 

Y= Xfault

Xno-fault

=1+(a1+a2TOD+a3F)∙F (3.8)

where a1

, a2, and a3 are correlation coefficients, Xfault and Xno-fault are performance parameters for a faulty 

and fault-free heat pump, and Y is a dimensionless parameter representing the ratio of the faulty 

performance from that of the fault-free heat pump. 

Tables 3.7 shows the correlation coefficients. They were determined by means of a multivariate 

polynomial regression method using the normalized values of performance parameters determined from 

heat pump test data. If the heat pump is fault free, values of all normalized parameters equal unity. 

3.3.2 Heating Mode Charts with Normalized Performance Parameters

Figure 3.9 shows the effects of reduced airflow over the indoor coil during heating mode operations. The 

airflow rate through the indoor heat exchanger was controlled by changing the speed of the nozzle 

chamber booster fan. As shown in the graphs, effects of this fault condition for COP and power are 

noticeable. Especially, for the higher outdoor temperature condition (TOD=8.3 °C (47 °F)) with a 50 % 

reduced airflow rate, COP was degraded by over 30 % and total power increased by more than 20 %. 

Figure 3.10 shows the effects of refrigerant undercharge. At the maximum fault level of 30 %, COP 

decreased by more than 8 % for the higher outdoor temperature condition (8.3 °C (47 °F)). The decrease 

was greater for the lower temperature lift case due to the lower pressure ratio and resulting lower mass 

flow rate potential (pressure drop) across the expansion valve as compared to the -8.3 °C (17 °F) case.

Mass flow rate is proportional to the square root of the pressure drop. Therefore the reduction in mass 

flow rate due to removing refrigerant and lowering liquid line subcooling (lowering liquid line pressure) 

will have a greater effect upon mass flow rate at higher condenser pressure (higher outdoor temperatures). 

Capacity reduction had a greater effect upon COP than compressor power demand due to undercharge; 

refrigerant-side capacity decreased by an average of 22 % while outdoor unit power demand decreased an 

average of only 5 % for this maximum fault level and 8.3 °C (47 °F) test condition. 

Figure 3.11 shows the effects of refrigerant overcharge. The control effect of the TXV is seen in the 

refrigerant-side capacity; capacity remains nearly constant (±1 %) while compressor power demand 

increases to approximately 15 % at 30 % fault level. The TXV maintains outdoor coil exit superheat by 

increasing pressure drop and limiting mass flow. Compressor power demand increases, being more 

pronounced at the lower temperature lift (lower pressure ratio), highest outdoor temperature. At the lower 

pressure ratio case, system capacity and refrigerant mass flow are already greater than the higher pressure 

ratio case, and the addition of refrigerant produces a greater change in power demand for a given fault 

level.
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 Table 3.7. Correlations for non-dimensional performance parameters in the heating mode

 * All temperatures are in Celsius ** FSE (fit standard error) equals the square root of the sum of the squared errors divided by the degrees of freedom

Refrigerant overcharge demonstrates itself in increased refrigerant subcooling in the liquid line. When 

subcooling was doubled from its nominal value (a fault level of 100 %), compressor power demand 

increased by approximately 15 % with little change in capacity (Figure 3.12). This resulted in an almost 

12 % decrease in COP. Increased subcooling (increased refrigerant charge) affects compressor power 

demand more than capacity due to the TXV control of evaporator exit superheat. 

Figure 3.13 shows the effects of non-condensable gas. The non-condensable gas will accumulate in the 

condenser (indoor coil) and thus reduce the heat transfer area available and raise the condenser pressure in 

direct proportion to the volume of the non-condensable gas. At the highest fault level of approximately 

20 %, the COP decreases by approximately 8 % at the lowest outdoor test temperature. The noncondensable gas appears to have equal effect upon compressor power demand at all fault levels and 

outdoor temperatures, while capacity is more affected at the higher pressure ratio produced at the lowest 

outdoor temperature. 

Figure 3.14 shows the effects of varying the system working voltage above and below the nominal value 

of 208 VAC. The changes in compressor power demand are a result of increased evaporator refrigerant 

saturation temperature at the higher indoor airflow rates. Changing the supply voltage changes all of the 

electric motors’ rotational speeds; therefore, lowering the voltage is equivalent to reducing compressor 

pumping capacity while leaving heat transfer area constant. At higher voltages, the higher compressor 

Fault Performance

Parameter Y

Y=1+(a

1

+ a

2

TOD + a

3

F)·F*

FSE**

a

1 a2 a3

Improper indoor 

airflow rate (AF)

COP YCOP = YQtot / YWtot 3.27E-02

Qtot 0.1545961 0.0078768 -0.1746421 2.72E-02

QR 0.0009404 0.0065171 -0.3464391 1.82E-02

WODU -0.177359 -0.0125111 0.4784914 1.21E-02

Wtot 0.0311053 -0.009332 0.7942998 2.87E-02

Refrigerant 

undercharge (UC)

COP YCOP = YQtot / YWtot 2.68E-02

Qtot -0.104922 0.0156348 -1.3702726 8.02E-03

QR -0.0338595 0.0202827 -2.6226343 2.55E-02

WODU 0.0615649 0.0044554 -0.2598507 8.79E-03

Wtot 0.0537015 0.004334 -0.2272758 7.85E-03

Refrigerant 

overcharge (OC)

COP YCOP = YQtot / YWtot 6.08E-03

Qtot -0.1198701 -0.0004505 0.5052803 5.20E-03

QR -0.0029514 0.0007379 -0.0064112 3.14E-03

WODU -0.0594134 0.0159205 1.8872153 9.19E-03

Wtot -0.053594 0.0140041 1.6948771 8.43E-03

Improper liquid 

line refrigerant 

subcooling (SC)

COP YCOP = YQtot / YWtot 1.77E-02

Qtot -0.0369891 0.0014081 0.0113751 1.06E-02

QR -0.0389621 0.0019259 0.0079344 1.41E-02

WODU 0.1353483 -0.001264 0.008241 8.45E-03

Wtot 0.1023326 -0.0007392 0.0128456 6.11E-03

Noncondensable 

gas (NC)

COP YCOP = YQtot / YWtot 9.69E-03

Qtot 0.0852956 0.0058473 -0.9522349 9.37E-03

QR -0.2081656 0.0058006 0.6035798 2.48E-03

WODU 0.181571 0.0008425 0.6093669 3.95E-03

Wtot 0.1840392 -0.0001309 0.3935121 3.92E-03

Improper line 

voltage (VOL)

COP YCOP = YQtot / YWtot 1.13E-02

Qtot 0.1107829 -0.0040167 -0.1347848 9.87E-03

QR 0.0912687 -0.0006155 -0.2343559 5.60E-03

WODU 0.1604092 0.0011052 0.9262117 1.80E-03

Wtot 0.283868 0.0009125 0.7759193 3.61E-03
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25

speed produces more of an effect on power demand than the indoor airflow rate produces on capacity. 

The TXV regulates refrigerant flow to maintain superheat at the higher indoor airflow rates. Capacity 

increases less than 2 % at the highest voltage while compressor power demand increases by more than 

10 % resulting in an almost 10 % decrease in COP. 
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26

-50 -40 -30 -20 -10 0 10 20 0.6

0.7

0.8

0.9

1.0

1.1

 -8.3

 8.3

 1.7

 Fit -8.3

 Fit 8.3

 Fit 1.7

COP (Normalized value)

Fault level (%)

-50 -40 -30 -20 -10 0 10 20 0.6

0.7

0.8

0.9

1.0

1.1

 -8.3

 8.3

 1.7

 Fit -8.3

 Fit 8.3

 Fit 1.7

QR (Normalized value)

Fault level (%)

-50 -40 -30 -20 -10 0 10 20 0.9

1.0

1.1

1.2

1.3

1.4

 -8.3

 8.3

 1.7

 Fit -8.3

 Fit 8.3

 Fit 1.7

WODU (Normalized value)

Fault level (%)

-50 -40 -30 -20 -10 0 10 20 0.6

0.7

0.8

0.9

1.0

1.1

 -8.3

 8.3

 1.7

 Fit -8.3

 Fit 8.3

 Fit 1.7

Qtot (Normalized value)

Fault level (%)

-50 -40 -30 -20 -10 0 10 20 0.9

1.0

1.1

1.2

1.3

1.4

 -8.3

 8.3

 1.7

 Fit -8.3

 Fit 8.3

 Fit 1.7

Wtot (Normalized value)

Fault level (%)

Figure 3.9. Normalized heating performance parameters for improper indoor airflow

(The number in the legend denotes TOD (°C))
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27

-30 -20 -10 0 0.6

0.7

0.8

0.9

1.0

1.1

 -8.3

8.3

 Fit 8.3

 Fit 8.3

COP (Normalized value)

Fault level (%)

-30 -20 -10 0 0.6

0.7

0.8

0.9

1.0

1.1

 -8.3

8.3

 Fit 8.3

 Fit 8.3

Qtot (Normalized value)

Fault level (%)

-30 -20 -10 0 0.6

0.7

0.8

0.9

1.0

1.1

 -8.3

8.3

 Fit 8.3

 Fit 8.3

WODU (Normalized value)

Fault level (%)

-30 -20 -10 0 0.6

0.7

0.8

0.9

1.0

1.1

 -8.3

8.3

 Fit 8.3

 Fit 8.3

QR (Normalized value)

Fault level (%)

-30 -20 -10 0 0.6

0.7

0.8

0.9

1.0

1.1

 -8.3

8.3

 Fit 8.3

 Fit 8.3

Wtot (Normalized value)

Fault level (%)

Figure 3.10. Normalized heating performance parameters for refrigerant undercharge

(The number in the legend denotes TOD (°C))
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0 5 10 15 20 25 30 35 0.8

0.9

1.0

1.1

 -8.3

 8.3

 Fit -8.3

 Fit 8.3

COP (Normalized value)

Fault level (%)

0 5 10 15 20 25 30 35 0.8

0.9

1.0

1.1

 -8.3

 8.3

 Fit -8.3

 Fit 8.3

Qtot (Normalized value)

Fault level (%)

0 5 10 15 20 25 30 35 0.9

1.0

1.1

1.2

1.3

 -8.3

 8.3

 Fit -8.3

 Fit 8.3

WODU (Normalized value)

Fault level (%)

0 5 10 15 20 25 30 35 0.8

0.9

1.0

1.1

 -8.3

 8.3

 Fit -8.3

 Fit 8.3

QR (Normalized value)

Fault level (%)

0 5 10 15 20 25 30 35 0.9

1.0

1.1

1.2

1.3

 -8.3

 8.3

 Fit -8.3

 Fit 8.3

Wtot (Normalized value)

Fault level (%)

Figure 3.11. Normalized heating performance parameters for refrigerant overcharge

(The number in the legend denotes TOD (°C))
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29

-50 0 50 100 150 200 0.7

0.8

0.9

1.0

1.1

 -8.3

 8.3

 Fit -8.3

 Fit 8.3

COP (Normalized value)

Fault level (%)

-50 0 50 100 150 200 0.70

0.75

0.80

0.85

0.90

0.95

1.00

1.05

1.10

 -8.3

 8.3

 Fit -8.3

 Fit 8.3

Qtot (Normalized value)

Fault level (%)

-50 0 50 100 150 200 0.9

1.0

1.1

1.2

1.3

 -8.3

 8.3

 Fit -8.3

 Fit 8.3

WODU (Normalized value)

Fault level (%)

-50 0 50 100 150 200 0.7

0.8

0.9

1.0

1.1

 -8.3

 8.3

 Fit -8.3

 Fit 8.3

QR (Normalized value)

Fault level (%)

-50 0 50 100 150 200 0.9

1.0

1.1

1.2

1.3

 -8.3

 8.3

 Fit -8.3

 Fit 8.3

Wtot (Normalized value)

Fault level (%)

Figure 3.12. Normalized heating performance parameters for improper refrigerant subcooling

(The number in the legend denotes TOD (°C))
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30

0 5 10 15 20 25 0.8

0.9

1.0

1.1

1.2

 -8.3

 8.3

 Fit -8.3

 Fit 8.3

COP (Normalized value)

Fault level (%)

0 5 10 15 20 0.8

0.9

1.0

1.1

1.2

 -8.3

 8.3

 Fit -8.3

 Fit 8.3

Qtot (Normalized value)

Fault level (%)

0 5 10 15 20 25 0.8

0.9

1.0

1.1

1.2

 -8.3

 8.3

 Fit -8.3

 Fit 8.3

WODU (Normalized value)

Fault level (%)

0 5 10 15 20 25 0.8

0.9

1.0

1.1

1.2

 -8.3

 8.3

 Fit -8.3

 Fit 8.3

QR (Normalized value)

Fault level (%)

0 5 10 15 20 25 0.80

0.85

0.90

0.95

1.00

1.05

1.10

1.15

1.20

 -8.3

 8.3

 Fit -8.3

 Fit 8.3

Wtot (Normalized value)

Fault level (%)

Figure 3.13. Normalized heating performance parameters for the presence of non-condensable gas

(The number in the legend denotes TOD (°C))
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31

-10 0 10 20 30 0.8

0.9

1.0

1.1

1.2

 -8.3

 8.3

 Fit -8.3

 Fit 8.3

COP (Normalized value)

Fault level (%)

-10 0 10 20 30 0.8

0.9

1.0

1.1

1.2

 -8.3

 8.3

 Fit -8.3

 Fit 8.3

Qtot (Normalized value)

Fault level (%)

-10 0 10 20 30 0.8

0.9

1.0

1.1

1.2

 -8.3

 8.3

 Fit -8.3

 Fit 8.3

WODU (Normalized value)

Fault level (%)

-10 0 10 20 30 0.8

0.9

1.0

1.1

1.2

 -8.3

 8.3

 Fit -8.3

 Fit 8.3

QR (Normalized value)

Fault level (%)

-10 0 10 20 30 0.8

0.9

1.0

1.1

1.2

 -8.3

 8.3

 Fit -8.3

 Fit 8.3

Wtot (Normalized value)

Fault level (%)

Figure 3.14. Normalized heating performance parameters for improper line voltage

(The number in the legend denotes TOD (°C))
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4. BUILDING/HEAT PUMP MODELING APPROACH

4.1 Building/Heat Pump Systems Simulation Models

Several building simulation models are available for modeling residential buildings. Many include welldeveloped user interfaces aimed at specific audiences – such as residential home energy raters in the 

United States who seek to determine the Home Energy Rating System score (HERS) (RESNET, 2006). 

Table 4.1 summarizes the features of these mainstream software tools. Energy Gauge USA, RemRate, 

and TREAT all have hundreds of users and are widely known in the residential energy efficiency 

community. However, while these tools include models for commonly-used systems and equipment 

operating at or near their nominal performance ratings, they do not have the flexibility to consider 

degraded, abnormal or off-design performance. 

Table 4.1. Comparison of residential building simulation software tools

Energy Gauge USA

Fully developed hour-by-hour building simulation model (based on DOE2.1e). Tool is commonly used by energy raters to develop a Home 

Energy Rating System (HERS) score. www.energygauge.com 

Advantages: good well-documented, building model with sound 

equipment components

Disadvantages: no flexibility to add extra correlations or components 

RemRATE

Building simulation model (using temperature bin calculations). Tool is 

commonly used by energy raters to develop a Home Energy Rating 

System (HERS) score. www.archenergy.com/products/remrate

Advantages: good well-documented, building model with models for 

common equipment components

Disadvantages: no flexibility to add extra correlations or components

TREAT

Hourly building simulation model aimed at residential energy analysis, 

for both single-family and multi-family homes. Used widely in the multifamily energy efficiency sector. www.psdconsulting.com/software/treat

Advantages: robust, well-documented, building model focused on multifamily issues

Disadvantages: no flexibility to consider alternate technologies

DOE-2 is the original, U.S. federally-funded building simulation model or calculation engine developed 

in the 1970s that is still used as the basis of some of the mainstream residential software tools (i.e., 

Energy Gauge USA). The DOE-2 software offers some flexibility but is no longer maintained or 

supported. 

EnergyPlus is a state-of-the-art, very flexible building simulation tool used for research evaluations and 

mainstream energy analysis and design assistance. Its development is supported by the U.S. Department 

of Energy (DOE). This detailed calculation engine works at sub-hourly time steps and can consider both 

residential and commercial buildings. 

TRNSYS is a highly flexible transient simulation tool that focuses on thermal systems, primarily aimed at 

building and HVAC applications (Klein et al., 2007). TRNSYS was originally developed at the 

University of Wisconsin to simulate the transient performance of solar thermal systems 

(http://sel.me.wisc.edu/trnsys/). TRNSYS is a modular tool where multiple components can be combined 

to build up a complex thermal system. TRNSYS includes several components necessary to simulate the 

transient performance of a building, including building envelope components, HVAC equipment, and 

utilities to read hourly weather data from TMY files. Because of its flexibility, this tool is uniquely able 
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33

to consider new concepts and technologies – such as the research evaluation of this project. The core of 

the TRNSYS simulation model is the building envelope model based on the Type 56 multi-zone building 

model. The inputs to Type 56 are defined using the TRNBuild software tool (see Figure 4.1) and then 

saved in a .BUI file. Type 56 then reads this file at runtime to generate the detailed building description. 

The building model includes all the basic characteristics of a residential building:

• Heat loss and gains through building walls, roof and floor

• Solar gains through windows

• Interactions between multiple zones (house, attic, rooms)

• Scheduled internal sensible and moisture loads for people, equipment, etc.

• Interactions with the heating, ventilation, and air conditioning equipment

• Scheduled set points for temperature and humidity 

Table 4.2 summarizes the advantages and disadvantages for each of these software tools. Because of its 

flexibility, we selected the building model developed in TRNSYS to study the integrated performance of 

a heat pump in residential application.

Table 4.2. Comparison of general building calculation models

DOE-2

An hour-by-hour building simulation model developed by the national 

laboratories in the US in the mid-1970s to predict energy use in 

commercial and residential buildings (http://gundog.lbl.gov/). DOE-2.1e 

is no longer under active maintenance, but is still the underlying 

calculation engine for several software packages, including Energy 

Gauge. A private software developer (JJ Hirsh and Associates) owns and 

maintains the newest version of the DOE-2.2 calculation engine and the 

widely used interface program (eQuest). http://www.doe2.com/

Advantages: well understood, flexible simulation program

Disadvantages: no longer updated or supported 

EnergyPlus

Flexible building simulation model for commercial and residential 

buildings. Public domain calculation engine developed by the US 

Department of Energy (DOE). www.energyplus.gov

Advantages: state of the art building model with robust, well-developed 

equipment components

Disadvantages: limited flexibility to add correlations to degrade 

performance

TRNSYS

Highly flexible, research grade package for analyzing transient thermal 

systems. Includes pre-developed models for building envelope and other 

HVAC components. www.trnsys.com

Advantages: highly flexible can consider any user-defined equation or 

component models

Disadvantages: difficult to use, and cumbersome to define building 

envelope details 
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34

Figure 4.1. Screen shot of TRNBuild used to define the building envelope details

In this study, we used a building model developed in TRNSYS to simulate the integrated performance of 

heat pumps in residential applications (CDH Energy Corp., 2010). This model was originally applied to 

simulate an integrated desiccant system’s performance (Henderson and Sand, 2003), and it was later 

refined to consider several issues germane to this residential study, including duct leakage and the part 

load performance of air conditioners (Henderson et al., 2007) and refrigerant charge impacts (Sachs et al., 

2009). The model is driven by typical meteorological year weather data sets TMY3 (Wilcox and Marion, 

2008) on a small time-step (e.g., 1.2 minutes). A detailed thermostat model turns the mechanical systems 

‘on’ and ‘off’ at the end of each time step depending on the calculated space conditions. 

4.2 Building and Weather City Definitions

Table 4.3 lists the climates with representative locations and house structures considered in this study. 

Two houses were modeled: a slab-on-grade house and a house with a basement. The simulated residential 

buildings corresponded to a code-compliant house with a HERS score of approximately 100 with

appropriate levels of insulation and other features corresponding to each climate. The slab-on-grade 

houses were modeled with ducts located in the attic. The houses with basements were modeled with

ducts located in a semi-conditioned space. For Houston, TX, only a slab-on-grade house was studied 

because houses with basements are rarely built in this location.
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35

The selected cities represent each of the International Energy Conservations Code (IECC) climate zones 2 

through 6 shown in Figure 4.2, from hot and humid climate to a heating dominated climate. This 

selection enabled prediction on how different faults will affect air conditioner and heat pump performance 

in the most prevalent climates in the U.S. TMY3 weather data were used for each location.

Table 4.3. Climates, locations and structures considered

Zone Climate Location Slab-on-grade house House with basement

2 Hot and humid Houston, TX Yes No

3 Hot and dry climate Las Vegas, NV Yes Yes

4 Mixed climate Washington, DC Yes Yes

5 Heating dominated Chicago, IL Yes Yes

6 Cold Minneapolis, MN Yes Yes

Figure 4.2. IECC climate zone map

4.3 Building and Enclosure Thermal Details

A 185.8 m2 (2,000 ft2

) three-bedroom house was modeled as a slab-on-grade with a separate attic zone –

or a 2-zone model – in TRNSYS Type 56. This house is similar to that simulated by Rudd et al. (2013)

for a recently completed ASHRAE research project (RP-1449). Also, a 3-zone model was developed for 

the house with a basement zone. The basement was not directly conditioned but coupled to the main zone 

via zone-to-zone air exchange. The characteristics of the buildings are listed in Table 4.4 for each 

climate. 
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36

Table 4.4. Specifications for simulated houses (HERS Index ≈100)

a) I-P units

Parameter Houston, TX 

(Climate Zone 2)

Las Vegas, NV 

(Climate Zone 3)

Washington, DC 

(Climate Zone 4)

Chicago, IL 

(Climate Zone 5)*

Wall insulation R-value (nominal) 13 13 13 19

 Cavity 13 13 13 19

 Sheathing 0 0 0 0

 framing factor 0.23 0.23 0.23 0.23

Ceiling insulation R-value 30 30 38 38

Slab insulation R-value (2' down) 0 0 0 0

Basement Walls na na na na

Window U-value (Btu∙h-1∙ft-2∙F-1) 0.75 0.65 0.40 0.35

Window SHGC 0.40 0.40 0.40 0.40

Building enclosure air leakage 

(ACH50) 7 7 7 7

Enclosure ELA (in2) 98.1 98.1 98.1 98.1

Duct air leakage to outside (%) 6% sup, 4% ret 6% sup, 4% ret 6% sup, 4% ret 6% sup, 4% ret

Supply duct area in attic (ft2) 544 544 544 544

Return duct area in attic (ft2) 100 100 100 100

Duct R-value 6 6 6 6

SEER, EER 13, 9.6 13, 9.6 13, 9.6 13, 9.6

HSPF, COP 7.7, 2.3 7.7, 2.3 7.7, 2.3 7.7, 2.3

Internal heat gain (lumped)**

(people+lighting+appliances) 72.70 kBtu/day 72.70 kBtu/day 72.70 kBtu/day 72.70 kBtu/day

Internal moisture generation 12 lb/day 12 lb/day 12 lb/day 12 lb/day

HERS 106 108 108 107

*This house was also used in simulations for Minneapolis, MN (Climate Zone 6)

**DOE Building America benchmark (Hendron, 2008) 

b) SI units

Parameter Houston, TX 

(Climate Zone 2)

Las Vegas, NV 

(Climate Zone 3)

Washington, DC 

(Climate Zone 4)

Chicago, IL 

(Climate Zone 5)*

Wall insulation R(SI)-value (nominal) 2.29 2.29 2.29 3.35

 Cavity 2.29 2.29 2.29 3.35

 Sheathing 0 0 0 0

 framing factor 0.23 0.23 0.23 0.23

Ceiling insulation R(SI)-value 5.38 5.38 6.69 6.69

Slab insulation R(SI)-value (2' down) 0 0 0 0

Basement Walls na na na na

Window U-value (W∙m-2∙K-1) 4.3 3.7 2.3 2.0

Window SHGC 0.40 0.40 0.40 0.40

Building enclosure air leakage 

(ACH50) 7 7 7 7

Enclosure ELA (m2) 0.063 0.063 0.063 0.063

Duct air leakage to outside (%) 6% sup, 4% ret 6% sup, 4% ret 6% sup, 4% ret 6% sup, 4% ret

Supply duct area in attic (m2) 50.5 50.5 50.5 50.5

Return duct area in attic (m2) 9.3 9.3 9.3 9.3

Duct R(SI)-value 1.1 1.1 1.1 1.1

SEER (I-P), COP 13, 9.6 13, 9.6 13, 9.6 13, 9.6

HSPF (I-P), COP 7.7, 2.3 7.7, 2.3 7.7, 2.3 7.7, 2.3

Internal heat gain (lumped)**

(people+lighting+appliances) 76.70 MJ/day 76.70 MJ/day 76.70 MJ/day 76.70 MJ/day

Internal moisture generation 5.4 kg/day 5.4 kg/day 5.4 kg/day 5.4 kg/day

*This house was also used in simulations for Minneapolis, MN (Climate Zone 6)

**DOE Building America benchmark (Hendron, 2008) 
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37

The slab-on-grade house only has perimeter slab insulation in climate zones 4 and 5 (Figure 4.3). For the 

house with a basement (Figure 4.4), the basement is connected to the main house by openings that are 

assumed to allow zone-to-zone air exchange of heat and moisture equivalent to 849.4 m3

∙h-1 (500 cfm).

