Applications Engineering Manual

Dehumidification in HVAC Systems

December 2002 SYS-APM004-EN

Dehumidification in HVAC Systems

John Murphy, senior applications Brenda Bradley, information designer Preface

As a leading HVAC manufacturer, we believe that it is our responsibility to serve the building industry by regularly disseminating information gathered through laboratory research, testing programs, and practical experience. Trane publishes a variety of educational materials for this purpose. Applications engineering manuals, such as this document, can serve as comprehensive reference guides for professionals who design building comfort systems.

This manual focuses on dehumidification (the process of removing moisture from air), as performed by HVAC systems in commercial comfort-cooling applications. Using basic psychrometric analyses, it reviews the dehumidification performance of various types of “cold-coil” HVAC systems, including constant-volume, variable-volume, and dedicated outdoor-air systems. In each case, full-load and part-load dehumidification performance is compared with the 60 percent-relative- limit that is currently recommended by ANSI/ASHRAE/IESNA Standard 62–2001. This manual also identifies ways to improve dehumidification performance, particularly at part- load conditions.

We encourage you to familiarize yourself with the contents of this manual and to review the appropriate sections when designing a comfort-system application with specific dehumidification requirements.

Note: This manual does not address residential applications, nor does it discuss the particular dehumidification requirements for process applications, such as supermarkets, manufacturing, or industrial drying. ■

Trane, in proposing these system design and application concepts, assumes no responsibility for the performance or desirability of any resulting system design. Design of the HVAC system is the prerogative and responsibility of the engineering professional.

“Trane” and the Trane logo are registered trademarks of Trane, which is a business of American Standard Companies.

© 2002 American Standard Inc. All rights reserved SYS-APM004-EN Contents

Introduction ...... 1

Sources and Effects of Indoor Moisture ...... 2 Why be Concerned about Indoor Humidity? ...... 3 ...... 3 Occupant Comfort and Productivity ...... 4 Building Maintenance ...... 5 Climate Considerations ...... 5 Energy Use ...... 7

Dehumidification Primer ...... 9 Types of Dehumidification ...... 9 Local Dehumidification ...... 9 Remote Dehumidification ...... 10 Processes for Dehumidification ...... 10 Condensation on a Cold Coil ...... 10 Adsorption Using a Desiccant ...... 13 Implications for HVAC Control ...... 17 Humidity Control during Unoccupied Periods ...... 17 Building Pressurization ...... 18 Airside Economizing ...... 18

Dehumidifying with Constant-Volume Mixed Air ...... 19 Analysis of Dehumidification Performance ...... 19 Application Considerations ...... 22 Ventilation ...... 22 Climate ...... 24 Packaged DX Equipment ...... 24 Total-Energy Recovery ...... 27 Cold Supply Air ...... 29 Humidity Control during Unoccupied Periods ...... 30 Building Pressurization ...... 30 Airside Economizing ...... 31 Improving Coincidental Dehumidification ...... 32 Adjustable Speed ...... 32 Mixed-Air Bypass ...... 34 Return-Air Bypass ...... 37 DX Coil Circuiting ...... 41 “Direct” Control of Humidity ...... 44 Separate Air Paths ...... 44 Supply-Air Tempering ...... 50

SYS-APM004-EN iii Contents

Dehumidifying with Variable-Volume Mixed Air ...... 61 Analysis of Dehumidification Performance ...... 61 Application Considerations ...... 63 Minimum Airflow Settings ...... 63 Supply-Air-Temperature Reset ...... 64 Supply-Air Tempering at VAV Terminals ...... 65 Humidity Control during Unoccupied Periods ...... 68 Building Pressurization ...... 69 Airside Economizing ...... 69 Improving Dehumidification Performance ...... 70 Condition Outdoor Air Separately ...... 70 Deliver Colder Supply Air ...... 73

Dehumidifying with Dedicated Outdoor Air ...... 75 System Configurations ...... 75 Design Objectives for Conditioned Outdoor Air ...... 77 Moisture Content ...... 77 Dry-Bulb Temperature ...... 80 Application Considerations ...... 86 Humidity Control during Unoccupied Periods ...... 86 Building Pressurization ...... 86 Cooling ...... 87 Reset Control Strategies ...... 90 Reheating Conditioned Air with Recovered Heat ...... 94 Preconditioning Outdoor Air with Recovered Energy ...... 98

Afterword ...... 100

Appendix A: Psychrometric Analysis ...... 101 Full-Load, Peak Dry-Bulb Condition ...... 102 Part-Load, Peak Dew-Point Condition ...... 107

Appendix B: Designing a Dedicated OA System ...... 111 Selecting the Dedicated Outdoor- ...... 112 Selecting the Local HVAC Terminals ...... 116

Glossary ...... 125

References ...... 129

Index ...... 131

iv SYS-APM004-EN Introduction

Uncontrolled moisture can reduce the quality of indoor air, make occupants uncomfortable, and damage a building’s structure and furnishings. One form of moisture is water vapor entrained in the air.

Before the widespread use of , humid weather meant high moisture levels indoors; indoor relative humidity remained acceptable, however, because the dry-bulb temperature indoors also increased. During warm weather, interior surfaces were only slightly cooler than the ambient temperature, so indoor condensation seldom occurred. The presence of any microbial growth primarily resulted from water leaks or spills, or from condensation on poorly insulated walls during cold weather.

Until 1970, designers typically chose constant-volume reheat or dual- systems to provide mechanical ventilation and air conditioning in commercial and institutional buildings. Both types of systems effectively (albeit coincidentally) controlled indoor humidity while regulating dry-bulb temperature. As the 1970s drew to a close, heightened concern about the availability and cost of energy prompted designers to choose system designs that neither used “wasteful” reheat energy nor mixed hot and cold air streams.

Although many of today’s HVAC systems adequately control the indoor dry-bulb temperature, the lack of reheat or mixing allows humidity in the space to “float.” High humidity levels can develop, especially during part-load operation. When coupled with the cold indoor surfaces that result from mechanical cooling, high humidity may lead to unwanted condensation on building surfaces.

The HVAC system and application influence the severity and duration of high indoor humidity. This manual therefore compares the dehumidification performance of several common types of HVAC systems. ■

SYS-APM004-EN 1 Sources and Effects of Indoor Moisture

Moisture can enter a building as a liquid or a vapor via several paths (Figure 1). Refer to Managing Building Moisture, It can cause problems in either form, and after it is inside the building, it can Trane applications engineering manual change readily from liquid to vapor (evaporation) or from vapor to liquid SYS-AM-15, for more information on (condensation). To assure that the conditioned environment inside the building sources of moisture in buildings, remains within the acceptable range, carefully evaluate all sources of moisture methods for calculating moisture- at all operating conditions when designing the HVAC system. related HVAC loads, and techniques for managing moisture in the building Liquid sources include ground-water seepage, leaks in the , envelope, occupied space, and spills, condensation on cold surfaces, and wet-cleaning processes (such as mechanical equipment room. ■ carpet shampooing). Roof leaks are a common source of unwanted water, especially in large low-rise buildings like schools. Leaking pipes, another common source, can be particularly troublesome because the leaks often develop in inaccessible areas of the building.

Water vapor develops inside the building or it can enter the building from outdoors. Indoor sources include respiration from people, evaporation from open water surfaces (such as pools, fountains, and aquariums), combustion, cooking, and evaporation from wet-cleaning. Outdoor sources include vapor pressure diffusion through the building envelope, outdoor air brought in by the HVAC system for ventilation, and air through cracks and other openings in the building envelope, including open doors and windows.

Figure 1. Sources of moisture in buildings

2 SYS-APM004-EN Sources and Effects of Indoor Moisture

Proper practices of design, construction, and operation can help minimize unwanted moisture inside the building. For example, proper landscaping can provide good drainage, periodic roof maintenance can help eliminate roof leaks, the building envelope can include a weather barrier to keep rain from penetrating the wall structure, and (depending on the season and climate) positive building pressurization can minimize the infiltration of humid outdoor air.

Why be Concerned about Indoor Humidity?

Indoor Air Quality

Scientists agree that excess water or “dampness” can contribute significantly to mold growth inside buildings. An article in the November 2002 issue of the ASHRAE Journal notes that:

While it has been difficult for epidemioligic studies to definitively link and human illness, there are indications that indoor mold is responsible for such health concerns as nasal irritation, allergic and non-allergic rhinitis, malaise, and hypersensitivity pneumonitis.1

It is virtually impossible to avoid contact with the spores produced by fungi The Web site hosted by the U.S. (including molds). Fungi exist everywhere: in the air, in and on plants and Environmental Protection Agency animals, on soil, and inside buildings. They extract the nutrients that they need (EPA) is a good source for information to survive from almost any carbon-based material, including dust. Excessive about indoor air quality and related indoor humidity, especially at surfaces, encourages fungi and other health effects (www.epa.gov/iaq). ■ microorganisms, such as bacteria and dust mites, to colonize and grow.

Minimizing sources of moisture is the best way to help minimize microbial growth. Scientist/authors Sarah Armstrong and Jane Liaw recommend that:

In the absence of clear guidance regarding what types of indoor fungi, or concentrations thereof in air, are safe or risky, one may wish simply to prevent mold from growing in buildings by acting quickly [drying water-damaged areas within 24 to 48 hours] when water leaks, spills, or floods occur indoors, being alert to condensation, and filtering air.

1 S. Armstrong and J. Liaw. “The Fundamentals of Fungi,” ASHRAE Journal 44 no. 11: 18–23.

SYS-APM004-EN 3 Sources and Effects of Indoor Moisture

ANSI/ASHRAE Standard 62–2001, Ventilation for Acceptable Indoor Air Quality, If approved, a proposed addendum to addresses the link between indoor moisture and microbial growth in this Standard 62 would require that systems recommendation: be designed to limit the relative humidity in occupied spaces to Relative humidity in habitable spaces preferably should be 65 percent or less at the design outdoor maintained between 30 percent and 60 percent to minimize dew-point condition. The design dew- the growth of allergenic and pathogenic organisms. point condition, however, does not (Section 5.10) necessarily coincide with the worst-case condition for indoor relative humidity. The U.S. Environmental Protection Agency (EPA) adopts a similar stance in its As the examples presented later in this publication titled Mold Remediation in Schools and Commercial Buildings: manual demonstrate, even higher indoor relative can occur on The key to mold control is moisture control. Solve moisture mild, rainy days during the cooling problems before they become mold problems! … [One way to season. The proposal was still under help prevent mold is to] maintain low indoor humidity, below debate when this manual went to press. 60 percent relative humidity (ideally 30–50 percent, if possible). Check ASHRAE’s Web site, www.ashrae.org, for more This publication, which was published in March 2001 and is identified as information. ■ EPA 402-K-01-001, is available from www.epa.gov/iaq/molds. For more information about the mechanics of mold growth and how it affects buildings and HVAC systems, review Chapter 7 in Humidity Control Design Guide for Commercial and Institutional Buildings (ISBN 1-883413-98-2). It was published by ASHRAE in 2001, and is available from their online bookstore at www.ashrae.org.

Occupant Comfort and Productivity

In addition to curbing microbial growth, limiting indoor humidity to an acceptable level helps assure consistent within occupied spaces, which:

■ Reduces occupant complaints ■ Improves worker productivity ■ Increases rental potential and market value

Figure 2. Summer “comfort zone” defined ANSI/ASHRAE Standard 55–1992, Thermal Environmental Conditions for by ASHRAE Standard 55–1992 Human Occupancy, specifies thermal environmental conditions that are acceptable to 80 percent or more of the occupants within a space. The “comfort zone” (Figure 2) defined by Standard 55 represents a range of environmental conditions based on dry-bulb temperature, humidity, thermal radiation, and air movement. Depending on the utility of the space, maintaining the relative comfort humidity between 30 percent and 60 percent keeps most occupants zone comfortable.

Note: A proposed revision to ASHRAE Standard 55 suggests redefining the upper humidity limit for thermal comfort as a humidity ratio of 84 gr/lb (12 g/kg). This approximates a dew point of 62°F (16.7°C) or a relative humidity

4 SYS-APM004-EN Sources and Effects of Indoor Moisture

of 65 percent when the dry-bulb temperature is 75°F (23.9°C). The proposal was still under debate when this manual went to press.

Building Maintenance

For more information about problems The same fungi (mold and mildew) that cause people discomfort and/or harm resulting from moisture in buildings, also can irreversibly damage building materials, structural components, and refer to Preventing Indoor Air Quality furnishings through premature failure, rot, corrosion, or other degeneration. Problems in Hot, Humid Climates: Moisture-related deterioration affects maintenance costs and operating costs Design and Construction Guidelines, by increasing the frequency of normal cleaning and by requiring periodic published by CH2M Hill, and to replacement of damaged furnishings, such as moldy carpet and wallpaper. Humidity Control Design Guide for Commercial and Institutional Buildings, published by ASHRAE. ■ Climate Considerations

The ASHRAE Handbook—Fundamentals is a popular source for tabular, climatic data representing the outdoor design conditions of many locations. Peak dry-bulb conditions for cooling systems appear under the heading “Cooling DB/MWB” (dry bulb and mean-coincident wet bulb). The ASHRAE weather tables also indicate how often each condition occurs. For example, the 0.4 percent, peak dry-bulb condition for Jacksonville, Florida, is 96°F DB and 76°F MWB (35.7°C DB, 24.5°C MWB). In other words, the outdoor dry-bulb temperature exceeds 96°F (35.7°C) for 0.4 percent of the time, or 35 hours, in an average year. Also, the average, coincident wet-bulb temperature at this dry bulb is 76°F (24.5°C WB).

The sensible load caused by the introduction of outdoor air and weather- dependent space loads, such as conduction, is greatest when the outdoor dry-bulb temperature is highest. Consequently, who design HVAC systems typically and (most of the time) appropriately use the peak dry-bulb condition to determine the required capacity for the cooling coil. The peak latent load resulting from the introduction of outdoor air, however, does not coincide with the highest outdoor dry-bulb temperature; instead, it occurs when the dew point of the outdoor air is highest.

Beginning with the 1997 edition, the design weather data in the ASHRAE Handbook—Fundamentals includes the peak dew-point condition for each location. Although peak dew-point data is seldom used for design purposes, it helps designers analyze the dehumidification performance of HVAC systems and, at the same time, provides a more complete picture of the relevant weather conditions. According to the 2001 Handbook:

The [extreme dew-point] values are used as a check point when analyzing the behavior of cooling systems at part-load conditions, particularly when such systems are used for humidity control as a secondary (or indirect) function. (p. 27.3)

SYS-APM004-EN 5 Sources and Effects of Indoor Moisture

Peak dew-point design conditions for cooling systems appear under the heading “Dehumidification DP/MDB and HR” (dew point/mean-coincident dry bulb and humidity ratio). The 0.4 percent, peak dew-point condition for Jacksonville, Florida, is 76°F DP and 84°F MDB (24.6°C DP, 28.8°C MDB). Outdoor air is cooler at this condition, but contains more moisture than outdoor air at the peak dry-bulb condition.

For outdoor air used for ventilation, the peak sensible load rarely coincides with the peak latent load. Consequently, coils selected for the highest sensible load may not provide sufficient latent capacity when the highest latent load occurs. More often, however, coils controlled to maintain the dry-bulb temperature in the space (sensible capacity) operate with inadequate latent capacity at part- load conditions, even though the latent capacity may be available. Therefore, it is important to evaluate system performance at full-load and part-load conditions, based on the humidity-control requirements of the application.

Moisture problems aren’t confined to hot, humid climates. Too often, indoor humidity problems are incorrectly associated only with buildings located in hot, humid climates. While it is true that such areas experience elevated outdoor humidity levels for a higher percentage of the year, the absolute amount of moisture in the air is comparable to that experienced in many other climates. To illustrate this fact, Table 1 shows the peak dry-bulb and peak dew-point conditions for several cities across the United States. Although these cities are located in different regions, the peak dew-point conditions for most of these locations are remarkably similar.

Table 1. Cooling design conditions for various U.S. cities 1 0.4% Peak 0.4% Peak dry-bulb condition dew-point condition Baltimore, Maryland 93°F DB (34.0°C) 75°F DP (23.8°C) 75°F WB (23.7°C) 83°F DB (28.1°C) Dallas, Texas 100°F DB (37.8°C) 75°F DP (23.7°C) 74°F WB (23.6°C) 82°F DB (28.0°C) Denver, Colorado 93°F DB (33.8°C) 60°F DP (15.6°C) 60°F WB (15.3°C) 69°F DB (20.4°C) Jacksonville, Florida 96°F DB (35.7°C) 76°F DP (24.6°C) 76°F WB (24.5°C) 84°F DB (28.8°C) Los Angeles, California 85°F DB (29.2°C) 67°F DP (19.4°C) 64°F WB (17.7°C) 75°F DB (23.6°C) Minneapolis, Minnesota 91°F DB (32.8°C) 73°F DP (22.5°C) 73°F WB (22.7°C) 83°F DB (28.5°C) San Francisco, California 83°F DB (28.4°C) 59°F DP (15.2°C) 63°F WB (17.0°C) 67°F DB (19.4°C)

1 Source: 2001 ASHRAE Handbook–Fundamentals, Chapter 27 (Table 1B)

6 SYS-APM004-EN Sources and Effects of Indoor Moisture

It is important to understand that indoor humidity problems are not solely attributable to outdoor air brought into the building for ventilation, however. Indoor humidity levels typically depend as much on the sensible and latent loads in the space (and the resulting space ratio), the type of HVAC system, and the method of controlling that system as they do on outdoor conditions. Moisture-related problems therefore can occur in any geographic region where buildings are mechanically ventilated and cooled.

