Humidity Control in Humid Climates

Armin Rudd

Building Science Corporation

August 2013

NOTICE

This report was prepared as an account of work sponsored by an agency of the United States government. Neither the United States government nor any agency thereof, nor any of their employees, subcontractors, or affiliated partners makes any warranty, express or implied, or assumes any legal liability or responsibility for the accuracy, completeness, or usefulness of any information, apparatus, product, or process disclosed, or represents that its use would not infringe privately owned rights. Reference herein to any specific commercial product, process, or service by trade name, trademark, manufacturer, or otherwise does not necessarily constitute or imply its endorsement, recommendation, or favoring by the United States government or any agency thereof. The views and opinions of authors expressed herein do not necessarily state or reflect those of the United States government or any agency thereof.

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Task Order 3

Task 11: Additional Research Activities

Deliverable 11.2.3 Humidity Control in Humid Climates: Technical

Report

Prepared for:

The National Renewable Energy Laboratory

On behalf of the U.S. Department of Energy’s Building America Program

Office of Energy Efficiency and Renewable Energy

15013 Denver West Parkway

Golden, CO 80401

Prepared by:

Armin Rudd

Building Science Corporation

30 Forest Street

Somerville, MA 02143

NREL Technical Monitor: Cheryn Metzger

Prepared under Subcontract No. TASK ORDER NO. KNDJ-0-40337-03

August 2013

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Contents

List of Figures ............................................................................................................................................ vi List of Tables ............................................................................................................................................ viii Definitions .................................................................................................................................................... x Executive Summary .................................................................................................................................. xii 1  Problem Statement ............................................................................................................................. 14 

1.1 Introduction and Background ............................................................................................14 1.2 Research Questions ............................................................................................................15 1.3 Relevance to Building America’s Goals ............................................................................15 1.4 Tradeoffs and Other Benefits .............................................................................................15 1.5 Technical Approach ...........................................................................................................15 

2  Results ................................................................................................................................................. 15 2.1 Dehumidification Study in Houston, Texas .......................................................................15 

2.1.1  Research Approach ................................................................................................15 2.1.2  Results ....................................................................................................................16 

2.2 Enhanced Cooling System Study .......................................................................................18 2.3 Advanced Cooling with Dedicated Dehumidifier Mode System ......................................24 

2.3.1  Design Approach ...................................................................................................24 2.3.2  Prototype Construction ..........................................................................................26 2.3.3  Bench-top Testing Results .....................................................................................27 

2.4 TRYNSYS Computer Simulation Study ...........................................................................28 2.4.1  Climates .................................................................................................................28 2.4.2  Building and Enclosure Thermal Details ...............................................................29 2.4.3  Cooling System Details..........................................................................................32 2.4.4  Mechanical Ventilation Options ............................................................................35 2.4.5  Space Conditioning Systems ..................................................................................37 2.4.6  Electric and Natural Gas Costs ..............................................................................42 2.4.7  Simulation Results .................................................................................................45 2.4.8  Evaluating Humidity Levels ..................................................................................49 2.4.9  Location of Ducts in the Conditioned Space .........................................................54 2.4.10  Impact of Lower Duct Leakage and R-value .........................................................55 2.4.11  Impact of Ventilation Options ...............................................................................57 2.4.12  Impact of Ventilation Rate (50% and 150% of ASHRAE 62.2 Requirements) ....67 2.4.13  Single Speed, Two Speed, and Variable Speed Air Conditioners .........................71 2.4.14  Enhanced Air Conditioner Control Options ..........................................................73 2.4.15  Cooling Systems with Further Enhancements .......................................................75 2.4.16  Dehumidifiers ........................................................................................................79 2.4.17  Advanced Dehumidifiers .......................................................................................82 2.4.18  Comparing the Best Overall Technologies ............................................................85 2.4.19  Impact of House Size and Other Factors ...............................................................88 2.4.20  Conclusions ............................................................................................................94 

2.5 Evaluation of a Method to Process Indoor and Outdoor Temperature and Relative Humidity Data to Estimate Supplemental Dehumidification Energy ................................96 

2.6 Equipment Cost ................................................................................................................100 3  Conclusions ...................................................................................................................................... 101 Acknowledgements ................................................................................................................................ 103 References ............................................................................................................................................... 103 

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Appendix A.  Presentation of Additional Analysis of ASHRAE RP-1449 Data, Summarizing Supplemental Dehumidification Energy and Cost ........................................................................ 106 

List of Figures

Figure 1. Photo of Enhanced Cooling PSC System test house in Houston, TX ................................ 19 Figure 2. Photo of Enhanced Cooling ECM System test house in Houston, TX ................................ 20 Figure 3. Indoor temperature and relative humidity at the PSC enhanced cooling system test

house (PSC blower motor) ................................................................................................................ 21 Figure 4. Indoor environmental conditions and equipment operation for a representative 3-day

period in April 2004 at the PSC System house in Houston ............................................................ 22 Figure 5. Indoor temperature and relative humidity for the ECM enhanced cooling system test

house (ECM blower) ........................................................................................................................... 23 Figure 6. Indoor environmental conditions and equipment operation for a representative 3-day

period in April at the ECM test house in Houston ........................................................................... 23 Figure 7: Final design schematic showing cooling operation (greyed-out lines are inactive) ......... 25 Figure 8: Final design schematic showing dehumidification-only operation (greyed-out lines are

inactive) ............................................................................................................................................... 26 Figure 9: (photo left) 2.2 Advanced Cooling with Dedicated Dehumidifier Mode System as bench-

top tested. (photo right) Close-up of the add-on dehumidifier components fitted (from left to right: cased condenser reheat coil, 3-way diverting valve with suction bleed, receiver, and reversing valve (required only for heat pump heating operation) ................................................. 27 

Figure 10. IECC Climate Zone Map ......................................................................................................... 29 Figure 11. Schematic of Mechanical Ventilation System Options ...................................................... 36 Figure 12. Configuration of Natural Gas-Fired Desiccant Unit (pulling air from the cooling outlet;

supplying to supply duct) .................................................................................................................. 42 Figure 13. Schematic of Configurations for Dehumidifier and Desiccant Systems.......................... 44 Figure 14. Comparing Total Costs for Different Climates and HERS Levels ..................................... 47 Figure 15. Comparing Cooling and Fan Energy Use for Different Climates and HERS Levels ....... 47 Figure 16. Comparing Air Conditioner Runtime for Different Climates and HERS Levels ............... 48 Figure 17. Comparing Air Conditioner “Gross” EER for Different Climates and HERS Levels ....... 48 Figure 18. Psychrometric Chart Showing Space Conditions for HERS 100, Miami, System 1,

Exhaust Fan ........................................................................................................................................ 50 Figure 19. Shade Plot showing Humidity “bins” for each Hour of Year for HERS 100, Miami,

System 1, Exhaust Fan ...................................................................................................................... 50 Figure 20. Histogram of Relative Humidity for HERS 100, Miami, System 1, Exhaust Fan .............. 51 Figure 21. Comparing High Humidity Levels for Different Climates and HERS Levels .................... 52 Figure 22. Comparing Number of High Humidity Events (exceeding 60% RH) for Different Climates

and HERS Levels ................................................................................................................................ 53 Figure 23. Impact of Duct Location on Space Humidity Levels for HERS 50 and 70 Houses .......... 55 Figure 24. Comparing Total Ventilation Rates (mechanical & natural ACH) for Different Ventilation

Systems in Miami................................................................................................................................ 58 Figure 25. Comparing Total Ventilation Rates (mechanical & natural cfm) for Different Ventilation

Systems in Miami................................................................................................................................ 59 Figure 26. Comparing Total Electric Use for Ventilation Options with Different Systems ............... 64 Figure 27. Comparing Total Costs for Different Ventilation Options in each City ............................ 65 Figure 28. Comparing High Humidity Levels for Different Ventilation Options in each City ........... 66 Figure 29. Psychrometric Plots Showing Impact of Enhanced Control in Miami, HERS 100 House75 Figure 30. Psychrometric Charts Comparing the Degree of Humidity Control at Two RH Set Points

for Standard DH unit, HERS 100, Miami ........................................................................................... 79 Figure 31. Total Operating Costs for Standalone DH at Various RH Set Points, HERS 100 House 80 Figure 32. Total Operating Costs All DH Options, 50% RH Set Point, HERS 100 House ................. 83 Figure 33. Total Operating Costs All DH Options, 50% RH Set Point, HERS 70 House ................... 85 Figure 34. Total Operating Costs Best Technologies, HERS 100 House ........................................... 86 

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Figure 35. Total Operating Costs Best Technologies, HERS 70 House ............................................. 87 Figure 36. Shade Plots Showing Windows Openings Across the Year, HERS 100 House .............. 90 Figure 37. Psychrometric Plots Showing Impact of Operable Windows, HERS 100 House ............ 90 Figure 38. Attic Dew Point in Base Model (black) and with an Imposed “Dew Point Bump” (red) . 92 Figure 39. Impact of Moisture and Sensible Heat Gains - HERS100 with Exhaust Fan ................... 94 Figure 40. Indoor temperature and relative humidity, and outdoor dry bulb temperature and dew

point temperature ............................................................................................................................... 98 Figure 41. Net moisture load for hours above 60% relative humidity indoors for 8/25/2000 to

10/31/200 for Ft. Myers, FL house ..................................................................................................... 99 Figure 42. Indoor relative humidity and predicted supplemental dehumidification energy

consumption from 8/25/2000 to 10/31/2000 for Ft. Meyers, FL house ......................................... 100 Figure A 1. Supplemental dehumidification energy consumption and cost, along with hours above

60% relative humidity, for a HERS 50 house in Orlando with three different ventilation systems111 Figure A 2. Supplemental dehumidification energy consumption and cost, along with hours above

60% relative humidity, for a HERS 50 house in Miami with three different ventilation systems112 Figure A 3. Supplemental dehumidification energy consumption and cost, along with hours above

60% relative humidity, for a HERS 50 house in Houston with three different ventilation systems113 

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List of Tables

Table 1. Monitoring results for Stand-alone dehumidifier system and the Ducted dehumidifier system in the Houston study ............................................................................................................ 18 

Table 2. Climates Selected for Simulation ............................................................................................. 29 Table 3. Description of Performance Levels ......................................................................................... 30 Table 4. Enclosure Leakage and Duct Performance ............................................................................ 30 Table 5. Cooling Unit Sizing for each Climate and HERS Level .......................................................... 33 Table 6. Cooling Unit Characteristics .................................................................................................... 33 Table 7. Additional cooling Unit Characteristics for Two-Speed and Variable Speed Systems ...... 34 Table 8. Heating Cooling and Dehumidification Set Points ................................................................. 34 Table 9. Summary of Space Conditioning Systems ............................................................................. 38 Table 10. Matrix of cooling Conditioner Systems Used with each HERS Level and System ........... 39 Table 11. Electric and Natural Gas Costs .............................................................................................. 43 Table 12. Performance Results for Different HERS Levels and Climates .......................................... 46 Table 13. Comparing Relative Annual Costs for each HERS Level and Climate ............................... 46 Table 14. Total Space Conditioning Cost Reduction Impact of Moving Ducts from Attic to the

Conditioned Space ............................................................................................................................. 54 Table 15. Impact of Duct Location on Space Humidity Levels for HERS 50 and 70 Houses ........... 55 Table 16. Performance Results for Different Duct Leakage Rates in HERS 100 and HERS 130

Houses ................................................................................................................................................. 56 Table 17. Performance Results for Different Duct Insulation Levels in HERS 70 and HERS 85

Houses ................................................................................................................................................. 56 Table 18. Performance Results for Different Ventilation System and Climates (for HERS 100,

System 1: 13 SEER cooling) .............................................................................................................. 60 Table 19. Performance Results for Different Ventilation System and Climates (for HERS 85,

System 1: 14.5 SEER cooling, ECM Fan) ......................................................................................... 61 Table 20. Performance Results for Different Ventilation System and Climates (for HERS 100,

System 3: Two-Speed Cooling) ........................................................................................................ 62 Table 21. Performance Results with Different Ventilation Rates, HERS 100, Exhaust Fan ............. 67 Table 22. Performance Results with Different Ventilation Rates, HERS 100, CFIS ........................... 68 Table 23. Performance Results with Different Ventilation Rates, HERS 100, ERV ........................... 68 Table 24. Performance Results with Different Ventilation Rates, HERS 70, Exhaust Fan ............... 69 Table 25. Performance Results with Different Ventilation Rates, HERS 70, CFIS ............................. 69 Table 26. Performance Results with Different Ventilation Rates, HERS 70, ERV.............................. 70 Table 27. Performance Results for Different Cooling Units and Climates (for HERS 100 and HERS

50) ......................................................................................................................................................... 72 Table 28. Performance Results with Enhanced Cooling Unit .............................................................. 74 Table 29. Performance Results with Various Cooling Unit Enhancements ....................................... 75 Table 30. Performance Results for Cooling Units with Further Enhancements, HERS 100 ............ 77 Table 31. Performance Results for Cooling Units with Further Enhancements, HERS 70 (two-

speed) .................................................................................................................................................. 78 Table 32. Performance Results with Standalone DH (System 5) with Different Set Points ............. 80 Table 33. Humidity Threshold Results for Standalone DH (System 5) with Different Set Points .... 81 Table 34. Performance Results with Various DH Units (Systems, 5, 6, 7, 13 & 14) with HERS 100 83 Table 35. Performance Results with Various DH Units (Systems, 5, 6, 7, 13 & 14) with HERS 70 .. 84 and 50% RH Setpoint ................................................................................................................................ 84 Table 36. Performance Results with Best DH and Enhanced Units, HERS 100 ............................... 86 Table 37. Performance Results with Best DH and Enhanced Units, HERS 70 (two-speed) ............. 87 Table 38. Changes to House and Mechanical Systems ....................................................................... 88 Table 39. Rules for Changing House Characteristics with House Size .............................................. 88 Table 40. Impact of House Size on Humidity Levels and Energy Use ................................................ 89 Table 41. Impact of Window Openings on Humidity Levels and Energy Use – HERS 100, Exh Fan91 

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Table 42. Impact of Window Openings on Humidity Levels and Energy Use – HERS 130, No Ventilation ........................................................................................................................................... 91 

Table 43. Impact of Attic Dew Point “Bump” on Humidity Levels and Energy Use – HERS 100, Exh Fan ....................................................................................................................................................... 93 

Table 44. Impact of Attic Dew Point “Bump” on Humidity Levels and Energy Use – HERS 130, No Ventilation ........................................................................................................................................... 93 

Table 45. First-Cost Estimates for Supplemental Dehumidification Systems ................................. 101 Table A 1. Supplemental dehumidification energy and cost over a conventional cooling system,

for a HERS 50 house with different mechancial ventilation systems ......................................... 107 Table A 2. Supplemental dehumidification energy and cost over a conventional cooling system,

for a HERS 70 house with different mechanical ventilation systems ......................................... 108 Table A 3. Supplemental dehumidification energy and cost over a conventional cooling system,

for a HERS 85 house with different mechanical ventilation systems ......................................... 109 Table A 4. Supplemental dehumidification energy and cost over a conventional cooling system,

for a HERS 100 house with different mechanical ventilation systems ....................................... 110 

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Definitions

ach air changes per hour

ach50 air changes per hour at 50 Pascal pressure differential

AHAM Association of Home Appliance Manufacturers

AHRI Air Conditioning Heating and Refrigeration Institute

AHU Air handler unit

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

ASTM American Society for Testing And Materials

BA Building America Program. More information about BA can be found at www.buildingamerica.gov

BeOpt Building Energy Optimization Program – House energy simulation program and primary analysis tool for Building America homes.

BNL Brookhaven National Laboratory

BSC Building Science Corporation. More information about BSC can be found at www.buildingscience.com

CFI Central-fan-integrated

CFM 50 Cubic feet per minute at 50 Pascal test pressure differential

DOE U.S. Department of Energy

EPA U.S. Environmental Protection Agency

ERV Energy recovery ventilator

HRV Heat recovery ventilator

HUD U.S. Department of Housing and Urban Development

ICC International Code Council

IECC International Energy Conservation Code. More information can be found at http://www.energycodes.gov/

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IMC International Mechanical Code

IRC International Residential Code

NREL National Renewable Energy Laboratory

Pa Pascal; SI unit of pressure (equivalent to one newton per square meter)

RH relative humidity

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Executive Summary

In the last decade, building codes and market demands have been pushing residential buildings to be more energy efficient. Overall, this is good, and produces a significant net energy and cost savings. However, this situation has forced us to rethink the way we have traditionally thought about conventional residential space conditioning system design in humid climates. Most building efficiency improvements brought about by code requirements and above code incentive programs, such as more insulation, better windows, and low-power lighting and appliances, are directed at lowering sensible gains while latent gains remain mostly unchanged. Latent gains are related to internal moisture generation by occupants and their activities, and ventilation requirements -- which require exchanging conditioned but stale indoor air with unconditioned but fresher outdoor air. Because conventional cooling systems are directed to control to a temperature setpoint, cooling systems in these more efficient, low sensible gain houses, have longer off-times. During those longer off-times, which can be hours to days, indoor moisture can build up and cause elevated levels of indoor relative humidity (RH). Typically this occurs during spring/fall swing seasons, summer nights and rainy periods. Elevated RH impacts comfort, indoor air quality, and sometimes material durability if mold growth occurs. Therefore, at times when there is no need to lower the space air temperature but outdoor absolute humidity is still higher than indoors, supplemental dehumidification will be required to maintain the indoor relative humidity below acceptable levels. Similar conclusion was found in Rudd et al. 2003, Rudd et al. 2005, Rudd and Henderson 2007, BSC 2007, Rudd and Henderson et al. 2013. Maximum indoor RH thresholds vary depending on the criteria. For example, to control for dust mite allergen, a maximum of 50% RH is recommended. To control for comfort, at typical indoor cooling season temperatures, a maximum of 60% RH is often proposed. To avoid summertime mold conditions on surfaces cooler than the indoor air, RH above 70% should be avoided, but to avoid wintertime condensation and mold potential the threshold would be much lower, such as 40% RH or lower depending many building enclosure insulation and vapor control factors. The approach taken with this research report was to go back in time to look at the earliest research BSC conducted in humidity control, to the present time, to draw a current evaluation of what are the best mechanical systems for controlling indoor humidity in hot-humid climates and how much supplemental dehumidification energy is required? A number of BSC studies on supplemental dehumidification techniques in hot-humid climates were discussed here, ranging from large field and simulation studies, to product development and testing, to an example of a data processing approach to predict supplemental dehumidification requirements from indoor and outdoor temperature and relative humidity data alone. The most important conclusions from this study are as follows:

In a multi-home study in Houston, TX (Rudd et al. 2003; Rudd et al. 2005), measured supplemental dehumidification energy consumption from two mechanically ventilated and relatively airtight homes (about 3.5 ach50) was 209 kWh/yr for a representative home with a stand-alone dehumidifier and 463 kWh/yr for another representative home with a ducted dehumidifier. The ducted dehumidifier was more efficient, and the homes had similar temperature and relative humidity control, but variability in occupant behaviors has a strong impact on internal moisture generation which has a strong impact on supplemental dehumidification requirements. Internal moisture generation was not

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measured in that study, and is nearly impossible to measure except when under controlled simulation in a lab house environment.

Detailed simulations showed that a number of humidity control solutions can be effective in hot-humid climates. The most effective solutions, having relatively low operating cost and essentially eliminating indoor humidity above 60% RH, were: full condensing reheat integrated with the central cooling system, ducted dehumidifier, stand-alone dehumidifier with central system mixing, and condenser regenerated desiccant dehumidifier. About 170 kWh/yr could be expected for a HERS 50 house (having ducts inside conditioned space) with a 60% RH setpoint. About five times that could be expected with a 50% RH setpoint. A close second was central cooling system with subcooling reheat but it showed more elevated RH hours. A more distant third place was enhanced cooling with 2oF over-cooling and lower airflow (200 cfm/ton), if more hours above 60% RH and over-cooling discomfort can be tolerated, and note that it only works well to reduce elevated RH if a 50% RH setpoint is used. Two-speed and variable speed systems did little to reduce hours of elevated relative humidity in hot-humid climates unless coupled with the enhanced cooling methods listed above.

An Energy Recovery Ventilator (ERV) does little to reduce moisture loads when supplemental dehumidification is needed, which mostly occurs when there is little difference in absolute humidity between indoors and outdoors. With little absolute humidity to exchange between, ERVs have little impact on reducing elevated indoor relative humidity hours. In some hot-humid climates, including Orlando and Houston, energy recovery ventilation actually increases hours of elevated indoor humidity over exhaust and central-fan-integrated supply because the ERV sometimes keeps moisture in the house when drier outdoor air could reduce indoor humidity. If an ERV is operated in conjunction with supplemental dehumidification operated with a 50% RH setpoint, then the ERV does help reduce supplemental dehumidification energy consumption. That is because the dehumidifier forces a greater indoor to outdoor absolute humidity difference, allowing the ERV to reject some outdoor moisture with house exhaust air.

An analysis approach using only hourly indoor and outdoor measured temperature and relative humidity data, from a mechanically ventilated test house in Ft. Meyers, FL, showed predicted supplemental dehumidification energy consumption of 344 kWh/yr for a ducted dehumidifier (2.5 L/kWh). This compared reasonably well with 410 kWh/yr from detailed simulations of a similar mechanically ventilated HERS 100 house in Miami. This analysis approach should be investigated further. All that is required is hourly or sub-hourly indoor and outdoor measured temperature and RH data, and some basic house characteristics information, from houses in hot-humid climates without supplemental dehumidification. Using available data in this way may inexpensively continue to improve predictions of supplemental dehumidification requirements for high performance homes.

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1 Problem Statement

1.1 Introduction and Background This research project focused on indoor humidity control in humid climates. In the last decade, building codes and market demands have been pushing residential buildings to be more energy efficient. Overall, this is good, and produces a significant net energy and cost savings. However, this situation has forced us to rethink the way we have traditionally thought about conventional residential space conditioning system design in humid climates. Most building efficiency improvements brought about by code requirements and above code incentive programs, such as more insulation, better windows, and low-power lighting and appliances, are directed at lowering sensible gains while latent gains remain mostly unchanged. Latent gains are related to internal moisture generation by occupants and their activities, and ventilation requirements -- which require exchanging conditioned but stale indoor air with unconditioned but fresher outdoor air. Because conventional cooling systems are directed to control to a temperature setpoint, cooling systems in these more efficient, low sensible gain houses, have longer off-times. During those longer off-times, which can be hours to days, indoor moisture can build up and cause elevated levels of indoor relative humidity (RH). Typically this occurs during spring/fall swing seasons, summer nights and rainy periods. Elevated RH impacts comfort, indoor air quality, and sometimes material durability if mold growth occurs. Therefore, at times when there is no need to lower the space air temperature but outdoor absolute humidity is still higher than indoors, supplemental dehumidification will be required to maintain the indoor relative humidity below acceptable levels. Similar conclusion was found in Rudd et al. 2003, Rudd et al. 2005, Rudd and Henderson 2007, BSC 2007, Rudd and Henderson et al. 2013. Maximum indoor RH thresholds vary depending on the criteria. For example, to control for dust mite allergen, a maximum of 50% RH is recommended. To control for comfort, at typical indoor cooling season temperatures, a maximum of 60% RH is often proposed. To avoid summertime mold conditions on surfaces cooler than the indoor air, RH above 70% should be avoided, depending on many building enclosure insulation and vapor control factors. To avoid wintertime condensation on metal window frames and single-glazed windows, often resulting in mold on sills, the threshold would be 50% RH or lower.

Extensive field testing was done with builder partners in Texas and Florida in 2001 to 2007 (Rudd et al. 2003, Rudd 2004, Rudd et al. 2005, Rudd 2006, Rudd and Henderson 2007, Rudd 2007(b)). In part, that testing revealed that supplemental dehumidification was required in high performance Building America level homes (roughly 30% above code) in order to maintain indoor relative humidity below 60% year-round. Off-the-shelf, stand-alone supplemental dehumidification systems were employed to address this problem while working with manufacturing partners on supplemental dehumidification integrated with the central space conditioning system. These companies began to offer integrated supplemental dehumidification solutions that allow year-round indoor relative humidity control between 50% and 60%. The supplemental dehumidification solution is intended to enable further reduction in sensible cooling load, through further efficiency improvements, without the risk of elevated indoor humidity.

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While these advancements have been important and needed in the residential space conditioning industry, supplemental dehumidification technology continues to improve and evolve, and the market for these products is still in its infancy. Design capacity prediction is subject to many unknowns and requires continued research to fully quantify.

1.2 Research Questions The research presented in this report is intended to help develop a better understanding of whole-building dehumidification to improve comfort, durability, and indoor air quality in low energy homes. BSC seeks to address the research questions, “What are the best mechanical means to accomplish indoor humidity control in high performance homes in hot-humid climates and how much supplemental dehumidification is needed?”

1.3 Relevance to Building America’s Goals Overall, the goal of the U.S. Department of Energy's (DOE) Building America program is to “reduce home energy use by 30%-50% (compared to 2009 energy codes for new homes and pre-retrofit energy use for existing homes).” To this end, we conduct research to “develop market-ready energy solutions that improve efficiency of new and existing homes in each U.S. climate zone, while increasing comfort, safety, and durability.”1 As described above, in order to accomplish these goals, latent load (moisture) control will need to be elevated to that commensurate with sensible cooling load (thermal) control to maintain acceptable comfort, indoor air quality, and material durability.

1.4 Tradeoffs and Other Benefits Achieving deep energy savings by deep reduction in sensible cooling load requires coincident latent load control. Otherwise, risk of elevated indoor relative humidity will result. As design and application of latent load control comes into focus for builders, they will find that they can market added value in that control of indoor moisture levels will be assured year-round just as temperature control is assured year-round.

1.5 Technical Approach The approach taken with this research report was to go back in time to look the earliest research BSC conducted in humidity control, to the present time, and evaluate what are the best mechanical systems for controlling indoor humidity in hot-humid climates, and how much supplemental dehumidification energy is required, and what is the expected operating cost?

2 Results

2.1 Dehumidification Study in Houston, Texas 2.1.1 Research Approach In 2001 to 2002, a study was conducted in cooperation with Pulte Home Corporation in Houston, Texas. Six different humidity control systems were evaluated in homes that were at least 30% better than Model Energy Code 1995. The homes were constructed with unvented attics (insulation under the roof sheathing, therefore the attics were semi-conditioned). Three reference

1 http://www1.eere.energy.gov/buildings/building_america/program_goals.html

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houses had the same energy efficiency measures and controlled mechanical ventilation, but no dehumidification separate from cooling. Three other reference houses met code minimums for energy efficiency and did not have mechanical ventilation or dehumidification separate from cooling, and had conventional vented attics.

Two of the six humidity control strategies studied were found to be most successful and are discussed here.

System 1: Stand-alone dehumidifier

The stand-alone dehumidifier system was an off-the-shelf 50 pint/day rated dehumidifier installed in an interior closet with a louvered door near the central air return. The dehumidistat built into the dehumidifier energized the dehumidifier whenever the humidity level rose above the user setting. A fan cycling control wired to the central air distribution fan was set to 33% duty cycle (on for 10 minutes if it had not been on for 20 minutes), to intermittently average temperature and humidity conditions throughout the house and distribute ventilation air.