The basement walls are modeled as 102 mm (4 inch) thick concrete, with R(SI)-1.76 (R-10) exterior foam 

insulation in climate zones 3, 4 and 5. 

Figure 4.3. Schematic of a slab-on-grade house (ducts located in the unconditioned attic)

Both the slab-on-grade and basement homes are modeled by adding a ‘fictitious layer’ into the resistance 

between the zone and ground temperature. This fictitious R-value is added to provide the amount of heat 

loss through the surfaces determined by the F-factor method (Reffective) as recommended by Winkelmann 

(1998). A schematic of this model is shown in Figures 4.3 and 4.4. Tables 4.5 and 4.6 summarize the 

calculations to determine the necessary R-value for the fictitious layer.

The above-ground portions of the slab-on-grade and basement houses are identical for each climate zone. 

Each model has exterior walls with layers of drywall, insulation (R(SI)-2.3 (R-13) or R(SI)-3.3 (R-19), 

depending on the climate zone), and stucco as the outside surface. Windows take up approximately 22 % 

of all of the exterior walls; 10.2 m2 (109.6 ft2

) on the north and south facing walls, and 6.5 m2 (70.4 ft2) on 

east and west facing walls. 

24.5 mm(1 in) carpet

Tground

0.3 m (1 ft)

soil

101 mm (4 in) concrete

Rfic-floor

Supply leak 

to attic

Return leak 

from attic

AHU
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Figure 4.4. Schematic of a house with basement (ducts located in the semi-conditioned basement)

The ceiling (i.e. boundary between main zone and attic) is made up of a layer of drywall, framing and 

insulation (R(SI)-5.3 (R-30) or R(SI)-6.7 (R-38), depending on climate zone). The attic has gable walls 

on the east and west sides and roof surface on the north and south sides. The roof is sheathed in plywood 

and then covered with asphalt shingles. The east and west surfaces (gables) are made up of plywood on 

the inside surface with stucco on the outside surface. 

0.3 m (1 ft) soil 0.3 m (1ft) soil

101 mm (4 in)

Tground

Tground

Tground

Rfic-wall Rfic-wall

Rfic-floor

25.4 mm (1 in) 

plywood

AHU

Return 

leak from 

basement

Supply leak 

to basement

Interzonal 

Air 

Exchange
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39

Table 4.5. Calculation of R-values for basement walls and floor

Basement Wall 

Material Thickness Resistance Total R-Value

R(SI) R

m ft K·m·W-1 h·ft·°F·Btu-1 K·m2 W-1 h·ft2·°F·Btu-1

Concrete 0.10 0.33 0.775 1.33 0.0775 0.44

Soil 0.30 1.00 1.18 2.01 0.354 2.00

Foam 0.035 0.115 25.1 43.5 0.881 5.00

Rfic Massless 0.111 0.63

Reffective 1.42 8.08

Basement Floor

Material Thickness Resistance Total R-Value

R(SI) R

m ft K·m·W-1 h·ft·°F·Btu-1 K·m2 W-1 h·ft2·°F·Btu-1

Concrete 0.10 0.33 0.775 1.33 0.0775 0.44

Soil 0.30 1.00 1.18 2.01 0.354 2.00

Rfic Massless 3.26 18.5

Reffective 3.69 20.95

Table 4.6. Calculation of R-values for slab-on-grade floor

Slab Resistance – Climate Zones 2 and 3

Material Thickness Resistance Total R-Value

R(SI) R

m ft K·m·W-1 h·ft·°F·Btu-1 K·m2 W-1 h·ft2·°F·Btu-1

Carpet 0.025 0.083 14.52 25.13 0.363 2.06

Concrete 0.10 0.33 0.775 1.33 0.0775 0.44

Soil 0.30 1.00 1.18 2.01 0.354 2.01

Rfic Massless 0.958* 5.44*

Reffective 1.75 9.95

Slab Resistance – Climate Zones 4 and 5

Material Thickness Resistance Total R-Value

R(SI) R

m ft K·m·W-1 h·ft·°F·Btu-1 K·m2 W-1 h·ft2·°F·Btu-1

Carpet 0.025 0.083 14.52 25.13 0.363 2.06

Concrete 0.10 0.33 0.775 1.33 0.0775 0.44

Soil 0.30 1.00 1.18 2.01 0.354 2.01

Rfic Massless 2.19* 12.42*

Reffective 2.98 16.93

*The difference in Rfic between climate zones 2/3 and 4/5 is due to the perimeter insulation of the slab in 

climate zones 4 and 5.
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4.3.1 Building Enclosure Air Leakage

The AIM-2 infiltration model (Walker and Wilson, 1998; ASHRAE, 2009a) relates infiltration to wind 

and indoor-outdoor temperature difference for each time step. All simulations in this study used 

coefficients representing shelter from buildings located across the street. An equivalent leakage area 

(ELA) of 0.0633 m2 (98.1 in2

) was chosen to provide the desired seven air changes per hour (ACH) at 50 

pascal pressure differential (ACH50 for the main zone in each building model). 

The attic used the same AIM-2 equations to determine leakage as a function of wind and temperature 

difference. The attic ELA was set to be 0.366 m2 (567 in2

) for each of the climate zones, or about 5 times 

the leakage rate for the HERS 100 house (Fugler, 1999). In houses with basements, that zone was 

assumed to have no leakage to outdoors. 

4.3.2 Duct Leakage and Thermal Losses

For the slab-on-grade houses, the ducts were modeled to be in the attic space and all the air leakage and 

thermal losses/gains go into that zone. The details of the duct model are given in Appendix A. For 

houses with basements, there is no duct leakage to the attic (all leaks are assumed to be into the 

conditioned space, so they are ignored). Duct leakage was assumed to be 10 % of flow, or 6 % on the 

supply side and 4 % on the return side. Duct insulation was assumed to be R(SI)-1.1 (R-6) with a supply 

duct area of 50.5 m2 (544 ft2 ) and a return duct area of 9.3 m2 (100 ft2

) for a 10.6 kW (3-ton) unit. The 

duct areas were increased and decreased proportionally based on the size (or nominal tonnage) of the heat 

pump unit.

4.3.3 Moisture and Thermal Gains

The scheduling or profile of internal heat and moisture generation was taken from the Building America 

Benchmark Definition (Hendron, 2008). Sensible gains from all sources were assumed to be 76.7 MJ/day 

(72.7 kBtu/day).

Internal moisture generation from all sources was specified as 5.4 kg/day (12 lb/day), or less than half of 

the ASHRAE Standard 160 moisture generation rate of 14.2 kg/day (31.2 lb/day) for a three-bedroom 

house (ASHRAE, 2009b). The ASHRAE 160 value is meant to be a ‘worst case’ design condition and 

therefore would not be expected to correspond to average conditions. 

4.3.4 Moisture and Thermal Capacitance

Moisture storage in the building materials and furnishings and the rate of mass transfer into storage are 

important hygrothermal parameters affecting the diurnal swings in indoor humidity. Building material 

moisture storage was modeled with a simple lumped parameter method with mass factor added to the air 

node in the zone model: 

i o internal AC,latent

i ( ) d

d

m w w Q Q

t

w C = ⋅ − + − (4.1)

The moisture capacitance term is usually set to a multiple of the air mass inside the house. The Florida 

Solar Energy Center used more detailed moisture models including Effective Moisture Penetration Depth 

(EMPD) to show that reasonable factors for the air mass multiplier are 20 to 30 times the air mass (EPA, 

2001). 

As a result of the calibration efforts (Appendix C in Rudd et al., 2013), a 30x multiplier for moisture 

capacitance was used for the main zone and the basement. The attic used a moisture capacitance factor of 

15x. 
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41

Thermal capacitance was simulated by adding internal walls to the model with 371.6 m2 (4000 ft2

) of 

exposed wall surface area. The thermal mass of the air node was also increased by a factor of 20 to 

12,331 kJ∙K-1 (6494 Btu∙F-1

) to reflect the impact of furniture and other material in the space. The attic 

was assumed to have a thermal capacitance of 1x and the basement (where applicable) was assumed have 

a thermal capacitance multiplier of 10x. 

4.3.5 Window Performance

The window model in Type 56 uses the window parameters generated by LBNLs WINDOW5 software, 

which is considerably more detailed than the NFRC rating values generally used in residential practice 

and building codes. The LBNL WINDOW5 inputs for this project were determined following the 

methodology developed by Arasteh et al. (2009) for use in EnergyPlus. 

4.4 Mechanical Ventilation

The only mechanical ventilation option considered in this study is an exhaust fan. The fan operated 

continuously to provide sufficient ventilation to the house. Figure 4.5 shows the airflow configuration 

used in this study. The fans provided an average rate of 98.5 m3

∙h-1 (58 cfm) required by ASHRAE 

Standard 62.2 (ASHRAE, 2013) for the 185.8 m2 (2000 ft2

) three-bedroom house. The exhaust fan power 

was assumed to be 0.85 kJ∙m-3 (0.4 W∙cfm-1

).

AHU

Return air

mechanical 

exhaust 

Induced 

infiltration 

infiltration exfiltration

Figure 4.5. Schematic of a mechanical exhaust system 

The combined impact of infiltration, ventilation, and duct leakage were considered by using the equations 

below. The duct leakage was always a net out, so that additional net flow was an exhaust. 

Vin = incoming ventilation flow

Vout = sum of all exhaust flows (exhaust fan, net duct leakage, etc.) 

Vbalanced = MIN (Vin, Vout)

Vunbalanced = MAX (Vin, Vout) – Vbalanced

Vinf = infiltration flow calculated for building for the timestep

Vcombined = MAX (Vunbalanced, Vinf + 0.5∙Vunbalanced) + Vbalanced

The net mechanical inlet flows were subtracted from Vcombined to determine the remaining non-mechanical 

ventilation (or infiltration) rate acting on the building envelope. A mass balance tracked CO2 levels in the 

space and confirmed the net impact of ventilation to be similar between the cases. 
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42

4.5 Air Flow Imbalance

Duct leakage is often exacerbated by interactions with building envelope leakage, depressurization caused 

by exhaust fan operation, and supply and return imbalances caused by closing interior doors (for central 

return systems). Cummings and Tooley (1989) and Modera (1989) both showed that the pressures 

induced by air handlers were much greater than the naturally-induced pressures from wind and stack 

effects in cooling dominated climates. Pressure mapping by Cummings and Tooley (1989) also showed 

that the supply/return imbalances caused by closing interior doors were also substantial.

One option for considering these interactions is to use a multi-zone, flow-pressurization model such as 

CONTAM 3.0 (Walton and Dols, 2010). A model can be developed to evaluate the interactions of 

building envelope leakage paths, duct leakage, and zone pressurization with the supply air (when doors 

are closed). These models can track airflow but cannot consider the thermal performance of the building 

envelope nor the energy use of the space-conditioning systems. 

In a small time-step thermal building simulation model, it is possible to properly account for the 

combined effects of ‘unbalanced’ duct leakage, unbalanced ventilation, and infiltration using a simpler 

approach. The following procedure accounts for the interactions of unbalanced ventilation and duct 

leakage with infiltration. The calculation is based on the approach summarized in Barnaby and Spitler 

(2004) as well as the ASHRAE Handbook of Fundamentals, Chapter 17 (ASHRAE 2009a):

Vin = incoming ventilation airflow

V out = sum of all exhaust airflows (exhaust fan, supply duct leaks, etc.) 

Vbalanced = MIN (Vin, Vout)

Vunbalanced = MAX (Vin, Vout) − Vbalanced

Vinf = infiltration flow calculated for building for the timestep

Vcombined = MAX (Vunbalanced, Vinf + 0.5· Vunbalanced ) + Vbalanced

4.6 Heat Pump Specifications and Modeling

A conventional heat pump unit with a 13 SEER and 7.7 HSPF rating was used in the simulations. The 

cyclic degradation coefficient, CD, of the heat pump was 0.15 in both cooling and heating. The required 

size of the unit was determined for each climate using ACCA Manual J (ACCA 2011a). Houses in 

Houston and Las Vegas had a heat pump with cooling capacity of 10.6 kW (3 ton) and 12.3 kW (3.5 ton), 

respectively. The Washington, D.C., Chicago, and Minneapolis houses had 8.8 kW (2.5-ton) units.

The detailed heat pump model required separate inputs for the gross COP at nominal conditions, sensible 

heat ratio (SHR), and indoor fan power. Table 4.7 lists the rated parameters and corresponding inputs to 

the heat pump model. The fan power assumed for rated conditions and used to calculate SEER is listed 

along with the actual fan power assumed for operation. The fan power at rated conditions was assumed to 

be 0.53 kJ∙m-3 (0.25 W∙cfm-1

), while the actual fan power was 1.06 kJ∙m-3 (0.5 W∙cfm-1).

Table 4.7. Heat pump cooling characteristics

Note: Gross COP is a ratio of gross cooling capacity (refrigerant-side capacity) and outdoor unit power (includes compressor, 

outdoor fan, and controls powers) at the nominal rating point: 35 °C (95 °F) outdoor dry-bulb temperature , 26.7 °C/19.4 °C 

(80 °F/67 °F) indoor dry-bulb/wet-bulb temperature, and 217.4 m3

∙h-1kW-1 (450 cfm/ton) supply airflow.

Unit Description Rated Performance Input Parameters

SEER 13 unit; 

Single-speed

PSC fan motor

Rated SEER

Btu∙W-1∙h-1

Rated COP Rated Fan 

Power

kJ∙m-3

 Gross COP Actual Fan 

Power

kJ∙m-3

SHR

13 2.81 0.53 4.05 1.06 0.77
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43

The airflow in the cooling and heating mode was assumed to be 181.1 m3

∙h-1∙kW-1 (375 cfm/ton). Data 

from the laboratory testing at NIST was used to correct the normalized fan power from the nominal value 

of 1.06 kJ∙m-3 (0.5 W∙cfm-1

) as the airflow changes from the nominal value of 181.1 m3∙h-1∙kW-1

(375 cfm/ton). The data showed a linear trend. The best fit to the measure data (Eq. 4.2) was used to 

predict the variation in fan power as the airflow varies.

 �

Wfan

V � =1.06 -�� V

Qtot

� -181.1�∙0.00316 (4.2)

where �

Wfan

V � = normalized fan power from the nominal value, kJ∙m-3

� V

Qtot

� = airflow to system capacity ratio, m3∙h-1∙kW-1

The heating performance for the heat pump used the generic performance curves developed for 

EnergyGauge (Parker et al., 1999). The generic model is based on catalog data from a series of singlespeed heat pump products (ranging from 10 to 14.5 SEER) and was shown to be appropriate over a range 

of heat pump efficiency levels. The generic model predicts the variation in heating capacity and power 

input as a function of outdoor dry-bulb temperature, indoor entering temperature, and the airflow ratio 

(actual airflow divided by nominal airflow). 

The impact of defrost operation was determined by the defrost degradation function shown in Figure 4.6, 

which is also used in EnergyGauge. This simple function1 predicts the degradation as a function of 

outdoor dry-bulb temperature for a time-initiated, temperature-terminated defrost controller. The impact 

of defrost starts at temperatures below 8.3 °C (47 °F), peaks at 11 % by 2.8 °C (37 °F), and tapers off to 

7.5 % at lower ambient temperatures. For comparison, the graph also includes the degradation rate 

implied by ASHRAE Standard 116 (ASHRAE, 2010). 

The heat pump gross COP at 8.3 °C (47 °F) was 2.7. The nominal gross heating capacity, also at 8.3 °C 

(47 °F), was 10 % greater than the nominal gross cooling capacity. A supplemental 10 kW electric heater

was activated if the space temperature dropped 0.28 °C (0.5 °F) below the heating set point or to 20.3 °C

(68.5 °F) in Chicago, Washington, DC, and Minneapolis. The degraded performance of a heat pump due 

to faults was modeled by applying the heat pump normalized performance parameters described in 

Section 3.

Table 4.8 lists thermostat set points for heating and cooling. The 21.1 °C (70 °F) heating set point was 

selected as appropriate for temperate climates, while the 22.2 °C (72 °F) set point was deemed as more 

appropriate for the warmer climates. The cooling set point of 25.6 °C (78 °F) was selected as most 

consistent with homeowner preferences in warm climates. In colder climates 24.4 °C (76 °F) was used.

The impact of thermostat deadband and anticipator were explicitly considered in this short time-step 

model in the cooling mode as per Henderson (1992). The deadband was ±0.56 °C (1.0 °F) around the 

desired temperature point. The anticipator temperature gain was 1.4 °C (2.5 °F), and the time constant of 

the anticipator was 90 seconds. The sensing element of the thermostat had a time constant of 300 

seconds. The result was the temperature ‘droop’ with runtime fraction of about 1.1 °C (2.0 °F). In the 

heating mode a simple deadband of ±0.6 °C (1.0 °F) around the set point was used without an anticipator

or sensing element time constant.

 1 Actually, defrost is a function of both temperature and ambient humidity. While more sophisticated defrost models 

are available in EnergyPlus (see the 2012 Engineering Reference Manual), these approaches were found to have 

flaws and could not be successfully implemented here for this study.
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44

Figure 4.6. Capacity degradation due to defrost as a function of outdoor temperature

(The different color lines on the plot show the defrost degradation from catalog data. The table of values 

summarizes the average values used in the simulations.)

Table 4.8. Thermostat cooling and heating set points

Zone Location Cooling Set Point 

°C (°F)

Heating Set Point 

°C (°F)

2 Houston, TX 25.6 (78) 22.2 (72) 3 Las Vegas, NV

4 Washington, DC

5 Chicago, IL 24.4 (76) 21.1 (70)

6 Minneapolis, MN

4.7 Cost of Electricity

Total heat pump operating costs were determined using the electric rates listed in Table 4.9.

Table 4.9. Cost of electricity

Zone Location Electric Utility

Cost

$/MJ $/kWh

2 Houston, TX Entergy 0.306 0.085

3 Las Vegas, NV NV Energy 0.454 0.126

4 Washington, DC Pepco 0.508 0.141

5 Chicago, IL ComEd 0.461 0.128

6 Minneapolis, MN Northern States Power 0.389 0.108

Note: Electric costs are from Form 826 data for local utility in 2010 for residential sector (EIA, 2012)

 

-20 0 20

Outdoor Air Temperature (C)

0

5

10

15

Degradation (%)

Defrost Penalty 

from ASHRAE 

Standard 116 

Defrost 

Degradation 

Curve

7.5% - -16.7°C 

8.5% - -8.3°C

11% - -2.8°C

 9% - +2.8°C

 0% - +8.3°C
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45

5. SIMULATIONS OF BUILDING/HEAT PUMP SYSTEMS WITH INSTALLATION 

 FAULTS

Section 4.2 discussed the IECC climate zones and baseline houses considered in this study. The selected 

house options include a slab-on-grade house and a house with a basement for Las Vegas, Washington, 

DC, Chicago, and Minneapolis, and a slab-on-grade house only for Houston. 

The following sections present results of annual simulations of energy consumption for a heat pump 

operating under different levels of different installation faults. These annual simulations focused on 

performance issues of the house/heat pump systems related to heat pump capacity and energy 

consumption while maintaining the target indoor dry-bulb temperature (shown in Table 4.8) within the 

temperature band imposed by the thermostat. For a few faults, we performed additional annual 

simulations with a lowered thermostat set-point temperature to mimic this common response to elevated 

indoor humidity levels caused by installation faults.

Results of annual simulations of energy consumptions are presented in the format consistent with Table 

5.1. The threshold 55 % relative humidity value used in the third column was selected as the level above 

which humidity might start to be a concern. This threshold is slightly lower than the limit of 60 % relative 

humidity, which has historically been identified as the space condition where mold growth can occur in 

the building envelope (Sterling et al., 1985). The ‘Space Temp Max’ column contains the highest indoor 

temperature reached during the cooling season. The column ‘AC Energy’ contains the energy used by the 

compressor and outdoor fan to provide cooling; the column ‘Htg Energy’ contains the energy used by the 

compressor, outdoor fan, and backup heat to provide heating; and the column ‘AHU Fan Energy’ contains 

the energy used by the indoor fan during the whole year. The column ‘TOTAL ENERGY’ contains the 

total energy used by the heat pump throughout the entire year, which consists of the energy use listed in 

the three previous columns, and the energy used by the home exhaust fan.

5.1 Annual Energy Consumption in Baseline Houses 

Table 5.1 presents simulation results of the annual heat pump operating energy consumption, energy cost, 

and relative energy cost referenced to that of the slab-on-grade house for each locality. The energy use of 

the basement house is from 17 % to 19 % lower than that for the slab-on-grade house in most climates. 

Most of this difference is due to duct leakage: the basement house has the ducts in the basement (with no 

losses) instead of ducts in the attic for the slab-on-grade house with the assumed typical leakage of 10 % 

(Section 4.3.2). 

Table 5.1 also includes results from additional runs for the slab-on-grade house without duct leakage (i.e., 

treated as ducts in the conditioned space) denoted in the table as ‘Slab, Ducts Inside’. When the duct 

leakage and duct thermal losses are eliminated, the slab-on-grade and basement houses perform within 

3 % for Las Vegas and within 9 % for the cold climates. The basement house does have higher energy 

use in the colder climates. 
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46

Table 5.1. Annual energy consumption and cost in baseline houses

5.2 Simulation with Single Faults

5.2.1 Studied Faults

Table 5.2 summarizes the studied faults and their level values used in simulations.

Table 5.2. Studied faults in the cooling and heating mode

Fault Type Fault Levels (%)

Cooling mode Heating mode

Heat Pump Sizing (SIZ) -20, 25, 50, 75, 100 -20, 25, 50, 75, 100

Duct Leakage (DUCT) 0, 10, 20, 30, 40, 50 0, 10, 20, 30, 40, 50

Indoor Coil Airflow (AF) -36, -15, 7, 28 -36, -15, 7, 28

Refrigerant Undercharge (UC) -10, -20, -30 -10, -20, -30

Refrigerant Overcharge (OC) 10, 20, 30 10, 20, 30

Excessive Refrigerant Subcooling (SC) 100, 200 -

Non-Condensable Gases (NC) 10, 20 10, 20

Electric Voltage (VOL) -8, 8, 25 -8, 8, 25

TXV Undersizing (TXV) -60, -40, -20 -

5.2.2 Effect of Heat Pump Sizing

Changing the size of the heat pump for a given house – either undersizing or oversizing – impacts the heat 

pump performance in several ways:

• Cycling losses increase as the unit gets larger; the unit runs for shorter periods and the degraded 

performance at startup has more impact (parameters used in simulations are: time constant = 45 

seconds, or CD ~ 0.15). 

