Humidity Control in Humid Climates

Transcription

1 Humidity Control in Humid Climates Armin Rudd Building Science Corporation August 2013

2 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. Available electronically at Available for a processing fee to U.S. Department of and its contractors, in paper, from: U.S. Department of Office of Scientific and Technical Information P.O. Box 62 Oak Ridge, TN phone: fax: mailto:reports@adonis.osti.gov Available for sale to the public, in paper, from: U.S. Department of Commerce National Technical Information Service 5285 Port Royal Road Springfield, VA phone: fax: orders@ntis.fedworld.gov online ordering: Printed on paper containing at least 50% wastepaper, including 20% postconsumer waste

3 Task Order 3 Task 11: Additional Research Activities Deliverable Humidity Control in Humid Climates: Technical Report Prepared for: The National Renewable Laboratory On behalf of the U.S. Department of s Building America Program Office of Efficiency and Renewable Denver West Parkway Golden, CO Prepared by: Armin Rudd Building Science Corporation 30 Forest Street Somerville, MA NREL Technical Monitor: Cheryn Metzger Prepared under Subcontract No. TASK ORDER NO. KNDJ August 2013 iii

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5 Contents List of Figures... vi List of Tables... viii Definitions... x Executive Summary... xii 1 Problem Statement Introduction and Background Research Questions Relevance to Building America s Goals Tradeoffs and Other Benefits Technical Approach Results Dehumidification Study in Houston, Texas Research Approach Results Enhanced Cooling System Study Advanced Cooling with Dedicated Dehumidifier Mode System Design Approach Prototype Construction Bench-top Testing Results TRYNSYS Computer Simulation Study Climates Building and Enclosure Thermal Details Cooling System Details Mechanical Ventilation Options Space Conditioning Systems Electric and Natural Gas Costs Simulation Results Evaluating Humidity Levels Location of Ducts in the Conditioned Space Impact of Lower Duct Leakage and R-value Impact of Ventilation Options Impact of Ventilation Rate (50% and 150% of ASHRAE 62.2 Requirements) Single Speed, Two Speed, and Variable Speed Air Conditioners Enhanced Air Conditioner Control Options Cooling Systems with Further Enhancements Dehumidifiers Advanced Dehumidifiers Comparing the Best Overall Technologies Impact of House Size and Other Factors Conclusions Evaluation of a Method to Process Indoor and Outdoor Temperature and Relative Humidity Data to Estimate Supplemental Dehumidification Equipment Cost Conclusions Acknowledgements References v

6 Appendix A. Presentation of Additional Analysis of ASHRAE RP-1449 Data, Summarizing Supplemental Dehumidification and Cost List of Figures Figure 1. Photo of Enhanced Cooling PSC System test house in Houston, TX Figure 2. Photo of Enhanced Cooling ECM System test house in Houston, TX Figure 3. Indoor temperature and relative humidity at the PSC enhanced cooling system test house (PSC blower motor) Figure 4. Indoor environmental conditions and equipment operation for a representative 3-day period in April 2004 at the PSC System house in Houston 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 Figure 7: Final design schematic showing cooling operation (greyed-out lines are inactive) Figure 8: Final design schematic showing dehumidification-only operation (greyed-out lines are inactive) Figure 9: (photo left) 2.2 Advanced Cooling with Dedicated Dehumidifier Mode System as benchtop 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) Figure 10. IECC Climate Zone Map Figure 11. Schematic of Mechanical Ventilation System Options Figure 12. Configuration of Natural Gas-Fired Desiccant Unit (pulling air from the cooling outlet; supplying to supply duct) Figure 13. Schematic of Configurations for Dehumidifier and Desiccant Systems Figure 14. Comparing Costs for Different Climates and HERS Levels Figure 15. Comparing Cooling and Fan Use for Different Climates and HERS Levels Figure 16. Comparing Air Conditioner Runtime for Different Climates and HERS Levels Figure 17. Comparing Air Conditioner Gross EER for Different Climates and HERS Levels 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 Figure 20. Histogram of Relative Humidity for HERS 100, Miami, System 1, Exhaust Fan Figure 21. Comparing High Humidity Levels for Different Climates and HERS Levels Figure 22. Comparing Number of High Humidity Events (exceeding 60% RH) for Different Climates and HERS Levels Figure 23. Impact of Duct Location on Space Humidity Levels for HERS 50 and 70 Houses Figure 24. Comparing Ventilation Rates (mechanical & natural H) for Different Ventilation Systems in Miami Figure 25. Comparing Ventilation Rates (mechanical & natural cfm) for Different Ventilation Systems in Miami Figure 26. Comparing Electric Use for Ventilation Options with Different Systems Figure 27. Comparing Costs for Different Ventilation Options in each City Figure 28. Comparing High Humidity Levels for Different Ventilation Options in each City 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 Figure 31. Operating Costs for Standalone DH at Various RH Set Points, HERS 100 House 80 Figure 32. Operating Costs All DH Options, 50% RH Set Point, HERS 100 House Figure 33. Operating Costs All DH Options, 50% RH Set Point, HERS 70 House Figure 34. Operating Costs Best Technologies, HERS 100 House vi

7 Figure 35. Operating Costs Best Technologies, HERS 70 House Figure 36. Shade Plots Showing Windows Openings Across the Year, HERS 100 House Figure 37. Psychrometric Plots Showing Impact of Operable Windows, HERS 100 House 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 Figure 40. Indoor temperature and relative humidity, and outdoor dry bulb temperature and dew point temperature 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 Figure 42. Indoor relative humidity and predicted supplemental dehumidification energy consumption from 8/25/2000 to 10/31/2000 for Ft. Meyers, FL house 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 vii

8 List of Tables Table 1. Monitoring results for Stand-alone dehumidifier system and the Ducted dehumidifier system in the Houston study Table 2. Climates Selected for Simulation Table 3. Description of Performance Levels Table 4. Enclosure Leakage and Duct Performance Table 5. Cooling Unit Sizing for each Climate and HERS Level Table 6. Cooling Unit Characteristics Table 7. Additional cooling Unit Characteristics for Two-Speed and Variable Speed Systems Table 8. Heating Cooling and Dehumidification Set Points Table 9. Summary of Space Conditioning Systems Table 10. Matrix of cooling Conditioner Systems Used with each HERS Level and System Table 11. Electric and Natural Gas Costs Table 12. Performance Results for Different HERS Levels and Climates Table 13. Comparing Relative Annual Costs for each HERS Level and Climate Table 14. Space Conditioning Cost Reduction Impact of Moving Ducts from Attic to the Conditioned Space Table 15. Impact of Duct Location on Space Humidity Levels for HERS 50 and 70 Houses Table 16. Performance Results for Different Duct Leakage Rates in HERS 100 and HERS 130 Houses Table 17. Performance Results for Different Duct Insulation Levels in HERS 70 and HERS 85 Houses Table 18. Performance Results for Different Ventilation System and Climates (for HERS 100, System 1: 13 SEER cooling) Table 19. Performance Results for Different Ventilation System and Climates (for HERS 85, System 1: 14.5 SEER cooling, ECM Fan) Table 20. Performance Results for Different Ventilation System and Climates (for HERS 100, System 3: Two-Speed Cooling) Table 21. Performance Results with Different Ventilation Rates, HERS 100, Exhaust Fan Table 22. Performance Results with Different Ventilation Rates, HERS 100, CFIS Table 23. Performance Results with Different Ventilation Rates, HERS 100, ERV Table 24. Performance Results with Different Ventilation Rates, HERS 70, Exhaust Fan Table 25. Performance Results with Different Ventilation Rates, HERS 70, CFIS Table 26. Performance Results with Different Ventilation Rates, HERS 70, ERV Table 27. Performance Results for Different Cooling Units and Climates (for HERS 100 and HERS 50) Table 28. Performance Results with Enhanced Cooling Unit Table 29. Performance Results with Various Cooling Unit Enhancements Table 30. Performance Results for Cooling Units with Further Enhancements, HERS Table 31. Performance Results for Cooling Units with Further Enhancements, HERS 70 (twospeed) Table 32. Performance Results with Standalone DH (System 5) with Different Set Points Table 33. Humidity Threshold Results for Standalone DH (System 5) with Different Set Points Table 34. Performance Results with Various DH Units (Systems, 5, 6, 7, 13 & 14) with HERS Table 35. Performance Results with Various DH Units (Systems, 5, 6, 7, 13 & 14) with HERS and 50% RH Setpoint Table 36. Performance Results with Best DH and Enhanced Units, HERS Table 37. Performance Results with Best DH and Enhanced Units, HERS 70 (two-speed) Table 38. Changes to House and Mechanical Systems Table 39. Rules for Changing House Characteristics with House Size Table 40. Impact of House Size on Humidity Levels and Use Table 41. Impact of Window Openings on Humidity Levels and Use HERS 100, Exh Fan91 viii

9 Table 42. Impact of Window Openings on Humidity Levels and Use HERS 130, No Ventilation Table 43. Impact of Attic Dew Point Bump on Humidity Levels and Use HERS 100, Exh Fan Table 44. Impact of Attic Dew Point Bump on Humidity Levels and Use HERS 130, No Ventilation Table 45. First-Cost Estimates for Supplemental Dehumidification Systems Table A 1. Supplemental dehumidification energy and cost over a conventional cooling system, for a HERS 50 house with different mechancial ventilation systems Table A 2. Supplemental dehumidification energy and cost over a conventional cooling system, for a HERS 70 house with different mechanical ventilation systems Table A 3. Supplemental dehumidification energy and cost over a conventional cooling system, for a HERS 85 house with different mechanical ventilation systems Table A 4. Supplemental dehumidification energy and cost over a conventional cooling system, for a HERS 100 house with different mechanical ventilation systems ix

10 Definitions ach ach50 AHAM AHRI AHU ASHRAE ASTM BA BeOpt BNL BSC CFI CFM 50 DOE EPA ERV HRV HUD ICC IECC air changes per hour air changes per hour at 50 Pascal pressure differential Association of Home Appliance Manufacturers Air Conditioning Heating and Refrigeration Institute Air handler unit American Society of Heating, Refrigerating and Air- Conditioning Engineers American Society for Testing And Materials Building America Program. More information about BA can be found at Building Optimization Program House energy simulation program and primary analysis tool for Building America homes. Brookhaven National Laboratory Building Science Corporation. More information about BSC can be found at Central-fan-integrated Cubic feet per minute at 50 Pascal test pressure differential U.S. Department of U.S. Environmental Protection Agency recovery ventilator Heat recovery ventilator U.S. Department of Housing and Urban Development International Code Council International Conservation Code. More information can be found at x

11 IMC IRC NREL Pa RH International Mechanical Code International Residential Code National Renewable Laboratory Pascal; SI unit of pressure (equivalent to one newton per square meter) relative humidity xi

12 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 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 xii

13 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 2 o F 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 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. xiii

14 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 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. 14

15 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 wholebuilding 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 's (DOE) Building America program is to reduce home energy use by 30%-50% (compared to 2009 energy codes for new homes and preretrofit energy use for existing homes). To this end, we conduct research to develop marketready 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 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 Code The homes were constructed with unvented attics (insulation under the roof sheathing, therefore the attics were semi-conditioned). Three reference

16 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 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. -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 16

17 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. 17

18 Table 1. Monitoring results for Stand-alone dehumidifier system and the Ducted dehumidifier system in the Houston study Stand-alone Dehumidifier Ducted Dehumidifier Monitoring period 10/4/2001-7/31/ /17/2001-8/1/2002 days with data House dry bulb temperature (avg of 3 points) min max avg std dev 2 3 House relative humidity (avg of 3 points) min max avg std dev 8 7 House dew point temperature (avg of 3 points) min max avg std dev 5 6 Hours above 60% RH 7% 4% Supplemental dehumidification energy consumption 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 HV 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 18

19 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, Each of the homes was a single-story plan of about 2500 ft 2, with similar solar exposure and occupancy. PSC System 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/ton 2. 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 Figure 1. Photo of Enhanced Cooling PSC System test house in Houston, TX 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. 19

20 the requested setpoint by more than 3 o F. This arrangement allows the evaporator coil to run very cold, in the range of 38 o F. 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 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. Figure 2. Photo of Enhanced Cooling ECM System test house in Houston, TX 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. 20

