Energy efficiency and cost assessment of humidity control options for residential buildings.
Traditionally, the conventional residential air-conditioning approach has been deemed adequate in terms of dehumidification for many situations and climates. In homes, air-conditioning systems with a conventional thermostat provide dehumidification as a consequence of sensible cooling. Since these systems are controlled based on a dry-bulb temperature setpoint, humidity is not explicitly controlled to a precise level but may still be maintained within an acceptable range (e.g., below 60%). However, as homes become very energy efficient and dedicated outdoor ventilation air is mechanically provided, this passive approach to humidity control may no longer be adequate.
Energy-efficient houses in warm, humid climates have low sensible heat gain. Low sensible heat gain is good for reducing cooling costs, but can contribute to part-load moisture control challenges (Rudd et al. 2003, Rudd et al. 2005). Especially during spring and fall seasons, as well as rainy periods and summer nights, there are a significant number of hours where little or no sensible cooling is needed but moisture removal is still needed due to internal moisture generation and outside air exchange. Conventional cooling equipment is designed mostly for reducing air temperature (sensible cooling), with only about 20% to 30% of its capacity typically devoted to removing moisture (latent cooling). Field measurements have shown that this results in periods of high indoor relative humidity in residences (Rudd and Henderson 2007).
Conventional air-conditioning systems both cool and dehumidify air as a result of system operation. Some enhanced cooling technologies are available to provide more moisture removal and less sensible cooling, for instance, by lowering supply airflow rates or continuing cooling operation below the cooling setpoint. These enhanced control approaches can reduce space humidity levels. However, these systems cannot solely dehumidify without providing sensible cooling. In efficient, low heat gain houses in warm, humid climates, there can be times when no cooling is needed but moisture removal is still required. In these cases, the solution has been to employ dehumidification systems that can remove moisture while supplying room neutral temperature or warmer air (e.g., a standalone room air dehumidifier that adds sensible heat to the space).
Separate or integrated (BSC 2006) dehumidification equipment can be employed to supplement the operation of a normal cooling system during periods when moisture removal alone is needed. The conventional wisdom assumes that the heat added to the space by a dehumidifier provides additional cooling load that increases cooling operation (Proctor and Piro 2005). However, simulation results from analysis by Henderson et al. (2008) showed that the energy penalty due to dehumidifier heat addition is smaller than expected because the dehumidifier often operates when the space dry-bulb temperature is below the cooling setpoint.
This research project developed and used a detailed building simulation model to accurately predict the annual energy use and humidity control performance of a typical residence with various mechanical systems, house configuration options, and climates. The simulation results provided the basis to compare the operating costs and humidity control performance of various system options. We also complete a range of sensitivity analysis to understand the situations and conditions when humidity control starts to become a concern.
The first step was to understand the current state of the art in dehumidification options for residential space conditioning through an extended literature search (BSC 2009). This process identified the most commonly used (or most promising) dehumidification approaches and technologies. The review also identified the information gaps in detailed performance data necessary to accurately simulate equipment performance.
The next step was to identify and understand any gaps or shortfalls in the ability of simulation models to accurately predict the performance of systems. The goal was to develop a robust simulation model to meet the stated objective of this research: to evaluate, assess, and rank the humidity control performance, energy efficiency, and cost of space conditioning systems for their capability to control indoor humidity throughout the year in humid climate residences.
Several model development needs were addressed in order to meet the objectives of this research effort. Some of the most important issues related to the ability to properly model (1) variable-speed ductless air-conditioning systems and ducted two-speed systems; (2) cooling systems using full condensing reheat; (3) cooling coil moisture evaporation (Shirey et all 2006) for dehumidifiers that integrate ventilation; and (4) full accounting for the combined effects of unbalanced duct leakage, unbalanced ventilation, and infiltration (Walker and Wilson 1998, ASHRAE 2009). These model improvements were implemented to provide more accurate and meaningful simulation results so that the performance of the different systems could be properly compared.
This study used TRNSYS (Klein et al. 2000) as the basis for the building and system simulations. Specifically a model originally developed to simulate desiccant-based dehumidification systems (Henderson and Sand 2003) and then residential AC and dehumidification systems (Henderson et al. 2007) was improved to meet the needs of this project. The TRNSYS-based building energy simulation tool was used to simulate the HVAC technologies being investigated as part of this project. The software tool already included 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 gas-fired or condenser-regenerated desiccant dehumidification equipment. It also includes robust models for conventional AC 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 of the TRNSYS 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.
The cities listed in Table 1 were chosen to represent locations that have significant latent loads for at least part of the year. Climates Z1 through Z5 represent each of the IECC Climate Zones 1A through 5A. Orlando was not originally part of the study but was added because this humid location was found to have unique humidity issues compared to the other modeled cities in Zones 1A or 2A. TMY3 typical year hourly weather data was used to represent each of these cities in the simulations.
The 2000 [ft.sup.2] (200 [m.sup.2]) 3 bedroom house was modeled as slab-on-grade building with a separate attic zone (a 2-zone model in TRNSYS Type 56). A range of building enclosures were simulated corresponding to Home Energy Rating System (HERS) levels (RESNET 2013) that were selected to correspond to common industry benchmarks (see Table 2). The combination of building and mechanical system characteristics were chosen to reflect typical practice at each HERS level. Therefore, higher-efficiency AC systems were specified at lower HERS indices.
The AIM-2 infiltration model (Walker and Wilson 1998, ASHRAE 2009) relates infiltration to wind and indoor/ outdoor temperature difference for each time step. All simulations in this study used coefficients representing shelter from buildings across the street. The effective leakage area (ELA) was varied to provide the desired ACH50 at each HERS level, as shown in Table 3.