• In the cooling mode, the shorter run periods impact the moisture removal capability (i.e., ability 

to control indoor humidity levels) because operational steady-state conditions are an even smaller 

portion of the runtime fraction.

• In the cooling mode, continuous fan operation with compressor cycling greatly increases moisture 

evaporation from the cooling coil. However, this impact is minimal with auto fan control (indoor 

fan time ‘on’ and ‘off’ the same as that of the compressor), since only a small amount of 

evaporation occurs with the assumed 4 % airflow during the off-cycle with the indoor fan off. If 

the air conditioner controls include an off-cycle fan delay – that keeps the fan on for 30-90 

 Hours 

Above 

55 % 

RH 

 Space 

Temp 

Max 

(C) 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC 

SHR (-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU 

Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

Slab-on-Grade 1,512 26.6 1,981 749 5.1 2,730.5 4.3 0.785 16,660 8,537 5,529 31,457 $743 100%

Basement

Slab, Ducts inside 1,715 25.2 1,555 588 0.3 2,142.9 4.3 0.789 13,007 6,623 4,339 24,700 $583 79%

Slab-on-Grade - 27.0 1,966 865 0.3 2,831.1 3.7 0.999 20,531 11,251 6,687 39,200 $1,372 100%

Basement - 25.3 1,552 718 0.3 2,269.9 3.7 1.000 16,107 9,407 5,362 31,607 $1,106 81%

Slab, Ducts inside - 25.3 1,536 668 0.3 2,204.5 3.7 1.000 15,941 8,763 5,207 30,642 $1,072 78%

Slab-on-Grade 253 25.1 1,207 1,971 89.0 3,178.0 4.5 0.809 8,098 21,759 5,363 35,952 $1,408 100%

Basement 654 24.0 742 1,907 27.0 2,649.5 4.4 0.775 5,008 19,120 4,471 29,330 $1,149 82%

Slab, Ducts inside 280 24.1 944 1,532 12.9 2,476.3 4.4 0.801 6,301 15,111 4,179 26,322 $1,031 73%

Slab-on-Grade 189 25.0 1,031 2,833 281.2 3,863.9 4.5 0.827 6,816 37,118 6,520 51,186 $1,820 100%

Basement 289 24.0 631 2,785 129.8 3,416.1 4.4 0.797 4,198 31,565 5,765 42,259 $1,503 83%

Slab, Ducts inside 203 24.0 815 2,288 70.0 3,103.7 4.5 0.819 5,369 24,753 5,238 36,092 $1,283 71%

Slab-on-Grade 13 25.2 897 3,432 612.5 4,328.9 4.5 0.846 5,912 55,105 7,305 69,053 $2,072 100%

Basement 61 24.0 515 3,424 354.2 3,939.8 4.4 0.810 3,428 46,239 6,648 57,048 $1,711 83%

Slab, Ducts inside 15 24.1 711 2,902 216.9 3,613.5 4.5 0.838 4,670 36,410 6,098 47,909 $1,437 69%

Chicago

Washington, DC

Las Vegas

Houston

Minneapolis
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47

seconds after the compressor stops – then the impact of off-cycle evaporation is in between these 

two extremes (Shirey et al., 2006). The results in this study assumed auto fan operation with no 

fan delay.

• In the heating mode, the backup heater runtime is lower for the oversized unit since the larger 

heat pump meets more of the winter heating needs.

Heat pump sizing also affects the level of duct losses. This study considered two heat pump sizing 

scenarios with regard to the sizing of the air duct. In scenario (1), the heat pump and air duct are 

proportionally undersized or oversized, i.e., the duct flow area increases proportionally to the increase of 

heat pump capacity. As a result, the air mass flux through the duct remains unchanged, and the duct 

surface area increases with the square root of capacity ratio (unit capacity/design building load). The duct 

losses to the attic (thermal and air leak losses) tend to increase with the unit size since the surface area of 

the duct and the amount of airflow increases; however, the lower indoor fan runtime associated with an 

oversized heat pump has the opposing influence (reduces duct losses to the attic) since in the model the 

losses only occur when the fan is ‘on’. Not included in this analysis is the impact that oversizing has on 

moisture control, especially at part load (see Sonne et al. (2006) for an in-depth review on this topic).

In scenario (2), the duct has been sized for a heat pump of nominal capacity and remains unchanged for 

different size heat pumps. When the heat pump is oversized, the fan speed is increased but the airflow 

does not reach the target flow rate because the unit is not capable of overcoming the increased external 

static pressure. Since the indoor fan works against increased static pressure, the fan power changes per the 

fan curve, i.e., fan power increases with an increasing unit size. The increased pressure in the duct 

increases the duct leakage. Table 5.3 shows the realized airflow per unit capacity, external static pressure, 

and duct leakage for scenario (1) and scenario (2).

Table 5.4 compares the effect of 100 % oversizing on the cooling and the heating performance for the

slab-on-grade house for the five studied cities and two oversizing scenarios. For scenario (1) - duct size 

changes - oversizing degrades the cooling COP only modestly (about 2 %). The thermostat has ‘droop’

that causes the average space temperature to drop by (1.1 ~ 1.7) °C ((2 ~ 3) °F) with lower runtime 

fractions. In addition, the larger ducts have more losses to the uninsulated attic, but the shorter indoor 

runtime has the opposing effect. The net effect is that the energy use in the cooling mode increases by 

(2 ~ 3) %. In the heating mode, the larger heat pump meets more of the space heating load, so less 

operation of the inefficient auxiliary resistance heater is required. As a result, the heating energy 

decreases by (3 ~ 4) % in the cooling-dominated climates and almost 9 % in the heating-dominated 

climates. Overall, the total annual energy use is barely affected in the cooling-dominated climates and 

decreases in the heating dominated climates by about 4 %. Note, that the simulations in this section use a 

duct leakage rate of 10 %, which is assumed to be a ‘no fault’ installation condition. For scenario (2) - no 

change in duct size -, the increased fan power (while working against increased static pressure) and fan 

heat added to the load are the main factors contributing to the significant increase in energy used in 

cooling-dominated climates (Houston, Las Vegas, Washington, DC). 
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48

Tables 5.3. Indoor airflow information for heat pump sizing scenario (1) and scenario (2)

a) SI units Heat Pump 

Sizing

(%)

Fan Speed

(%)

Normalized 

Airflow

(m3∙h-1∙kW-1)

Normalized 

Fan Power 

(kJ∙m-3)

Static 

Pressure

(Pa)

Duct Leakage 

(%)

Supply Return

Scenario (1):

Duct size 

changes

proportionally 

with HP size

80 100 181.1 1.06 167 6 4

100 100 181.1 1.06 167 6 4

125 100 181.1 1.06 167 6 4

150 100 181.1 1.06 167 6 4

175 100 181.1 1.06 167 6 4

200 100 181.1 1.06 167 6 4

Scenario (2):

Duct size 

stays the same 

as HP size 

changes

80 90 202.4 0.80 137 5.4 3.6

100 100 181.1 1.06 167 6.0 4.0

125 115 168.1 1.45 224 7.0 4.6

150 120 145.5 1.68 249 7.3 4.9

175 125 130.9 1.90 274 7.7 5.1

200 130 120.8 2.11 299 8.0 5.4

b) I-P units Heat Pump 

Sizing

(%)

Fan Speed

(%)

Normalized 

Airflow

(cfm/ton)

Normalized 

Fan Power 

(W∙cfm-1)

Static 

Pressure

(inch)

Duct Leakage 

(%)

Supply Return

Scenario (1):

Duct size 

changes

proportionally 

with HP size

80 100 375 0.50 0.76 6 4

100 100 375 0.50 0.76 6 4

125 100 375 0.50 0.76 6 4

150 100 375 0.50 0.76 6 4

175 100 375 0.50 0.76 6 4

200 100 375 0.50 0.76 6 4

Scenario (2):

Duct size 

stays the same 

as HP size 

changes

80 90 419 0.38 0.55 5.4 3.6

100 100 375 0.50 0.67 6.0 4.0

125 115 348 0.68 0.90 7.0 4.6

150 120 301 0.79 1.00 7.3 4.9

175 125 271 0.89 1.10 7.7 5.1

200 130 250 0.99 1.20 8.0 5.4
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49

Table 5.4. Effect of 100 % unit oversizing on annual energy use for a slab-on-grade house for scenario (1) 

and scenario (2)

Scenario (1):

Duct size 

changes

proportionally 

with HP size

Cooling COP 

(%)

Cooling

Load

(%)

Cooling 

Energy

(%)

Heating 

Energy

(%)

Total

Energy

(%)

Houston -2.0 1.2 3.3 -4.1 0.9

Las Vegas -2.5 -0.6 1.9 -3.3 0.1

Washington -1.9 0.3 2.2 -7.9 -3.6

Chicago -1.8 0.0 1.8 -8.9 -4.6

Minneapolis -1.7 0.2 2.0 -8.6 -4.3

Scenario (2):

Duct size 

stays the same 

as HP size 

changes

Cooling COP

(%)

Cooling

Load

(%)

Cooling 

Energy

(%)

Heating 

Energy

(%)

Total

Energy

(%)

Houston -10.3 9.6 22.2 -0.6 24.2

Las Vegas -11.9 5.6 19.8 2.2 21.7

Washington -10.3 9.6 22.1 -10.9 8.0

Chicago -10.2 10.2 22.7 -13..5 2.1

Minneapolis -10.2 10.8 23.4 -14.2 -0.9

Tables 5.5 and 5.6 show in detail the effect of heat pump sizing on the total energy performance for 

scenario (1). The impact of oversizing is modest for the house with the basement (Table 5.6) since the 

ducts are in the conditioned space. In this case, oversizing increases cooling energy because of efficiency 

losses from cyclic degradation, therefore overall energy use in cooling-dominated locations such as 

Houston and Las Vegas increases. In the heating-dominated climates, such as Chicago, the heating 

energy is affected by cyclic degradation as well; however, the larger heat pump meets more of the heating 

load, which reduces the need for backup heating. The net effect is a slight decrease in overall energy use.

For the slab-on-grade house (Table 5.5) the impact of duct leakage further complicates the situation. In 

addition to the factors discussed for the house with the basement, oversized heat pumps have reduced 

runtimes, which reduce duct losses and result in a less energy being used than by the baseline system. 

Combining all effects, the net impact on energy use in Houston and Las Vegas is neutral. In Chicago, 

oversizing actually reduces energy use by as much as 5 % for the slab-on-grade house. 

Tables 5.7 and 5.8 show in detail the effect of sizing on the total performance for scenario (2), and 

Figures 5.1 and 5.2 show relative energy input for the slab-on-grade house and house with a basement, 

respectively. The indoor fan power changes associated with heat pump sizing have proportionally bigger 

impact in the basement house then the slab-on-grade house since the cooling loads are smaller. In heating,

the added fan power from oversizing in the basement house attenuates the drop in heating energy. The 

houses located in cooling dominated climates use less energy when the heat pump is undersized because 

the heat pump does not handle all the cooling load (the indoor temperature increases on hot days). For the 

heating dominated climates, the energy use is increased because of the significantly increased use of the 

resistant heater.
This publication is available free of charge from: http://dx.doi.org/10.6028/NIST.TN.1848

50

Table 5.5. Effect of heat pump sizing on annual energy use for a slab-on-grade house with duct sized to 

match heat pump size (scenario (1))

Houston

 Hours 

Above 

55 % 

RH 

 Space 

Temp 

Max 

(C) 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC 

SHR (-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU 

Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

Undersized 20 % 1,521 27.7 2,401 918 15.1 3,319.0 4.4 0.784 16,078 8,710 5,377 30,897 $730 98%

Normal 1,512 26.6 1,981 749 5.1 2,730.5 4.3 0.785 16,660 8,537 5,529 31,457 $743 100%

Oversized 25 % 1,527 25.6 1,606 597 1.0 2,202.5 4.3 0.785 16,901 8,369 5,575 31,577 $746 100%

Oversized 50 % 1,544 25.3 1,347 493 0.3 1,840.0 4.3 0.784 17,012 8,283 5,589 31,616 $746 101%

Oversized 75 % 1,561 25.1 1,162 420 0.2 1,581.6 4.3 0.784 17,119 8,232 5,605 31,687 $748 101%

Oversized 100 % 1,587 25.1 1,022 365 0.2 1,387.3 4.3 0.785 17,213 8,191 5,618 31,754 $750 101%

Las Vegas

 Hours 

Above 

55 % 

RH 

 Space 

Temp 

Max 

(C) 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC 

SHR (-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU 

Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

Undersized 20 % - 28.5 2,376 1,095 2.4 3,470.2 3.8 0.999 19,716 11,448 6,559 38,455 $1,346 98%

Normal - 27.0 1,966 865 0.3 2,831.1 3.7 0.999 20,531 11,251 6,687 39,200 $1,372 100%

Oversized 25 % - 25.6 1,587 680 0.3 2,267.3 3.7 0.999 20,758 11,070 6,696 39,256 $1,374 100%

Oversized 50 % - 25.2 1,326 562 0.3 1,887.8 3.7 0.999 20,806 10,983 6,690 39,210 $1,372 100%

Oversized 75 % - 25.1 1,140 479 0.2 1,618.7 3.7 0.999 20,863 10,927 6,692 39,215 $1,373 100%

Oversized 100 % - 25.1 1,000 417 0.2 1,417.4 3.6 1.000 20,926 10,877 6,697 39,232 $1,373 100%

Washington, DC

 Hours 

Above 

55 % 

RH 

 Space 

Temp 

Max 

(C) 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC 

SHR (-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU 

Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

Undersized 20 % 238 25.9 1,480 2,330 147.7 3,810.1 4.5 0.809 7,909 22,789 5,144 36,573 $1,432 102%

Normal 253 25.1 1,207 1,971 89.0 3,178.0 4.5 0.809 8,098 21,759 5,363 35,952 $1,408 100%

Oversized 25 % 276 24.3 974 1,633 46.6 2,606.4 4.4 0.808 8,178 20,958 5,498 35,365 $1,385 98%

Oversized 50 % 280 24.0 815 1,385 22.3 2,199.9 4.4 0.809 8,216 20,487 5,568 35,004 $1,371 97%

Oversized 75 % 287 24.0 701 1,197 9.0 1,898.1 4.4 0.809 8,251 20,223 5,605 34,811 $1,363 97%

Oversized 100 % 303 23.9 616 1,049 1.8 1,664.9 4.4 0.809 8,280 20,044 5,619 34,674 $1,358 96%

Chicago

 Hours 

Above 

55 % 

RH 

 Space 

Temp 

Max 

(C) 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC 

SHR (-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU 

Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

Undersized 20 % 189 26.0 1,269 3,178 421.6 4,446.0 4.6 0.827 6,690 39,279 6,002 52,703 $1,874 103%

Normal 189 25.0 1,031 2,833 281.2 3,863.9 4.5 0.827 6,816 37,118 6,520 51,186 $1,820 100%

Oversized 25 % 193 24.2 830 2,430 178.3 3,259.1 4.5 0.827 6,863 35,494 6,875 49,963 $1,776 98%

Oversized 50 % 193 24.0 694 2,120 112.0 2,813.7 4.5 0.827 6,892 34,615 7,122 49,361 $1,755 96%

Oversized 75 % 190 24.0 597 1,860 76.2 2,456.5 4.4 0.827 6,916 34,162 7,254 49,065 $1,745 96%

Oversized 100 % 197 23.9 524 1,648 53.4 2,171.7 4.4 0.827 6,941 33,826 7,329 48,828 $1,736 95%

Minneapolis

 Hours 

Above 

55 % 

RH 

 Space 

Temp 

Max 

(C) 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC 

SHR (-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU 

Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

Undersized 20 % 12 26.1 1,107 3,727 827.2 4,833.6 4.5 0.847 5,819 58,359 6,525 71,436 $2,143 103%

Normal 13 25.2 897 3,432 612.5 4,328.9 4.5 0.846 5,912 55,105 7,305 69,053 $2,072 100%

Oversized 25 % 15 24.3 722 3,066 432.9 3,788.6 4.5 0.846 5,958 52,707 7,992 67,388 $2,022 98%

Oversized 50 % 15 24.1 604 2,735 322.0 3,339.1 4.5 0.846 5,981 51,531 8,452 66,696 $2,001 97%

Oversized 75 % 15 24.0 521 2,440 253.2 2,960.7 4.4 0.847 6,012 50,862 8,743 66,349 $1,990 96%

Oversized 100 % 16 23.9 457 2,195 203.5 2,651.8 4.4 0.848 6,028 50,388 8,950 66,098 $1,983 96%
This publication is available free of charge from: http://dx.doi.org/10.6028/NIST.TN.1848

51

Table 5.6. Effect of heat pump sizing on annual energy use for a house with basement with duct sized to 

match heat pump size (scenario (1))

Las Vegas

 Hours 

Above 

55 % 

RH 

 Space 

Temp 

Max 

(C) 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC 

SHR (-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU 

Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

Undersized 20 % - 26.0 1,908 898 0.3 2,806.8 3.7 1.000 15,839 9,411 5,305 31,287 $1,095 99%

Normal - 25.3 1,552 718 0.3 2,269.9 3.7 1.000 16,107 9,407 5,362 31,607 $1,106 100%

Oversized 25 % - 25.2 1,260 574 0.2 1,834.1 3.6 1.000 16,354 9,413 5,416 31,915 $1,117 101%

Oversized 50 % - 25.1 1,064 479 0.2 1,542.6 3.6 1.000 16,568 9,430 5,467 32,196 $1,127 102%

Oversized 75 % - 25.0 920 411 0.2 1,331.1 3.6 1.000 16,735 9,434 5,503 32,404 $1,134 103%

Oversized 100 % - 24.9 812 360 0.2 1,171.9 3.6 1.000 16,871 9,452 5,537 32,592 $1,141 103%

Washington, DC

 Hours 

Above 

55 % 

RH 

 Space 

Temp 

Max 

(C) 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC 

SHR (-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU 

Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

Undersized 20 % 647 24.2 914 2,294 62.1 3,208.5 4.4 0.773 4,930 19,645 4,331 29,638 $1,161 101%

Normal 654 24.0 742 1,907 27.0 2,649.5 4.4 0.775 5,008 19,120 4,471 29,330 $1,149 100%

Oversized 25 % 666 23.9 603 1,562 6.5 2,164.2 4.4 0.776 5,085 18,867 4,565 29,249 $1,146 100%

Oversized 50 % 669 23.8 507 1,313 0.5 1,820.1 4.3 0.778 5,142 18,855 4,607 29,336 $1,149 100%

Oversized 75 % 677 23.8 439 1,128 0.1 1,566.8 4.3 0.779 5,196 18,899 4,627 29,454 $1,154 100%

Oversized 100 % 694 23.7 387 986 0.1 1,373.4 4.3 0.780 5,236 18,897 4,635 29,500 $1,155 101%

Chicago

 Hours 

Above 

55 % 

RH 

 Space 

Temp 

Max 

(C) 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC 

SHR (-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU 

Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

Undersized 20 % 276 24.2 777 3,199 240.2 3,976.0 4.5 0.795 4,131 33,239 5,368 43,470 $1,546 103%

Normal 289 24.0 631 2,785 129.8 3,416.1 4.4 0.797 4,198 31,565 5,765 42,259 $1,503 100%

Oversized 25 % 287 23.9 512 2,351 60.6 2,862.8 4.4 0.799 4,259 30,692 6,039 41,721 $1,483 99%

Oversized 50 % 285 23.8 431 2,010 31.1 2,441.1 4.4 0.800 4,308 30,474 6,179 41,693 $1,482 99%

Oversized 75 % 285 23.8 373 1,745 15.4 2,117.8 4.4 0.801 4,344 30,363 6,254 41,693 $1,482 99%

Oversized 100 % 292 23.7 328 1,541 6.5 1,868.8 4.4 0.803 4,373 30,362 6,307 41,774 $1,485 99%

Minneapolis

 Hours 

Above 

55 % 

RH 

 Space 

Temp 

Max 

(C) 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC 

SHR (-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU 

Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

Undersized 20 % 61 24.3 635 3,793 551.4 4,428.1 4.4 0.809 3,375 49,371 5,978 59,456 $1,784 104%

Normal 61 24.0 515 3,424 354.2 3,939.8 4.4 0.810 3,428 46,239 6,648 57,048 $1,711 100%

Oversized 25 % 64 24.0 418 2,984 219.1 3,401.9 4.4 0.812 3,480 44,581 7,176 55,969 $1,679 98%

Oversized 50 % 66 23.9 352 2,602 147.8 2,954.8 4.4 0.814 3,520 43,888 7,479 55,619 $1,669 97%

Oversized 75 % 70 23.8 304 2,299 99.1 2,603.2 4.4 0.816 3,548 43,438 7,688 55,405 $1,662 97%

Oversized 100 % 70 23.7 268 2,057 65.8 2,324.7 4.4 0.817 3,570 43,260 7,846 55,408 $1,662 97%
This publication is available free of charge from: http://dx.doi.org/10.6028/NIST.TN.1848

52

Table 5.7. Effect of heat pump sizing on annual energy use for a slab-on-grade house with fixed duct size

(scenario (2)) 

Houston

 Hours 

Above 

55 % 

RH 

 Space 

Temp 

Max (C) 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC 

SHR (-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU 

Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

Undersized 20 % 1,692 27.6 2,317 906 14.4 3,223.4 4.5 0.790 15,621 8,583 4,292 29,227 $690 93%

Normal 1,512 26.6 1,981 749 5.1 2,730.5 4.3 0.785 16,660 8,537 5,529 31,457 $743 100%

Oversized 25 % 1,443 25.8 1,687 605 1.0 2,292.1 4.3 0.780 17,709 8,481 7,330 34,252 $809 109%

Oversized 50 % 1,320 25.3 1,492 505 0.3 1,996.8 4.1 0.774 18,713 8,470 8,147 36,062 $851 115%

Oversized 75 % 1,244 25.2 1,343 433 0.2 1,775.8 4.0 0.769 19,587 8,482 8,832 37,634 $889 120%

Oversized 100 % 1,205 25.1 1,224 379 0.2 1,602.7 3.9 0.766 20,351 8,483 9,513 39,078 $923 124%

Las Vegas

 Hours 

Above 

55 % 

RH 

 Space 

Temp 

Max (C) 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC 

SHR (-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU 

Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

Undersized 20 % - 28.3 2,303 1,109 3.0 3,411.4 3.9 1.000 19,197 11,620 5,060 36,608 $1,281 93%

Normal - 27.0 1,966 865 0.3 2,831.1 3.7 0.999 20,531 11,251 6,687 39,200 $1,372 100%

Oversized 25 % - 25.9 1,658 708 0.3 2,366.3 3.6 0.999 21,651 11,509 8,447 42,339 $1,482 108%

Oversized 50 % - 25.2 1,461 588 0.3 2,048.9 3.5 0.997 22,773 11,483 9,346 44,335 $1,552 113%

Oversized 75 % - 25.2 1,311 505 0.2 1,815.7 3.4 0.995 23,742 11,495 10,108 46,077 $1,613 118%

Oversized 100 % - 25.2 1,194 442 0.2 1,635.9 3.3 0.993 24,605 11,504 10,878 47,719 $1,670 122%

Washington, DC

 Hours 

Above 

55 % 

RH 

 Space 

Temp 

Max (C) 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC 

SHR (-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU 

Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

Undersized 20 % 326 25.7 1,424 2,266 138.8 3,689.5 4.6 0.816 7,658 22,274 4,795 35,459 $1,389 99%

Normal 253 25.1 1,207 1,971 89.0 3,178.0 4.5 0.809 8,098 21,759 5,363 35,952 $1,408 100%