Energy Use

Heightened concern about the cost and availability of energy is hastening the For more information on Standard 90.1 obsolescence of HVAC systems that reheat cold supply air using “new energy” and its effect on the design of HVAC or that mix hot and cold air streams to achieve the desired space temperature. systems, see the Trane Engineers Newsletter titled “90.1 Ways to Save In the United States, the primary standard related to energy consumption in Energy” (ENEWS-30/1). This newsletter commercial buildings is ANSI/ASHRAE/IESNA Standard 90.1–2001, Energy is available at www.trane.com. Standard for Buildings Except Low-Rise Residential Buildings. It provides minimum requirements for energy-efficient building design, including the Standard 90.1 is available from building envelope, system, motors, HVAC system, and service-water ASHRAE’s online bookstore at www.ashrae.org. A user’s guide . accompanies the standard. ■ Some people believe that the requirements of Standard 90.1 make it impossible to maintain indoor humidity within the ranges recommended by Standard 62 and the U.S. EPA (p. 4). Section 6.3.2.1 and Section 6.3.2.3 of Standard 90.1 restrict the use of “new energy” for reheat and limit mixing of hot and cold air streams; the intent is to restrict dehumidification systems and control strategies that waste energy.

Section 6.3.2.3, (excerpted on the next page) is particularly relevant because it specifically addresses HVAC systems that regulate indoor humidity. We address its implications throughout this manual, and describe system designs and control strategies that comply with Standard 90.1 while properly regulating indoor humidity. ■

SYS-APM004-EN 7 Sources and Effects of Indoor Moisture

from ANSI/ASHRAE/IESNA Standard 90.1–2001 on Dehumidification

6.3.2.3 Dehumidification. Where d) Systems serving spaces where specific humidistatic controls are provided, such humidity levels are required to satisfy controls shall prevent reheating, mixing process needs, such as computer of hot and cold airstreams, or other means rooms, museums, surgical suites, and of simultaneous heating and cooling of buildings with refrigerating systems, the same airstream. such as supermarkets, refrigerated warehouses, and arenas. This Exceptions to 6.3.2.3: exception also applies to other a) The system is capable of reducing applications for which fan volume supply air volume to 50% or less of the controls listed in accordance with design airflow rate or the minimum Exception (a) are proven to be rate specified in 6.1.3 of ASHRAE impractical to the enforcement agency. Standard 62, whichever is larger, e) At least 75% of the energy for before simultaneous heating and reheating or for providing warm air in cooling takes place. mixing systems is provided from a site- b) The individual fan cooling unit has a recovered (including condenser heat) or design cooling capacity of 80,000 Btu/h site solar energy source. (23 kW) or less and is capable of f) Systems where the heat added to the unloading to 50% capacity before airstream is the result of the use of a simultaneous heating and cooling desiccant system and 75% of the heat takes place. added by the desiccant system is c) The individual mechanical cooling unit removed by a , either has a design cooling capacity of 40,000 before or after the desiccant system Btu/h (12 kW) or less. An individual with energy recovery. ■ mechanical cooling unit is a single system composed of a fan or fans and a cooling coil capable of providing mechanical cooling.

8 SYS-APM004-EN Dehumidification Primer

Types of Dehumidification

Maintaining the indoor humidity within the desired range requires a means of either locally removing moisture from the air that is already in the space, or replacing that moisture-laden air with drier air that was dehumidified elsewhere.

Local Dehumidification

Figure 3. Local dehumidification Portable , like those used in many residential basements, provide dedicated, local dehumidification. These devices (Figure 3) commonly use mechanical to remove moisture from the air in the space: an coil dehumidifies and coincidentally cools the entering air, and a condenser coil reheats the leaving air. Humidity in the space decreases, while the dry-bulb temperature increases.

Simple, in-space air conditioners often coincidentally dehumidify the space as they cool it; but do not confuse these devices with dedicated dehumidification equipment. The evaporator coil in a packaged terminal air conditioner (“PTAC,” Figure 4) responds to the room , directly cooling a mixture of recirculated return air and outdoor air, and removing moisture in the process. At full load, the air conditioner usually provides adequate dehumidification because the thermostat keeps the unit running and the coil cold.

To avoid overcooling the space at part load, however, the thermostat reduces the sensible-cooling capacity of the coil by cycling it on and off. Cycling raises the average temperature of the coil, which significantly reduces its Figure 4. Packaged terminal air conditioner dehumidification (latent-cooling) capacity. Simple air conditioners, such as the PTAC, may provide adequate coincidental dehumidification for spaces with constant cooling loads. When the varies widely, however, additional equipment and/or controls may be required for adequate dehumidification at part-load conditions.

SYS-APM004-EN 9 Dehumidification Primer

Remote Dehumidification

Figure 5. Remote dehumidification The central air-conditioning system commonly serves as a remote source of dehumidification for the occupied spaces in a commercial or industrial building. To maintain an acceptable indoor humidity, the system must be properly designed and controlled so that the air it supplies is drier than the air in the space (Figure 5). In effect, the supply air must be dry enough to “soak up” the water vapor in the space; the absorbed moisture is then carried from the space in the return air.

Depending on the type of system and method of control, central air- conditioning units may or may not be able to adequately dehumidify the space at all load conditions. The dehumidification performance of various system types and control methods is discussed in the next three chapters.

Processes for Dehumidification

An air-conditioning system typically uses one of two processes to dehumidify the supply air that ultimately reaches the space: condensation on a cold coil or adsorption via a desiccant.

Condensation on a Cold Coil

Figure 6. “Cold-coil” dehumidification Water vapor condenses on a surface if the temperature of the surface is colder than the dew point of the moist air in contact with it. Controlled condensation dehumidifies an air stream by directing it across the cold surfaces of a finned- tube coil. Circulating either or evaporating through the coil makes the coil surfaces cold enough to induce condensation. As warm, moist air passes through the coil, water vapor condenses on the cold surfaces (Figure 6); the condensate (liquid water) then drains down the coil fins and collects in the drain pan, where it is piped from the air handler. The air leaves the coil cooler and drier.

A psychrometric chart can illustrate how “cold-coil” dehumidification works. This special-purpose chart (Figure 7) represents the interrelated physical properties of moist air: dry-bulb (DB), wet-bulb (WB), and dew-point (DP) temperatures; relative humidity (RH), (h), and humidity ratio (W). For example, if sensible heat is added or removed with no change in moisture content, the condition of the air moves horizontally on the chart. Conversely, if moisture is added or removed without changing the dry-bulb temperature, then the condition of the air moves vertically on the chart.

Figure 8 (p. 12) illustrates what happens when a mixture of outdoor air and recirculated return air at 80°F DB, 60°F DP (26.7°C DB, 15.6°C DP), enters a cold coil. The temperature of the coil surface is well below the dew point of the

10 SYS-APM004-EN Dehumidification Primer

Figure 7. Psychrometric chart

entering air. Sensible cooling occurs as the air passes through the coil; on the chart, the air condition moves horizontally to the left. When the condition of the air nears the saturated state (100 percent-relative humidity), moisture begins to condense on the cold surface of the coil. The condition of the air now moves The Trane psychrometric chart includes diagonally down and to the left on the chart, representing the removal of both a series of “coil curves” that depict the sensible heat and moisture. Cool, dry air leaves the coil in this example at approximate performance of a wide 55°F DB, 53°F DP (12.8°C DB, 11.7°C DP). range of coil configurations (Figure 19, p. 26). These curved lines, established No moisture removal occurs unless the temperature of the coil surface is from hundreds of laboratory tests of lowered below the dew point of the entering air. If the coil surface is not colder various coil geometries at different air than the dew point, only sensible cooling takes place. Sensible cooling without and temperatures, represent dehumidification is especially common during part-load operation of a the changes in dry-bulb and dew-point constant-volume system. That’s because constant-volume systems (discussed temperatures as air passes through a in the next chapter) respond to part-load conditions by reducing coil capacity, “typical” cooling coil. Of course, exact coil performance depends on actual which raises the temperature of the coil surface and of the supply air. coil geometry and can be precisely For comfort-cooling applications that do not require a supply-air dew point determined by software that accurately models the performance of the lower than 40°F to 45°F (4.5°C to 7°C), cold-coil condensation is the traditional specific coils. ■ choice for dehumidification because of its low first cost and low operating cost. Given that decision, the next choice is whether to use chilled water or refrigerant to make the coil cold.

SYS-APM004-EN 11 Dehumidification Primer

Figure 8. Psychrometric analysis of “cold-coil” dehumidification

Chilled water systems, with their individually selected components, provide the necessary design flexibility for applications that require low supply-air dew points, that is, dew points approaching 40°F to 45°F (4.5°C to 7°C).

By contrast, most DX systems are packaged. Although prematched refrigeration and air-handling components lower the initial cost of the system, they also make the system less flexible by deferring certain design decisions to the manufacturer. A traditional, “off-the-shelf” packaged DX system is optimized for operation at about 400 cfm/ton (0.054 m³/s/kW), which prevents it from achieving “low” dew points. Specially designed DX equipment can reach dew points of 45°F to 50°F (7°C to 10°C) because they are designed to deliver less airflow (cfm) per cooling ton (L/s per kW).

When space loads or process requirements dictate an even lower supply-air dew point, moisture adsorption is preferred for dehumidification.

Condensate management

When a cold coil is used for dehumidification, moisture condenses from the air Managing Building Moisture, Trane onto the surface of the coil and falls into the drain pan, where it is piped from applications engineering manual the air handler. Too often, inattention to proper trapping of the condensate line SYS-AM-15, discusses proper design causes “spitting,” which dampens the insulation inside the air handler and and installation of condensate traps for ductwork, or restricts flow from the drain pan, causing it to overflow. Both draw-through and blow-through coil situations create opportunities for microbial growth. To assure proper configurations. ■ condensate removal under all operating conditions, comply with the manufacturer’s instructions for drain-line installation and trapping.

12 SYS-APM004-EN Dehumidification Primer

Adsorption Using a Desiccant

Solid desiccants are typically used for Desiccants used for commercial dehumidification are selected for their ability to dehumidification equipment applied in collect large quantities of water vapor. The porous surface of the desiccant commercial and institutional buildings. attracts and retains water molecules from the passing air stream. This Liquid desiccants are also available, but dehumidification process is described as adsorption because the collected they are traditionally used in industrial moisture does not chemically or physically alter the desiccant. applications. Refer to the “Desiccant Dehumidification and Pressure-Drying Vapor pressure at the desiccant surface is directly proportional to the surface Equipment” chapter of the ASHRAE temperature of the desiccant and the amount of moisture adsorbed there. Handbook–HVAC Systems and When the desiccant is cool and dry, its surface vapor pressure is low; when the Equipment for more information. ■ desiccant is warm and moist, its surface vapor pressure is high. Water vapor migrates from areas of high vapor pressure to areas of low vapor pressure. Consequently, a desiccant with a low surface vapor pressure will adsorb water molecules from the surrounding air, while a desiccant with a high surface vapor pressure will reject water molecules to the surrounding air.

The most common application of adsorption for commercial dehumidification uses a rotating wheel that contains a fluted, desiccant-coated medium. The wheel rotates between two air streams: the “process” air stream and the “regeneration” air stream. Warm, moist process air enters one side of the rotating wheel, where water vapor collects on the desiccant surface. As the wheel rotates, the moisture-laden portion moves into the regeneration air stream, where the collected water vapor is released and transported outdoors. The cycle repeats with each rotation, providing continuous dehumidification.

The temperature of the regeneration air determines whether the adsorption process is passive or active.

Passive adsorption

Figure 9. Total-energy wheel When the regeneration air is drier than the process air, but is not heated to drive the moisture from the desiccant, the dehumidification process is considered passive adsorption.

An example of passive adsorption is the use of building exhaust air to regenerate the desiccant of a total-energy/enthalpy wheel (Figure 9). The wheel is mounted so that the minimum outdoor (process) airflow required for ventilation passes through half of the wheel, while exhaust (regeneration) air passes through the other half. The wheel rotates quickly—between 20 rpm and 60 rpm—alternately exposing the desiccant to process air and regeneration air.

In the summer, when the outdoor air is hot and humid, the total-energy wheel cools and dehumidifies the entering outdoor air by transferring sensible heat and moisture to the cooler, drier exhaust air (Figure 10, p. 14). Desiccant regeneration occurs at a low temperature—78°F (25.6°C) in this example— without additional heat. In the winter, when the outdoor air is cold and dry, the

SYS-APM004-EN 13 Dehumidification Primer

Figure 10. Example of passive adsorption performed by a total-energy wheel Jacksonville, Florida

total-energy wheel warms and humidifies the entering outdoor air by transferring sensible heat and moisture from the warmer, moister exhaust air.

Although desiccant-coated devices, such as the total-energy wheel, reduce the sensible heat and moisture content of entering outdoor air, these passive Refer to Air-to-Air Energy Recovery in adsorption devices are not considered as dehumidification equipment. Such HVAC Systems, Trane applications devices are less than 100 percent effective: When it is humid outside, process engineering manual SYS-APM003-EN, air leaving the wheel always contains more moisture than regeneration air for more information about using the (from the space) entering the exhaust side of the wheel. By definition, a passive passive adsorption of total-energy adsorption device cannot dehumidify the space because the air leaving the wheels to precondition outdoor air. ■ supply side of the device never can be drier than the space. As demonstrated in “Dehumidifying with Constant-Volume Mixed Air” (pp. 27–29), a space under these conditions will always require additional dehumidification.

Active adsorption

In the active adsorption process, the moisture-collecting ability of the desiccant is improved by adding sensible heat to the regeneration air before it enters the desiccant. Figure 11 depicts the active desiccant wheel mounted so that the outdoor (process) air for ventilation passes through half of the wheel, while regeneration air (either a separate outdoor air stream or exhaust air from the building) passes through the other half.

As the active desiccant wheel slowly rotates between 10 rph and 30 rph, it removes moisture from the outdoor (process) air stream and releases sensible heat (Figure 12). The resulting temperature increase is directly proportional to the amount of moisture removed from the process air. In this example, active adsorption dehumidifies the process air to 44°F DP (6.7°C DP) and raises the temperature of the process air to 120°F DB (48.9°C DB). Consequently, the process air must be cooled before it is delivered to the building’s occupied

14 SYS-APM004-EN Dehumidification Primer

Figure 11. Active adsorption system

spaces. The psychrometric analysis (Figure 12) for this example system shows that the cooling coil lowers the temperature of the process air to 80°F DB (26.7°C DB).

On the regeneration side of the system, a gas-fired heater raises the temperature of the regeneration air. Depending on the dew-point target for the process air, regeneration air temperatures typically range from 130°F to 250°F (54°C to 121°C). The warmer that the regeneration air is, the drier the resulting process air will be.

Recall that Section 6.3.2.3 of ASHRAE Standard 90.1 requires that humidistatic controls prevent simultaneous heating and cooling of the same air stream. It therefore addresses active-adsorption dehumidification, which heats the process air and requires downstream cooling. Exception F of Section 6.3.2.3

Figure 12. Example performance for an active adsorption system

SYS-APM004-EN 15 Dehumidification Primer

(p. 8 in this manual) defines the conditions for compliance; that is, an active desiccant system must recover 75 percent of the heat that adsorption adds to the process air.

For example, if the adsorption process adds 100,000 Btu/hr (29.3 kW) of Sensible heat added by the adsorption process: sensible heat to the process air, then 75,000 Btu/hr (22.0 kW) of energy must be × , × ()° removed from that same air. One possible design solution places a sensible- Qs = 1.085 2 634 cfm 120 F – 85°F energy, air-to-air heat exchanger downstream of the active desiccant wheel to = 100,000 Btu/hr transfer at least 75,000 Btu/hr (22.0 kW) of heat from the hot, dry process air to the regeneration air. Another possible solution adds an air-to-air energy- ()× × []° Qs = 1.21 1.24 m³/s 48.9 C – 29.4°C recovery device, such as a total-energy wheel, upstream of the active desiccant ()= 29.3 kW wheel to precondition the outdoor air and transfer at least 75,000 Btu/hr (22.0 kW) of heat (sensible plus latent energy) from the process air to another air stream.

Typical applications for adsorption dehumidification

Total-energy wheels and other types of passive adsorption devices are used in all types of HVAC systems to precondition outdoor air. This practice enables downsizing of cooling and heating equipment, which reduces the initial cost of the system; it also saves energy by reducing the cooling and heating loads associated with ventilation.

Active adsorption systems are primarily used in applications where high internal latent loads or process requirements dictate a lower-than-normal dew point (below a threshold of 40°F to 45°F [4.5°C to 7°C]) for the supply air. Typical applications include supermarkets, ice rinks, museums, industrial drying processes, and other spaces that require exceptionally dry air. Given the relatively high first cost, the energy required to heat the regeneration air, and the additional energy needed to post-cool the process air, active adsorption systems are seldom used in comfort-cooling applications. The succeeding chapters of this manual therefore focus exclusively on comfort-cooling systems that use “cold coil” condensation for dehumidification.

16 SYS-APM004-EN Dehumidification Primer

Implications for HVAC Control

The next three chapters examine three types of HVAC systems, which are Even when the average relative distinguished from one another by how each system delivers ventilation air humidity in a conditioned space is low, to the space: constant-volume mixed air, variable-volume mixed air, and high relative humidities can develop dedicated outdoor air. In each case, the central theme is “cold coil” near cold surfaces and increase the dehumidification during full-load and part-load comfort cooling. The likelihood of condensation. Enforcing a performance benchmark is a relative humidity of 60 percent, which is the upper maximum relative humidity of limit currently recommended by ASHRAE Standard 62. 60 percent or 65 percent should make most surfaces 12ºF to 15ºF (6.7ºC to Certain control strategies will affect the dehumidification performance of any of 8.3ºC) warmer than the space dew point these HVAC systems: and generally avoid concentrations of water vapor near surfaces. ■ ■ Humidity control during unoccupied periods ■ Building pressurization ■ Airside economizing

Brief descriptions of how each of these control strategies affects dehumidification performance follow. Specific application considerations by system type are discussed within the appropriate chapter.