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..sup.2] (0.366 [m.sup.2]) for all the HERS levels, or about five times the leakage rate of the building enclosure of the HERS 100 house (Fugler 1999).
Duct Leakage and Thermal Losses. Table 3 also lists the characteristics of the duct system. For the HERS 130 to HERS 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 apportioned 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 (RSI-1.1) for the HERS 130 and 100 houses and R-8 (RSI-1.4) for the HERS 85 and 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). Internal sensible gains from all sources are assumed to be:
* 72.70 MBtu/day (21.3 kWh/day) for HERS 130 and 100
* 65.43 MBtu/day (19.2 kWh/day) for HERS 85
* 58.16 MBtu/day (17 kWh/day) for HERS 70
* 50.89 MBtu/day (14.9 kWh/day) for HERS 50
Internal moisture generation from all sources was specified as 12 lb/day (5.5 kg/day) or less than half of the ASHRAE Standard 160 moisture generation rate of 31.2 lb/day (14.2 kg/ day) for a three-bedroom, four-person 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 (5.5 kg/day) was selected based on a calibration effort where we compared the model to measured data (Rudd et al. 2013). Based 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.
The moisture capacitance term is usually set to a multiple of moisture holding capacity 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.
As a result of the calibration efforts described in Rudd et al. (2013), a 30x multiplier (30 times the house air moisture storage capacity) 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 4000 [ft.sup.2] (371.6 [m.sup.2]) of exposed wall surface area. The thermal mass of the air node was itself also increased by 20x to 12,331.2 kJ/K (21,418 Btu/[degrees]F) to reflect the impact of furniture and other material in the space. The attic was 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 (Arasteh er al. 2009). The WINDOW inputs for this project were determined following the methodology developed by Arasteh et al. (2009) for use in EnergyPlus.
The suitability of this model to TRNSYS was checked by comparing the fraction of transmitted solar radiation as calculated by TRNSYS to the nominal solar heat gain coefficient (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 was 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 all cases except the HERS 50 case, 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 x [m.sup.2] x K (1.6Btu/h x [ft.sup.2] x [degrees]F) was selected to be most consistent with the expected attic temperatures (Gu 2011).
AC System Details
The study considered both conventional cooling systems as well as other systems that provided improved humidity control. The size of the conventional AC unit was determined for each climate and HERS level using ACCA Manual J (2003). The resulting unit sizes ranged from 3.5 to 2 tons (12.3 to 7 kW); the smallest available unit was assumed to be 2 tons (7 kW). This sizing was used for all systems.
A range of air-conditioner efficiencies 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 three values are required to determine the resulting SEER for each AC system. These values are given in Table 4. The nominal SHR was always set to 0.77 (and the performance map used an apparatus dew-point/bypass factor calculation approach to predict SHR at other conditions). Two-speed and variable-speed systems require additional parameters at low speed to define their performance. Table 4 lists these additional low speed parameters. Note that the low stage fan power (Watt [cfm]) was also used for operation at reduced airflow on single-speed systems (i.e., System 2).
The nominal airflow in the cooling mode is assumed to be 375 cfm per ton (181 liter/h-W). In the heating mode, the fan airflow is slightly lower at 275 cfm per nominal ton (133 liter/h x W) of cooling. For the single-speed units, 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 heating airflow rate or the airflow corresponding to the lowest stage (from the table above) is assumed to be used for ventilation or mixing as described in more detail below.
Setpoints. A 70[degrees]F (21.1[degrees]C) heating setpoint was selected as appropriate for temperate climates while 72[degrees]F (22.2[degrees]C) was deemed as more appropriate for the warm, humid climates. The cooling setpoint of 78[degrees]F (25.6[degrees]C) was selected as most consistent with homeowner preferences in warm climates and is also consistent with the HERS Reference House according to the 2013 Mortgage Industry National Home Energy Rating Systems Standards (RESNET 2013). When humidistat control was required, setpoints of 50% and 60% rh were used.
The impact of thermostat deadband and anticipator were explicitly considered in this short timestep model for the cooling mode using the method developed by Henderson (1992). Droop is an important performance characteristic of both conventional and digital thermostats (Honeywell 2002). The deadband was [+ or -] 1[degrees]F ([+ or -] 0.6[degrees]C) with other parameters and time constants selected to provide a resulting temperature droop with runtime fraction that was about 2[degrees]F ([+ or -] 1.2[degrees]C). For heating, a simple deadband with anticipator dynamics was used. The humidity control deadband was also a simple deadband of [+ or -] 1.5% rh around the target value.
Mechanical Ventilation Options
Several ventilation options were considered in this study, including:
* exhaust fan only (EXH)
* supply: central fan integrated (CFIS)
* heat recovery ventilator (HRV) with AHU distribution
* enthalpy recovery ventilator (ERV), with AHU distribution
* independent enthalpy recovery ventilator (Independent ERV), without AHU operation
Figure 1 shows the airflow configuration used for each ventilation system. All these mechanical ventilation options provide the average rate of 58 cfm (98.5 [m.sup.3]/h) required by ASHRAE Standard 62.2 for the 2000 [ft.sup.2] (200 [m.sup.2]) three bedroom house. The HERS 130 was also modeled with no mechanical ventilation (infiltration only).
The combined impact of infiltration, ventilation, and duct leakage was considered by using the following equations. Since the supply duct leak was dominant, the duct leakage was always a net out in this case, i.e. exhaust.