21 PSC, Enhanced cooling system house, Houston, TX Temperature (F) or RH (%) Jun-03 5-Jul-03 5-Aug-03 5-Sep-03 6-Oct-03 6-Nov-03 7-Dec-03 7-Jan-04 7-Feb-04 9-Mar-04 9-Apr May Jun-04 Tdb_tstat RH_CR bed RH_mbed RH_tstat 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 3 o F 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. 21

22 PSC, Enhanced cooling system house, Houston, TX :00 6:00 12:00 18:00 0:00 6:00 12:00 18:00 0:00 6:00 12:00 18:00 0:00 6:00 12:00 18:00 Temperature (F) or RH (%) Minutes or Cycles per hour Tdb_avg RH_avg Tdp_avg Fan_RT Cool_RT Cool cycles Heat_RT 10 0 Hour of day (4/23 to 4/ ) 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. 22

23 ECM, Enhanced cooling system house, Houston, TX Temperature (F) Oct-03 4-Nov-03 5-Dec-03 5-Jan-04 5-Feb-04 7-Mar-04 7-Apr-04 8-May-04 8-Jun RH (%) Tdb_tstat RH_CR bed RH_mbed RH_tstat Figure 5. Indoor temperature and relative humidity for the ECM enhanced cooling system test house (ECM blower) ECM, Enhanced cooling test house, Houston Temperature (F) or RH (%) Minutes or Cycles per hour RH_avg Tdb_avg Tdp_avg Tsuppy_min Tcoil_min Fan_RT Cool_RT Heat_RT Cool cycles Hour of day (4/23 to 4/ ) 0 Figure 6. Indoor environmental conditions and equipment operation for a representative 3-day period in April at the ECM test house in Houston 23

24 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 49 o F, 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 3 o F 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 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 24

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) 25

26 Figure 8: Final design schematic showing dehumidification-only operation (greyed-out lines are inactive) 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. 26

27 Figure 9: (photo left) 2.2 Advanced Cooling with Dedicated Dehumidifier Mode System as benchtop 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) 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 (50 o F - 65 o F). In standard cooling mode and 50 o F 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 80 o F 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 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 6 o F 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 27

28 refrigerant condensing temperature was 125 o F. 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 100 o F 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 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-ResDH 3 (Henderson and Sand 2003; Henderson et al 2007), was improved to meet the needs of this project. TRNSYS-based hour-byhour building energy simulation tool was used to simulate the HV 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 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 An updated version and documentation will be available to the PMS at the conclusion of this project. 28

29 Table 2. Climates Selected for Simulation Zone ID IECC Climate City Zone Z0 2A Orlando Z1 1A Miami Z2 2A Houston Z3 3A Atlanta Z4 4A Nashville Z5 5A Indianapolis Figure 10. IECC Climate Zone Map Building and Enclosure Thermal Details The 2,000 ft 2 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 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. 29

30 Table 3. Description of Performance Levels Description or Benchmark HERS Index Cooling Efficiency Typical Existing House SEER HERS Reference House SEER Star House SEER (BPM fan) Builders Challenge House SEER (two-speed) Building America Prototype 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 H50 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 in 2 for all the HERS levels, or about 5 times the leakage rate for the HERS 100 house (Fugler 1999). HERS Level Target H at 50 Pa Table 4. Enclosure Leakage and Duct Performance ELA at 4 Pa (in 2 ) Duct Leakage, Supply and Return Combined (% of flow) Duct Insulation R-Value (h-f-ft 2 /Btu) Sup / Ret Duct Area (ft 2 ) % / % / % / % / None Na Na Note: Duct leaks are assumed to be 60% on supply side, 40% on return side. 30

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: MBtu/day (21.3 kwh/day) for HERS 130 and MBtu/day (19.2 kwh/day) for HERS MBtu/day (17 kwh/day) for HERS 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: dw dt i C m ( wi wo ) Qint ernal Q, latent where: w - humidity ratio (lb/lb), i=indoor, o=outdoor m - net indoor-to-outdoor airflow rate (lb/h) Q internal - Internal moisture gains (lb/h) Q, latent - moisture removal by air conditioner or dehumidifier (lb/h) C - a capacitance term: air mass in space x multiplication factor 31

32 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 mass 4. 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 ft 2 (371.6 m 2 ) 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 Plus. 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 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 32

33 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 Miami Houston Atlanta Nashville Indianapolis 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 Input Parameters Rated Performance 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) SEER 13 Unit (Single Spd, PSC fan mtr) SEER 14.5 Unit (Single Spd, BPM fan mtr) Two Speed (BPM fan Mtr) Variable Speed, Ductless 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. 33

34 Description Table 7. Additional cooling Unit Characteristics for Two-Speed and Variable Speed Systems Input Parameters Gross EER (Btu/Wh) SHR (-) Actual Fan Power (W/cfm) Low Stage Gross EER (Btu/Wh) Low Stage Input Parameters Low Low Low Stage Stage Fan Stage Capacity Power SHR (-) Ratio (-) (W/cfm) SEER 10 Unit (Single Spd, PSC fan mtr) SEER 13 Unit (Single Spd, PSC fan mtr) SEER 14.5 Unit (Single Spd, BPM fan mtr) Two Speed (BPM fan Mtr) Variable Speed, Ductless 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. Low Stage Airflow Ratio (-) 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 Rating Systems Standards. 0 Orlando 1 Miami 2 Houston 3 Atlanta 4 Nashville 5 Indianapolis Table 8. Heating Cooling and Dehumidification Set Points Heating Set Point ( F) Cooling Set Point ( F) Dehumidification Set Point (%) 78 50% or 60% 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. 34

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) 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 ft 2 3 bedroom house. The HERS 130 was also modeled with no mechanical ventilation (infiltration only). infiltration mechanical exhaust exfiltration infiltration Induced exfiltration AHU Return air Induced infiltration Mechanical ventilation AHU Return air exfiltration Exhaust Only (exh fan operates 100% of hour) Supply - Central Fan Integrated (damper/ahu operates 34% of 3x flow) 35

36 infiltration exfiltration infiltration exfiltration AHU Return air AHU HRV Return air HRV Balanced HRV / ERV Ducted AHU (HRV & AHU operates 50% of 2x flow) Balanced HRV / ERV Ductless AHU (Sys 4) (HRV operates 50% of 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. cfm in cfm out = sum of all incoming ventilation flows (vent cfm at, dehum/vent unit, etc) = sum of all exhaust flows (exhaust fan, net duct leakage, etc) cfm balanced = MIN(cfm in, cfm out ) cfm unbalanced = MAX(cfm in, cfm out ) - cfm balanced cfm inf = infiltration flow calculated for building for the timestep cfm combined = MAX(cfm unbalanced, cfm inf + 0.5* cfm unbalanced ) + cfm balanced Some fresh air is provided as part of the mechanical system. Therefore, the net mechanical inlet flows are subtracted from cfm combined to determine the remaining non-mechanical ventilation (or infiltration) rate acting on the building enclosure. A mass balance tracks CO 2 levels in the space and confirms that the net impact of ventilation is similar between all the cases. Exhaust Only. 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 Central Fan Integrated. 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). 36

37 Balanced Ventilation ERV/HRV. An 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) 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. 37

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 Conventional DX System with Lower Airflow and 2 Yes (reset) No Thermostat Overcooling 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 6 Ducted Dehumidifier with Conventional System Mixing Yes Yes 7 Ducted Dehumidifier with Outdoor Air Preconditioning Yes Yes 8 9 Enhanced Cooling with Subcooling Reheat (single speed & two-stage compressor) Enhanced Cooling with Full Condensing/Subcooling Reheat (single-speed and two-stage compressor) 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 System 1 Conventional DX System. A conventional, single-speed DX air conditioner (aircooled 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. System 2 Conventional DX System with Lower Airflow and Thermostat Overcooling. 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. 38 Yes Yes Yes Yes No No

39 System 3 Two Speed Conventional DX System. A two-speed cooling unit with a gross EER of Btu/Wh at high speed and 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 System 1 Conven System 2 Enhanced System 3 Two-Spd System 4 Var-Spd Systems 5-14 HERS 130 SEER 10 1-Spd (16) SEER 10 1-Spd (16) SEER Spd (19) SEER 19 V-Spd (20) SEER 10 1-Spd (16) HERS 100 SEER 13 1-Spd (17) SEER 13 1-Spd (17) SEER Spd (19) SEER 19 V-Spd (20) SEER 13 1-Spd (17) HERS 85 SEER Spd (18) SEER Spd (18) SEER Spd (19) SEER 19 V-Spd (20) SEER Spd (18) HERS 70 HERS 50 Note: SEER Spd (18) SEER Spd (18) SEER Spd (19) SEER Spd (19) SEER Spd (19) SEER Spd (19) The number in parentheses is the unit index in TRNSYS. Grey shaded cells indicate the default for each HERS level SEER 19 V-Spd (20) SEER 19 V-Spd (20) SEER Spd (19) SEER Spd (19) System 4 Variable Speed Mini Split DX System. 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. System 5 Stand Alone Dehumidifier with Conventional System Mixing. A portable, standalone dehumidifier (DH) supplements the air conditioner. A small 50 pint/day dehumidifier with 39

40 an Factor (EF) of 1.55 liters/kwh at AHAM rating conditions 5. 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. System 6 Ducted Dehumidifier with Conventional System Mixing. 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 HV 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. System 7 Ducted Dehumidifier with Outdoor Air Preconditioning. 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 HV 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. System 8 Enhanced Cooling with Subcooling Reheat. 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. 40

41 System 9 Enhanced Cooling with Full Condensing/Subcooling Reheat. 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). System 10 Conventional DX System with Lower Airflow. 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. System 11 Conventional DX System with Thermostat Overcooling. 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. System 12 Conventional DX System with Sensible Only AAHX. 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 noncondensing, 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). System 13 Conventional DX System with Natural Gas Regenerated Desiccant Dehumidifier. 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. 41

42 Figure 12. Configuration of Natural Gas-Fired Desiccant Unit (pulling air from the cooling outlet; supplying to supply duct) System 14 DX Condenser Regenerated Desiccant Dehumidifier. 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 H). 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 Electric and Natural Gas Costs HV 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

43 Table 11. Electric and Natural Gas Costs Electric Utility $/kwh $/therm Z0 Orlando FPL Z1 Miami FPL Z2 Houston Entergy Z3 Atlanta Georgia Power Z4 Nashville Nashville Z5 Indianapolis Indianapolis PL 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

44 infiltration exfiltration infiltration exfiltration DH AHU DH AHU Return air Return air Stand-Alone Dehumidifier (System 5) Ducted Dehumidifier (System 6) infiltration exfiltration DH AHU Return air Ducted Dehumidifier with OA (System 7) infiltration exfiltration infiltration exfiltration Des AHU Des AHU Return air Natural Gas-Fired Desiccant DH (System 13) Return air Desiccant Dehumidifier (System 14) Figure 13. Schematic of Configurations for Dehumidifier and Desiccant Systems 44

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 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). 45

46 Table 12. Performance Results for Different HERS Levels and Climates Orlando Miami Houston Atlanta Nashville Indianapolis Hours Above 60% RH Runtime (hrs) EER (Btu/Wh) Htg AHU Fan Exh Fan Electric w/o HT Costs w Furnace ($) $ $ $ $ $ HERS 50 1,011 3, , , HERS , , , HERS 85 1,121 2, ,611 2, , HERS 100 1,645 1, ,302 5,642 1, , HERS 130 1,361 2, ,629 6,876 1, ,190 1,352 HERS , , ,893 $ 295 HERS , , ,192 $ 426 HERS , , ,214 $ 657 HERS 100 1,303 2, ,232 1,513 1, ,007 $ 917 HERS 130 1,313 3, ,505 1,944 1, ,495 $ 1,309 HERS , ,963 1, ,347 $ 269 HERS , ,856 2, ,410 $ 376 HERS , ,876 3, ,838 $ 540 HERS , ,572 7,687 1, ,061 $ 849 HERS , ,823 14,261 1, ,581 $ 1,386 HERS , ,149 3, ,488 $ 341 HERS , ,815 4, ,289 $ 510 HERS , ,364 6, ,082 $ 689 HERS , ,546 13, ,631 $ 1,168 HERS , ,530 18, ,668 $ 1,673 HERS 50-2, ,248 4, ,611 $ 320 HERS 70-2, ,859 5, ,361 $ 475 HERS 85-1, ,379 8, ,118 $ 632 HERS , ,938 10, ,069 $ 823 HERS , ,246 18,838 1, ,594 $ 1,387 HERS 50-1, , ,213 $ 364 HERS 70-1, ,206 12, ,736 $ 553 HERS 85-1, ,582 13, ,238 $ 634 HERS 100-1, ,960 18, ,013 $ 862 HERS , ,090 22,902 1, ,322 $ 1,177 Table 13. Comparing Relative Annual Costs for each HERS Level and Climate HERS50 HERS70 HERS85 HERS100 HERS130 HERS130 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% 46