Humidity Control during Unoccupied Periods

Latent loads associated with occupants and their activities make humidity control important during scheduled operation. But after-hours humidity control is also important in facilities, such as schools, with few or no occupants for extended periods. ASHRAE offers the following recommendation:

In humid climates, serious consideration should be given to dehumidification during the summer months, when the school is unoccupied, to prevent the growth of mold and mildew. (1999 ASHRAE Handbook–Applications, Chapter 6, p. 6.3)

Controlling humidity at all times of the day can greatly reduce the risk of microbial growth on building surfaces and furnishings. Wet-cleaning procedures (mopping floors, shampooing carpets) bring large amounts of moisture into the building and usually take place when the building is unoccupied. Drying wet surfaces is critical to prevent microbial growth. For shampooed carpets, this is best accomplished by providing adequate air motion and dehumidification during unoccupied hours.

SYS-APM004-EN 17 Dehumidification Primer

Building Pressurization

HVAC systems do more than provide heating, cooling, and ventilation; they also Refer to Building Pressurization bring makeup air into the building to replace the air removed by local exhaust Control, Trane applications engineering fans (in restrooms and kitchens, for example) and combustion equipment manual AM-CON-17, for additional (, ). Turning off the ventilation system during unoccupied information about how to regulate periods while allowing these devices to continue operating creates negative building pressure through design and pressure inside the building. Unconditioned outdoor air infiltrates the building, control of the HVAC system. ■ which can raise the dew point in the envelope (risking condensation) and increase the humidity in the occupied space (perhaps beyond the limit recommended by ASHRAE).

One solution is to design the building control system so that it turns off all local exhaust fans and combustion equipment whenever the ventilation system is off. However, this approach may require a manual override to accommodate after-hours cleaning.

Wind, variable operation of local exhaust fans, and “” in multistory buildings can create building pressure fluctuations despite a properly balanced HVAC system. Therefore, controlling building pressure directly may be desirable to prevent negative pressure from developing inside the building… and it may be necessary during economizer operation to prevent overpressurization.

Airside Economizing

An airside economizer can lower operating costs by using outdoor air to help offset building cooling loads. When outdoor conditions are suitable for natural cooling, the outdoor-air opens fully, assisting the mechanical cooling equipment by offsetting as much of the cooling load as possible. At cooler outdoor conditions, the outdoor-air damper maintains the target temperature in the space by modulating between its full-open and minimum-open positions.

When the outdoor air is too warm or too cold for economizing, the outdoor-air damper remains at the minimum-open position to provide the necessary quantity of outdoor air for ventilation; meanwhile, the cooling or heating coil satisfies the space load.

Proper control of the airside economizer is critical to maximize energy savings without creating potential humidity problems. ■

18 SYS-APM004-EN Dehumidifying with Constant-Volume Mixed Air

Mixed-air systems use an air handler to condition a combination of outdoor air and recirculated return air before delivering this mixed air to each space. A constant-volume, mixed-air system supplies an unchanging quantity of air, usually to a single space or thermal zone. The temperature of the supply air modulates in response to the varying sensible-cooling load in the space.

“Basic” constant-volume systems, which consist of an air handler containing a fan and a cold coil (Figure 13), indirectly affect indoor humidity. A thermostat compares the dry-bulb temperature in the space to the setpoint; it then modulates the cooling coil until the cooling capacity matches the sensible load—that is, until the space temperature and setpoint match. Reducing the capacity of the cooling coil results in a warmer coil surface and less dehumidification. Similarly, increasing the coil capacity makes the coil surface colder and provides more dehumidification.

The peak sensible load on the cooling coil rarely coincides with the peak latent load. So, a cooling coil selected for the highest sensible load (in some air- handling arrangements) may not provide sufficient capacity when the highest latent load occurs. More often, however, a cooling coil that is controlled to maintain the space dry-bulb temperature often operates without adequate moisture-removal capacity at peak latent-load conditions. As the following examples reveal, accurate predictions of dehumidification performance require an analysis of system operation at both full-load and part-load conditions.

Figure 13. Basic, constant-volume HVAC system

Analysis of Dehumidification Performance

10, 000 cfm × 9 air changes/hr V = ------sa 60 min/hr Consider a 10,000 ft³ (283 m³), 30-occupant classroom in Jacksonville, Florida. = 1,500 cfm For thermal comfort, the space setpoint is 74°F DB (23.3°C DB). Supply airflow

Vsa is based on nine and is 1,500 cfm (0.7 m³/s). ASHRAE §·283 m³ × 9 air changes/hr V = ------Standard 62 requires 15 cfm (8 L/s) of outdoor air per person for adequate ©¹sa 3, 600 sec/hr ventilation; so, 450 cfm (0.21 m³/s) of the supply air must be outdoor air. ()= 0.7 m³/s

SYS-APM004-EN 19 Dehumidifying with Constant-Volume Mixed Air

Figure 14. Dehumidification performance of a basic, constant-volume HVAC system at various outdoor conditions

Design Full load Part load condition Peak dry bulb Peak dew point Mild, rainy OA 96.0°F DB, 76.0°F WB 76.0°F DP, 84.0°F DB 70.0°F DB, 69.0°F WB RA 74.0°F DB, 52.4% RH 74.0°F DB, 67.0% RH 74.0°F DB, 73.0% RH MA 80.6°F DB 77.0°F DB 72.8°F DB SA 55.7°F DB 63.0°F DB 66.5°F DB

The classroom is air conditioned by a basic constant-volume system, which uses a chilled water coil to cool and dehumidify the supply air. A modulating valve controls coil capacity.

Performance at peak dry-bulb (full-load) condition. According to At the peak dry-bulb condition: the 2001 ASHRAE Handbook—Fundamentals, the peak dry-bulb condition for 29, 750 Btu/hr ------Jacksonville is 96°F DB, 76°F WB (35.7°C DB, 24.5°C WB). At this condition, the SHR = , , 29 750 Btu/hr + 5 250 Btu/hr sensible and latent loads calculated for the classroom—29,750 Btu/hr (8.7 kW) = 0.85 and 5,250 Btu/hr (1.5 kW), respectively—yield a sensible-heat ratio (SHR) of 0.85 §·8.7 kW in that space. Given the supply airflow of 1,500 cfm (0.7 m³/s), satisfying the ©¹SHR = ------8.7 kW + 1.5 kW sensible-cooling load and maintaining the space at 74°F DB (23.3°C DB) ()= 0.85 requires 55.7°F (13.1°C) supply air.

Figure 14 summarizes the psychrometric analysis of this system’s full-load × , × ()° Qs = 1.085 1 500 cfm 74 FT– supply dehumidification performance. At the peak dry-bulb condition, controlling the , ∴ ° temperature in the space to 74°F (23.3°C) will result in a comfortable relative = 29 750 Btu/hr Tsupply = 55.7 F humidity of 52 percent. ()× × []° Qs = 1.21 0.7 m³/s 23.3 CT– supply Note: To simplify the analysis, which is detailed in Appendix A of this manual, ()∴ ° = 8.7 kW Tsupply = 13.1 C the latent load in the classroom is limited to moisture generated by the occupants. A higher relative humidity would result if other sources of indoor moisture, such as infiltration and vapor-pressure diffusion, were considered. The cooling coil is expected to offset the “non-space” latent load that results from ventilating the classroom with outdoor air.

20 SYS-APM004-EN Dehumidifying with Constant-Volume Mixed Air

× , × () QT = 4.5 1 500 cfm 31.4– 22.9 Btu/lb The total capacity required from the cooling coil at the peak dry-bulb condition = 57,375 Btu/hr= 4.78 tons is 4.78 tons (16.8 kW). At full load, the cooling coil removes both sensible heat and moisture (latent ()× × [] QT = 1.2 0.7 m³/s 73.1– 53.1 kJ/kg heat), directly controlling space temperature and indirectly affecting space ()= 16.8 kW humidity.

Performance at peak dew-point (part-load) condition. As the sensible- cooling load in the space decreases, a constant-volume HVAC system allows The peak dew-point condition does not the supply-air temperature to rise by reducing the capacity of the cooling coil. In necessarily represent the worst-case this example system, coil capacity is reduced by modulating the water valve. condition for humidity control. It simply Although this control action successfully maintains the desired dry-bulb is an easy “test case” for analyzing part- temperature for the space, raising the supply-air temperature also reduces the load dehumidification performance. ■ amount of moisture that condenses on the coil; space humidity rises. In other words, making the coil surface warmer decreases the rate at which moisture condenses from the mixed air.

To determine whether a system will provide adequate dehumidification at part load, analyze performance at the peak dew-point condition. For our Jacksonville classroom, the peak dew-point condition is 76°F DP, 84°F DB (24.6°C DP, 28.8°C DB). The cooler outdoor dry-bulb temperature and correspondingly lower solar and conducted heat gains reduce the sensible load in the classroom to 17,850 Btu/hr (5.2 kW). Because the classroom’s latent load At the peak dew-point condition: remains unchanged at 5,250 Btu/hr (1.5 kW), however, the sensible-heat ratio × , × ()° Qs = 1.085 1 500 cfm 74 FT– supply (SHR) for the space drops to 0.77. Consequently, the 1,500 cfm (0.7 m³/s) of , ∴ ° supply air must be delivered at a higher temperature, 63°F (17.2°C) in this case, = 17 850 Btu/hr Tsupply = 63 F to avoid overcooling the space. ()× × []° Qs = 1.21 0.7 m³/s 23.3 CT– supply Warmer supply air, combined with the lower space SHR, raises the relative ()∴ ° = 5.2 kW Tsupply = 17.2 C humidity in the classroom from 52 percent to 67 percent (Figure 14)—well above the 60 percent limit that ASHRAE recommends. Although the cooling coil could provide additional cooling (up to 4.78 tons [16.8 kW] if sized for the design dry-bulb condition), the thermostat reduces coil capacity to 3.66 tons (12.9 kW). This control action maintains the dry-bulb temperature in the classroom at setpoint, but space humidity rises. Oversizing the cooling coil will not prevent the shortfall in latent capacity if system control is based solely on the dry-bulb temperature in the space.

Performance on a mild, rainy day (part-load condition). Although the peak dew-point condition is helpful for analyzing the part-load dehumidification On a mild, rainy day: performance of an HVAC system, do not assume that it represents the worst- × , × ()° Qs = 1.085 1 500 cfm 74 FT– supply case condition for space humidity control. Most of the time, the humidity in the , ∴ ° space depends more on the space SHR and the system control strategy than on = 12 250 Btu/hr Tsupply = 66.5 F outdoor conditions. ()× × []° Qs = 1.21 0.7 m³/s 23.3 CT– supply Consider our example Jacksonville classroom on a mild, rainy day. At 70°F DB, ()∴ = 3.6 kW Tsupply = 19.2°C 69°F WB (21.2°C DB, 20.6°C WB), the sensible load in the classroom drops

SYS-APM004-EN 21 Dehumidifying with Constant-Volume Mixed Air

further—this time to 12,250 Btu/hr (3.6 kW). Given an unchanged latent load of 5,250 Btu/hr (1.5 kW) due to occupants, the SHR in the classroom drops to 0.70. To prevent overcooling, the thermostat reduces the cooling coil capacity to 1.63 tons (5.74 kW) so that the 1,500 cfm (0.7 m³/s) of supply air is delivered to the classroom at 66.5°F (19.2°C). How is humidity in the classroom affected? Relative humidity climbs to 73 percent!

Application Considerations

Ventilation

The 1989 revision of ASHRAE Standard 62 increased the required per-person ventilation rate from 5 cfm to 20 cfm (from 3 L/s to 10 L/s) for office buildings, and from 5 cfm to 15 cfm (from 3 L/s to 8 L/s) for schools. Bringing more outdoor air into the building to satisfy ventilation requirements significantly increases the cooling and heating loads on the HVAC system. But, does bringing more outdoor air into the building for ventilation cause moisture- related IAQ problems? Some people think so. Let’s examine what happens if the example classroom receives only 150 cfm (0.07 m³/s) of outdoor air for ventilation rather than the 450 cfm (0.21 m³/s) that ASHRAE Standard 62 requires. Only the ventilation load differs from the previous examples (Figure 14, p. 20); the sensible- and latent-cooling loads for the classroom are unchanged, as are the supply-air temperatures and sensible-heat ratios (SHRs) for that space.

Figure 15 illustrates the effect of underventilating the classroom:

■ At the peak dry-bulb condition, the relative humidity drops from 52 percent to approximately 50 percent.

Figure 15. Dehumidification performance of a basic, constant-volume HVAC system with underventilation

Design Full load Part load condition Peak dry bulb Peak dew point Mild, rainy OA 96.0°F DB, 76.0°F WB 76.0°F DP, 84.0°F DB 70.0°F DB, 69.0°F WB RA 74.0°F DB, 50.4% RH 74.0°F DB, 64.6% RH 74.0°F DB, 70.0% RH MA 76.2°F DB 75.0°F DB 73.6°F DB SA 55.7°F DB 63.0°F DB 66.5°F DB

22 SYS-APM004-EN Dehumidifying with Constant-Volume Mixed Air

■ At the peak dew-point condition, the relative humidity drops from 67 percent to approximately 65 percent. ■ On a mild and rainy day, the relative humidity drops from 73 percent to 70 percent.

In this example, a lower ventilation rate slightly reduces space humidity; however, the reduction may not remediate the humidity problems inherent to a constant-volume control strategy that is based only on the space dry-bulb temperature. More importantly, lowering the ventilation rate can create other IAQ problems.

Lowering the ventilation rate does not significantly improve dehumidification performance because space humidity depends on the dew point of supply air leaving the cooling coil—not on the dew point of mixed air entering the coil. Lowering the ventilation rate introduces a smaller quantity (and percentage) of outdoor air, which lowers the dew point of the mixed air entering the cooling coil. But as Figure 16 shows, the dew point of the supply air leaving the coil is not much lower than when the system treats a larger quantity of outdoor air.

A lower sensible load, however, will result in a supply-air temperature that is too warm to induce moisture condensation from the air passing through the coil. Such conditions will also produce a considerably different supply-air dew point.

In other words, when the outdoor air is more humid than the desired indoor humidity, the extent to which ventilation (outdoor air) affects indoor humidity depends on the loads in the space and the supply-air condition.

Figure 16. Effect of ventilation on the dehumidification performance of a basic, constant-volume HVAC system

Design Ventilation airflow condition 450 cfm OA 150 cfm OA OA 84.0°F DB, 84.0°F DB, (peak DP) 76.0°F DP 76.0°F DP RA 74.0°F DB, 74.0°F DB, 67.0% RH 64.6% RH MA 77.0°F DB, 75.0°F DB, 67.0°F DP 63.0°F DP SA 63.0°F DB, 63.0°F DB, 61.0°F DP 59.5°F DP

SYS-APM004-EN 23 Dehumidifying with Constant-Volume Mixed Air

Climate

The previous example demonstrated that the quantity of outdoor air is not necessarily the primary cause of indoor humidity problems. Can the same be said for the condition of the outdoor air? Table 2 shows peak dew-point and mild, rainy conditions for seven U.S. cities with differing climates. The table also shows the relative humidity that would result in the same example classroom, with its basic constant-volume HVAC system, at each condition.

Notice the similarity of the peak dew-point conditions for most locations. In these regions, the resulting space relative humidity is similar at peak dew-point conditions. In the dry climates (Denver and San Francisco), the system performs better because the outdoor air is dry enough to provide a dehumidifying effect. Although the increased frequency and duration of humid conditions is greater in hot, humid climates, the conditions capable of causing moisture-related problems occur in many regions. Ignoring system operation at part-load conditions can lead to high indoor humidity in dry climates as well as in hot, humid locales.

Table 2. Constant-volume system performance for various cities in the United States Peak dew-point condition Mild, rainy condition Location Outdoors Indoor RH Outdoors Indoor RH Baltimore, Maryland 75°F (23.8°C) DP, 62% 70°F (21.2°C) DB, 65% 83°F (28.1°C) DB 69°F (20.6°C) WB Dallas, Texas 75°F (23.7°C) DP, 66% 70°F (21.2°C) DB, 68% 82°F (28.0°C) DB 69°F (20.6°C) WB Denver, Colorado 60°F (15.6°C) DP, 55% 63°F (17.2°C) DB, 58% 69°F (20.4°C) DB 61°F (16.1°C) WB Jacksonville, Florida 76°F (24.6°C) DP, 67% 70°F (21.2°C) DB, 73% 84°F (28.8°C) DB 69°F (20.6°C) WB Los Angeles, California 67°F (19.4°C) DP, 62% 63°F (17.2°C) DB, 65% 75°F (23.6°C) DB 62°F (16.7°C) WB Minneapolis, Minnesota 73°F (22.5°C) DP, 66% 70°F (21.2°C) DB, 70% 83°F (28.5°C) DB 69°F (20.6°C) WB San Francisco, California 59°F (15.2°C) DP, 56% 54°F (12.2°C) DB, 56% 67°F (19.4°C) DB 53°F (11.7°C) WB

Packaged DX Equipment

In constant-volume applications with high ventilation requirements, packaged direct-expansion (DX) air-conditioning equipment can compound indoor humidity problems. More outdoor air, especially in humid climates, increases the required cooling and dehumidification capacity.

24 SYS-APM004-EN Dehumidifying with Constant-Volume Mixed Air

The key to designing a system with adequate dehumidification capability at all load conditions lies with determining the proper relationship between airflow and cooling capacity. Sensible cooling loads in the space—not ventilation requirements—dictate airflow for the space (unless the required ventilation exceeds the airflow needed to cool the space). The increase in outdoor air required for ventilation requires more cooling capacity. For a given space load, an increase in the ventilation load results in less airflow per cooling ton (m³/s per kW). The flexibility of applied systems, such as chilled-water air handlers, normally lets you select equipment based on a specific airflow rate (cfm [m³/s]) and a specific cooling capacity (tons [kW]). By contrast, packaged unitary systems (a direct-expansion rooftop air conditioner, for example) typically limit your selection to a cfm/ton (m³/s/kW) range of application.