[Q.sub.in] = sum of all incoming ventilation flows (ventilation airflow at AC, ventilating dehumifier unit, etc.)
[Q.sub.out] = sum of all exhaust flows (exhaust fan, net duct leakage, etc.)
[Q.sub.balanced] = MIN([Q.sub.in], [Q.sub.out])
[Q.sub.unbalanced] = MAX([Q.sub.in], [Q.sub.out]) - [Q.sub.balanced]
[Q.sub.inf] = infiltration flow calculated for building for the timestep
[Q.sub.combined] = MAX([Q.sub.unbalanced], [Q.sub.inf] + 0.5 x [Q.sub.unbalanced]) + [Q.sub.balanced]
Some fresh air is provided as part of the mechanical system. Therefore, the net mechanical inlet flows are subtracted from [Q.sub.combined] to determine the remaining nonmechanical ventilation (or infiltration) rate acting on the building envelope. A mass balance tracks CO2 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 AC unit) to provide the necessary ventilation for the house. The exhaust fan power is assumed to be 0.4 Watt/cfm (0.85 J/liter).
Supply: Central Fan Integrated. A fresh air damper is installed on the return side of the AHU. When there is a call for cooling or heating, fresh air is provided. Including calls for cooling or heating, the supply air fan runs a minimum of 34% of each hour (at heating airflow, fan power is 0.4 Watt/cfm [0.85 J/liter]) 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 (296 [m.sup.3]/h) of fresh air regardless of the airflow rate (heating or cooling airflow).
Balanced Ventilation: ERV/HRV. An energy recovery ventilator (enthalpy wheel) or heat recovery ventilator (sensible heat exchanger) is installed to provide balanced ventilation. The assumed performance is 70% sensible, 60% moisture effectiveness for the ERV, and 75% sensible effectiveness for the HRV.
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 (1.06 J/liter) for the ventilation supply 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 1). The air handler unit (AHU) supply fan is interlocked with HRV/ ERV fan to ensure adequate distribution of the ventilation air. When there is not a call for cooling or heating, the AHU fan operates at the heating airflow rate for single-speed cooling units and at the lowest available airflow rate for two-speed cooling units.
These mechanical ventilation options were implemented for each of the AC system options described below. One modeling change implemented for this study was to allow for a dehumidifier to operate simultaneously with the HRV/ERV option, so this combination could be considered. The ventilation options are not required in combination with System 7 (ventilating dehumidifier). This system provides ventilation in conjunction with dehumidification, so it did not require other ventilation options.
The other slight variations or exceptions are for the two-speed air conditioner (System 3) which has a very small low-stage airflow, therefore the ventilation rate for the central fan integrated supply option is assumed to provide 116 cfm (197 [m.sup.3]/h) of ventilation 50% of the time (instead of 174 cfm [296 [m.sup.3]/h] 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 AC is still assumed to run at low speed in conjunction with the HRV/ERV to provide some degree of mixing.
As a sensitivity, we also considered an ERV that was fully decoupled from AHU operation in selected situations. The ERV pulls air directly from the space and supplies tempered ventilation air directly to the space (therefore, typical interlocked AHU fan operation is not required).
Space Conditioning and Dehumidification Systems
Fourteen cooling and dehumidification systems were identified for evaluation in the Task 1 report (BSC 2009). These systems are described in detail below. Some systems only operate based on a thermostat signal to control space temperature (Systems 1, 3, 4, and 12). The other systems also have a humidistat to control dehumidifiers or change system operation in response to the space-relative humidity level. The air-conditioner supply or air-handling unit (AHU) fan in some cases runs in conjunction with the dehumidification unit to ensure adequate space mixing (Systems 5, 6, 7, 13, and 14).
System 1: Conventional DX System. A conventional, single-speed DX air conditioner (air-cooled condenser) with a PSC fan motor (SEER 13). The system has a gross EER of 13.8 Btu/Wh (cooling COP = 4) at rated conditions. In cooling, the system operates at a supply airflow rate of 375 cfm/ton (181 liter/h-W) with corresponding fan power of 0.5 Watt/cfm (1.06 J/liter). The heating airflow is 275 cfm/ton (133 liter/h x W).
System 2: Conventional DX System with Lower Airflow and Thermostat Overcooling. This system is the same as System 1 above but with controls to provide lower airflow (reduced to 200 cfm/ton [97 liter/h-W]) and space overcooling (as much as 2[degrees]F [1.2[degrees]C] below setpoint) when space humidity is high. When space humidity increases above the RH setpoint, the cooling setpoint is reduced up to 2[degrees]F (1.2[degrees]C) to continue cooling operation. While in the overcooling mode, compressor, and fan operation is limited to 50% runtime (10 minutes ON, 10 minutes OFF) to avoid evaporator coil icing, the 50% runtime limit was only found to be necessary in the most humid climates with humidity setpoints of 50% rh. The operating fan power for PSC motor fan drops from 0.5 to 0.3 W/cfm (1.06 to 0.64 J/liter) at low airflow. The operating fan power for brushless permanent magnet (BPM) motor drops from 0.35 to 0.1 W/cfm (0.74 to 0.21 J/liter) at low airflow.
System 3: Two-Speed Conventional DX System. This system is a two-speed AC unit with a gross EER of 14.85 Btu/ Wh at high speed and 18.56 Btu/Wh at low speed (cooling COP of 4.4 and 5.4). 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 (0.74 J/liter at 181 liter/h-W) for high speed. Compressor capacity and airflow drop at the same proportion to maintain the same cfm per delivered ton.