47 Costs w Furnace ($) 1,800 1,600 1,400 1,200 1, System 1 Exh Fan (V1) Z0 Orlando Z1 Miami Z2 Houston Z3 Atlanta Z4 Nashville Z5 Indianapolis HERS50 HERS70 HERS85 HERS100 HERS130 HERS130 No Vnt Figure 14. Comparing Costs for Different Climates and HERS Levels Figure 15 shows that when heating energy is excluded (designated as 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. 14,000 System 1 Exh Fan (V1) Electric w/o HT 12,000 10,000 8,000 6,000 4,000 2,000 Z0 Orlando Z1 Miami Z2 Houston Z3 Atlanta Z4 Nashville Z5 Indianapolis HERS50 HERS70 HERS85 HERS100 HERS130 HERS130 No Vnt Figure 15. Comparing Cooling and Fan Use for Different Climates and HERS Levels 47

48 6,000 System 1 Exh Fan (V1) 5,000 Runtime (hrs) 4,000 3,000 2,000 1,000 Z0 Orlando Z1 Miami Z2 Houston Z3 Atlanta Z4 Nashville Z5 Indianapolis HERS50 HERS70 HERS85 HERS100 HERS130 HERS130 No Vnt 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 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. 25 System 1 Exh Fan (V1) 20 EER (Btu/Wh) Z0 Orlando Z1 Miami Z2 Houston Z3 Atlanta Z4 Nashville Z5 Indianapolis HERS50 HERS70 HERS85 HERS100 HERS130 HERS130 No Vnt Figure 17. Comparing Air Conditioner Gross EER for Different Climates and HERS Levels 48

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 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. 49

50 Daily Indoor Space Conditions: z1h100s1rh50v Hours Above 50% = 5361 Hours Above 55% = 2399 Hours Above 60% = 1303 Hours Above 65% = 543 Hours Above 70% = 107 All Hrs Cooling Hrs 80% 70% 60% 50% 40% Days with Cooling Operation Humidity Ratio (lb/lb) Dry Bulb Temperature (F) Figure 18. Psychrometric Chart Showing Space Conditions for HERS 100, Miami, System 1, Exhaust Fan Hour of Day Shades of Gray 0 below 55% % % % 5 above 75% Jan Feb Mar Apr May Jun Jul Aug Sep Oct Nov Dec 2006 Figure 19. Shade Plot showing Humidity bins for each Hour of Year for HERS 100, Miami, System 1, Exhaust Fan 50

51 400 RH: z1h100s1rh50v1 Number of Hours 200 Min RH = 42.0 Avg RH = 53.1 Max RH = Indoor Relative Humidity (%) Figure 20. Histogram of Relative Humidity for HERS 100, Miami, System 1, Exhaust Fan 51

52 Hours Above 60% RH 1,800 1,600 1,400 1,200 1, System 1 Exh Fan (V1) Above 60% Z0 Orlando Z1 Miami Z2 Houston Z3 Atlanta Z4 Nashville Z5 Indianapolis HERS50 HERS70 HERS85 HERS100 HERS130 HERS130 No Vnt Hours Above 55% RH 4,000 3,500 3,000 2,500 2,000 1,500 1, System 1 Exh Fan (V1) Above 55% Z0 Orlando Z1 Miami Z2 Houston Z3 Atlanta Z4 Nashville Z5 Indianapolis HERS50 HERS70 HERS85 HERS100 HERS130 HERS130 No Vnt Hours Above 50% RH 8,000 7,000 6,000 5,000 4,000 3,000 2,000 1,000 System 1 Exh Fan (V1) Above 50% Z0 Orlando Z1 Miami Z2 Houston Z3 Atlanta Z4 Nashville Z5 Indianapolis HERS50 HERS70 HERS85 HERS100 HERS130 HERS130 No Vnt Figure 21. Comparing High Humidity Levels for Different Climates and HERS Levels 52

53 No of Events > 4 hrs System 1 Exh Fan (V1) Z0 Orlando Z1 Miami Z2 Houston Z3 Atlanta Z4 Nashville Z5 Indianapolis HERS50 HERS70 HERS85 HERS100 HERS130 HERS130 No Vnt 60 System 1 Exh Fan (V1) 50 No of Events > 8 hrs Z0 Orlando Z1 Miami Z2 Houston Z3 Atlanta Z4 Nashville Z5 Indianapolis HERS50 HERS70 HERS85 HERS100 HERS130 HERS130 No Vnt Figure 22. Comparing Number of High Humidity Events (exceeding 60% RH) for Different Climates and HERS Levels 53

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. Space Conditioning Cost Reduction Impact of Moving Ducts from Attic to the Conditioned Space Cost Reduction HERS 70 HERS 50 Z0-Orlando 22% 26% Z1-Miami 27% 28% Z2-Houston 26% 26% Z3-Atlanta 24% 24% Z4-Nashville 24% 16% Z5-Indianapolis 23% 10% 54

55 1,200 Location of Ducts Hours Above 60% RH 1, Z0 Orlando Z1 Miami Z2 Houston Z3 Atlanta Z4 Nashville Z5 Indianapolis HERS 50 HERS 70 In Attic In Space In Attic In Space 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 Hours >60% RH Ducts inside Ducts in attic 1 % Diff. wrt ducts in attic HERS 50 Orlando % Miami % Houston % HERS 70 Orlando % Miami % Houston % 1 ducts in attic have 5% leakage (3% supply, 2% return) 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%. 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. 55

56 Table 16. Performance Results for Different Duct Leakage Rates in HERS 100 and HERS 130 Houses HERS 100, Exh Fan Houston Hours Above 60% RH Runtime (hrs) EER (Btu/Wh) Htg AHU Fan Exh Fan Electric w/o HT Costs w Furnace ($) 10% Duct Leak 628 1, ,572 7,687 1, , % Duct Leak 628 1, ,337 7,410 1, , Nashville 10% Duct Leak 4 1, ,938 10, , % Duct Leak 4 1, ,797 9, , HERS 130, no vent Houston 20% Duct Leak 778 2, ,530 13,589 1,498-8,028 1,309 5% Duct Leak 849 1, ,736 12,237 1,340-7,076 1,165 Nashville 20% Duct Leak 25 1, ,119 17,533 1,103-5,222 1,292 5% Duct Leak 17 1, ,634 15, ,621 1,151 The HERS 70 and HERS 85 houses by default use R8 duct insulation. The results in Table 17 show the impact of decreasing the insulation level to R6. As expected, energy use increases with less duct insulation. The number of high humidity hours decrease slightly as more sensible heat gain through the ducts increases cooling runtime and removes more moisture. Table 17. Performance Results for Different Duct Insulation Levels in HERS 70 and HERS 85 Houses Houston HERS 70 Hours Above 60% RH Runtime (hrs) EER (Btu/Wh) Htg AHU Fan Exh Fan Electric w/o HT Costs w Furnace ($) Ducts R , ,403 2, , Ducts R , ,554 2, , Nashville Ducts R8-1, ,215 5, , Ducts R6-1, ,327 6, , HERS 85 Houston Ducts R , ,876 3, , Ducts R , ,020 3, , Nashville Ducts R8-1, ,379 8, , Ducts R6-1, ,487 8, ,

57 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 ft 2, 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) 57

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. Number of Hours z1h100s1rh50v0 - No Vent Min H = Avg H = Max H = Number of Hours z1h100s1rh50v1 - Exh Fan Min H = Avg H = Max H = H (1/h) H (1/h) Number of Hours z1h100s1rh50v2 - CFIS Min H = Avg H = Max H = Number of Hours z1h100s1rh50v4 - ERV Min H = Avg H = Max H = H (1/h) H (1/h) Figure 24. Comparing Ventilation Rates (mechanical & natural H) for Different Ventilation Systems in Miami 58

59 Vent & Inf (cfm) Vent & Inf (cfm) Week of 01/01/06 - Winter : 0: 12: 0: 12: 0: 12: 0: 12: 0: 12: 0: 12: 0: 12: 0: Black = Exh Fan Green = CFIS Blue = ERV Week of 06/16/06 - Summer : 12: 0: 12: 0: 12: 0: 12: 0: 12: 0: 12: 0: 12: 0: 12: 0: Figure 25. Comparing Ventilation Rates (mechanical & natural cfm) for Different Ventilation Systems in Miami 59

60 Table 18. Performance Results for Different Ventilation System and Climates (for HERS 100, System 1: 13 SEER cooling) System 1, HERS 100 Orlando Hours Above 60% RH Runtime (hrs) EER (Btu/Wh) Htg AHU Fan Exh Fan HRV Electric w/o HT Costs w Furnace ($) V0 - Natural 1,833 1, ,001 5,260 1, , V1 - Exh Fan 1,645 1, ,302 5,642 1, , V2 - CFIS 1,627 1, ,406 5,543 1, ,242 1,029 V3 - HRV V4 - ERV 1,742 1, ,300 5,312 2, ,797 1,068 V5 - ERV ind 1,623 1, ,221 5,447 1, , Miami V0 - Natural 1,119 2, ,821 1,382 1, , V1 - Exh Fan 1,303 2, ,232 1,513 1, , V2 - CFIS 1,394 2, ,330 1,474 2, , V3 - HRV V4 - ERV 1,737 2, ,178 1,396 2, , V5 - ERV ind 1,206 2, ,124 1,447 1, , Houston V0 - Natural 625 1, ,284 7,060 1, , V1 - Exh Fan 628 1, ,572 7,687 1, , V2 - CFIS 728 1, ,646 7,666 1, , V3 - HRV V4 - ERV 793 1, ,536 7,391 2, , V5 - ERV ind 590 1, ,489 7,399 1, , Atlanta V0 - Natural 226 1, ,435 11, ,276 1,070 V1 - Exh Fan 291 1, ,546 13, ,631 1,168 V2 - CFIS 163 1, ,598 12,864 1, ,043 1,202 V3 - HRV 157 1, ,650 12,475 1, ,715 1,247 V4 - ERV 62 1, ,542 12,556 1, ,593 1,239 V5 - ERV ind 194 1, ,519 12, ,642 1,140 Nashville V0 - Natural - 1, ,812 8, , V1 - Exh Fan 4 1, ,938 10, , V2 - CFIS 6 1, ,962 10,225 1, , V3 - HRV 15 1, ,019 9,908 1, , V4 - ERV Indianapolis V0 - Natural - 1, ,914 16, , V1 - Exh Fan - 1, ,960 18, , V2 - CFIS - 1, ,980 17,938 1, , V3 - HRV 9 1, ,033 17,669 1, , V4 - ERV 60

61 Table 19. Performance Results for Different Ventilation System and Climates (for HERS 85, System 1: 14.5 SEER cooling, ECM Fan) System 1, HERS 85 Orlando Hours Above 60% RH Runtime (hrs) EER (Btu/Wh) Htg AHU Fan Exh Fan HRV Electric w/o HT Costs w Furnace ($) V0 - Natural 1,449 1, ,314 1, , V1 - Exh Fan 1,121 2, ,611 2, , V2 - CFIS 1,105 2, ,704 2,367 1, , V3 - HRV V4 - ERV 1,304 1, ,534 2,111 1, , V5 - ERV ind 1,029 2, ,524 2, , Miami V0 - Natural 578 2, , , V1 - Exh Fan 756 2, , , V2 - CFIS 865 2, , , , V3 - HRV V4 - ERV 1,149 2, , , , V5 - ERV ind 616 2, , , Houston V0 - Natural 278 2, ,595 2, , V1 - Exh Fan 228 2, ,876 3, , V2 - CFIS 304 2, ,937 3,314 1, , V3 - HRV V4 - ERV 460 2, ,771 2,982 1, , V5 - ERV ind 182 2, ,788 2, , Atlanta V0 - Natural 55 1, ,262 5, , V1 - Exh Fan 20 1, ,364 6, , V2 - CFIS 10 1, ,399 6, , V3 - HRV 2 1, ,434 6,010 1, , V4 - ERV - 1, ,318 6,082 1, , V5 - ERV ind 10 1, ,338 5, , Nashville V0 - Natural 6 1, ,260 6, , V1 - Exh Fan - 1, ,379 8, , V2 - CFIS 4 1, ,406 8, , V3 - HRV 4 1, ,441 7,777 1, , V4 - ERV Indianapolis V0 - Natural - 1, ,545 11, , V1 - Exh Fan - 1, ,582 13, , V2 - CFIS - 1, ,598 14, , V3 - HRV 2 1, ,641 13, , V4 - ERV 61