At the peak dry-bulb condition: Recall that a chilled-water coil provides the air conditioning for the classroom in × , × ()° Qs = 1.085 1 750 cfm 74 FT– supply the preceding examples. The coil was selected to deliver 4.78 tons (16.8 kW) of , ∴ ° cooling capacity at 1,500 cfm (0.7 m³/s) supply airflow, resulting in a flow-to- = 29 750 Btu/hr Tsupply = 58.3 F capacity ratio of 314 cfm/ton (0.042 m³/s/kW). ()× × []° Qs = 1.21 0.83 m³/s 23.3 CT– supply Most packaged DX air conditioners, however, are designed to operate between ()∴ ° = 8.7 kW Tsupply = 14.6 C 350 and 450 cfm/ton (0.047 and 0.060 m³/s/kW). The classroom in our example would require a nominal 5-ton (17.6 kW) air conditioner that delivers no less than 350 cfm/ton (0.047 m³/s/kW), or 1,750 cfm (0.83 m³/s). To assure adequate cooling capacity at full-load conditions, you must accept this higher-than- At the peak dew-point condition: required supply airflow instead of the desired 1,500 cfm (0.7 m³/s). Because the × , × ()° Qs = 1.085 1 750 cfm 74 FT– supply sensible load is unchanged, however, the 1,750 cfm (0.83 m³/s) of supply air , ∴ ° must be delivered at 58.3°F (14.6°C) to avoid overcooling the space. = 17 850 Btu/hr Tsupply = 64.6 F

()× × []° Oversized supply airflow results in warmer air leaving the cooling coil. In non- Qs = 1.21 0.83 m³/s 23.3 CT– supply arid climates, this higher supply-air temperature reduces the dehumidification ()= 5.2 kW∴ T = 18.1°C supply capacity of the system. At the peak dry-bulb condition, the relative humidity in the example classroom increases from 52 percent to 56 percent (Figure 17).

Figure 17. Dehumidification performance of a basic, constant-volume, packaged DX air conditioner

Design Full load Part load condition Peak dry bulb Peak dew point Mild, rainy OA 96.0°F DB, 76.0°F WB 76.0°F DP, 84.0°F DB 70.0°F DB, 69.0°F WB RA 74.0°F DB, 56.2% RH 74.0°F DB, 68.7% RH 74.0°F DB, 74.0% RH MA 79.7°F DB 76.6°F DB 73.0°F DB SA 58.3°F DB 64.6°F DB 67.5°F DB

SYS-APM004-EN 25 Dehumidifying with Constant-Volume Mixed Air

Not surprisingly, the classroom becomes even more humid at the peak dew-point condition. With the thermostat throttling the capacity of the cooling coil to meet the smaller space sensible load, the 64.6°F (18.1°C) supply air offers even less dehumidification; relative humidity climbs to about 69 percent.

On a mild, rainy day: As for the mild and rainy day, the supply-air temperature rises to 67.5°F (19.7°C) × , × ()° Qs = 1.085 1 750 cfm 74 FT– supply and the humidity in the space increases to 74 percent. = 12,250 Btu/hr∴ T = 67.5°F supply Selecting larger packaged unitary equipment to provide additional cooling ()× × []° capacity can yield a higher supply airflow, correspondingly warmer supply air, Qs = 1.21 0.83 m³/s 23.3 CT– supply and an elevated indoor humidity. ()∴ ° = 3.6 kW Tsupply = 19.7 C Note: Excess supply airflow and the increased humidity that accompanies it also can result from a conservative estimate of the space sensible load. When selecting cooling equipment (whether chilled water or DX) for constant-volume applications, exercise particular care to avoid oversizing the supply airflow.

Figure 18. Henderson’s “latent capacity cycling in DX equipment further complicates humidity problems. degradation” model When the turn off, condensate on the cooling coil re-evaporates and the supply fan “pushes” the moisture downstream to the occupied space. Recent research (Henderson, 1998) led to the development of a “latent capacity degradation model” for DX equipment in which the compressors cycle and the supply fan runs constantly. This model (Figure 18) predicts the latent cooling (dehumidification) capacity of the equipment as a function of the run-time fraction, which represents how long the compressor operates during an hour.

Plotting the latent capacity degradation model on the psychrometric chart (Figure 19) reveals that, over time, there is little difference in performance for cycling DX systems versus chilled-water coils with modulating valves.2 In other words, given the same supply airflow, the resulting relative humidity indoors will be essentially the same regardless of which type of system (DX or applied chilled water) is used.

Figure 19. Comparison of “latent capacity degradation” model to a Trane coil curve

26 SYS-APM004-EN Dehumidifying with Constant-Volume Mixed Air

Total-Energy Recovery

Some design engineers believe that passive energy-recovery devices provide Refer to Air-to-Air Energy Recovery in adequate space dehumidification. A passive, total-energy-recovery device, such HVAC Systems, Trane applications as a total-energy wheel (Figure 20), revolves through the parallel outdoor- and engineering manual SYS-APM003-EN, exhaust-air streams, preconditioning the outdoor air and reducing the capacity for more information about total-energy required from the cooling and heating coils. During the cooling season, the wheels and other types of air-to-air desiccant-coated wheel removes both sensible heat and moisture from the energy recovery. ■ outdoor air and rejects it to the exhaust air. During the heating season, the sensible heat and moisture that the wheel collects from the exhaust air preheats and prehumidifies the entering outdoor air. Transferring energy between the air streams provides two benefits: downsized equipment for cooling, heating, and humidification; and reduced operating costs.

Figure 20. Constant-volume system with total-energy recovery

Chilled water applications. Figure 21 (p. 28) illustrates the effect of adding a total-energy wheel, which has an effectiveness rating of 70 percent, to the basic chilled water, constant-volume system in previous examples. At the peak dry-bulb condition, the total-energy wheel preconditions the outdoor air to 81°F DB, 66.8°F WB (27.2°C DB, 19.3°C WB). The resulting mixed-air condition reduces the total load on the cooling coil from 4.66 tons (16.4 kW) to 3.5 tons

2 H. Henderson. “The Impact of Part-Load Air-Conditioner Operation on Dehumidification Performance: Validating a Latent Capacity Degradation Model,” Conference Proceedings from IAQ & Energy 1998: Using ASHRAE Standards 62 and 90.1, (Atlanta, GA: American Society of Heating, Refrigerating, and Air-Conditioning Engineers, Inc., 1998).

SYS-APM004-EN 27 Dehumidifying with Constant-Volume Mixed Air

Figure 21. Dehumidification performance of a constant-volume system with total-energy recovery

Design Full load Part load condition Peak dry bulb Peak dew point Mild, rainy OA 96.0°F DB, 76.0°F WB 76.0°F DP, 84.0°F DB 70.0°F DB, 69.0°F WB OA' 81.0°F DB, 66.8°F WB 77.0°F DB, 70.0°F WB 73.0°F DB, 67.5°F WB RA 74.0°F DB, 50.4% RH 74.0°F DB, 64.6% RH 74.0°F DB, 70.0% RH SA 55.7°F DB 63.0°F DB 66.5°F DB

(12.3 kW); however, the wheel does not affect the cooling loads in the space, so the supply-air temperature and supply airflow remain unchanged. The resulting relative humidity in the classroom drops from 52 percent to 50 percent at full load.

With a smaller sensible load in the space at the peak dew-point condition, the thermostat reduces coil capacity from 3.66 tons to 2.47 tons (from 12.9 kW to 8.7 kW) to deliver 63°F (17.2°C) supply air. The resulting relative humidity is 65 percent, compared to 67 percent without the wheel. On a mild, rainy day, the resulting relative humidity is 70 percent, compared to 73 percent without the wheel.

Preconditioning the outdoor air with a total-energy wheel significantly reduces mechanical cooling requirements at both full-load and part-load conditions, but it does little to lower indoor humidity.

Even if the total-energy wheel could be 100 percent effective, when it is humid outside, the air passing through the supply side of the wheel would be only as dry as—but never drier than—the air passing through the exhaust side of the wheel. Because the exhaust air stream originates in the space, the air leaving the supply side of the wheel will not be drier than the space.

Packaged DX applications. Adding a total-energy wheel to a constant- volume, packaged DX air conditioner reduces the cooling capacity required from the mechanical cooling system and, therefore, increases the required cfm/ ton (m3/s/kW). Consequently, the wheel may make it possible to select a smaller DX unit with an airflow that more closely matches space requirements (within the constraints of the flow-to-capacity ratio) in lieu of a larger unit with too much airflow.

28 SYS-APM004-EN Dehumidifying with Constant-Volume Mixed Air

At the peak dry-bulb condition in our example, preconditioning the outdoor air with a 70 percent-effective, total-energy wheel reduces the coil load from 4.66 tons (16.4 kW) to 3.5 tons (12.3 kW). Based on the desired supply airflow of 1,500 cfm (0.7 m³/s), the new airflow-to-capacity ratio of 428 cfm/ton (0.057 m³/s/kW) now falls within the design operating range for packaged DX equipment. Basing the selection on 1,500 cfm (0.7 m³/s) rather than 1,750 cfm (0.83 m³/s) lowers the supply-air temperature from 58.3°F (14.6°C) to 55.7°F (13.1°C) and provides more dehumidification. The resulting relative humidity in the classroom is 50 percent, rather than the 56 percent that resulted from the oversized packaged DX unit without a wheel. At the part-load, peak dew-point condition, the wheel results in a relative humidity of 65 percent, compared to 69 percent without the wheel. On a mild and rainy day, the resulting space relative humidity is 70 percent, compared to 74 percent without the wheel.

So, the total-energy wheel improves the dehumidification performance of constant-volume, packaged DX equipment by avoiding oversizing the airflow, but it does not eliminate the problem of high indoor humidity.

Cold Supply Air

Lowering the leaving-air temperature of the cooling coil removes more moisture from the air and requires less airflow to offset the sensible-cooling Peak dry-bulb condition: load in the space. At the peak dry-bulb condition, delivering supply air to the × × ()° Qs = 1.085 Vsa 74– 50 F classroom at 50°F (10°C) rather than 55.7°F (13.1°C) reduces the required , ∴ , supply airflow from 1,500 cfm (0.7 m³/s) to 1,142 cfm (0.54 m³/s). Supplying = 29 750 Btu/hr Vsa = 1 142 cfm colder, drier air also reduces the relative humidity from 52 percent to 47 percent ( × × []° ) Qs = 1.21 Vsa 23.3– 10 C (Figure 22, p. 30). The total cooling capacity required from the coil increases to ()∴ 5.1 tons (17.9 kW). = 8.7 kW Vsa = 0.54 m³/s The classroom requires warmer supply air at the peak dew-point condition Peak dew-point condition: because the sensible-cooling load is less. But the supply air is still significantly × , × ()° Qs = 1.085 1 142 cfm 74 FT– supply cooler—59.6°F (15.3°C) versus 63°F (17.2°C)—than when the system delivers a , ∴ ° higher supply airflow. The resulting relative humidity also improves slightly = 17 850 Btu/hr Tsupply = 59.6 F (62 percent versus 67 percent), and the required cooling capacity is 3.8 tons ()× × []° Qs = 1.21 0.54 m³/s 23.3 CT– supply (13.4 kW). For a mild, rainy day, the supply-air temperature increases to 64.1°F ()∴ ° (17.8°C) and requires a cooling capacity of 1.8 tons (6.3 kW). The resulting = 5.2 kW Tsupply = 15.3 C relative humidity in the classroom reaches 69 percent.

Mild, rainy day: Delivering “cold” supply air can improve the dehumidification performance × , × ()° Qs = 1.085 1 142 cfm 74 FT– supply of a constant-volume system, but it will not solve the problem of high indoor , ∴ ° humidity. = 12 250 Btu/hr Tsupply = 64.1 F

()× × []° Note: Applied chilled water systems typically work best for “cold air” Qs = 1.21 0.54 m³/s 23.3 CT– supply distribution because the designer can match the design requirements for ()= 3.6 kW∴ T = 17.8°C supply airflow and cooling capacity. Packaged DX systems defer many design

SYS-APM004-EN 29 Dehumidifying with Constant-Volume Mixed Air

Figure 22. Dehumidification performance of a constant-volume system that delivers “cold” supply air

Design Full load Part load condition Peak dry bulb Peak dew point Mild, rainy OA 96.0°F DB, 76.0°F WB 76.0°F DP, 84.0°F DB 70.0°F DB, 69.0°F WB RA 74.0°F DB, 47.0% RH 74.0°F DB, 62.0% RH 74.0°F DB, 69.0% RH MA 82.7°F DB 77.9°F DB 72.4°F DB SA 50.0°F DB 59.6°F DB 64.1°F DB

decisions to the manufacturer, which reduces initial cost; however, the limitations of a fixed design may make it difficult to achieve the desired “cold coil” temperature.

Humidity Control during Unoccupied Periods

Constant-volume systems that base control solely on the space dry-bulb temperature (indirect dehumidification) are unlikely to remove much moisture at night or during periods of low occupancy when space sensible loads are likely to be very low. Constant-volume systems that include a means to directly control space humidity—“dual-path” air handlers (p. 47) or supply-air tempering (p. 50), for example—will require an after-hours source of reheat energy. Dedicated outdoor-air systems, which are discussed in a later chapter (pp. 75–99), may be best suited to address dehumidification needs during unoccupied periods.

Building Pressurization

Maintaining an appropriate indoor–outdoor pressure difference is generally Consult Building Pressurization straightforward during mechanical cooling operation because the airflows in a Control, Trane applications constant-volume system do not change. Most problems occur at night (when engineering manual AM-CON-17, for exhaust systems are left on, but the ventilation system is off) or during information about regulating building economizer cooling. pressure through design and control of the HVAC system. ■ Systems with airside may require some method of building- pressure control to avoid overpressurization.

30 SYS-APM004-EN Dehumidifying with Constant-Volume Mixed Air

Airside Economizing

In constant-volume systems, the economizer cycle (p. 18) is often controlled by monitoring the outdoor-air dry-bulb temperature and comparing it with a fixed, predetermined limit. The following example illustrates the potential humidity problems associated with this control strategy.

Suppose that the outdoor-air condition for our Jacksonville school is 65ºF DB, 64ºF WB (18.3ºC DB, 17.8ºC WB). The constant-volume system responds to the less-than-design, sensible-cooling load in the space by reducing cooling Figure 23. Effect of airside economizer capacity, which raises the supply-air temperature to 68ºF (20ºC). If the control on space humidity economizer setpoint is 65ºF (18.3ºC), the outdoor- and return-air dampers modulate to mix 1,000 cfm (0.47 m3/s) of outdoor air with recirculated return air to maintain the space temperature at setpoint. Although this method of economizer control allows the cooling coil to shut off, the high moisture content of the outdoor air increases the indoor humidity to 75 percent (Figure 23).

Application considerations ■ When using the “fixed dry bulb” method of economizer control, pick a limit that is low enough to avoid bringing moisture-laden outdoor air indoors. ■ When designing a constant-volume system that requires an economizer to comply with Standard 90.1, investigate “fixed enthalpy” and “electronic enthalpy” control (which are allowed by the standard). Alternatively, consider selecting cooling equipment with an efficiency that is high enough to exempt the system from the economizer requirement. ■ To determine the appropriate economizer control, consider the climate, hours of occupancy, and potential operating-cost savings. ■ In most climates, avoid using a “differential (comparative) enthalpy” strategy to control the economizer in a constant-volume system. If you do opt to use this strategy, install a humidity sensor in the space to disable the economizer whenever the indoor relative humidity exceeds 60 percent.

Implications of ANSI/ASHRAE/IESNA Standard 90.1. Standard 90.1–2001 (Section 6.3.1) contains requirements for economizers in HVAC systems, including when they are required and how they should be controlled. When the cooling capacity of the constant-volume air handler is less than either 65,000 Btu/hr (19 kW) or 135,000 Btu/hr (38 kW), depending on the climate, an economizer is not required. If you choose to use one anyway, requirements related to the control of that economizer no longer apply (because the standard did not require the economizer in the first place). Although compliance with Section 6.3.1 should minimize energy use, it may not acceptably control indoor humidity at all operating conditions in all climates.

Section 6.3.1 defines high-limit-shutoff requirements for airside economizers. These requirements are based on climate and control method (fixed dry bulb, differential enthalpy, and so on). “Fixed dry bulb” control of economizers is

SYS-APM004-EN 31 Dehumidifying with Constant-Volume Mixed Air

allowed in any climate, with this stipulation: When used in a humid climate, the economizer can only operate when the outdoor dry-bulb temperature is less than or equal to 65°F (18.3°C). (The preceding example shows what can happen when the “fixed dry bulb” economizer setpoint is too high.) Of course, the effect of this control method for a particular installation will depend on the number of hours that the system operates in the “economizer” mode.

Improving Coincidental Dehumidification

The design of a basic constant-volume HVAC system can be altered to improve coincidental dehumidification performance without directly controlling space humidity. Table 3 compares the effect of these modifications (described on pp. 32–43), when applied to the example classroom in Jacksonville, Florida.

Table 3. Comparison of coincidental (“indirect”) dehumidification1 Resulting relative humidity, % Design of constant-volume system Peak dry bulb Peak dew point Mild, rainy day Basic (unaltered) 52 67 73 Basic plus… adjustable fan speed (p. 32) 52 60 68 mixed-air bypass (p. 34) 52 65 68 mixed-air bypass and adjustable 52 58 65 fan speed (p. 36) return-air bypass with full 52 55 60 coil face at part load (p. 37) return-air bypass with reduced 52 64 66 coil face at part load (p. 39)

1 Comparison of dehumidification performance is based on a classroom in Jacksonville, Florida. See “Analysis of Dehumidification Performance” (p. 19) for a description of the room and the constant-volume HVAC system serving it.