Since the different HERS levels require different air-conditioner efficiencies, the difference between HERS level and system are not always clear. Table 5 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 that listed under System 2. 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 AC system for HERS 50 and 70 (except for System 12, which is discussed below).
System 4: Variable-Speed Mini-Split DX System. This system uses 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 near 19. Table 4 shows the detailed characteristics that define this unit. The operating fan power is 0.1 W/cfm at 375 cfm/ton (0.21 J/liter at 181 liter/h-W) 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 in the final report provides more detail on how this system was modeled (Rudd et al. 2013).
System 5: Stand-Alone Dehumidifier with Conventional System Mixing. This system includes a portable, standalone dehumidifier (DH) supplements the air conditioner. It is a small 50 pint/day (23.1 liter/day) dehumidifier with an energy factor (EF) of 1.55 liter/kWh at AHAM rating conditions. The dehumidifier operates to maintain the dehumidification setpoint independently of the conventional air conditioner. Fan controls ensure that the AHU supply fan runs a minimum of 7 to 11 minutes of at each hour (depending on the supply airflow) to provide mixing in the space corresponding to a recirculation turnover rate of 0.5 air changes per hour. Since this DH unit is ductless, it does not have separate fan power. Appendix E in the final report provides the detailed performance curves for this system (Rudd et al. 2013).
System 6: Ducted Dehumidifier with Conventional System Mixing. This system uses a slightly larger 64 pint/ day (29.6 liter/day) ducted dehumidifier with an EF of 1.98 liter/kWh at AHAM rating conditions. This ducted unit includes a fan that is integrated with the AC 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 DH unit when the DH fan is off. The dehumidifier operates to maintain the required humidity setpoint. The equipment performance map for this unit was based on a unit that was tested in the NREL HVAC lab; see Appendix E in the final report (Rudd et al. 2013).
The airflow through the 64 pint/day (29.6 liter/day) dehumidifier was assumed to provide 135 cfm (229 [m.sup.3]/h) with a fan power of 0.7 W/cfm (1.5 J/liter). 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 AC supply air fan runs a minimum of 7 to 11 minutes of each hour (depending on the supply airflow) to provide mixing in the space corresponding to a recirculation turnover rate of 0.5 air changes per hour.
System 7: Ducted Dehumidifier with Outdoor Air Preconditioning. Larger 82 pint/day (37.9 liter/day) dehumidifier with an EF of 1.98 liter/kWh at AHAM rating conditions. This ducted unit includes a fan that is integrated with the AC unit in a ventilation air preconditioning configuration, as shown in Figure 2. 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 (296 [m.sup.3]/ h) through the dehumidifier includes 58 cfm (98.5 [m.sup.3]/h) 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 AC supply duct. The dehumidifier compressor operates in response to the humidity setpoint. The equipment performance map for this unit was based on a unit that was tested in the NREL HVAC lab, see Appendix E in the final report (Rudd et al. 2013).
Fan controls ensure that the AC supply air fan runs a minimum of 7 to 11 minutes of each hour (depending on the supply airflow) to provide mixing in the space corresponding to a recirculation turnover rate of 0.5 air changes per hour.
System 8: Enhanced Cooling with Partial-Condensing/ Subcooling Reheat. This system uses a central DX cooling system with refrigerant subcooling/reheat coil. It is based on either a single-speed or a two-speed AC unit depending on the HERS level, but with enhanced mode operation when humidity is high. Specifics regarding this system's dehumidification operating mode are provided in the Task 1 report (BSC 2009) and in Appendix F of the final report (Rudd et al. 2013).
System 9: Enhanced Cooling with Full Condensing/ Subcooling Reheat. This system uses a central DX cooling system with modulating hot gas reheat providing full condensing at an indoor reheat coil. It is based on either a single-speed or a two-speed AC unit depending on the HERS level. The fan power is increased by 0.05 Watt/cfm (0.11 J/liter) 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 (BSC 2009) and in Appendix G of the final report (Rudd et al. 2013).
System 10: Conventional DX System with Lower Airflow. This system is similar to System 2 but only includes controls to provide lower airflow (reduced to 200 cfm/ton [97 liter/h-W]), orto 53.3%) when space humidity rises above the setpoint. This system was only considered in a subset of the climates and HERS levels.
System 11: Conventional DX System with Thermostat Overcooling. This system is similar to System 2, but only includes controls to allow for space overcooling (by as much as 2[degrees]F [1.2[degrees]C] below the cooling setpoint) when space humidity increases above the RH setpoint. The reset schedule linearly varies the about overcooling in proportion to the increase in RH above the dehumidification setpoint. This system was only considered in a subset of the climates and HERS levels.
System 12: Conventional DX System with Sensible-Only AAHX. This system is a conventional air conditioner with sensible air-to-air heat exchanger (AAHX) added for improved dehumidification performance (e.g., heat pipes). The heat pipe is assumed to be a noncondensing, aluminum fin, 1/2 inch O.D. copper tube, 2 row, 11 fpi. The face area of the heat pipe is set equal to the nominal AC tonnage. The operating fan power increases to 0.55 Watt/cfm (1.17 J/liter) and the airflow drops to 350 cfm/ton (169 liter/h-W) to account for the additional heat exchanger pressure drop compared to System 1. This system uses the 14.5 SEER single-speed AC in the HERS 50 and HERS 70 houses (since two-speed operation would result in condensation on the heat pipe at low speed).