62 Table 20. Performance Results for Different Ventilation System and Climates (for HERS 100, System 3: Two-Speed Cooling) System 3, HERS 100 Orlando Hours Above 60% RH Runtime (hrs) EER (Btu/Wh) Htg AHU Fan Exh Fan HRV Electric w/o HT Costs w Furnace ($) V0 - Natural 1,786 2, ,914 5, , V1 - Exh Fan 1,635 3, ,147 5, , V2 - CFIS 1,603 3, ,167 5, , V3 - HRV V4 - ERV 1,595 3, ,091 5, , Miami V0 - Natural 1,148 4, ,227 1, , V1 - Exh Fan 1,313 4, ,544 1, , V2 - CFIS 1,289 4, ,542 1, , V3 - HRV V4 - ERV 1,360 4, ,442 1, , Houston V0 - Natural 634 3, ,214 7, , V1 - Exh Fan 620 3, ,452 7, , V2 - CFIS 701 3, ,440 7, , V3 - HRV V4 - ERV 675 3, ,368 7, , Atlanta V0 - Natural 223 2, ,850 11, , V1 - Exh Fan 295 2, ,946 12, ,628 1,062 V2 - CFIS 143 2, ,954 12, ,542 1,050 V3 - HRV 192 2, ,014 12, ,863 1,074 V4 - ERV 85 2, ,917 12, ,753 1,067 Nashville V0 - Natural - 2, ,152 8, , V1 - Exh Fan 3 2, ,258 10, , V2 - CFIS 4 2, ,241 10, , V3 - HRV 7 2, ,311 10, , V4 - ERV Indianapolis V0 - Natural - 1, ,457 16, , V1 - Exh Fan - 1, ,492 18, , V2 - CFIS - 1, ,487 18, , V3 - HRV 3 1, ,539 17, , V4 - ERV 62

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). 63

64 Electric 10,000 9,000 8,000 7,000 6,000 5,000 4,000 3,000 2,000 1,000 Miami, System 1, HERS 100 V0 Natural V1 Exh Fan V2 CFIS V3 HRV V4 ERV V5 ERV ind AHU Fan Exh Fan HRV Fan 13 SEER Electric 7,000 6,000 5,000 4,000 3,000 2,000 1,000 Miami, System 1, HERS 85 V0 Natural V1 Exh Fan V2 CFIS V3 HRV V4 ERV V5 ERV ind AHU Fan Exh Fan HRV Fan 14.5 SEER, ECM Electric 6,000 5,000 4,000 3,000 2,000 1,000 Miami, System 3, HERS 100 V0 Natural V1 Exh Fan V2 CFIS V3 HRV V4 ERV V5 ERV ind AHU Fan Exh Fan HRV Fan Two-Speed Cooling Figure 26. Comparing Electric Use for Ventilation Options with Different Systems 64

65 Costs w Furnace ($) 1,400 1,200 1, HERS 100 System 1 13 SEER Z0 Orlando Z1 Miami Z2 Houston Z3 Atlanta Z4 Nashville Z5 Indianapolis V0 Natural V1 EXH Fan V2 CFIS V3 HRV V4 ERV Costs w Furnace ($) HERS 85 System SEER, ECM Z0 Orlando Z1 Miami Z2 Houston Z3 Atlanta Z4 Nashville Z5 Indianapolis Costs w Furnace ($) 1,200 1, V0 Natural V1 EXH Fan V2 CFIS V3 HRV V4 ERV HERS 100 System 3 Two-Speed Z0 Orlando Z1 Miami Z2 Houston Z3 Atlanta Z4 Nashville Z5 Indianapolis V0 Natural V1 EXH Fan V2 CFIS V3 HRV V4 ERV Figure 27. Comparing Costs for Different Ventilation Options in each City 65

66 Hours Above 60% RH 2,000 1,800 1,600 1,400 1,200 1, HERS 100 System 1 13 SEER Z0 Orlando Z1 Miami Z2 Houston Z3 Atlanta Z4 Nashville Z5 Indianapolis 1,600 1,400 V0 Natural V1 EXH Fan V2 CFIS V3 HRV V4 ERV HERS 85 System SEER, ECM Hours Above 60% RH 1,200 1, Z0 Orlando Z1 Miami Z2 Houston Z3 Atlanta Z4 Nashville Z5 Indianapolis V0 Natural V1 EXH Fan V2 CFIS V3 HRV V4 ERV Hours Above 60% RH 2,000 1,800 1,600 1,400 1,200 1, HERS 100 System 3 Two-Speed Z0 Orlando Z1 Miami Z2 Houston Z3 Atlanta Z4 Nashville Z5 Indianapolis V0 Natural V1 EXH Fan V2 CFIS V3 HRV V4 ERV Figure 28. Comparing High Humidity Levels for Different Ventilation Options in each City 66

67 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 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 Exh Vent, HERS 100 Orlando Hours Above 60% RH Runtime (hrs) EER (Btu/Wh) Htg AHU Fan Exh Fan HRV Electric w/o HT Costs w Furnace ($) 50% Vent 1,688 1, ,124 5,433 1, , % 100% Vent 1,645 1, ,302 5,642 1, , % 150% Vent 1,742 1, ,478 5,892 1, ,003 1, % Miami 50% Vent 1,188 2, ,992 1,441 1, , % 100% Vent 1,303 2, ,232 1,513 1, , % 150% Vent 1,425 2, ,474 1,609 1, , % Houston 50% Vent 580 1, ,405 7,348 1, , % 100% Vent 628 1, ,572 7,687 1, , % 150% Vent 731 2, ,741 8,035 1, , % Atlanta 50% Vent 240 1, ,482 12, ,431 1,112 95% 100% Vent 291 1, ,546 13, ,631 1, % 150% Vent 388 1, ,612 13, ,820 1, % Nashville 50% Vent - 1, ,863 9, , % 100% Vent 4 1, ,938 10, , % 150% Vent 42 1, ,005 10, , % Indianapolis 50% Vent - 1, ,933 17, , % 100% Vent - 1, ,960 18, , % 150% Vent 6 1, ,982 19, , % 67

68 Table 22. Performance Results with Different Ventilation Rates, HERS 100, CFIS CFIS, HERS 100 Orlando Hours Above 60% RH Runtime (hrs) EER (Btu/Wh) Htg AHU Fan Exh Fan HRV Electric w/o HT Costs w Furnace ($) 50% Vent 1,728 1, ,190 5,260 1, , % 100% Vent 1,627 1, ,406 5,543 1, ,242 1, % 150% Vent 1,612 1, ,608 5,897 1, ,482 1, % Miami 50% Vent 1,367 2, ,030 1,388 2, , % 100% Vent 1,394 2, ,330 1,474 2, , % 150% Vent 1,487 2, ,610 1,572 2, , % Houston 50% Vent 744 1, ,441 7,188 1, , % 100% Vent 728 1, ,646 7,666 1, , % 150% Vent 801 2, ,834 8,216 1, , % Atlanta 50% Vent 63 1, ,506 12,186 1, ,929 1,148 96% 100% Vent 163 1, ,598 12,864 1, ,043 1, % 150% Vent 281 1, ,680 13,669 1, ,146 1, % Nashville 50% Vent - 1, ,868 9,411 1, , % 100% Vent 6 1, ,962 10,225 1, , % 150% Vent 50 1, ,044 11,163 1, , % Indianapolis 50% Vent - 1, ,937 16,900 1, , % 100% Vent - 1, ,980 17,938 1, , % 150% Vent 10 1, ,019 19,427 1, , % Table 23. Performance Results with Different Ventilation Rates, HERS 100, ERV ERV, HERS 100 Orlando Hours Above 60% RH Runtime (hrs) EER (Btu/Wh) Htg AHU Fan Exh Fan HRV Electric w/o HT Costs w Furnace ($) 50% Vent 1,803 1, ,213 5,238 2, ,557 1,039 97% 100% Vent 1,742 1, ,300 5,312 2, ,797 1, % 150% Vent 1,760 1, ,390 5,399 2, ,027 1, % Miami 50% Vent 1,751 2, ,057 1,381 2, , % 100% Vent 1,737 2, ,178 1,396 2, , % 150% Vent 1,719 2, ,296 1,412 2, ,128 1, % Houston 50% Vent 829 1, ,454 7,262 2, , % 100% Vent 793 1, ,536 7,391 2, , % 150% Vent 803 1, ,616 7,547 2, , % Atlanta 50% Vent 27 1, ,502 12,359 1, ,406 1,208 97% 100% Vent 62 1, ,542 12,556 1, ,593 1, % 150% Vent 99 1, ,581 12,746 1, ,765 1, % 68

69 Table 24. Performance Results with Different Ventilation Rates, HERS 70, Exhaust Fan Exh Vent, HERS 70 Orlando Hours Above 60% RH Runtime (hrs) EER (Btu/Wh) Htg AHU Fan Exh Fan HRV Electric w/o HT Costs w Furnace ($) 50% Vent 678 1, , , % 100% Vent 606 1, , , % 150% Vent 858 1, , , % Miami 50% Vent 159 2, , , % 100% Vent 433 3, , , % 150% Vent 763 3, , , % Houston 50% Vent 164 2, ,241 1, , % 100% Vent 214 2, ,403 2, , % 150% Vent 288 2, ,537 2, , % Atlanta 50% Vent 27 1, ,098 4, , % 100% Vent 15 1, ,167 4, , % 150% Vent 40 1, ,214 5, , % Nashville 50% Vent - 1, ,137 5, , % 100% Vent - 1, ,215 5, , % 150% Vent 16 1, ,266 6, , % Indianapolis 50% Vent - 1, ,426 11, , % 100% Vent - 1, ,451 12, , % 150% Vent 4 1, ,463 13, , % Table 25. Performance Results with Different Ventilation Rates, HERS 70, CFIS CFIS, HERS 70 Orlando Hours Above 60% RH Runtime (hrs) EER (Btu/Wh) Htg AHU Fan Exh Fan HRV Electric w/o HT Costs w Furnace ($) 50% Vent 861 1, , , % 100% Vent 745 1, , , , % 150% Vent 882 1, , , , % Miami 50% Vent 438 2, , , % 100% Vent 602 3, , , , % 150% Vent 793 3, , , , % Houston 50% Vent 300 2, ,275 1, , % 100% Vent 255 2, ,455 2, , % 150% Vent 350 2, ,609 2, , % Atlanta 50% Vent - 1, ,106 4, , % 100% Vent - 1, ,186 4, , % 150% Vent 32 1, ,256 5, , % Nashville 50% Vent - 1, ,145 5, , % 100% Vent - 1, ,229 6, , % 150% Vent 16 1, ,299 7, , % Indianapolis 50% Vent - 1, ,426 11, , % 100% Vent - 1, ,462 13, , % 150% Vent 4 1, ,493 14, , % 69

70 Table 26. Performance Results with Different Ventilation Rates, HERS 70, ERV ERV, HERS 70 Orlando Hours Above 60% RH Runtime (hrs) EER (Btu/Wh) Htg AHU Fan Exh Fan HRV Electric w/o HT Costs w Furnace ($) 50% Vent 1,179 1, , , , % 100% Vent 984 1, , , , % 150% Vent 927 1, , , , % Miami 50% Vent 721 2, , , , % 100% Vent 770 3, , , , % 150% Vent 827 3, , , , % Houston 50% Vent 533 2, ,204 1,891 1, , % 100% Vent 415 2, ,283 1,989 1, , % 150% Vent 397 2, ,359 2,098 1, , % Atlanta 50% Vent - 1, ,061 4,267 1, , % 100% Vent - 1, ,103 4,438 1, , % 150% Vent - 1, ,142 4,631 1, , % 70