System 2: Ducted dehumidifier

The ducted dehumidifier system was a 90 pint/day rated high-efficiency ventilating dehumidifier located in the unvented attic. The dehumidifier blower operated continuously on low speed, drawing in about 40 cfm of outside air and about 120 cfm of recirculated house air. The mixed air was filtered by the dehumidifier unit and supplied to the main supply air trunk of the central air distribution system. A remote dehumidistat located in the living space activated the dehumidifier compressor if the humidity level in that space rose above the user setting. A fan cycling control wired to the central air distribution fan was set to 17% duty cycle (on for 10 minutes if it had not been on for 50 minutes) to average temperature and humidity conditions throughout the house and distribute ventilation air. 2.1.2 Results Standard reference houses Monitoring data from all three Standard Reference Houses was analyzed to quantify the humidity control performance of homes that just met code requirements for energy-efficiency, and had no whole-house mechanical ventilation system or dehumidification separate from the central cooling system. While the cooling system runtimes were often predictably short due to cooling system over-sizing, there was little correlation between cooling system short-cycling and uncomfortably high relative humidity. Humidity control performance was good in these houses, but cooling energy consumption was high relative to the energy-efficient reference houses. Energy-efficient reference houses For the mechanically ventilated energy-efficient reference houses, a stronger relationship between elevated indoor relative humidity and outdoor dew point temperature was observed compared to the Standard Reference houses. This indicates that the energy-efficient houses were more affected by outdoor air exchange. With similar indoor dry bulb temperatures, increased

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outdoor air exchange added moisture and increased indoor relative humidity. Other dominant factors in the elevated relative humidity conditions found in these homes were:

1. Lower sensible heat gain causing the cooling systems to operate less often (longer off times). Indoor relative humidity was generally higher with low cooling system on-time fraction.

2. Interior moisture generation with little source control by local exhaust fan usage.

System 1 - Stand-alone dehumidifier in hall closet The stand-alone dehumidifier in an interior hall closet system had the lowest initial cost and operating cost while providing reasonably good humidity control. During the 300 day monitoring period starting October 2001, 7% of the hours were above 60% RH. During that same period, the dehumidifier consumed 209 kWh of electricity for supplemental dehumidification. The system required the loss of a lower shelf in the hall closet, and some occupants may be sensitive to the new noise. System 2 – Ducted dehumidifier

The ventilating ducted dehumidifier system showed good humidity control but had high first cost and high operating cost due to the continuously operating ventilation fan. During the 288 day monitoring period starting October 2001, 4% of the hours were above 60% RH. During that same period, the dehumidifier consumed 463 kWh of electricity for supplemental dehumidification.

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Table 1. Monitoring results for Stand-alone dehumidifier system and the Ducted dehumidifier system in the Houston study

Table 1 provides a summary of the monitored data for the two dehumidification systems. The indoor environment conditions were quite similar for both houses, yet the wide range of energy consumption for supplemental dehumidification illustrates the sensitivity to occupancy and occupant behaviors causing different internal moisture generation rates in homes. The fact that both of the houses showed hours above 60% RH was not because of too little dehumidification capacity but because of the dehumidifier setpoints being near 60% RH and the fairly wide control deadband associated with the dehumidifier controls.

2.2 Enhanced Cooling System Study

In 2003 and 2004 BSC began to work with builder partner David Weekley Homes and manufacturing partner Carrier Corporation to evaluate system-integrated dehumidification for humidity control without over-cooling the space. This was determined to be the most significant climate-specific need. The following testing regime was established for two new Carrier advanced HVAC systems. Testing, and one year of monitoring of the houses and systems was as follows:

Each house was tested for building air leakage and duct leakage. The standard for building leakage was: building leakage less than or equal to 0.35 cfm per square foot of building surface area at 50 Pa pressure differential (Environments for Living “Gold”

Stand-aloneDehumidifier

DuctedDehumidifier

Monitoring period 10/4/2001 - 7/31/2002 10/17/2001 - 8/1/2002Total days with data 300 288House dry bulb temperature(avg of 3 points)

min 67 63max 80 80avg 73 72

std dev 2 3House relative humidity(avg of 3 points)

min 26 25max 70 69avg 50 51

std dev 8 7House dew point temperature(avg of 3 points)

min 35 32max 64 66avg 53 53

std dev 5 6Hours above 60% RH 7% 4%Supplemental dehumidificationenergy consumption (kWh) 209 463

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level). The standard for duct leakage was: duct leakage to outside less than or equal to 5% of high speed air handler flow.

Each house was tested to assure that under normal air handler operation, that neither the main living area nor closed rooms were more than plus or minus 3 Pascal with respect to outside. That assures free air movement throughout the house without mechanically inducing air exchange by infiltration.

Temperature and relative humidity (RH) were monitored in 3 areas in the house and one location in the attic where the mechanical equipment and ducts were located. The 3 house locations were the thermostat, master bedroom, another occupied bedroom or another main body location.

The mechanical equipment and process air conditions were monitored as follows: hourly min and max supply air temperature leaving the evaporator coil; hourly min and max return air temperature entering the air handler unit; hourly average outside air flow; hourly cooling runtime, number of cycles, and kW-h consumption hourly heating runtime hourly fan runtime, number of cycles, and kW-h consumption

Two houses in Houston were outfitted with different Carrier cooling systems, each with enhanced dehumidification capacity. The two systems were installed and commissioned on September 24 and 25, 2003. Each of the homes was a single-story plan of about 2500 ft2, with similar solar exposure and occupancy. PSC System - 15251 Henderson Point: This system utilized a Carrier 58DLA furnace with a standard 4-speed PSC blower motor controlled by a Thermidistat-2. A dehumidify speed change was built into the 58DLA furnace logic board. The Thermidistat-2 control had three cooling/dehumidification modes:

The first mode was “normal cooling” where the blower will run at high speed to achieve between 350 to 400 cfm/ton2. This normal cooling mode achieves dehumidification according to the standard equipment design.

The second mode was “cool to dehumidify” where the blower ran at reduced speed (about 280 cfm/ton) only when there was a call for cooling. This mode could operate continuously.

The third mode was “super dehumidify” where, in addition to the “cool to dehumidify mode,” if the actual relative humidity remained higher than the humidity setpoint, then cooling was cycled intermittently for 10 minutes ON and ten minutes OFF at further reduced blower speed (as low as 210 cfm/ton), without a call for cooling, but limited to not cool below

2 According to Carrier’s typical practice, System 1, with the PSC AHU, was ARI rated at 400 cfm/ton to anticipate lower installed air flow due to ducting flow resistance, while System 2 was ARI rated at 350 cfm/ton because the torque-sensing capability of the ECM control reliably gives the desired flow rate.

Figure 1. Photo of Enhanced Cooling PSC System test

house in Houston, TX

20

the requested setpoint by more than 3oF. This arrangement allows the evaporator coil to run very cold, in the range of 38oF. Cooling coil freeze protection results from proper set up of airflow and the intermittent operation, but an optional freeze-stat which monitors coil temperature can be installed. TXV refrigerant control is mandatory to assure required superheat at the compressor.

An R-22 (38BRC*) outdoor condensing unit was installed, matched to indoor coil size with ARI rating.

Central-fan-integrated supply ventilation was installed, comprising an outside air duct to the main return air duct with a manual balancing damper for flow adjustment and a motorized damper to limit over-ventilation. An Aprilaire model 8120 controller periodically operated the blower if it has not operated enough for the ventilation design (50 cfm, 33% duty cycle) and controlled the motorized damper for limiting over-ventilation during long cooling or heating cycles. ECM System - 7511 Golden Thistle Drive: This system utilized a Carrier 58CVA furnace with an ECM air handler, controlled by a Thermidistat-2. Depending on the indoor relative humidity measured by the control, the combination of these units will allow the cooling to dehumidify operation as described above, with the exception that the ECM blower speed control allowed lower fan speeds/airflow than the relay arrangement used for the PSC motor system. This system had the same mechanical ventilation system as System 1. Referring to Figure 3, data collection for the PSC System house started in September 2003, however, since prior data had been taken at that house, going back to June 2003, the additional data is also shown for comparative purposes. Indoor relative humidity remained reasonably well below 60% RH until spring of 2004, when excursions above 60% RH at the thermostat location, and above 70% RH in the master bedroom, occurred frequently and for multiple hours at a time. The wide temperature swing within the space was a result of the occupants using a programmable thermostat to set up the temperature while they were away at work, and to set back the temperature while at home, especially at night.

Figure 2. Photo of Enhanced Cooling ECM System test

house in Houston, TX

21

Figure 3. Indoor temperature and relative humidity at the PSC enhanced cooling system test

house (PSC blower motor)

The representative period shown in Figure 4 shows the indoor environmental conditions and space conditioning equipment operation at the PSC enhanced cooling system house. Sustained periods of 10 to 18 hours with relative humidity between 60% and 70% occurred. Use of the programmable thermostat was clear, which seemed to limit the cooling systems ability to energize cooling for dehumidification according to the 3oF over-cooling limitation. Central system fan operation was controlled to a minimum 33% duty cycle for distributing ventilation air and overall whole-house mixing and filtration to improve indoor air quality. Cooling cycles (activations) per hour ranged up to 5 which showed that the length of cooling cycle was often less than 5 minutes (with a minimum 5 minutes off between cycles for compressor protection), which was too short for achieving much moisture removal.

PSC, Enhanced cooling system house, Houston, TX

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Figure 4. Indoor environmental conditions and equipment operation for a representative 3-day

period in April 2004 at the PSC System house in Houston

Similar to the PSC enhanced cooling system house, the ECM enhanced cooling system test house also encountered humidity control problems by spring of 2004, as shown in Figure 5, but in this case, the problems were more pronounced. Relative humidity excursions between 60% and 80% occurred frequently and for long duration (days). As it turned out for this house, an exacerbating factor was that, although they had been initially informed, the occupants inadvertently changed the fan mode from AUTO to ON. That condition existed for much of the time between March and June 2004 between site visits for data retrieval. The constant fan operation (Fan ON) defeated the intended humidity control enhancement due to water evaporation from the indoor coil each time the compressor turned off.

PSC, Enhanced cooling system house, Houston, TX

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Figure 5. Indoor temperature and relative humidity for the ECM enhanced cooling system test

house (ECM blower)

Figure 6. Indoor environmental conditions and equipment operation for a representative 3-day

period in April at the ECM test house in Houston

ECM, Enhanced cooling system house, Houston, TX

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The data detailed in Figure 6 gives a representative snapshot illustrating the humidity control inadequacy of the ECM enhanced cooling system. During this 3-day period of operation, with the ECM System set for “super dehumidify” at 50% RH, the indoor relative humidity was consistently over 70% most of the time. The minimum temperature of refrigerant leaving the evaporator coil for each hour of cooling system operation ranged between 42 and 49oF, indicating that the system was operating well enough to achieve a cold coil, but it could not operate long enough to adequately reduce the indoor humidity level without over-cooling the space more than 3oF from the thermostat setpoint. Cooling cycles per hour ranged from 1 to 5, also indicating that the cooling system did not operate long enough for optimal moisture removal. Even with the thermostat fan switch unfortunately set to ON by the occupant, the fan runtime was never a full 60 minutes per hour because the system control turned the fan off for 5 minutes after each cooling cycle if the dehumidify mode was activated and the relative humidity reading was higher than the setpoint. A final recommendation made to the manufacturer was to disable the thermostat Fan-ON capability if the controls were set in any of the dehumidify modes.

2.3 Advanced Cooling with Dedicated Dehumidifier Mode System Available enhanced cooling systems were found to be inadequate to reliably control indoor humidity throughout the year. So, in 2005, BSC developed and bench-top tested a new type of central air conditioning system, one that could do conventional cooling and but also serve as a special type of dehumidifier -- delivering dry but room neutral temperature air. This system accomplished supplemental dehumidification using the same refrigeration equipment that also provided central cooling (or central cooling and heating in the case of a heat pump unit). This system was designed to eliminate the need for a separate dehumidifier, that also had the potential disadvantage of delivering air warmer than room temperature, and it was designed to eliminate the need to overcool to dehumidify using a conventional central cooling system. The result was a dual-mode system providing a standard cooling mode with incidental dehumidification and dehumidification only mode. The ultimate goal was a single-compressor refrigeration system that could provide the same temperature control as a conventional system but also control indoor humidity throughout the year, to at least below 60% RH, for lower first cost and operating cost compared to a conventional central cooling system plus a separate dehumidifier integrated with the central system. 2.3.1 Design Approach The design approach included modification of a conventional cooling system to include an additional refrigerant condensing coil (reheat coil) in the process air stream after the evaporator coil. The control strategy was that when a temperature set point was satisfied, but a humidity set point was not satisfied, the dedicated dehumidification mode would be activated. The design schematic of Figure 7 shows the system operating in cooling mode (which also provides some incidental dehumidification), and the design schematic of Figure 8 shows the

25

system operating in dehumidification-only mode. These schematics reflect the as-built prototype configurations.

Figure 7: Final design schematic showing cooling operation (greyed-out lines are inactive)

26

Figure 8: Final design schematic showing dehumidification-only operation (greyed-out lines are

inactive)

2.3.2 Prototype Construction A nominal 2.0 ton Goodman heat pump split air conditioning system was purchased as the base, off-the-shelf system. It was rated at 14 SEER with a variable speed (ECM) indoor fan. The outdoor unit had a scroll compressor, an accumulator, and low pressure controls to aid in low ambient temperature operation. Normal installation called for a 3/4 inch vapor line and a 3/8 inch liquid line between the outdoor and indoor units. However, a 5/8 inch line was substituted for the 3/8 inch line in anticipation that the larger size would be needed when the line would carry a mixture of gas and liquid while in dehumidifier mode. A relay was added to the outdoor unit wiring to allow the condenser fan to be turned off during dehumidification mode while the compressor continued to operate.

27

Figure 9: (photo left) 2.2 Advanced Cooling with Dedicated Dehumidifier Mode System as bench-top tested. (photo right) Close-up of the add-on dehumidifier components fitted (from left to right:

cased condenser reheat coil, 3-way diverting valve with suction bleed, receiver, and reversing valve (required only for heat pump heating operation)

2.3.3 Bench-top Testing Results All of the indoor and outdoor operating conditions during the available period of bench-top testing were on the low end of normal for this type of equipment (50oF - 65oF). In standard cooling mode and 50 oF outdoors, the measured EER was 17 for the SEER 14 rated unit, with a sensible heat ratio (SHR) of 0.73 and a moisture removal efficacy of 2.06 L/kW-h. Although testing was not done at the AHAM dehumidifier standard rating conditions of 80oF and 60% RH, moisture removal efficacy was listed here, in liters of water removed per kilowatt-hour of electrical energy consumed, for rough comparison sake. The EPA Energy Star requirement for high capacity dehumidifiers was 2.25 L/kW-h. While in dehumidifier-only mode, moisture removal was about 1.5 L/kWh. The outdoor condenser fan was de-energized to allow full condensing at the indoor condenser/reheat coil. Turning off the outdoor condenser fan reduced the amount of outdoor condensing, but it did not eliminate it. To try to simulate higher ambient temperatures in two of the tests, a thick insulating blanket was laid over the outdoor condenser to cause the condenser to heat up. This raised the supply air temperature to about 6oF above the return air temperature, whereas the supply air temperature had been close to the same as the return air temperature (room neutral temperature) without the blanket. In this configuration, the increase in system pressure and compressor power reduced the moisture removal efficacy to about 1.0 L/kWh. The refrigerant coil manufacturer’s predicted air pressure drop across the indoor reheat coil was 0.03 inch w.c. or 7.5 Pa at 800 cfm. This proved very realistic in testing, which showed an air pressure drop of 6 Pa. That airflow resistance was small and inconsequential, especially when compared to the pressure drop across a wet evaporator coil normally in the range of 40 to 50 Pa. Based on prior measurements from cooling systems and dehumidifiers, the initial prediction for

28

refrigerant condensing temperature was 125oF. That turned out to be too high because of heat loss from the outdoor condenser even with the condenser fan off. The actual refrigerant condensing temperature was about 100oF which reduced the expected capacity of the reheat coil. A reheat coil with smaller tubing and closer fin spacing, to increase reheat capacity while maintaining acceptable air pressure drop, would have been an improvement. In 2007, working with manufacturing partner AAON, Inc. a full condensing/subcooling version of this partial condensing/subcooling design was commercialized and made available for sale. It differs in design from the designs shown in Figure 7 and Figure 8 in that a third refrigerant line is used to directly move hot gas refrigerant between the compressor and the indoor condenser reheat coil. The system has been evaluated at homes in Florida, Louisiana, and Texas. The Florida and Louisiana sites used 2-speed scroll compressor systems, while the Texas site used a digital scroll compressor system. Some data taken from those evaluations was analyzed and used to produce performance mapping for the simulation study described in the next section. 2.4 TRYNSYS Computer Simulation Study This study (Rudd and Henderson et al. 2013) was based in ASHRAE Research Project 1449 “Energy Efficiency and Cost Assessment of Humidity Control Options for Residential Buildings” and leveraged by this project. It used TRNSYS as the basis for building and mechanical system simulations. Specifically TRN-ResDH3 (Henderson and Sand 2003; Henderson et al 2007), was improved to meet the needs of this project. TRNSYS-based hour-by-hour building energy simulation tool was used to simulate the HVAC technologies being investigated as part of this project. TRN-ResDH includes component models for various dehumidification systems of interest in this study, including conventional standalone room air dehumidifiers, high-efficiency mechanical dehumidifiers, subcool/condenser reheat systems, and desiccant dehumidification equipment. It also includes robust models for conventional cooling components that accurately predict performance at part-load conditions. The impact of moisture capacitance in building materials and furnishings, moisture evaporation from the cooling coil when the compressor is off, and other impacts of fan performance are all considered in this simulation tool. For this study a short-time-step version was developed (with a time step of 0.02 hour, or 72 seconds) in order to properly consider all of the system control and performance interactions of multiple, simultaneously-operating machines. The proper consideration of these interactions is critical to predicting energy consumption and resulting indoor humidity conditions. 2.4.1 Climates The cities shown below, representing each of the IECC Climate Zones 1A through 5A shown in Figure 10, were chosen to represent locations that have, at least for part of the year, substantial latent loads, and cover a temperature range that includes both conditions with little coincident sensible load and significant coincident sensible load. TMY3 weather data is used. Orlando was not originally included, but was added in the middle of the study.

3 A publicly-available version of TRN-ResDH is available at http://www.cdhenergy.com/trn-resdh2/. An updated version and documentation will be available to the PMS at the conclusion of this project.

29

Table 2. Climates Selected for Simulation

Zone ID IECC Climate Zone

City

Z0 2A Orlando Z1 1A Miami Z2 2A Houston Z3 3A Atlanta Z4 4A Nashville Z5 5A Indianapolis

Figure 10. IECC Climate Zone Map

2.4.2 Building and Enclosure Thermal Details The 2,000 ft2 3 bedroom house was modeled as slab-on-grade with a separate attic zone (a 2-zone model in TRNSYS Type 56). A range of building enclosures were simulated corresponding to Home Energy Rating System (HERS) levels (RESNET 2013) that were selected to correspond to common industry benchmarks (see Table 3). The combination of building and mechanical system characteristics were chosen to reflect typical practice at each level. Therefore, higher efficiency cooling systems were specified at lower HERS Indices. Appendix B of Rudd and Henderson et al. (2013) provides the thermal enclosure details and mechanical system efficiency level for each house.

30

Table 3. Description of Performance Levels

Description or Benchmark HERS Index Cooling Efficiency Typical Existing House 130 10 SEER HERS Reference House 100 13 SEER Energy Star House 85 14.5 SEER (BPM fan) Builders Challenge House 70 17.7 SEER (two-speed) Building America Prototype 50 17.7 SEER (two-speed) Building Enclosure Air Leakage The AIM-2 infiltration model (Walker and Wilson 1998, ASHRAE 2009) relates infiltration to wind and indoor-outdoor temperature difference for each timestep. All simulations in this study used coefficients representing shelter from buildings across the street. The equivalent leakage area (ELA) was varied to provide the desired ACH50 at each HERS level, as shown in Table 4. The attic uses the same AIM-2 equations to determine leakage as a function of wind and temperature difference. The attic ELA was set to be 567 in2 for all the HERS levels, or about 5 times the leakage rate for the HERS 100 house (Fugler 1999).

Table 4. Enclosure Leakage and Duct Performance

HERS Level

Target ACH at 50 Pa

ELA at 4 Pa (in2)

Duct Leakage, Supply and

Return Combined (% of

flow)

Duct Insulation R-Value

(h-F-ft2/Btu)

Sup / Ret Duct Area

(ft2)

130 10 140.1 20% 6 544 / 100 100 7 98.1 10% 6 544 / 100 85 5 70.1 5% 8 544 / 100 70 4 56.1 5% 8 544 / 100 50 3 42 None Na Na

Note: Duct leaks are assumed to be 60% on supply side, 40% on return side.

31

Duct Leakage and Thermal Losses Table 4 also lists the characteristics of the duct system. For the HERS 130 to 70 Houses, the ducts were assumed to be located in the attic space and all the air leakage and thermal losses/gains went into that zone. For the HERS 50 house the ducts were assumed to be in the conditioned space so leakage and thermal conduction had no impact. The duct leakage rates are assumed to be 60% on the supply side and 40% on the return side for all the cases. Duct insulation is assumed to be R-6 for the HERS 130, 100 houses and R-8 for the HERS 85, 70 houses. The duct R-value was not given for the HERS 50 house since all ducts were inside conditioned space. 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 are assumed to be:

72.70 MBtu/day (21.3 kWh/day) for HERS 130 and 100 65.43 MBtu/day (19.2 kWh/day) for HERS.85 58.16 MBtu/day (17 kWh/day) for HERS 70 50.89 MBtu/day (14.9 kWh/day) for HERS 50

Internal moisture generation from all sources was specified as 12 lb/day, or less than half of the ASHRAE Standard 160 moisture generation rate of 31.2 lb/day (1.3 lb/h) for a 3 bedroom house. The ASHRAE 160 value is meant to be a “worst case” design condition and therefore would not be expected to correspond to average conditions. This value of 12 lb/day was selected based on a calibration effort where we compared the model to measured data (see Appendix C and Appendix J of Rudd and Henderson et al. (2013)). Base on that analysis we selected key values that resulted in humidity distributions similar to those observed from monitored homes. 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 required capacity of dehumidification equipment along with the diurnal swings in indoor humidity. Building material moisture storage was modeled as a simple lumped parameter method with mass factor added to the air node in the zone model:

latentACernaloii QQwwm

dt

dwC ,int)(

where: w - humidity ratio (lb/lb), i=indoor, o=outdoor m - net indoor-to-outdoor airflow rate (lb/h) Qinternal - Internal moisture gains (lb/h) QAC, latent - moisture removal by air conditioner or dehumidifier (lb/h) C - a capacitance term: air mass in space x multiplication factor

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The moisture capacitance term is usually set to a multiple of the air mass inside the house. Shirey et al (2001) 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 mass4. As a result of the calibration efforts described in Appendix C of Rudd and Henderson et al. (2013), a 30x multiplier for moisture capacitance was used for the main zone. The attic used a moisture capacitance factor of 15x. Thermal capacitance was simulated by adding internal walls to the model with 4,000 ft2 (371.6 m2) of exposed wall surface area. The thermal mass of the air node was also increased by 20x to 12,331.2 kJ/K to reflect the impact of furniture and other material in the space. The attic is assumed to have a thermal capacitance multiplier of 1x. 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 discussed in residential practice and building codes. The LBNL WINDOW inputs for this project were determined following the methodology developed by Arasteh et al. (2009) for use in EnergyPlus. The suitability of this model to TRNSYS was checked by comparing the fraction of transmitted solar radiation as calculated by TRNSYS to the nominal SHGC. The actual solar gains for the windows for this south facing wall were typically within -21% to +14% of the expected performance based on the nominal SHGC. The average difference for all the windows was zero. Attic and Roof Heat Transfer The house for this study has been constructed with a two-zone model: 1) the main zone and 2) the unconditioned, vented attic zone. The two zones are connected via an insulated ceiling. For some cases air conditioner ducts are located in the attic and interact with that zone. Return air leaks pull air from the attic zone. Supply air leaks and duct conduction losses tend to cool and dehumidify the attic zone. The roof absorbance was set to 0.9 consistent with dark asphalt shingles. The convection coefficient from the exterior roof surface was adjusted to bring the peak temperature in the attic close to values observed in real homes. 32 kJ/h-m2-K was selected to be most consistent with the expected attic temperatures. Similarly, Lixing Gu, a simulation expert and researcher at FSEC, reported that he typically uses 10 W-m2-K (36 kJ-h-m2-K) for the exterior heat transfer coefficient for modeling Florida homes. 2.4.3 Cooling System Details The study considered both conventional cooling systems as well as other systems that provided improved humidity control. The size of the conventional cooling unit was determined for each

4 Reference Manual for IHAT simulation tool

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climate and HERS level. This resulting manual J-sizing is given in Table 5. The smallest available unit was assumed to be 2 tons. This sizing was used for all systems.

Table 5. Cooling Unit Sizing for each Climate and HERS Level

Cooling Size per Manual J (tons) HERS

130 HERS

100 HERS

85 HERS

70 HERS

50 0 – Orlando 3 3 2.5 2 2 1 – Miami 3 3 2.5 2 2 2 – Houston 3 3 2.5 2 2 3 – Atlanta 2.5 2.5 2 2 2 4 – Nashville 2.5 2.5 2 2 2 5 – Indianapolis 2.5 2.5 2 2 2 Note: Shading indicates where the sizing was actually lower than assumed minimum of 2 tons A range or air conditioners were required for this study. This detailed air conditioner model required separate inputs for the gross EER at nominal conditions, sensible heat ratio (SHR), and fan power. These 3 values are required to determine the resulting SEER for each cooling system. These values are given in Table 6. The fan power assumed for rated conditions – and used to calculate SEER – is listed along with the actual fan power assumed for operation. For instance for the SEER 13 unit the fan power at rated conditions was assumed to be 0.25 W/cfm, while the actual fan power was 0.5 W/cfm. Two-speed and variable-speed systems require additional parameters at low speed to define their performance. Table 7 lists these additional low speed parameters. Note that the low stage fan power (Watt/cfm) was also used for reduced airflow on single speed systems (i.e., System 2).

Table 6. Cooling Unit Characteristics

Note: Gross EER is total coil cooling capacity divided by compressor and condenser fan power at the nominal rating point: 95°F outdoors, 80°F/67°F entering coil and 450 cfm per ton supply airflow.

Description

Gross EER

(Btu/Wh) SHR (-)

Actual Fan

Power (W/cfm)

Actual Airflow

(cfm/ton)

Rated Fan

Power (W/cfm)

Rated SEER

(Btu/Wh)

Rated EER

(Btu/Wh)SEER 10 Unit (Single Spd, PSC fan mtr) 10.35 0.77 0.50 375 0.25 10 7.8SEER 13 Unit (Single Spd, PSC fan mtr) 13.80 0.77 0.50 375 0.25 13 9.6SEER 14.5 Unit (Single Spd, BPM fan mtr) 14.85 0.77 0.35 375 0.18 14.5 11.2Two Speed (BPM fan Mtr) 14.85 0.77 0.35 375 0.18 17.7 11.2Variable Speed, Ductless 14.85 0.77 0.10 375

Input Parameters Rated Performance

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Table 7. Additional cooling Unit Characteristics for Two-Speed and Variable Speed Systems

Note: Gross EER is total coil cooling capacity divided by compressor and condenser fan power at the nominal rating point: 95°F outdoors, 80°F/67°F entering coil and 450 cfm per ton supply airflow.

The airflow in the cooling mode is assumed to be 375 cfm per ton. In the heating mode the fan airflow is 27% lower at 275 cfm per ton. Any other fan operation for ventilation or mixing is also assumed to use the heating airflow rate. For the two-speed and variable speed units the airflow corresponding to the lowest stage (from the table above) is assumed to be used for ventilation or mixing. Set Points Thermostat set points for heating and cooling are listed in Table 8. The 70°F heating set point was selected as appropriate for temperate climates while the 72°F was deemed as more appropriate for the warm, humid climates (the original plan had been to use 68°F heating set point). The cooling set point of 78°F was selected as most consistent with homeowner preferences in warm climates and is also consistent with the HERS Reference House according to the 2006 Mortgage Industry National Home Energy Rating Systems Standards.