System 13: Conventional DX System with Gas-Regenerated Desiccant Dehumidifier. This system is a conventional AC unit with gas-regenerated desiccant dehumidifier. The unit pulls in regeneration air from outdoors and then exhausts it back to outdoors. The 400 cfm (680 [m.sup.3]/h) unit has a dehumidification capacity of 6.3 lb/h (145 pint/day or 67 liter/day) at AHAM rating conditions. Gas consumption is 10,000 Btu/h (2,930 W). 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 AC supply duct, as shown in Figure 2. The process side (or supply) fan power is 0.6 Watt/cfm (1.27 J/liter). Fan controls ensure that the conventional AC supply air fan runs at least 7 to 11 minutes of each hour (depending on the supply airflow) to provide mixing in the space corresponding to a recirculation turnover rate of 0.5 air changes per hour. Appendix H in the final report (Rudd et al. 2013) provides the detailed performance curves to model this unit.
System 14: DX Condenser-Regenerated Desiccant
Dehumidifier. This system is a conventional mechanical cooling system with a desiccant system regenerated by condenser waste heat from its small internal compressor. The unit's energy factor is 2.6 liter/kWh (excluding supply fan power). The supply fan is 300 cfm (510 [m.sup.3]/h) at 210 Watts (0.7 W/cfm or 1.5 J/liter). Controls ensure that the conventional AC supply air fan runs at least 7 to 11 minutes out of every hour (depending on the supply airflow) to provide mixing in the space (minimum zone air mixing rate of 0.5 ACH). A performance map for the this desiccant unit has been developed based on data in Appendix H of the final report (Rudd et al. 2013).
Summary of Dehumidifier Configurations. Figure 2 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 recirculation turnover rate of 0.5 air changes per hour.
Electric and Gas Costs
Total HVAC costs (cooling, heating, fans) were determined using the electric and gas rates from Table 6. The heating load was tracked for each hour and then heating costs for the gas furnace case were determined by post-processing the results assuming a furnace efficiency and using the gas costs from the table below. Furnace efficiencies were assumed to be 80% for HERS 130, 85% for HERS 100, 87% for HERS 85 and HERS 70, and 93% for HERS 50.
Several hundred annual simulation runs were completed for the five different HERS levels, six climates, fourteen different systems and five different mechanical ventilation options. Systems with active dehumidification controls were also run at two different humidity setpoints (50% and 60% rh). The detailed results are all available at a project website http:/ /cloud.cdhenergy.com/rp1449/. The local utility rates given in Table 6 were used to calculate operating costs for each simulation run.
HERS Level and Climate
Simulations were run at the 5 different HERS levels and in six climates. Abbreviated summary results are listed in Table 7 for System 1, the baseline or conventional AC. The HERS 130 level was run with no ventilation (V0) to represent the current housing stock and with the ASHRAE Standard 62.2-2010 ventilation rate via an exhaust fan (V1). For all other HERS levels shown in Table 7, only the results with the exhaust ventilation system (V1) are provided here. Comparison of all ventilations systems is provided later.
Figure 3 compares the total costs for the different HERS and climates using the data from Table 7. The relative costs area pproximately (but not precisely) in line with the ratios implied by the HERS level.
The air-conditioner operating hours are also significantly greater in Miami, as shown in Figure 3 and Table 7. The annual runtime for the two-speed AC units in the HERS 50 and 70 houses is also much greater than the single-speed units in the other HERS level houses since the unit runs at low speed for longer periods of time.
Evaluating Humidity Levels
Humidity in the space is not directly controlled by the AC unit, so humidity levels typically vary across the year. The psychrometric chart in Figure 4 shows the average daily conditions observed across the year for the HERS 100 house in Miami with System 1 and exhaust ventilation. Each point on the plot represents the average conditions for the day. The days with any cooling operation are shown as blue. The annual total number of hours over 60% rh was 1303. Most of these hours occur in Miami when the AC unit is off for the day. Figure 5 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, or light gray, indicates the hours below 55% rh; 5, or black, indicates the hours above 75% rh). This plot confirms that high humidity does not occur much in the summer but tends to happen in swing seasons and in the winter when little or no cooling operation is required. Later night and morning hours are also predominant.
The most common metric for gauging the prevalence of high humidity is the annual number of hours above some threshold. Figure 6 compares the number of hours above 60% rh across the climates and HERS levels with the conventional cooling system and exhaust ventilation. While there is not a hard requirement or limit on humidity, the authors believe that the hours above 60% is a reasonable gauge when comparing different systems. Other thresholds were also considered in the final report (Rudd et al. 2013). Other metrics, such as the number of high humidity events of a certain duration (e.g. 4 hours and 8 hours) were also shown to indicate the same trends.
The results in Figure 6 show the expected trends consistent with previous studies for conventional cooling systems (Henderson et al. 2007; Henderson et al. 2008). The HERS 130 house without mechanical ventilation typically has lower humidity levels than the same house with mechanical ventilation. When ventilation is provided to the house, the hours of high humidity tend to increase slightly. 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 impact of duct location is considered in greater detail below.
Impact of Duct Location
One of the main differences between the HERS 70 and HERS 50 houses is the elimination of ducts in the attic. Locating the space conditioning air distribution 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. Figure 7 also shows that moving the ducts inside the conditioned space clearly results in more hours with high humidity in all cases. Removing ducts from the attic decreases total space conditioning costs by 10% to 28% depending on climate and HERS level. The reduction was 22% to 28% in the three hot-humid climates. 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. In the hot-humid climates, when ducts were moved inside the conditioned space and heat gain to ducts was eliminated, the hours above 60% rh increased 27%-37% for HERS 70 and 33%-54% for HERS 50.