Oversized 25 % 216 24.4 1,025 1,604 42.9 2,628.8 4.4 0.804 8,581 20,448 7,423 37,184 $1,456 103%

Oversized 50 % 134 24.1 907 1,363 17.4 2,269.7 4.2 0.796 9,077 19,830 8,047 37,685 $1,476 105%

Oversized 75 % 89 24.1 817 1,181 5.6 1,998.2 4.1 0.790 9,511 19,567 8,464 38,273 $1,499 106%

Oversized 100 % 59 24.0 745 1,036 0.7 1,781.8 4.0 0.785 9,892 19,396 8,822 38,841 $1,521 108%

Chicago

 Hours 

Above 

55 % 

RH 

 Space 

Temp 

Max (C) 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC 

SHR (-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU 

Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

Undersized 20 % 187 25.8 1,215 3,123 399.8 4,338.1 4.6 0.836 6,451 38,633 5,807 51,622 $1,835 101%

Normal 189 25.0 1,031 2,833 281.2 3,863.9 4.5 0.827 6,816 37,118 6,520 51,186 $1,820 100%

Oversized 25 % 192 24.3 874 2,398 164.7 3,272.0 4.4 0.821 7,209 34,553 9,278 51,772 $1,841 101%

Oversized 50 % 183 24.1 777 2,086 97.3 2,862.8 4.3 0.811 7,657 33,246 10,145 51,779 $1,841 101%

Oversized 75 % 183 24.0 700 1,829 63.6 2,529.5 4.1 0.803 8,033 32,604 10,640 52,008 $1,849 102%

Oversized 100 % 173 24.0 640 1,623 40.8 2,262.7 4.1 0.798 8,365 32,100 11,041 52,238 $1,857 102%

Minneapolis

 Hours 

Above 

55 % 

RH 

 Space 

Temp 

Max (C) 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC 

SHR (-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU 

Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

Undersized 20 % 16 26.0 1,059 3,673 797.0 4,732.5 4.6 0.857 5,604 57,677 6,432 70,445 $2,113 102%

Normal 13 25.2 897 3,432 612.5 4,328.9 4.5 0.846 5,912 55,105 7,305 69,053 $2,072 100%

Oversized 25 % 13 24.5 761 3,033 408.9 3,793.4 4.4 0.840 6,255 51,321 10,781 69,089 $2,073 100%

Oversized 50 % 11 24.1 677 2,694 293.2 3,371.5 4.3 0.828 6,659 49,336 11,945 68,672 $2,060 99%

Oversized 75 % 8 24.0 612 2,403 224.3 3,014.5 4.1 0.820 7,000 48,228 12,630 68,589 $2,058 99%

Oversized 100 % 5 24.0 560 2,161 171.5 2,720.2 4.0 0.814 7,295 47,255 13,164 68,445 $2,053 99%
This publication is available free of charge from: http://dx.doi.org/10.6028/NIST.TN.1848

53

Table 5.8. Effect of heat pump sizing on annual energy use for a house with basement with fixed duct 

size (scenario (2)) 

Figure 5.1. Annual energy use for slab-on-grade houses for different heat pump sizings, scenario (2) 

Las Vegas

 Hours 

Above 

55 % 

RH 

 Space 

Temp 

Max (C) 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC 

SHR (-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU 

Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

Undersized 20 % - 25.8 1,835 904 0.3 2,739.1 3.8 1.000 15,274 9,464 4,062 29,531 $1,034 93%

Normal - 25.3 1,552 718 0.3 2,269.9 3.7 1.000 16,107 9,407 5,362 31,607 $1,106 100%

Oversized 25 % - 25.2 1,313 591 0.3 1,903.5 3.6 1.000 17,007 9,695 6,773 34,207 $1,197 108%

Oversized 50 % - 25.1 1,167 497 0.2 1,664.2 3.4 1.000 18,072 9,796 7,564 36,163 $1,266 114%

Oversized 75 % - 25.1 1,054 429 0.2 1,483.1 3.3 0.999 18,965 9,870 8,223 37,790 $1,323 120%

Oversized 100 % - 25.0 963 377 0.2 1,340.7 3.2 0.997 19,736 9,918 8,875 39,260 $1,374 124%

Washington, DC

 Hours 

Above 

55 % 

RH 

 Space 

Temp 

Max (C) 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC 

SHR (-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU 

Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

Undersized 20 % 863 24.0 873 2,212 53.8 3,085.1 4.5 0.779 4,734 18,958 4,126 28,550 $1,118 97%

Normal 654 24.0 742 1,907 27.0 2,649.5 4.4 0.775 5,008 19,120 4,471 29,330 $1,149 100%

Oversized 25 % 514 24.0 630 1,535 4.6 2,165.2 4.3 0.775 5,301 18,466 6,135 30,634 $1,200 104%

Oversized 50 % 359 23.9 563 1,302 0.5 1,865.1 4.1 0.768 5,663 18,592 6,610 31,596 $1,238 108%

Oversized 75 % 263 23.9 509 1,128 0.1 1,637.6 4.0 0.764 5,965 18,725 6,900 32,323 $1,266 110%

Oversized 100 % 216 23.8 467 995 0.1 1,461.3 3.9 0.761 6,231 18,818 7,159 32,940 $1,290 112%

Chicago

 Hours 

Above 

55 % 

RH 

 Space 

Temp 

Max (C) 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC 

SHR (-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU 

Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

Undersized 20 % 318 24.1 741 3,119 213.1 3,860.3 4.5 0.804 3,963 32,140 5,292 42,126 $1,498 100%

Normal 289 24.0 631 2,785 129.8 3,416.1 4.4 0.797 4,198 31,565 5,765 42,259 $1,503 100%

Oversized 25 % 263 24.0 536 2,313 54.7 2,848.6 4.3 0.795 4,446 29,982 8,099 43,258 $1,538 102%

Oversized 50 % 237 23.9 480 1,991 28.8 2,470.1 4.2 0.787 4,758 29,840 8,751 44,080 $1,567 104%

Oversized 75 % 224 23.8 435 1,745 13.8 2,179.5 4.1 0.780 5,019 29,858 9,130 44,738 $1,591 106%

Oversized 100 % 213 23.8 398 1,550 6.1 1,948.1 4.0 0.776 5,243 29,944 9,426 45,345 $1,612 107%

Minneapolis

 Hours 

Above 

55 % 

RH 

 Space 

Temp 

Max (C) 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC 

SHR (-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU 

Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

Undersized 20 % 72 24.2 604 3,724 505.2 4,327.8 4.5 0.818 3,230 47,917 6,016 57,894 $1,737 101%

Normal 61 24.0 515 3,424 354.2 3,939.8 4.4 0.810 3,428 46,239 6,648 57,048 $1,711 100%

Oversized 25 % 58 24.0 438 2,943 204.3 3,381.6 4.3 0.808 3,636 43,494 9,633 57,495 $1,725 101%

Oversized 50 % 54 24.0 392 2,580 139.4 2,972.0 4.2 0.798 3,894 42,806 10,527 57,959 $1,739 102%

Oversized 75 % 50 23.9 356 2,297 93.9 2,653.1 4.1 0.791 4,110 42,474 11,075 58,390 $1,752 102%

Oversized 100 % 49 23.8 327 2,067 62.8 2,393.8 4.0 0.786 4,300 42,343 11,497 58,871 $1,766 103%

90

95

100

105

110

115

120

125

130

Houston Las Vegas Wash., DC Chicago Minneapolis

Relative energy use (%)

-20% Nominal 25% 50% 75 100%
This publication is available free of charge from: http://dx.doi.org/10.6028/NIST.TN.1848

54

Figure 5.2. Annual energy use for houses with basement for different heat pump sizings, scenario (2) 

5.2.3 Effect of Duct Leakage

Per the earlier discussion in Section 4.3.2, the effect of duct leakage has been evaluated only for slab-ongrade houses where ducts were installed in the attic (i.e., in the unconditioned space). The baseline 

houses include ducts in the attic with a leakage rate of 10 % (leakage distributed 60 % on the supply side 

and 40 % on the return side) as well as thermal losses through the duct wall. Table 5.9 compares this base 

case to other levels of duct leakage with the thermostat set at the default set point temperature (Table 4.8). 

The entry ‘0 % & No thermal’ in the left most column denotes an idealistic installation with zero air 

leakage and no thermal loss (i.e., an insulation with an infinite R). For all other simulation cases the duct 

insulation is assumed to be R(SI)-1.1 (R-6). 

As expected, the baseline duct losses increase energy use in the baseline houses; our simulations showed 

a 20 % and 30 % increase for the cooling climates and heating climates, respectively, compared to the 

0 % leak case. As the duct leakage increases, energy use increases by at least 8 % for the cooling 

climates and by 12 % for the heating climates for each 10 % increment in the duct leakage fault. A slight 

improvement of the cooling COP shown with the increasing fault level is caused by a somewhat higher

refrigerant saturation temperature (and pressure) in the evaporator when the air returning to the indoor 

section is at higher temperature due to duct losses. This COP improvement, however, can’t compensate 

for the significant increase in the cooling load, which is the cause of the increased energy use.

Table 5.10 shows the effect of duct leakage on annual energy use for the slab-on-grade house from

lowering the cooling set point by 1.1 °C (2.0 °F). For completeness, the table includes all studied 

locations, although houses in Houston and Washington, DC, are most likely to be operated at a lowered 

set point temperature to improve the indoor comfort. Table 5.11 shows simulation results for the indoor 

set point temperature lowered by an additional 1.1 °C (2.0 °F), i.e., by 2.2 °C (4.0 °F) below the default 

value for the house in Houston.

Reducing the set point results in a lower number of hours with relative humidity above 55 % for small 

levels of duct leaks only (Tables 5.10 and 5.11). For large levels of duct leakage, the number of hours 

with relative humidity above 55 % actually increases. This result is caused by the fact that lowering the 

set point requires longer operational runtimes (with correspondingly higher energy consumption and duct 

leakage) and, depending on the ratio of sensible to latent capacities, lowering the indoor temperature may 

actually increase the relative humidity, although the indoor comfort might improve due to a lower drybulb temperature.

90

95

100

105

110

115

120

125

130

Las Vegas Wash., DC Chicago Minneapolis

Relative energy use (%)

-20% Nominal 25% 50% 75 100%
This publication is available free of charge from: http://dx.doi.org/10.6028/NIST.TN.1848

55

Table 5.9. Effect of duct leakage on annual energy use for a slab-on-grade house at default cooling set 

point 

Note: All simulation cases account for thermal losses along with leakage losses except the case denoted ‘0 % & No thermal’.

Houston

 Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC SHR 

(-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

0 % & No thermal 1,715 1,555 588 0.3 2,142.9 4.3 0.789 13,007 6,623 4,339 24,700 $583 79%

0 % Leak 1,537 1,794 685 2.1 2,479.0 4.3 0.812 15,046 7,761 5,020 28,559 $674 91%

10 % Leak 1,512 1,981 749 5.1 2,730.5 4.3 0.785 16,660 8,537 5,529 31,457 $743 100%

20 % Leak 1,632 2,160 815 9.4 2,975.1 4.4 0.767 18,179 9,383 6,025 34,317 $810 109%

30 % Leak 1,922 2,327 883 17.5 3,209.7 4.5 0.753 19,574 10,393 6,500 37,198 $878 118%

40 % Leak 2,738 2,489 953 35.5 3,441.7 4.5 0.743 20,922 11,773 6,970 40,397 $954 128%

50 % Leak 3,364 2,649 1,032 61.8 3,681.0 4.6 0.734 22,231 13,578 7,454 43,995 $1,039 140%

Las Vegas

 Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC SHR 

(-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

0 % & No thermal - 1,536 668 0.3 2,204.5 3.7 1.000 15,941 8,763 5,207 30,642 $1,072 78%

0 % Leak - 1,817 786 0.3 2,602.5 3.7 1.000 18,952 10,273 6,147 36,104 $1,264 92%

10 % Leak - 1,966 865 0.3 2,831.1 3.7 0.999 20,531 11,251 6,687 39,200 $1,372 100%

20 % Leak - 2,114 951 1.2 3,065.4 3.8 0.998 22,081 12,339 7,241 42,393 $1,484 108%

30 % Leak - 2,261 1,054 3.7 3,315.3 3.8 0.998 23,580 13,718 7,831 45,861 $1,605 117%

40 % Leak - 2,405 1,170 8.6 3,575.4 3.9 0.997 25,028 15,353 8,445 49,558 $1,735 126%

50 % Leak - 2,549 1,290 22.7 3,838.7 3.9 0.996 26,444 17,362 9,067 53,605 $1,876 137%

Washington, DC

 Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC SHR 

(-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

0 % & No thermal 280 944 1,532 12.9 2,476.3 4.4 0.801 6,301 15,111 4,179 26,322 $1,031 73%

0 % Leak 175 1,100 1,803 54.5 2,902.7 4.4 0.823 7,361 19,093 4,898 32,084 $1,257 89%

10 % Leak 253 1,207 1,971 89.0 3,178.0 4.5 0.809 8,098 21,759 5,363 35,952 $1,408 100%

20 % Leak 368 1,314 2,133 134.8 3,446.8 4.5 0.799 8,825 24,760 5,817 40,133 $1,572 112%

30 % Leak 523 1,419 2,294 192.5 3,712.5 4.6 0.791 9,528 28,180 6,265 44,704 $1,751 124%

40 % Leak 814 1,523 2,457 270.0 3,979.2 4.6 0.786 10,216 32,335 6,715 49,997 $1,958 139%

50 % Leak 1,165 1,625 2,595 382.3 4,219.9 4.7 0.781 10,884 37,541 7,121 56,278 $2,204 157%

Chicago

 Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC SHR 

(-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

0 % & No thermal 203 815 2,288 70.0 3,103.7 4.5 0.819 5,369 24,753 5,238 36,092 $1,283 71%

0 % Leak 190 943 2,639 187.4 3,582.0 4.5 0.839 6,217 32,197 6,045 45,190 $1,607 88%

10 % Leak 189 1,031 2,833 281.2 3,863.9 4.5 0.827 6,816 37,118 6,520 51,186 $1,820 100%

20 % Leak 192 1,119 3,007 394.4 4,125.5 4.6 0.818 7,410 42,561 6,962 57,664 $2,050 113%

30 % Leak 220 1,208 3,150 532.6 4,358.0 4.6 0.812 8,003 48,636 7,354 64,725 $2,301 126%

40 % Leak 310 1,296 3,285 697.0 4,581.3 4.7 0.806 8,591 55,589 7,731 72,642 $2,583 142%

50 % Leak 427 1,386 3,408 900.9 4,793.8 4.7 0.801 9,174 63,893 8,090 81,888 $2,912 160%

Minneapolis

 Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC SHR 

(-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

0 % & No thermal 15 711 2,902 216.9 3,613.5 4.5 0.838 4,670 36,410 6,098 47,909 $1,437 69%

0 % Leak 13 822 3,258 443.5 4,079.8 4.4 0.856 5,407 47,766 6,885 60,789 $1,824 88%

10 % Leak 13 897 3,432 612.5 4,328.9 4.5 0.846 5,912 55,105 7,305 69,053 $2,072 100%

20 % Leak 15 973 3,577 802.2 4,550.0 4.5 0.839 6,421 62,936 7,678 77,767 $2,333 113%

30 % Leak 27 1,050 3,698 1,009.5 4,748.5 4.6 0.833 6,937 71,179 8,013 86,861 $2,606 126%

40 % Leak 48 1,127 3,816 1,234.7 4,942.6 4.6 0.829 7,444 80,060 8,341 96,576 $2,897 140%

50 % Leak 89 1,207 3,946 1,483.7 5,152.5 4.7 0.825 7,964 89,955 8,695 107,345 $3,220 155%
This publication is available free of charge from: http://dx.doi.org/10.6028/NIST.TN.1848

56

Table 5.10. Effect of duct leakage on annual energy use for a slab-on-grade house at lowered cooling set 

point by 1.1 °C (2.0 °F)

Figures 5.3 and 5.4 present the number of hours above 55 % relative humidity and relative energy use, 

respectively, for a slab-on-grade house in Houston with different duct leak rates at the three studied

thermostat set point temperatures. The energy use is related to that of a house with 10 % leak rate 

(assumed as a representative of no-fault duct installation) at the default thermostat set point (Table 5.9). 

At a leak rate greater than 20 %, the heat pump was unable to lower the number of hours above 55 % 

Houston

 Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC SHR 

(-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

0 % & No thermal 1,186 1,929 610 0.3 2,539.6 4.2 0.801 15,943 6,870 5,143 28,687 $677 79%

0 % Leak 988 2,220 710 2.1 2,930.4 4.2 0.822 18,386 8,042 5,934 33,093 $781 91%

10 % Leak 1,035 2,451 777 5.1 3,227.6 4.3 0.792 20,333 8,844 6,536 36,445 $861 100%

20 % Leak 1,213 2,663 845 9.5 3,508.7 4.4 0.772 22,105 9,724 7,105 39,666 $937 109%

30 % Leak 1,867 2,858 915 18.0 3,773.2 4.5 0.757 23,717 10,759 7,641 42,848 $1,012 118%

40 % Leak 2,851 3,051 989 36.0 4,040.3 4.5 0.746 25,288 12,191 8,182 46,392 $1,095 127%

50 % Leak 3,336 3,237 1,069 63.5 4,306.1 4.6 0.736 26,785 14,046 8,720 50,283 $1,187 138%

Las Vegas

 Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC SHR 

(-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

0 % & No thermal - 1,788 684 0.3 2,472.5 3.7 1.000 18,346 8,965 5,840 33,883 $1,186 78%

0 % Leak - 2,114 805 0.3 2,918.3 3.7 1.000 21,779 10,517 6,893 39,920 $1,397 92%

10 % Leak - 2,280 884 0.3 3,164.2 3.7 0.999 23,494 11,496 7,474 43,196 $1,512 100%

20 % Leak - 2,444 973 1.2 3,416.7 3.8 0.998 25,155 12,625 8,070 46,581 $1,630 108%

30 % Leak - 2,603 1,079 3.7 3,681.6 3.8 0.997 26,742 14,031 8,696 50,201 $1,757 116%

40 % Leak - 2,760 1,198 8.8 3,957.3 3.9 0.996 28,275 15,712 9,347 54,067 $1,892 125%

50 % Leak - 2,917 1,323 22.6 4,239.9 3.9 0.995 29,786 17,787 10,015 58,319 $2,041 135%

Washington, DC

 Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC SHR 

(-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

0 % & No thermal 157 1,171 1,554 13.0 2,725.8 4.4 0.813 7,717 15,317 4,600 28,365 $1,111 74%

0 % Leak 65 1,364 1,831 54.2 3,195.0 4.4 0.835 9,008 19,345 5,392 34,477 $1,350 89%

10 % Leak 158 1,499 2,001 89.0 3,500.4 4.5 0.818 9,918 22,035 5,907 38,592 $1,512 100%

20 % Leak 301 1,632 2,170 134.5 3,802.2 4.5 0.806 10,802 25,092 6,416 43,042 $1,686 112%

30 % Leak 563 1,758 2,331 192.6 4,089.5 4.6 0.797 11,632 28,528 6,901 47,793 $1,872 124%

40 % Leak 1,015 1,883 2,500 270.1 4,383.0 4.6 0.791 12,442 32,734 7,396 53,304 $2,088 138%

50 % Leak 1,311 2,008 2,647 382.6 4,654.5 4.7 0.785 13,246 38,021 7,854 59,853 $2,344 155%

Chicago

 Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC SHR 

(-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

0 % & No thermal 182 1,002 2,303 69.9 3,304.9 4.4 0.828 6,521 24,884 5,577 37,714 $1,341 71%

0 % Leak 173 1,159 2,657 187.4 3,815.3 4.4 0.847 7,554 32,354 6,438 47,078 $1,674 88%

10 % Leak 176 1,267 2,849 281.2 4,115.8 4.5 0.833 8,277 37,266 6,945 53,220 $1,892 100%

20 % Leak 175 1,375 3,024 394.4 4,398.2 4.5 0.823 8,994 42,715 7,422 59,863 $2,128 112%

30 % Leak 246 1,483 3,169 533.5 4,651.8 4.6 0.815 9,705 48,830 7,850 67,117 $2,386 126%

40 % Leak 365 1,591 3,311 697.0 4,901.8 4.6 0.809 10,407 55,823 8,272 75,233 $2,675 141%

50 % Leak 498 1,699 3,438 901.3 5,136.4 4.7 0.803 11,098 64,171 8,668 84,668 $3,010 159%

Minneapolis

 Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC SHR 

(-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

0 % & No thermal 8 884 2,919 216.9 3,802.3 4.4 0.845 5,730 36,561 6,416 49,439 $1,483 70%

0 % Leak 4 1,021 3,276 443.2 4,297.0 4.4 0.864 6,627 47,932 7,251 62,541 $1,876 88%

10 % Leak 6 1,114 3,449 612.5 4,563.3 4.5 0.852 7,251 55,263 7,701 70,946 $2,128 100%

20 % Leak 7 1,209 3,598 802.2 4,807.8 4.5 0.843 7,883 63,128 8,113 79,855 $2,396 113%

30 % Leak 8 1,304 3,724 1,009.6 5,028.1 4.6 0.836 8,501 71,417 8,485 89,134 $2,674 126%

40 % Leak 48 1,399 3,845 1,234.7 5,244.2 4.6 0.831 9,116 80,330 8,850 99,027 $2,971 140%

50 % Leak 129 1,497 3,979 1,484.0 5,476.5 4.7 0.826 9,745 90,267 9,242 109,985 $3,300 155%
This publication is available free of charge from: http://dx.doi.org/10.6028/NIST.TN.1848

57

relative humidity although the amount of moisture in the air was lowered and a lower indoor air 

temperature improved indoor thermal comfort to some degree. For the house with a 40 % duct leakage, 

the energy use is predicted to be 47 % and 97 % higher than for the reference house if the set point 

temperature is lowered by 1.1 °C and 2.2 °C, respectively (Figure 5.4). 

The results contained in Table 5.12 (derived from Tables 5.10 and 5.11) present a change in the annual 

energy use for the baseline houses due to lowering the cooling set point. For Las Vegas, Washington, 

Chicago, and Minneapolis, the change in energy use is the same for the slab-on-grade house and the 

house with a basement. The use of energy increased by the same percentage for a slab-on-grade house 

and a house with a basement located in the same climate. As expected, the effect of lowering the set point 

temperature was small on the total energy use in houses located in heating dominated climates. 