Table 8. Heating Cooling and Dehumidification Set Points

Heating Set Point (°F)

Cooling Set Point

(°F)

Dehumidification Set Point (%)

0 – Orlando 72

78

50% or 60% 1 – Miami 2 – Houston 3 – Atlanta

70 4 – Nashville 5 – Indianapolis

The impact of thermostat deadband and anticipator was explicitly considered in this short timestep model in the cooling mode as per Henderson (1992). The deadband was ±1°F around the desired temperature point. The anticipator temperature gain was 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 a temperature “droop” with runtime fraction that was about 2°F. In the heating mode a simple deadband of ±1°F around the set point was used without an anticipator.

Description

Gross EER

(Btu/Wh) SHR (-)

Actual Fan

Power (W/cfm)

Low Stage Gross

EER (Btu/Wh)

Low Stage

SHR (-)

Low Stage

Capacity Ratio (-)

Low Stage

Fan Power

(W/cfm)

Low Stage

Airflow Ratio (-)

SEER 10 Unit (Single Spd, PSC fan mtr) 10.35 0.77 0.50 0.3SEER 13 Unit (Single Spd, PSC fan mtr) 13.80 0.77 0.50 0.3SEER 14.5 Unit (Single Spd, BPM fan mtr) 14.85 0.77 0.35 0.1Two Speed (BPM fan Mtr) 14.85 0.77 0.35 18.56 0.77 0.5 0.1 0.5Variable Speed, Ductless 14.85 0.77 0.10 20.56 0.77 0.33 0.07 0.30

Input Parameters Low Stage Input Parameters

35

For explicit dehumidification systems as well as for air conditioners with an enhanced dehumidification mode, the dehumidification set point was specified in terms of relative humidity. A deadband of ±1.5% around the target value was used. Two different humidity set points (50% and 60% RH) were considered corresponding to the often-stated expectations of different groups within the industry. One variation on traditional dehumidification control is to reset the cooling setpoint based on the space humidity level. This overcooling control method (used in System 2 below) lowers the cooling set point by as much 2°F as the RH increases 12% above the set point. The degree of overcooling is linearly reset as the humidity increases above the set point. While in the overcooling mode, compressor operation is limited to 50% (10 minutes ON, 10 minutes OFF) 2.4.4 Mechanical Ventilation Options Several ventilation options were considered in this study, including:

Exhaust fan only

Supply - Central fan integrated (CFIS)

Heat recovery ventilator (HRV), with AHU mixing

Enthalpy Recovery Ventilator (ERV), with AHU mixing

Independent Enthalpy Recovery Ventilator (ERV), without AHU operation Figure 11 shows the airflow configuration used for each ventilation system. All these mechanical ventilation options provide the average rate of 58 cfm required by ASHRAE Standard 62.2 for the 2000 ft2 3 bedroom house. The HERS 130 was also modeled with no mechanical ventilation (infiltration only).

AHU

Return air

mechanical exhaust

Induced infiltration

infiltration exfiltration

Exhaust Only (exh fan operates 100% of hour)

AHU

Return air

Induced exfiltration

Mechanical ventilation

infiltration

exfiltration

Supply - Central Fan Integrated (damper/AHU operates 34% of hour @ 3x flow)

36

AHU

Return air

infiltration exfiltration

HRV

Balanced HRV / ERV – Ducted AHU (HRV & AHU operates 50% of hr @ 2x flow)

Return air

infiltration exfiltration

HRV

AHU

Balanced HRV / ERV – Ductless AHU (Sys 4) (HRV operates 50% of hr @ 2x flow)

Figure 11. Schematic of Mechanical Ventilation System Options

The combined impact of infiltration, ventilation, and duct leakage was considered by using the equations below. The duct leakage is always a net out, so that additional net flow is exhaust. cfmin = sum of all incoming ventilation flows (vent cfm at AC, dehum/vent unit, etc) cfmout = sum of all exhaust flows (exhaust fan, net duct leakage, etc) cfmbalanced = MIN(cfmin, cfmout) cfmunbalanced = MAX(cfmin, cfmout) - cfmbalanced cfminf = infiltration flow calculated for building for the timestep cfmcombined = MAX(cfmunbalanced , cfminf + 0.5* cfmunbalanced ) + cfmbalanced Some fresh air is provided as part of the mechanical system. Therefore, the net mechanical inlet flows are subtracted from cfmcombined to determine the remaining non-mechanical ventilation (or infiltration) rate acting on the building enclosure. A mass balance tracks CO2 levels in the space and confirms that the net impact of ventilation is similar between all the cases. ExhaustOnly. The exhaust fan runs 100% of the time (independent of the cooling unit) to provide the necessary ventilation for the house. The exhaust fan power is assumed to be 0.4 Watt/cfm. Supply–CentralFanIntegrated. A fresh air damper is installed on the return side of the air conditioner. Including calls for heating and cooling, the supply air fan runs a minimum of 34% of each hour (at heating speed, fan power is 0.4 Watt/cfm) to provide the necessary ventilation for the house. The fresh air damper is open during fan operation for no more than 34% of each hour. The ventilation damper provides 174 cfm of fresh air regardless of the airflow rate (heating or cooling airflow).

37

BalancedVentilation–ERV/HRV. An Energy Recovery Ventilator (enthalpy wheel) or Heat Recovery Ventilator (sensible heat exchanger) is installed to provide balanced ventilation. The assumed performance is

ERV effectiveness: 70% sensible, 60% moisture effectiveness HRV effectiveness: 75% sensible

The airflow through the ventilator unit was twice the continuous ventilation requirement, but the fan only runs 50% of each hour (initiated at the top of the hour). The fan power is 0.5 Watt/cfm based on the ventilation flow. The ERV/HRV pulls exhaust air from the return side of the central air distribution system (or from the house) and supplies tempered fresh air back to the return duct (see Figure 11). The air handler unit (AHU) supply fan is interlocked with HRV/ERV fan to ensure distribution of the ventilation air. The AHU fan operates at the lowest required airflow when there is not a call for cooling or heating. These mechanical ventilation options were implemented for each of the cooling system options described below. One change implemented in this newest version of TRN-RESDH was to allow for a Dehumidifier to be combined with the HRV/ERV option, so this combination could be considered. The ventilation options are not required in combination with System 7 (Dehumidifier with ventilation). This system provides ventilation in conjunction with dehumidification, so it did not consider other ventilation options. The other slight variations or exceptions, are for the two-speed air conditioner (System 3) which has a very small minimum airflow, therefore the ventilation rate for the central fan integrated supply option is assumed to provide 116 cfm of ventilation 50% of the time (instead of 174 cfm 34% of the time). The variable speed air conditioner (System 4) is a ductless unit, so the CFIS option is not possible. The fan on the ductless cooling is still assumed to run at low speed in conjunction with the HRV/ERV to provide some degree of mixing. As a sensitivity study, we also considered an ERV that was fully decoupled from the central air distribution system ducts and AHU operation in selected situations. The ERV pulls air directly from the space and supplies conditioned ventilation air directly back to the space (therefore, interlocked AHU fan operation is not required). 2.4.5 Space Conditioning Systems Fourteen (14) cooling and dehumidification systems were identified for evaluation in the Task 1 report. These systems are summarized in Table 9. Specific details for each system are given below.

38

Table 9. Summary of Space Conditioning Systems

System No

System Description

Control Based on Humidity Set Pt?

AHU Mixing Required

1 Conventional DX System No No

2 Conventional DX System with Lower Airflow and Thermostat Overcooling

Yes (reset) No

3 Two-Speed Conventional DX System No No

4 Variable Speed Mini-Split DX System No No

5 Stand-Alone Dehumidifier with Conventional System Mixing

Yes Yes

6 Ducted Dehumidifier with Conventional System Mixing Yes Yes

7 Ducted Dehumidifier with Outdoor Air Preconditioning Yes Yes

8 Enhanced Cooling with Subcooling Reheat (single speed & two-stage compressor)

Yes No

9 Enhanced Cooling with Full Condensing/Subcooling Reheat (single-speed and two-stage compressor)

Yes No

10 Conventional DX System with Lower Airflow Yes No

11 Conventional DX System with Thermostat Overcooling reset No

12 Conventional DX System with Sensible-Only AAHX No No

13 Natural Gas-Regenerated Desiccant Dehumidifier Yes Yes

14 DX Condenser-Regenerated Desiccant Dehumidifier Yes Yes

System1–ConventionalDXSystem. A conventional, single-speed DX air conditioner (air-cooled condenser) with a PSC fan motor (SEER 13). The system has a “gross” EER of 13.8 Btu/Wh at Rated Conditions. In cooling the system operates at a supply airflow rate of 375 cfm/ton with corresponding fan power of 0.5 Watts/cfm. The Heating airflow is 275 cfm/ton. System2–ConventionalDXSystemwithLowerAirflowandThermostatOvercooling. Same as System 1 above but with controls to provide lower airflow (reduced to 200 cfm/ton and space overcooling, as much as 2°F below setpoint, when space humidity is high). When space humidity increases above the RH set point, the cooling set point is reduced to continue cooling operation. While in the overcooling mode, compressor operation is limited to 50% runtime (10 minutes ON, 10 minutes OFF). The operating fan power for PSC motor fan drops from 0.5 to 0.3 W/cfm at low airflow. The operating fan power for brushless permanent magnet (BPM) motor drops from 0.35 to 0.1 W/cfm at low airflow. 

39

System3–Two‐SpeedConventionalDXSystem. A two-speed cooling unit with a gross EER of 14.85 Btu/Wh at high speed and 18.56 Btu/Wh at low speed. Low speed is half of high speed capacity. The resulting SEER is 17.7 Btu/Wh. The operating supply air fan power with the ECM motor is 0.35 W/cfm at 375 cfm/ton for high speed. Compressor capacity and airflow drop at the same proportion to maintain the same cfm per “delivered ton”. Table 7 shows the detailed characteristics that define this unit. Appendix D of Rudd and Henderson et al. (2013) provides further details about how this system was controlled and modeled. Since the different HERS levels required different air conditioner efficiencies, the difference between HERS level and System are not always clear. Table 10 summarizes which air conditioner was used with each HERS level and System. The HERS 50 and 70 actually used the 14.5 SEER single-speed unit for System 1 as a sensitivity. Therefore, the default air conditioner for HERS 50 and 70 are actually the listed under System 3. The gray-shaded entries in the table show the default air conditioner used for the conventional air conditioner at each HERS level. Systems 5 and higher use the 17.7 SEER two-speed as the standard or conventional cooling system for HERS 50 and 70 (except for System 12, which is discussed below).

Table 10. Matrix of cooling Conditioner Systems Used with each HERS Level and System

Note: The number in parentheses is the unit index in TRNSYS.

Grey shaded cells indicate the default for each HERS level

System4–VariableSpeedMini‐SplitDXSystem. A variable-speed, ductless air conditioner unit with the ability to modulate airflow with compressor speed. The system characteristics were selected to achieve an SEER over 20. Table 7 shows the detailed characteristics that define this unit. The operating fan power is 0.1 W/cfm at 375 cfm/ton for high speed compressor operation (Larson 2010). Lower compressor speeds allow for a lower supply airflow rate per “delivered ton” with correspondingly lower fan power. Appendix D of Rudd and Henderson et al. (2013) provides further details about how this system was modeled. System5–Stand‐AloneDehumidifierwithConventionalSystemMixing. A portable, stand-alone dehumidifier (DH) supplements the air conditioner. A small 50 pint/day dehumidifier with

System 1 Conven

AC

System 2 Enhanced

AC

System 3 Two-Spd

AC

System 4 Var-Spd

AC

Systems 5-14

HERS 130SEER 10 1-Spd (16)

SEER 10 1-Spd (16)

SEER 17.7 2-Spd (19)

SEER 19 V-Spd (20)

SEER 10 1-Spd (16)

HERS 100SEER 13 1-Spd (17)

SEER 13 1-Spd (17)

SEER 17.7 2-Spd (19)

SEER 19 V-Spd (20)

SEER 13 1-Spd (17)

HERS 85SEER 14.5 1-Spd (18)

SEER 14.5 1-Spd (18)

SEER 17.7 2-Spd (19)

SEER 19 V-Spd (20)

SEER 14.5 1-Spd (18)

HERS 70SEER 14.5 1-Spd (18)

SEER 17.7 2-Spd (19)

SEER 17.7 2-Spd (19)

SEER 19 V-Spd (20)

SEER 17.7 2-Spd (19)

HERS 50SEER 14.5 1-Spd (18)

SEER 17.7 2-Spd (19)

SEER 17.7 2-Spd (19)

SEER 19 V-Spd (20)

SEER 17.7 2-Spd (19)

40

an Energy Factor (EF) of 1.55 liters/kWh at AHAM rating conditions5 . The dehumidifier operates to maintain the dehumidification set point independently of the conventional air conditioner. Fan controls ensure that the AHU supply fan runs 7 to 11 minutes of at each hour (depending on the supply airflow) to provide mixing in the space corresponding to a turnover rate of 0.5 Air changes per hour. Since this DH unit is ductless it does not have separate fan power. Appendix E of Rudd and Henderson et al. (2013) provides the detailed performance curves for this system. System6–DuctedDehumidifierwithConventionalSystemMixing. A slightly larger 64 pint/day ducted dehumidifier with an EF of 1.98 liters/kWh at AHAM rating conditions. This ducted unit includes a fan that is integrated with the cooling unit in a recirculation configuration (pulling air from the main zone and then supplying air into the AHU supply duct). This configuration requires that the dehumidifier unit have a back draft damper to ensure the AHU supply fan does not drive airflow backwards through the unit when the DH is off. The dehumidifier operates to maintain the required humidity set point. The equipment performance map for this unit was based on an UltraAire 70H unit that was tested in the NREL HVAC lab (see Appendix E of Rudd and Henderson et al. (2013)). The airflow through the 64 pint/day dehumidifier was assumed to provide 135 cfm with a fan power of 0.7 W/cfm. The extra fan power for this ducted system reflects the airflow and fan data provided by the manufacturer (Thermastor 2010). Fan controls ensure that the conventional cooling supply air fan runs 7 to 11 minutes of each hour (depending on the supply airflow) to provide mixing in the space corresponding to a turnover rate of 0.5 Air changes per hour. System7–DuctedDehumidifierwithOutdoorAirPreconditioning. Larger 82 pint/day dehumidifier with an EF of 1.98 liters/kWh at AHAM rating conditions. This ducted unit includes a fan that is integrated with the cooling unit in a ventilation air pre-conditioning configuration, as shown in Figure 13. The dehumidifier fan operates to bring in 67% return air and 33% ventilation air (a ratio of 2 to 1). Therefore, the total flow of 174 cfm through the dehumidifier includes 58 cfm of ventilation air. The dehumidifier fan runs continuously to provide the required ventilation airflow. The supply air from the dehumidifier unit is discharged into the cooling supply duct. The dehumidifier compressor operates in response to the humidity set point. The equipment performance map for this unit was based on an UltraAire 70H unit that was tested in the NREL HVAC lab (see Appendix E of Rudd and Henderson et al. (2013)). Fan controls ensure that the cooling supply air fan runs 7 to 11 minutes of each hour (depending on the supply airflow) to provide mixing in the space corresponding to a turnover rate of 0.5 Air changes per hour. System8–EnhancedCoolingwithSubcoolingReheat. Central DX cooling system with refrigerant subcooling/reheat coil. Based on either a single-speed or a two-speed cooling unit depending on the HERS level, but with enhanced mode operation when humidity is high (i.e., control scheme based on Lennox Humiditrol system). Specifics regarding this system’s dehumidification operating mode are provided in the Rudd and Henderson et al. (2013) Task 1 report (Section 4.4.1) and in Appendix F of the same reference. 5 AHAM rating point is 80°F and 60%RH. Systems are rated at zero static.

41

System9–EnhancedCoolingwithFullCondensing/SubcoolingReheat. Central DX cooling system with modulating hot gas reheat providing full condensing at an indoor reheat coil. Based on either a single-speed or a two-speed cooling unit depending on the HERS level. The fan power is increased by 0.05 Watt/cfm compared to the conventional units to account for the extra pressure drop of the reheat coil. Specifics regarding this system’s dehumidification operating mode are provided in the Task 1 report (Section 4.4.2) and in Appendix G of Rudd and Henderson et al. (2013). System10‐ConventionalDXSystemwithLowerAirflow. Similar to System 2 but only includes controls to provide lower airflow (reduced to 200 cfm/ton, or to 53.3%) when space humidity rises above the set point. This system was only considered in a subset of the climates and HERS levels. System11‐ConventionalDXSystemwithThermostatOvercooling. Similar to System 2, but only includes controls to allow for space overcooling (by as much as 2 F below the cooling setpoint) when space humidity increases above the RH set point. The reset schedule linearly varies the about overcooling in proportion to the increase in RH above the dehumidification set point. This system was only considered in a subset of the climates and HERS levels.  

System12–ConventionalDXSystemwithSensible‐OnlyAAHX. Conventional DX cooling system with a refrigerant heat pipe sensible air-to-air heat exchanger (AAHX) “wrapped around” the DX evaporator coil for improved dehumidification performance. The non-powered heat pipe AAHX lowers the temperature of the air entering the DX evaporator coil and raises the temperature of the air leaving the evaporator coil. The heat pipe is assumed to be a non-condensing, aluminum fin, 1/2 inch O.D. copper tube, 2 row, 11 fpi. Face area of the heat pipe is set equal to the nominal cooling tonnage. The operating fan power increases to 0.55 Watt/cfm and the airflow drops to 350 cfm/ton to account for the additional heat exchanger pressure drop compared to System 1. This system uses the 14.5 SEER single-speed cooling in the HERS 50 and HERS 70 houses (since two-speed operation would result in condensing on the heat pipe at low speed). System13–ConventionalDXSystemwithNaturalGas‐RegeneratedDesiccantDehumidifier. Conventional cooling unit with natural gas-regenerated desiccant dehumidifier (e.g., NovelAire ComfortDry 400). The unit pulls in regeneration air from outdoors and then exhausts it back to outdoors. The 400 cfm unit has a dehumidification capacity of 6.3 lb/h (145 pint/day) at AHAM rating conditions. Natural gas consumption is 10,000 Btu/h. The process side of the desiccant unit is arranged to pull air from the supply duct (downstream of the cooling coil) and provide dehumidified process air back to the cooling supply duct, as shown in Figure 12. The process side (or supply) fan power is 0.6 Watt/cfm. Fan controls ensure that the conventional cooling supply air fan runs 7 to 11 minutes of each hour (depending on the supply airflow) to provide mixing in the space corresponding to a turnover rate of 0.5 Air changes per hour. Appendix H of Rudd and Henderson et al. (2013) provides the detailed performance curves to model this unit.

42

Figure 12. Configuration of Natural Gas-Fired Desiccant Unit (pulling air from the cooling outlet;

supplying to supply duct)

System14‐DXCondenser‐RegeneratedDesiccantDehumidifier. Conventional air conditioner with a desiccant system regenerated by condenser waste heat from its small internal compressor (e.g., NovelAire 300). The unit’s energy factor is 2.6 liters/kWh (excluding supply fan power). The supply fan is 300 cfm at 210 Watts (0.7 W/cfm). Controls ensure that the conventional cooling supply air fan runs at least 7 minutes out of every hour to provide mixing in the space (to provide a zone air mixing rate of 0.5 ACH). A performance map for the this desiccant unit has been developed based on data from the NovelAire 300 (see Appendix H of Rudd and Henderson et al. (2013)) Summary of Dehumidifier Configurations Figure 13 schematically shows the configuration for the dehumidifier and desiccant systems. We have assumed that ducted dehumidifiers and desiccant systems provide air into the supply trunk of the main air handler unit (AHU). This helps to distribute dehumidified air throughout the house. However, practical experiences with these systems have shown that some AHU fan operation is required to provide adequate air distribution. Therefore, the supply air fan is required to operate for a minimum fraction of each hour to provide a desired air mixing rate for the space. The AHU supply fan is operated to provide a turnover rate of 0.5 air changes per hour. 2.4.6 Electric and Natural Gas Costs Total HVAC costs (cooling, heating, fans) were determined using the electric and natural gas rates from Table 11. The heating load was tracked as kWh with no efficiency losses. Therefore, heating costs for the natural gas furnace case were determined by assuming a furnace efficiency and using the natural gas costs from the table below. Furnace efficiencies are 80% for HERS 130, 85% for HERS 100, 87% for HERS 85 and HERS 70, and 93% for HERS 50.

43

Table 11. Electric and Natural Gas Costs

Electric Utility $/kWh $/therm Z0 – Orlando FPL 0.101 1.79 Z1 – Miami FPL 0.101 1.79

Z2 – Houston Entergy 0.085 1.08 Z3 – Atlanta Georgia Power 0.103 1.52

Z4 – Nashville Nashville 0.097 1.05 Z5 – Indianapolis Indianapolis PL 0.078 0.86

Notes: Electric costs are from Form 826 data for the local utility in 2010 for residential. Natural gas costs are average residential rate for each state in 2010.

44

AHU

Return air

infiltration exfiltration

DH

Stand-Alone Dehumidifier (System 5)

AHU

Return air

infiltration exfiltration

DH

Ducted Dehumidifier (System 6)

AHU

Return air

infiltration exfiltration

DH

Ducted Dehumidifier with OA (System 7)

AHU

Return air

infiltration exfiltration

De

s

Natural Gas-Fired Desiccant DH (System 13)

AHU

Return air

infiltration exfiltration

De

s

Desiccant Dehumidifier (System 14)

Figure 13. Schematic of Configurations for Dehumidifier and Desiccant Systems

45

2.4.7 Simulation Results Several hundred simulations were run for the various climates, buildings, systems, RH set points and ventilation options. The designation for the each run follows the nomenclature shown below:

Z1 H100 S1 RH50 V1 Z0 = Orlando Z1 = Miami Z2 = Houston Z3 = Atlanta Z4 = Nashville Z5 = Indianapolis

HERS50 HERS70 HERS85 HERS100 HERS130

S1 = System 1 …. S14 = System 14

RH50 = 50% RH RH60 = 60% RH

V0 = No Vent V1 = Exh Fan V2 = CFIS V3 = HRV V4 = ERV

Summary data from all the annual simulation runs are available at the website http://cloud.cdhenergy.com/rp1449/. The sections below summarize the results focusing on the observed trends and themes.

HERS Level and Climate with Conventional Cooling Simulations were run for houses at 5 different HERS levels and in 6 climates. The results are listed in Table 12. The HERS 130 level was run for the case with the ASHRAE 62.2 recommended ventilation rate via an exhaust fan (V1) as well as for the case without mechanical ventilation (V0), to represent the current housing stock. All other HERS levels used an exhaust fan for ventilation. Figure 14 compares the total costs for the different HERS and climates using the data from Table 12. The relative costs are approximately (but not precisely) in line with the ratios implied by the HERS level (see Table 13).

46

Table 12. Performance Results for Different HERS Levels and Climates

Table 13. Comparing Relative Annual Costs for each HERS Level and Climate

Hours Above

60% RH

AC Runtime

(hrs) AC EER

(Btu/Wh)

AC Energy

(kWh)

Htg Energy

(kWh)

AHU Fan Energy

(kWh)

Exh Fan Energy

(kWh)

Total Electric w/o HT

(kWh)

Total Costs w

Furnace ($)

HERS 50 1,011 3,266 21.3 1,869 633 140 203 2,211 265$ HERS 70 575 3,491 21.0 2,615 673 217 203 3,036 354$ HERS 85 1,121 2,039 16.6 3,611 2,247 700 203 4,514 614$

HERS 100 1,645 1,860 15.4 4,302 5,642 1,179 203 5,684 980$ HERS 130 1,361 2,127 11.7 6,629 6,876 1,358 203 8,190 1,352$

HERS 50 822 4,378 21.0 2,516 38 173 203 2,893 295$ HERS 70 382 5,303 20.0 3,629 42 360 203 4,192 426$ HERS 85 756 2,859 16.5 5,064 420 947 203 6,214 657$

HERS 100 1,303 2,700 15.3 6,232 1,513 1,572 203 8,007 917$ HERS 130 1,313 3,054 11.6 9,505 1,944 1,786 203 11,495 1,309$

HERS 50 380 3,252 20.4 1,963 1,758 181 203 2,347 269$ HERS 70 191 3,844 19.3 2,856 2,034 350 203 3,410 376$ HERS 85 228 2,185 16.3 3,876 3,047 759 203 4,838 540$

HERS 100 628 1,966 15.0 4,572 7,687 1,286 203 6,061 849$ HERS 130 942 2,170 11.4 6,823 14,261 1,555 203 8,581 1,386$

HERS 50 40 2,054 21.1 1,149 3,370 135 203 1,488 341$ HERS 70 15 2,604 19.7 1,815 4,594 271 203 2,289 510$ HERS 85 20 1,703 16.6 2,364 6,228 515 203 3,082 689$

HERS 100 291 1,338 15.2 2,546 13,004 881 203 3,631 1,168$ HERS 130 426 1,726 11.7 3,530 18,387 935 203 4,668 1,673$

HERS 50 - 2,184 21.0 1,248 4,254 160 203 1,611 320$ HERS 70 - 2,641 19.7 1,859 5,975 299 203 2,361 475$ HERS 85 - 1,705 16.6 2,379 8,000 535 203 3,118 632$

HERS 100 4 1,555 15.4 2,938 10,150 928 203 4,069 823$ HERS 130 119 1,658 11.6 4,246 18,838 1,145 203 5,594 1,387$

HERS 50 - 1,457 21.5 812 8,523 198 203 1,213 364$ HERS 70 - 1,828 20.5 1,206 12,390 327 203 1,736 553$ HERS 85 - 1,154 17.0 1,582 13,605 452 203 2,238 634$

HERS 100 - 1,057 15.8 1,960 18,145 850 203 3,013 862$ HERS 130 7 1,240 12.0 3,090 22,902 1,029 203 4,322 1,177$

Orlando

Miami

Houston

Atlanta

Nashville

Indianapolis

HERS50 HERS70 HERS85 HERS100 HERS130HERS130 No Vnt

Z0-Orlando 27% 36% 63% 100% 138% 129%Z1-Miami 32% 46% 72% 100% 143% 134%Z2-Houston 32% 44% 64% 100% 163% 154%Z3-Atlanta 29% 44% 59% 100% 143% 135%Z4-Nashville 39% 58% 77% 100% 169% 157%Z5-Indianapolis 42% 64% 74% 100% 137% 126%

47

Figure 14. Comparing Total Costs for Different Climates and HERS Levels

Figure 15 shows that when heating energy is excluded (designated as “Total Electric w/o HT”), then the electric use in Miami is greater than the other climates, followed by Houston and Orlando. The air conditioner operating hours are also significantly greater in Miami, as shown in Figure 16. The annual runtime for the two-speed cooling units in the HERS 50 and 70 houses is also much greater than the single speed units in the other houses since the unit runs at low speed for longer periods of time.