Comparing 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 the HERS 130 house only.
* Exhaust Fan (V1). 58 cfm (98.5 [m.sup.3]/h) is continuously exhausted from the house. Fan power is 0.4 W/cfm (0.85 J/liter).
* CFIS (V2). 174 cfm (296 [m.sup.3]/h) 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 (98.5 [m.sup.3]/h) of outside air on average. A ventilation damper shuts if the fan runtime exceeds 34% of the time in the hour. For two-speed AC units, the ventilation rate is changed to be 116 cfm (197 [m.sup.3]/h) for 50% of each hour since the minimum 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 (197 [m.sup.3]/h) of outside air for 50% of each hour. The AHU fan is interlocked with HRV fan to provide distribution to the space. HRVs are used in Atlanta, Nashville, and Indianapolis.
* ERV (V4). Same operation and control as HRV. ERVs are used in Orlando, Miami, Houston, and Atlanta.
All the ventilation options provided the equivalent of 58 cfm (98.5 [m.sup.3]/h) of continuous fresh air 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 8 (a) compares the total costs, including natural gas costs for space heating, for the different ventilation options. The lowest cost option is typically the exhaust fan since this does not require additional AHU fan operation. 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 8(b) 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, 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 disbenefitofkeepingmoistureinsidethehouseattimeswhen 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 off-cycle evaporation from the cooling coil (Shirey et al. 2006).
Enhanced Control Approaches
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 AC unit. Another potential strategy is to continue cooling the space below the setpoint 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% when the space humidity is above the RH setpoint.
2. Reset the space cooling setpoint down by as much as 2[degrees]F (1.2[degrees]C) as the space humidity increases by 10% rh above the setpoint. Limit AC operation to no more than 50% of each hour while overcooling is occurring.
These two strategies together are called an enhanced AC unit (System 2) for this study. An activation setpoint of 50% rh was used. Figure 9 includes psychrometric charts for HERS 100 in Miami of conventional and enhanced systems that show how the enhanced control approach tends to eliminate or "sweep down" many of the high humidity hours that occur just below the cooling setpoint by continuing cooling operation at lower temperatures. The performance results for the conventional and enhanced AC unit are compared in Table 8 for both the single-speed and two-speed units in Houston for the HERS 100 and HERS 70 houses. Enhanced control provides benefits for both these system types.
Best Humidity Control Technologies
Various dehumidifier and enhanced systems were considered in this analysis. All these results are available at the website referenced above. The best performance option from each class of system were identified in the analysis on the final report (Rudd et al. 2013). The best enhanced and dehumidifier technologies are compared in Table 9 for the HERS 100 house in the humid climates. Total operating costs for the systems that provide full control of the humidity setpoint are typically a 10%-20% cost premium compared to cooling alone. System 9, which has full condensing reheat control, has the lowest operating cost premium.
Analysis of all the simulation results from the final report (Rudd el al. 2013) led us to identify several themes and trends that are discussed below.
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 setpoint.
A More Efficient Building Envelope 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 website mentioned above.
Moving Ducts into the Conditioned Space Saves Energy but Increases Relative Humidity Levels. When duct losses to the attic are eliminated, space-conditioning energy decreases (22% to 28% reduction in hot-humid climates) but the number of hours over 60% rh increases. A sensitivity in the HERS 50 and HERS 70 houses showed this as well as the comparison of ductless variable-speed systems and the two-speed systems with ducts in the attic. At the HERS 50 level--when ducts come inside the conditioned space and heat gain to ducts is eliminated--the hours above 60% rh increased significantly in the three hot-humid climates.
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 Atlanta. However, CFIS ventilation slightly increased high humidity hours in Miami and Houston because it provided more fresh air and because the part-time off-cycle operation of the AHU fan sometimes resulted in increased evaporation from the cooling coil.
Energy Recovery Ventilators Have Little to No Dehumidification 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. 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 warm-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.
Air Conditioners with Enhanced Controls Significantly Reduce the Number of High Relative Humidity Hours. The AC 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 setpoint was 50% rh. An activation setpoint of 60% rh resulted in almost no benefit. The enhanced controls strategies are to (1) reduce the relative airflow (cfm/ton) at high relative humidity levels and (2) allow for 2[degrees]F (1.2[degrees]C) overcooling to increase compressor runtime at mild conditions. This relatively simple technology provides a cost effective means to mitigate the number of 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. 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 Gains, and Cooling/ Heating Setpoint Significantly Impact Relative Humidity Levels. Several model parameters have a large impact on the prevalence of high relative humidity hours across the year (see Appendix J in the final report [Rudd et al. 2013]). Internal moisture gain and capacitance are difficult to determine by direct measurement or observation while internal sensible gain and the heating setpoint can be determined; all these factors have a significant impact on the number of high humidity hours predicted with the model. This finding is consistent with the occupant-based variability found in field studies by Rudd et al. 2003 and 2005 and Rudd and Henderson 2007. For this study, we chose the moisture gain (12 lb/day [5.5 kg/day]) and moisture capacitance (30x) by comparing the simulation model to measured data (see Appendix C in the final report). The cooling setpoint was 78[degrees]F (25.6[degrees]C). The heating setpoint was 70[degrees]F (21.1[degrees]C) except in the hot-humid climates where it was 72[degrees]F (22.2[degrees]C); raising the heating setpoint even more continued to reduce high RH hours.
The Choice of Relative Humidity Setpoint Affects Energy Use. The energy required for dehumidification strongly depends on the choice of humidity setpoint. For instance, decreasing the dehumidifier (DH) setpoint from 60% to 50% can increase DH energy use by a factor of five.