Figure 24. Classroom unit ventilator Adjustable Fan Speed

Many room terminals, such as fan–coils and classroom unit ventilators (Figure 24), include fans that can run at different speeds. Depending on the equipment, fan speed is controlled manually by a switch or automatically by a unit controller.

Slowing the fan speed improves the coincidental dehumidification provided by constant-volume room terminals; it is also the first step to reduce cooling capacity. Let’s use the example classroom to demonstrate the effect of less airflow. Assume that the HVAC system provides 1,500 cfm (0.7 m³/s) of supply airflow when the fan operates at its highest speed. As the sensible-cooling load in the space decreases (Figure 25), the system initially responds by switching to

32 SYS-APM004-EN Dehumidifying with Constant-Volume Mixed Air

Figure 25. Example of automatic low-speed fan operation, which reduces the supply airflow to 1,025 cfm (0.48 fan-speed adjustment m³/s). As the space load decreases further, the control valve modulates the chilled water flow through the coil to appropriately reduce cooling capacity.

Because the fan operates at high speed when the peak dry-bulb condition exists, dehumidification performance matches that of the basic constant- volume system described at the beginning of our analysis (p. 20). That is, the system supplies 55.7°F (13.1°C) air to satisfy the sensible-cooling load in the space and maintain the 74°F DB (23.3°C) target; the resulting humidity is 52 percent.

Peak dew-point condition: At the part-load, peak dew-point condition (Figure 26), the reduced supply × , × ()° Qs = 1.085 1 025 cfm 74 FT– supply airflow results in a lower supply-air temperature, 57.9°F (14.3°C) versus 63°F , ∴ ° (17.2°C). Reducing the airflow allows the coil to remove more moisture, = 17 850 Btu/hr Tsupply = 57.9 F improving the dehumidification performance of the system. At this condition, ()× × []° Qs = 1.21 0.48 m³/s 23.3 CT– supply the relative humidity in the space improves from 67 percent to 60 percent; the ()∴ ° required cooling capacity is 3.9 tons (13.7 kW). = 5.2 kW Tsupply = 14.3 C On the mild and rainy day, the supply-air temperature rises to 63°F (17.2°C) and Mild, rainy day: the resulting space relative humidity climbs to 68 percent. The required cooling × , × ()° Qs = 1.085 1 025 cfm 74 FT– supply capacity is 1.7 tons (6.0 kW). = 12,250 Btu/hr∴ T = 63.0°F supply Application considerations ()× × []° ■ Adjusting the fan speed offers an important acoustical benefit, particularly Qs = 1.21 0.48 m³/s 23.3 CT– supply for room terminals located within the occupied space: Fans operate quieter ()= 3.6 kW∴ T = 17.2°C supply at low speed. ■ Unit controls should automatically adjust the position of the outdoor-air damper whenever the fan speed changes, thereby assuring that the space continues to receive the proper amount of outdoor air. As part of the system

Figure 26. Constant-volume dehumidification performance at low fan speed

Design Part load condition Peak dew point Mild, rainy OA 76.0°F DP, 70.0°F DB, 84.0°F DB 69.0°F WB RA 74.0°F DB, 74.0°F DB, 60.0% RH 68.0% RH SA 57.9°F DB 63.0°F DB

SYS-APM004-EN 33 Dehumidifying with Constant-Volume Mixed Air

air-balancing procedure and with the exhaust fan operating, determine the appropriate damper position for each fan speed.

Mixed-Air Bypass

Face-and-bypass dampers, arranged to allow mixed air to bypass the cooling coil, are often used to improve the indirect dehumidification performance of a constant-volume system. Simple and inexpensive, mixed-air bypass blends cold, dry air leaving the cooling coil with bypassed mixed air. The space thermostat controls cooling capacity by adjusting the positions of the linked face-and-bypass dampers, regulating airflow through and around the coil to achieve the proper supply-air temperature (Figure 27); chilled water flow through the coil remains constant. This control method is sometimes described as letting the cooling coil “run wild.”

At the peak dry-bulb condition, the face damper is wide open and the bypass damper is closed. All of the mixed air passes through the cooling coil, so dehumidification performance is identical to that of the basic constant-volume system without mixed-air bypass (p. 20).

At the part-load, peak dew-point condition, the face damper modulates closed and the linked bypass damper modulates open to satisfy the space-thermostat setpoint. The entering water temperature and water-flow rate through the coil are unchanged. Diverting some of the mixed air around the coil slows the velocity of the air passing through the coil; more of the entrained moisture condenses, so the conditioned air (CA) leaves the coil drier and colder—that is, at 52°F (11.1°C) for our example classroom in Jacksonville (Figure 28), as determined with the help of a coil-performance program. The conditioned air then blends with the bypassed, mixed air to achieve the desired supply-air temperature of 63°F (17.2°C).

Figure 27. Basic constant-volume HVAC system with mixed-air bypass

34 SYS-APM004-EN Dehumidifying with Constant-Volume Mixed Air

Figure 28. Constant-volume dehumidification performance with mixed-air bypass

Design Part load condition Peak dew point Mild, rainy OA 76.0°F DP, 70.0°F DB, 84.0°F DB 69.0°F WB RA 74.0°F DB, 74.0°F DB, 64.5% RH 68.0% RH CA 52.0°F DB 46.1°F DB SA 63.0°F DB 66.5°F DB

Do not assume that the cold, dry air leaving the coil will adequately dehumidify the space. At the peak dew-point condition, moisture in the bypassed air prevents more than a slight decrease in relative humidity—from 67 percent to 65 percent, in this case. The total coil load increases from 3.68 tons (12.9 kW) to 3.98 tons (14.0 kW).

On the mild and rainy day, the air leaves the cooling coil at 46.1°F (7.8°C) and blends with the bypassed, mixed air to achieve the required supply-air temperature of 66.5°F (19.2°C). The resulting relative humidity rises to 68 percent, and the required cooling capacity is 1.8 tons (6.3 kW). Again, mixed- air bypass improves the indirect (coincidental) dehumidification of this basic constant-volume system … but only slightly.

Application considerations ■ Resetting the chilled water temperature or varying the water flow through the cooling coil will degrade the dehumidification performance of a constant-volume system equipped with mixed-air bypass. ■ Mixed-air bypass works with either blow-through or draw-through supply fans. It also requires less space than true (full coil face active at part load) return-air bypass, making it advantageous for terminal-unit applications. ■ Mixed-air bypass dehumidifies more effectively than return-air bypass in arid climates, where outdoor air is usually drier than return air. For climates where return air is usually drier than outdoor air, however, return-air bypass (p. 37) dehumidifies more effectively than mixed-air bypass. ■ To prevent water vapor from condensing on the interior and/or exterior of the air-handler casing, close the chilled-water control valve when the fan is off, or when the face damper is nearly closed.

SYS-APM004-EN 35 Dehumidifying with Constant-Volume Mixed Air

Mixed-air bypass plus adjustable fan speed. Indirect dehumidification provided by a constant-volume system with mixed-air bypass improves when combined with low-speed fan operation as the first step of capacity control. Recall from an earlier example (p. 32) that the system first responds to a decrease in the space sensible-cooling load by switching the fan to low speed, which reduces supply airflow to 1,025 cfm (0.48 m³/s).

Peak dew-point condition: When the space load decreases further, the face-and-bypass dampers vary the × , × ()° Qs = 1.085 1 025 cfm 74 FT– supply cooling capacity accordingly by diverting some of the mixed air around the coil. , ∴ ° At the part-load, peak dew-point condition (Figure 29), less airflow lowers the = 17 850 Btu/hr Tsupply = 57.9 F supply-air temperature from 63°F (17.2°C) to 57.9°F (14.3°C). Reduced supply ( × × []° ) Qs = 1.21 0.48 m³/s 23.3 CT– supply airflow also means that more mixed air passes through the coil before blending ()∴ ° with the bypassed mixed air. The air that passes through the cooling coil is = 5.2 kW Tsupply = 14.3 C cooled to 51.5°F DB (10.8°C DB). At this condition, the relative humidity in the space improves to 58 percent, compared to 65 percent with mixed-air bypass Mild, rainy day: only. The required cooling capacity is 4.1 tons (14.4 kW). × , × ()° Qs = 1.085 1 025 cfm 74 FT– supply On the mild and rainy day, the air leaves the cooling coil at 46.1°F (7.8°C) and = 12,250 Btu/hr∴ T = 63.0°F supply blends with the bypassed, mixed air to achieve the required supply-air ( × × []° ) temperature of 63°F (17.2°C). The resulting relative humidity is 65 percent, Qs = 1.21 0.48 m³/s 23.3 CT– supply compared to 68 percent with mixed-air bypass only. The required cooling ()= 3.6 kW∴ T = 17.2°C supply capacity is 1.8 tons (6.3 kW).

Application considerations ■ Adjusting the fan speed offers an important acoustical benefit, particularly for in-room terminals: Fans operate quieter at low speed. ■ Unit controls should automatically adjust the position of the outdoor-air damper whenever the fan speed changes, thereby assuring that the space continues to receive the proper amount of outdoor air. As part of the system

Figure 29. Constant-volume dehumidification performance with mixed-air bypass and low-speed fan operation

Design Part load condition Peak dew point Mild, rainy OA 76.0°F DP, 70.0°F DB, 84.0°F DB 69.0°F WB RA 74.0°F DB, 74.0°F DB, 58.0% RH 65.0% RH CA 51.5°F DB 46.1°F DB SA 57.9°F DB 63.0°F DB

36 SYS-APM004-EN Dehumidifying with Constant-Volume Mixed Air

air-balancing procedure and with the exhaust fan operating, determine the appropriate damper position for each fan speed.

Return-Air Bypass

Full coil face area. Face-and-bypass dampers also can be arranged to bypass only return air (Figure 30) instead of a mixture of outdoor and return air. Return-air bypass requires additional space and ductwork, which may increase initial cost; however, it also limits humidity better than other methods of coincidental dehumidification, making return-air bypass a cost-effective enhancement for constant-volume systems.

As with mixed-air bypass, the space thermostat controls cooling capacity by adjusting the position of the linked face-and-bypass dampers, regulating airflow through and around the cooling coil. Because the coil “runs wild,” its surface can be very cold at part-load conditions, enhancing the system’s ability to dehumidify the space without directly controlling humidity. What makes return-air bypass more effective than mixed-air bypass at most loads, however, is that it directs all of the moist outdoor air through the cooling coil. The cold, dry air leaving the cooling coil then blends with relatively dry return air (rather than a mixture of humid outdoor and return air).

At the peak dry-bulb condition, the face damper is wide open and the bypass damper is closed; therefore, all of the mixed air passes through the cooling coil. At this condition, dehumidification performance is identical to that of the basic constant-volume system in our example Jacksonville classroom (p. 20). As the sensible-cooling load in the space decreases, the face damper modulates toward closed and the linked bypass damper modulates open. Some of the recirculated return air bypasses the cooling coil, while the rest mixes with the outdoor air and passes through the cooling coil. The bypassed return air and conditioned air mix downstream of the coil (Figure 31, p. 38). The entering water temperature and water flow rate through the coil remain unchanged, so

Figure 30. Basic constant-volume HVAC system with return-air bypass

SYS-APM004-EN 37 Dehumidifying with Constant-Volume Mixed Air

Figure 31. Effect of return-air bypass on part-load airflows (peak dew-point condition)

with less airflow through the coil, the air leaving the coil is colder than at the peak dry-bulb condition. It is difficult to accurately model the part-load performance of an HVAC At the part-load, peak dew-point condition (Figure 31, Figure 32), all 450 cfm system with return-air bypass. Both the (0.21 m³/s) of the outdoor air brought into the classroom for ventilation passes airflow through the coil and the through the cooling coil, as does 300 cfm (0.14 m³/s) of the recirculated return temperature of the mixed air as it air; the resulting coil load is 4.2 tons (14.7 kW). The rest of the recirculated enters the coil change as the face-and- return air, 750 cfm (0.35 m³/s), bypasses the coil and mixes with the 51.8°F bypass dampers modulate. Analyzing (11.0°C) conditioned air leaving the coil. When the blended, 63°F (17.2°C) supply part-load performance is a trial-and- air reaches the classroom, the resulting relative humidity is 55 percent. error process that is best accomplished using coil modeling software and a On the mild and rainy day (Figure 32, Figure 33), only 400 cfm (0.19 m³/s) of the psychrometric chart. ■ outdoor air passes through the cooling coil, which results in a cooling load of 2.2 tons (7.7 kW). The remaining 50 cfm (0.02 m³/s) of outdoor air mixes with

Figure 32. Constant-volume dehumidification performance with return-air bypass and entire coil surface

Design Part load condition Peak dew point Mild, rainy OA 76.0°F DP, 70.0°F DB, 84.0°F DB 69.0°F WB RA 74.0°F DB, 74.0°F DB, 55.2% RH 60.0% RH CA 51.8°F DB 46.7°F DB SA 63.0°F DB 66.5°F DB

38 SYS-APM004-EN Dehumidifying with Constant-Volume Mixed Air

Figure 33. Effect of return-air bypass on part-load airflows (mild, rainy condition)

1,050 cfm (0.5 m³/s) of recirculated return air and bypasses the coil. In the classroom, the blended 66.5°F (19.2°C) supply air results in a relative humidity of 60 percent.

Reduced coil face area. There are various ways to implement return-air bypass in HVAC equipment. Because of the limited space within terminals, such as unit ventilators, operating the face-and-bypass dampers often reduces the usable face area of the coil. In other words, as the face damper opens (in response to a diminishing sensible-cooling load), it actually blocks part of the cooling-coil surface (Figure 34). Less dehumidification occurs because less of the coil is available for cooling and because the velocity of the air passing through the coil is essentially unchanged.

Figure 34. Operation of face-and-bypass dampers that reduce available coil surface

SYS-APM004-EN 39 Dehumidifying with Constant-Volume Mixed Air

Figure 35. Constant-volume dehumidification performance with return-air bypass and reduced coil surface

Design Part load condition Peak dew point Mild, rainy OA 76.0°F DP, 70.0°F DB, 84.0°F DB 69.0°F WB RA 74.0°F DB, 74.0°F DB, 64.0% RH 66.0% RH CA 59.6°F DB 59.8°F DB SA 63.0°F DB 66.5°F DB

Our example classroom helps demonstrate the effect of less coil surface during return-air-bypass operation (Figure 35). If the face area of the coil decreases as the face damper closes, the relative humidity in the classroom climbs to 64 percent at the part-load, peak dew-point condition, as compared to 55 percent when the entire coil face is available. The required cooling capacity is 3.9 tons (13.7 kW). Mild and rainy conditions require 1.9 tons (6.7 kW) of cooling capacity, and the resulting relative humidity increases to 66 percent.

Application considerations ■ Resetting the chilled water temperature or varying the water flow through the cooling coil degrades the dehumidification performance of a constant- volume system equipped with return-air bypass.

Figure 36. Examples of air handlers with ■ Return-air bypass requires a draw-through supply-fan arrangement and properly configured return-air bypass more space, typically, than mixed-air bypass. For proper dehumidification, the air-handler configuration must mix the bypassed return air with conditioned air downstream of the cooling coil (Figure 36). ■ Because it usually directs all incoming outdoor air through the cooling coil, return-air bypass dehumidifies better than mixed-air bypass when the outdoor air contains more moisture than the return air. For arid climates where outdoor air is usually drier than return air, however, mixed-air bypass (p. 34) works best. ■ At very low loads, the airflow through the coil may be less than the outdoor airflow. In this case, the remaining portion of the outdoor air mixes with the recirculated return air to bypass the cooling coil. Even though some of the outdoor air bypasses the cooling coil, the percentage of outdoor air in the bypass path generally is very low when compared to mixed-air bypass.

40 SYS-APM004-EN Dehumidifying with Constant-Volume Mixed Air

■ When considering equipment that uses return-air bypass, determine whether the face of the cooling coil is blocked at part-load conditions. Reducing the available face area significantly impairs dehumidification performance. ■ To prevent water vapor from condensing on the interior and/or exterior of the air-handler casing, close the chilled-water control valve when the fan is off, or when the face damper is nearly closed.

DX Coil Circuiting

Figure 37. Common DX coil configurations When more than one liquid-refrigerant distributor serves a direct-expansion (DX) cooling coil, the coil is divided into sections (Figure 37). Each section is independently controlled by its own expansion device. The most common configurations for divided, finned-tube evaporator coils are:

■ Face-split, which divides the coil into parallel sections. Both sections are active when the cooling load is high, providing a uniform leaving-air temperature. At low loads, only one coil section is active to cool and dehumidify the air passing through it. Air passing through the inactive (top) section remains unconditioned. The two air streams mix downstream of the coil, producing average temperature and humidity conditions. ■ Intertwined, which divides the coil by feeding alternate tubes of the coil via two distributors. At high loads, both distributors feed liquid refrigerant to all of the tubes. At low loads, only one distributor operates, and the coil performs as though its fin surface area were substantially greater. Therefore, the coil surface can be warmer at part load (reducing the risk of frost), and still provide a uniform leaving-air temperature. This performance characteristic makes intertwined coils well-suited for VAV applications.

Which of these DX coil-circuiting arrangements, face-split or intertwined, provides the best constant-volume dehumidification? To find out, we used a computer model to simulate performance at the part-load, peak dew-point condition for our example Jacksonville classroom. The HVAC system in this analysis contained two equally sized compressors and a coil served by two solenoid valves. Whenever the thermostat turned off a compressor, the corresponding solenoid valve stopped refrigerant flow through half of the coil. Table 4 compares the results of the analysis.