Table 5.11. Effect of duct leakage on annual energy use for a slab-on-grade house in Houston at lowered 

cooling set point by 2.2 °C (4.0 °F) 

Figure 5.3. Number of hours above 55 % relative humidity for a slab-on-grade house in Houston with 

duct leak rates from 10 % to 50 % at three thermostat set point temperatures

Houston

 Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC SHR (-

) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

0 % & No thermal 536 2,759 995 0.3 3,511.5 4.3 0.841 22,498 11,221 7,602 42,053 $993 82%

0 % Leak 389 3,106 1,116 2.3 3,948.8 4.3 0.855 25,383 12,592 8,550 47,256 $1,116 93%

10 % Leak 435 3,376 1,193 5.3 4,276.4 4.4 0.825 27,602 13,468 9,253 51,054 $1,205 100%

20 % Leak 737 3,627 1,280 9.7 4,601.1 4.4 0.803 29,645 14,516 9,936 54,829 $1,295 107%

30 % Leak 1,889 3,842 1,352 18.0 4,895.4 4.5 0.786 31,384 15,547 10,519 58,182 $1,374 114%

40 % Leak 2,685 4,054 1,438 35.3 5,194.0 4.5 0.772 33,084 17,064 11,122 62,002 $1,464 121%

50 % Leak 3,197 4,250 1,517 61.9 5,475.9 4.6 0.760 34,637 18,858 11,677 65,904 $1,556 129%

0

500

1000

1500

2000

2500

3000

3500

Default 1.1 °C 2.2 °C

Hours above 55 % RH

50%

40%

30%

20%

10%
This publication is available free of charge from: http://dx.doi.org/10.6028/NIST.TN.1848

58

Figure 5.4. . Energy use for a slab-on-grade house in Houston with duct leak rates from 10 % to 50 % at 

three thermostat set point temperatures related to energy use for the house at the default set point and 

10 % leak rate (shown in Table 5.9)

 

80

100

120

140

160

180

200

220

Default 1.1 °C 2.2 °C

Relative energy use (%)

50%

40%

30%

20%

10%
This publication is available free of charge from: http://dx.doi.org/10.6028/NIST.TN.1848

59

Table 5.12. Effect of lowering cooling set point by 1.1 °C (2.0 °F) on annual energy use of a baseline slab-on-grade house and a house with 

basement

 Hours 

Above 

55 % 

RH 

 Space 

Temp 

Max (C) 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC 

SHR (-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU 

Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

Cooling 

Energy

Normal Set Pt 1,512 26.6 1,981 749 5.1 2,730.5 4.3 0.785 16,660 8,537 5,529 31,457 $743 100% 100%

Lower Set Pt 1,035 26.1 2,451 777 5.1 3,227.6 4.3 0.792 20,333 8,844 6,536 36,445 $861 116% 122%

Normal Set Pt - 27.0 1,966 865 0.3 2,831.1 3.7 0.999 20,531 11,251 6,687 39,200 $1,372 100% 100%

Lower Set Pt - 26.6 2,280 884 0.3 3,164.2 3.7 0.999 23,494 11,496 7,474 43,196 $1,512 110% 114%

Normal Set Pt 253 25.1 1,207 1,971 89.0 3,178.0 4.5 0.809 8,098 21,759 5,363 35,952 $1,408 100% 100%

Lower Set Pt 158 24.4 1,499 2,001 89.0 3,500.4 4.5 0.818 9,918 22,035 5,907 38,592 $1,512 107% 122%

Normal Set Pt 189 25.0 1,031 2,833 281.2 3,863.9 4.5 0.827 6,816 37,118 6,520 51,186 $1,820 100% 100%

lower Set Pt 176 24.4 1,267 2,849 281.2 4,115.8 4.5 0.833 8,277 37,266 6,945 53,220 $1,892 104% 121%

Normal Set Pt 13 25.2 897 3,432 612.5 4,328.9 4.5 0.846 5,912 55,105 7,305 69,053 $2,072 100% 100%

lower Set Pt 6 24.5 1,114 3,449 612.5 4,563.3 4.5 0.852 7,251 55,263 7,701 70,946 $2,128 103% 123%

Houston

Las Vegas

Washington, DC

Chicago

Minneapolis

 Hours 

Above 

55 % 

RH 

 Space 

Temp 

Max (C) 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC 

SHR (-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU 

Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

Cooling 

Energy

Normal Set Pt - 25.3 1,552 718 0.3 2,269.9 3.7 1.000 16,107 9,407 5,362 31,607 $1,106 100% 100%

Lower Set Pt - 24.2 1,809 731 0.3 2,539.6 3.7 1.000 18,571 9,573 5,999 34,874 $1,221 110% 115%

Normal Set Pt 654 24.0 742 1,907 27.0 2,649.5 4.4 0.775 5,008 19,120 4,471 29,330 $1,149 100% 100%

Lower Set Pt 306 23.0 976 1,925 27.0 2,900.5 4.4 0.793 6,484 19,284 4,895 31,394 $1,230 107% 129%

Normal Set Pt 289 24.0 631 2,785 129.8 3,416.1 4.4 0.797 4,198 31,565 5,765 42,259 $1,503 100% 100%

lower Set Pt 160 23.0 826 2,796 129.7 3,621.3 4.4 0.810 5,419 31,661 6,111 43,923 $1,562 104% 129%

Normal Set Pt 61 24.0 515 3,424 354.2 3,939.8 4.4 0.810 3,428 46,239 6,648 57,048 $1,711 100% 100%

lower Set Pt 43 23.0 689 3,434 354.2 4,123.5 4.4 0.824 4,520 46,328 6,958 58,538 $1,756 103% 132%

Las Vegas

Washington, DC

Chicago

Minneapolis

Slab-on-grade house

House with basement
This publication is available free of charge from: http://dx.doi.org/10.6028/NIST.TN.1848

60

5.2.4 Effect of Indoor Coil Airflow

This fault covers the case where a heat pump properly sized for the building load operates with improperly 

sized ductwork. As a result, the indoor coil airflow is not nominal. The effect of improper airflow in the 

cooling mode was determined using the baseline performance maps for the air conditioner used in a past 

study because they were shown to be very close to the correlations derived from NIST lab testing (Section 

3.2.1). The impact of indoor airflow on heat pump performance in the heating mode was not considered in 

the heat pump baseline performance maps, therefore, the NIST correlations were used to determine this 

impact. The simulated indoor airflows, ranging from -36 % to +28 % of the nominal flow corresponded to

external static pressures of (177, 171, 168, 165, and 149) Pa ((0.71, 0.69, 0.67, 0.66, 0.60) inch H2O), 

respectively. 

Reduced airflow results in an increase in energy consumption, and this effect is similar for all houses in all 

climates studied (Tables 5.13 and 5.14). Figure 5.5, generated for slab-on-grade houses, also provides a 

good representation of simulation results for houses with a basement. For the lowest airflow, 36 % below 

the nominal value, the energy use increased from 11 % to 14 %. 

Figure 5.5. Annual energy use for slab-on-grade houses for different indoor coil airflows relative to energy 

use for the house in the same location with nominal airflow rate

In the cooling mode, reducing the airflow below the nominal value of 181.1 m3

∙h-1∙kW-1 (375 cfm/ton) 

causes a decrease in the indoor coil temperature and provides better humidity control, but results in higher 

energy use because the sensible capacity is reduced and running time increased. Conversely, providing 

more airflow hurts humidity control in the house but decreases energy use. The efficiency of the system 

goes up, and more importantly, the latent removal decreases so energy use decreases. To account for a 

possible scenario where the homeowner lowers the temperature setting on the thermostat in an effort to 

make the indoor environment more comfortable, Tables 5.15 and 5.16 provide simulation results for both 

houses for cases where the thermostat set point is reduced 1.1 °C (2.0 °F) below the ‘default’ values shown 

in Table 4.8. 

Tables 5.13 and 5.15 show the energy usage penalties associated with lowering the airflow and reducing 

the thermostat set point to aid in humidity control. In Table 5.13 for Houston, a hot and humid climate, the 

slab-on-grade house spends 1183 hours above 55 % RH even with the airflow reduced by 36 %, resulting 

in a 12 % increase in annual energy usage (The total energy draw was 35334 MJ). Keeping the airflow at 

the nominal value but lowering the thermostat set point by 1.1 °C (2.0 °F), as shown in table 5.15, reduces 

the number of hours above 55 % RH to a comparable number of hours of 1035 while increasing the energy 

90

95

100

105

110

115

120

125

130

Houston Las Vegas Wash., DC Chicago Minneapolis

Relative energy use (%)

-36% -15% Nominal 7% 28%
This publication is available free of charge from: http://dx.doi.org/10.6028/NIST.TN.1848

61

Table 5.13. Effect of indoor coil airflow on annual energy use for a slab-on-grade house when operating 

at the default cooling set point

use by 16 % (36445 MJ compared to 31457 MJ). Reduced airflow or lowered cooling set point in other 

climates - in which the number of hours above 55 % was small - resulted in significant energy use 

penalties and a small reduction of high RH hours.

Houston

 Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC SHR 

(-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

-36 % flow 1,183 2,272 853 9.7 3,125.5 3.9 0.770 18,783 10,982 4,838 35,334 $834 112%

-15 % flow 1,364 2,074 785 6.6 2,858.7 4.2 0.780 17,332 9,405 5,331 32,800 $774 104%

nominal flow 1,512 1,981 749 5.1 2,730.5 4.3 0.785 16,660 8,537 5,529 31,457 $743 100%

7 % flow 1,617 1,951 743 4.9 2,693.9 4.4 0.787 16,455 8,465 5,609 31,262 $738 99%

28 % flow 2,026 1,878 726 4.7 2,603.3 4.5 0.793 16,080 8,259 5,727 30,798 $727 98%

 

 

 Las Vegas

 Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC SHR 

(-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

-36 % flow - 2,268 1,000 1.3 3,268.2 3.3 0.992 23,192 14,547 5,902 44,373 $1,553 113%

-15 % flow - 2,057 910 0.6 2,966.6 3.6 0.998 21,369 12,396 6,454 40,951 $1,433 104%

nominal flow - 1,966 865 0.3 2,831.1 3.7 0.999 20,531 11,251 6,687 39,200 $1,372 100%

7 % flow - 1,933 856 0.4 2,789.0 3.8 1.000 20,232 11,126 6,775 38,865 $1,360 99%

28 % flow - 1,866 837 0.3 2,702.4 3.9 1.000 19,667 10,875 6,936 38,211 $1,337 97%

 

 

 Washington, DC

 Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC SHR 

(-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

-36 % flow 26 1,409 2,175 119.1 3,583.9 4.0 0.786 9,295 26,391 4,623 41,041 $1,607 114%

-15 % flow 153 1,271 2,042 98.0 3,312.6 4.3 0.801 8,476 23,334 5,148 37,689 $1,476 105%

nominal flow 253 1,207 1,971 89.0 3,178.0 4.5 0.809 8,098 21,759 5,363 35,952 $1,408 100%

7 % flow 305 1,184 1,959 87.0 3,143.5 4.5 0.812 7,974 21,618 5,455 35,778 $1,401 100%

28 % flow 520 1,132 1,931 83.7 3,063.8 4.6 0.821 7,738 21,376 5,617 35,463 $1,389 99%

 

 

 Chicago

 Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC SHR 

(-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

-36 % flow 160 1,217 3,046 350.6 4,263.5 4.1 0.798 7,920 43,548 5,500 57,699 $2,052 113%

-15 % flow 183 1,089 2,909 301.3 3,997.4 4.4 0.816 7,159 39,221 6,212 53,323 $1,896 104%

nominal flow 189 1,031 2,833 281.2 3,863.9 4.5 0.827 6,816 37,118 6,520 51,186 $1,820 100%

7 % flow 190 1,009 2,817 277.3 3,826.5 4.6 0.831 6,695 36,913 6,640 50,980 $1,813 100%

28 % flow 216 960 2,781 270.0 3,740.4 4.6 0.844 6,462 36,577 6,858 50,628 $1,800 99%

 

 

Minneapolis Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC SHR 

(-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

-36 % flow 3 1,066 3,613 727.2 4,679.7 4.1 0.813 6,917 62,805 6,037 76,491 $2,295 111%

-15 % flow 9 950 3,496 646.3 4,446.1 4.4 0.834 6,225 57,542 6,909 71,408 $2,142 103%

nominal flow 13 897 3,432 612.5 4,328.9 4.5 0.846 5,912 55,105 7,305 69,053 $2,072 100%

7 % flow 15 878 3,418 606.3 4,295.6 4.6 0.851 5,803 54,874 7,454 68,863 $2,066 100%

28 % flow 27 832 3,387 592.6 4,218.8 4.6 0.866 5,581 54,499 7,735 68,546 $2,056 99%
This publication is available free of charge from: http://dx.doi.org/10.6028/NIST.TN.1848

62

Table 5.14. Effect of indoor coil airflow on annual energy use for a house with basement when operating 

at the default cooling set point

Las Vegas

 Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC SHR 

(-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

-36 % flow - 1,765 805 0.3 2,569.9 3.3 0.994 18,011 11,795 4,641 35,178 $1,231 111%

-15 % flow - 1,616 750 0.3 2,366.2 3.5 1.000 16,707 10,302 5,148 32,889 $1,151 104%

nominal flow - 1,552 718 0.3 2,269.9 3.7 1.000 16,107 9,407 5,362 31,607 $1,106 100%

7 % flow - 1,529 713 0.3 2,241.9 3.7 1.000 15,890 9,341 5,446 31,409 $1,099 99%

28 % flow - 1,478 698 0.3 2,175.8 3.8 1.000 15,443 9,145 5,585 30,905 $1,082 98%

 

 

 

Chicago

 Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC SHR 

(-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

-36 % flow 218 729 2,992 170.2 3,721.0 4.0 0.761 4,784 37,000 4,800 47,316 $1,682 112%

-15 % flow 250 663 2,862 142.6 3,524.9 4.3 0.784 4,388 33,460 5,478 44,058 $1,567 104%

nominal flow 289 631 2,785 129.8 3,416.1 4.4 0.797 4,198 31,565 5,765 42,259 $1,503 100%

7 % flow 299 620 2,768 126.9 3,387.0 4.5 0.802 4,131 31,369 5,877 42,108 $1,497 100%

28 % flow 377 590 2,727 118.4 3,317.3 4.6 0.818 3,992 30,921 6,082 41,726 $1,484 99%

 

 

 

Washington, DC

 Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC SHR 

(-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

-36 % flow 245 851 2,091 39.7 2,941.7 4.0 0.746 5,661 22,995 3,795 33,183 $1,300 113%

-15 % flow 463 777 1,974 30.8 2,751.8 4.3 0.765 5,217 20,502 4,276 30,727 $1,203 105%

nominal flow 653 742 1,907 27.0 2,649.5 4.4 0.775 5,007 19,120 4,471 29,330 $1,149 100%

7 % flow 743 729 1,895 25.7 2,624.5 4.4 0.779 4,933 18,999 4,554 29,218 $1,144 100%

28 % flow 1,030 699 1,861 22.7 2,560.1 4.5 0.791 4,790 18,680 4,694 28,895 $1,132 99%

 

 

 

Minneapolis

 Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC SHR 

(-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

-36 % flow 49 597 3,615 428.4 4,211.9 4.0 0.770 3,921 52,744 5,433 62,829 $1,885 110%

-15 % flow 55 542 3,497 377.7 4,039.3 4.3 0.796 3,589 48,444 6,277 59,042 $1,771 103%

nominal flow 61 515 3,424 354.2 3,939.8 4.4 0.810 3,428 46,239 6,648 57,048 $1,711 100%

7 % flow 68 506 3,410 348.1 3,915.8 4.5 0.816 3,372 46,004 6,795 56,902 $1,707 100%

28 % flow 78 482 3,373 330.3 3,854.6 4.5 0.834 3,255 45,391 7,067 56,445 $1,693 99%
This publication is available free of charge from: http://dx.doi.org/10.6028/NIST.TN.1848

63

Table 5.15. Effect of indoor coil airflow on annual energy use for a slab-on-grade house when operating at

a cooling set point that is 1.1 °C (2.0 °F) lower than the default value

Note: Although the relative energy use shown in this table is equal or less than the values shown in Table 5.13 

(baseline), the total energy use for cases presented in Table 5.15 is higher than those presented in Table 5.13. 

Houston

 Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC SHR 

(-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

-36 % flow 572 2,789 884 9.8 3,672.5 3.9 0.779 22,733 11,373 5,685 40,522 $957 111%

-15 % flow 846 2,556 813 6.6 3,369.4 4.2 0.788 21,082 9,739 6,283 37,836 $893 104%

nominal flow 1,035 2,451 777 5.1 3,227.6 4.3 0.792 20,333 8,844 6,536 36,445 $861 100%

7 % flow 1,139 2,413 770 4.9 3,183.6 4.4 0.794 20,083 8,766 6,629 36,209 $855 99%

28 % flow 1,628 2,326 752 4.7 3,078.5 4.5 0.799 19,631 8,556 6,773 35,692 $843 98%

Las Vegas

 Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC SHR 

(-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

-36 % flow - 2,613 1,022 1.3 3,635.2 3.4 0.991 26,339 14,874 6,565 48,509 $1,698 112%

-15 % flow - 2,382 931 0.6 3,312.5 3.6 0.998 24,409 12,684 7,207 45,031 $1,576 104%

nominal flow - 2,280 884 0.3 3,164.2 3.7 0.999 23,494 11,496 7,474 43,196 $1,512 100%

7 % flow - 2,242 874 0.4 3,116.5 3.8 0.999 23,156 11,360 7,571 42,818 $1,499 99%

28 % flow - 2,166 855 0.3 3,021.1 3.9 1.000 22,516 11,112 7,754 42,114 $1,474 97%

Washington, DC

 Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC SHR 

(-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

-36 % flow 4 1,742 2,212 119.1 3,953.8 4.0 0.798 11,319 26,773 5,100 43,923 $1,720 114%

-15 % flow 58 1,576 2,075 97.8 3,650.9 4.3 0.811 10,365 23,644 5,674 40,415 $1,583 105%

nominal flow 158 1,499 2,001 89.0 3,500.4 4.5 0.818 9,918 22,035 5,907 38,592 $1,512 100%

7 % flow 203 1,473 1,989 87.4 3,461.8 4.5 0.820 9,777 21,902 6,007 38,418 $1,505 100%

28 % flow 461 1,410 1,960 83.6 3,369.8 4.6 0.828 9,487 21,640 6,178 38,036 $1,490 99%

Chicago

 Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC SHR 

(-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

-36 % flow 131 1,490 3,067 350.6 4,556.5 4.1 0.808 9,571 43,759 5,878 59,939 $2,131 113%

-15 % flow 160 1,336 2,927 301.3 4,263.4 4.3 0.824 8,683 39,398 6,625 55,439 $1,971 104%

nominal flow 176 1,267 2,849 281.2 4,115.8 4.5 0.833 8,277 37,266 6,945 53,220 $1,892 100%

7 % flow 176 1,240 2,833 277.4 4,073.6 4.5 0.837 8,131 37,061 7,068 52,992 $1,884 100%

28 % flow 199 1,183 2,799 270.0 3,981.5 4.6 0.848 7,859 36,740 7,300 52,631 $1,871 99%

Minneapolis

 Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC SHR 

(-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

-36 % flow - 1,318 3,636 727.0 4,954.4 4.0 0.822 8,438 63,034 6,391 78,595 $2,358 111%

-15 % flow 1 1,178 3,515 646.2 4,693.4 4.3 0.841 7,628 57,722 7,294 73,375 $2,201 103%

nominal flow 6 1,114 3,449 612.5 4,563.3 4.5 0.852 7,251 55,263 7,701 70,946 $2,128 100%

7 % flow 6 1,091 3,436 606.3 4,526.0 4.5 0.856 7,118 55,039 7,854 70,742 $2,122 100%

28 % flow 13 1,036 3,404 592.6 4,440.5 4.6 0.869 6,853 54,661 8,141 70,387 $2,112 99%
This publication is available free of charge from: http://dx.doi.org/10.6028/NIST.TN.1848

64

Table 5.16. Effect of indoor coil airflow on annual energy use for a house with basement when operating 

at cooling set point that is 1.1 °C (2.0 °F) lower than the default value

Note: Although the relative energy use shown in this table is equal or less than the values shown in Table 5.14

(baseline), the total energy use for cases presented in Table 5.16 is higher than those presented in Table 5.14. 

5.2.5 Effect of Refrigerant Undercharge

When the amount of refrigerant charge in the TXV-controlled system is below the nominal value, the 

performance of the unit is degraded. Tables 5.17 and 5.18 show the results for the slab-on-grade house 

and the basement house, respectively. Figure 5.6 shows the relative energy use for the slab-on-grade 

house, which provides a good representation of the energy use in the house with a basement as well. The 

figure indicates that the energy use increases exponentially with increasing refrigerant undercharge. For 

the 30 % refrigerant undercharge level the energy use increases by as much as (17 ~ 23) %. The moisture 

removal capacity of the unit is also degraded when the unit is undercharged. 

Las Vegas

 Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC SHR 

(-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

-36 % flow - 1,765 805 0.3 2,569.9 3.3 0.994 18,011 11,795 4,641 35,178 $1,231 111%

-15 % flow - 1,616 750 0.3 2,366.2 3.5 1.000 16,707 10,302 5,148 32,889 $1,151 104%

nominal flow - 1,552 718 0.3 2,269.9 3.7 1.000 16,107 9,407 5,362 31,607 $1,106 100%

7 % flow - 1,529 713 0.3 2,241.9 3.7 1.000 15,890 9,341 5,446 31,409 $1,099 99%

28 % flow - 1,478 698 0.3 2,175.8 3.8 1.000 15,443 9,145 5,585 30,905 $1,082 98%

 

 

 Washington, DC

 Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC SHR 

(-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

-36 % flow 67 1,118 2,111 39.9 3,229.1 3.9 0.767 7,332 23,207 4,165 35,435 $1,388 113%

-15 % flow 184 1,021 1,993 30.8 3,013.7 4.2 0.784 6,755 20,680 4,683 32,850 $1,287 105%

nominal flow 306 976 1,925 27.0 2,900.5 4.4 0.793 6,484 19,284 4,895 31,394 $1,230 100%

7 % flow 378 959 1,913 25.7 2,871.9 4.4 0.797 6,387 19,164 4,983 31,266 $1,225 100%

28 % flow 666 918 1,878 22.7 2,796.0 4.5 0.808 6,191 18,831 5,126 30,880 $1,209 98%

 

 

 Chicago Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC SHR 

(-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

-36 % flow 111 952 3,004 170.2 3,956.5 4.0 0.777 6,166 37,130 5,104 49,132 $1,747 112%

-15 % flow 142 867 2,872 142.6 3,738.2 4.3 0.798 5,663 33,556 5,809 45,759 $1,627 104%

nominal flow 160 826 2,796 129.7 3,621.3 4.4 0.810 5,419 31,661 6,111 43,923 $1,562 100%

7 % flow 165 811 2,777 127.2 3,588.5 4.5 0.815 5,337 31,467 6,227 43,762 $1,556 100%

28 % flow 193 774 2,738 118.5 3,511.6 4.5 0.829 5,153 31,024 6,438 43,346 $1,541 99%

 

 

Minneapolis Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC SHR 

(-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

-36 % flow 17 799 3,627 428.5 4,425.4 4.0 0.788 5,165 52,870 5,709 64,475 $1,934 110%

-15 % flow 31 726 3,508 377.7 4,233.4 4.2 0.811 4,735 48,549 6,579 60,595 $1,818 104%

nominal flow 43 689 3,434 354.2 4,123.5 4.4 0.824 4,520 46,328 6,958 58,538 $1,756 100%

7 % flow 46 677 3,421 347.9 4,097.9 4.4 0.829 4,448 46,096 7,111 58,386 $1,752 100%

28 % flow 54 644 3,383 330.3 4,027.2 4.5 0.845 4,285 45,481 7,383 57,881 $1,736 99%
This publication is available free of charge from: http://dx.doi.org/10.6028/NIST.TN.1848

65

Table 5.17. Effect of refrigerant undercharge on annual energy use for a slab-on-grade house

Table 5.18. Effect of refrigerant undercharge on annual energy use for a house with basement

Figure 5.6. Annual energy use for slab-on-grade houses at different levels of refrigerant undercharge 

relative to the annual energy use for the house in the same location when the heat pump operates with the 

nominal refrigerant charge

Under Charge Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC 

SHR (-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

0 % 1,512 1,981 749 5.1 2,730.5 4.3 0.785 16,660 8,537 5,529 31,457 $743 100%

-10 % 1,581 2,052 778 5.8 2,830.4 4.2 0.787 17,098 8,787 5,731 32,348 $764 103%

-20 % 1,676 2,176 855 8.5 3,031.2 4.0 0.789 17,901 9,562 6,138 34,333 $811 109%

-30 % 1,811 2,366 1,000 20.2 3,366.3 3.8 0.792 19,131 11,284 6,817 37,963 $896 121%

0 % - 1,966 865 0.3 2,831.1 3.7 0.999 20,531 11,251 6,687 39,200 $1,372 100%

-10 % - 2,044 900 0.4 2,944.4 3.6 1.000 21,109 11,573 6,955 40,369 $1,413 103%

-20 % - 2,177 1,000 1.0 3,176.8 3.5 1.000 22,133 12,652 7,504 43,021 $1,506 110%

-30 % - 2,379 1,199 3.7 3,578.0 3.2 1.000 23,671 14,919 8,451 47,773 $1,672 122%

0 % 253 1,207 1,971 89.0 3,178.0 4.5 0.809 8,098 21,759 5,363 35,952 $1,408 100%

-10 % 281 1,246 2,020 91.5 3,266.2 4.4 0.811 8,304 22,133 5,512 36,680 $1,437 102%