Several hundred annual simulation runs were completed for these various scenarios. The detailed results are available at the project website http://cloud.cdhenergy.com/rp1449/. Local utility rates were used to calculate operating costs for each run. We generally found that systems that properly control humidity throughout the year have a 10% to 30% higher space conditioning operating cost than uncontrolled conventional systems, depending on the dehumidification system and the relative humidity control setpoint 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 subcool/reheat coil (System 8) has just slightly higher operating costs. The air conditioner with enhanced controls (System 2) eliminated many of the high humidity hours with a very modest cost premium over the conventional system.
The key findings from the analysis of the simulation results are summarized below:
* Periods of high relative humidity mostly occur at mild conditions in the winter and swing seasons when there is little of no need for sensible cooling. Humidity is rarely out of control during the summer periods.
* With mechanical ventilation provided, a more efficient building envelope generally reduces the prevalence of high humidity levels as the HERS level decreases from 130 to 70. Moving from HERS 70 to HERS 50 typically increases humidity levels since the ducts are moved inside the conditioned space at the HERS 50 level (see below).
* Moving ducts from the attic to the conditioned space saves energy but increases space humidity levels in warm, humid climates. The reduction in sensible heat gains is greater than the reduction in latent loads, resulting in a mix of sensible and latent loads that increases space humidity levels.
* Different ventilation systems have different impacts on space humidity levels. The CFIS approach was found to have slightly more high humidity hours in some humid climates compared to the exhaust-only approach because it provided more fresh air and because the periodic fan operation sometimes resulted in increased moisture evaporation from the deactivated cooling coil. Energy recovery ventilators (ERVs) provide little or no dehumidification benefit in humid climates. Since most hours with high humidity levels occur at mild conditions--when indoor and outdoor conditions are similar--there is little humidity reduction provided by the ERV at that time.
* Air conditioners with enhanced controls significantly reduce the number of high humidity hours. The enhancements of lower airflow and overcooling when humidity is high generally cut the number of high humidity hours in half when the activation setpoint was 50% rh. This relatively simple control enhancement to a conventional air conditioner provides a cost effective means to mitigate many of the high humidity hours.
* Variable capacity systems do not improve high humidity control unless the lower airflow and overcooling control enhancements mentioned above are implemented. Two-speed and variable-speed systems showed no significant difference from conventional systems until the airflow was reduced. The enhancements may have slightly more humidity control benefit when implemented on a variable-capacity system since the longer runtimes at low speed provide more time for the control enhancements to be effective.
AC = air conditioner
AHU = air handling unit or indoor section of air conditioner
HRV = heat recovery ventilator
ERV = energy recovery ventilator
Q = airflow rate
in = airflow in
out = airflow out
inf = infiltration
This paper is based on findings resulting from ASHRAE Research Project RP-1449.
This project was supported, in part, by the Air-Conditioning, Heating and Refrigeration Technology Institute (AHRTI). The authors wish to thank the members of the project monitoring subcommittee (PMS) for their valuable guidance and oversight of this research project, including Mark Olsen, John Andrews, Dane Christensen, Xudong Wang, and chairman Michael Lubliner. Don Shirey was an integral part of this project through the Task 4 simulation plan. Also thanks to Jon Douglas, Jim Cummings, and Rob Vieira for their feedback and insights.
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Hugh I. Henderson, Jr, PE
Hugh I. Henderson Jr. is a principal at CDH Energy Corp. in Cazenovia, NY. Armin Rudd is a principal at ABT Systems, LLC, Annville, PA.