Figure 15. Comparing Cooling and Fan Energy Use for Different Climates and HERS Levels

‐

200 

400 

600 

800 

1,000 

1,200 

1,400 

1,600 

1,800 

Z0‐Orlando Z1‐Miami Z2‐Houston Z3‐Atlanta Z4‐Nashville Z5‐Indianapolis

Total Costs w Furnace ($) 

System 1 ‐ Exh Fan (V1)

HERS50 HERS70 HERS85 HERS100 HERS130 HERS130 No Vnt

‐

2,000 

4,000 

6,000 

8,000 

10,000 

12,000 

14,000 

Z0‐Orlando Z1‐Miami Z2‐Houston Z3‐Atlanta Z4‐Nashville Z5‐Indianapolis

Total Electric w/o

 HT (kWh) 

System 1 ‐ Exh Fan (V1)

HERS50 HERS70 HERS85 HERS100 HERS130 HERS130 No Vnt

48

Figure 16. Comparing Air Conditioner Runtime for Different Climates and HERS Levels

The resulting annual EER for the cooling system in each scenario are compared in Figure 17. The EERs are based on gross cooling capacity (w/o fan heat) and using power for the outdoor unit only. The HERS 130 cases use a SEER 10 unit and result in gross annual EERs around 11-12 Btu/Wh. The HERS 85 and 100 houses (with 13 and 14.5 SEER units) results in gross EERs near 15 and 16 Btu/Wh. The HERS 70 and 50 houses use two speed units with a 17.7 SEER that result in gross EERs exceeding 21 Btu/Wh.

Figure 17. Comparing Air Conditioner “Gross” EER for Different Climates and HERS Levels

‐

1,000 

2,000 

3,000 

4,000 

5,000 

6,000 

Z0‐Orlando Z1‐Miami Z2‐Houston Z3‐Atlanta Z4‐Nashville Z5‐Indianapolis

AC Runtime (hrs) 

System 1 ‐ Exh Fan (V1)

HERS50 HERS70 HERS85 HERS100 HERS130 HERS130 No Vnt

‐

5 

10 

15 

20 

25 

Z0‐Orlando Z1‐Miami Z2‐Houston Z3‐Atlanta Z4‐Nashville Z5‐Indianapolis

AC EER

 (Btu/W

h) 

System 1 ‐ Exh Fan (V1)

HERS50 HERS70 HERS85 HERS100 HERS130 HERS130 No Vnt

49

2.4.8 Evaluating Humidity Levels Humidity in the space is not directly controlled by the conventional cooling unit, so humidity levels typically vary across the year. The psychrometric chart in Figure 18 shows the average daily conditions observed across the year for the HERS 100 house in Miami with System 1. Each point on the plot represents the average conditions for the day. The days with any cooling operation are shown as blue. The total number of hours over 60% RH was 1,303 per year. Most of these hours occur in Miami when the cooling unit is off for the day. Figure 19 is a shade plot that shows the humidity bins for each hour of the year with shades of gray. Each day is shown as a vertical stripe on the plot. Successive days are shown along the X axis. Darker shades indicate hours with higher humidity levels (0, which indicates the hours below 55% RH, corresponds to light gray; 5, which indicates the hours above 75% RH is black). This plot confirms that high humidity does not occur in the main part of summer but tends to happen in swing seasons and in the winter when little or no cooling operation is required. Finally, Figure 20 is a histogram showing the distribution of hours across the year at each humidity level for Miami, HERS 100, System 1, with exhaust ventilation. All the plots discussed in this section are also available in PDF form for each run at the web site referenced above. The most common metric for gauging the prevalence of high humidity is the number of hours above some threshold per year. Figure 21 includes three plots comparing the number of hours above 60%, 55%, and 50% RH, respectively. While there is not a hard requirement or limit on humidity, we believe that the hours above 60% is a reasonable gauge when comparing different systems. The three different plots essentially show the same patterns when comparing the HERS level and the climates. Other metrics such as the number of high humidity events of certain duration also show the same result. Figure 22 shows the number of events where the humidity was above 60% RH for periods of 4 hours and 8 hours, respectively. The results in Figure 21 show the expected trends for conventional systems, consistent with previous studies by Henderson et al 2007 and Henderson et al 2008. The HERS 130 house without ventilation typically has lower humidity levels. When ventilation is provided to the house, the hours of high humidity tend to increase. Orlando generally has the highest levels, followed by Miami and Houston. As the HERS level decreases, the hours of high humidity tend to decrease and are lowest for HERS levels of 70 and 85. The hours of high humidity increase again for the HERS 50 house, when ducts are moved into the conditioned space. The next section explores the impact of duct location in greater detail.

50

Figure 18. Psychrometric Chart Showing Space Conditions for HERS 100, Miami, System 1, Exhaust Fan

Figure 19. Shade Plot showing Humidity “bins” for each Hour of Year for HERS 100, Miami,

System 1, Exhaust Fan

Daily Indoor Space Conditions: z1h100s1rh50v1

60 65 70 75 80 85 90

Dry Bulb Temperature (F)

0.000

0.005

0.010

0.015

0.020

0.025

Hum

idity

Rat

io (

lb/lb

)

40%

50%

60%

70%

80%

Hours Above 50% = 5361Hours Above 55% = 2399Hours Above 60% = 1303Hours Above 65% = 543Hours Above 70% = 107All HrsCooling Hrs

Jan Feb Mar Apr May Jun Jul Aug Sep Oct Nov Dec

2006

0

2

4

6

8

10

12

14

16

18

20

22

24

Hou

r o

f D

ay

Days with Cooling Operation

Shades of Gray 0 – below 55% 1 – 55-60% 2 – 60-65% 3 – 65-70 4 – 70-75% 5 – above 75%

51

Figure 20. Histogram of Relative Humidity for HERS 100, Miami, System 1, Exhaust Fan

RH: z1h100s1rh50v1

30 40 50 60 70 80

Indoor Relative Humidity (%)

0

200

400N

umbe

r of

Hou

rs

Min RH = 42.0

Avg RH = 53.1

Max RH = 75.8

52

Figure 21. Comparing High Humidity Levels for Different Climates and HERS Levels

‐

200 

400 

600 

800 

1,000 

1,200 

1,400 

1,600 

1,800 

Z0‐Orlando Z1‐Miami Z2‐Houston Z3‐Atlanta Z4‐Nashville Z5‐Indianapolis

Hours Above

 60% RH 

System 1 ‐ Exh Fan (V1)

HERS50 HERS70 HERS85 HERS100 HERS130 HERS130 No Vnt

‐

500 

1,000 

1,500 

2,000 

2,500 

3,000 

3,500 

4,000 

Z0‐Orlando Z1‐Miami Z2‐Houston Z3‐Atlanta Z4‐Nashville Z5‐Indianapolis

Hours Above

 55% RH 

System 1 ‐ Exh Fan (V1)

HERS50 HERS70 HERS85 HERS100 HERS130 HERS130 No Vnt

‐

1,000 

2,000 

3,000 

4,000 

5,000 

6,000 

7,000 

8,000 

Z0‐Orlando Z1‐Miami Z2‐Houston Z3‐Atlanta Z4‐Nashville Z5‐Indianapolis

Hours Above

 50% RH 

System 1 ‐ Exh Fan (V1)

HERS50 HERS70 HERS85 HERS100 HERS130 HERS130 No Vnt

Above 60%

Above 55%

Above 50%

53

Figure 22. Comparing Number of High Humidity Events (exceeding 60% RH) for Different Climates

and HERS Levels

‐

10 

20 

30 

40 

50 

60 

70 

80 

Z0‐Orlando Z1‐Miami Z2‐Houston Z3‐Atlanta Z4‐Nashville Z5‐Indianapolis

No of Events > 4 hrs 

System 1 ‐ Exh Fan (V1)

HERS50 HERS70 HERS85 HERS100 HERS130 HERS130 No Vnt

‐

10 

20 

30 

40 

50 

60 

Z0‐Orlando Z1‐Miami Z2‐Houston Z3‐Atlanta Z4‐Nashville Z5‐Indianapolis

No of Events > 8 hrs 

System 1 ‐ Exh Fan (V1)

HERS50 HERS70 HERS85 HERS100 HERS130 HERS130 No Vnt

54

2.4.9 Location of Ducts in the Conditioned Space One of the main differences between the HERS 70 and HERS 50 houses is the elimination of ducts in the attic. Locating the supply ducts in the conditioned space clearly reduces energy use. To further evaluate the impact of duct location on costs and humidity, we repeated the HERS 70 run without ducts in the attic and the HERS 50 run was repeated with ducts in the attic. Removing ducts from the attic decreases total space conditioning costs by 10 to 28% depending on climate and HERS level, as shown by Table 14. The reduction was 22 to 28% in the three hot-humid climates. Figure 23 also shows that moving the ducts inside the conditioned space clearly results in more hours with high humidity in all cases. The reduction in sensible cooling loads is greater than the latent load reduction, leaving a mix of latent and sensible loads that is poorly matched to the sensible heat ratio of conventional air conditioning systems, so there is a net increase in humidity levels.

Table 14. Total Space Conditioning Cost Reduction Impact of Moving Ducts from Attic to the Conditioned Space

HERS 70 HERS 50Z0-Orlando 22% 26%Z1-Miami 27% 28%Z2-Houston 26% 26%Z3-Atlanta 24% 24%Z4-Nashville 24% 16%Z5-Indianapolis 23% 10%

Cost Reduction

55

Figure 23. Impact of Duct Location on Space Humidity Levels for HERS 50 and 70 Houses

Table 15. Impact of Duct Location on Space Humidity Levels for HERS 50 and 70 Houses

2.4.10 Impact of Lower Duct Leakage and R-value The HERS 130 and HERS 100 houses have total duct leakage rates of 20% and 10% of cooling system airflow, respectively (60% on the supply side; 40% on the return side). Table 16 shows the impact of decreasing the leakage rate to 5%. Energy use and operating costs are lower as expected. However, there is very little impact on humidity levels for the HERS 100 house. Reducing leakage in the HERS 130 house for Houston slightly increased the hours above 60% RH.

‐

200 

400 

600 

800 

1,000 

1,200 

Z0‐Orlando Z1‐Miami Z2‐Houston Z3‐Atlanta Z4‐Nashville Z5‐Indianapolis

Hours Above

 60% RH 

Location of Ducts

In Attic In Space In Attic In Space

% Diff. wrt

Ducts inside Ducts in attic1 ducts in attic

HERS 50Orlando 770 580 33%Miami 600 390 54%Houston 290 190 53%

HERS 70Orlando 1000 790 27%Miami 820 600 37%Houston 380 300 27%

1 ducts in attic have 5% leakage (3% supply, 2% return)

Hours >60% RH

HERS 50 HERS 70

56

Table 16. Performance Results for Different Duct Leakage Rates in HERS 100 and HERS 130 Houses

The HERS 70 and HERS 85 houses by default use R8 duct insulation. The results in Table 17 show the impact of decreasing the insulation level to R6. As expected, energy use increases with less duct insulation. The number of high humidity hours decrease slightly as more sensible heat gain through the ducts increases cooling runtime and removes more moisture.

Table 17. Performance Results for Different Duct Insulation Levels in HERS 70 and HERS 85 Houses

Hours Above

60% RH

AC Runtime

(hrs) AC EER

(Btu/Wh)

AC Energy

(kWh)

Htg Energy

(kWh)

AHU Fan Energy

(kWh)

Exh Fan Energy

(kWh)

Total Electric w/o HT

(kWh)

Total Costs w

Furnace ($)

10% Duct Leak 628 1,966 15.0 4,572 7,687 1,286 203 6,061 849 5% Duct Leak 628 1,877 14.9 4,337 7,410 1,229 203 5,769 812

10% Duct Leak 4 1,555 15.4 2,938 10,150 928 203 4,069 823 5% Duct Leak 4 1,489 15.3 2,797 9,776 889 203 3,889 789

20% Duct Leak 778 2,096 11.3 6,530 13,589 1,498 - 8,028 1,309 5% Duct Leak 849 1,871 11.1 5,736 12,237 1,340 - 7,076 1,165

20% Duct Leak 25 1,622 11.5 4,119 17,533 1,103 - 5,222 1,292 5% Duct Leak 17 1,451 11.3 3,634 15,698 987 - 4,621 1,151

HERS 130, no vent

Houston

Nashville

HERS 100, Exh Fan

Houston

Nashville

Hours Above

60% RH

AC Runtime

(hrs) AC EER

(Btu/Wh)

AC Energy

(kWh)

Htg Energy

(kWh)

AHU Fan Energy

(kWh)

Exh Fan Energy

(kWh)

Total Electric w/o HT

(kWh)

Total Costs w

Furnace ($)

Ducts R8 214 2,391 16.4 3,403 2,045 661 203 4,267 449 Ducts R6 195 2,498 16.4 3,554 2,172 691 203 4,449 470

Ducts R8 - 1,586 16.6 2,215 5,989 515 203 2,933 531 Ducts R6 - 1,666 16.6 2,327 6,350 542 203 3,071 559

Ducts R8 228 2,185 16.3 3,876 3,047 759 203 4,838 540 Ducts R6 218 2,266 16.3 4,020 3,188 787 203 5,010 561

Ducts R8 - 1,705 16.6 2,379 8,000 535 203 3,118 632 Ducts R6 - 1,782 16.6 2,487 8,402 560 203 3,250 661

Nashville

Houston

Nashville

HERS 85

Houston

HERS 70

57

2.4.11 Impact of Ventilation Options Several ventilation options were considered in this study. In all cases the interactions between the ventilation option, infiltration, and duct leakage were considered.

No Ventilation (V0). Fresh air to the house is only provided by infiltration. Used for HERS 130 and 100 house only.

Exhaust Fan (V1). 58 cfm is continuously exhausted from the house. Fan power is 0.4 W/cfm.

CFIS (V2). 174 cfm of fresh air is introduced into the return side of the AHU fan. Including calls for heating and cooling, the AHU fan operates at least 34% of time to provide 58 cfm on average. A ventilation damper shuts if the fan runtime exceeds 34% of the time in the hour. For two-speed cooling units the ventilation rate is changed to be 116 cfm for 50% of each hour since the low-stage airflow is 50% of full speed flow.

HRV (V3). The inlet to the HRV draws from the return side of the AHU. The outlet from the HRV is introduced into the return side of the AHU downstream of the inlet. The HRV supplies 116 cfm of outside air for 50% of each hour. The AHU fan is interlocked with HRV fan so that fresh air is distributed to the space. HRVs are used in the colder climates (Atlanta, Nashville, Indianapolis)

ERV (V4). Same operation and control as HRV. ERVs are used in the warmer climates (Orlando, Miami, Houston, Atlanta).

Independent ERV (V5). This ERV is decoupled from the central air distribution ducts and the AHU fan. It runs at 116 cfm 50% of each hour. The AHU fan does not run with the ERV fan. 

All these mechanical ventilation options provide the average rate of 58 cfm of outside air required by ASHRAE Standard 62.2 for the 2000 ft2, 3-bedroom house in addition to the interactions with natural infiltration and duct leakage. The duct leakage occurs only when the AHU is on and always results in net exhaust (since supply leaks are always greater than return leaks). Figure 24 shows the total combined air change rate for each ventilation system for Miami, HERS 100, System 1. Figure 25 shows a time series plot of the total combined airflow rate (as cfm) for typical winter and summer weeks. The combined ventilation for the balanced ERV system is the highest since mechanical ventilation and infiltration are directly added. Compared to exhaust, the CFIS provides more unbalanced ventilation part of the time, which overwhelms infiltration, especially when the forces driving infiltration (i.e., wind speed and temperature difference) are low. When infiltration rates are higher (later in the winter week, earlier in the summer week) the combined airflows for CFIS and exhaust cases tend to be similar. Table 18, Table 19 and Table 20 summarize the results for the various ventilation systems in all the climates for three different air conditioners (with different AHU fan options):

HERS 100, System 1. A 13 SEER cooling with a conventional AHU fan (0.5 W/cfm).

HERS 85, System 1. A 14.5 SEER cooling with ECM fan motor (0.35 W/cfm)

58

HERS 100, System 3. A two-speed cooling with ECM fan motor (0.35 W/cfm, 50% flow for vent) 

The different airflows and fan powers associated with these different systems slightly changes the operating cost ranking of the different ventilation options.

Figure 24. Comparing Total Ventilation Rates (mechanical & natural ACH) for Different Ventilation

Systems in Miami

z1h100s1rh50v0 - No Vent

0.00 0.10 0.20 0.30 0.40 0.50 0.60

ACH (1/h)

0

200

400

Nu

mb

er o

f Ho

urs

Min ACH = 0.010

Avg ACH = 0.160

Max ACH = 0.495

z1h100s1rh50v1 - Exh Fan

0.0 0.2 0.4 0.6 0.8

ACH (1/h)

0

100

200

300

400

500

Nu

mb

er o

f Ho

urs

Min ACH = 0.216

Avg ACH = 0.287

Max ACH = 0.603

z1h100s1rh50v2 - CFIS

0.0 0.2 0.4 0.6 0.8

ACH (1/h)

0

100

200

300

400

500

Nu

mbe

r of

Ho

urs

Min ACH = 0.190

Avg ACH = 0.323

Max ACH = 0.573

z1h100s1rh50v4 - ERV

0.0 0.2 0.4 0.6 0.8

ACH (1/h)

0

50

100

150

200

250

300N

um

ber

of H

our

sMin ACH = 0.253

Avg ACH = 0.389

Max ACH = 0.724

59

Figure 25. Comparing Total Ventilation Rates (mechanical & natural cfm) for Different Ventilation

Systems in Miami

Week of 01/01/06 - Winter

12: 0: 12: 0: 12: 0: 12: 0: 12: 0: 12: 0: 12: 0: 12: 0:

31 1 2 3 4 5 6 7 8

0

50

100

150

200

Tot

al V

ent

& In

f (c

fm)

Week of 06/16/06 - Summer

0: 12: 0: 12: 0: 12: 0: 12: 0: 12: 0: 12: 0: 12: 0: 12: 0:

16 17 18 19 20 21 22 23 24

0

50

100

150

200

Tot

al V

en

t & In

f (c

fm)

Black = Exh Fan Green = CFIS Blue = ERV

60

Table 18. Performance Results for Different Ventilation System and Climates (for HERS 100, System 1: 13 SEER cooling)

Hours Above

60% RH

AC Runtime

(hrs) AC EER

(Btu/Wh)

AC Energy

(kWh)

Htg Energy

(kWh)

AHU Fan Energy

(kWh)

Exh Fan Energy

(kWh)

HRV Energy

(kWh)

Total Electric w/o HT

(kWh)

Total Costs w

Furnace ($)

V0 - Natural 1,833 1,753 15.2 4,001 5,260 1,110 - - 5,110 894 V1 - Exh Fan 1,645 1,860 15.4 4,302 5,642 1,179 203 - 5,684 980

V2 - CFIS 1,627 1,889 15.5 4,406 5,543 1,837 - - 6,242 1,029 V3 - HRVV4 - ERV 1,742 1,850 15.4 4,300 5,312 2,243 - 254 6,797 1,068

V5 - ERV ind 1,623 1,834 15.3 4,221 5,447 1,159 - 254 5,634 961

V0 - Natural 1,119 2,553 15.2 5,821 1,382 1,484 - - 7,305 837 V1 - Exh Fan 1,303 2,700 15.3 6,232 1,513 1,572 203 - 8,007 917

V2 - CFIS 1,394 2,721 15.4 6,330 1,474 2,091 - - 8,422 956 V3 - HRVV4 - ERV 1,737 2,663 15.4 6,178 1,396 2,434 - 254 8,866 996

V5 - ERV ind 1,206 2,664 15.3 6,124 1,447 1,550 - 254 7,927 905

V0 - Natural 625 1,867 14.9 4,284 7,060 1,216 - - 5,500 774 V1 - Exh Fan 628 1,966 15.0 4,572 7,687 1,286 203 - 6,061 849

V2 - CFIS 728 1,983 15.1 4,646 7,666 1,890 - - 6,536 888 V3 - HRVV4 - ERV 793 1,943 15.0 4,536 7,391 2,286 - 254 7,076 922

V5 - ERV ind 590 1,941 15.0 4,489 7,399 1,265 - 254 6,008 832

V0 - Natural 226 1,294 15.1 2,435 11,997 841 - - 3,276 1,070 V1 - Exh Fan 291 1,338 15.2 2,546 13,004 881 203 - 3,631 1,168

V2 - CFIS 163 1,355 15.3 2,598 12,864 1,445 - - 4,043 1,202 V3 - HRV 157 1,380 15.3 2,650 12,475 1,812 - 254 4,715 1,247 V4 - ERV 62 1,335 15.2 2,542 12,556 1,796 - 254 4,593 1,239

V5 - ERV ind 194 1,330 15.2 2,519 12,541 868 - 254 3,642 1,140

V0 - Natural - 1,509 15.3 2,812 8,988 883 - - 3,695 737 V1 - Exh Fan 4 1,555 15.4 2,938 10,150 928 203 - 4,069 823

V2 - CFIS 6 1,556 15.4 2,962 10,225 1,471 - - 4,433 861 V3 - HRV 15 1,582 15.5 3,019 9,908 1,838 - 254 5,111 913 V4 - ERV

V0 - Natural - 1,045 15.7 1,914 16,336 809 - - 2,723 777 V1 - Exh Fan - 1,057 15.8 1,960 18,145 850 203 - 3,013 862

V2 - CFIS - 1,060 15.8 1,980 17,938 1,366 - - 3,346 880 V3 - HRV 9 1,085 15.9 2,033 17,669 1,733 - 254 4,019 924 V4 - ERV

Indianapolis

System 1, HERS 100

Orlando

Miami

Houston

Atlanta

Nashville

61

Table 19. Performance Results for Different Ventilation System and Climates (for HERS 85, System 1: 14.5 SEER cooling, ECM Fan)

Hours Above

60% RH

AC Runtime

(hrs) AC EER

(Btu/Wh)

AC Energy

(kWh)

Htg Energy

(kWh)

AHU Fan Energy

(kWh)

Exh Fan Energy

(kWh)

HRV Energy

(kWh)

Total Electric w/o HT

(kWh)

Total Costs w

Furnace ($)

V0 - Natural 1,449 1,906 16.4 3,314 1,946 652 - - 3,966 537 V1 - Exh Fan 1,121 2,039 16.6 3,611 2,247 700 203 - 4,514 614

V2 - CFIS 1,105 2,069 16.7 3,704 2,367 1,090 - - 4,794 650 V3 - HRVV4 - ERV 1,304 1,987 16.6 3,534 2,111 1,315 - 254 5,102 664

V5 - ERV ind 1,029 2,004 16.6 3,524 2,099 686 - 254 4,464 598

V0 - Natural 578 2,670 16.4 4,653 312 882 - - 5,536 581 V1 - Exh Fan 756 2,859 16.5 5,064 420 947 203 - 6,214 657

V2 - CFIS 865 2,893 16.6 5,179 433 1,241 - - 6,420 679 V3 - HRVV4 - ERV 1,149 2,775 16.6 4,938 367 1,422 - 254 6,613 694

V5 - ERV ind 616 2,806 16.5 4,941 367 928 - 254 6,123 644

V0 - Natural 278 2,064 16.2 3,595 2,551 712 - - 4,307 474 V1 - Exh Fan 228 2,185 16.3 3,876 3,047 759 203 - 4,838 540

V2 - CFIS 304 2,196 16.4 3,937 3,314 1,123 - - 5,059 570 V3 - HRVV4 - ERV 460 2,117 16.3 3,771 2,982 1,345 - 254 5,370 583

V5 - ERV ind 182 2,151 16.3 3,788 2,814 744 - 254 4,787 526

V0 - Natural 55 1,656 16.5 2,262 5,361 493 - - 2,755 603 V1 - Exh Fan 20 1,703 16.6 2,364 6,228 515 203 - 3,082 689

V2 - CFIS 10 1,712 16.7 2,399 6,532 848 - - 3,247 724 V3 - HRV 2 1,736 16.7 2,434 6,010 1,056 - 254 3,744 744 V4 - ERV - 1,670 16.6 2,318 6,082 1,043 - 254 3,615 735

V5 - ERV ind 10 1,697 16.6 2,338 5,812 509 - 254 3,101 666

V0 - Natural 6 1,648 16.5 2,260 6,913 508 - - 2,769 553 V1 - Exh Fan - 1,705 16.6 2,379 8,000 535 203 - 3,118 632

V2 - CFIS 4 1,708 16.7 2,406 8,415 849 - - 3,255 662 V3 - HRV 4 1,731 16.7 2,441 7,777 1,054 - 254 3,749 684 V4 - ERV

V0 - Natural - 1,146 17.0 1,545 11,919 431 - - 1,976 556 V1 - Exh Fan - 1,154 17.0 1,582 13,605 452 203 - 2,238 634

V2 - CFIS - 1,155 17.1 1,598 14,055 767 - - 2,365 659 V3 - HRV 2 1,185 17.2 1,641 13,244 981 - 254 2,876 671 V4 - ERV

Indianapolis

System 1, HERS 85

Orlando

Miami

Houston

Atlanta

Nashville

62

Table 20. Performance Results for Different Ventilation System and Climates (for HERS 100, System 3: Two-Speed Cooling)

Hours Above

60% RH

AC Runtime

(hrs) AC EER

(Btu/Wh)

AC Energy

(kWh)

Htg Energy

(kWh)

AHU Fan Energy

(kWh)

Exh Fan Energy

(kWh)

HRV Energy

(kWh)

Total Electric w/o HT

(kWh)

Total Costs w

Furnace ($)

V0 - Natural 1,786 2,948 19.9 2,914 5,300 459 - - 3,374 722 V1 - Exh Fan 1,635 3,012 19.8 3,147 5,676 511 203 - 3,862 798

V2 - CFIS 1,603 3,037 20.0 3,167 5,719 627 - - 3,794 794 V3 - HRVV4 - ERV 1,595 3,000 19.9 3,091 5,596 615 - 254 3,960 802

V0 - Natural 1,148 4,292 19.9 4,227 1,359 440 - - 4,667 569 V1 - Exh Fan 1,313 4,386 19.8 4,544 1,491 497 203 - 5,244 637

V2 - CFIS 1,289 4,399 19.9 4,542 1,512 580 - - 5,122 626 V3 - HRVV4 - ERV 1,360 4,354 19.9 4,442 1,470 567 - 254 5,263 637

V0 - Natural 634 3,032 19.0 3,214 7,019 496 - - 3,710 620 V1 - Exh Fan 620 3,078 18.9 3,452 7,659 553 203 - 4,209 690

V2 - CFIS 701 3,104 19.1 3,440 7,770 660 - - 4,100 685 V3 - HRVV4 - ERV 675 3,068 19.0 3,368 7,638 651 - 254 4,273 694

V0 - Natural 223 2,060 19.2 1,850 11,966 438 - - 2,289 966 V1 - Exh Fan 295 2,049 19.0 1,946 12,966 479 203 - 2,628 1,062

V2 - CFIS 143 2,073 19.2 1,954 12,919 588 - - 2,542 1,050 V3 - HRV 192 2,102 19.2 2,014 12,759 594 - 254 2,863 1,074 V4 - ERV 85 2,060 19.1 1,917 12,839 582 - 254 2,753 1,067

V0 - Natural - 2,511 19.6 2,152 8,964 403 - - 2,555 626 V1 - Exh Fan 3 2,488 19.4 2,258 10,124 448 203 - 2,909 709

V2 - CFIS 4 2,484 19.6 2,241 10,301 554 - - 2,795 705 V3 - HRV 7 2,524 19.6 2,311 10,124 561 - 254 3,125 730 V4 - ERV

V0 - Natural - 1,801 20.5 1,457 16,292 466 - - 1,923 713 V1 - Exh Fan - 1,767 20.3 1,492 18,107 510 203 - 2,206 797

V2 - CFIS - 1,762 20.4 1,487 18,004 614 - - 2,101 786 V3 - HRV 3 1,795 20.4 1,539 17,889 619 - 254 2,412 806 V4 - ERV

Indianapolis

System 3, HERS 100

Orlando

Miami

Houston

Atlanta

Nashville

63

The breakdown of energy use for each component is shown for the different cooling systems in Figure 26 for Miami. The electric use of the cooling unit, AHU Fan, exhaust fan, and HRV are shown separately. The CFIS and ERV options operate the AHU fan to provide ventilation air distribution in the space so AHU fan power use is greater. The two-speed and 14.5 SEER cooling units have ECM fan motors so the energy penalty for mixing with the AHU fan is smaller. Figure 27 compares the total costs, including natural gas costs for space heating, for the different ventilation options. A different plot is given for each cooling system. The lowest cost option is typically the exhaust fan since this does not require additional AHU fan operation. With the two-speed cooling, low speed fan power is much lower so the CFIS case has lower net costs in some cases. The HRV/ERV does not provide net savings in the humid climates; however, it does provide heating savings in the colder climates. In Atlanta – where both the HRV and ERV were simulated – costs are slightly lower for the ERV. Figure 28 compares the impact of each ventilation option on the high humidity levels. The annual number of hours exceeding 60% RH for the no ventilation case was the lowest for Miami but the highest for Orlando for the no ventilation case, though this scenario does not satisfy ASHRAE 62.2 ventilation requirements. The ERV results in higher humidity levels in Orlando, Miami and Houston, as expected based on previous studies (Henderson et al 2007). In those climates, the ERV is ineffective in reducing high humidity hours since most of the high humidity hours occur at part load conditions – when there is little or no difference between indoor and outdoor humidity levels to drive moisture exchange in the ERV. Also, in winter for Orlando, Miami, and Houston, ERV operation has the disbenefit of keeping moisture inside the house at times when drier outside air might have helped reduce indoor high humidity. In Atlanta, the ERV shows the lowest high humidity hours of all options, even slightly reducing humidity levels compared to the HRV. The CFIS option slightly reduced high humidity hours compared to exhaust ventilation in Orlando, and more so in Atlanta. The CFIS option slightly increased high humidity hours in Miami and Houston, compared to an exhaust fan, because it provided more fresh air and because the part-time off-cycle operation of the AHU fan sometimes resulted in increased evaporation from the cooling coil (Shirey et al 2006).