At the peak dew-point condition, one compressor operates continuously while the other cycles on and off. With both compressors operating and the entire coil surface active, the leaving-air condition—58.7°F DB, 58.4°F DP (14.8°C DB, 14.7°C DP)—is identical for both coil circuiting arrangements. Offsetting the space sensible-cooling load and maintaining the thermostat setpoint, however, requires 63°F DB (17.2°C DB) supply air. Therefore, operating both compressors

SYS-APM004-EN 41 Dehumidifying with Constant-Volume Mixed Air

eventually overcools the space. The thermostat then stops the second compressor, and one of the solenoid valves prevents refrigerant flow through half of the coil.

Table 4. Performance comparison of DX coil-circuiting options at peak dew point1 Operating condition Face-split coil Intertwined coil Coil leaving-air temperature with: 58.7°F (14.8°C) DB, 58.7°F (14.8°C) DB, ■ Both compressors operating 58.4°F (14.7°C) DP 58.4°F (14.7°C) DP ■ Full coil surface active Coil leaving-air temperature active coil section only 57.8°F (14.3°C) DB, not applicable with: 57.5°F (14.2°C) DP ■ One compressor operating mixed air downstream 67.4°F (19.7°C) DB, 64.0°F (17.8°C) DB, ■ Half of coil surface active of coil 62.7°F (17.1°C) DP 63.5°F (17.5°C) DP Averaged supply-air condition to 63.0°F (17.2°C) DB, 63.0°F (17.2°C) DB, match sensible-cooling load 60.6°F (15.9°C) DP 62.6°F (17.0°C) DP Resulting condition in space 74.0°F (23.3°C) DB, 74.0°F (23.3°C) DB, 68% RH 71% RH

1 Coil performance is based on the peak dew-point condition for Jacksonville, Florida: 76°F DP, 84°F DB (24.6°C DP, 28.8°C DB).

Face-split coil. With only one compressor operating, air passing through the active (bottom) half of the face-split coil is cooled and dehumidified to 57.8°F DB, 57.5°F DP (14.3°C DB, 14.2°C DP). When this conditioned air mixes with the unconditioned air that passed through the inactive (top) half of the coil, the resulting mixed air downstream of the coil is 67.4°F DB, 62.7°F DP (19.7°C DB, 17.1°C DP). The supply air eventually undercools the space because it is too warm to offset the sensible-cooling load. At this point, the thermostat restarts the second compressor and the entire coil surface is active again.

With the second compressor cycling on and off, the average supply-air condition during an hour of HVAC operation is 63.0°F DB, 60.6°F DP (17.2°C DB, 15.9°C DP).

Intertwined coil. Unlike face-split circuiting, every other row in an intertwined coil still receives refrigerant after the second compressor stops. Air passing through the coil is cooled and dehumidified to 64.0°F DB, 63.5°F DP (17.8°C DB, 17.5°C DP). Because the supply air is too warm to offset the sensible- cooling load, the second compressor cycles as necessary to maintain the thermostat setpoint. However, because the supply-air condition is close to the desired condition of 63°F (17.2°C), the system with the intertwined coil operates the second compressor much less throughout the hour than the system with the face-split coil.

42 SYS-APM004-EN Dehumidifying with Constant-Volume Mixed Air

With the second compressor cycling on and off, the average supply-air condition for an hour of operation yields 63.0°F DB, 62.6°F DP (17.2°C DB, 17.0°C DP).

Conclusion. At the peak dew-point condition, with one compressor operating continuously and the other cycling on and off, the constant-volume performance of face-split and intertwined coils is similar. The average supply- air dew point for the face-split coil is only 2°F (1.1°C) less than that of the intertwined coil, and it results in a space relative humidity that is only 3 percent less. To represent a broader range of loads, performance was also compared for mild, rainy conditions (Table 5) and yielded similar results. Again, the face-split coil provided a slightly lower space relative humidity, that is, 71 percent versus 73 percent for the intertwined coil.

None of the performance models accounted for the transient effect of moisture re-evaporating from the inactive half of the face-split coil; so, the already small difference in performance is likely to be even less.

Application consideration ■ For constant-volume DX applications in which humidity control is important, use supply-air tempering (pp. 50–52), or separately condition the incoming outdoor air (pp. 44–50), to improve dehumidification performance.

Table 5. Performance comparison of DX coil-circuiting options on a mild, rainy day1 Operating condition Face-split coil Intertwined coil Coil leaving-air temperature active coil section only 55.0°F (12.8°C) DB, not applicable with: 54.8°F (12.7°C) DP ■ One compressor operating mixed air downstream 63.9°F (17.7°C) DB, 60.5°F (15.8°C) DB, ■ Half of coil surface active of coil 61.0°F (16.1°C) DP 60.3°F (15.7°C) DP Coil entering- and leaving-air 72.8°F (22.7°C) DB, 72.8°F (22.7°C) DB, temperatures with: 66.1°F (18.9°C) DP 66.1°F (18.9°C) DP ■ No compressors operating ■ Entire coil surface inactive Averaged supply-air condition to 66.5°F (19.2°C) DB, 66.5°F (19.2°C) DB, match sensible-cooling load 62.5°F (16.9°C) DP 63.3°F (17.4°C) DP Resulting condition in space 74.0°F (23.3°C) DB, 74.0°F (23.3°C) DB, 71% RH 73% RH

1 Coil performance is based on a mild, rainy condition for Jacksonville, Florida: 70°F DB, 69°F WB (21.2°C DB, 20.6°C WB).

SYS-APM004-EN 43 Dehumidifying with Constant-Volume Mixed Air

“Direct” Control of Humidity

Coincidental (indirect) dehumidification enhancements may work well for comfort-cooling applications in certain indoor environments and in certain climates. When latent loads and sensible loads vary significantly, however, or when it is necessary to maintain a low humidity in the occupied space, it may be necessary to directly control both dry-bulb temperature and humidity. This is usually accomplished either by separately conditioning the outdoor air and return air or by overcooling and tempering the supply air.

Note: For the analyses discussed in this section, we assumed that the HVAC system directly controls the indoor relative humidity, keeping it below 60 percent.

Separate Air Paths

Providing separate treatment paths for the outdoor air entering the building and for return air from the space can enable direct humidity control. These paths may reside in individual air handlers, as in a “dedicated outdoor-air system,” or within the same air-handler casing, as in a “dual-path air handler.”

Dedicated outdoor-air system

As its name implies, a dedicated outdoor-air (OA) system devotes one air Dedicated outdoor-air systems can be handler exclusively to cooling and dehumidifying all outdoor air so that it is arranged in several ways. These drier than the air in the space. The conditioned outdoor air is then delivered configurations are discussed in more either directly to the space or to other air handlers. Common names for an air detail in “Dehumidifying with handler that serves this purpose include “dedicated outdoor-air unit,” Dedicated Outdoor Air,” pp. 75–99. ■ “100 percent-outdoor-air unit,” “fresh-air unit,” and “makeup-air unit.”

Figure 38 shows a dedicated outdoor-air system that delivers 450 cfm

Figure 38. Dedicated outdoor-air system (0.21 m³/s) of dry, neutral-temperature outdoor air directly to our example Jacksonville classroom. The dedicated OA unit dehumidifies the entering dedicated OA unit outdoor air to a low dew point and then reheats it to the approximate dry-bulb- temperature target for the space. In this example, the dedicated OA system dehumidifies the outdoor air to 52°F DP (11.1°C DP) and then reheats it to 71°F DB (21.7°C DB).

The dedicated OA unit modulates cooling-coil capacity to maintain the desired leaving dew-point temperature. This dew point is determined during the design process to assure that the conditioned outdoor air is dry enough to properly dehumidify the space at all load conditions. With the addition of a space humidity sensor, the leaving dew-point temperature can be reset in response to actual space conditions. Meanwhile, a fan–coil in the classroom cools 1,500 cfm (0.7 m³/s) of recirculated air from the space to offset the local cooling loads.

44 SYS-APM004-EN Dehumidifying with Constant-Volume Mixed Air

Figure 39. Dehumidification performance of a dedicated outdoor-air system Peak dry-bulb condition: at peak dry-bulb condition

Qs,fc = Qs,space – Qs,doa = 29,750 Btu/hr – []1.085× 450cfm × ()74– 71°F = 28,285 Btu/hr × , × () = 1.085 1 500 cfm 74°F– Tsupply ∴ Tsupply = 56.6°F

() Qs,fc = Qs,space – Qs,doa (= 8.7 kW – [])1.21× 0.21 m³/s × ()23.3– 21.7°C ()= 8.29 kW ()× × [] = 1.21 0.7 m³/s 23.3°C– Tsupply ()∴ Tsupply = 13.5°C

At the peak dry-bulb condition (Figure 39), the fan–coil cools 100 percent- recirculated air to 56.6°F DB (13.5°C DB); the resulting relative humidity is 50 percent. (This supply-air temperature is slightly warmer than in a basic constant-volume system because the 71°F DB [21.7°C DB] air, supplied by the dedicated outdoor-air unit, offsets part of the sensible-cooling load in the space.) Together, the dedicated outdoor-air unit and the fan–coil—which provide 3.0 tons (10.6 kW) and 2.8 tons (9.8 kW), respectively—offset the total cooling load of 5.8 tons (20.4 kW).

At the part-load, peak dew-point condition (Figure 40, p. 46), the dedicated outdoor-air unit delivers the outdoor air at the same conditions, 71°F DB (21.7°C DB) and 52°F DP (11.1°C DP). Fan–coil capacity modulates to match the lower sensible-cooling load in the space, which raises that unit’s supply-air temperature to 63.9°F (17.7°C) and results in a relative humidity of 56 percent. Again, the total cooling load of 4.8 tons (16.9 kW) is divided between the Appendix B in this manual describes dedicated outdoor-air unit and the fan–coil, which handle 3.4 tons (12.0 kW) the design process that we used to select and 1.4 tons (4.9 kW), respectively. the dedicated outdoor-air unit and the local terminal(s). For this example, we sized the dedicated OA unit to maintain the relative humidity below 60 percent in the classroom at full- and part-load conditions. ■

SYS-APM004-EN 45 Dehumidifying with Constant-Volume Mixed Air

Figure 40. Dehumidification performance of a dedicated outdoor-air system at peak dew-point condition Peak dew-point condition:

Qs,fc = Qs,space – Qs,doa = 17,850 Btu/hr – []1.085× 450cfm × ()74– 71°F = 16,385 Btu/hr × , × () = 1.085 1 500 cfm 74°F– Tsupply ∴ Tsupply = 63.9°F

() Qs,fc = Qs,space – Qs,doa (= 5.2 – []1.21× 0.21 m³/s × ()23.3– 21.7°C ) ()= 4.8 kW ()× × [] = 1.21 0.7 m³/s 23.3°C– Tsupply ()∴ Tsupply = 17.7°C

The sensible-cooling load in the space is even lower on the mild and rainy day, so the supply-air temperature from the fan–coil increases to 67.4°F (19.5°C), while the conditioned air from the dedicated outdoor-air unit remains unchanged. The resulting relative humidity in the space is 60 percent. Of the total cooling load, which is 2.9 tons (10.2 kW), the dedicated outdoor-air unit handles 2.0 tons (7.0 kW) and the fan–coil handles 0.9 tons (3.2 kW).

Figure 41. Dehumidification performance of a dedicated outdoor-air system on a mild, rainy day Mild, rainy day:

Qs,fc = Qs,space – Qs,doa = 12, 250 Btu/hr – []1.085× 450cfm × ()74– 71°F = 10,785 Btu/hr × , × () = 1.085 1 500 cfm 74°F– Tsupply ∴ Tsupply = 67.4°F

() Qs,fc = Qs,space – Qs,doa (= 3.6 kW – []1.21× 0.21 m³/s × ()23.3– 21.7°C ) ()= 3.2 kW ()× × [] = 1.21 0.7 m³/s 23.3°C– Tsupply ()∴ Tsupply = 19.5°C

46 SYS-APM004-EN Dehumidifying with Constant-Volume Mixed Air

“Dual-path” air handler

A dual-path air handler, which separately conditions both return air and outdoor air, offers an alternative to one dedicated outdoor-air unit for the system and a room terminal in each space. Each air path includes a dedicated cooling coil, but the same constant-volume fan serves both paths (Figure 42).

Working together, the cooling coils in the return-air (RA) and outdoor-air (OA) paths maintain the dry-bulb temperature and humidity in the space:

■ Arranging the air paths in a stacked The OA cooling coil prevents the humidity in the space from exceeding a configuration (below) reduces the predefined limit, dehumidifying the outdoor air enough to offset the latent footprint of the “dual-path” air handler. load. A humidity sensor in the space directly controls coil capacity. ■ The RA cooling coil provides the additional cooling needed to offset the sensible load. A thermostat in the space directly controls coil capacity to maintain the space dry-bulb temperature at setpoint.

At the example classroom’s peak dry-bulb condition (Figure 43, p. 48), each coil conditions the air to 55.7°F (13.1°C). When the combined airflows are supplied to the space, the resulting relative humidity is about the same as that achieved by the basic, single-coil HVAC system. The total load of 4.81 tons (16.9 kW) is split between the outdoor-air coil, which handles 2.15 tons (7.6 kW), and the return-air coil, which handles 2.66 tons (9.4 kW).

At the part-load, peak dew-point condition (Figure 44, p. 48), the humidity sensor modulates the capacity of the OA cooling coil to maintain the relative humidity at the desired upper limit—60 percent in this case. This is done by opening the chilled water valve and reducing the leaving-coil temperature to 52°F (11.1°C). The thermostat reduces the capacity of the RA cooling coil to match the diminished sensible load; the resulting temperature of the blended

Figure 42. Dual-path air handler

SYS-APM004-EN 47 Dehumidifying with Constant-Volume Mixed Air

Figure 43. Dehumidification performance of a dual-path air handler at peak dry-bulb condition

supply air is 63°F (17.2°C). Again, the total cooling load of 4.07 tons (14.3 kW) is divided between the two coils: the OA coil handles 3.44 tons (12.1 kW) and the RA coil handles 0.63 tons (2.2 kW). The system directly controls the relative humidity in the space to the desired upper limit of 60 percent.

Maintaining 60 percent-relative humidity when it is mild and rainy outside (Figure 45) requires a leaving-air temperature of 51°F (10.6°C) from the OA coil. To achieve a blended supply-air temperature of 66.5°F (19.2°C) and avoid overcooling the space, the RA coil cools the recirculated return air from 74°F (23.3°C) to 73.1°F (22.8°C). The OA coil load is 2.09 tons (7.4 kW), and the RA coil load is 0.09 tons (0.3 kW).

Figure 44. Dehumidification performance of a dual-path air handler at peak dew-point condition

48 SYS-APM004-EN Dehumidifying with Constant-Volume Mixed Air

Figure 45. Dehumidification performance of a dual-path air handler on a mild, rainy day

Application considerations Table 6. Summary of cooling-coil loads Table 6 summarizes the cooling-coil loads at the peak dry-bulb, peak dew-point, for example dual-path air handler1 and mild, rainy conditions. In this case, the highest load on the RA cooling coil Cooling load, tons (kW) occurs at the peak dry-bulb condition, while the highest load on the OA cooling OA coil RA coil Block coil occurs at the peak dew-point condition. However, neither of these Peak 2.15 2.66 4.81 conditions necessarily represents the worst-case combined load for these coils. DB (7.6) (9.4) (16.9) ■ Peak 3.44 0.63 4.07 Size each cooling coil for its individual peak load. DP (12.1) (2.2) (14.3) ■ Because the peak loads on the two coils occur at different times, size the Mild, 2.09 0.09 2.18 cooling equipment (a central chilled water plant, for example) based on the rainy (7.4) (0.3) (7.7) block load rather than the sum of the peak loads.

1 Outdoor conditions refer to the classroom examples set ■ If the risk of below-freezing outdoor temperatures exists during occupied in Jacksonville, Florida: Peak dry bulb 96°F (35.7°C) DB, 76°F (24.5°C) WB periods, consider protecting the chilled water coil in the outdoor air path by: Peak dew point 76°F (24.6°C) DP, 84°F (28.8°C) DB installing a preheat coil upstream of the chilled water coil; using an air-to-air, Mild, rainy 70°F (21.2°C) DB, 69°F (20.6°C) WB energy-recovery device to recover heat from the exhaust air; or adding glycol to the chilled water system. ■ When the sensible load is low enough, the RA cooling coil may actually turn off while the OA cooling coil continues to produce cold air. A heating coil is generally required to heat the recirculated return air and avoid overcooling the space. Section 6.3.2.3 of ASHRAE Standard 90.1–2001 (see p. 8 in this manual) restricts the mixing of hot and cold air. If the dual-path air handler adds heat to the return-air path at certain part-load conditions, and if it meets one of the following criteria, it may be exempt from this restriction: ❚ Per Exception B, the dual-path air handler’s design cooling capacity is 6.67 tons (23 kW) or less, and the combined load on both coils is less than 50 percent of design capacity when heat is added to the return-air path.

SYS-APM004-EN 49 Dehumidifying with Constant-Volume Mixed Air

❚ Per Exception D, the dual-path air handler serves a space that requires specific humidity levels to satisfy process needs; examples include computer rooms, museums, surgical suites, supermarkets, refrigerated warehouses, and ice arenas. ❚ Per Exception E, at least 75 percent of the heat added to the return-air path originates from an on-site source of recovered (condenser heat, for example) or solar energy.

■ For cold-weather climates, size the capacity of the RA heating coil to offset the heating loads in the space. ■ Size the OA cooling coil for the minimum outdoor airflow required for ventilation. Provide an additional outdoor air path if the system includes an airside economizer. ■ For proper system control, install a humidity sensor in the space and assure that the air-handler controls can independently modulate the capacity of each cooling coil.