71 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. 71

72 Table 27. Performance Results for Different Cooling Units and Climates (for HERS 100 and HERS 50) Orlando HERS 100 Hours Above 60% RH Runtime (hrs) EER (Btu/Wh) Htg AHU Fan Exh Fan Electric w/o HT Electric w/o HT Costs w Furnace ($) Constant Speed 1,645 1, ,302 5,642 1, ,684 5, Two-Speed 1,635 3, ,147 5, ,862 3, Variable Spd Ductless 1,874 3, ,026 4, ,487 2, Var Spd - Lower flw 1,805 3, ,070 4, ,523 2, Miami Constant Speed 1,303 2, ,232 1,513 1, ,007 8, Two-Speed 1,313 4, ,544 1, ,244 5, Variable Spd Ductless 1,538 4, ,948 1, ,337 3, Var Spd - Lower flw 1,382 4, ,012 1, ,387 3, Houston Constant Speed 628 1, ,572 7,687 1, ,061 6, Two-Speed 620 3, ,452 7, ,209 4, Variable Spd Ductless 739 3, ,227 6, ,684 2, Var Spd - Lower flw 684 3, ,272 6, ,719 2, Atlanta Constant Speed 291 1, ,546 13, ,631 3,631 1,168 Two-Speed 295 2, ,946 12, ,628 2,628 1,062 Variable Spd Ductless 329 2, ,217 10, ,681 1, Nashville Constant Speed 4 1, ,938 10, ,069 4, Two-Speed 3 2, ,258 10, ,909 2, Variable Spd Ductless 4 2, ,400 7, ,827 1, Indianapolis Constant Speed - 1, ,960 18, ,013 3, Two-Speed - 1, ,492 18, ,206 2, Variable Spd Ductless - 1, , ,449 1, HERS 50 (Ducts in Space) Orlando Hours Above 60% RH Runtime (hrs) EER (Btu/Wh) Htg AHU Fan Exh Fan Electric w/o HT Electric w/o HT Costs w Furnace ($) Constant Speed 975 1, , ,226 3, Two-Speed 1,011 3, , ,211 2, Variable Spd Ductless 976 4, , ,007 2, Var Spd - Lower flw 880 4, , ,040 2, Miami Constant Speed 777 2, , ,257 4, Two-Speed 822 4, , ,893 2, Variable Spd Ductless 787 5, , ,633 2, Var Spd - Lower flw 670 5, , ,678 2, Houston Constant Speed 362 1, ,623 1, ,341 3, Two-Speed 380 3, ,963 1, ,347 2, Variable Spd Ductless 371 3, ,841 1, ,164 2, Var Spd - Lower flw 330 4, ,881 1, ,196 2, Atlanta Constant Speed 35 1, ,554 3, ,106 2, Two-Speed 40 2, ,149 3, ,488 1, Variable Spd Ductless 35 2, ,060 3, ,371 1, Nashville Constant Speed - 1, ,679 4, ,269 2, Two-Speed - 2, ,248 4, ,611 1, Variable Spd Ductless - 2, ,167 4, ,499 1, Indianapolis Constant Speed ,095 8, ,647 1, Two-Speed - 1, , ,213 1, Variable Spd Ductless - 1, , ,130 1,

73 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. 73

74 Orlando HERS 100, Conv Table 28. Performance Results with Enhanced Cooling Unit Hours Above 60% RH Runtime (hrs) EER (Btu/Wh) Htg AHU Fan Electric w/o HT Costs w Furnace ($) Conv 1,645 1, ,302 5,642 1,179 5, Enhanced (60% RH) 1,571 1, ,317 5,737 1,248 5, Enhanced (50% RH) 928 2, ,570 5,841 1,144 5,917 1,017 Miami Conv 1,303 2, ,232 1,513 1,572 8, Enhanced (60% RH) 1,205 2, ,256 1,517 1,570 8, Enhanced (50% RH) 471 2, ,513 1,588 1,478 8, Houston Conv 628 1, ,572 7,687 1,286 6, Enhanced (60% RH) 604 1, ,577 7,693 1,286 6, Enhanced (50% RH) 231 2, ,707 7,745 1,245 6, Atlanta Conv 291 1, ,546 13, ,631 1,168 Enhanced (60% RH) 291 1, ,546 13, ,631 1,168 Enhanced (50% RH) 183 1, ,576 13, ,656 1,170 Nashville Conv 4 1, ,938 10, , Enhanced (60% RH) 4 1, ,938 10, , Enhanced (50% RH) 2 1, ,947 10, , Indianapolis Conv - 1, ,960 18, , Enhanced (60% RH) - 1, ,960 18, , Enhanced (50% RH) - 1, ,964 18, , Orlando HERS 70, 2-Spd Hours Above 60% RH Runtime (hrs) EER (Btu/Wh) Htg AHU Fan Electric w/o HT Costs w Furnace ($) Conv 575 3, , , Enhanced (60% RH) 571 3, , , Enhanced (50% RH) 308 3, , , Miami Conv 382 5, , , Enhanced (60% RH) 337 5, , , Enhanced (50% RH) 20 5, , , Houston Conv 191 3, ,856 2, , Enhanced (60% RH) 188 3, ,856 2, , Enhanced (50% RH) 53 4, ,914 2, , Atlanta Conv 15 2, ,815 4, , Enhanced (60% RH) 15 2, ,815 4, , Enhanced (50% RH) 15 2, ,819 4, , Nashville Conv - 2, ,859 5, , Enhanced (60% RH) - 2, ,859 5, , Enhanced (50% RH) - 2, ,863 5, , Indianapolis Conv - 1, ,206 12, , Enhanced (60% RH) - 1, ,206 12, , Enhanced (50% RH) - 1, ,208 12, ,

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 Daily Indoor Space Conditions: z1h100s1rh50v Daily Indoor Space Conditions: z1h100s2rh50v Hours Above 50% = 5361 Hours Above 55% = 2399 Hours Above 60% = 1303 Hours Above 65% = 543 Hours Above 70% = 107 All Hrs Cooling Hrs 80% 70% 60% Humidity Ratio (lb/lb) Hours Above 50% = 3939 Hours Above 55% = 1520 Hours Above 60% = 471 Hours Above 65% = 37 Hours Above 70% = 0 All Hrs Cooling Hrs 80% 70% 60% Impact of Enhanced Control Humidity Ratio (lb/lb) 50% 50% 40% 40% Dry Bulb Temperature (F) Dry Bulb Temperature (F) 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. HERS 100 Single Spd Houston Table 29. Performance Results with Various Cooling Unit Enhancements Hours Above 60% RH Runtime (hrs) EER (Btu/Wh) Htg AHU Fan Electric w/o HT Costs w Furnace ($) S1 - Conv 628 1, ,572 7,687 1,286 6, S10 - Lower Airflow 476 2, ,617 7,699 1,227 6, S11 - Overcooling 326 2, ,713 7,766 1,328 6, S2 - Both 231 2, ,707 7,745 1,245 6, HERS 70 2-Stage S1 - Conv 191 3, ,856 2, , S10 - Lower Airflow 112 3, ,893 2, , S11 - Overcooling 77 4, ,376 2, , S2 - Both 53 4, ,914 2, , 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 75

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 2 o F 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). 76

77 Table 30. Performance Results for Cooling Units with Further Enhancements, HERS 100 Orlando HERS 100 Hours Above 60% RH Runtime (hrs) EER (Btu/Wh) Htg AHU Fan Electric w/o HT Costs w Furnace ($) Conv 1,645 1, ,302 5,642 1,179 5, % Enhanced 928 2, ,570 5,841 1,144 5,917 1, % w/ HPs 1,162 2, ,808 5,649 1,423 6,434 1, % w/ sc-rheat 205 2, ,840 5,836 1,394 7,437 1, % w/ full reheat - 2, ,202 6,022 1,463 6,868 1, % Miami Conv 1,303 2, ,232 1,513 1,572 8, % Enhanced 471 2, ,513 1,588 1,478 8, % w/ HPs 758 3, ,953 1,515 1,926 9,082 1, % w/ sc-rheat 22 3, ,702 1,563 1,709 9,615 1, % w/ full reheat - 3, ,987 1,628 1,811 9,001 1, % Houston Conv 628 1, ,572 7,687 1,286 6, % Enhanced 231 2, ,707 7,745 1,245 6, % w/ HPs 338 2, ,050 7,694 1,536 6, % w/ sc-rheat 9 2, ,397 7,742 1,365 6, % w/ full reheat - 2, ,997 7,786 1,450 6, % Atlanta Conv 291 1, ,546 13, ,631 1, % Enhanced 183 1, ,576 13, ,656 1, % w/ HPs w/ sc-rheat 54 1, ,772 13, ,879 1, % w/ full reheat 1 1, ,682 13, ,847 1, % Nashville Conv 4 1, ,938 10, , % Enhanced 2 1, ,947 10, , % w/ HPs w/ sc-rheat - 1, ,061 10, , % w/ full reheat - 1, ,004 10,144 1,010 4, % Indianapolis Conv - 1, ,960 18, , % Enhanced - 1, ,964 18, , % w/ HPs w/ sc-rheat - 1, ,031 18, , % w/ full reheat - 1, ,998 18, , % 77

78 Table 31. Performance Results for Cooling Units with Further Enhancements, HERS 70 (two-speed) Orlando HERS 70 Hours Above 60% RH Runtime (hrs) EER (Btu/Wh) Htg AHU Fan Electric w/o HT Costs w Furnace ($) Conv 575 3, , , % Enhanced 308 3, , , % w/ HPs 418 2, , ,126 5, % w/ sc-rheat 32 4, , , % w/ full reheat - 4, , , % Miami Conv 382 5, , , % Enhanced 20 5, , , % w/ HPs 46 3, , ,482 6, % w/ sc-rheat - 5, , , % w/ full reheat - 6, , , % Houston Conv 191 3, ,856 2, , % Enhanced 53 4, ,914 2, , % w/ HPs 93 2, ,837 2,044 1,157 5, % w/ sc-rheat - 4, ,233 2, , % w/ full reheat - 4, ,000 2, , % Atlanta Conv 15 2, ,815 4, , % Enhanced 15 2, ,819 4, , % w/ HPs w/ sc-rheat - 2, ,879 4, , % w/ full reheat - 2, ,856 4, , % Nashville Conv - 2, ,859 5, , % Enhanced - 2, ,863 5, , % w/ HPs w/ sc-rheat - 2, ,914 5, , % w/ full reheat - 2, ,890 5, , % Indianapolis Conv - 1, ,206 12, , % Enhanced - 1, ,208 12, , % w/ HPs w/ sc-rheat - 1, ,239 12, , % w/ full reheat - 1, ,225 12, , % 78

79 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. operating costs go up by about $ per year compared to cooling only in the hot-humid climates. Set Point = 60% RH Set Point = 50% RH Daily Indoor Space Conditions: z1h100s5rh60v Daily Indoor Space Conditions: z1h100s5rh50v Hours Above 50% = 5508 Hours Above 55% = 2186 Hours Above 60% = 598 Hours Above 65% = 15 Hours Above 70% = 0 All Hrs Cooling Hrs 80% 70% 60% 50% Humidity Ratio (lb/lb) Hours Above 50% = 2968 Hours Above 55% = 168 Hours Above 60% = 0 Hours Above 65% = 0 Hours Above 70% = 0 All Hrs Cooling Hrs 80% 70% 60% 50% Humidity Ratio (lb/lb) 40% % Dry Bulb Temperature (F) Dry Bulb Temperature (F) Figure 30. Psychrometric Charts Comparing the Degree of Humidity Control at Two RH Set Points for Standard DH unit, HERS 100, Miami 79