64

Figure 26. Comparing Total Electric Use for Ventilation Options with Different Systems

‐

1,000 

2,000 

3,000 

4,000 

5,000 

6,000 

7,000 

8,000 

9,000 

10,000 

V0 ‐ Natural V1 ‐ Exh Fan V2 ‐ CFIS V3 ‐ HRV V4 ‐ ERV V5 ‐ ERV ind

Total Electric (kWh)

Miami, System 1, HERS 100

AC AHU Fan Exh Fan HRV Fan

‐

1,000 

2,000 

3,000 

4,000 

5,000 

6,000 

7,000 

V0 ‐ Natural V1 ‐ Exh Fan V2 ‐ CFIS V3 ‐ HRV V4 ‐ ERV V5 ‐ ERV ind

Total Electric (kWh)

Miami, System 1, HERS 85

AC AHU Fan Exh Fan HRV Fan

‐

1,000 

2,000 

3,000 

4,000 

5,000 

6,000 

V0 ‐ Natural V1 ‐ Exh Fan V2 ‐ CFIS V3 ‐ HRV V4 ‐ ERV V5 ‐ ERV ind

Total Electric (kWh)

Miami, System 3, HERS 100

AC AHU Fan Exh Fan HRV Fan

13 SEER AC

Two-Speed Cooling

14.5 SEER, ECM

65

Figure 27. Comparing Total Costs for Different Ventilation Options in each City

‐

200 

400 

600 

800 

1,000 

1,200 

1,400 

Z0‐Orlando Z1‐Miami Z2‐Houston Z3‐Atlanta Z4‐Nashville Z5‐Indianapolis

Total Costs w Furnace ($) 

HERS 100 ‐ System 1

V0‐Natural V1‐EXH Fan V2‐CFIS V3‐HRV V4 ‐ ERV

‐

100 

200 

300 

400 

500 

600 

700 

800 

Z0‐Orlando Z1‐Miami Z2‐Houston Z3‐Atlanta Z4‐Nashville Z5‐Indianapolis

Total Costs w Furnace ($) 

HERS 85 ‐ System 1

V0‐Natural V1‐EXH Fan V2‐CFIS V3‐HRV V4 ‐ ERV

‐

200 

400 

600 

800 

1,000 

1,200 

Z0‐Orlando Z1‐Miami Z2‐Houston Z3‐Atlanta Z4‐Nashville Z5‐Indianapolis

Total Costs w Furnace ($) 

HERS 100 ‐ System 3

V0‐Natural V1‐EXH Fan V2‐CFIS V3‐HRV V4 ‐ ERV

13 SEER AC

Two-Speed

14.5 SEER, ECM

66

Figure 28. Comparing High Humidity Levels for Different Ventilation Options in each City

‐

200 

400 

600 

800 

1,000 

1,200 

1,400 

1,600 

1,800 

2,000 

Z0‐Orlando Z1‐Miami Z2‐Houston Z3‐Atlanta Z4‐Nashville Z5‐Indianapolis

Hours Above

 60% RH 

HERS 100 ‐ System 1

V0‐Natural V1‐EXH Fan V2‐CFIS V3‐HRV V4 ‐ ERV

‐

200 

400 

600 

800 

1,000 

1,200 

1,400 

1,600 

Z0‐Orlando Z1‐Miami Z2‐Houston Z3‐Atlanta Z4‐Nashville Z5‐Indianapolis

Hours Above

 60% RH 

HERS 85 ‐ System 1

V0‐Natural V1‐EXH Fan V2‐CFIS V3‐HRV V4 ‐ ERV

‐

200 

400 

600 

800 

1,000 

1,200 

1,400 

1,600 

1,800 

2,000 

Z0‐Orlando Z1‐Miami Z2‐Houston Z3‐Atlanta Z4‐Nashville Z5‐Indianapolis

Hours Above

 60% RH 

HERS 100 ‐ System 3

V0‐Natural V1‐EXH Fan V2‐CFIS V3‐HRV V4 ‐ ERV

13 SEER AC

Two-Speed

14.5 SEER, ECM

67

2.4.12 Impact of Ventilation Rate (50% and 150% of ASHRAE 62.2 Requirements) The section above shows results for various ventilation system options that meet the requirements of ASHRAE 62.2-2010. The tables below show the impact of providing 50% above and 50% below the normally required ventilation rate. Table 21 shows results for the Exhaust Fan, Table 22 shows the CFIS options, and Table 23 shows the results for HRV cases for the HERS 100 house. Table 24, Table 25, and Table 26 show the same results for the HERS 70 house (with constant speed cooling). As expected more ventilation increases energy costs by 2 to 6% and less ventilation has the opposite impact. Generally, more ventilation increases the number of hours over 60% RH. Orlando was an exception to that trend.

Table 21. Performance Results with Different Ventilation Rates, HERS 100, Exhaust Fan

Hours Above

60% RH

AC Runtime

(hrs) AC EER

(Btu/Wh)

AC Energy

(kWh)

Htg Energy

(kWh)

AHU Fan Energy

(kWh)

Exh Fan

Energy (kWh)

HRV Energy

(kWh)

Total Electric w/o HT

(kWh)

Total Costs w

Furnace ($)

50% Vent 1,688 1,797 15.3 4,124 5,433 1,138 91 - 5,354 931 95%100% Vent 1,645 1,860 15.4 4,302 5,642 1,179 203 - 5,684 980 100%150% Vent 1,742 1,923 15.5 4,478 5,892 1,220 305 - 6,003 1,030 105%

50% Vent 1,188 2,614 15.2 5,992 1,441 1,521 91 - 7,604 872 95%100% Vent 1,303 2,700 15.3 6,232 1,513 1,572 203 - 8,007 917 100%150% Vent 1,425 2,788 15.4 6,474 1,609 1,625 305 - 8,404 964 105%

50% Vent 580 1,909 14.9 4,405 7,348 1,246 91 - 5,743 807 95%100% Vent 628 1,966 15.0 4,572 7,687 1,286 203 - 6,061 849 100%150% Vent 731 2,025 15.1 4,741 8,035 1,328 305 - 6,374 890 105%

50% Vent 240 1,312 15.1 2,482 12,437 858 91 - 3,431 1,112 95%100% Vent 291 1,338 15.2 2,546 13,004 881 203 - 3,631 1,168 100%150% Vent 388 1,364 15.3 2,612 13,506 903 305 - 3,820 1,218 104%

50% Vent - 1,527 15.3 2,863 9,515 902 91 - 3,856 775 94%100% Vent 4 1,555 15.4 2,938 10,150 928 203 - 4,069 823 100%150% Vent 42 1,579 15.4 3,005 10,750 950 305 - 4,260 866 105%

50% Vent - 1,049 15.8 1,933 17,145 827 91 - 2,851 814 95%100% Vent - 1,057 15.8 1,960 18,145 850 203 - 3,013 862 100%150% Vent 6 1,061 15.8 1,982 19,070 870 305 - 3,158 905 105%

Exh Vent, HERS 100

Nashville

Indianapolis

Orlando

Miami

Houston

Atlanta

68

Table 22. Performance Results with Different Ventilation Rates, HERS 100, CFIS

Table 23. Performance Results with Different Ventilation Rates, HERS 100, ERV

Hours Above

60% RH

AC Runtime

(hrs) AC EER

(Btu/Wh)

AC Energy

(kWh)

Htg Energy

(kWh)

AHU Fan Energy

(kWh)

Exh Fan

Energy (kWh)

HRV Energy

(kWh)

Total Electric w/o HT

(kWh)

Total Costs w

Furnace ($)

50% Vent 1,728 1,813 15.4 4,190 5,260 1,796 - - 5,986 983 96%100% Vent 1,627 1,889 15.5 4,406 5,543 1,837 - - 6,242 1,029 100%150% Vent 1,612 1,961 15.6 4,608 5,897 1,874 - - 6,482 1,079 105%

50% Vent 1,367 2,615 15.3 6,030 1,388 2,039 - - 8,069 915 96%100% Vent 1,394 2,721 15.4 6,330 1,474 2,091 - - 8,422 956 100%150% Vent 1,487 2,821 15.5 6,610 1,572 2,142 - - 8,752 997 104%

50% Vent 744 1,912 15.0 4,441 7,188 1,851 - - 6,292 847 95%100% Vent 728 1,983 15.1 4,646 7,666 1,890 - - 6,536 888 100%150% Vent 801 2,047 15.2 4,834 8,216 1,927 - - 6,761 931 105%

50% Vent 63 1,319 15.2 2,506 12,186 1,423 - - 3,929 1,148 96%100% Vent 163 1,355 15.3 2,598 12,864 1,445 - - 4,043 1,202 100%150% Vent 281 1,387 15.4 2,680 13,669 1,467 - - 4,146 1,261 105%

50% Vent - 1,521 15.3 2,868 9,411 1,448 - - 4,316 815 95%100% Vent 6 1,556 15.4 2,962 10,225 1,471 - - 4,433 861 100%150% Vent 50 1,586 15.5 3,044 11,163 1,492 - - 4,536 911 106%

50% Vent - 1,047 15.8 1,937 16,900 1,348 - - 3,285 840 95%100% Vent - 1,060 15.8 1,980 17,938 1,366 - - 3,346 880 100%150% Vent 10 1,073 15.9 2,019 19,427 1,388 - - 3,408 937 106%

Indianapolis

CFIS, HERS 100

Nashville

Atlanta

Orlando

Miami

Houston

Hours Above

60% RH

AC Runtime

(hrs) AC EER

(Btu/Wh)

AC Energy

(kWh)

Htg Energy

(kWh)

AHU Fan Energy

(kWh)

Exh Fan

Energy (kWh)

HRV Energy

(kWh)

Total Electric w/o HT

(kWh)

Total Costs w

Furnace ($)

50% Vent 1,803 1,819 15.4 4,213 5,238 2,231 - 114 6,557 1,039 97%100% Vent 1,742 1,850 15.4 4,300 5,312 2,243 - 254 6,797 1,068 100%150% Vent 1,760 1,883 15.5 4,390 5,399 2,257 - 381 7,027 1,098 103%

50% Vent 1,751 2,620 15.3 6,057 1,381 2,417 - 114 8,588 967 97%100% Vent 1,737 2,663 15.4 6,178 1,396 2,434 - 254 8,866 996 100%150% Vent 1,719 2,706 15.4 6,296 1,412 2,451 - 381 9,128 1,023 103%

50% Vent 829 1,914 15.0 4,454 7,262 2,274 - 114 6,841 896 97%100% Vent 793 1,943 15.0 4,536 7,391 2,286 - 254 7,076 922 100%150% Vent 803 1,970 15.1 4,616 7,547 2,298 - 381 7,295 947 103%

50% Vent 27 1,319 15.2 2,502 12,359 1,790 - 114 4,406 1,208 97%100% Vent 62 1,335 15.2 2,542 12,556 1,796 - 254 4,593 1,239 100%150% Vent 99 1,351 15.2 2,581 12,746 1,803 - 381 4,765 1,269 102%

Orlando

Miami

Houston

Atlanta

ERV, HERS 100

69

Table 24. Performance Results with Different Ventilation Rates, HERS 70, Exhaust Fan

Table 25. Performance Results with Different Ventilation Rates, HERS 70, CFIS

Hours Above

60% RH

AC Runtime

(hrs) AC EER

(Btu/Wh)

AC Energy

(kWh)

Htg Energy

(kWh)

AHU Fan Energy

(kWh)

Exh Fan

Energy (kWh)

HRV Energy

(kWh)

Total Electric w/o HT

(kWh)

Total Costs w

Furnace ($)

50% Vent 678 1,774 16.6 3,093 554 594 91 - 3,778 420 91%100% Vent 606 1,851 16.7 3,270 686 622 203 - 4,095 462 100%150% Vent 858 1,915 16.8 3,419 862 646 305 - 4,370 502 109%

50% Vent 159 2,981 16.6 4,181 25 783 91 - 5,055 512 93%100% Vent 433 3,122 16.7 4,430 42 820 203 - 5,453 554 100%150% Vent 763 3,247 16.7 4,647 55 853 305 - 5,805 590 107%

50% Vent 164 2,307 16.3 3,241 1,792 635 91 - 3,967 413 92%100% Vent 214 2,391 16.4 3,403 2,045 661 203 - 4,267 449 100%150% Vent 288 2,461 16.5 3,537 2,522 687 305 - 4,529 492 109%

50% Vent 27 1,525 16.5 2,098 4,126 468 91 - 2,657 520 91%100% Vent 15 1,558 16.6 2,167 4,612 485 203 - 2,855 569 100%150% Vent 40 1,577 16.6 2,214 5,251 500 305 - 3,019 624 110%

50% Vent - 1,550 16.5 2,137 5,376 495 91 - 2,724 486 91%100% Vent - 1,586 16.6 2,215 5,989 515 203 - 2,933 531 100%150% Vent 16 1,606 16.6 2,266 6,854 534 305 - 3,105 584 110%

50% Vent - 1,050 17.0 1,426 11,462 464 91 - 1,981 541 92%100% Vent - 1,055 17.0 1,451 12,419 481 203 - 2,135 585 100%150% Vent 4 1,053 17.0 1,463 13,555 499 305 - 2,266 634 108%

Exh Vent, HERS 70

Nashville

Indianapolis

Orlando

Miami

Houston

Atlanta

Hours Above

60% RH

AC Runtime

(hrs) AC EER

(Btu/Wh)

AC Energy

(kWh)

Htg Energy

(kWh)

AHU Fan Energy

(kWh)

Exh Fan

Energy (kWh)

HRV Energy

(kWh)

Total Electric w/o HT

(kWh)

Total Costs w

Furnace ($)

50% Vent 861 1,774 16.7 3,135 606 988 - - 4,123 459 93%100% Vent 745 1,856 16.8 3,322 780 1,013 - - 4,335 493 100%150% Vent 882 1,931 16.9 3,490 974 1,037 - - 4,527 526 107%

50% Vent 438 2,999 16.6 4,246 30 999 - - 5,245 532 94%100% Vent 602 3,156 16.8 4,524 55 1,036 - - 5,560 565 100%150% Vent 793 3,296 16.9 4,772 99 1,070 - - 5,841 597 106%

50% Vent 300 2,307 16.4 3,275 1,921 922 - - 4,197 438 92%100% Vent 255 2,404 16.5 3,455 2,369 949 - - 4,404 475 100%150% Vent 350 2,486 16.6 3,609 2,800 974 - - 4,583 508 107%

50% Vent - 1,516 16.6 2,106 4,307 814 - - 2,919 558 92%100% Vent - 1,555 16.7 2,186 4,992 830 - - 3,016 608 100%150% Vent 32 1,588 16.7 2,256 5,763 847 - - 3,103 663 109%

50% Vent - 1,539 16.6 2,145 5,612 823 - - 2,968 519 91%100% Vent - 1,579 16.7 2,229 6,537 844 - - 3,073 567 100%150% Vent 16 1,613 16.7 2,299 7,472 862 - - 3,162 614 108%

50% Vent - 1,040 17.0 1,426 11,615 781 - - 2,207 564 92%100% Vent - 1,053 17.0 1,462 13,012 798 - - 2,260 615 100%150% Vent 4 1,064 17.1 1,493 14,531 825 - - 2,318 671 109%

Indianapolis

CFIS, HERS 70

Nashville

Atlanta

Orlando

Miami

Houston

70

Table 26. Performance Results with Different Ventilation Rates, HERS 70, ERV

Hours Above

60% RH

AC Runtime

(hrs) AC EER

(Btu/Wh)

AC Energy

(kWh)

Htg Energy

(kWh)

AHU Fan Energy

(kWh)

Exh Fan

Energy (kWh)

HRV Energy

(kWh)

Total Electric w/o HT

(kWh)

Total Costs w

Furnace ($)

50% Vent 1,179 1,740 16.7 3,078 576 1,242 - 114 4,433 488 95%100% Vent 984 1,777 16.8 3,159 622 1,250 - 254 4,663 515 100%150% Vent 927 1,812 16.8 3,237 656 1,258 - 381 4,875 538 105%

50% Vent 721 2,934 16.6 4,151 30 1,153 - 114 5,417 549 95%100% Vent 770 3,002 16.7 4,271 36 1,166 - 254 5,690 577 100%150% Vent 827 3,068 16.7 4,384 41 1,179 - 380 5,944 603 105%

50% Vent 533 2,260 16.3 3,204 1,891 1,098 - 114 4,416 455 95%100% Vent 415 2,304 16.4 3,283 1,989 1,108 - 254 4,645 479 100%150% Vent 397 2,345 16.5 3,359 2,098 1,117 - 381 4,856 502 105%

50% Vent - 1,490 16.5 2,061 4,267 1,009 - 114 3,184 582 95%100% Vent - 1,512 16.6 2,103 4,438 1,015 - 254 3,372 612 100%150% Vent - 1,532 16.6 2,142 4,631 1,020 - 381 3,542 641 105%

Orlando

Miami

Houston

Atlanta

ERV, HERS 70

71

2.4.13 Single Speed, Two Speed, and Variable Speed Air Conditioners Three different cooling unit types were considered:

System 1. Single Speed Cooling Unit (SEER 13, 0.5 W/cfm fan)

System 3. Two-Speed Cooling Unit (SEER 17.7, ECM fan, 50% capacity at low speed)

System 4. Variable Speed Cooling Unit (0.1 W/cfm, 33% capacity at lowest speed, ductless)

System4L. Variable Speed Cooling Unit with lower airflow at low speed (275 cfm/ton) 

The data for these systems is summarized in Table 27. The two-speed and variable-speed cooling units have commensurately higher EERs as expected. The runtime of the cooling units also increases as expected because the units spend significant amounts of time at the lowest speed (the variable-speed cooling with 33% minimum capacity has even greater runtime). The degree of humidity control is similar for single-speed and two-speed cooling units. This result is not surprising given that the airflow per actual ton for the two-speed is similar in both low and high speed. Therefore the latent removal fraction is similar at all conditions. The ductless variable-speed system results in higher humidity levels in the HERS 100 house compared to the single and two-speed systems that have the ducts located in the attic. In the HERS 50 house, all the cooling units have similar humidity control performance since all the systems do not have ducts in the attic. These results confirm the finding above that locating the ducts in the conditioned space (and eliminating duct leakage) tends to increase humidity levels. It is often stated that two-speed systems provide better humidity control because the cooling coil runs longer. This is not necessarily true, unless the airflow per ton is lower at low speed. The “Variable Spd – Lower flw” option in the table does actually provide better humidity control than the other options since the low speed airflow is 275 cfm/ton. For the HERS 50 case (where all ducts are in the conditioned space) the trends are clearer, and the low airflow, variable speed system shows the lowest number of hours over 60% RH.

72

Table 27. Performance Results for Different Cooling Units and Climates (for HERS 100 and HERS 50)

Hours Above

60% RH

AC Runtime

(hrs) AC EER

(Btu/Wh)

AC Energy

(kWh)

Htg Energy

(kWh)

AHU Fan Energy

(kWh)

Exh Fan Energy

(kWh)

Total Electric w/o HT

(kWh)

Total Electric w/o HT

(kWh)

Total Costs w

Furnace ($)

Constant Speed 1,645 1,860 15.4 4,302 5,642 1,179 203 5,684 5,684 980 Two-Speed 1,635 3,012 19.8 3,147 5,676 511 203 3,862 3,862 798

Variable Spd Ductless 1,874 3,218 22.1 2,026 4,540 258 203 2,487 2,487 578 Var Spd - Lower flw 1,805 3,308 21.7 2,070 4,541 249 203 2,523 2,523 581

Constant Speed 1,303 2,700 15.3 6,232 1,513 1,572 203 8,007 8,007 917 Two-Speed 1,313 4,386 19.8 4,544 1,491 497 203 5,244 5,244 637

Variable Spd Ductless 1,538 4,678 22.0 2,948 1,189 186 203 3,337 3,337 422 Var Spd - Lower flw 1,382 4,804 21.6 3,012 1,190 172 203 3,387 3,387 428

Constant Speed 628 1,966 15.0 4,572 7,687 1,286 203 6,061 6,061 849 Two-Speed 620 3,078 18.9 3,452 7,659 553 203 4,209 4,209 690

Variable Spd Ductless 739 3,282 21.2 2,227 6,168 254 203 2,684 2,684 496 Var Spd - Lower flw 684 3,370 20.8 2,272 6,166 244 203 2,719 2,719 499

Constant Speed 291 1,338 15.2 2,546 13,004 881 203 3,631 3,631 1,168 Two-Speed 295 2,049 19.0 1,946 12,966 479 203 2,628 2,628 1,062

Variable Spd Ductless 329 2,143 21.3 1,217 10,256 261 203 1,681 1,681 799

Constant Speed 4 1,555 15.4 2,938 10,150 928 203 4,069 4,069 823 Two-Speed 3 2,488 19.4 2,258 10,124 448 203 2,909 2,909 709

Variable Spd Ductless 4 2,636 22.0 1,400 7,853 224 203 1,827 1,827 508

Constant Speed - 1,057 15.8 1,960 18,145 850 203 3,013 3,013 862 Two-Speed - 1,767 20.3 1,492 18,107 510 203 2,206 2,206 797

Variable Spd Ductless - 1,867 22.8 928 13,834 317 203 1,449 1,449 591 Indianapolis

HERS 100

Orlando

Miami

Houston

Atlanta

Nashville

Hours Above

60% RH

AC Runtime

(hrs) AC EER

(Btu/Wh)

AC Energy

(kWh)

Htg Energy

(kWh)

AHU Fan Energy

(kWh)

Exh Fan Energy

(kWh)

Total Electric w/o HT

(kWh)

Total Electric w/o HT

(kWh)

Total Costs w

Furnace ($)

Constant Speed 975 1,801 16.7 2,540 635 483 203 3,226 3,226 368 Two-Speed 1,011 3,266 21.3 1,869 633 140 203 2,211 2,211 265

Variable Spd Ductless 976 4,170 22.7 1,710 625 94 203 2,007 2,007 244 Var Spd - Lower flw 880 4,294 22.2 1,751 625 86 203 2,040 2,040 247

Constant Speed 777 2,416 16.6 3,419 39 635 203 4,257 4,257 433 Two-Speed 822 4,378 21.0 2,516 38 173 203 2,893 2,893 295

Variable Spd Ductless 787 5,494 22.3 2,316 38 114 203 2,633 2,633 268 Var Spd - Lower flw 670 5,642 21.9 2,371 38 103 203 2,678 2,678 273

Constant Speed 362 1,849 16.3 2,623 1,766 514 203 3,341 3,341 354 Two-Speed 380 3,252 20.4 1,963 1,758 181 203 2,347 2,347 269

Variable Spd Ductless 371 3,984 21.4 1,841 1,753 119 203 2,164 2,164 253 Var Spd - Lower flw 330 4,091 21.0 1,881 1,752 112 203 2,196 2,196 256

Constant Speed 35 1,120 16.5 1,554 3,369 349 203 2,106 2,106 405 Two-Speed 40 2,054 21.1 1,149 3,370 135 203 1,488 1,488 341

Variable Spd Ductless 35 2,646 22.4 1,060 3,367 108 203 1,371 1,371 329

Constant Speed - 1,206 16.5 1,679 4,254 386 203 2,269 2,269 384 Two-Speed - 2,184 21.0 1,248 4,254 160 203 1,611 1,611 320

Variable Spd Ductless - 2,734 22.1 1,167 4,254 128 203 1,499 1,499 309

Constant Speed - 798 16.9 1,095 8,526 349 203 1,647 1,647 398 Two-Speed - 1,457 21.5 812 8,523 198 203 1,213 1,213 364

Variable Spd Ductless - 1,870 22.9 749 8,513 178 203 1,130 1,130 357 Indianapolis

HERS 50 (Ducts in Space)

Orlando

Miami

Houston

Atlanta

Nashville

73

2.4.14 Enhanced Air Conditioner Control Options Several control strategies are available to enhance the moisture removal performance of conventional air conditioners. For instance, lowering the airflow tends to reduce the evaporator coil temperature and decrease the sensible heat ratio of the cooling unit. Another potential strategy is to continue cooling the space below the set point when humidity levels are high. This overcooling strategy can result in occupant discomfort and in extreme cases it can cause even higher relative humidity. Typically, the runtime of the air conditioner is also limited to avoid coil icing. These strategies are often implemented only when humidity levels are high. The combined strategies listed below:

1. Lower the airflow by 53% from 375 cfm/ton to 200 cfm/ton when the space humidity is 

above the RH setpoint.    

2. Reset the space cooling set point down by as much as 2°F as the space humidity 

increases by 10% RH above the set point.  Limit cooling operation to no more than 50% 

of each hour while overcooling is occurring. 

These two strategies together are called an “enhanced” cooling unit (System 2) for this study. The performance results for the conventional and enhanced cooling unit are compared in Table 28 for both the single-speed and two-speed units. Activation set points of 50% RH and 60% RH were used. Enhanced cooling operation has a significant impact on humidity control as long as the activation set point is 50% RH. The number of hours above 60% RH was reduced by a factor of 2 to 3 depending on the climate for both the conventional and two-speed cooling systems. If the activation set point is set to 60% RH, there is almost no impact on the number hours above 60% RH. Clearly these more “passive” humidity control approaches must be activated at lower humidity levels to be effective.