Supply-Air Tempering

Using a single cooling coil in series with a source of heat for tempering (Figure 46) provides an alternative means of directly controlling indoor humidity. The cooling coil dehumidifies the air to a dew point that is dry enough to maintain the space at an acceptable humidity level. The downstream heating coil “tempers” (raises) the supply-air dry-bulb temperature just enough to maintain the thermostat setpoint and avoid overcooling the space.

Note: In this manual, we use the term “tempering” instead of “reheat” because the heating coil only moderates the cooling effect of the dry supply air. The space still requires cooling, but not as much as the dehumidification process provides.

Figure 46. Constant-volume HVAC system with supply-air tempering

50 SYS-APM004-EN Dehumidifying with Constant-Volume Mixed Air

Figure 47. Dehumidification performance of supply-air tempering at part-load conditions

At the peak dry-bulb condition for our Jacksonville classroom, a constant- volume system with supply-air tempering performs identically to the basic system in our original example (Figure 14, p. 20). Both systems also respond identically when the sensible-cooling load in the space decreases; that is, they raise the supply-air temperature by reducing the cooling capacity of the coil. If the space humidity rises above the maximum limit, however, the system with supply-air tempering overcools the supply air by modulating the cooling-coil capacity to enforce the maximum humidity limit. Another common control strategy operates the cooling coil at full capacity when the humidity in the space exceeds the specified maximum limit, which lowers the relative humidity more quickly so that the system can return to the normal cooling mode.

In either case, the overcooled supply air then passes through the heating coil, which adds a small amount of heat to temper the air and avoid overcooling the space.

Recall that the basic, constant-volume system achieved a relative humidity of 67 percent at the peak dew-point condition. In the system with supply-air tempering (Figure 47), however, the humidity sensor increases the capacity of the cooling coil to avoid violating the 60 percent maximum limit for relative humidity; the resulting leaving-coil temperature is 58°F DB (14.4°C DB). The thermostat prevents overcooling by increasing the capacity of the heating coil, which tempers the supply air to 63°F (17.2°C). The total load on the cooling coil at this condition is 4.74 tons (16.7 kW); there is also a 8.14 MBh (2.4 kW) load on the heating coil.

On the mild and rainy day, the air leaves the cooling coil at 59°F (15°C), and is then tempered to 66.5°F (19.2°C) to enforce the 60-percent-humidity limit in the space. At this condition, the total cooling-coil load is 3.15 tons (11.1 kW) and the heating coil-load is 12.2 MBh (3.6 kW).

SYS-APM004-EN 51 Dehumidifying with Constant-Volume Mixed Air

Application considerations. Table 7 and Figure 48 summarize the coil loads at the peak dry-bulb, peak dew-point, and mild, rainy conditions. In this example, the largest cooling burden occurs at the peak dry-bulb condition. Table 7. Summary of coil loads for supply- Although the enthalpy of the outdoor air is higher at the peak dew-point air-tempering example1 condition, the humidity sensor in the space permits the relative humidity (RH) Coil load (kW) to reach the maximum limit; therefore, the enthalpy of the air leaving the coil is Cooling, tons Heating, MBh higher, too. Lowering the maximum RH limit may cause the highest cooling Peak DB 4.78 0.00 load to occur at a condition other than the peak dry bulb. (16.8) (0.0) ■ Peak DP 4.74 8.14 When designing a constant-volume system that includes supply-air (16.7) (2.4) tempering, size the cooling coil and central plant (in the case of a chilled Mild, rainy 3.15 12.20 water system) to handle the largest cooling load. Remember that the largest (11.1) (3.6) load may occur at full-load or part-load conditions, depending on the desired humidity limit and the load characteristics of the space. 1 Outdoor conditions refer to the classroom examples set in Jacksonville, Florida: ■ Adding a reheat coil to the supply-air path increases the fan power Peak dry bulb 96°F (35.7°C) DB, 76°F (24.5°C) WB requirement. Peak dew point 76°F (24.6°C) DP, 84°F (28.8°C) DB Mild, rainy 70°F (21.2°C) DB, 69°F (20.6°C) WB ■ Supply-air tempering using recovered heat can reduce system operating costs by avoiding the use of new energy for heat. Furthermore, it may allow the system to meet the requirements of energy standards (ASHRAE Standard 90.1, for example) and codes. ■ For proper system control, install a humidity sensor in the space. Also, assure that the air-handler controls can determine when to switch between the “standard cooling” and “dehumidification” modes, as well as modulate the capacity of each coil independently.

Figure 48. Comparison of coil loads for supply-air tempering at various conditions

52 SYS-APM004-EN Dehumidifying with Constant-Volume Mixed Air

Recovered heat

Tempering supply air requires a source of heat, but ASHRAE Standard 90.1– 2001 prohibits the use of “new” energy for tempering or reheat in constant- volume systems, doesn’t it? Not necessarily. Section 6.3.2.3 (see p. 8 in this manual) defines several exceptions for which new-energy reheat is permitted— smaller terminal equipment, midsize equipment that is capable of unloading to 50 percent capacity before reheat is used, and systems that serve certain types of spaces (museums, surgical suites, and supermarkets, for example). Furthermore, tempering is always permissible if at least 75 percent of the energy required for reheat is recovered on-site.

Restricted use of new energy for tempering will probably result in HVAC system designs that temper supply air by recovering heat from the cooling process, particularly in dehumidification applications. Recovering sensible heat from another part of the HVAC system reduces operating costs by avoiding the use of new energy (electricity, hot water, steam, gas) for that purpose. Sources of recoverable heat include:

■ Condenser water in a water-cooled, chilled water system ■ Hot refrigerant in a refrigeration system ■ Another air stream or another location in the same air stream (using an air- to-air heat exchanger)

Condenser-water heat recovery. In a water-cooled- application of For more information on recovering the vapor-compression refrigeration cycle, the compressor discharges hot heat from a water-cooled chiller, refrigerant vapor into a shell-and-tube heat exchanger or “condenser.” Heat including system configurations that transfers from the hot, high-pressure refrigerant vapor inside the condenser optimize this arrangement, refer to the shell to the relatively cool water flowing through the tubes; the loss of heat following manuals: causes the refrigerant to condense into a liquid. The warm condenser water is ■ Waterside Heat Recovery then pumped to a , where it is cooled by the outdoor air. Instead in HVAC Systems, Trane of rejecting the condenser heat to the cooling tower, it can be recovered and applications engineering manual used to temper supply air. SYS-APM005-EN Condenser-water heat recovery is especially cost-effective in supply-air- ■ Multiple-Chiller-System tempering applications. It not only provides sufficient heat for tempering, but Design and Control, Trane applications engineering manual allows the primary heating equipment () to be turned off during the SYS-APM001-EN summer. Of course, recovered heat is only available while the chiller operates.

■ Application Guide: Chiller Heat Any water-cooled chiller can provide sensible heat for supply-air tempering. Recovery, ASHRAE publication ISBN 1-8883413-74-5 ■ Chiller with a standard condenser The leaving-air temperatures in supply-air-tempering applications typically range from 55°F (13°C) to 75°F (24°C), so the water used for tempering need not need be hot. Most standard water-cooled can provide suitable condenser-water temperatures if operated at a slightly elevated refrigerant- condensing temperature.

SYS-APM004-EN 53 Dehumidifying with Constant-Volume Mixed Air

Figure 49. Condenser-water heat recovery using a plate-and-frame heat exchanger

Recirculating the same condenser water through an air handler’s hot water coil and an open cooling tower increases the potential for tube fouling. Adding a second water loop, another pump, and a plate-and-frame heat exchanger can eliminate this risk. One loop circulates water through the chiller condenser, the plate-and-frame heat exchanger, and the hot water coil; the other loop circulates water through the plate-and-frame heat exchanger and the open cooling tower (Figure 49).

This arrangement also accommodates closed-circuit cooling towers or other types of evaporative fluid coolers. However, it is less efficient than other methods of condenser-water heat recovery because it adds an intermediate heat-transfer step and uses more pump energy.

“Heating” chiller in sidestream position For systems with multiple chillers, using one of the chillers as a “heater” offers a more efficient alternative to an intermediate heat exchanger. The “heating” chiller does not require an additional condenser; instead, it is installed and controlled for the condenser heat that it rejects. The evaporator is connected to the chilled water loop, typically in the sidestream configuration (Figure 50); it provides only the cooling needed to satisfy the heating load on the condenser. The temperature of the water leaving the evaporator is a by-product, letting the more efficient cooling-only chillers meet the rest of the cooling load. In effect, the evaporator of the “heating” chiller precools the system return water, reducing the load on the downstream chillers.

This arrangement is well-suited for the year-round heating loads associated with supply-air tempering to control humidity…especially for buildings that require significantly less heating than cooling (buildings in hot, humid climates, for example). The “heating” chiller operates more efficiently than other heat- producing devices, and it is less expensive than a chiller equipped with a second, heat-recovery condenser.

54 SYS-APM004-EN Dehumidifying with Constant-Volume Mixed Air

Figure 50. Condenser-water heat recovery using a sidestream “heating” chiller

Chiller with heat-recovery condenser Figure 51. Centrifugal chiller with Another means of condenser-water heat recovery requires a chiller that is heat-recovery condenser equipped with either two separate condensers (Figure 51) or a single condenser containing two separate tube bundles. Both scenarios require two condenser- water loops: one loop circulates water through a cooling tower to reject heat from the standard condenser, and the other carries water from the dedicated heat-recovery condenser to the hot-water coil for supply-air tempering.

The hot refrigerant vapor discharged by the compressor migrates to the condenser with the lowest pressure. Condenser pressure is a function of that condenser’s leaving water temperature. Raising the leaving water temperature of the standard condenser—reducing the flow rate or increasing the entering water temperature—increases the heat available from the heat-recovery condenser. This approach eliminates the intermediate heat exchanger, but the additional condenser increases the initial cost of the chiller.

In a multiple-chiller system, installing the heat-recovery chiller in the sidestream configuration (Figure 52) provides two notable benefits. First, the

SYS-APM004-EN 55 Dehumidifying with Constant-Volume Mixed Air

Figure 52. Condenser-water heat recovery using a sidestream, heat-recovery chiller

heat-recovery chiller can supply warmer water than the other operating chillers. It provides only the cooling needed to offset the supply-air-tempering load on the heat-recovery condenser, effectively precooling the returning system water and letting the more efficient cooling-only chillers meet the rest of the cooling load.

Second, if piped in a primary–secondary (“decoupled”) configuration, the heat-recovery chiller receives the warmest system water. This arrangement maximizes recoverable heat as well as the efficiency of this chiller. Because it is positioned upstream of the bypass line, the heat-recovery chiller is not affected by excess flow from the supply side of the system (which would otherwise lower that chiller’s entering-water temperature).

56 SYS-APM004-EN Dehumidifying with Constant-Volume Mixed Air

Refrigerant heat recovery. Heat generated by the vapor-compression For more information about recovering refrigeration cycle is also recoverable from direct-expansion (DX) refrigerating heat from a DX refrigeration system, equipment and air-cooled chillers. This is typically accomplished by piping a refer to Refrigerant Heat Recovery, heat-recovering refrigerant coil downstream of the compressor, either in series Trane applications engineering manual with, or parallel to, the standard condenser coil. Sometimes described as a SYS-AM-5. ■ “hot-gas reheat coil” or “desuperheater,” the heat-recovery coil collects sensible heat from the hot refrigerant vapor and transfers it to the supply air stream. When the humidity indoors exceeds the desired upper limit, the evaporator (DX cooling) coil dehumidifies the supply air; the heat-recovery coil then tempers the cold, dry supply air to avoid overcooling the space.

Figure 53 depicts a refrigerant, heat-recovery coil that is piped in series with a standard air-cooled condenser. If the humidity exceeds the maximum limit, cooling capacity is increased in response to indoor humidity. The linked face- and-bypass dampers modulate the capacity of the heat-recovery coil to maintain the space temperature at setpoint. If the humidity does not exceed the maximum limit, then the cooling coil is controlled to maintain the space temperature at setpoint. Because tempering is unnecessary, the face dampers modulate closed and the bypass dampers modulate open. Directing the air stream around the heat-recovery coil lowers the pressure drop.

Figure 54 (p. 58) shows a refrigerant heat-recovery coil piped in parallel with the condenser in a water-source . If the humidity exceeds the maximum limit, the compressor operates to lower the indoor humidity. A two-position refrigerant valve opens to divert the hot, high-pressure refrigerant vapor from the compressor to the refrigerant heat-recovery coil, which is located downstream of the refrigerant-to-air heat exchanger. (The heat exchanger acts as an “evaporator” during the cooling mode.) This refrigerant valve cycles open and closed to maintain the space temperature at setpoint.

Figure 53. Direct-expansion HVAC system with series-piped condenser and refrigerant, heat-recovery coil

SYS-APM004-EN 57 Dehumidifying with Constant-Volume Mixed Air

Figure 54. Water-source heat pump with parallel-piped condenser and heat-recovery coil

Application considerations ■ Refrigerant heat recovery is readily packaged within an air handler and is not susceptible to freezing, which makes it a convenient and relatively inexpensive means to temper supply air. ■ If not factory-engineered and -installed, use care when selecting the heat- recovery coil and installing the refrigerant piping and controls. Pipe the heat-recovery coil in series with the condenser, and use face-and-bypass dampers to control capacity. Doing so simplifies the refrigeration circuit and facilitates proper compressor lubrication. ■ Heat for tempering is only available while the compressor operates. ■ The additional coil increases the airside pressure drop and associated fan- energy consumption.

Air-to-air heat recovery. Sensible heat for supply-air tempering can be recovered from another air stream (or another location in the same air stream) Air-to-Air Energy Recovery in by using an air-to-air heat exchanger. The air-to-air heat exchanger can be a coil HVAC Systems (Trane applications loop, a fixed-plate heat exchanger, a , or a sensible rotary heat engineering manual SYS-APM003-EN) exchanger (heat wheel). There are two configurations for using a sensible air-to- discusses the use of air-to-air heat air heat exchanger for supply-air tempering: series and parallel. exchangers for tempering supply air. ■

58 SYS-APM004-EN Dehumidifying with Constant-Volume Mixed Air

Figure 55. Air-to-air heat recovery applied in a series configuration

Series configuration Figure 55 depicts a constant-volume, mixed-air system. The air-to-air, sensible-energy-recovery device is applied in a series (or “wrap-around”) configuration. To temper the supply air, the device removes sensible heat from the air upstream of the dehumidifying/cooling coil and releases it downstream of the coil. Technically, this arrangement transfers heat from one location to another within the same air stream, rather than “recovering” it from elsewhere in the system.

Parallel configuration Figure 56 shows the same constant-volume, mixed-air system; this time, the sensible-energy-recovery device is applied in a “parallel” configuration. The device collects sensible heat from the return air stream and releases it downstream of the dehumidifying/cooling coil, warming the supply air.

Figure 56. Air-to-air heat recovery applied in a parallel configuration

SYS-APM004-EN 59 Dehumidifying with Constant-Volume Mixed Air

Series or parallel? In constant-volume, mixed-air systems, both series and parallel configurations reduce the heating energy required for tempering. When comparing the amount of recoverable heat, however, the return air stream (parallel configuration) is a more constant source of heat than the outdoor air (series configuration). If recovered heat is needed when the outdoor air is warmer than the return air, the series configuration transfers more heat; but if recovered heat is needed when the outdoor air is cooler than the return air, the parallel configuration transfers more heat.

Although both configurations “precool” the entering air, which saves cooling energy when tempering, neither configuration permits downsizing of the cooling and heating plants. At the full-load, peak dry-bulb condition, the air is supplied to the space at the design (cold) supply-air temperature. Tempering is unnecessary, so no precooling occurs. Therefore, the cooling coil and cooling plant must be sized to handle the total design-cooling load.

Finally, either configuration may require supplemental heat at certain conditions; also, both require a method for modulating the capacity of the air- to-air heat exchanger to avoid overheating the space during dehumidification. The right choice for a given project depends on the balance of initial cost, energy savings for cooling and heating, and increased fan energy resulting from the additional static-pressure loss through the heat exchanger. ■

60 SYS-APM004-EN Dehumidifying with Variable-Volume Mixed Air

Mixed-air systems use an air handler to condition a combination of outdoor air and recirculated return air before delivering the mixed air to each space. The variable-air-volume (VAV) version of a mixed-air system (Figure 57) consists of a central air handler and multiple VAV terminals, each of which is controlled by a space thermostat. Unlike a constant-volume system, which delivers a constant amount of air at varying temperatures, a VAV system delivers varying amounts of constant-temperature air, typically 45°F to 55°F DB (7°C to 13°C DB).

A thermostat in each space compares the dry-bulb temperature to a setpoint, and the VAV terminal responds by modulating the volume of supply air to match the changing sensible-cooling load in the space. Meanwhile, the central supply fan modulates to maintain the static-pressure setpoint in the duct system, and the capacity of the central cooling coil modulates to maintain a constant supply-air dry-bulb temperature.

VAV systems typically provide effective, coincidental dehumidification over a wide range of indoor load conditions (sensible-heat ratios). If supply-air- temperature reset is not used, and as long as any space needs cooling, the VAV air handler will supply dry (low-dew-point) air to all VAV terminals. Moisture generated within the space is absorbed by the dry supply air—offsetting the latent load—then removed from the space by the return air stream.

Analysis of Dehumidification Performance

Accurate predictions of coincidental dehumidification require an analysis of system operation at both full-load and part-load conditions. The following examples are based on the 10,000 ft³ (283 m³), 30-occupant classroom in Jacksonville, Florida. Unlike the previous constant-volume examples, which featured a single-space system, the classroom is served by a multiple-space system. To provide thermal comfort, the target condition is 74°F DB (23.3°C DB)

Figure 57. Basic, variable-air-volume HVAC system

SYS-APM004-EN 61 Dehumidifying with Variable-Volume Mixed Air

and 50 percent-relative humidity, with a design supply airflow of nine air changes or 1,500 cfm (0.7 m³/s) per hour.