Table 1. Climates Selected for Simulation Zone ID IECC Climate Zone City Z0 2A Orlando Z1 1A Miami Z2 2A Houston Z3 3A Atlanta Z4 4A Nashville Z5 5A Indianapolis Table 2. Description of House Performance Levels Description HERS AC Efficiency or Benchmark Index Typical Older Existing 130 10 SEER House HERS Reference House 100 13 SEER Energy Star House 85 14.5 SEER (BPM fan) Builders Challenge House 70 17.7 SEER (2-speed compressor) Building America House 50 17.7 SEER (2-speed compressor) Notes: BPM--Brushless Permanent Magnet fan motor Table 3. Envelope Leakage and Duct Performance HERS Target ELA Duct Level ACH at 4 Pa, Leakage, at [in.sup.2] Supply 50 Pa ([m.sup.2]) and Return Combined 130 10 140.1/0.090 20% 100 7 98.1/0.063 10% 85 5 70.1/0.045 5% 70 4 56.1/0.036 5% 50 3 42/0.027 na HERS Duct Supply: Level Insulation Return Duct R([R.sub.SI]) Area, [ft.sup.2] ([m.sup.2]) 130 R6/1.1 544:100/50.5:9.3 100 R6/1.1 544:100/50.5:9.3 85 R8/1.4 544:100/50.5:9.3 70 R8/1.4 544:100/50.5:9.3 50 na na Table 4. AC Unit Performance Characteristics Description Gross EER, Actual Fan Power, Btu/Wh (W/W) W/cfm (J/liter) SEER 10 (Single Speed, PSC Fan) 10.35/3.0 0.50/1.06 SEER 13 (Single Speed, PSC Fan) 13.80/4.0 0.50/1.06 SEER 14.5 (Single Speed, BPM Fan) 14.85/4.4 0.35/0.74 SEER 17.7 (Two-Speed) 14.85/4.4 0.35/0.74 SEER 19 (Variable Speed Ductless) 14.85/4.4 0.10/0.21 Description Actual Low Stage Low Stage Fan Power, Capacity : W/cfm (J/liter) Fan Speed Ratio SEER 10 (Single Speed, PSC Fan) 0.30/0.64 -- SEER 13 (Single Speed, PSC Fan) 0.30/0.64 -- SEER 14.5 (Single Speed, BPM Fan) 0.10/0.21 -- SEER 17.7 (Two-Speed) 0.10/0.21 0.5 : 0.5 SEER 19 (Variable Speed Ductless) 0.07/0.15 0.33 : 0.30 Notes: Gross EER is total cooling capacity of the coil divided by compressor and condenser fan power at ARI rating point and with 450 cfm/ton (218 liter/h-W). The ARI Rating Point is 95degrees]F outdoor db/80[degrees]F entering db /67[degrees]F entering wb (35[degrees]C/ 26.7[degrees]C/19.4[degrees]C). Fan power corresponding to SEER rating conditions is 0.25 Watt/cfm (0.53 J/liter) for PSC fan motor and 0.18 Watt/cfm (0.38 J/liter) for BPM fan motor. Table 5. Matrix of AC Units Used with Each HERS Level and System HERS Level System 1 System 2 System 3 System 4 Systems ConvAC Enhanced Two-Speed Variable- 5-14 AC AC Speed AC SEER 130 10# 10 17.7 19 10 100 13# 13 17.7 19 13 85 14.5 14.5# 17.7 19 14.5 70 14.5 17.7# 17.7 19 17.7 50 14.5 17.7# 17.7 19 17.7 Note: gray shaded cells indicate the default AC unit for each HERS level. Note: Default AC unit for each HERS are indicated #. Table 6. Electric and Gas Costs Zone ID Electric Cost, Gas Cost, $/therm $/kWh Z0: Orlando 0.101 1.79 Z1: Miami 0.101 1.79 Z2: Houston 0.085 1.08 Z3: Atlanta 0.103 1.52 Z4: Nashville 0.097 1.05 Z5: Indianapolis 0.078 0.86 Notes: Electric costs from Form 826 EIA data for residential customers for local utility. Gas costs from EIA data for residential customers in each state. Table 7. Performance Results for Different HERS Levels and Climates for System 1 HERS Level Climate Hours AC Runtime, Total Above h Costs 60% rh with Gas Furnace 50 Orlando 1011 3266 $265 70 575 3491 $354 85 1121 2039 $614 100 1645 1860 $980 130 1361 2127 $1352 130 no vent 1296 2042 $1260 50 Miami 822 4378 $295 70 382 5303 $426 85 756 2859 $657 100 1303 2700 $917 130 1313 3054 $1309 130 not vent 1059 2936 $1226 50 Houston 380 3252 $269 70 191 3844 $376 85 228 2185 $540 100 628 1966 $849 130 942 2170 $1386 130 no vent 778 2096 $1309 50 Atlanta 40 2054 $341 70 15 2604 $510 85 20 1703 $689 100 291 1338 $1168 130 426 1726 $1673 130 no vent 314 1702 $1573 50 Nashville -- 2184 $320 70 -- 2641 $475 85 -- 1705 $632 100 4 1555 $823 130 119 1658 $1387 130 not vent 25 1622 $1292 50 Indianapolis -- 1457 $364 70 -- 1828 $553 85 -- 1154 $634 100 -- 1057 $862 130 7 1240 $1177 130 no vent -- 1237 $1083 Table 8. Performance Results with Various AC Enhancements In Houston System Hours AC Total Above Runtime, Costs 60% rh h with Gas Furnace System 1: Conventional AC HERS 100, 628 1,966 $849 System 10: Lower Airflow Single- 476 2,006 $848 System 12: Overcooling Speed AC 326 2,037 $868 System 2: Both 231 2,053 $859 Enhancements System 1: Conventional AC HERS 70, 191 3,844 $376 System 10: Lower Airflow Two-Speed 112 3,961 $378 System 12: Overcooling AC 77 4,204 $323 System 2: Both 53 4,016 $381 Enhancements Table 9. Dehumidification Performance Results for HERS 100 and Different Climates HERS Level 100 Climate Hours AC Run- Relative Above time, Total 60% rh h Costs System 1: Conventional AC Orlando 1,645 1,860 100% System 2: Enhanced AC 928 2,038 104% System 8: Partial RH/SC 205 2,320 120% System 9: Full Reheat -- 2,455 115% System 6: Ducted DH -- 1,996 119% System 14: Condenser DES -- 1,818 120% System 1: Conventional AC Miami 1,303 2,700 100% System 2: Enhanced AC 471 2,881 103% System 8: Partial RH/SC 22 3,138 118% System 9: Full Reheat -- 3,178 112% System 6: Ducted DH -- 2,829 117% System 14: Condenser DES -- 2,672 116% System 1: Conventional AC Houston 628 1,966 100% System 2: Enhanced AC 231 2,053 101% System 8: Partial RH/SC 9 2,211 109% System 9: Full Reheat -- 2,250 106% System 6: Ducted DH -- 2,033 110% System 14: Condenser DES -- 1,953 110% System 1: Conventional AC Atlanta 291 1,338 100% System 2: Enhanced AC 183 1,359 100% System 8: Partial RH/SC 54 1,417 102% System 9: Full Reheat 1 1,446 102% System 6: Ducted DH -- 1,352 104% System 14: Condenser DES -- 1.336 104%
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|Author:||Henderson, Hugh I., Jr.; Rudd, Armin|
|Date:||Jan 1, 2014|
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