74

Table 28. Performance Results with Enhanced Cooling Unit

Hours Above

60% RH

AC Runtime

(hrs) AC EER

(Btu/Wh)

AC Energy

(kWh)

Htg Energy

(kWh)

AHU Fan Energy

(kWh)

Total Electric w/o HT

(kWh)

Total Costs w

Furnace ($)

Conv AC 1,645 1,860 15.4 4,302 5,642 1,179 5,684 980 Enhanced AC (60% RH) 1,571 1,868 15.4 4,317 5,737 1,248 5,769 995 Enhanced AC (50% RH) 928 2,038 15.2 4,570 5,841 1,144 5,917 1,017

Conv AC 1,303 2,700 15.3 6,232 1,513 1,572 8,007 917 Enhanced AC (60% RH) 1,205 2,714 15.3 6,256 1,517 1,570 8,029 920 Enhanced AC (50% RH) 471 2,881 15.2 6,513 1,588 1,478 8,195 942

Conv AC 628 1,966 15.0 4,572 7,687 1,286 6,061 849 Enhanced AC (60% RH) 604 1,969 15.0 4,577 7,693 1,286 6,066 849 Enhanced AC (50% RH) 231 2,053 15.0 4,707 7,745 1,245 6,155 859

Conv AC 291 1,338 15.2 2,546 13,004 881 3,631 1,168 Enhanced AC (60% RH) 291 1,338 15.2 2,546 13,004 881 3,631 1,168 Enhanced AC (50% RH) 183 1,359 15.2 2,576 13,006 877 3,656 1,170

Conv AC 4 1,555 15.4 2,938 10,150 928 4,069 823 Enhanced AC (60% RH) 4 1,555 15.4 2,938 10,150 928 4,069 823 Enhanced AC (50% RH) 2 1,563 15.4 2,947 10,149 922 4,072 823

Conv AC - 1,057 15.8 1,960 18,145 850 3,013 862 Enhanced AC (60% RH) - 1,057 15.8 1,960 18,145 850 3,013 862 Enhanced AC (50% RH) - 1,060 15.8 1,964 18,145 847 3,015 862

Indianapolis

HERS 100, Conv AC

Orlando

Miami

Houston

Atlanta

Nashville

Hours Above

60% RH

AC Runtime

(hrs) AC EER

(Btu/Wh)

AC Energy

(kWh)

Htg Energy

(kWh)

AHU Fan Energy

(kWh)

Total Electric w/o HT

(kWh)

Total Costs w

Furnace ($)

Conv AC 575 3,491 21.0 2,615 673 217 3,036 354 Enhanced AC (60% RH) 571 3,493 21.0 2,617 673 217 3,037 354 Enhanced AC (50% RH) 308 3,826 20.8 2,756 699 196 3,155 368

Conv AC 382 5,303 20.0 3,629 42 360 4,192 426 Enhanced AC (60% RH) 337 5,313 20.0 3,633 42 360 4,196 427 Enhanced AC (50% RH) 20 5,721 19.9 3,772 48 339 4,315 439

Conv AC 191 3,844 19.3 2,856 2,034 350 3,410 376 Enhanced AC (60% RH) 188 3,844 19.3 2,856 2,034 350 3,410 376 Enhanced AC (50% RH) 53 4,016 19.3 2,914 2,045 342 3,459 381

Conv AC 15 2,604 19.7 1,815 4,594 271 2,289 510 Enhanced AC (60% RH) 15 2,604 19.7 1,815 4,594 271 2,289 510 Enhanced AC (50% RH) 15 2,619 19.7 1,819 4,594 270 2,292 510

Conv AC - 2,641 19.7 1,859 5,975 299 2,361 475 Enhanced AC (60% RH) - 2,641 19.7 1,859 5,973 299 2,361 475 Enhanced AC (50% RH) - 2,659 19.7 1,863 5,973 297 2,364 475

Conv AC - 1,828 20.5 1,206 12,390 327 1,736 553 Enhanced AC (60% RH) - 1,828 20.5 1,206 12,390 327 1,736 553 Enhanced AC (50% RH) - 1,834 20.5 1,208 12,390 326 1,738 554

Indianapolis

HERS 70, 2-Spd

Orlando

Miami

Houston

Atlanta

Nashville

75

Figure 29 shows the impact that enhanced control has on space conditions. The 2°F of overcooling tends to “sweep” down (indicated by the thick black line) the high humidity days when the space temperature is just below the cooling set point. The only remaining high humidity days correspond to times when the space temperature is near the heating setpoint.

Conventional Cooling Enhanced Cooling – 50% RH Set Pt

Figure 29. Psychrometric Plots Showing Impact of Enhanced Control in Miami, HERS 100 House

System 10 and System 11 implement each portion of the enhancements separately for the HERS 100, Conventional cooling and the HERS 70, Two-Speed system in Houston. Overcooling alone (S11) provides slightly more humidity control benefit than lower airflow alone (S10). Though the two strategies combined do provide greater benefit than the individual strategies alone.

Table 29. Performance Results with Various Cooling Unit Enhancements

2.4.15 Cooling Systems with Further Enhancements Some cooling systems include additional enhancements such as air-to-air heat exchangers (e.g., heat pipes) and either full condensing reheat or subcooling reheat coils to further improve

Daily Indoor Space Conditions: z1h100s1rh50v1

60 65 70 75 80 85 90

Dry Bulb Temperature (F)

0.000

0.005

0.010

0.015

0.020

0.025

Hum

idity

Rat

io (

lb/lb

)

40%

50%

60%

70%

80%

Hours Above 50% = 5361Hours Above 55% = 2399Hours Above 60% = 1303Hours Above 65% = 543Hours Above 70% = 107All HrsCooling Hrs

Daily Indoor Space Conditions: z1h100s2rh50v1

60 65 70 75 80 85 90

Dry Bulb Temperature (F)

0.000

0.005

0.010

0.015

0.020

0.025

Hum

idity

Rat

io (

lb/lb

)

40%

50%

60%

70%

80%

Hours Above 50% = 3939Hours Above 55% = 1520Hours Above 60% = 471Hours Above 65% = 37Hours Above 70% = 0All HrsCooling Hrs

Hours Above

60% RH

AC Runtime

(hrs) AC EER

(Btu/Wh)

AC Energy

(kWh)

Htg Energy

(kWh)

AHU Fan Energy

(kWh)

Total Electric w/o HT

(kWh)

Total Costs w

Furnace ($)

S1 - Conv AC 628 1,966 15.0 4,572 7,687 1,286 6,061 849 S10 - Lower Airflow 476 2,006 14.9 4,617 7,699 1,227 6,048 848 S11 - Overcooling 326 2,037 15.1 4,713 7,766 1,328 6,245 868

S2 - Both 231 2,053 15.0 4,707 7,745 1,245 6,155 859

S1 - Conv AC 191 3,844 19.3 2,856 2,034 350 3,410 376 S10 - Lower Airflow 112 3,961 19.2 2,893 2,034 342 3,439 378 S11 - Overcooling 77 4,204 21.7 2,376 2,063 192 2,770 323

S2 - Both 53 4,016 19.3 2,914 2,045 342 3,459 381

HERS 100 Single Spd

HERS 70 2-Stage

Houston

Impact of Enhanced Control

76

dehumidification performance while in cooling. Other systems can provide dehumidification without any sensible cooling. The following systems fall into this category:

System 8. A cooling unit that uses a subcooling/reheat coil to provide condenser reheat while boosting refrigeration cycle efficiency with the additional refrigerant subcooling. The controls switch the unit into this low SHR mode when humidity is high and the space temperature drops below the cooling set point. Overcooling of the space, up to 2oF below the cooling set point, is allowed.

System 9. A cooling unit that uses modulating condenser reheat to maintain the supply air at space neutral conditions. The reheat mode is activated when humidity is high and the space drops below the cooling set point. The condensing temperature is slightly lower in this mode which slightly boosts efficiency. Since this system provides space neutral temperature supply air, by definition it does not overcool the space, but the system is allowed to operate (as a dehumidifier) when the space temperature is up to 5°F below the cooling set.

System 12. A cooling unit with a sensible heat exchanger (SHX) configured around the cooling coil to pre-cool incoming air and reheat air leaving the coil. The sensible heat exchanger increases pressure drop on the air side of the system.

Table 30 compares the results for the conventional cooling, the cooling with enhanced controls, and the cooling units with additional hardware enhancements in the HERS 100 house. These systems are all based on a constant speed (13 SEER) cooling unit. The cooling with a SHX performs only slightly better than the conventional cooling unit in terms of humidity control and has considerably higher operating costs. The cooling with enhanced controls provides better humidity control with lower operating costs. Table 31 compares the same systems in the HERS 70 house. In this case the base cooling unit is two-speed cooling (except for System 12, which uses the 14.5 SEER constant-speed system).

77

Table 30. Performance Results for Cooling Units with Further Enhancements, HERS 100

Hours Above

60% RH

AC Runtime

(hrs) AC EER

(Btu/Wh)

AC Energy

(kWh)

Htg Energy

(kWh)

AHU Fan Energy

(kWh)

Total Electric w/o HT

(kWh)

Total Costs w

Furnace ($)

Conv AC 1,645 1,860 15.4 4,302 5,642 1,179 5,684 980 100%Enhanced AC 928 2,038 15.2 4,570 5,841 1,144 5,917 1,017 104%

AC w/ HPs 1,162 2,234 14.8 4,808 5,649 1,423 6,434 1,056 108%AC w/ sc-rheat 205 2,320 12.1 5,840 5,836 1,394 7,437 1,171 120%

AC w/ full reheat - 2,455 13.8 5,202 6,022 1,463 6,868 1,126 115%

Conv AC 1,303 2,700 15.3 6,232 1,513 1,572 8,007 917 100%Enhanced AC 471 2,881 15.2 6,513 1,588 1,478 8,195 942 103%

AC w/ HPs 758 3,242 14.7 6,953 1,515 1,926 9,082 1,026 112%AC w/ sc-rheat 22 3,138 13.0 7,702 1,563 1,709 9,615 1,083 118%

AC w/ full reheat - 3,178 14.3 6,987 1,628 1,811 9,001 1,026 112%

Conv AC 628 1,966 15.0 4,572 7,687 1,286 6,061 849 100%Enhanced AC 231 2,053 15.0 4,707 7,745 1,245 6,155 859 101%

AC w/ HPs 338 2,348 14.5 5,050 7,694 1,536 6,790 911 107%AC w/ sc-rheat 9 2,211 13.2 5,397 7,742 1,365 6,965 928 109%

AC w/ full reheat - 2,250 14.3 4,997 7,786 1,450 6,651 903 106%

Conv AC 291 1,338 15.2 2,546 13,004 881 3,631 1,168 100%Enhanced AC 183 1,359 15.2 2,576 13,006 877 3,656 1,170 100%

AC w/ HPsAC w/ sc-rheat 54 1,417 14.2 2,772 13,014 903 3,879 1,194 102%

AC w/ full reheat 1 1,446 14.8 2,682 13,013 962 3,847 1,191 102%

Conv AC 4 1,555 15.4 2,938 10,150 928 4,069 823 100%Enhanced AC 2 1,563 15.4 2,947 10,149 922 4,072 823 100%

AC w/ HPsAC w/ sc-rheat - 1,596 14.8 3,061 10,145 938 4,202 835 102%

AC w/ full reheat - 1,607 15.2 3,004 10,144 1,010 4,218 837 102%

Conv AC - 1,057 15.8 1,960 18,145 850 3,013 862 100%Enhanced AC - 1,060 15.8 1,964 18,145 847 3,015 862 100%

AC w/ HPsAC w/ sc-rheat - 1,080 15.3 2,031 18,145 856 3,090 868 101%

AC w/ full reheat - 1,087 15.6 1,998 18,140 905 3,106 869 101%

HERS 100

Miami

Houston

Atlanta

Orlando

Nashville

Indianapolis

78

Table 31. Performance Results for Cooling Units with Further Enhancements, HERS 70 (two-speed)

Hours Above

60% RH

AC Runtime

(hrs) AC EER

(Btu/Wh)

AC Energy

(kWh)

Htg Energy

(kWh)

AHU Fan Energy

(kWh)

Total Electric w/o HT

(kWh)

Total Costs w

Furnace ($)

Conv AC 575 3,491 21.0 2,615 673 217 3,036 354 100%Enhanced AC 308 3,826 20.8 2,756 699 196 3,155 368 104%

AC w/ HPs 418 2,311 16.0 3,799 689 1,126 5,128 566 160%AC w/ sc-rheat 32 4,000 17.0 3,374 703 265 3,843 438 124%

AC w/ full reheat - 4,220 19.4 2,955 720 309 3,467 401 113%

Conv AC 382 5,303 20.0 3,629 42 360 4,192 426 100%Enhanced AC 20 5,721 19.9 3,772 48 339 4,315 439 103%

AC w/ HPs 46 3,846 16.0 5,064 42 1,482 6,749 685 161%AC w/ sc-rheat - 5,686 16.9 4,413 42 432 5,048 513 120%

AC w/ full reheat - 6,013 19.2 3,901 47 469 4,573 465 109%

Conv AC 191 3,844 19.3 2,856 2,034 350 3,410 376 100%Enhanced AC 53 4,016 19.3 2,914 2,045 342 3,459 381 101%

AC w/ HPs 93 2,918 15.9 3,837 2,044 1,157 5,197 528 141%AC w/ sc-rheat - 4,061 17.3 3,233 2,039 380 3,817 411 109%

AC w/ full reheat - 4,235 18.8 3,000 2,052 439 3,642 396 105%

Conv AC 15 2,604 19.7 1,815 4,594 271 2,289 510 100%Enhanced AC 15 2,619 19.7 1,819 4,594 270 2,292 510 100%

AC w/ HPsAC w/ sc-rheat - 2,662 19.2 1,879 4,595 274 2,356 517 101%

AC w/ full reheat - 2,698 19.5 1,856 4,595 331 2,390 520 102%

Conv AC - 2,641 19.7 1,859 5,975 299 2,361 475 100%Enhanced AC - 2,659 19.7 1,863 5,973 297 2,364 475 100%

AC w/ HPsAC w/ sc-rheat - 2,678 19.2 1,914 5,973 302 2,419 481 101%

AC w/ full reheat - 2,708 19.6 1,890 5,973 359 2,452 484 102%

Conv AC - 1,828 20.5 1,206 12,390 327 1,736 553 100%Enhanced AC - 1,834 20.5 1,208 12,390 326 1,738 554 100%

AC w/ HPsAC w/ sc-rheat - 1,856 20.0 1,239 12,390 328 1,771 556 100%

AC w/ full reheat - 1,870 20.4 1,225 12,390 367 1,795 558 101%

HERS 70

Miami

Houston

Atlanta

Orlando

Nashville

Indianapolis

79

2.4.16 Dehumidifiers The 50 pint per day stand-alone dehumidifier was simulated with a set point of 50% and 60% RH. The unit was more than adequate to maintain the set point of 60% RH (even though nearly 600 hours slightly exceeded that mark by 1-2 %RH). However, the unit was not fully able to maintain the space at the 50% RH set point (the result was closer to 55% RH). In general the DH unit tended to increase cooling energy requirements slightly while also decreasing heating requirements (at least in humid climates). Lowering the set point by 10% RH increased dehumidifier energy use by a factor of 5 or more. Total operating costs go up by about $100-200 per year compared to cooling only in the hot-humid climates.

Set Point = 60% RH Set Point = 50% RH

Figure 30. Psychrometric Charts Comparing the Degree of Humidity Control at Two RH Set Points for Standard DH unit, HERS 100, Miami

Daily Indoor Space Conditions: z1h100s5rh60v1

60 65 70 75 80 85 90

Dry Bulb Temperature (F)

0.000

0.005

0.010

0.015

0.020

0.025

Hum

idity

Rat

io (

lb/lb

)

40%

50%

60%

70%

80%

Hours Above 50% = 5508Hours Above 55% = 2186Hours Above 60% = 598Hours Above 65% = 15Hours Above 70% = 0All HrsCooling Hrs

Daily Indoor Space Conditions: z1h100s5rh50v1

60 65 70 75 80 85 90

Dry Bulb Temperature (F)

0.000

0.005

0.010

0.015

0.020

0.025

Hum

idity

Rat

io (

lb/lb

)

40%

50%

60%

70%

80%

Hours Above 50% = 2968Hours Above 55% = 168Hours Above 60% = 0Hours Above 65% = 0Hours Above 70% = 0All HrsCooling Hrs

80

Table 32. Performance Results with Standalone DH (System 5) with Different Set Points

Figure 31. Total Operating Costs for Standalone DH at Various RH Set Points, HERS 100 House

Hours Above

60% RH

AC Runtime

(hrs) AC EER

(Btu/Wh)

AC Energy

(kWh)

Htg Energy

(kWh)

AHU Fan Energy

(kWh)

DH Energy

(kWh)

Exh Fan Energy

(kWh)

Total Electric w/o HT

(kWh)

Total Costs w

Furnace ($)

Conv AC 1,645 1,860 15.4 4,302 5,642 1,179 - 203 5,684 980 Standalone DH 60% 664 1,872 15.4 4,332 5,430 1,435 321 203 6,291 1,026 Standalone DH 50% - 2,016 15.4 4,627 5,060 1,492 1,788 203 8,110 1,183

Conv AC 1,303 2,700 15.3 6,232 1,513 1,572 - 203 8,007 917 Standalone DH 60% 598 2,716 15.3 6,268 1,356 1,720 247 203 8,438 950 Standalone DH 50% - 2,849 15.3 6,540 1,214 1,778 1,428 203 9,949 1,092

Conv AC 628 1,966 15.0 4,572 7,687 1,286 - 203 6,061 849 Standalone DH 60% 301 1,974 15.0 4,592 7,608 1,463 129 203 6,387 873 Standalone DH 50% - 2,042 15.0 4,728 7,482 1,491 850 203 7,272 943

Conv AC 291 1,338 15.2 2,546 13,004 881 - 203 3,631 1,168 Standalone DH 60% 103 1,340 15.2 2,551 12,969 1,084 34 203 3,873 1,190 Standalone DH 50% - 1,354 15.2 2,573 12,893 1,087 292 203 4,155 1,215

Conv AC 4 1,555 15.4 2,938 10,150 928 - 203 4,069 823 Standalone DH 60% 4 1,550 15.4 2,932 10,202 1,120 - 203 4,255 843 Standalone DH 50% - 1,560 15.4 2,947 10,193 1,123 122 203 4,394 856

Conv AC - 1,057 15.8 1,960 18,145 850 - 203 3,013 862 Standalone DH 60% 2 1,052 15.8 1,954 18,216 1,031 - 203 3,188 878 Standalone DH 50% - 1,058 15.8 1,963 18,211 1,033 70 203 3,270 884

Indianapolis

HERS 100

Orlando

Miami

Houston

Atlanta

Nashville

‐

200 

400 

600 

800 

1,000 

1,200 

1,400 

Z0‐Orlando Z1‐Miami Z2‐Houston Z3‐Atlanta Z4‐Nashville Z5‐Indianapolis

Total Costs w Furnace ($) 

Dehumidifier Set Point

Conv AC DH 60% RH DH 50% RH

81

Table 33. Humidity Threshold Results for Standalone DH (System 5) with Different Set Points

Hours Above

60% RH

Hours Above

55% RH

Hours Above

50% RH

Conv AC 1,645 3,074 6,737 Standalone DH 60% 664 2,588 6,660 Standalone DH 50% - 314 3,428

Conv AC 1,303 2,399 5,361 Standalone DH 60% 598 2,186 5,508 Standalone DH 50% - 168 2,968

Conv AC 628 1,500 3,716 Standalone DH 60% 301 1,277 3,693 Standalone DH 50% - 51 1,835

Conv AC 291 722 1,479 Standalone DH 60% 103 531 1,313 Standalone DH 50% - 15 652

Conv AC 4 105 586 Standalone DH 60% 4 112 593 Standalone DH 50% - 1 287

Conv AC - 20 352 Standalone DH 60% 2 22 371 Standalone DH 50% - - 188

Indianapolis

HERS 100

Orlando

Miami

Houston

Atlanta

Nashville

82

2.4.17 Advanced Dehumidifiers Various types of high performance dehumidifiers were also simulated, including:

System 6. A high efficiency, 65 pint per day, ducted DH unit with EF of 2.0 L/kWh and fan power of about 0.7 W/cfm.

System 7. A high efficiency, 82 pint per day unit (2.0 L/kWh) set up to also provide ventilation to the home. The unit uses its fan to provide 174 cfm to the space year-round, with 33% of that airflow from outdoors.

System 13. A natural gas-fired desiccant unit installed in the AHU supply duct. The unit has a capacity of 145 pint per day and the airflow is 400 cfm. The unit fans only run when there is a call for dehumidification. Natural gas input is 10 MBtu/h.

System 14. A condenser-regenerated desiccant unit with a compressor and a DX coil. The unit has a capacity of 120 pint per day (2.6 L/kWh) and airflow of 300 cfm.  

Table 34 and Table 35 summarize the results for these DH systems and compare them to the conventional cooling option. Table 34 compares the different systems for the HERS 100 house and Table 35 compares them for the HERS 70 house. Figure 32 compares the operating costs for the HERS 100 house and Figure 33 compares costs for the HERS 70 house. Generally, the high efficiency ducted DH unit (System 6) and the condenser regenerated desiccant unit (System 14) have the lowest energy costs. The DH providing ventilation (System 7) typically is a bit more expensive due to the need to operate that fan continuously. The natural gas-fired desiccant is significantly more expensive to operate in part due to the higher cost of natural gas in humid climates (see Table 11). However, even with 50% lower natural gas costs these systems would still generally have higher operating costs.  

83

Table 34. Performance Results with Various DH Units (Systems, 5, 6, 7, 13 & 14) with HERS 100

Figure 32. Total Operating Costs All DH Options, 50% RH Set Point, HERS 100 House

Hours Above

55% RH

AC Runtime

(hrs) AC EER (Btu/Wh)

AC Energy

(kWh)

Htg Energy

(kWh)

AHU Fan Energy

(kWh)

DH Energy

(kWh)

DH Fan Energy

(kWh)

Exh Fan Energy

(kWh)

DH Gas Use

(therms)

Total Electric w/o HT

(kWh)

Total Costs w

Furnace ($)

S1 - Conv AC 3,074 1,860 15.4 4,302 5,642 1,179 - - 203 - 5,684 980 S5 - DH Unit 314 2,016 15.4 4,627 5,060 1,492 1,788 - 203 - 8,110 1,183

S6 - Ducted DH 219 1,996 15.3 4,582 5,097 1,483 1,366 281 203 - 7,916 1,166 S7 - Vent DH 3 1,960 15.3 4,479 4,626 1,451 867 1,067 - - 7,865 1,127

S13 - Gas-fired DES - 2,038 15.4 4,662 5,354 1,505 204 544 203 227 7,118 1,509 S14 - Cond DES - 1,818 15.3 4,173 5,429 1,406 1,707 260 203 - 7,749 1,173

S1 - Conv AC 2,399 2,700 15.3 6,232 1,513 1,572 - - 203 - 8,007 917 S5 - DH Unit 168 2,849 15.3 6,540 1,214 1,778 1,428 - 203 - 9,949 1,092

S6 - Ducted DH 116 2,829 15.3 6,496 1,226 1,769 1,084 221 203 - 9,773 1,075 S7 - Vent DH 5 2,803 15.3 6,403 1,079 1,746 699 1,067 - - 9,915 1,079

S13 - Gas-fired DES - 2,919 15.4 6,694 1,317 1,808 216 576 203 240 9,497 1,483 S14 - Cond DES - 2,672 15.3 6,139 1,377 1,696 1,326 201 203 - 9,565 1,065

S1 - Conv AC 1,500 1,966 15.0 4,572 7,687 1,286 - - 203 - 6,061 849 S5 - DH Unit 51 2,042 15.0 4,728 7,482 1,491 850 - 203 - 7,272 943

S6 - Ducted DH 35 2,033 15.0 4,708 7,494 1,487 648 134 203 - 7,180 935 S7 - Vent DH 6 2,033 15.0 4,680 7,115 1,476 428 1,067 - - 7,650 959

S13 - Gas-fired DES - 2,082 15.1 4,820 8,326 1,522 131 349 203 146 7,026 1,116 S14 - Cond DES - 1,953 15.0 4,527 7,590 1,450 770 117 203 - 7,067 930

S1 - Conv AC 722 1,338 15.2 2,546 13,004 881 - - 203 - 3,631 1,168 S5 - DH Unit 15 1,354 15.2 2,573 12,893 1,087 292 - 203 - 4,155 1,215

S6 - Ducted DH 14 1,352 15.2 2,570 12,902 1,087 229 48 203 - 4,137 1,214 S7 - Vent DH - 1,388 15.2 2,616 12,250 1,088 139 1,067 - - 4,910 1,253

S13 - Gas-fired DES - 1,331 15.2 2,549 13,594 1,085 27 72 203 30 3,936 1,280 S14 - Cond DES - 1,336 15.2 2,540 12,937 1,081 253 39 203 - 4,115 1,213

S1 - Conv AC 105 1,555 15.4 2,938 10,150 928 - - 203 - 4,069 823 S5 - DH Unit 1 1,560 15.4 2,947 10,193 1,123 122 - 203 - 4,394 856

S6 - Ducted DH - 1,559 15.4 2,945 10,187 1,122 95 20 203 - 4,384 855 S7 - Vent DH - 1,589 15.3 2,981 9,829 1,127 45 1,067 - - 5,221 921

S13 - Gas-fired DES - 1,522 15.4 2,901 11,065 1,117 32 85 203 35 4,338 924 S14 - Cond DES - 1,547 15.3 2,924 10,191 1,118 118 18 203 - 4,381 855

S1 - Conv AC 20 1,057 15.8 1,960 18,145 850 - - 203 - 3,013 862 S5 - DH Unit - 1,058 15.8 1,963 18,211 1,033 70 - 203 - 3,270 884

S6 - Ducted DH - 1,057 15.8 1,962 18,208 1,033 55 11 203 - 3,264 883 S7 - Vent DH - 1,097 15.8 2,018 17,590 1,035 20 1,067 - - 4,140 930

S13 - Gas-fired DES - 1,021 15.8 1,912 19,483 1,035 17 46 203 19 3,214 940 S14 - Cond DES - 1,050 15.8 1,947 18,210 1,030 69 10 203 - 3,260 883

Indianapolis

HERS 100, 50% RH Set Pt

Orlando

Miami

Houston

Atlanta

Nashville

‐

200 

400 

600 

800 

1,000 

1,200 

1,400 

1,600 

Z0‐Orlando Z1‐Miami Z2‐Houston Z3‐Atlanta Z4‐Nashville Z5‐Indianapolis

Total Costs w Furnace ($) 

Comparing DH Units, HERS 100

Conv AC Standalone DH Ducted DH Vent DH Gas‐fired DES Cond DES

84

Table 35. Performance Results with Various DH Units (Systems, 5, 6, 7, 13 & 14) with HERS 70

and 50% RH Setpoint

Hours Above

55% RH

AC Runtime

(hrs) AC EER (Btu/Wh)

AC Energy

(kWh)

Htg Energy

(kWh)

AHU Fan Energy

(kWh)

DH Energy

(kWh)

DH Fan Energy

(kWh)

Exh Fan Energy

(kWh)

DH Gas Use

(therms)

Total Electric w/o HT

(kWh)

Total Costs w

Furnace ($)

S3 - 2-Spd 1,904 3,491 21.0 2,615 673 217 - - 203 - 3,036 354 S5 - DH Unit - 3,718 21.0 2,765 552 250 1,086 - 203 - 4,304 474

S6 - Ducted DH - 3,701 21.0 2,746 570 247 821 167 203 - 4,184 463 S7 - Vent DH - 3,720 21.1 2,728 510 240 530 1,067 - - 4,565 497

S13 - Gas-fired DES - 3,642 21.1 2,686 819 245 105 280 203 116 3,518 621 S14 - Cond DES - 3,442 20.9 2,539 654 231 992 151 203 - 4,116 462

S3 - 2-Spd 1,429 5,303 20.0 3,629 42 360 - - 203 - 4,192 426 S5 - DH Unit - 5,563 20.0 3,778 35 387 823 - 203 - 5,191 527

S6 - Ducted DH - 5,543 20.0 3,760 35 385 623 125 203 - 5,096 517 S7 - Vent DH - 5,549 20.1 3,714 33 374 397 1,067 - - 5,552 563

S13 - Gas-fired DES - 5,490 20.1 3,709 53 378 83 222 203 93 4,596 634 S14 - Cond DES - 5,274 20.0 3,559 42 363 805 122 203 - 5,052 513

S3 - 2-Spd 732 3,844 19.3 2,856 2,034 350 - - 203 - 3,410 376 S5 - DH Unit - 3,951 19.3 2,912 2,063 378 461 - 203 - 3,955 424

S6 - Ducted DH - 3,939 19.3 2,904 2,075 378 344 70 203 - 3,900 419 S7 - Vent DH - 3,969 19.4 2,885 2,001 370 230 1,067 - - 4,553 472

S13 - Gas-fired DES - 3,860 19.3 2,864 2,644 383 50 133 203 56 3,634 481 S14 - Cond DES - 3,826 19.2 2,824 2,117 370 400 61 203 - 3,858 418

S3 - 2-Spd 240 2,604 19.7 1,815 4,594 271 - - 203 - 2,289 510 S5 - DH Unit - 2,601 19.7 1,812 4,699 301 104 - 203 - 2,421 530

S6 - Ducted DH - 2,600 19.7 1,812 4,700 301 82 17 203 - 2,415 529 S7 - Vent DH - 2,632 19.8 1,801 4,520 294 41 1,067 - - 3,203 599

S13 - Gas-fired DES - 2,434 19.5 1,742 5,195 306 6 16 203 7 2,273 554 S14 - Cond DES - 2,587 19.7 1,803 4,720 301 81 12 203 - 2,400 529

S3 - 2-Spd 33 2,641 19.7 1,859 5,975 299 - - 203 - 2,361 475 S5 - DH Unit - 2,641 19.7 1,855 6,120 328 71 - 203 - 2,457 490

S6 - Ducted DH - 2,640 19.7 1,855 6,126 328 55 11 203 - 2,452 490 S7 - Vent DH - 2,663 19.7 1,846 5,964 322 29 1,067 - - 3,263 562

S13 - Gas-fired DES - 2,493 19.5 1,795 6,737 334 10 26 203 11 2,368 518 S14 - Cond DES - 2,623 19.7 1,845 6,125 327 65 10 203 - 2,450 490

S3 - 2-Spd 11 1,828 20.5 1,206 12,390 327 - - 203 - 1,736 553 S5 - DH Unit - 1,822 20.5 1,201 12,571 356 49 - 203 - 1,809 565

S6 - Ducted DH - 1,821 20.5 1,201 12,571 356 39 8 203 - 1,808 565 S7 - Vent DH - 1,857 20.6 1,203 12,200 347 15 1,067 - - 2,632 617

S13 - Gas-fired DES - 1,701 20.3 1,147 13,429 366 5 14 203 6 1,735 594 S14 - Cond DES - 1,814 20.5 1,195 12,577 355 46 7 203 - 1,806 565

Indianapolis

HERS 70, 50% RH Set Pt

Orlando

Miami

Houston

Atlanta

Nashville

85

Figure 33. Total Operating Costs All DH Options, 50% RH Set Point, HERS 70 House

2.4.18 Comparing the Best Overall Technologies The best enhanced and dehumidifier technologies are compared in Table 36 and Figure 34 for the HERS 100 house and in Table 37 and Figure 35 for the HERS 70 house in the humid climates. Total operating costs for the Systems that provide full control of the humidity set point are typically a 10-30% cost premium compared to cooling alone. System 9, which has full condensing reheat control, has the lowest operating cost premium.