-20 % 312 1,317 2,168 109.9 3,485.6 4.2 0.815 8,690 23,868 5,882 39,172 $1,534 109%

-30 % 382 1,433 2,450 154.3 3,882.8 3.9 0.819 9,319 27,533 6,552 44,135 $1,729 123%

0 % 189 1,031 2,833 281.2 3,863.9 4.5 0.827 6,816 37,118 6,520 51,186 $1,820 100%

-10 % 189 1,063 2,886 286.2 3,948.6 4.4 0.830 6,984 37,586 6,663 51,964 $1,848 102%

-20 % 193 1,123 3,035 327.6 4,158.4 4.2 0.834 7,311 40,065 7,017 55,125 $1,960 108%

-30 % 188 1,221 3,281 433.9 4,502.1 3.9 0.841 7,842 45,504 7,597 61,674 $2,193 120%

0 % 13 897 3,432 612.5 4,328.9 4.5 0.846 5,912 55,105 7,305 69,053 $2,072 100%

-10 % 15 925 3,475 621.3 4,400.0 4.4 0.850 6,056 55,608 7,425 69,821 $2,095 101%

-20 % 15 977 3,604 687.0 4,581.0 4.2 0.855 6,342 58,734 7,730 73,538 $2,206 106%

-30 % 15 1,062 3,804 839.1 4,866.2 3.9 0.862 6,802 65,356 8,212 81,101 $2,433 117%

Houston

Las Vegas

Washington, DC

Chicago

Minneapolis

Under Charge Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC 

SHR (-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

0 % - 1,552 718 0.3 2,269.9 3.7 1.000 16,107 9,407 5,362 31,607 $1,106 100%

-10 % - 1,606 745 0.3 2,350.5 3.6 1.000 16,512 9,651 5,552 32,448 $1,136 103%

-20 % - 1,703 817 0.3 2,520.0 3.4 1.000 17,302 10,405 5,952 34,391 $1,204 109%

-30 % - 1,861 966 0.3 2,827.1 3.2 1.000 18,620 12,013 6,678 38,042 $1,331 120%

0 % 654 742 1,907 27.0 2,649.5 4.4 0.775 5,008 19,120 4,471 29,330 $1,149 100%

-10 % 694 762 1,954 27.5 2,715.9 4.3 0.777 5,111 19,412 4,583 29,838 $1,169 102%

-20 % 755 800 2,099 35.1 2,898.9 4.1 0.780 5,321 20,749 4,892 31,693 $1,241 108%

-30 % 851 863 2,384 55.2 3,246.7 3.8 0.785 5,673 23,624 5,479 35,508 $1,391 121%

0 % 289 631 2,785 129.8 3,416.1 4.4 0.797 4,198 31,565 5,765 42,259 $1,503 100%

-10 % 294 647 2,838 131.2 3,485.1 4.3 0.800 4,283 31,904 5,881 42,800 $1,522 101%

-20 % 295 679 2,995 156.3 3,673.9 4.2 0.804 4,453 33,893 6,200 45,277 $1,610 107%

-30 % 304 732 3,282 221.2 4,014.3 3.9 0.810 4,749 38,240 6,774 50,495 $1,795 119%

0 % 61 515 3,424 354.2 3,939.8 4.4 0.810 3,428 46,239 6,648 57,048 $1,711 100%

-10 % 65 529 3,472 357.3 4,001.0 4.3 0.813 3,497 46,578 6,752 57,559 $1,727 101%

-20 % 68 554 3,616 403.2 4,169.8 4.1 0.818 3,636 49,132 7,036 60,536 $1,816 106%

-30 % 69 597 3,861 518.1 4,457.6 3.9 0.825 3,871 54,824 7,522 66,950 $2,008 117%

Las Vegas

Washington, DC

Chicago

Minneapolis

90

95

100

105

110

115

120

125

130

Houston Las Vegas Wash., DC Chicago Minneapolis

Relative energy use (%)

Nominal -10% -20% -30%
This publication is available free of charge from: http://dx.doi.org/10.6028/NIST.TN.1848

66

5.2.6 Effect of Refrigerant Overcharge

When the amount of refrigerant charge in the system is above the correct (nominal) value, the performance 

of the unit is degraded. Table 5.19 and 5.20 show the results for the slab-on-grade house and for the 

basement house, respectively. The heat pump uses (10 ~ 16) % more energy when overcharged by 30 %, 

with somewhat higher increases in energy use occurring in localities with a significant heating season (i.e., 

Chicago, Washington, DC, and Minneapolis). Figure 5.7 shows the relative energy use for the slab-ongrade house, which provides a good representation of the energy use in the house with a basement as well. 

The moisture removal capability of the unit is not affected by the overcharge fault. 

Table 5.19. Effect of refrigerant overcharge on annual energy use for a slab-on-grade house

Table 5.20. Effect of refrigerant overcharge on annual energy use for a house with basement

Over Charge Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC 

SHR (-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

0 % 1,512 1,981 749 5.1 2,730.5 4.3 0.785 16,660 8,537 5,529 31,457 $743 100%

10 % 1,553 1,955 764 5.8 2,718.5 4.3 0.786 16,743 8,912 5,505 31,891 $753 101%

20 % 1,572 1,937 778 6.5 2,714.9 4.2 0.787 17,006 9,616 5,498 32,851 $776 104%

30 % 1,547 1,932 796 7.3 2,728.4 4.1 0.786 17,486 10,736 5,525 34,478 $814 110%

0 % - 1,966 865 0.3 2,831.1 3.7 0.999 20,531 11,251 6,687 39,200 $1,372 100%

10 % - 1,945 884 0.4 2,828.3 3.7 0.999 20,689 11,742 6,681 39,843 $1,394 102%

20 % - 1,929 904 0.6 2,833.4 3.6 0.999 21,042 12,711 6,693 41,178 $1,441 105%

30 % - 1,919 925 0.7 2,843.8 3.5 0.999 21,577 14,180 6,717 43,206 $1,512 110%

0 % 253 1,207 1,971 89.0 3,178.0 4.5 0.809 8,098 21,759 5,363 35,952 $1,408 100%

10 % 277 1,191 2,004 94.3 3,194.7 4.4 0.810 8,144 22,476 5,391 36,744 $1,439 102%

20 % 281 1,183 2,037 100.6 3,220.6 4.3 0.811 8,296 23,977 5,435 38,439 $1,506 107%

30 % 264 1,181 2,074 106.4 3,255.3 4.2 0.809 8,544 26,260 5,493 41,029 $1,607 114%

0 % 189 1,031 2,833 281.2 3,863.9 4.5 0.827 6,816 37,118 6,520 51,186 $1,820 100%

10 % 190 1,016 2,871 293.4 3,887.7 4.5 0.828 6,849 38,136 6,561 52,277 $1,859 102%

20 % 191 1,009 2,904 308.1 3,913.7 4.4 0.829 6,976 40,241 6,604 54,552 $1,940 107%

30 % 189 1,008 2,945 321.4 3,953.0 4.3 0.827 7,188 43,466 6,671 58,056 $2,064 113%

0 % 13 897 3,432 612.5 4,328.9 4.5 0.846 5,912 55,105 7,305 69,053 $2,072 100%

10 % 15 885 3,465 633.6 4,350.4 4.5 0.848 5,945 56,322 7,341 70,339 $2,110 102%

20 % 15 879 3,494 656.0 4,372.9 4.4 0.848 6,053 58,820 7,379 72,984 $2,190 106%

30 % 13 878 3,528 678.6 4,405.8 4.3 0.847 6,236 62,694 7,435 77,096 $2,313 112%

Houston

Las Vegas

Washington, DC

Chicago

Minneapolis

Over Charge Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC 

SHR (-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

0 % - 1,552 718 0.3 2,269.9 3.7 1.000 16,107 9,407 5,362 31,607 $1,106 100%

10 % - 1,538 733 0.3 2,270.4 3.6 1.000 16,261 9,814 5,363 32,170 $1,126 102%

20 % - 1,527 747 0.3 2,274.2 3.6 1.000 16,565 10,587 5,372 33,256 $1,164 105%

30 % - 1,521 763 0.3 2,283.8 3.5 1.000 17,021 11,782 5,394 34,929 $1,223 111%

0 % 654 742 1,907 27.0 2,649.5 4.4 0.775 5,008 19,120 4,471 29,330 $1,149 100%

10 % 695 734 1,940 28.9 2,674.8 4.3 0.776 5,050 19,735 4,514 30,031 $1,176 102%

20 % 695 730 1,972 32.5 2,702.7 4.3 0.776 5,154 21,110 4,561 31,557 $1,236 108%

30 % 658 730 2,009 35.0 2,739.6 4.1 0.775 5,317 23,264 4,623 33,936 $1,329 116%

0 % 289 631 2,785 129.8 3,416.1 4.4 0.797 4,198 31,565 5,765 42,259 $1,503 100%

10 % 295 624 2,823 138.0 3,447.7 4.4 0.799 4,234 32,451 5,818 43,234 $1,537 102%

20 % 294 621 2,862 146.8 3,482.6 4.3 0.798 4,319 34,408 5,877 45,335 $1,612 107%

30 % 285 621 2,902 156.0 3,523.1 4.2 0.797 4,457 37,502 5,945 48,636 $1,729 115%

0 % 61 515 3,424 354.2 3,939.8 4.4 0.810 3,428 46,239 6,648 57,048 $1,711 100%

10 % 65 510 3,462 369.4 3,971.6 4.4 0.812 3,456 47,300 6,702 58,190 $1,746 102%

20 % 65 507 3,497 385.6 4,004.3 4.3 0.812 3,528 49,673 6,757 60,690 $1,821 106%

30 % 62 507 3,534 402.7 4,040.4 4.2 0.810 3,640 53,431 6,818 64,621 $1,939 113%

Las Vegas

Washington, DC

Chicago

Minneapolis
This publication is available free of charge from: http://dx.doi.org/10.6028/NIST.TN.1848

67

Figure 5.7. Annual energy use for slab-on-grade houses at different levels of refrigerant overcharge 

relative to the annual energy use for the house in the same location when the heat pump operates with the 

nominal refrigerant charge

5.2.7 Effect of Excessive Refrigerant Subcooling

The level of this fault was determined by an increase of refrigerant subcooling at the TXV inlet at the 

operating condition defined by the AHRI Standard 210/240 test-A (AHRI, 2008). Refrigerant subcooling 

is indicative of refrigerant charge in a TXV-equipped system, and excessive subcooling is equivalent to the 

fault of refrigerant overcharge. When the amount of subcooling at the TXV inlet is increased the cooling 

system performance is degraded. Table 5.21 shows the results for the slab-on-grade house, and Table 5.22

shows the results for the basement house. Figure 5.8 shows the relative energy use for the slab-on-grade 

house, which provides a good representation of the energy use in the house with a basement as well. In 

general, increasing subcooling increases the capacity of the unit but degrades its efficiency. Both the 

cooling and heating energy use increased by about 20 % at the maximum fault level (200 %, i.e., an 

increase of subcooling from 4.4 °C (8.0 °F) to 13.2 °C (24.0 °F)). We may note that a 100 % increase in 

subcooling corresponds approximately to the 20 % overcharge fault discussed in Section 5.2.6. 

Table 5.21. Effect of excessive refrigerant subcooling on annual energy use for a slab-on-grade house

Note: Subcooling of 4.4 °C (8.0 °F) was used as a no-fault condition.

Excessive 

Sub-Cooling

 Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC 

SHR (-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

0 % 1,512 1,981 749 5.1 2,730.5 4.3 0.785 16,660 8,537 5,529 31,457 $743 100%

100 % 1,432 1,964 735 4.9 2,699.3 4.1 0.782 17,560 9,496 5,466 33,253 $785 106%

200 % 1,483 1,976 710 4.5 2,686.0 3.5 0.786 20,480 10,377 5,439 37,028 $874 118%

0 % - 1,966 865 0.3 2,831.1 3.7 0.999 20,531 11,251 6,687 39,200 $1,372 100%

100 % - 1,950 848 0.4 2,797.3 3.6 0.998 21,599 12,522 6,607 41,460 $1,451 106%

200 % - 1,971 818 0.3 2,789.5 3.0 1.000 25,241 13,716 6,589 46,277 $1,620 118%

0 % 253 1,207 1,971 89.0 3,178.0 4.5 0.809 8,098 21,759 5,363 35,952 $1,408 100%

100 % 194 1,199 1,954 90.1 3,153.1 4.3 0.803 8,565 24,299 5,321 38,916 $1,524 108%

200 % 223 1,208 1,912 88.8 3,119.5 3.6 0.807 10,023 26,696 5,264 42,714 $1,673 119%

0 % 189 1,031 2,833 281.2 3,863.9 4.5 0.827 6,816 37,118 6,520 51,186 $1,820 100%

100 % 180 1,024 2,812 286.4 3,836.3 4.3 0.820 7,217 41,052 6,474 55,475 $1,972 108%

200 % 183 1,031 2,769 282.3 3,799.9 3.7 0.824 8,446 44,775 6,412 60,365 $2,146 118%

0 % 13 897 3,432 612.5 4,328.9 4.5 0.846 5,912 55,105 7,305 69,053 $2,072 100%

100 % 11 892 3,416 620.7 4,308.1 4.3 0.839 6,266 60,185 7,270 74,453 $2,234 108%

200 % 12 898 3,375 614.9 4,272.7 3.6 0.843 7,332 64,959 7,210 80,233 $2,407 116%

Las Vegas

Washington, DC

Chicago

Minneapolis

Houston

90

95

100

105

110

115

120

125

130

Houston Las Vegas Wash., DC Chicago Minneapolis

Relative energy use (%)

Nominal -10% -20% -30%
This publication is available free of charge from: http://dx.doi.org/10.6028/NIST.TN.1848

68

Table 5.22. Effect of excessive refrigerant subcooling on annual energy use for a house with basement

Note: Subcooling of 4.4 °C (8.0 °F) was used as a no-fault condition.

Figure 5.8. Annual energy use for slab-on-grade houses at different level of refrigerant subcooling relative

to the annual energy use for the house in the same location with the heat pump operating with the nominal 

refrigerant charge and subcooling

5.2.8 Effect of Non-Condensable Gases

If the refrigerant system gets non-condensable gases (e.g., air) mixed in with the refrigerant, the 

performance of the unit is degraded. Table 5.23 shows the results for the slab-on-grade house, and Table 

5.24 shows the results for the basement house. The overall results show a (1 ~ 2) % energy use increase in 

climates with a significant heating season and a 4 % increase in the warmer climates. The moisture 

removal capability of the unit is only minimally affected by the non-condensable gases in the system. 

Excessive 

Sub-Cooling

 Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC 

SHR (-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

0 % - 1,552 718 0.3 2,269.9 3.7 1.000 16,107 9,407 5,362 31,607 $1,106 100%

100 % - 1,533 705 0.3 2,237.5 3.5 0.999 16,878 10,496 5,285 33,391 $1,169 106%

200 % - 1,554 682 0.3 2,236.6 3.0 1.000 19,787 11,530 5,283 37,332 $1,307 118%

0 % 654 742 1,907 27.0 2,649.5 4.4 0.775 5,008 19,120 4,471 29,330 $1,149 100%

100 % 532 737 1,891 28.0 2,628.0 4.2 0.770 5,293 21,594 4,435 32,054 $1,255 109%

200 % 620 741 1,850 27.6 2,591.2 3.6 0.774 6,182 23,960 4,373 35,246 $1,380 120%

0 % 289 631 2,785 129.8 3,416.1 4.4 0.797 4,198 31,565 5,765 42,259 $1,503 100%

100 % 260 628 2,767 134.2 3,395.2 4.2 0.791 4,450 35,482 5,729 46,393 $1,650 110%

200 % 278 631 2,721 133.7 3,352.8 3.6 0.794 5,197 39,277 5,658 50,863 $1,808 120%

0 % 61 515 3,424 354.2 3,939.8 4.4 0.810 3,428 46,239 6,648 57,048 $1,711 100%

100 % 57 513 3,410 362.3 3,922.9 4.2 0.803 3,633 51,396 6,620 62,380 $1,871 109%

200 % 60 516 3,372 359.6 3,887.3 3.6 0.808 4,244 56,387 6,560 67,922 $2,038 119%

Las Vegas

Washington, DC

Chicago

Minneapolis

90

95

100

105

110

115

120

125

130

Houston Las Vegas Wash., DC Chicago Minneapolis

Relative energy use (%)

Nominal 100% 200%
This publication is available free of charge from: http://dx.doi.org/10.6028/NIST.TN.1848

69

Table 5.23. Effect of non-condensable gases on annual energy use for a slab-on-grade house

Table 5.24. Effect of non-condensable gases on annual energy use for a house with basement

5.2.9 Effect of Voltage

When input voltage to the unit is changed from the nominal value, the performance of the unit is degraded. 

Tables 5.25 and 5.26 show the results for the slab-on-grade house and the basement house, respectively. 

The condition of 25 % overvoltage results in a (9 ~10) % increase in annual energy consumption. This

effect on the energy use does not include an adjustment for indoor fan power change with voltage. The 

undervoltage of 8 % resulted in an insignificant (within 1 %) change in the energy use. Higher levels of 

undervoltage were not studied because of a possible heat pump catastrophic failure.

Non 

Condensibles

 Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC 

SHR (-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

0 % 1,512 1,981 749 5.1 2,730.5 4.3 0.785 16,660 8,537 5,529 31,457 $743 100%

10 % 1,527 2,006 735 4.9 2,740.9 4.2 0.785 17,359 8,579 5,550 32,220 $761 102%

20 % 1,579 1,985 713 4.3 2,697.7 4.0 0.787 17,947 8,598 5,463 32,739 $773 104%

0 % - 1,966 865 0.3 2,831.1 3.7 0.999 20,531 11,251 6,687 39,200 $1,372 100%

10 % - 1,976 848 0.3 2,823.9 3.6 0.999 21,368 11,295 6,670 40,065 $1,402 102%

20 % - 1,949 821 0.3 2,769.7 3.5 1.000 22,127 11,328 6,542 40,730 $1,426 104%

0 % 253 1,207 1,971 89.0 3,178.0 4.5 0.809 8,098 21,759 5,363 35,952 $1,408 100%

10 % 255 1,234 1,947 86.0 3,180.8 4.3 0.809 8,468 21,875 5,368 36,442 $1,427 101%

20 % 277 1,233 1,901 81.1 3,133.9 4.1 0.810 8,793 21,906 5,289 36,719 $1,438 102%

0 % 189 1,031 2,833 281.2 3,863.9 4.5 0.827 6,816 37,118 6,520 51,186 $1,820 100%

10 % 186 1,055 2,802 275.8 3,856.9 4.3 0.827 7,126 37,276 6,508 51,642 $1,836 101%

20 % 188 1,055 2,754 264.2 3,808.7 4.2 0.829 7,395 37,352 6,427 51,905 $1,846 101%

0 % 13 897 3,432 612.5 4,328.9 4.5 0.846 5,912 55,105 7,305 69,053 $2,072 100%

10 % 13 918 3,406 603.0 4,324.7 4.3 0.847 6,182 55,304 7,298 69,515 $2,085 101%

20 % 14 919 3,366 582.3 4,284.4 4.1 0.848 6,416 55,348 7,230 69,726 $2,092 101%

Las Vegas

Washington, DC

Chicago

Minneapolis

Houston

Non 

Condensibles

 Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC 

SHR (-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

0 % - 1,552 718 0.3 2,269.9 3.7 1.000 16,107 9,407 5,362 31,607 $1,106 100%

10 % - 1,565 705 0.3 2,270.4 3.5 1.000 16,771 9,474 5,363 32,339 $1,132 102%

20 % - 1,550 685 0.3 2,235.4 3.4 1.000 17,390 9,540 5,280 32,941 $1,153 104%

0 % 654 742 1,907 27.0 2,649.5 4.4 0.775 5,008 19,120 4,471 29,330 $1,149 100%

10 % 649 760 1,882 26.0 2,641.7 4.2 0.775 5,236 19,279 4,458 29,704 $1,163 101%

20 % 677 761 1,841 23.2 2,602.0 4.0 0.776 5,438 19,434 4,391 29,995 $1,175 102%

0 % 289 631 2,785 129.8 3,416.1 4.4 0.797 4,198 31,565 5,765 42,259 $1,503 100%

10 % 288 647 2,753 127.3 3,400.0 4.2 0.797 4,387 31,817 5,737 42,674 $1,517 101%

20 % 287 649 2,706 119.9 3,354.6 4.1 0.799 4,560 32,046 5,661 42,999 $1,529 102%

0 % 61 515 3,424 354.2 3,939.8 4.4 0.810 3,428 46,239 6,648 57,048 $1,711 100%

10 % 61 528 3,399 347.5 3,927.5 4.2 0.810 3,584 46,560 6,628 57,503 $1,725 101%

20 % 64 530 3,354 333.6 3,884.3 4.1 0.812 3,726 46,815 6,555 57,828 $1,735 101%

Las Vegas

Washington, DC

Chicago

Minneapolis
This publication is available free of charge from: http://dx.doi.org/10.6028/NIST.TN.1848

70

Table 5.25. Effect of voltage on annual energy use for a slab-on-grade house

Table 5.26. Effect of voltage on annual energy use for a house with basement

Figure 5.9. Annual energy use for slab-on-grade houses at different levels of input voltages relative to the 

energy use for the house in the same location when the heat pump operates with nominal voltage

Voltage Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC 

SHR (-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

-8 % 1,508 1,992 748 5.0 2,740.1 4.3 0.785 16,677 8,464 5,549 31,421 $742 100%

0 % 1,512 1,981 749 5.1 2,730.5 4.3 0.785 16,660 8,537 5,529 31,457 $743 100%

8 % 1,519 1,974 752 5.3 2,725.8 4.3 0.785 16,970 8,733 5,520 31,954 $754 102%

25 % 1,547 1,966 767 5.9 2,733.2 3.9 0.786 18,676 9,616 5,535 34,559 $816 110%

-8 % - 1,977 863 0.3 2,840.0 3.7 0.999 20,715 11,143 6,708 39,299 $1,375 100%

0 % - 1,966 865 0.3 2,831.1 3.7 0.999 20,531 11,251 6,687 39,200 $1,372 100%

8 % - 1,958 870 0.3 2,827.2 3.7 0.999 20,741 11,523 6,678 39,674 $1,389 101%

25 % - 1,947 888 0.4 2,835.3 3.4 0.999 22,465 12,694 6,697 42,587 $1,491 109%

-8 % 252 1,213 1,969 88.6 3,181.3 4.5 0.809 8,062 21,594 5,368 35,756 $1,400 99%

0 % 253 1,207 1,971 89.0 3,178.0 4.5 0.809 8,098 21,759 5,363 35,952 $1,408 100%

8 % 256 1,202 1,979 89.7 3,181.8 4.4 0.809 8,289 22,211 5,369 36,601 $1,434 102%

25 % 274 1,197 2,010 94.8 3,207.7 3.9 0.810 9,211 24,175 5,413 39,530 $1,548 110%

-8 % 188 1,035 2,830 280.3 3,865.2 4.5 0.827 6,770 36,879 6,522 50,904 $1,810 99%

0 % 189 1,031 2,833 281.2 3,863.9 4.5 0.827 6,816 37,118 6,520 51,186 $1,820 100%

8 % 189 1,027 2,842 283.5 3,868.8 4.4 0.827 6,988 37,781 6,529 52,030 $1,850 102%

25 % 189 1,022 2,879 294.2 3,901.4 3.9 0.828 7,786 40,678 6,584 55,779 $1,983 109%

-8 % 13 901 3,430 611.0 4,331.0 4.5 0.846 5,871 54,824 7,309 68,736 $2,062 100%

0 % 13 897 3,432 612.5 4,328.9 4.5 0.846 5,912 55,105 7,305 69,053 $2,072 100%

8 % 14 894 3,440 616.2 4,334.1 4.4 0.846 6,064 55,920 7,314 70,029 $2,101 101%

25 % 14 890 3,470 635.5 4,360.3 3.9 0.848 6,764 59,502 7,358 74,356 $2,231 108%

Washington, DC

Chicago

Minneapolis

Houston

Las Vegas

Voltage Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC 

SHR (-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

-8 % - 1,561 716 0.3 2,277.2 3.6 1.000 16,243 9,319 5,379 31,672 $1,109 100%

0 % - 1,552 718 0.3 2,269.9 3.7 1.000 16,107 9,407 5,362 31,607 $1,106 100%

8 % - 1,545 721 0.3 2,266.0 3.6 1.000 16,275 9,635 5,352 31,994 $1,120 101%

25 % - 1,536 736 0.3 2,272.1 3.4 1.000 17,649 10,604 5,367 34,351 $1,202 109%

-8 % 656 746 1,905 26.8 2,650.9 4.4 0.775 4,987 18,969 4,473 29,161 $1,142 99%

0 % 654 742 1,907 27.0 2,649.5 4.4 0.775 5,008 19,120 4,471 29,330 $1,149 100%

8 % 657 740 1,916 27.1 2,655.6 4.3 0.775 5,125 19,545 4,481 29,883 $1,170 102%

25 % 674 736 1,945 29.6 2,681.6 3.9 0.775 5,687 21,380 4,525 32,324 $1,266 110%

-8 % 286 634 2,782 129.7 3,415.8 4.5 0.797 4,175 31,353 5,764 42,023 $1,494 99%

0 % 289 631 2,785 129.8 3,416.1 4.4 0.797 4,198 31,565 5,765 42,259 $1,503 100%

8 % 289 629 2,793 131.8 3,422.0 4.3 0.797 4,302 32,212 5,775 43,020 $1,530 102%

25 % 295 627 2,829 138.4 3,455.6 3.9 0.798 4,792 34,940 5,831 46,295 $1,646 110%

-8 % 61 518 3,421 353.6 3,938.8 4.4 0.810 3,408 45,973 6,647 56,759 $1,703 99%

0 % 61 515 3,424 354.2 3,939.8 4.4 0.810 3,428 46,239 6,648 57,048 $1,711 100%

8 % 61 514 3,432 357.0 3,946.0 4.3 0.810 3,514 47,027 6,659 57,931 $1,738 102%

25 % 63 511 3,469 369.9 3,980.7 3.9 0.811 3,911 50,497 6,717 61,857 $1,856 108%

Washington, DC

Chicago

Minneapolis

Las Vegas

90

95

100

105

110

115

120

125

130

Houston Las Vegas Wash., DC Chicago Minneapolis

Relative energy use (%)

-8% Nominal 8% 25%
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71

5.2.10 Effect of TXV Sizing 

Only undersizing of the TXV in the cooling mode is considered in this study. When the size of the TXV 

does not match the compressor size, the performance of the system is degraded. Table 5.27 shows the 

results for the slab-on-grade houses and Table 5.28 shows the results for the basement houses. Generally,

the impact is modest at 20 % undersizing in any climate and remains relatively small for Minneapolis at 

even higher fault levels. However, the impact becomes significant at 40 % undersizing, particularly in hot 

climates where the energy use increases by (10 ~ 14) %. Moisture removal is only modestly affected. 