80 Table 32. Performance Results with Standalone DH (System 5) with Different Set Points Orlando HERS 100 Hours Above 60% RH Runtime (hrs) EER (Btu/Wh) Htg AHU Fan DH Exh Fan Electric w/o HT Costs w Furnace ($) Conv 1,645 1, ,302 5,642 1, , Standalone DH 60% 664 1, ,332 5,430 1, ,291 1,026 Standalone DH 50% - 2, ,627 5,060 1,492 1, ,110 1,183 Miami Conv 1,303 2, ,232 1,513 1, , Standalone DH 60% 598 2, ,268 1,356 1, , Standalone DH 50% - 2, ,540 1,214 1,778 1, ,949 1,092 Houston Conv 628 1, ,572 7,687 1, , Standalone DH 60% 301 1, ,592 7,608 1, , Standalone DH 50% - 2, ,728 7,482 1, , Atlanta Conv 291 1, ,546 13, ,631 1,168 Standalone DH 60% 103 1, ,551 12,969 1, ,873 1,190 Standalone DH 50% - 1, ,573 12,893 1, ,155 1,215 Nashville Conv 4 1, ,938 10, , Standalone DH 60% 4 1, ,932 10,202 1, , Standalone DH 50% - 1, ,947 10,193 1, , Indianapolis Conv - 1, ,960 18, , Standalone DH 60% 2 1, ,954 18,216 1, , Standalone DH 50% - 1, ,963 18,211 1, , ,400 Dehumidifier Set Point Costs w Furnace ($) 1,200 1, Z0 Orlando Z1 Miami Z2 Houston Z3 Atlanta Z4 Nashville Z5 Indianapolis Conv DH 60% RH DH 50% RH Figure 31. Operating Costs for Standalone DH at Various RH Set Points, HERS 100 House 80

81 Table 33. Humidity Threshold Results for Standalone DH (System 5) with Different Set Points Orlando HERS 100 Hours Above 60% RH Hours Above 55% RH Hours Above 50% RH Conv 1,645 3,074 6,737 Standalone DH 60% 664 2,588 6,660 Standalone DH 50% ,428 Miami Conv 1,303 2,399 5,361 Standalone DH 60% 598 2,186 5,508 Standalone DH 50% ,968 Houston Conv 628 1,500 3,716 Standalone DH 60% 301 1,277 3,693 Standalone DH 50% ,835 Atlanta Conv ,479 Standalone DH 60% ,313 Standalone DH 50% Nashville Conv Standalone DH 60% Standalone DH 50% Indianapolis Conv Standalone DH 60% Standalone DH 50%

82 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. 82

83 Table 34. Performance Results with Various DH Units (Systems, 5, 6, 7, 13 & 14) with HERS 100 HERS 100, 50% RH Set Pt Orlando Hours Above 55% RH Runtime (hrs) EER (Btu/Wh) Htg AHU Fan DH DH Fan Exh Fan DH Gas Use (therms) Electric w/o HT Costs w Furnace ($) S1 - Conv 3,074 1, ,302 5,642 1, , S5 - DH Unit 314 2, ,627 5,060 1,492 1, ,110 1,183 S6 - Ducted DH 219 1, ,582 5,097 1,483 1, ,916 1,166 S7 - Vent DH 3 1, ,479 4,626 1, , ,865 1,127 S13 - Gas-fired DES - 2, ,662 5,354 1, ,118 1,509 S14 - Cond DES - 1, ,173 5,429 1,406 1, ,749 1,173 Miami S1 - Conv 2,399 2, ,232 1,513 1, , S5 - DH Unit 168 2, ,540 1,214 1,778 1, ,949 1,092 S6 - Ducted DH 116 2, ,496 1,226 1,769 1, ,773 1,075 S7 - Vent DH 5 2, ,403 1,079 1, , ,915 1,079 S13 - Gas-fired DES - 2, ,694 1,317 1, ,497 1,483 S14 - Cond DES - 2, ,139 1,377 1,696 1, ,565 1,065 Houston S1 - Conv 1,500 1, ,572 7,687 1, , S5 - DH Unit 51 2, ,728 7,482 1, , S6 - Ducted DH 35 2, ,708 7,494 1, , S7 - Vent DH 6 2, ,680 7,115 1, , , S13 - Gas-fired DES - 2, ,820 8,326 1, ,026 1,116 S14 - Cond DES - 1, ,527 7,590 1, , Atlanta S1 - Conv 722 1, ,546 13, ,631 1,168 S5 - DH Unit 15 1, ,573 12,893 1, ,155 1,215 S6 - Ducted DH 14 1, ,570 12,902 1, ,137 1,214 S7 - Vent DH - 1, ,616 12,250 1, , ,910 1,253 S13 - Gas-fired DES - 1, ,549 13,594 1, ,936 1,280 S14 - Cond DES - 1, ,540 12,937 1, ,115 1,213 Nashville S1 - Conv 105 1, ,938 10, , S5 - DH Unit 1 1, ,947 10,193 1, , S6 - Ducted DH - 1, ,945 10,187 1, , S7 - Vent DH - 1, ,981 9,829 1, , , S13 - Gas-fired DES - 1, ,901 11,065 1, , S14 - Cond DES - 1, ,924 10,191 1, , Indianapolis S1 - Conv 20 1, ,960 18, , S5 - DH Unit - 1, ,963 18,211 1, , S6 - Ducted DH - 1, ,962 18,208 1, , S7 - Vent DH - 1, ,018 17,590 1, , , S13 - Gas-fired DES - 1, ,912 19,483 1, , S14 - Cond DES - 1, ,947 18,210 1, , ,600 Comparing DH Units, HERS 100 Costs w Furnace ($) 1,400 1,200 1, Z0 Orlando Z1 Miami Z2 Houston Z3 Atlanta Z4 Nashville Z5 Indianapolis Conv Standalone DH Ducted DH Vent DH Gas fired DES Cond DES Figure 32. Operating Costs All DH Options, 50% RH Set Point, HERS 100 House 83

84 Table 35. Performance Results with Various DH Units (Systems, 5, 6, 7, 13 & 14) with HERS 70 and 50% RH Setpoint HERS 70, 50% RH Set Pt Orlando Hours Above 55% RH Runtime (hrs) EER (Btu/Wh) Htg AHU Fan DH DH Fan Exh Fan DH Gas Use (therms) Electric w/o HT Costs w Furnace ($) S3-2-Spd 1,904 3, , , S5 - DH Unit - 3, , , , S6 - Ducted DH - 3, , , S7 - Vent DH - 3, , , , S13 - Gas-fired DES - 3, , , S14 - Cond DES - 3, , , Miami S3-2-Spd 1,429 5, , , S5 - DH Unit - 5, , , S6 - Ducted DH - 5, , , S7 - Vent DH - 5, , , , S13 - Gas-fired DES - 5, , , S14 - Cond DES - 5, , , Houston S3-2-Spd 732 3, ,856 2, , S5 - DH Unit - 3, ,912 2, , S6 - Ducted DH - 3, ,904 2, , S7 - Vent DH - 3, ,885 2, , , S13 - Gas-fired DES - 3, ,864 2, , S14 - Cond DES - 3, ,824 2, , Atlanta S3-2-Spd 240 2, ,815 4, , S5 - DH Unit - 2, ,812 4, , S6 - Ducted DH - 2, ,812 4, , S7 - Vent DH - 2, ,801 4, , , S13 - Gas-fired DES - 2, ,742 5, , S14 - Cond DES - 2, ,803 4, , Nashville S3-2-Spd 33 2, ,859 5, , S5 - DH Unit - 2, ,855 6, , S6 - Ducted DH - 2, ,855 6, , S7 - Vent DH - 2, ,846 5, , , S13 - Gas-fired DES - 2, ,795 6, , S14 - Cond DES - 2, ,845 6, , Indianapolis S3-2-Spd 11 1, ,206 12, , S5 - DH Unit - 1, ,201 12, , S6 - Ducted DH - 1, ,201 12, , S7 - Vent DH - 1, ,203 12, , , S13 - Gas-fired DES - 1, ,147 13, , S14 - Cond DES - 1, ,195 12, ,

85 700 Comparing DH Units, HERS 70 Costs w Furnace ($) Z0 Orlando Z1 Miami Z2 Houston Z3 Atlanta Z4 Nashville Z5 Indianapolis Conv Standalone DH Ducted DH Vent DH Gas fired DES Cond DES Figure 33. Operating Costs All DH Options, 50% RH Set Point, HERS 70 House 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. 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. 85

86 HERS 100, 50% RH Set Pt Orlando Table 36. Performance Results with Best DH and Enhanced Units, HERS 100 Hours Above 60% RH Runtime (hrs) EER (Btu/Wh) Htg AHU Fan DH DH Fan Electric w/o HT Costs w Furnace ($) S1 - Conv 1,645 1, ,302 5,642 1, , % S2 - low flw 928 2, ,570 5,841 1, ,917 1, % S8 - Partial SC/RH 205 2, ,840 5,836 1, ,437 1, % S9 - Full RH - 2, ,202 6,022 1, ,868 1, % S6 - Ducted DH - 1, ,582 5,097 1,483 1, ,916 1, % S14 - Cond DES - 1, ,173 5,429 1,406 1, ,749 1, % Miami S1 - Conv 1,303 2, ,232 1,513 1, , % S2 - low flw 471 2, ,513 1,588 1, , % S8 - Partial SC/RH 22 3, ,702 1,563 1, ,615 1, % S9 - Full RH - 3, ,987 1,628 1, ,001 1, % S6 - Ducted DH - 2, ,496 1,226 1,769 1, ,773 1, % S14 - Cond DES - 2, ,139 1,377 1,696 1, ,565 1, % Houston S1 - Conv 628 1, ,572 7,687 1, , % S2 - low flw 231 2, ,707 7,745 1, , % S8 - Partial SC/RH 9 2, ,397 7,742 1, , % S9 - Full RH - 2, ,997 7,786 1, , % S6 - Ducted DH - 2, ,708 7,494 1, , % S14 - Cond DES - 1, ,527 7,590 1, , % Atlanta S1 - Conv 291 1, ,546 13, ,631 1, % S2 - low flw 183 1, ,576 13, ,656 1, % S8 - Partial SC/RH 54 1, ,772 13, ,879 1, % S9 - Full RH 1 1, ,682 13, ,847 1, % S6 - Ducted DH - 1, ,570 12,902 1, ,137 1, % S14 - Cond DES - 1, ,540 12,937 1, ,115 1, % 1,400 Best Technologies, HERS 100 Costs w Furnace ($) 1,200 1, Z0 Orlando Z1 Miami Z2 Houston Z3 Atlanta Z4 Nashville Z5 Indianapolis Conv Enhanced w/ SC/Rht w/ Full Rht Ducted DH Cond DES Figure 34. Operating Costs Best Technologies, HERS 100 House 86

87 Table 37. Performance Results with Best DH and Enhanced Units, HERS 70 (two-speed) HERS 70, 50% RH Set Pt Orlando Hours Above 60% RH Runtime (hrs) EER (Btu/Wh) Htg AHU Fan DH DH Fan Electric w/o HT Costs w Furnace ($) S3-2-Spd 575 3, , , % S2 - low flw 308 3, , , % S8 - Partial SC/RH 32 4, , , % S9 - Full RH - 4, , , % S6 - Ducted DH - 3, , , % S14 - Cond DES - 3, , , % Miami S3-2-Spd 382 5, , , % S2 - low flw 20 5, , , % S8 - Partial SC/RH - 5, , , % S9 - Full RH - 6, , , % S6 - Ducted DH - 5, , , % S14 - Cond DES - 5, , , % Houston S3-2-Spd 191 3, ,856 2, , % S2 - low flw 53 4, ,914 2, , % S8 - Partial SC/RH - 4, ,233 2, , % S9 - Full RH - 4, ,000 2, , % S6 - Ducted DH - 3, ,904 2, , % S14 - Cond DES - 3, ,824 2, , % Atlanta S3-2-Spd 15 2, ,815 4, , % S2 - low flw 15 2, ,819 4, , % S8 - Partial SC/RH - 2, ,879 4, , % S9 - Full RH - 2, ,856 4, , % S6 - Ducted DH - 2, ,812 4, , % S14 - Cond DES - 2, ,803 4, , % 600 Best Technologies, HERS 70 Costs w Furnace ($) Z0 Orlando Z1 Miami Z2 Houston Z3 Atlanta Z4 Nashville Z5 Indianapolis Conv Enhanced w/ SC/Rht w/ Full Rht Ducted DH Cond DES Figure 35. Operating Costs Best Technologies, HERS 70 House 87

88 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 Small House Base House Large House Floor Area (ft2) Area Ratio Ventilation (cfm) Vent Ratio Cooling Size (tons) Houston HERS Nashville HERS Houston HERS Nashville HERS Heating (MBtu/h) HERS HERS 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. 88