‐

100 

200 

300 

400 

500 

600 

700 

Z0‐Orlando Z1‐Miami Z2‐Houston Z3‐Atlanta Z4‐Nashville Z5‐Indianapolis

Total Costs w Furnace ($) 

Comparing DH Units, HERS 70

Conv AC Standalone DH Ducted DH Vent DH Gas‐fired DES Cond DES

86

Table 36. Performance Results with Best DH and Enhanced Units, HERS 100

Figure 34. Total Operating Costs Best Technologies, HERS 100 House

Hours Above

60% RH

AC Runtime

(hrs) AC EER

(Btu/Wh)

AC Energy

(kWh)

Htg Energy

(kWh)

AHU Fan Energy

(kWh)

DH Energy

(kWh)

DH Fan Energy

(kWh)

Total Electric w/o HT

(kWh)

Total Costs w

Furnace ($)

S1 - Conv AC 1,645 1,860 15.4 4,302 5,642 1,179 - - 5,684 980 100%S2 - AC low flw 928 2,038 15.2 4,570 5,841 1,144 - - 5,917 1,017 104%

S8 - Partial SC/RH 205 2,320 12.1 5,840 5,836 1,394 - - 7,437 1,171 120%S9 - Full RH - 2,455 13.8 5,202 6,022 1,463 - - 6,868 1,126 115%

S6 - Ducted DH - 1,996 15.3 4,582 5,097 1,483 1,366 281 7,916 1,166 119%S14 - Cond DES - 1,818 15.3 4,173 5,429 1,406 1,707 260 7,749 1,173 120%

S1 - Conv AC 1,303 2,700 15.3 6,232 1,513 1,572 - - 8,007 917 100%S2 - AC low flw 471 2,881 15.2 6,513 1,588 1,478 - - 8,195 942 103%

S8 - Partial SC/RH 22 3,138 13.0 7,702 1,563 1,709 - - 9,615 1,083 118%S9 - Full RH - 3,178 14.3 6,987 1,628 1,811 - - 9,001 1,026 112%

S6 - Ducted DH - 2,829 15.3 6,496 1,226 1,769 1,084 221 9,773 1,075 117%S14 - Cond DES - 2,672 15.3 6,139 1,377 1,696 1,326 201 9,565 1,065 116%

S1 - Conv AC 628 1,966 15.0 4,572 7,687 1,286 - - 6,061 849 100%S2 - AC low flw 231 2,053 15.0 4,707 7,745 1,245 - - 6,155 859 101%

S8 - Partial SC/RH 9 2,211 13.2 5,397 7,742 1,365 - - 6,965 928 109%S9 - Full RH - 2,250 14.3 4,997 7,786 1,450 - - 6,651 903 106%

S6 - Ducted DH - 2,033 15.0 4,708 7,494 1,487 648 134 7,180 935 110%S14 - Cond DES - 1,953 15.0 4,527 7,590 1,450 770 117 7,067 930 110%

S1 - Conv AC 291 1,338 15.2 2,546 13,004 881 - - 3,631 1,168 100%S2 - AC low flw 183 1,359 15.2 2,576 13,006 877 - - 3,656 1,170 100%

S8 - Partial SC/RH 54 1,417 14.2 2,772 13,014 903 - - 3,879 1,194 102%S9 - Full RH 1 1,446 14.8 2,682 13,013 962 - - 3,847 1,191 102%

S6 - Ducted DH - 1,352 15.2 2,570 12,902 1,087 229 48 4,137 1,214 104%S14 - Cond DES - 1,336 15.2 2,540 12,937 1,081 253 39 4,115 1,213 104%

HERS 100, 50% RH Set Pt

Orlando

Miami

Houston

Atlanta

‐

200 

400 

600 

800 

1,000 

1,200 

1,400 

Z0‐Orlando Z1‐Miami Z2‐Houston Z3‐Atlanta Z4‐Nashville Z5‐Indianapolis

Total Costs w Furnace ($) 

Best Technologies, HERS 100

Conv AC Enhanced AC AC w/ SC/Rht AC w/ Full Rht Ducted DH Cond DES

87

Table 37. Performance Results with Best DH and Enhanced Units, HERS 70 (two-speed)

Figure 35. Total Operating Costs Best Technologies, HERS 70 House

Hours Above

60% RH

AC Runtime

(hrs) AC EER

(Btu/Wh)

AC Energy

(kWh)

Htg Energy

(kWh)

AHU Fan Energy

(kWh)

DH Energy

(kWh)

DH Fan Energy

(kWh)

Total Electric w/o HT

(kWh)

Total Costs w

Furnace ($)

S3 - 2-Spd 575 3,491 21.0 2,615 673 217 - - 3,036 354 100%S2 - AC low flw 308 3,826 20.8 2,756 699 196 - - 3,155 368 104%

S8 - Partial SC/RH 32 4,000 17.0 3,374 703 265 - - 3,843 438 124%S9 - Full RH - 4,220 19.4 2,955 720 309 - - 3,467 401 113%

S6 - Ducted DH - 3,701 21.0 2,746 570 247 821 167 4,184 463 131%S14 - Cond DES - 3,442 20.9 2,539 654 231 992 151 4,116 462 130%

S3 - 2-Spd 382 5,303 20.0 3,629 42 360 - - 4,192 426 100%S2 - AC low flw 20 5,721 19.9 3,772 48 339 - - 4,315 439 103%

S8 - Partial SC/RH - 5,686 16.9 4,413 42 432 - - 5,048 513 120%S9 - Full RH - 6,013 19.2 3,901 47 469 - - 4,573 465 109%

S6 - Ducted DH - 5,543 20.0 3,760 35 385 623 125 5,096 517 121%S14 - Cond DES - 5,274 20.0 3,559 42 363 805 122 5,052 513 120%

S3 - 2-Spd 191 3,844 19.3 2,856 2,034 350 - - 3,410 376 100%S2 - AC low flw 53 4,016 19.3 2,914 2,045 342 - - 3,459 381 101%

S8 - Partial SC/RH - 4,061 17.3 3,233 2,039 380 - - 3,817 411 109%S9 - Full RH - 4,235 18.8 3,000 2,052 439 - - 3,642 396 105%

S6 - Ducted DH - 3,939 19.3 2,904 2,075 378 344 70 3,900 419 112%S14 - Cond DES - 3,826 19.2 2,824 2,117 370 400 61 3,858 418 111%

S3 - 2-Spd 15 2,604 19.7 1,815 4,594 271 - - 2,289 510 100%S2 - AC low flw 15 2,619 19.7 1,819 4,594 270 - - 2,292 510 100%

S8 - Partial SC/RH - 2,662 19.2 1,879 4,595 274 - - 2,356 517 101%S9 - Full RH - 2,698 19.5 1,856 4,595 331 - - 2,390 520 102%

S6 - Ducted DH - 2,600 19.7 1,812 4,700 301 82 17 2,415 529 104%S14 - Cond DES - 2,587 19.7 1,803 4,720 301 81 12 2,400 529 104%

HERS 70, 50% RH Set Pt

Orlando

Miami

Houston

Atlanta

‐

100 

200 

300 

400 

500 

600 

Z0‐Orlando Z1‐Miami Z2‐Houston Z3‐Atlanta Z4‐Nashville Z5‐Indianapolis

Total Costs w Furnace ($) 

Best Technologies, HERS 70

Conv AC Enhanced AC AC w/ SC/Rht AC w/ Full Rht Ducted DH Cond DES

88

2.4.19 Impact of House Size and Other Factors The impact of house size on humidity levels was also considered. The sensitivity to the size of the house was evaluated by making a house that was bigger and smaller than the 2,016 sq ft base house for the HERS 70 and HERS 100 performance level. Table 38 summarizes the major changes to the house model. Wall areas were changed to maintain enclosure continuity. Table 39 summarizes the rules used to change other parameters in the house.

Table 38. Changes to House and Mechanical Systems

SmallHouse

BaseHouse

LargeHouse

Floor Area (ft2) 1198 2016 3495Area Ratio 0.59 1.00 1.73

Ventilation (cfm) 35 58 73Vent Ratio 0.60 1.00 1.26

Cooling Size (tons) Houston HERS 70 2 2 2.5

Nashville HERS 70 2 2 2

Houston HERS 100 2.5 3 4

Nashville HERS 100 2 2.5 3.5

Heating (MBtu/h) HERS 70 40 40 60

HERS 100 40 60 80

Table 39. Rules for Changing House Characteristics with House Size

Parameters that are Proportional to:

Floor Area ELA, supply duct area, return duct area, zone volumes

Ventilation Flow Rate

All ventilation flow rates for HRV, CFIS, fan power

Cooling Unit Size

Supply airflow (heat and cooling), fan power

The simulation results for the smaller and larger houses are given in Table 40. The energy use changed as expected. However the impact of these changes on space humidity levels was very modest. For Houston, the smaller and larger house both resulted in only slightly more hours over 60% RH.

89

Table 40. Impact of House Size on Humidity Levels and Energy Use

In the simulated base house for this study, the windows were closed during the entire year. Since most high humidity levels occur at mild conditions, having the windows closed year-round might impact humidity levels. Many homeowners open their windows during the swing season. Logic was added to open windows (i.e. increase the airchange rate) when certain conditions were met. The following logic was applied to each time step: If space temperature is 2°F lower than the bottom of the cooling deadband and 1°F higher than the top of the heating deadband, then:

Increase ELA by 800 square inches, and

disable the exhaust fan Figure 36 shows the resulting window openings for the HERS 100 house Miami. The shade plot qualitatively shows the hours when the windows were open throughout the year with shades of gray. Each day is shown as a vertical stripe on the plot. Successive days are shown along the X axis. Darker shades indicate hours with the windows open for a larger fraction of each hour. In total, the windows were open for more than 364 hours throughout the year, mostly in the swing months.

Hours Above

60% RH

AC Runtime

(hrs) AC EER

(Btu/Wh)

AC Energy

(kWh)

Htg Energy

(kWh)

AHU Fan Energy

(kWh)

DH Energy

(kWh)

Exh Fan Energy

(kWh)

Total Electric w/o HT

(kWh)

Total Costs w

Furnace ($)

Small House 660 1,532 14.9 2,964 4,789 859 - 123 3,945 543 Normal House 628 1,966 15.0 4,572 7,687 1,286 - 203 6,061 849 Large House 686 1,898 15.0 5,889 9,932 1,657 - 256 7,802 1,094

Small House 5 1,252 15.3 1,890 6,301 617 - 123 2,630 521 Normal House 4 1,555 15.4 2,938 10,150 928 - 203 4,069 823 Large House 3 1,347 15.3 3,578 15,073 1,193 - 256 5,027 1,123

Small House 228 2,908 20.5 1,707 1,227 143 - 123 1,973 220 Normal House 191 3,844 19.3 2,856 2,034 350 - 203 3,410 376 Large House 271 3,915 19.3 3,660 2,923 450 - 256 4,366 495

Small House - 1,968 20.8 1,119 3,550 138 - 123 1,379 280 Normal House - 2,641 19.7 1,859 5,975 299 - 203 2,361 475 Large House - 2,884 19.2 2,303 9,848 400 - 256 2,959 693

HERS 100

Houston

Nashville

HERS 70

Houston

Nashville

90

Figure 36. Shade Plots Showing Windows Openings Across the Year, HERS 100 House

The psychrometric plots in Figure 37 show the impact of operable windows. The hours of high humidity increase slightly. Table 41 and Table 42 show the impact of opening windows in all the climates for both the HERS 100 and the HER 130 houses. Opening the windows in the swing season typically increases the hours above 60% RH.

Windows Closed Operable Windows

Figure 37. Psychrometric Plots Showing Impact of Operable Windows, HERS 100 House

Window Opening - z1h100s1rh50v1w ( 364.8 hrs)

Day (MAX/MIN = 1.00/ 0.00 )

Jan Feb Mar Apr May Jun Jul Aug Sep Oct Nov Dec

2006

0

2

4

6

8

10

12

14

16

18

20

22

24

Hou

r of

Da

y

Daily Indoor Space Conditions: z1h100s1rh50v1

60 65 70 75 80 85 90

Dry Bulb Temperature (F)

0.000

0.005

0.010

0.015

0.020

0.025

Hum

idity

Rat

io (

lb/lb

)

40%

50%

60%

70%

80%

Hours Above 50% = 5361Hours Above 55% = 2399Hours Above 60% = 1303Hours Above 65% = 543Hours Above 70% = 107All HrsCooling Hrs

Daily Indoor Space Conditions: z1h100s1rh50v1w

60 65 70 75 80 85 90

Dry Bulb Temperature (F)

0.000

0.005

0.010

0.015

0.020

0.025

Hum

idity

Rat

io (

lb/lb

)

40%

50%

60%

70%

80%

Hours Above 50% = 5408Hours Above 55% = 2552Hours Above 60% = 1359Hours Above 65% = 592Hours Above 70% = 128All HrsCooling Hrs

91

Table 41. Impact of Window Openings on Humidity Levels and Energy Use – HERS 100, Exh Fan

Table 42. Impact of Window Openings on Humidity Levels and Energy Use – HERS 130, No Ventilation

Hours Above

60% RH

AC Runtime

(hrs) AC EER

(Btu/Wh)

AC Energy

(kWh)

Htg Energy

(kWh)

AHU Fan Energy

(kWh)

Exh Fan

Energy (kWh)

Total Electric w/o HT

(kWh)

Total Costs w

Furnace ($)

Windows Closed 1,645 1,860 15.4 4,302 5,642 1,179 203 5,684 980 Win. Operable 1,729 1,864 15.4 4,314 5,709 1,182 194 5,689 985

Windows Closed 1,303 2,700 15.3 6,232 1,513 1,572 203 8,007 917 Win. Operable 1,359 2,700 15.3 6,233 1,553 1,573 195 8,001 920

Windows Closed 628 1,966 15.0 4,572 7,687 1,286 203 6,061 849 Win. Operable 825 1,962 15.0 4,563 7,764 1,286 194 6,043 850

Windows Closed 291 1,338 15.2 2,546 13,004 881 203 3,631 1,168 Win. Operable 532 1,311 15.2 2,499 13,149 872 181 3,552 1,168

Windows Closed 4 1,555 15.4 2,938 10,150 928 203 4,069 823 Win. Operable 132 1,511 15.3 2,862 10,335 910 185 3,958 820

Windows Closed - 1,057 15.8 1,960 18,145 850 203 3,013 862 Win. Operable 13 1,002 15.7 1,866 18,358 829 186 2,881 859

Indianapolis

HERS 100, Exh vent

Orlando

Miami

Houston

Atlanta

Nashville

Hours Above

60% RH

AC Runtime

(hrs) AC EER

(Btu/Wh)

AC Energy

(kWh)

Htg Energy

(kWh)

AHU Fan Energy

(kWh)

Exh Fan

Energy (kWh)

Total Electric w/o HT

(kWh)

Total Costs w

Furnace ($)

Windows Closed 1,296 2,042 11.6 6,307 6,434 1,300 - 7,606 1,260 Win. Operable 1,435 2,051 11.6 6,340 6,551 1,307 - 7,647 1,273

Windows Closed 1,059 2,936 11.5 9,064 1,798 1,715 - 10,779 1,226 Win. Operable 1,286 2,939 11.5 9,078 1,834 1,718 - 10,795 1,230

Windows Closed 778 2,096 11.3 6,530 13,589 1,498 - 8,028 1,309 Win. Operable 1,109 2,099 11.3 6,542 13,662 1,501 - 8,043 1,313

Windows Closed 314 1,702 11.6 3,455 17,327 909 - 4,364 1,573 Win. Operable 582 1,679 11.6 3,413 17,498 903 - 4,316 1,579

Windows Closed 25 1,622 11.5 4,119 17,533 1,103 - 5,222 1,292 Win. Operable 247 1,590 11.5 4,048 17,735 1,092 - 5,140 1,293

Windows Closed - 1,237 11.9 3,058 20,920 989 - 4,047 1,083 Win. Operable 4 1,170 11.9 2,905 21,196 963 - 3,868 1,079

Nashville

Indianapolis

HERS 130, no vent

Orlando

Miami

Houston

Atlanta

92

Another issue of interest is the impact that moisture adsorption and desorption in the attic has on space humidity levels. This simulation model did not consider this physical phenomena. However, as a means to estimate an order of magnitude of this impact, an artificial dew point increase was imposed on the attic air node during the late morning hours. The “dew point bump” in the attic from desorption was arbitrarily set to be 0.0065 lb/lb from 10 am to 1 pm each day, whenever the attic temperature was over 90°F dry bulb. This increase in humidity ratio corresponds to a dew point increase of 10°F starting from 70°F dp. Figure 38 shows the resulting increase in dew point imposed on the attic air node for a typical summer day (June 30).

Figure 38. Attic Dew Point in Base Model (black) and with an Imposed “Dew Point Bump” (red)

This higher dew point in the attic zone in late morning slightly increased the impact of duct leakage on space humidity levels, as shown in Table 43 and Table 44. This worst-case approximation of solar-driven moisture desorption in the attic is shown to only have a modest impact on space humidity levels and cooling operation.

22: 0: 2: 4: 6: 8: 10: 12: 14: 16: 18: 20: 22: 0:

29 30 1

June

60

80

100

Atti

c D

ew P

t (F

)

93

Table 43. Impact of Attic Dew Point “Bump” on Humidity Levels and Energy Use – HERS 100, Exh Fan

Table 44. Impact of Attic Dew Point “Bump” on Humidity Levels and Energy Use – HERS 130, No Ventilation

Hours Above

60% RH

AC Runtime

(hrs) AC EER

(Btu/Wh)

AC Energy

(kWh)

Htg Energy

(kWh)

AHU Fan Energy

(kWh)

Exh Fan

Energy (kWh)

Total Electric w/o HT

(kWh)

Total Costs w

Furnace ($)

Normal 1,645 1,860 15.4 4,302 5,642 1,179 203 5,684 980 w/ Dewpt "Bump" 1,647 1,869 15.4 4,329 5,638 1,184 203 5,716 982

Normal 1,303 2,700 15.3 6,232 1,513 1,572 203 8,007 917 w/ Dewpt "Bump" 1,307 2,714 15.3 6,271 1,513 1,580 203 8,054 922

Normal 628 1,966 15.0 4,572 7,687 1,286 203 6,061 849 w/ Dewpt "Bump" 630 1,976 15.0 4,600 7,687 1,292 203 6,095 851

Normal 291 1,338 15.2 2,546 13,004 881 203 3,631 1,168 w/ Dewpt "Bump" 292 1,343 15.2 2,559 13,004 884 203 3,647 1,169

Normal 4 1,555 15.4 2,938 10,150 928 203 4,069 823 w/ Dewpt "Bump" 4 1,563 15.4 2,957 10,149 931 203 4,091 825

Normal - 1,057 15.8 1,960 18,145 850 203 3,013 862 w/ Dewpt "Bump" - 1,061 15.8 1,969 18,145 852 203 3,025 862

HERS 100, Exh vent

Orlando

Miami

Houston

Atlanta

Nashville

Indianapolis

Hours Above

60% RH

AC Runtime

(hrs) AC EER

(Btu/Wh)

AC Energy

(kWh)

Htg Energy

(kWh)

AHU Fan Energy

(kWh)

Exh Fan

Energy (kWh)

Total Electric w/o HT

(kWh)

Total Costs w

Furnace ($)

Normal 1,296 2,042 11.6 6,307 6,434 1,300 - 7,606 1,260 w/ Dewpt "Bump" 1,299 2,060 11.6 6,379 6,434 1,310 - 7,689 1,268

Normal 1,059 2,936 11.5 9,064 1,798 1,715 - 10,779 1,226 w/ Dewpt "Bump" 1,074 2,962 11.6 9,167 1,799 1,730 - 10,897 1,238

Normal 778 2,096 11.3 6,530 13,589 1,498 - 8,028 1,309 w/ Dewpt "Bump" 790 2,111 11.3 6,594 13,584 1,506 - 8,100 1,314

Normal 314 1,702 11.6 3,455 17,327 909 - 4,364 1,573 w/ Dewpt "Bump" 319 1,710 11.7 3,478 17,327 912 - 4,390 1,576

Normal 25 1,622 11.5 4,119 17,533 1,103 - 5,222 1,292 w/ Dewpt "Bump" 31 1,635 11.5 4,162 17,532 1,109 - 5,271 1,297

Normal - 1,237 11.9 3,058 20,920 989 - 4,047 1,083 w/ Dewpt "Bump" - 1,246 11.9 3,085 20,920 993 - 4,078 1,086

Nashville

Indianapolis

HERS 130, no vent

Orlando

Miami

Houston

Atlanta

94

There are several key assumptions in the building simulation model that significantly impact the number of high humidity hours. These assumptions include:

the moisture capacitance of the building enclosure and its furnishings,

the internal moisture gains in the space,

the sensible heat gains in the space,

the heat set point. Figure 39 shows the significant impact that moisture and sensible gains have on the number of hours above 60% RH for Orlando in HERS 100 house. Appendix J of Rudd and Henderson et al. (2013) presents a more comprehensive analysis of the results from several hundred runs that varied these model parameters in the HERS 100 and HERS 130 houses in both Houston and Orlando.

Figure 39. Impact of Moisture and Sensible Heat Gains - HERS100 with Exhaust Fan

2.4.20 Conclusions Several hundred annual simulation runs were completed for these various scenarios. The detailed results are available at a project website http://cloud.cdhenergy.com/rp1449/. Local utility rates were used to calculate operating costs for each run. In general, systems that properly control humidity throughout the year had a 10 to 30% higher space conditioning operating cost than uncontrolled conventional systems, depending on the dehumidification system and the relative humidity control set point between 50 to 60% RH. When explicit humidity control is desired, the lowest cost options were the air conditioner with full condenser reheat (System 9) and the condenser-driven desiccant dehumidifier (System 14). The ducted, high-efficiency dehumidifier (System 6) and the air conditioner with a subcooling reheat coil (System 8) had just slightly higher operating costs. In addition, the following observations were made.

‐

500 

1,000 

1,500 

2,000 

2,500 

3,000 

3,500 

10.7 kWh/day 21.3 kWh/day 32 kWh/day

Hours Above

 60% RH 

Orlando:  Moisture & Sensible Gains

6 lb/day 12 lb/day 18 lb/day 24 lb/day

95

Elevated Relative Humidity Levels Typically Occur at Mild Conditions in the Winter and Swing Seasons. Periods of high humidity rarely occur during the main cooling season but instead tend to happen on days when little or no cooling is required. Similarly, dehumidifier operation would be expected to be required at these transition times when space temperatures are just at or below the cooling set point. A More Efficient Building Enclosure Reduces Elevated Relative Humidity Levels – to a Point. With mechanical ventilation provided, as the house efficiency level decreases from HERS 130 to HERS 70, the number of elevated humidity hours generally decreases. However, at HERS 50 bringing the ducts inside tends to increase humidity levels. Hours Above a Certain Relative Humidity Threshold is a Reasonable Metric. The hours above some relative humidity level, say 60% RH, is a reasonably good metric to compare the performance of different systems. In this report we chose 60% RH as the most commonly used limit. However, we also looked at other RH thresholds as well as metrics such as the number of events of some duration above a certain RH level. All these various metrics generally showed the same trends when comparing systems. All these statistics are available in the data set at the web site mentioned in the previous section. Moving Ducts into the Conditioned Space Increases Relative Humidity Levels. When duct losses to the attic are eliminated the number of hours over 60% RH generally increases. A sensitivity in the HERS 50 and HERS 70 houses showed this as well as the comparison of ductless variable speed system and the two-speed system with ducts in the attic. In the hot-humid climates, when ducts were moved inside the conditioned space and heat gain to ducts was eliminated, the hours above 60% RH increased 27%-37% for HERS 70 and 33%-54% for HERS 50. Different Ventilation Systems Have Different Impacts on Relative Humidity Levels. It is generally understood that different types of ventilation system (exhaust, AHU supply, and balanced) combine with infiltration to provide different overall ventilation impacts. We confirmed this finding here and also quantified the impact that different ventilation approaches had on the prevalence of elevated relative humidity. Exhaust ventilation was considered to be the baseline approach in this study. Central fan integrated supply (CFIS) slightly reduced high humidity hours compared to exhaust ventilation in Orlando, and more so in Atlanta. CFIS ventilation slightly increased high humidity hours in Miami and Houston because it provided more fresh air and because the part-time off-cycle operation of the AHU fan sometimes resulted in increased evaporation from the cooling coil. Energy Recovery Ventilators have Little to No High Humidity Control Benefit in Hot-Humid Climates. Since most high relative humidity hours occur at mild conditions – when indoor and outdoor absolute humidity levels are similar – there is very little humidity reduction benefit at that time. The ability of the ERV to exchange moisture between the two air streams is modest at these conditions when indoor and outdoor absolute humidity are nearly the same. Also, in winter in hot-humid climates, ERV operation has the disbenefit of keeping moisture inside the house at times when drier outside air might have helped reduce indoor high humidity.