Table 5.27. Effect of TXV sizing on annual energy use for a slab-on-grade house

Table 5.28. Effect of TXV sizing on annual energy use for a house with basement

Undersized 

TXV

 Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC 

SHR (-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

0 % 1,512 1,981 749 5.1 2,730.5 4.3 0.785 16,660 8,537 5,529 31,457 $743 100%

20 % 1,516 2,000 749 5.1 2,748.6 4.3 0.785 16,855 8,536 5,566 31,688 $748 101%

40 % 1,534 2,312 749 5.1 3,060.6 3.6 0.784 20,357 8,533 6,198 35,819 $846 114%

60 % 1,575 2,767 749 5.1 3,515.9 2.8 0.780 25,508 8,531 7,120 41,890 $989 133%

0 % - 1,966 865 0.3 2,831.1 3.7 0.999 20,531 11,251 6,687 39,200 $1,372 100%

20 % - 1,973 865 0.3 2,837.4 3.7 0.999 20,623 11,242 6,702 39,298 $1,375 100%

40 % - 2,210 865 0.3 3,074.8 3.3 1.000 23,723 11,242 7,263 42,959 $1,504 110%

60 % - 2,647 864 0.3 3,511.2 2.6 1.000 29,509 11,235 8,294 49,770 $1,742 127%

0 % 253 1,207 1,971 89.0 3,178.0 4.5 0.809 8,098 21,759 5,363 35,952 $1,408 100%

20 % 257 1,234 1,971 89.0 3,204.0 4.3 0.809 8,341 21,754 5,407 36,233 $1,419 101%

40 % 260 1,449 1,971 89.0 3,420.1 3.6 0.810 10,317 21,758 5,771 38,577 $1,511 107%

60 % 258 1,751 1,970 88.9 3,720.8 2.8 0.810 13,097 21,748 6,279 41,855 $1,639 116%

0 % 189 1,031 2,833 281.2 3,863.9 4.5 0.827 6,816 37,118 6,520 51,186 $1,820 100%

20 % 188 1,058 2,833 281.2 3,890.7 4.4 0.827 7,064 37,117 6,566 51,478 $1,830 101%

40 % 188 1,246 2,833 281.2 4,079.1 3.6 0.830 8,792 37,116 6,884 53,523 $1,903 105%

60 % 182 1,512 2,833 281.2 4,344.2 2.8 0.834 11,229 37,113 7,331 56,405 $2,006 110%

0 % 13 897 3,432 612.5 4,328.9 4.5 0.846 5,912 55,105 7,305 69,053 $2,072 100%

20 % 13 922 3,432 612.5 4,354.1 4.3 0.847 6,139 55,106 7,348 69,324 $2,080 100%

40 % 13 1,087 3,431 612.5 4,518.4 3.5 0.851 7,649 55,099 7,625 71,104 $2,133 103%

60 % 11 1,321 3,431 612.5 4,751.9 2.8 0.856 9,787 55,097 8,019 73,634 $2,209 107%

Houston

Las Vegas

Washington, DC

Chicago

Minneapolis

Undersized 

TXV

 Hours 

Above 

55 % RH 

 AC 

Runtime 

(h) 

 Htg 

Runtime 

(h) 

 Backup 

Heat 

Runtime 

(h) 

 AHU Fan 

Runtime 

(h) 

 AC COP 

(-) 

 AC 

SHR (-) 

 AC 

Energy 

(MJ) 

 Htg 

Energy 

(MJ) 

 AHU Fan 

Energy 

(MJ) 

 TOTAL 

ENERGY 

(MJ) 

Total 

Costs

Relative 

Energy

0 % - 1,552 718 0.3 2,269.9 3.7 1.000 16,107 9,407 5,362 31,607 $1,106 100%

20 % - 1,558 718 0.3 2,276.2 3.7 1.000 16,187 9,408 5,376 31,704 $1,110 100%

40 % - 1,738 718 0.3 2,455.9 3.2 1.000 18,575 9,408 5,801 34,516 $1,208 109%

60 % - 2,117 717 0.3 2,834.1 2.5 1.000 23,631 9,403 6,694 40,460 $1,416 128%

0 % 654 742 1,907 27.0 2,649.5 4.4 0.775 5,008 19,120 4,471 29,330 $1,149 100%

20 % 653 756 1,907 27.0 2,663.0 4.3 0.775 5,132 19,120 4,494 29,477 $1,155 101%

40 % 649 877 1,907 27.0 2,784.1 3.5 0.778 6,269 19,121 4,698 30,819 $1,207 105%

60 % 635 1,066 1,907 27.0 2,972.8 2.8 0.782 8,022 19,120 5,017 32,890 $1,288 112%

0 % 289 631 2,785 129.8 3,416.1 4.4 0.797 4,198 31,565 5,765 42,259 $1,503 100%

20 % 283 645 2,785 129.8 3,430.6 4.3 0.797 4,327 31,569 5,789 42,418 $1,508 100%

40 % 284 750 2,785 129.8 3,534.9 3.5 0.801 5,312 31,564 5,965 43,573 $1,549 103%

60 % 282 908 2,785 129.7 3,692.9 2.7 0.808 6,779 31,559 6,232 45,302 $1,611 107%

0 % 61 515 3,424 354.2 3,939.8 4.4 0.810 3,428 46,239 6,648 57,048 $1,711 100%

20 % 61 527 3,424 354.2 3,950.5 4.3 0.811 3,531 46,236 6,667 57,165 $1,715 100%

40 % 59 611 3,424 354.2 4,035.1 3.5 0.815 4,326 46,235 6,809 58,102 $1,743 102%

60 % 56 739 3,424 354.2 4,163.3 2.7 0.822 5,516 46,235 7,026 59,507 $1,785 104%

Las Vegas

Washington, DC

Chicago

Minneapolis
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72

Figure 5.10. Annual energy use for slab-on-grade houses at different levels of TXV undersizing relative to 

the annual energy use for the house when the heat pump operates with a properly sized TXV

5.2.11 Discussion of the Effects of Single Faults

Figure 5.11 shows examples of annual energy used by a heat pump installed with different installation 

faults in a slab-on-grade house. The levels of individual faults were selected to reflect, to some degree, the

installation condition which might not be noticed by a poorly trained technician. (The authors recognize 

the speculative aspect of this selection.)

Figure 5.11. Annual energy use by a heat pump in a slab-on-grade house resulting from a single-fault 

installation, referenced to a fault-free installation. (Table 5.29 shows the selected fault levels)

90

95

100

105

110

115

120

125

130

135

140

Houston Las Vegas Wash., DC Chicago Minneapolis

Relative energy use (%)

Nominal 20% 40% 60%

90

95

100

105

110

115

120

125

130

SIZ DUCT AF UC OC NC VOL TXV

Relative energy use (%)

Houston Las Vegas Wash. DC Chicago Minneapolis
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73

Table 5.29 Levels of individual faults used in Figure 5.11

Fault Type Fault Level

(%)

Heat Pump Sizing (SIZ)(a) + 50

Duct Leakage (DUCT) 30

Indoor Coil Airflow (AF) - 36

Refrigerant Undercharge (UC) - 30

Refrigerant Overcharge (OC) + 30

Non-Condensable Gases (NC) 10

Electric Voltage (VOL) + 8

TXV Undersizing (TXV) - 40

 (a) Oversize scenario (2) described in Section 5.2.2.

Simulation results show no drastic differences in the effect of installation faults on energy use in a slab-ongrade house and a basement house, except for the duct leakage fault. For the slab-on-grade house, this fault 

has the potential to result in a higher increase in energy use that any other fault. The impact of this fault is 

higher for the heating dominated climate (Chicago and Minneapolis, 26 %) than for the cooling dominated 

climate (Houston, 18 %). Obviously, duct leakage will also result in some increase of energy use for the 

basement house; however, the model we used would not discern this effect. 

The second most influential fault is refrigerant undercharge. For the 30 % undercharge fault level, the 

energy use increase is of the order of 20 % irrespective of the climate and building type. Refrigerant 

overcharge can also result in a significant increase in energy use, (10 ~16) % at the 30 % overcharge fault 

level. Improper indoor airflow can affect similar performance degradation. 

Equipping a house with an oversized heat pump has a small effect if the air duct is oversized accordingly 

(which may be the case with a new construction). However, if the air duct is too restrictive and the 

nominal indoor airflow is maintained by adjusting the fan speed (scenario (2)), a 15 % increase in energy 

use for the house in Houston is predicted. 

The cooling TXV undersized fault has also the potential to significantly increase the energy use. The effect 

of this fault will be most pronounced in localities with a high number of cooling mode operating hours. 

The cooling mode TXV undersized by 40 % results in (9 ~ 14) % more energy used in Houston as

compared to a (3 ~ 5) % in Chicago.

The impact of the remaining faults – non-condensables and improper voltage – is under 4 %. The noncondensables and improper voltage faults, however, represent a substantial risk for durability of equipment 

and are very important to be diagnosed during a heat pump installation. 
This publication is available free of charge from: http://dx.doi.org/10.6028/NIST.TN.1848

74

5.3 Simulations with Dual Faults

5.3.1 Studied Fault Combinations

The analysis in this section considers the combination of two faults, A and B. Each set of faults was 

considered in four combinations (Table 5.30).

Table 5.30. Combinations of studied faults

Fault combination 

case

Level of fault A Level of fault B 

A moderate moderate

B moderate worst

C worst moderate

D worst worst

The moderate level will be the value at the middle of the range, while the worst level will be the highest 

(or lowest) probable level of the fault value. Table 5.31 defines the set or combinations of dual faults 

simulated for cases where heating and cooling were considered together. Table 5.32 defines the sets of 

faults that apply for the cooling-only case. The most right-hand column in both tables shows an 

approximate effect of the studied fault sets on the energy use: the faults effects may be additive (A+B), 

less than additive (<A+B), or greater the additive (>A+B).

Table 5.31. Dual fault sets considered in simulations (heating and cooling) and their approximate 

collective effect on annual energy use

Fault set

#

Fault A 

(moderate & worst level)(a)

Fault B 

(moderate & worst level) Effect on energy use

1 Duct leakage 

(20 %, 40 %)

Oversize(b)

(25 %, 50 %) A+B

2 Duct leakage

(20 %, 40 %)

Indoor coil airflow

(-15 %, -36 %) < A+B

3 Duct leakage

(20 %, 40 %)

Refrigerant undercharge

(-15 %, -30 %) A+B or > A+B

4 Duct leakage 

(20 %, 40 %)

Refrigerant overcharge

(15 %, 30 %) A+B

5 Duct leakage

(20 %, 40 %)

Non-condensables

(10 %, 20 %) A+B

6 Oversize(b)

(25 %, 50 %)

Refrigerant undercharge

(-15 %, -30 %) A+B

7 Oversize(b)

(25 %, 50 %)

Refrigerant overcharge

(15 %, 30 %) A+B

8 Oversize(b)

(25 %, 50 %)

Non-condensables 

(10 %, 20 %) A+B

9 Indoor coil airflow

(-15 %, -36 %) 

Refrigerant undercharge 

(-15 %, -30 %) < A+B

10 Indoor coil airflow

(-15 %, -36 %) 

Refrigerant overcharge 

(15 %, 30 %) < A+B

11 Indoor coil airflow

(-15 %, -36 %) 

Non-condensables 

(10 %, 20 %) < A+B

(a) moderate = mid-level value, worst = lowest/highest level value

(b) Oversize scenario (2) was selected because it covers the prevalent field bias (undersized ducts)
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75

Table 5.32. Dual fault sets considered in simulations (heating and cooling) and their approximate 

collective effect on annul energy use; TXV fault existing in cooling only (a)

(a) Faults listed as Faults A exist in cooling and heating

(b) moderate = mid-level value, worst = lowest/highest level value

(c) Oversize scenario (2) was selected because it covers the prevalent field bias (undersized ducts)

5.3.2. Effects of Dual Faults

Simulations were performed for 14 dual fault sets, with 4 runs per set, in the 9 house/climate combinations 

for a total of 504 runs. Because of similarity between the obtained results, the tables below are limited to 

representative cases, which include the slab-on-grade house for Houston, Washington, DC, and 

Minneapolis, and the house with a basement for Washington, DC. For the Houston house, Table 5.33

shows results for dual fault sets 1 through 5, which represent all studied dual faults involving duct leakage;

Table 5.34 shows results for dual fault sets 6 through 8, which represent all studied dual faults involving 

the oversized heat pump, except the case with duct leakage presented in Table 5.33; and Table 5.35

presents the remaining three studied cases with dual faults present in both cooling and heating. Table 

5.36 presents the effect on annual energy use of the undersized cooling TXV with either duct leakage, 

oversized heat pump, or low airflow rate faults, which occur in both cooling and heating mode. Tables 

5.37 through 5.47 present simulation results for the remaining cases. For nine out of fourteen sets studied, 

the effect of dual faults was approximately additive (Table 5.31). For the remaining five sets – all 

involving indoor coil airflow – the effect was less than additive. A few results that are not immediately 

intuitive are discussed at the end of this section.

Table 5.33. Relative energy use for dual fault sets 1 to 5 for the slab-on-grade house in Houston

Duct leakage with oversized heat pump, low airflow rate, undercharged, overcharged, and non-condensable 

gases faults

Dual Fault Set: 1

20% Duct 

Leakage

40% Duct 

Leakage

100% 109% 128%

25% Oversized 109% 121% 148%

50% Oversized 115% 128% 159%

Dual Fault Set: 2

20% Duct 

Leakage

40% Duct 

Leakage

100% 109% 128%

-15% Airflow 104% 110% 129%

-36% Airflow 112% 114% 130%

Dual Fault Set: 3

20% Duct 

Leakage

40% Duct 

Leakage

100% 109% 128%

15% Undercharged 105% 115% 136%

30% Undercharged 121% 132% 156%

Dual Fault Set: 4

20% Duct 

Leakage

40% Duct 

Leakage

100% 109% 128%

15% Overcharged 103% 112% 132%

30% Overcharged 110% 119% 141%

Dual Fault Set: 5

20% Duct 

Leakage

40% Duct 

Leakage

100% 109% 128%

10% Non-Condensibles 102% 112% 131%

20% Non-Condensibles 104% 113% 133%

Fault set # Fault A 

(moderate & worst level)(b)

Fault B

(moderate & worst level) Effect on energy use

12 Duct leakage 

(20 %, 40 %)

Cooling TXV undersizing

(-20 %, -60 %) A+B

13 Oversize(c)

(25 %, 50 %)

Cooling TXV undersizing

(-20 %, -60 %)) A+B

14 Indoor coil airflow

(-15 %, -36 %) 

Cooling TXV undersizing

(-20 %, -60 %) < A+B
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76

Table 5.34. Relative energy use for dual fault sets 6 to 8 for the slab-on-grade house in Houston

Oversized heat pump with undercharged, overcharged, and non-condensable gases faults

Table 5.35. Relative energy use for dual fault sets 9 to 11 for the slab-on-grade house in Houston

Low airflow rate with undercharged, overcharged, and non-condensable gases faults

Table 5.36. Relative energy use for dual fault sets 12 to 14 involving cooling mode TXV for the slab-ongrade house in Houston 

Undersized TXV with duct leakage, oversized heat pump, and low airflow rate faults

Mult Fault Set: 10 -15% Airflow -36% Airflow

100% 104% 112%

15% Undercharged 105% 107% 111%

30% Undercharged 121% 123% 127%

Mult Fault Set: 11 -15% Airflow -36% Airflow

100% 104% 112%

15% Overcharged 103% 105% 109%

30% Overcharged 110% 112% 116%

Mult Fault Set: 12 -15% Airflow -36% Airflow

100% 104% 112%

10% Non-Condensibles 102% 104% 109%

20% Non-Condensibles 104% 106% 111%

Dual Fault Set: 13

20% Duct 

Leakage

40% Duct 

Leakage

100% 109% 128%

20% TXV Undersizing 101% 110% 129%

60% TXV Undersizing 133% 143% 163%

Dual Fault Set: 14

25% 

Oversized

50% 

Oversized

100% 109% 115%

20% TXV Undersizing 101% 110% 116%

60% TXV Undersizing 133% 150% 161%

Dual Fault Set: 15 -15% Airflow -36% Airflow

100% 104% 112%

20% TXV Undersizing 101% 103% 107%

60% TXV Undersizing 133% 135% 139%

Dual Fault Set: 7

25% 

Oversized

50% 

Oversized

100% 109% 115%

15% Undercharged 105% 115% 121%

30% Undercharged 121% 133% 140%

Dual Fault Set: 8

25% 

Oversized

50% 

Oversized

100% 109% 115%

15% Overcharged 103% 112% 118%

30% Overcharged 110% 119% 125%

Dual Fault Set: 9

25% 

Oversized

50% 

Oversized

100% 109% 115%

10% Non-Condensibles 102% 112% 118%

20% Non-Condensibles 104% 113% 120%

Dual Fault Set: 6 Dual Fault Set: 7

Dual Fault Set: 8

Dual Fault Set: 9 Dual Fault Set: 10

Dual Fault Set: 11

Dual Fault Set: 12 Dual Fault Set: 13

Dual Fault Set: 14
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77

Table 5.37. Relative energy use for dual fault sets 1 to 5 for the slab-on-grade house in Washington, DC

Duct leakage with oversized heat pump , low airflow rate, undercharged, overcharged, and non-condensable

gases faults

 

Table 5.38. Relative energy use for dual fault sets 6 to 8 for the slab-on-grade house in Washington, DC

Oversized heat pump with undercharged, overcharged, and non-condensable gases faults

Table 5.39. Relative energy use for dual fault sets 9 to 11 for the slab-on-grade house in Washington, DC

Low airflow rate with undercharged, overcharged, and non-condensable gases faults

Dual Fault Set: 1

20% Duct 

Leakage

40% Duct 

Leakage

100% 112% 139%

25% Oversized 103% 117% 152%

50% Oversized 105% 119% 156%

Dual Fault Set: 2

20% Duct 

Leakage

40% Duct 

Leakage

100% 112% 139%

-15% Airflow 105% 112% 137%

-36% Airflow 114% 114% 137%

Dual Fault Set: 3

20% Duct 

Leakage

40% Duct 

Leakage

100% 112% 139%

15% Undercharged 105% 117% 146%

30% Undercharged 123% 137% 172%

Dual Fault Set: 4

20% Duct 

Leakage

40% Duct 

Leakage

100% 112% 139%

15% Overcharged 104% 116% 145%

30% Overcharged 114% 127% 157%

Dual Fault Set: 5

20% Duct 

Leakage

40% Duct 

Leakage

100% 112% 139%

10% Non-Condensibles 101% 113% 140%

20% Non-Condensibles 102% 114% 141%

Mult Fault Set: 10 -15% Airflow -36% Airflow

100% 105% 114%

15% Undercharged 105% 106% 109%

30% Undercharged 123% 124% 127%

Mult Fault Set: 11 -15% Airflow -36% Airflow

100% 105% 114%

15% Overcharged 104% 106% 109%

30% Overcharged 114% 116% 119%

Mult Fault Set: 12 -15% Airflow -36% Airflow

100% 105% 114%

10% Non-Condensibles 101% 103% 106%

20% Non-Condensibles 102% 104% 107%

Dual Fault Set: 7

25% 

Oversized

50% 

Oversized

100% 103% 105%

15% Undercharged 105% 108% 110%

30% Undercharged 123% 127% 129%

Dual Fault Set: 8

25% 

Oversized

50% 

Oversized

100% 103% 105%

15% Overcharged 104% 107% 109%

30% Overcharged 114% 118% 119%

Dual Fault Set: 9

25% 

Oversized

50% 

Oversized

100% 103% 105%

10% Non-Condensibles 101% 105% 107%

20% Non-Condensibles 102% 106% 108%

Dual Fault Set: 6 Dual Fault Set: 7

Dual Fault Set: 8

Dual Fault Set: 9 Dual Fault Set: 10

Dual Fault Set: 11
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Table 5.40. Relative energy use for dual fault sets 12 to 14 involving the cooling mode TXV for the slabon-grade house in Washington, DC

Undersized TXV with duct leakage, oversized heat pump, and low airflow rate faults

Table 5.41. Relative energy use for dual fault sets 1 to 5 for the slab-on-grade house in Minneapolis

Duct leakage with oversized heat pump , low airflow rate, undercharged, overcharged, and non-condensable

gases faults

Table 5.42. Relative energy use for dual fault sets 6 to 8 for the slab-on-grade house in Minneapolis