89 Houston HERS 100 Table 40. Impact of House Size on Humidity Levels and Use Hours Above 60% RH Runtime (hrs) EER (Btu/Wh) Htg AHU Fan DH Exh Fan Electric w/o HT Costs w Furnace ($) Small House 660 1, ,964 4, , Normal House 628 1, ,572 7,687 1, , Large House 686 1, ,889 9,932 1, ,802 1,094 Nashville Small House 5 1, ,890 6, , Normal House 4 1, ,938 10, , Large House 3 1, ,578 15,073 1, ,027 1,123 HERS 70 Houston Small House 228 2, ,707 1, , Normal House 191 3, ,856 2, , Large House 271 3, ,660 2, , Nashville Small House - 1, ,119 3, , Normal House - 2, ,859 5, , Large House - 2, ,303 9, , 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. 89

90 24 Window Opening - z1h100s1rh50v1w ( hrs) Hour of Day Jan Feb Mar Apr May Jun Jul Aug Sep Oct Nov Dec 2006 Day (MAX/MIN = 1.00/ 0.00 ) 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 Daily Indoor Space Conditions: z1h100s1rh50v Daily Indoor Space Conditions: z1h100s1rh50v1w Hours Above 50% = 5361 Hours Above 55% = 2399 Hours Above 60% = 1303 Hours Above 65% = 543 Hours Above 70% = 107 All Hrs Cooling Hrs 80% 70% 60% 50% Humidity Ratio (lb/lb) Hours Above 50% = 5408 Hours Above 55% = 2552 Hours Above 60% = 1359 Hours Above 65% = 592 Hours Above 70% = 128 All Hrs Cooling Hrs 80% 70% 60% 50% Humidity Ratio (lb/lb) 40% 40% Dry Bulb Temperature (F) Dry Bulb Temperature (F) Figure 37. Psychrometric Plots Showing Impact of Operable Windows, HERS 100 House 90

91 Table 41. Impact of Window Openings on Humidity Levels and Use HERS 100, Exh Fan HERS 100, Exh vent Orlando Hours Above 60% RH Runtime (hrs) EER (Btu/Wh) Htg AHU Fan Exh Fan Electric w/o HT Costs w Furnace ($) Windows Closed 1,645 1, ,302 5,642 1, , Win. Operable 1,729 1, ,314 5,709 1, , Miami Windows Closed 1,303 2, ,232 1,513 1, , Win. Operable 1,359 2, ,233 1,553 1, , Houston Windows Closed 628 1, ,572 7,687 1, , Win. Operable 825 1, ,563 7,764 1, , Atlanta Windows Closed 291 1, ,546 13, ,631 1,168 Win. Operable 532 1, ,499 13, ,552 1,168 Nashville Windows Closed 4 1, ,938 10, , Win. Operable 132 1, ,862 10, , Indianapolis Windows Closed - 1, ,960 18, , Win. Operable 13 1, ,866 18, , Table 42. Impact of Window Openings on Humidity Levels and Use HERS 130, No Ventilation HERS 130, no vent Orlando Hours Above 60% RH Runtime (hrs) EER (Btu/Wh) Htg AHU Fan Exh Fan Electric w/o HT Costs w Furnace ($) Windows Closed 1,296 2, ,307 6,434 1,300-7,606 1,260 Win. Operable 1,435 2, ,340 6,551 1,307-7,647 1,273 Miami Windows Closed 1,059 2, ,064 1,798 1,715-10,779 1,226 Win. Operable 1,286 2, ,078 1,834 1,718-10,795 1,230 Houston Windows Closed 778 2, ,530 13,589 1,498-8,028 1,309 Win. Operable 1,109 2, ,542 13,662 1,501-8,043 1,313 Atlanta Windows Closed 314 1, ,455 17, ,364 1,573 Win. Operable 582 1, ,413 17, ,316 1,579 Nashville Windows Closed 25 1, ,119 17,533 1,103-5,222 1,292 Win. Operable 247 1, ,048 17,735 1,092-5,140 1,293 Indianapolis Windows Closed - 1, ,058 20, ,047 1,083 Win. Operable 4 1, ,905 21, ,868 1,079 91

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 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). 100 Attic Dew Pt (F) : 0: 2: 4: 6: 8: 10: 12: 14: 16: 18: 20: 22: 0: June 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. 92

93 Table 43. Impact of Attic Dew Point Bump on Humidity Levels and Use HERS 100, Exh Fan HERS 100, Exh vent Orlando Hours Above 60% RH Runtime (hrs) EER (Btu/Wh) Htg AHU Fan Exh Fan Electric w/o HT Costs w Furnace ($) Normal 1,645 1, ,302 5,642 1, , w/ Dewpt "Bump" 1,647 1, ,329 5,638 1, , Miami Normal 1,303 2, ,232 1,513 1, , w/ Dewpt "Bump" 1,307 2, ,271 1,513 1, , Houston Normal 628 1, ,572 7,687 1, , w/ Dewpt "Bump" 630 1, ,600 7,687 1, , Atlanta Normal 291 1, ,546 13, ,631 1,168 w/ Dewpt "Bump" 292 1, ,559 13, ,647 1,169 Nashville Normal 4 1, ,938 10, , w/ Dewpt "Bump" 4 1, ,957 10, , Indianapolis Normal - 1, ,960 18, , w/ Dewpt "Bump" - 1, ,969 18, , Table 44. Impact of Attic Dew Point Bump on Humidity Levels and Use HERS 130, No Ventilation HERS 130, no vent Orlando Hours Above 60% RH Runtime (hrs) EER (Btu/Wh) Htg AHU Fan Exh Fan Electric w/o HT Costs w Furnace ($) Normal 1,296 2, ,307 6,434 1,300-7,606 1,260 w/ Dewpt "Bump" 1,299 2, ,379 6,434 1,310-7,689 1,268 Miami Normal 1,059 2, ,064 1,798 1,715-10,779 1,226 w/ Dewpt "Bump" 1,074 2, ,167 1,799 1,730-10,897 1,238 Houston Normal 778 2, ,530 13,589 1,498-8,028 1,309 w/ Dewpt "Bump" 790 2, ,594 13,584 1,506-8,100 1,314 Atlanta Normal 314 1, ,455 17, ,364 1,573 w/ Dewpt "Bump" 319 1, ,478 17, ,390 1,576 Nashville Normal 25 1, ,119 17,533 1,103-5,222 1,292 w/ Dewpt "Bump" 31 1, ,162 17,532 1,109-5,271 1,297 Indianapolis Normal - 1, ,058 20, ,047 1,083 w/ Dewpt "Bump" - 1, ,085 20, ,078 1,086 93

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. 3,500 Orlando: Moisture & Sensible Gains 3,000 Hours Above 60% RH 2,500 2,000 1,500 1, kwh/day 21.3 kwh/day 32 kwh/day 6 lb/day 12 lb/day 18 lb/day 24 lb/day Figure 39. Impact of Moisture and Sensible Heat Gains - HERS100 with Exhaust Fan Conclusions Several hundred annual simulation runs were completed for these various scenarios. The detailed results are available at a project website 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. 94

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 hothumid 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. 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. 95

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 2 o F 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 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 78 o F. The heating set point was 70 o F except in the hot-humid climates where it was 72 o F; raising the heat set point even more continued to reduce high RH hours. The Choice of Relative Humidity Set Point Affects 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 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 HV 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. 96

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. 97

98 100 Indoor Conditions Ft. Myers site Temperature (F) or Relative Humidity (%) /2000 9/2000 9/2000 9/ / / / / /2000 Indoor RH Indoor T 100 Outdoor Conditions Ft. Myers site 90 Temperature (F) /2000 9/2000 9/2000 9/ / / / / /2000 Outdoor T Outdoor Tdp 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 98

99 exchange was based on a constant 50 cfm, as that relates to the ASHRAE Standard 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. 1.8 Net moisture load for hours above 60% RH indoors Ft. Myers site Moisture load (lb/h) (0.2) (0.4) 8/2000 9/ / / /2000 Infil moisture Internal generation moisture 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 Star dehumidifier 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. 99

100 Indoor RH and predicted dehumidifier energy Ft. Myers site Indoor Relative Humidity (%) Hourly Dehumidifier /2000 9/ / / /2000 Indoor RH Deh 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. 100

101 Table 45. First-Cost Estimates for Supplemental Dehumidification Systems Supplemental Dehumidification System First-Cost Estimate Stand-alone Dehumidifier with Remote Dehumidistat Integrated Ducted Dehumidifier $300 $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 101

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 2 o F 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 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 (80 o F) 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. 102

103 Acknowledgements This project was supported by the U.S. Department of, Office of Building Technologies, Building America Program, with contract management by the National Renewable 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. References Arasteh, D., C. Kohler, and B. Griffith Modeling Windows in Plus with Simple Performance Indices (LBNL, NREL, October) ceindices.pdf. ASHRAE ASHRAE Handbook Fundamentals, Chapter 16 (Ventilation and Infiltration), pp Atlanta, GA: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. BSC (2007). Whole House Ventilation System Options Phase 1 Simulation Study. Building Science Corporation, ARTI Report No , Final Report, March. Air-Conditioning and Refrigeration Technology Institute, Arlington, VA. BSC Systems Engineering Approach To Development Of Advanced Residential Buildings, 11.B.1 Results Of Advanced System Research. Building Science Corporation, Somerville, MA. Project 6 Enhanced Dehumidifying Air Conditioner of Final Report to U.S. Dept. of under Task Order No. Kaax Under Task Ordering Agreement No. Kaax Midwest Research Institute, National Renewable Laboratory Division, Golden, CO. Fugler, D Conclusions from Ten Years of Canadian Attic Research. ASHRAE Transactions. CH Vol Pt. 1. P 819. Gu, L Personal communication between A. Rudd and Lixing Gu. May Henderson, H.I., 'Simulating Combined Thermostat, Air Conditioner and Building Performance in a House,' ASHRAE Transactions, Vol. 98 Part 1, January Henderson, H.I. and J. Sand 'An Hourly Building Simulation Tool to Evaluate Hybrid Desiccant System Configuration Options. KC , ASHRAE Transactions, Vol Pt. 2. June. Henderson, H.I., D.B. Shirey and R. Raustad Closing the Gap: Getting Full Performance from Residential Central Air Conditioners, Task 4 Develop New Climate-Sensitive Air Conditioner, Simulation Results and Cost Benefit Analysis, Final Report, FSEC-CR Cocoa, FL: Florida Solar Center. Henderson, H.I., D.B. Shirey and C.K. Rice Can Conventional Cooling Equipment Meet Dehumidification Needs for Homes in Humid Climates, Proceedings of the EEE Summer Study on Efficiency in Buildings. Washington, D.C.: American Council for an -Efficient Economy. 103

104 Hendron, R Building America Research Benchmark Definition. Technical Report, NREL/TP , Updated December 19, 2008, National Renewable Laboratory, Golden, CO. Larson, B Informal data exchange from test run on a mini-split system at Purdue University, Herrick Labs. Ecotope Inc., Seattle, WA. Proctor, J. and J. Piro System Optimization of Residential Ventilation, Space Conditioning, and Thermal Distribution. ARTI-21CR/ RESNET Mortgage Industry National Home Rating Systems Standards. Residential Services Network, Inc., Oceanside, CA. January. Rudd, Armin, Joseph Lstiburek and Kohta Ueno Residential dehumidification and ventilation systems research for hot-humid climates, Proceedings of 24th AIVC and BETEC Conference, Ventilation, Humidity Control, and, Washington, US, pp October. Air Infiltration and Ventilation Centre, Brussels, Belgium. Rudd, Armin Results of Advanced Systems Research, Deliverable Number 5.C.1 Project 3 Supplemental Humidity Control Systems, pg of Final Report to U.S. Dept. of under Task Order No. KAAX under Task Ordering Agreement No. KAAX , pg. 8-10, October 29. Midwest Research Institute, National Renewable Laboratory Division, Golden, CO. Rudd, A., J. Lstiburek, and K. Ueno Residential dehumidification systems research for hot-humid climates. U.S. Department of, Efficiency and Renewable, NREL/SR Rudd, Armin Systems Engineering Approach To Development Of Advanced Residential Buildings, 11.B.1 Results Of Advanced System Research. Project 6 Enhanced Dehumidifying Air Conditioner of Final Report to U.S. Dept. of under Task Order No. Kaax under Task Ordering Agreement No. Kaax Midwest Research Institute, National Renewable Laboratory Division, 1617 Cole Boulevard, Golden, CO. Rudd, A. and H.I. Henderson Monitored Indoor Moisture and Temperature Conditions in Humid Climate U.S. Residences. DA ASHRAE Transactions Vol Pt. 1. January. Rudd, Armin 2007(b). Systems Engineering Approach To Development Of Advanced Residential Buildings, 14.B.1 Results Of Advanced Systems Research, Project 1 Enhanced Dehumidifying Air Conditioning of Final Report To U.S. Dept. of under Task Order No. Kaax under Task Ordering Agreement No. Kaax Midwest Research Institute, National Renewable Laboratory Division, 1617 Cole Boulevard, Golden, CO. Rudd, Armin, Hugh I. Henderson, Jr., Daniel Bergey, Don B. Shirey ASHRAE 1449-RP: Efficiency and Cost Assessment of Humidity Control Options for Residential Buildings. Research Project Final Report submitted to American Society of Heating Refrigeration and Air-Conditioning Engineers, Atlanta, GA. Shirey, D.B., H.I. Henderson and R. Raustad Understanding the Dehumidification Performance of Air-Conditioning Equipment at Part-Load Conditions, Final Report, FSEC-CR Cocoa, FL: Florida Solar Center. pdf/fsec-cr pdf. Shirey, D. B. and S. Carlson Appendix to Documentation for IHAT Simulation Software developed for EPA. 104