96

Air Conditioners with Enhanced Controls Significantly Reduce the Number of High Relative Humidity Hours. The cooling units with enhanced controls implemented generally cut the number of hours over 60% RH in half compared to a conventional cooling system when the activation set point was 50%. The enhanced controls strategies are to 1) reduce the relative airflow (cfm/ton) at high relative humidity levels, and 2) allow for 2oF overcooling to increase compressor runtime at mild conditions. This relatively simple technology provides a cost effective means to mitigate many high humidity hours, however, some questions about occupant acceptance of overcooling remain. Variable Capacity Systems (Alone) Do Not Improve Relative Humidity Control. Two-Speed and Variable-Speed Systems do not offer improved relative humidity control unless one or more of the enhancements described above are implemented (lower cfm/ton and overcooling). The benefits of those control enhancements can be realized with both constant speed and variable capacity systems alike. The benefits may be slightly greater for two-speed and variable-speed systems due to the longer runtimes available at low speed. Moisture Capacitance, Internal Moisture Gains, and Cooling/Heating Set Point Significantly Impact Relative Humidity Levels. Model parameters of moisture capacitance, internal moisture gains, and cooling/heating set point have a significant impact on the prevalence of high relative humidity hours across the year (see Appendix J of Rudd and Henderson et al. (2013)). Internal moisture gain and capacitance are difficult to determine by direct measurement or observation, while internal sensible gain and the heating set point can be determined, but all have a significant impact on the number of high humidity hours predicted with the model. This is consistent with the occupant-based variability found in field studies by Rudd et al. 2003, Rudd et al. 2005, and Rudd and Henderson 2007. For this study, we chose the moisture gain (12 lb/day) and moisture capacitance (30x) by comparing the simulation model to measured data (see Appendix C of Rudd and Henderson et al. (2013)). The cooling set point was 78oF. The heating set point was 70oF except in the hot-humid climates where it was 72oF; raising the heat set point even more continued to reduce high RH hours. The Choice of Relative Humidity Set Point Affects Energy Use. The energy required for dehumidification strongly depends on the choice of humidity set point. For instance, decreasing the dehumidifier (DH) set point from 60% to 50% can increase DH energy use by a factor of 5. Other factors may affect the choice of relative humidity set point as well, such as dust mite and condensation/mold avoidance, but those are factors are not within the scope of this report. 2.5 Evaluation of a Method to Process Indoor and Outdoor Temperature and

Relative Humidity Data to Estimate Supplemental Dehumidification Energy A fair amount of hourly or sub-hourly data exists for indoor and outdoor temperature and relative humidity in homes because it is relatively easy to get that data. It is a lot more difficult to get monitored data for HVAC and dehumidifier equipment in homes, so much less of that is available. In order to make better use of available temperature and RH data from hot-humid climates, an analysis process was developed to try to make reasonable predictions of supplemental dehumidification energy consumption based only on temperature and RH.

97

Indoor temperature and RH, and outdoor dry bulb and dew point temperatures are shown in Figure 40. These data were from a previous BSC Building America mechanically ventilated test house in Ft. Myers, FL without supplemental dehumidification. The data were analyzed to predict how much supplemental dehumidification would have been needed to keep the indoor humidity below 60% RH. The analysis approach assumed that for any hour that the indoor RH was above a given threshold value, say 60% RH, any net moisture gain due to air exchange with outdoors, and internal moisture generation, in that hour, would be removed by supplemental dehumidification. It was recognized that this approach ignored the heat and moisture transfer interactions between operating the supplemental dehumidifier and operating the central space conditioning system. However, simulations and testing experience typically have shown that, while not unimportant, that issue is of diminished importance because supplemental dehumidification is mostly needed when the house interior conditions are floating between the cooling and heating set points.

98

Figure 40. Indoor temperature and relative humidity, and outdoor dry bulb temperature and dew

point temperature

The hourly calculated data shown in Figure 41 comes from measured indoor and outdoor temperature and humidity conditions, and assumed internal generation and outdoor air exchange rates. Measured temperature and RH data for this test site was from August through October, which included most of the fall shoulder season between cooling and heating. Figure 41 gives the moisture load components from infiltration/air exchange (which can be positive or negative) and internal generation, and the total moisture which is the net sum of infiltration and internal generation. Internal moisture generation was held constant at 0.5 lb/h, coming from the assumption of 12 lb/day as consistent with the simulation study discussed above. Outside air

40

50

60

70

80

90

100

8/2000 9/2000 9/2000 9/2000 10/2000 10/2000 10/2000 10/2000 11/2000

Tem

per

atu

re (

F)

or

Rel

ativ

e H

um

idit

y (%

)

Indoor ConditionsFt. Myers site

Indoor RH Indoor T

40

50

60

70

80

90

100

8/2000 9/2000 9/2000 9/2000 10/2000 10/2000 10/2000 10/2000 11/2000

Tem

per

atu

re (

F)

Outdoor ConditionsFt. Myers site

Outdoor T Outdoor Tdp

99

exchange was based on a constant 50 cfm, as that relates to the ASHRAE Standard 62.2-2010 rate for the mechanically ventilated house. As displayed in Figure 41 this way, the moisture load in lb/hr can be related to equipment sizing predictions. In this case, with a peak total moisture load of 1.4 lb/h, it can be understood why experience has shown that a 40 pint/day (1.7 lb/h) dehumidifier seems to be able to control moisture below 60% RH for average homes in hot-humid climates.

Figure 41. Net moisture load for hours above 60% relative humidity indoors for 8/25/2000 to

10/31/200 for Ft. Myers, FL house

The resulting moisture gain from this analysis approach, in liters, was divided by the Energy Star dehumidifier Energy Factor (EF) of 2.5 L/kWh, which represents the EF of larger dehumidifiers than can be integrated with central air distribution systems. The result was the predicted supplemental dehumidification electrical energy consumption in kWh, shown in Figure 42. For the sake of this illustration, assuming the spring shoulder season would show a similar pattern, doubling the fall supplemental dehumidification energy of 172 kWh would result in 344 kWh of predicted supplemental dehumidification energy for the year. Of course, with full-year data, this prediction would be more accurate and the exact times of year when supplemental dehumidification was needed would be clear.

(0.4)

(0.2)

‐

0.2 

0.4 

0.6 

0.8 

1.0 

1.2 

1.4 

1.6 

1.8 

8/2000 9/2000 10/2000 10/2000 11/2000

Mo

istu

re l

oad

(lb

/h)

Net moisture load for hours above 60% RH indoorsFt. Myers site

Infil moisture Internal generation Total moisture

100

Figure 42. Indoor relative humidity and predicted supplemental dehumidification energy

consumption from 8/25/2000 to 10/31/2000 for Ft. Meyers, FL house

2.6 Equipment Cost The estimated equipment cost for supplemental dehumidification, or cost differential for the subcooling/condensing reheat systems, can range from $400 to $2,000 depending on the system solution chosen. A stand-alone dehumidifier will cost the least and a desiccant dehumidifier integrated with the central space conditioning system will cost the most. Table 45 provides brief cost estimate detail. Installation costs are complicated with new and retrofit issues, and can vary quite a bit depending on a contractors experience with specific systems, and were not intended to be a part of this report.

0.000

0.050

0.100

0.150

0.200

0.250

0.300

0.350

0.400

0.450

0.500

0

10

20

30

40

50

60

70

80

90

100

8/2000 9/2000 10/2000 10/2000 11/2000

Ho

url

y D

ehu

mid

ifie

r E

ner

gy

(kW

h)

Ind

oo

r R

elat

ive

Hu

mid

ity

(%)

Indoor RH and predicted dehumidifier energyFt. Myers site

Indoor RH Deh Energy

101

Table 45. First-Cost Estimates for Supplemental Dehumidification Systems

Supplemental Dehumidification System

First-Cost Estimate

Stand-alone Dehumidifier with Remote Dehumidistat

$300

Integrated Ducted

Dehumidifier $1,000

Sub-cooling Reheat $1,600

Full-condensing Reheat $1,750

Desiccant Dehumidifier $2,000

3 Conclusions

A number of BSC studies on supplemental dehumidification techniques were used to draw a current evaluation the question, “What are the best mechanical means to accomplish indoor humidity control in high performance homes in hot-humid climates and how much supplemental dehumidification is needed?” Those studies ranged from large field and simulation studies, to product development and testing, to a data processing approach to predict supplemental dehumidification requirements from indoor and outdoor temperature and relative humidity data alone. The most important conclusions from this study are as follows:

In a multi-home study in Houston, TX, measured supplemental dehumidification energy consumption from two mechanically ventilated homes was 209 kWh/yr for a representative home with a stand-alone dehumidifier and 463 kWh/yr for another representative home with a ducted dehumidifier. The ducted dehumidifier was more efficient, and the homes had similar temperature and relative humidity control, but variability in occupant behaviors has a strong impact on internal moisture generation which has a strong impact on supplemental dehumidification requirements. Internal moisture generation was not measured in that study, and is nearly impossible to measure except when under controlled simulation in a lab house environment.

Detailed simulations showed that a number of humidity control solutions can be effective in hot-humid climates. The most effective solutions, having relatively low operating cost and essentially eliminating indoor humidity above 60% RH, were: full condensing reheat integrated with the central cooling system, ducted dehumidifier, stand-alone dehumidifier with central system mixing, and condenser regenerated desiccant dehumidifier. About 170 kWh/yr could be expected for a HERS 50 house (having ducts inside conditioned space) with a 60% RH setpoint. About five times that could be expected with a 50% RH

102

setpoint. A close second was central cooling system with subcooling reheat but it showed more elevated RH hours. A more distant third place was enhanced cooling with 2oF over-cooling and lower airflow (200 cfm/ton), if more hours above 60% RH and over-cooling discomfort can be tolerated, and note that it only works well to reduce elevated RH if a 50% RH setpoint is used. Two-speed and variable speed systems did little to reduce hours of elevated relative humidity in hot-humid climates unless coupled with the enhanced cooling methods listed above.

An Energy Recovery Ventilator (ERV) does little to reduce moisture loads when supplemental dehumidification is needed, which mostly occurs when there is little difference in absolute humidity between indoors and outdoors. With little absolute humidity to exchange between, ERVs have little impact on reducing elevated indoor relative humidity hours. In some hot-humid climates, including Orlando and Houston, energy recovery ventilation actually increases hours of elevated indoor humidity over exhaust and central-fan-integrated supply because the ERV sometimes keeps moisture in the house when drier outdoor air could reduce indoor humidity. If an ERV is operated in conjunction with supplemental dehumidification operated with a 50% RH setpoint, then the ERV does help reduce supplemental dehumidification energy consumption. That is because the dehumidifier forces a greater indoor to outdoor absolute humidity difference, allowing the ERV to reject some outdoor moisture with house exhaust air.

An analysis approach using only hourly indoor and outdoor measured temperature and relative humidity data, from a mechanically ventilated test house in Ft. Meyers, FL, showed predicted supplemental dehumidification energy consumption of 344 kWh/yr for a ducted dehumidifier (2.5 L/kWh). This compared to 410 kWh/yr from detailed simulations of a mechanically ventilated HERS 100 house in Miami. This analysis approach should be investigated further. All that is required is hourly or sub-hourly indoor and outdoor measured temperature and RH data, and some basic house characteristics information, from houses in hot-humid climates without supplemental dehumidification. Using available data in this way may prove useful for predicting supplemental dehumidification requirements for high performance homes.

Based on that measured data analysis approach and BSC field experience, when controlling to 60% relative humidity, the required capacity for supplemental dehumidification in average homes in hot-humid climates is not large – about 1.5 lb/h, or that of a typical 40 to 50 pint/day unit. However, the capacity rating of dehumidifiers is made at higher temperature (80oF) than is typical in homes. With the same amount of moisture in the air, the actual capacity at lower temperatures will be less than the rated capacity.

103

Acknowledgements

This project was supported by the U.S. Department of Energy, Office of Building Technologies, Building America Program, with contract management by the National Renewable Energy Laboratory. The computer simulations part of this project (Rudd and Henderson et al. 2013) was a collaborative effort supported by the American Society of Heating, Refrigeration and Air Conditioning Engineers and, in part, by the Air-Conditioning, Heating and Refrigeration Technology Institute.

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106

Appendix A. Presentation of Additional Analysis of ASHRAE RP-1449 Data, Summarizing Supplemental Dehumidification Energy and Cost

107

Table A 1. Supplemental dehumidification energy and cost over a conventional cooling system, for a HERS 50 house with different mechancial ventilation systems

RH Setpoint: 60% 60% 50% 50% 60% 60% 50% 50% 60% 60% 50% 50%

Dehumification Energy over

Conv. System(kWh)

Dehumification Cost over

Conv. System($)

Dehumification Energy over

Conv. System(kWh)

Dehumification Cost over

Conv. System($)

Dehumification Energy over

Conv. System(kWh)

Dehumification Cost over

Conv. System($)

Dehumification Energy over

Conv. System(kWh)

Dehumification Cost over

Conv. System($)

Dehumification Energy over

Conv. System(kWh)

Dehumification Cost over

Conv. System($)

Dehumification Energy over

Conv. System(kWh)

Dehumification Cost over

Conv. System($)

Orlando Conv or 2-spd1 -- -- -- -- -- -- -- -- -- -- -- --

Low flow+overcool 2 3 0 110 12 4 0 77 9 3 0 63 10

Subcool Reheat 2 61 6 1004 103 76 7 596 61 62 7 433 46

Full Cond Reheat 4 118 12 471 48 160 16 325 34 151 16 270 29

Standalone DH 3 152 9 1418 126 132 8 915 76 118 6 604 47

Ducted DH 4 138 8 1293 114 114 6 836 69

Cond Desiccant 4 121 9 1233 119 99 8 786 74

Miami Conv or 2-spd1 -- -- -- -- -- -- -- -- -- -- -- --

Low flow+overcool 2 9 1 124 14 11 1 91 11 7 1 68 8

Subcool Reheat 2 81 8 1041 105 95 10 591 61 66 7 421 43

Full Cond Reheat 4 136 14 467 47 169 17 320 33 153 15 265 27

Standalone DH 3 119 11 1308 131 112 11 778 78 80 7 540 54

Ducted DH 4 106 10 1186 119 100 10 707 70

Cond Desiccant 4 92 9 1139 114 94 9 691 69

Houston Conv or 2-spd1 -- -- -- -- -- -- -- -- -- -- -- --

Low flow+overcool 2 1 0 57 6 3 0 46 6 3 0 33 3

Subcool Reheat 2 29 2 530 46 37 3 354 31 27 3 254 22

Full Cond Reheat 4 94 8 276 24 140 12 237 21 133 11 211 18

Standalone DH 3 82 4 735 56 64 3 463 34 51 2 362 25

Ducted DH 4 76 4 673 51 59 3 430 32

Cond Desiccant 4 63 4 604 49 43 3 378 30

Atlanta Conv or 2-spd1 -- -- -- -- -- -- -- -- -- -- -- --

Low flow+overcool 2 0 0 8 1 0 0 5 1 0 0 0 0

Subcool Reheat 2 0 0 98 10 0 0 86 9 0 0 34 4

Full Cond Reheat 4 47 5 94 10 98 9 129 12 98 9 114 11

Standalone DH 3 29 2 170 14 0 0 128 11 0 0 55 4

Ducted DH 4 29 2 160 13 0 0 121 10

Cond Desiccant 4 30 2 146 13 0 0 106 10

HERS 50

Exhaust Ventilation CFIS Ventilation ERV Ventilation

1 This system does not control to a RH setpoint2 These systems will attempt to control to an RH setpoint but will not always meet the setpoint due to a limit on overcooling3 This system will control to a RH setpoint but may not have the capacity to always meet the setpoint, especially without whole-house mixing4 These systems will control to a RH setpoint and will meet the setpoint

108

Table A 2. Supplemental dehumidification energy and cost over a conventional cooling system, for a HERS 70 house with different mechanical ventilation systems

RH Setpoint: 60% 60% 50% 50% 60% 60% 50% 50% 60% 60% 50% 50%Dehumification

Energy overConv. System

(kWh)

Dehumification Cost over

Conv. System($)

Dehumification Energy over

Conv. System(kWh)

Dehumification Cost over

Conv. System($)

Dehumification Energy over

Conv. System(kWh)

Dehumification Cost over

Conv. System($)

Dehumification Energy over

Conv. System(kWh)

Dehumification Cost over

Conv. System($)

Dehumification Energy over

Conv. System(kWh)

Dehumification Cost over

Conv. System($)

Dehumification Energy over

Conv. System(kWh)

Dehumification Cost over

Conv. System($)

Orlando Conv or 2-spd1 -- -- -- -- -- -- -- -- -- -- -- --

Low flow+overcool 2 1 0 120 14 2 0 79 10 1 0 63 9

Subcool Reheat 2 38 4 807 84 60 6 546 57 47 5 436 47

Full Cond Reheat 4 133 14 431 47 193 19 351 36 183 19 314 34

Standalone DH 3 108 11 1268 120 114 8 805 67 94 6 612 49

Ducted DH 4 102 10 1148 109 107 7 736 62

Cond Desiccant 4 87 10 1080 108 89 7 665 63

Miami Conv or 2-spd1 -- -- -- -- -- -- -- -- -- -- -- --

Low flow+overcool 2 4 0 123 13 4 0 68 8 2 0 56 6

Subcool Reheat 2 45 5 856 86 52 6 419 43 35 4 357 36

Full Cond Reheat 4 161 16 381 39 190 19 295 30 178 18 275 28

Standalone DH 3 72 7 1000 100 68 7 544 54 48 5 472 46

Ducted DH 4 65 7 904 91 63 6 499 50

Cond Desiccant 4 58 6 860 87 54 5 469 47

Houston Conv or 2-spd1 -- -- -- -- -- -- -- -- -- -- -- --

Low flow+overcool 2 0 0 50 5 1 0 32 4 1 0 26 3

Subcool Reheat 2 12 1 407 35 21 2 258 23 13 1 209 19

Full Cond Reheat 4 116 10 232 20 160 13 228 19 155 13 218 18

Standalone DH 3 51 8 545 48 37 2 358 26 32 2 300 21

Ducted DH 4 49 8 490 43 34 2 322 23

Cond Desiccant 4 46 8 448 42 28 2 278 22

Atlanta Conv or 2-spd1 -- -- -- -- -- -- -- -- -- -- -- --

Low flow+overcool 2 0 0 3 0 0 0 1 0 0 0 0 0

Subcool Reheat 2 0 0 67 7 0 0 49 5 0 0 20 3

Full Cond Reheat 4 74 8 101 10 120 11 140 13 117 11 130 12

Standalone DH 3 21 10 132 20 0 0 87 7 0 0 53 4

Ducted DH 4 21 10 126 19 0 0 83 7

Cond Desiccant 4 21 10 111 19 0 0 78 7

Exhaust Ventilation CFIS Ventilation ERV Ventilation

HERS 70

1 This system does not control to a RH setpoint2 These systems will attempt to control to an RH setpoint but will not always meet the setpoint due to a limit on overcooling3 This system will control to a RH setpoint but may not have the capacity to always meet the setpoint, especially without whole-house mixing4 These systems will control to a RH setpoint and will meet the setpoint

109

Table A 3. Supplemental dehumidification energy and cost over a conventional cooling system, for a HERS 85 house with different mechanical ventilation systems

RH Setpoint: 60% 60% 50% 50% 60% 60% 50% 50% 60% 60% 50% 50%Dehumification

Energy overConv. System

(kWh)

Dehumification Cost over

Conv. System($)

Dehumification Energy over

Conv. System(kWh)

Dehumification Cost over

Conv. System($)

Dehumification Energy over

Conv. System(kWh)

Dehumification Cost over

Conv. System($)

Dehumification Energy over

Conv. System(kWh)

Dehumification Cost over

Conv. System($)

Dehumification Energy over

Conv. System(kWh)

Dehumification Cost over

Conv. System($)

Dehumification Energy over

Conv. System(kWh)

Dehumification Cost over

Conv. System($)

Orlando Conv or 2-spd1 -- -- -- -- -- -- -- -- -- -- -- --

Low flow+overcool 2 19 3 127 18 17 3 105 17 8 1 102 18

Subcool Reheat 2 104 13 1019 110 141 16 943 102 139 16 1007 109

Full Cond Reheat 4 230 27 688 79 232 28 564 68 227 29 604 77

Standalone DH 3 347 30 1585 137 221 14 1316 105 216 13 1268 104

Ducted DH 4 333 28 1448 125 203 12 1212 97

Cond Desiccant 4 317 31 1332 127 194 16 1122 103

Miami Conv or 2-spd1 -- -- -- -- -- -- -- -- -- -- -- --

Low flow+overcool 2 6 1 114 13 9 1 104 14 13 1 116 17

Subcool Reheat 2 93 9 879 90 136 15 855 89 177 20 1017 106

Full Cond Reheat 4 230 24 592 61 242 26 516 55 255 28 590 64

Standalone DH 3 238 19 1205 112 164 12 1064 98 200 15 1186 109

Ducted DH 4 224 18 1103 102 154 11 968 89

Cond Desiccant 4 212 18 956 91 141 12 869 84

Houston Conv or 2-spd1 -- -- -- -- -- -- -- -- -- -- -- --

Low flow+overcool 2 0 0 41 4 2 0 48 6 6 1 53 7

Subcool Reheat 2 27 3 438 38 54 5 503 44 63 6 561 49

Full Cond Reheat 4 150 13 334 30 163 14 328 30 162 14 365 33

Standalone DH 3 171 15 676 54 67 4 588 45 76 5 647 48

Ducted DH 4 168 15 622 50 60 3 537 41

Cond Desiccant 4 163 15 545 46 53 4 473 39

Atlanta Conv or 2-spd1 -- -- -- -- -- -- -- -- -- -- -- --

Low flow+overcool 2 0 0 5 1 0 0 4 0 0 0 1 0

Subcool Reheat 2 0 0 103 11 0 0 107 11 0 0 64 6

Full Cond Reheat 4 83 8 132 14 81 8 125 13 78 8 110 11

Standalone DH 3 145 19 279 30 0 0 142 13 0 0 78 7

Ducted DH 4 145 19 271 30 0 0 132 11

Cond Desiccant 4 145 19 252 29 0 0 120 12

HERS 85

Exhaust Ventilation CFIS Ventilation ERV Ventilation

1 This system does not control to a RH setpoint2 These systems will attempt to control to an RH setpoint but will not always meet the setpoint due to a limit on overcooling3 This system will control to a RH setpoint but may not have the capacity to always meet the setpoint, especially without whole-house mixing4 These systems will control to a RH setpoint and will meet the setpoint

110

Table A 4. Supplemental dehumidification energy and cost over a conventional cooling system, for a HERS 100 house with different mechanical ventilation systems

RH Setpoint: 60% 60% 50% 50% 60% 60% 50% 50% 60% 60% 50% 50%Dehumification

Energy overConv. System

(kWh)

Dehumification Cost over

Conv. System($)

Dehumification Energy over

Conv. System(kWh)

Dehumification Cost over

Conv. System($)

Dehumification Energy over

Conv. System(kWh)

Dehumification Cost over

Conv. System($)

Dehumification Energy over

Conv. System(kWh)

Dehumification Cost over

Conv. System($)

Dehumification Energy over

Conv. System(kWh)

Dehumification Cost over

Conv. System($)

Dehumification Energy over

Conv. System(kWh)

Dehumification Cost over

Conv. System($)

Orlando Conv or 2-spd1 -- -- -- -- -- -- -- -- -- -- -- --

Low flow+overcool 2 85 15 233 38 75 14 171 35 49 11 130 30

Subcool Reheat 2 257 34 1753 191 291 40 1533 173 288 38 1779 196

Full Cond Reheat 4 405 54 1184 147 385 56 909 126 365 54 995 137

Standalone DH 3 607 46 2426 203 406 28 1973 155 386 21 2191 171

Ducted DH 4 581 43 2233 186 377 25 1803 140

Cond Desiccant 4 565 51 2065 193 371 34 1747 162

Miami Conv or 2-spd1 -- -- -- -- -- -- -- -- -- -- -- --

Low flow+overcool 2 22 2 187 24 20 2 140 20 37 4 137 21

Subcool Reheat 2 224 23 1607 166 288 31 1427 149 392 41 1843 192

Full Cond Reheat 4 347 38 993 109 353 40 786 92 390 45 956 110

Standalone DH 3 431 32 1942 175 303 20 1607 141 385 27 2051 181

Ducted DH 4 410 30 1766 158 276 18 1468 128

Cond Desiccant 4 393 34 1558 148 264 23 1347 128

Houston Conv or 2-spd1 -- -- -- -- -- -- -- -- -- -- -- --

Low flow+overcool 2 5 1 93 10 10 2 77 10 15 2 67 9

Subcool Reheat 2 95 9 903 79 147 13 896 79 167 15 1078 95

Full Cond Reheat 4 221 20 589 54 225 20 509 49 236 22 584 56

Standalone DH 3 326 24 1211 94 174 11 984 74 187 12 1185 89

Ducted DH 4 318 24 1119 87 156 10 906 68

Cond Desiccant 4 315 24 1006 81 149 11 838 68

Atlanta Conv or 2-spd1 -- -- -- -- -- -- -- -- -- -- -- --

Low flow+overcool 2 0 0 26 3 0 0 15 1 0 0 8 1

Subcool Reheat 2 17 1 248 26 15 2 239 25 6 1 194 20

Full Cond Reheat 4 97 10 217 23 92 10 187 20 85 9 170 18

Standalone DH 3 242 23 524 47 29 2 305 27 8 1 234 20

Ducted DH 4 239 22 507 46 27 2 286 25

Cond Desiccant 4 241 23 485 46 29 3 274 26

Exhaust Ventilation CFIS Ventilation ERV Ventilation

HERS 100

1 This system does not control to a RH setpoint2 These systems will attempt to control to an RH setpoint but will not always meet the setpoint due to a limit on overcooling3 This system will control to a RH setpoint but may not have the capacity to always meet the setpoint, especially without whole-house mixing4 These systems will control to a RH setpoint and will meet the setpoint

111

Figure A 1. Supplemental dehumidification energy consumption and cost, along with hours above 60% relative humidity, for a HERS 50 house in Orlando with three different ventilation systems

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Dehumidification Energy, Orlando, HERS 50, ERV

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Dehumidification Cost, Orlando, HERS 50, ERV

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112

Figure A 2. Supplemental dehumidification energy consumption and cost, along with hours above 60% relative humidity, for a HERS 50 house in

Miami with three different ventilation systems

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Dehumidification Energy, Miami, HERS 50, Exhaust

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Dehumidification Cost, Miami, HERS 50, ERV

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113

Figure A 3. Supplemental dehumidification energy consumption and cost, along with hours above 60% relative humidity, for a HERS 50 house in

Houston with three different ventilation systems

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DOE/GO-000000-0000 ▪ Month Year

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