Oversized heat pump with undercharged, overcharged, and non-condensable gases faults

Dual Fault Set: 13

20% Duct 

Leakage

40% Duct 

Leakage

100% 112% 139%

20% TXV Undersizing 101% 113% 140%

60% TXV Undersizing 116% 129% 157%

Dual Fault Set: 14

25% 

Oversized

50% 

Oversized

100% 103% 105%

20% TXV Undersizing 101% 104% 106%

60% TXV Undersizing 116% 123% 127%

Dual Fault Set: 15 -15% Airflow -36% Airflow

100% 105% 114%

20% TXV Undersizing 101% 102% 105%

60% TXV Undersizing 116% 118% 121%

Dual Fault Set: 1

20% Duct 

Leakage

40% Duct 

Leakage

100% 113% 140%

25% Oversized 100% 114% 148%

50% Oversized 99% 114% 149%

Dual Fault Set: 2

20% Duct 

Leakage

40% Duct 

Leakage

100% 113% 140%

-15% Airflow 103% 113% 138%

-36% Airflow 111% 113% 137%

Dual Fault Set: 3

20% Duct 

Leakage

40% Duct 

Leakage

100% 113% 140%

15% Undercharged 103% 116% 144%

30% Undercharged 117% 132% 162%

Dual Fault Set: 4

20% Duct 

Leakage

40% Duct 

Leakage

100% 113% 140%

15% Overcharged 104% 116% 144%

30% Overcharged 112% 125% 153%

Dual Fault Set: 5

20% Duct 

Leakage

40% Duct 

Leakage

100% 113% 140%

10% Non-Condensibles 101% 113% 140%

20% Non-Condensibles 101% 113% 140%

Dual Fault Set: 12 Dual Fault Set: 13

Dual Fault Set: 14

Dual Fault Set: 7

25% 

Oversized

50% 

Oversized

100% 100% 99%

15% Undercharged 103% 103% 103%

30% Undercharged 117% 118% 117%

Dual Fault Set: 8

25% 

Oversized

50% 

Oversized

100% 100% 99%

15% Overcharged 104% 103% 103%

30% Overcharged 112% 112% 112%

Dual Fault Set: 9

25% 

Oversized

50% 

Oversized

100% 100% 99%

10% Non-Condensibles 101% 101% 101%

20% Non-Condensibles 101% 101% 101%

Dual Fault Set: 6 Dual Fault Set: 7

Dual Fault Set: 8
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79

Table 5.43. Relative energy use for dual fault sets 9 to 11 for the slab-on-grade house in Minneapolis

Low airflow rate with undercharged, overcharged, and non-condensable gases faults

Table 5.44. Relative energy use for dual fault sets 12 to 14 involving the cooling mode TXV for the slabon-grade house in Minneapolis 

Undersized TXV with duct leakage, oversized heat pump, and low airflow rate faults

Table 5.45. Relative energy use for dual fault sets 6 to 8 for the basement house in Washington, DC

Oversized heat pump with undercharged, overcharged, and non-condensable gases faults

Dual Fault Set: 10 -15% Airflow -36% Airflow

100% 103% 111%

15% Undercharged 103% 104% 106%

30% Undercharged 117% 118% 121%

Dual Fault Set: 11 -15% Airflow -36% Airflow

100% 103% 111%

15% Overcharged 104% 104% 107%

30% Overcharged 112% 112% 115%

Dual Fault Set: 12 -15% Airflow -36% Airflow

100% 103% 111%

10% Non-Condensibles 101% 102% 104%

20% Non-Condensibles 101% 102% 104%

Dual Fault Set: 13

20% Duct 

Leakage

40% Duct 

Leakage

100% 113% 140%

20% TXV Undersizing 100% 113% 140%

60% TXV Undersizing 107% 120% 147%

Dual Fault Set: 14

25% 

Oversized

50% 

Oversized

100% 100% 99%

20% TXV Undersizing 100% 100% 100%

60% TXV Undersizing 107% 108% 108%

Dual Fault Set: 15 -15% Airflow -36% Airflow

100% 103% 111%

20% TXV Undersizing 100% 101% 103%

60% TXV Undersizing 107% 108% 110%

Dual Fault Set: 9 Dual Fault Set: 10

Dual Fault Set: 11

Dual Fault Set: 12 Dual Fault Set: 13

Dual Fault Set: 14

Dual Fault Set: 7

25% 

Oversized

50% 

Oversized

100% 104% 108%

15% Undercharged 104% 109% 112%

30% Undercharged 121% 126% 129%

Dual Fault Set: 8

25% 

Oversized

50% 

Oversized

100% 104% 108%

15% Overcharged 105% 109% 112%

30% Overcharged 116% 120% 124%

Dual Fault Set: 9

25% 

Oversized

50% 

Oversized

100% 104% 108%

10% Non-Condensibles 101% 106% 109%

20% Non-Condensibles 102% 107% 111%

Dual Fault Set: 6 Dual Fault Set: 7

Dual Fault Set: 8
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80

Table 5.46. Relative energy use for dual fault sets 9 to 11 for the basement house in Washington, DC

Low airflow rate with undercharged, overcharged, and non-condensable gases faults

Table 5.47. Relative energy use for dual fault sets 13 to 14 involving the cooling mode TXV for the 

basement house in Washington, DC

Undersized TXV with duct leakage, oversized heat pump, and low airflow rate faults

While reviewing the above results, a reader may be surprised to see that in a few cases the energy use with 

two simultaneous faults is as at a similar level as that for the more influential single fault. The most 

confounding are perhaps the results obtained for the dual fault set # 2 involving air duct leakage and 

reduced indoor coil airflow (Table 5.37). In this case, for the 40 % duct leakage existing alone the energy 

use increases by 39 %, and for the 36 % reduction in the airflow the energy use increases by 14 %; 

however, when these two faults exist simultaneously the combined effect is an increase of energy use by 

37 %, which is less than that when the duct leakage fault exists alone. This result can be explained by the 

fact that at a lowered airflow the heat pump satisfies the load using less air (it produces a larger 

temperature spread between the return and supply air). Hence, in absolute numbers, the amount of energy 

lost due duct leakage is smaller because the leaked air is a percentage of the total airflow. Simply, duct 

leakage is a dominating fault, and a reduction of the effect of this fault more than compensates for the 

losses associated with the reduced airflow (decreased air-side heat transfer coefficient and increased 

compressor power due to increased temperature lift).

Also interesting results for the low indoor airflow combined with either the refrigerant overcharge (dual set 

fault # 10) or non-condensable gases (dual set fault # 11) can be reviewed in Table 5.39. If the low airflow 

fault exists alone, the energy use increases by 14 % for the 36 % airflow reduction. This fault demonstrates 

itself in a lower temperature of the evaporator, which results in a somewhat lower sensible capacity and 

increased latent capacity of the air conditioner. Since in performed simulations the air conditioner had to 

satisfy the thermostat (i.e., the same sensitive load) and the rate of moisture removal increased, the energy 

use increased. Now, refrigerant overcharge fault or non-condensables fault causes the condenser pressure 

to increase. This pulls up the pressure (and temperature) of the evaporator, which reduces the latent load 

the air conditioner handles. At moderate levels of the overcharge and non-condensables faults, the 

energetic benefit of the lowered latent load is greater than that of a modest COP penalty associate with 

Dual Fault Set: 10 -15% Airflow -36% Airflow

100% 105% 113%

15% Undercharged 104% 105% 106%

30% Undercharged 121% 122% 124%

Dual Fault Set: 11 -15% Airflow -36% Airflow

100% 105% 113%

15% Overcharged 105% 106% 107%

30% Overcharged 116% 117% 118%

Dual Fault Set: 12 -15% Airflow -36% Airflow

100% 105% 113%

10% Non-Condensibles 101% 102% 103%

20% Non-Condensibles 102% 103% 105%

Dual Fault Set: 14

25% 

Oversized

50% 

Oversized

100% 104% 108%

20% TXV Undersizing 101% 105% 108%

60% TXV Undersizing 112% 118% 123%

Dual Fault Set: 15 -15% Airflow -36% Airflow

100% 105% 113%

20% TXV Undersizing 101% 101% 103%

60% TXV Undersizing 112% 113% 115%

Dual Fault Set: 9 Dual Fault Set: 10

Dual Fault Set: 11

Dual Fault Set: 13 Dual Fault Set: 14
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81

these faults. Consequently, moderate levels of refrigerant overcharge and non-condensables faults caused 

a reduction of energy used by the unit with 36 % reduced air flow. Greater levels of these faults reverse 

this energy use trend. (Note that the above explanation discusses the first order effects of a rather 

complicated reaction of the systems to these faults, e.g., a lower air-side heat transfer coefficient, lower 

indoor fan power, and the effect on performance in the heating mode).

Relatively less perplexing is the interaction between the low airflow fault and undersized TXV fault 

(Table 5.40, dual fault set # 14). In this case, a 20 % undersized cooling-mode TXV improved the 

performance of the system operated with a reduced indoor coil airflow. Since a reduced airflow reduces 

the system capacity, a TVX that was 20 % undersized for the rated capacity showed to be a better match 

for the ‘reduced capacity’ system than the TXV properly sized for the rated capacity.

It should be noted that airflow reduction lowers equipment capacity and may compromise occupant’s

comfort when approaching design conditions. Additionally, in extreme cases or in combination with other 

faults, it may lead to indoor coil frosting during cooling operation and equipment tripping or failure.

5.3.3. Discussion of the Effects of Dual Faults

As expected, the collective impact of two simultaneous faults on the energy consumption varies and 

depends on the faults considered. In most cases the collective effect can be described as being additive; 

however, the effect can exceed or be markedly below this additive value, including being approximately 

equal to the individual effect of one of the faults involved, as noted in Tables 5.31 and 5.32. The above 

characterization applies to all house/climate combinations. The relative impact on energy use also is 

similar for all cases studied (Figures 5.12, and 5.13).

Figure 5.12. Annual energy use for slab-on-grade houses with 14 dual-faults relative to the energy use for 

the houses with fault-free installations

(Faults defined in Tables 5.31 and 5.32; Table 5.30 case d, worst level for both faults) 

90

100

110

120

130

140

150

160

170

180

1 2 3 4 5 6 7 8 9 10 11 12 13 14

Relative energy use (%)

Houston Las Vegas Wash. DC Chicago Minneapolis
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82

Figure 5.13. Annual energy use for houses with basement with 8 dual-fault installations referenced to the 

energy use for the houses with fault-free installations

(Faults defined in Tables 5.31 and 5.32, Table 5.30 case d; worst level for both faults; the omitted dual 

faults involve duct leakage, which was not considered in houses with basement) 

5.4 Effects of Triple Faults

Triple faults were not simulated in this study because the open literature does not provide sufficient data 

on effects of multiple faults to allow for their characterization and use in annual simulations of 

building/heat pump systems. Nevertheless, the occurrence of three simultaneous faults is plausible,

particularly for the most common faults such as refrigerant undercharge, improper indoor airflow, or duct 

leakage. It is reasonable to assume that the effect of a triple fault will be as least as high as that of any of 

the possible three fault pairs considered individually; however, the effect of the third fault can increase the 

effect of the other two faults in an additive manner. As an example of a triple fault, SCE (2012) reported 

almost 70 % degradation in capacity for a split air conditioner operating under highly restricted airflow of 

the condenser, 40 % refrigerant undercharge, and 56 % reduction in the indoor airflow.

90

100

110

120

130

140

150

160

170

180

1 2 3 4 5 6 7 8 9 10 11 12 13 14

Relative energy use (%)

Las Vegas Wash. DC Chicago Minneapolis
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83

6. CONCLUDING REMARKS

Extensive simulations of house/heat pump systems in five climatic zones lead to the following 

conclusions:

o Effect of different installation faults on annual energy use is similar for a slab-on-grade house

(ducts located in the unconditioned attic) and a basement house (ducts located in the semiconditioned basement), except the duct leakage fault.

o Effect of different installation faults is similar in different climates except for the following cases:

 - Duct leakage: significant increase in the indoor RH for an installation in a hot & humid climate

- Heat pump oversizing with undersized air ducts: in heating-dominated climates, heat pump 

 oversizing reduces the use of backup heat, which compensates for the increased indoor 

 fan energy use associated with overcoming the higher external static pressure

o Undersized cooling mode TXV: little effect in heating-dominated climates, while a significant 

increase of energy use is possible in cooling-dominated climates.

The effect of simultaneous faults can be additive (e.g., duct leakage and non-condensable gases), little 

changed relative to the single fault condition (e.g., low indoor airflow and refrigerant undercharge), or 

well-beyond additive (duct leakage and refrigerant undercharge). 

The study found duct leakage, refrigerant undercharge, oversized heat pump with non-oversized ductwork, 

low indoor airflow due to undersized ductwork, and refrigerant overcharge to have the most potential for 

causing significant performance degradation and increased annual energy consumption. Increases of 

energy use by 30 % due to improper installation practices seem to be plausible. A well-designed and 

documented survey of heat pump installations would be helpful in establishing the prevalence of different 

installation faults and effective practices for their elimination. 

A significant increase in annual energy use can be caused by lowering the thermostat in the cooling mode 

to improve indoor comfort in cases of excessive indoor humidity levels. For Houston, TX, lowering the 

thermostat setting by 1.1 °C (2.0 °F) increased the annual energy use by 20 %, and the energy use increase 

rate is even higher due to further lowering the setting (the effect is not linear).

The authors contend that the laboratory and modeling results from this analysis using a 2.5 ton heat pump 

are representative of all unitary equipment including commercial split-systems and single package units 

(e.g., roof top units).

The goal of this study was to assess the impacts that HVAC system installation faults had on equipment 

electricity consumption. The effect of the installation faults on occupant comfort was not the main focus 

of the study, and this research did not seek to quantify any impacts on indoor air quality, or noise 

generation (e.g., airflow noise from air moving through restricted ducts). Additionally, the study does not 

address the effects that installation faults have on equipment reliability/robustness (number of starts/stops, 

etc.), maintainability (e.g., access issues), or costs of initial installation and ongoing maintenance.
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84

7. NOMENCLATURE

A = area [m2

, (ft2)]

ACH50 = air changes per hour at 50 pascal pressure differential

AF = improper indoor airflow rate fault

AHU = air handling unit

a = coefficient of multivariate polynomial

C = capacitance term: air mass in space multiplied by a multiplication factor in Eq. (4.1)

CD = heat pump cyclic degradation coefficient

CF = improper outdoor airflow rate (condenser fouling) fault

COP = coefficient of performance

cp = specific heat of air [J∙g-1

∙°C-1, (Btu∙lb-1∙°F-1)]

cfm = volumetric flow rate of air in I-P units (ft3

∙min-1)

DUCT = duct leakage fault

EER = energy efficiency ratio [Btu∙ h

-1

∙W-1]

FDD = fault detection and diagnosis

ELA = equivalent leakage area [m2

, (ft2)]

FSE = fit standard error; equal to the square root of the sum of the squared errors divided by 

 the degrees of freedom

F = fault level [% or dimensionless (fraction)]

FR = fraction of total return airflow (mR) from zone 2

FS = fraction of total supply airflow (mS) into zone 2

Gross capacity = total capacity (sensible and latent for evaporator) provided by the coil (does not 

 include indoor fan heat)

Gross COP = gross coil capacity divided by outdoor unit power. Outdoor unit power does not 

 include indoor fan power

HP = heat pump

HSPF = heating seasonal performance factor

HVAC = heating, ventilating, air conditioning

Htg = heating

hi = convective coefficient for exterior of duct [W∙m-2

∙°C-1, (Btu∙h-1·ft-2·°F-1)]

Latent capacity = portion of the cooling capacity that removes moisture (latent) energy (reduces the 

 moisture content (humidity ratio) of the air stream) 

LL = liquid line restriction fault

m = number of coefficients or mass flow rate [kg∙s-1

, (lb∙s

-1

); or kg∙h-1, (lb∙h-1)]

mR = return airflow to AHU [kg∙s-1, (lb∙s-1)]

m′R = airflow into return duct after accounting for leakage [kg∙s-1

, (lb∙s-1)],

 i.e., m′R = mR∙(1-FR)

mS = supply airflow from air-handling unit [kg∙s-1

, (lb∙s

-1

)]

N = number of data points

NC = presence of non-condensable gases fault

OC = refrigerant overcharge fault, % (or fraction) departure from the correct value

P = pressure [Pa, (mm H20)]

Q = capacity or heat loss or heat gain [W, (Btu∙h-1

)]

Qinternal = internal moisture gains [W, (Btu∙h-1

)]

QAC,latent = moisture removal by air conditioner [W, (Btu∙h-1)]

R = thermal resistance in I-P system of units [(h∙ft2∙°F∙Btu-1

)]

R(SI) = thermal resistance in SI system of units [K∙m2

∙W-1]

RH = relative humidity [%]
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85

SC = refrigerant subcooling at the liquid line service valve [°C, (°F)] or excessive 

 refrigerant subcooling fault, % (or fraction) departure from the correct value

SEER = seasonal energy efficiency ratio [(Btu∙W-1

∙h-1)]

Sensible capacity = portion of cooling capacity that removes sensible energy (decreases the temperature 

 of the air stream)

SHGC = solar heat gain coefficient

SHR = sensible heat ratio (sensible capacity divided by total capacity)

SIZ = heat pump sizing fault, % (or fraction) above or below the correct capacity

T = temperature [°C, (°F)]

TID = indoor dry-bulb temperature [°C, (°F)]

TIDP = indoor dew-point temperature [°C, (°F)]

TOD = outdoor dry-bulb temperature [°C, (°F)]

TMY3 = data set 3 with typical meteorological year weather data

TXV = thermostatic expansion valve or TXV undersizing fault in cooling

TACout = average temperature of air leaving AHU [°C]

Tdep = outdoor temperature at which a cooling mode TXV opens fully, as calculated 

 by Eq. (3.2) [(°C)]

t = time [s, (s)]

U = overall heat transfer coefficient [W∙m-2∙K-1, (Btu∙h-1∙ft-2∙F-1)]

UC = refrigerant undercharge fault, % (or fraction) departure from the correct value

V = volumetric flow rate [m3

∙h-1, (ft3∙min-1)]

VOL = electric line voltage fault

W = power [W, (W)]

WODU = power of outdoor unit; includes compressor, outdoor fan and control powers 

 [W, (W)]

Wtot = total power; includes WODU and indoor fan power [W, (W)]

w = humidity ratio [g∙g-1

, (lb∙lb-1)]

wACout = average humidity ratio of air leaving AHU [g∙g-1

, (lb∙lb-1)]

X = measured performance parameter

Y = normalized performance parameter

Greek Symbol

Δ = difference 

Subscripts

AR = air in the return duct

AS = air in the supply duct

i = indoor or feature index

in = incoming or inside

inf = infiltration

o = outdoor 

out = outcoming or outside

R = return duct or refrigerant

r = reduced

S = supply duct

sat = saturation

tot = total

z1 = zone 1

z2 = zone 2
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86

8. REFERENCES 

ACCA, 2009. ANSI/ACCA Standard 9 QIVP-2009, HVAC Quality Installation Verification Protocols. Air 

Conditioning Contractors of America, Arlington, VA., http://www.acca.org/

ACCA, 2010. ANSI/ACCA Standard 5 QI-2010, HVAC Quality Installation Specification. Air 

Conditioning Contractors of America, Arlington, VA., http://www.acca.org/quality

ACCA, 2011a. ANSI/ACCA 2 Manual J – 2011, Residential Load Calculation 8th Edition. Air 

Conditioning Contractors of America, Arlington, VA., http://www.acca.org/

ACCA, 2011b. ANSI/ACCA Standard 9 QIVP - 2011, HVAC Quality Installation Verification Protocols, 

Air Conditioning Contractors of America, Arlington, VA., http://www.acca.org/quality

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ACKNOWLEDGEMENT

This study was performed within Annex 36, Quality Installation/Quality Maintenance Sensitivity Study 

Analysis, of the International Energy Agency Heat Pump Program. The authors acknowledge Van Baxter 

of the Oak Ridge National Laboratory, Oak Ridge, TN, and Glenn Hourahan of the Air Conditioning 

Contractors of America, Arlington, VA, for organizing and managing the Annex. The authors also thank 

Glenn Hourahan for suggesting the scope of this study and for sharing his practical insights during 

different phases of the project, and Brian Dougherty of NIST for his expert review of the final manuscript.
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APPENDIX A: DUCT LOSSES

Duct losses − leakage and thermal − have been widely evaluated and studied in the field (Cummings and 

Tooley, 1989; Modera, 1989; Andrews, 1997; Siegel et al., 2003). The impacts of duct leakage and losses 

are especially significant in homes in the southern and western U.S. where ductwork is often installed 

outside the conditioned space (e.g., in the attic). Duct losses are complex phenomena where heat is lost to 

an unconditioned zone (typically the attic) and then in some cases ‘regained’ by reduced heat transfer 

between the conditioned and unconditioned zones (i.e., heat lost from attic ducts in the winter, tends to 

warm the attic and reduce heat loss through the ceiling). ASHRAE Standard 152 (ASHRAE 2004) has 

been developed to characterize the overall impact of thermal conduction and leak losses by determining 

the overall distribution efficiency (DE) for a system.

We used the leakage model developed for a prior TRNSYS-based simulation study of dehumidification 

systems (Henderson et al., 2007), as well as a study to evaluate the efficacy of a robust or ‘fault tolerant’

AC unit (Sachs et al., 2009). The model assumes all air leakage and conductions losses are from the 

ductwork to Zone 2 (the attic), as shown schematically in Figure A1. The following is the calculation 

scheme for the return duct and supply duct.

Attic (Zone 2)

House (Zone 1)

Supply 

air to 

Space

Return 

air from 

Space

Supply 

leak to 

attic

Return 

leak from

attic

Supply duct 

thermal losses 

Return duct 

thermal losses 

AHU

Figure A1. Schematic representation of duct leakage in a home with attic ducts

Return Duct

Air from the house zone (Zone 1) enters the return duct. According to evaluations of ASHRAE Standard 

152 by Francisco and Palmiter (2000), the temperature change of air in a duct that passes through an 

unconditioned space at a uniform temperate (To) is defined as:

UA m cp e

T T

T T ⋅ = −− − /

( )

( )

in o

out o (A1)

Applying Eq. (A1) to our case, the parameters of air arriving at the air handing unit (AHU) are given by:
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94



  − ⋅ ⋅ +

= + − ) 1 / ' p (

AR z2 z1 z2 i R R duct

( ) h A m c RT T T T e (A2)

wAR = wz1 (A3)

Then the air parameters at the end of the return duct after the thermal losses are:

TAR' = TAR·(1 − FR) + Tz2·FR (A4)

wAR' = wAR·(1 − FR) + wz2·FR (A5)

The heat gain to Zone 2 from thermal conduction is the same as the heat loss of the return air as it travels

through the duct, which is defined as 

QR = mR·(1 − FR) ·cp·(Tz1 – TAR) (A6)

Supply Duct

Supply air from the AHU unit (i.e., the average for the time step) enters the supply duct. The impact of 

thermal conduction losses are given by:

S S p duct /

AS z2 ACout z2 ( ) A m c R T T T T e

− ⋅ ⋅ = + − (A7)

wAS = wACout (A6)

A portion of the supply airflow goes to the space (zone 1), while the balance goes into the attic (zone 2)

To Space (Zone 1): mS-space = mS · (1 − FS) (A8)

To Zone 2: mS-z2 = mS ∙ FS (A9)

The heat gain to Zone 2 from thermal conduction is the same as the heat loss of the supply air as it travel 

through the duct, which is defined as 

QS = mS ·cp·(TACout – TAS) (A10)

Zone 2 has two impacts from the duct losses:

- supply air (airflow of mS-z2 at TAS and wAS) enters the zone to condition it,

- conduction losses from the return duct (QR) and the supply duct (QS) are added to the zone as a thermal 

gain.