105 Thermastor communication with factory engineers. Therma-Stor LLC, Madison, WI. Walker, I. and D. Wilson Field Validation of Algebraic Equations for Stack and Wind Driven Air Infiltration Calculations. International Journal of HV&R Research (now ASHRAE HV&R Research Journal), Vol. 4, No. 2, April. 105

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

107 Table A 1. Supplemental dehumidification energy and cost over a conventional cooling system, for a HERS 50 house with different mechancial ventilation systems HERS 50 Exhaust Ventilation CFIS Ventilation ERV Ventilation RH Setpoint: 60% 60% 50% 50% 60% 60% 50% 50% 60% 60% 50% 50% Dehumification over Conv. System Dehumification Cost over Conv. System ($) Dehumification over Conv. System Dehumification Cost over Conv. System ($) Dehumification over Conv. System Dehumification Cost over Conv. System ($) Dehumification over Conv. System Dehumification Cost over Conv. System ($) Dehumification over Conv. System Dehumification Cost over Conv. System ($) Dehumification over Conv. System Dehumification Cost over Conv. System ($) Orlando Conv or 2-spd Low flow+overcool Subcool Reheat Full Cond Reheat Standalone DH Ducted DH Cond Desiccant Miami Conv or 2-spd Low flow+overcool Subcool Reheat Full Cond Reheat Standalone DH Ducted DH Cond Desiccant Houston Conv or 2-spd Low flow+overcool Subcool Reheat Full Cond Reheat Standalone DH Ducted DH Cond Desiccant Atlanta Conv or 2-spd Low flow+overcool Subcool Reheat Full Cond Reheat Standalone DH Ducted DH Cond Desiccant This system does not control to a RH setpoint 2 These systems will attempt to control to an RH setpoint but will not always meet the setpoint due to a limit on overcooling 3 This system will control to a RH setpoint but may not have the capacity to always meet the setpoint, especially without whole-house mixing 4 These systems will control to a RH setpoint and will meet the setpoint 107

108 Table A 2. Supplemental dehumidification energy and cost over a conventional cooling system, for a HERS 70 house with different mechanical ventilation systems Exhaust Ventilation CFIS Ventilation ERV Ventilation RH Setpoint: 60% 60% 50% 50% 60% 60% 50% 50% 60% 60% 50% 50% Dehumification Dehumification Dehumification Dehumification Dehumification Dehumification Dehumification Dehumification Dehumification Dehumification Dehumification Dehumification HERS 70 over Cost over over Cost over over Cost over over Cost over over Cost over over Cost over Conv. System Conv. System Conv. System Conv. System Conv. System Conv. System Conv. System Conv. System Conv. System Conv. System Conv. System Conv. System ($) ($) ($) ($) ($) ($) Orlando Conv or 2-spd Low flow+overcool Subcool Reheat Full Cond Reheat Standalone DH Ducted DH Cond Desiccant Miami Conv or 2-spd Low flow+overcool Subcool Reheat Full Cond Reheat Standalone DH Ducted DH Cond Desiccant Houston Conv or 2-spd Low flow+overcool Subcool Reheat Full Cond Reheat Standalone DH Ducted DH Cond Desiccant Atlanta Conv or 2-spd Low flow+overcool Subcool Reheat Full Cond Reheat Standalone DH Ducted DH Cond Desiccant This system does not control to a RH setpoint 2 These systems will attempt to control to an RH setpoint but will not always meet the setpoint due to a limit on overcooling 3 This system will control to a RH setpoint but may not have the capacity to always meet the setpoint, especially without whole-house mixing 4 These systems will control to a RH setpoint and will meet the setpoint 108

109 Table A 3. Supplemental dehumidification energy and cost over a conventional cooling system, for a HERS 85 house with different mechanical ventilation systems Exhaust Ventilation CFIS Ventilation ERV Ventilation RH Setpoint: 60% 60% 50% 50% 60% 60% 50% 50% 60% 60% 50% 50% Dehumification Dehumification Dehumification Dehumification Dehumification Dehumification Dehumification Dehumification Dehumification Dehumification Dehumification Dehumification HERS 85 over Cost over over Cost over over Cost over over Cost over over Cost over over Cost over Conv. System Conv. System Conv. System Conv. System Conv. System Conv. System Conv. System Conv. System Conv. System Conv. System Conv. System Conv. System ($) ($) ($) ($) ($) ($) Orlando Conv or 2-spd Low flow+overcool Subcool Reheat Full Cond Reheat Standalone DH Ducted DH Cond Desiccant Miami Conv or 2-spd Low flow+overcool Subcool Reheat Full Cond Reheat Standalone DH Ducted DH Cond Desiccant Houston Conv or 2-spd Low flow+overcool Subcool Reheat Full Cond Reheat Standalone DH Ducted DH Cond Desiccant Atlanta Conv or 2-spd Low flow+overcool Subcool Reheat Full Cond Reheat Standalone DH Ducted DH Cond Desiccant This system does not control to a RH setpoint 2 These systems will attempt to control to an RH setpoint but will not always meet the setpoint due to a limit on overcooling 3 This system will control to a RH setpoint but may not have the capacity to always meet the setpoint, especially without whole-house mixing 4 These systems will control to a RH setpoint and will meet the setpoint 109

110 Table A 4. Supplemental dehumidification energy and cost over a conventional cooling system, for a HERS 100 house with different mechanical ventilation systems Exhaust Ventilation CFIS Ventilation ERV Ventilation RH Setpoint: 60% 60% 50% 50% 60% 60% 50% 50% 60% 60% 50% 50% Dehumification Dehumification Dehumification Dehumification Dehumification Dehumification Dehumification Dehumification Dehumification Dehumification Dehumification Dehumification HERS 100 over Cost over over Cost over over Cost over over Cost over over Cost over over Cost over Conv. System Conv. System Conv. System Conv. System Conv. System Conv. System Conv. System Conv. System Conv. System Conv. System Conv. System Conv. System ($) ($) ($) ($) ($) ($) Orlando Conv or 2-spd Low flow+overcool Subcool Reheat Full Cond Reheat Standalone DH Ducted DH Cond Desiccant Miami Conv or 2-spd Low flow+overcool Subcool Reheat Full Cond Reheat Standalone DH Ducted DH Cond Desiccant Houston Conv or 2-spd Low flow+overcool Subcool Reheat Full Cond Reheat Standalone DH Ducted DH Cond Desiccant Atlanta Conv or 2-spd Low flow+overcool Subcool Reheat Full Cond Reheat Standalone DH Ducted DH Cond Desiccant This system does not control to a RH setpoint 2 These systems will attempt to control to an RH setpoint but will not always meet the setpoint due to a limit on overcooling 3 This system will control to a RH setpoint but may not have the capacity to always meet the setpoint, especially without whole-house mixing 4 These systems will control to a RH setpoint and will meet the setpoint 110

111 Dehumidification, Orlando, HERS 50, Exhaust Dehumidification, Orlando, HERS 50, CFIS Dehumidification, Orlando, HERS 50, ERV Dehumidification Over Conventional System Hours Above RH Setpoint Dehumidification Over Conventional System Hours Above RH Setpoint Dehumidification Over Conventional System Hours Above RH Setpoint kwh 60% RH kwh 50% RH h>60% RH w/setpt=60% h>60% RH w/setpt=50% kwh 60% RH kwh 50% RH h>60% RH w/setpt=60% h>60% RH w/setpt=50% kwh 60% RH kwh 50% RH h>60% RH w/setpt=60% h>60% RH w/setpt=50% Dehumidification Cost, HERS 50, Exhaust Dehumidification Cost, Orlando, HERS 50, CFIS Dehumidification Cost, Orlando, HERS 50, ERV Dehumidification Cost Over Conventional System ($) Hours Above RH Setpoint Dehumidification Cost Over Conventional System ($) Hours Above RH Setpoint Dehumidification Cost Over Conventional System ($) Hours Above RH Setpoint $ 60% RH $ 50% RH h>60% RH w/setpt=60% h>60% RH w/setpt=50% $ 60% RH $ 50% RH h>60% RH w/setpt=60% h>60% RH w/setpt=50% $ 60% RH $ 50% RH h>60% RH w/setpt=60% h>60% RH w/setpt=50% 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 111

112 Dehumidification, Miami, HERS 50, Exhaust Dehumidification, Miami, HERS 50, CFIS Dehumidification, Miami, HERS 50, ERV Dehumidification Over Conventional System Hours Above RH Setpoint Dehumidification Over Conventional System Hours Above RH Setpoint Dehumidification Over Conventional System Hours Above RH Setpoint kwh 60% RH kwh 50% RH h>60% RH w/setpt=60% h>60% RH w/setpt=50% kwh 60% RH kwh 50% RH h>60% RH w/setpt=60% h>60% RH w/setpt=50% kwh 60% RH kwh 50% RH h>60% RH w/setpt=60% h>60% RH w/setpt=50% Dehumidification Cost, Miami, HERS 50, Exhaust Dehumidification Cost, Miami, HERS 50, CFIS Dehumidification Cost, Miami, HERS 50, ERV Dehumidification Cost Over Conventional System ($) Hours Above RH Setpoint Dehumidification Cost Over Conventional System ($) Hours Above RH Setpoint Dehumidification Cost Over Conventional System ($) Hours Above RH Setpoint $ 60% RH $ 50% RH h>60% RH w/setpt=60% h>60% RH w/setpt=50% $ 60% RH $ 50% RH h>60% RH w/setpt=60% h>60% RH w/setpt=50% $ 60% RH $ 50% RH h>60% RH w/setpt=60% h>60% RH w/setpt=50% 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 112

113 Dehumidification, Houston, HERS 50, Exhaust Dehumidification, Houston, HERS 50, CFIS Dehumidification, Houston, HERS 50, ERV Dehumidification Over Conventional System Hours Above RH Setpoint Dehumidification Over Conventional System Hours Above RH Setpoint Dehumidification Over Conventional System Hours Above RH Setpoint kwh 60% RH kwh 50% RH h>60% RH w/setpt=60% h>60% RH w/setpt=50% kwh 60% RH kwh 50% RH h>60% RH w/setpt=60% h>60% RH w/setpt=50% kwh 60% RH kwh 50% RH h>60% RH w/setpt=60% h>60% RH w/setpt=50% Dehumidification Cost, Houston, HERS 50, Exhaust Dehumidification Cost, Houston, HERS 50, CFIS Dehumidification Cost, Houston, HERS 50, ERV Dehumidification Cost Over Conventional System ($) Hours Above RH Setpoint Dehumidification Cost Over Conventional System ($) Hours Above RH Setpoint Dehumidification Cost Over Conventional System ($) Hours Above RH Setpoint $ 60% RH $ 50% RH h>60% RH w/setpt=60% h>60% RH w/setpt=50% $ 60% RH $ 50% RH h>60% RH w/setpt=60% h>60% RH w/setpt=50% $ 60% RH $ 50% RH h>60% RH w/setpt=60% h>60% RH w/setpt=50% 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 113

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