Performance at peak dry-bulb (full-load) condition. At the peak dry-bulb At the peak dry-bulb condition: condition, the space sensible-cooling load and supply-air temperature are the × , × ()° Qs = 1.085 1 500 cfm 74 FT– supply same as for a constant-volume system. Given the supply airflow of 1,500 cfm , ∴ ° (0.7 m³/s), the system must deliver 55.7°F (13.1°C) supply air to offset the = 29 750 Btu/hr Tsupply = 55.7 F sensible cooling load in the space and satisfy the thermostat setpoint of 74°F DB ()× × []° (23.3°C DB). Qs = 1.21 0.7 m³/s 23.3 CT– supply ()∴ ° = 8.7 kW Tsupply = 13.1 C Psychrometric analysis (Figure 58) reveals that the cooling coil removes both sensible heat and moisture from the air, directly controlling space temperature and coincidentally affecting space humidity. Maintaining the temperature in the space at 74°F (23.3°C) requires a total capacity of 4.78 tons (16.8 kW) from the cooling coil and results in a comfortable relative humidity of 52 percent.

As the sensible-cooling load decreases, the VAV system supplies less air to the space while maintaining a constant supply-air temperature.

Performance at peak dew-point (part-load) condition. Lower solar- and At the peak dew-point condition: conducted-heat gains and cooler outdoor air reduce the sensible-cooling load in × × () Qs = 1.085 Vsa 74– 55.7°F the classroom. With no change in the occupant-generated latent load, the = 17,850 Btu/hr sensible-heat ratio for the space drops to 0.77. The supply-air temperature remains constant at 55.7°F (13.1°C), so supply airflow is reduced to 899 cfm ∴ V = 899 cfm sa (0.42 m³/s). The required cooling capacity is 4.0 tons (14.2 kW). ( × × []) Qs = 1.21 Vsa 23.3– 13.1°C Because the supply air is still cool and dry, the relative humidity in the ()= 5.2 kW classroom only rises to 57 percent (Figure 58). By contrast, the relative humidity ()∴ Vsa = 0.42 m³/s reaches 67 percent when the classroom is served by a basic constant-volume system operating at the same condition (p. 21).

Figure 58. Dehumidification performance of a variable-air-volume HVAC system at various outdoor conditions

Design Full load Part load condition Peak dry bulb Peak dew point Mild, rainy OA 96.0°F DB, 76.0°F WB 76.0°F DP, 84.0°F DB 70.0°F DB, 69.0°F WB RA 74.0°F DB, 52.4% RH 74.0°F DB, 57.0% RH 74.0°F DB, 60.0% RH MA 80.6°F DB 79.0°F DB 71.1°F DB SA 55.7°F DB 55.7°F DB 55.7°F DB

62 SYS-APM004-EN Dehumidifying with Variable-Volume Mixed Air

Performance on a mild, rainy day (part-load condition). Although the peak dew-point condition is helpful for analyzing the part-load dehumidification Mild, rainy day: performance of an HVAC system, do not assume that it represents worst-case × × () Qs = 1.085 Vsa 74– 55.7°F conditions for humidity control. Typically, indoor humidity depends as much on = 12,250 Btu/hr the sensible and latent cooling loads in the space, the type of HVAC system, and ∴ the method of controlling that system, as it does on outdoor conditions. Vsa = 617 cfm Consider our example Jacksonville classroom on a mild, rainy day (Figure 58)— ()× × [] Qs = 1.21 Vsa 23.3– 13.1°C 70°F DB, 69°F WB (21.2°C DB, 20.6°C WB). With no change in the occupant- ()= 3.6 kW generated latent load, the space sensible load drops to 12,250 Btu/hr (3.6 kW) ()∴ Vsa = 0.29 m³/s and the space sensible-heat ratio drops to 0.70. Only 617 cfm (0.29 m³/s) of constant-temperature supply air is required to offset the sensible load without overcooling the space. At this airflow, the relative humidity increases to 60 percent while the required capacity from the cooling coil decreases to 2.1 tons (7.4 kW).

Application Considerations

Minimum Airflow Settings

In most applications, each VAV terminal has a minimum airflow setting that usually represents either the ventilation requirement for the space or the performance limits of the diffusers or VAV terminal. Providing less than the required minimum airflow may:

■ Underventilate the space and degrade indoor air quality. ■ “Dump” cold supply air into the space, making occupants uncomfortable. (Most diffusers require a minimum discharge velocity to properly mix the air within the space.) ■ Cause erroneous airflow readings that interfere with proper control. (The accuracy of the flow sensor in the VAV terminal is based upon a specific airflow range.)

The minimum airflow setting for the VAV terminal that serves our example classroom is 550 cfm (0.26 m³/s).

Eventually, the sensible-cooling load in the space becomes small enough Mild, rainy day and minimum airflow: that the required supply (primary) airflow is less than the minimum airflow × × () Qs = 1.085 700 cfm Tspace – 55.7°F setting of the VAV terminal. If we assume that the minimum airflow setting for , ∴ the example classroom is 700 cfm (0.33 m³/s), then this situation occurs on a = 12 250 Btu/hr Tspace = 71.8°F mild, rainy day. If the supply-air temperature is held constant at 55.7°F (13.1°C), the VAV system will overcool the space to 71.8°F (22.1°C). ()× × [] Qs = 1.21 0.33 m³/s Tspace – 13.1°C ()∴ = 3.6 kW = Tspace = 22.1°C A psychrometric analysis (Figure 59, p. 64) reveals that the relative humidity will climb to 66 percent because of the decreased dry-bulb temperature. As a

SYS-APM004-EN 63 Dehumidifying with Variable-Volume Mixed Air

Figure 59. Overcooling results when minimum airflow exceeds required airflow

result, the classroom will feel cool and damp even though the actual moisture content of the air is unchanged from the previous example in which only 617 cfm (0.29 m³/s) is supplied to the classroom.

Supply-Air-Temperature Reset

Mild, rainy day and SA temperature reset: One way to prevent overcooling is to reset the supply-air temperature upward × × () Qs = 1.085 700 cfm 74°F– Tsupply at low-load conditions. If the minimum airflow setting is 700 cfm (0.33 m³/s), , ∴ raising the supply-air temperature to 57.9°F (14.3°C), for example, avoids = 12 250 Btu/hr Tsupply = 57.9°F overcooling the classroom on a mild, rainy day…but the cooling coil also removes less moisture from the supply air. As a result, the relative humidity in ()Q = 1.21× 0.33 m³/s × []23.3°C– T s supply the space increases to 65 percent (Figure 60). ()∴ = 3.6 kW Tsupply = 14.3°C Resetting the supply-air temperature reduces the energy consumed by the mechanical cooling equipment—from 2.1 tons (7.2 kW) to 1.9 tons (6.7 kW) in

Figure 60. Effect of supply-air-temperature reset on dehumidification performance

64 SYS-APM004-EN Dehumidifying with Variable-Volume Mixed Air

this example. All spaces receive warmer air. Therefore, the spaces not only become more humid, but also require more air to offset the sensible-cooling loads. The fans therefore consume more energy.

Note: In VAV applications, supply-air-temperature reset is typically used to avoid excessive reheat during cold weather. Avoid using supply-air-temperature reset during the cooling season unless the system analysis indicates that the savings in mechanical cooling and reheat energy will outweigh the increase in supply-fan energy and space humidity.

Supply-Air Tempering at VAV Terminals

Figure 61. VAV terminal with heating coil Overcooling and increased humidity can be avoided by tempering the supply air when it diminishes to the minimum airflow setting of the VAV terminal. “Tempering” moderates the cooling effect by adding sensible heat to the supply air, either at the VAV terminal (Figure 61) or within the space.

Figure 62 illustrates the effect of adding supply-air tempering to the VAV system that serves our example classroom. When the supply airflow is reduced to the minimum airflow setting of 700 cfm (0.33 m³/s), a heating coil in the VAV terminal warms the 55.7°F (13.1°C) supply air to 57.9°F (14.3°C) before delivering it to the space. This avoids overcooling the classroom and, on a mild and rainy day, results in a relative humidity of 60 percent. The total load on the cooling coil is 2.1 tons (7.4 kW), while the load on the heating coil is 1.7 MBh (0.49 kW).

Application considerations ■ Certain zones in a VAV system typically require tempering, even when high sensible-cooling loads exist elsewhere. To curb operating costs, consider on-site recovered heat (discussed on p. 66) for supply-air tempering. ■ If the VAV control strategy includes supply-air-temperature reset, provide a humidity sensor to regulate the humidity in the space. Raising the supply-air

Figure 62. Dehumidification performance of a VAV system with supply-air tempering at VAV terminals

SYS-APM004-EN 65 Dehumidifying with Variable-Volume Mixed Air

temperature not only increases supply airflow but also the energy consumption of the fans. The increase in fan energy often exceeds the cooling and tempering energy saved as a result of resetting the supply-air temperature. ■ When using coils, comply with the manufacturer’s guidelines for minimum airflow limits across the heating elements to assure safe operation.

Options for supply-air tempering include radiant heaters in the space, heating coils mounted on the VAV terminals, fan-powered VAV terminals, and dual-duct air distribution.

Heating coils at VAV terminals

It is common to think of electricity, hot water, or steam as a source of heat for Doesn’t ASHRAE Standard 90.1–2001 supply-air tempering performed by a heating coil at the VAV terminal. However, prohibit the use of new-energy “reheat” recovering sensible heat from elsewhere in the HVAC system reduces operating in VAV terminals? costs. For example, the sensible heat collected from the condenser of a water- cooled chiller is easily distributed to VAV terminals throughout the building. Not necessarily. Section 6.3.2.3 (see p. 8 in this manual) defines several Condenser-water heat recovery is particularly well-suited for supply-air exceptions for which new-energy reheat tempering applications: It provides the relatively small amount of heat needed is permitted. Exception A in the standard permits the use of new energy for tempering and allows the primary heating equipment (boilers, for example) for reheat after the supply airflow is to be turned off during the summer. Any water-cooled chiller can be used to reduced to a defined limit. provide sensible heat for supply-air tempering. Examples of common system configurations were discussed in the previous chapter; see pp. 53–56. The minimum airflow setting for most zones in a VAV system is less than the Fan-powered VAV terminals limits defined by this section of Standard 90.1. ■ When return air from the space passes through an open ceiling plenum, it collects heat from the lights and roof. A fan-powered VAV terminal mixes this warm plenum air with cold primary air to provide local tempering at low-load conditions. Depending on the application, the plenum air may be warm enough to reduce or eliminate the need for a supplemental heating coil.

There are two types of fan-powered VAV terminals: “parallel” and “series.” These classifications describe the arrangement of the fan in the VAV terminal relative to the primary-air fan in the central air handler (Figure 63).

The small, constant-volume fan in a parallel, fan-powered VAV terminal is situated in the local recirculated-return-air path, parallel to the primary-air fan. As the cooling load in the space decreases, the central air handler delivers less primary air to the VAV terminal. The small fan in the VAV terminal only operates when the primary airflow drops to the minimum airflow setting. Mixing recirculated return air from the plenum with the cool primary air increases the total airflow delivered to the space and raises the supply-air temperature.

66 SYS-APM004-EN Dehumidifying with Variable-Volume Mixed Air

Figure 63. Fan-powered VAV terminals

The slightly larger, constant-volume fan in a series, fan-powered VAV terminal is positioned in the local supply-air path so that it is in series with the primary- air fan. Unlike the parallel configuration, the fan in a series, fan-powered VAV terminal operates continuously when the space is occupied. The fan draws air from both the primary air stream and the plenum to supply the space with a constant volume of air at all times. As the cooling load in the space decreases, the central air handler delivers less primary air to the VAV terminal. To maintain a constant supply airflow, the VAV terminal draws in more recirculated return air from the plenum. A heating coil can provide supplemental supply-air tempering if the cooling load drops below the VAV terminal’s minimum primary-airflow setting.

A system that uses series, fan-powered VAV terminals does a better job of dehumidifying the space at part-load conditions than systems equipped with other types of VAV terminals. Series, fan-powered VAV terminals require more primary airflow (PA) to offset the warm return air (RRA), which is recirculated to provide the space with a constant supply airflow. More of the dry primary air results in lower space humidity.

Dual-duct air distribution

VAV systems that are designed for dual-duct air distribution also temper supply air at part-load conditions by mixing warm air with cold primary air. Instead of using recirculated return air from the plenum, the VAV terminal receives warm primary air through separate ductwork. Two modulating devices, one for each air stream, control the amount of cool and/or warm primary air that enters the

SYS-APM004-EN 67 Dehumidifying with Variable-Volume Mixed Air

Figure 64. Dual-duct VAV terminal VAV terminal (Figure 64). The primary air streams mix inside the VAV terminal and are then delivered to the space (Figure 65).

As the cooling load in the space decreases, the modulation device that controls the cool primary air modulates toward its minimum-open position. When the cooling load drops to the point where the required amount of cool primary air is less than the VAV terminal’s minimum airflow setting, the second modulation device begins to open. This allows warm primary air to mix with and temper the cool primary air before it is supplied to the space.

With further decreases in the cooling load, the space will eventually require heating. To offset an increasing heating load, the VAV terminal mixes the minimum amount of cool primary air with ever-increasing amounts of warm primary air. When the heating load becomes large enough, the recirculated return air is heated before it is delivered to the VAV terminals as warm primary air.

Figure 65. VAV system with dual-duct air distribution

Humidity Control during Unoccupied Periods

Buildings that are served by VAV systems may require an after-hours source of “reheat” energy—for example, heating coils at the VAV terminals or warm plenum air if the VAV terminals are fan-powered. If the VAV system includes an air handler that exclusively conditions the outdoor air, then the dedicated outdoor-air handler also can provide after-hours dehumidification. (The next chapter discusses dedicated outdoor-air systems in detail.)

68 SYS-APM004-EN Dehumidifying with Variable-Volume Mixed Air

Building Pressurization

For most buildings, the difference between indoor and outdoor static pressures Consult Building Pressurization results directly from the combined effect of continuously changing conditions: Control, Trane applications engineering weather (“stack effect”), wind, and operation of the mechanical ventilation manual AM-CON-17, for information system (local exhaust fans, airside economizer). Maintaining the desired about regulating building pressure pressure difference, even during normal daytime operation, is particularly through design and control of the challenging in VAV applications because the supply airflow also changes. ■ HVAC system. Preventing both infiltration and economizer-induced overpressurization requires a control strategy that directly controls building pressure. Such strategies monitor building pressure and then modulate relief airflow accordingly—by either adjusting the capacity of the relief fan or the position of the relief damper—to maintain the desired pressure difference across the building envelope.

Airside Economizing

Climate, hours of occupancy, and potential savings in operating cost usually influence the choice between methods of economizer control. In most VAV applications, however, the differential (comparative) enthalpy economizer best balances dehumidification performance and operating-cost savings. As its name implies, comparative enthalpy control compares the enthalpy of the outdoor air to the enthalpy of the recirculated return air. When the outdoor air has a lower enthalpy than the return air, the outdoor-air damper fully opens. This strategy reduces cooling-energy consumption. Because the air that passes through the cooling coil in a VAV application is always dehumidified to the same low dew point, it avoids introducing unwanted moisture into the space.

Note: If the VAV system includes both an airside economizer and supply-air- temperature reset, make sure that the control scheme will not introduce humid outdoor air into the space.

SYS-APM004-EN 69 Dehumidifying with Variable-Volume Mixed Air

Improving Dehumidification Performance

VAV systems can provide effective coincidental dehumidification over a wide range of indoor load conditions; but the basic design of the system can be altered to enhance dehumidification performance.

Condition Outdoor Air Separately

Figure 66. Dual-duct VAV terminal used One way to improve indoor humidity control is to separately treat the outdoor with dedicated outdoor air air before mixing it with recirculated return air. A dedicated air handler cools and dehumidifies all of the outdoor air to a dew point that is drier (lower) than the space. The conditioned outdoor air then is delivered to one or more VAV air handlers, or directly to the individual, dual-duct VAV (Figure 66) terminals that serve each zone. The “ventilation” damper in the dual-duct VAV terminal maintains the required quantity of outdoor air from the dedicated outdoor-air unit, while the “primary-air” damper regulates the 100 percent-recirculated return air from the VAV air handler.

To demonstrate how separately conditioning the outdoor air affects dehumidification performance, let’s revisit the example classroom. Assume that the dedicated outdoor-air handler supplies 450 cfm (0.21 m³/s) of outdoor air to the dual-duct VAV terminal serving the classroom. The outdoor air is cooled and dehumidified to 52°F DP (11.1°C DP) and delivered—without tempering or reheat—to the “ventilation” damper in the dual-duct terminal.

At the full-load, peak dry-bulb condition (Figure 67), the VAV air handler delivers 1,050 cfm (0.5 m³/s) of 57.3°F (14.1°C), primary air to the second damper in the dual-duct VAV terminal. Inside the VAV terminal, primary air (PA) mixes with conditioned outdoor air (CA); the resulting supply air yields a 50-percent relative humidity in the classroom. The cooling-coil load is 3.1 tons (10.8 kW)

Figure 67. Dehumidification performance of a VAV system with separately conditioned outdoor air at peak dry-bulb condition

Space sensible-cooling load offset by dedicated outdoor-air handler: × × (), Qs,doa = 1.085 450 cfm 74– 52°F = 10 742 Btu/hr ()× × [] Qs,doa = 1.21 0.21 m³/s 23.3– 11.1°C = 3.1 kW

Recirculating air handler at peak dry-bulb condition: , , Qs = 29 750 – 10 742 Btu/hr × , × ()∴ = 1.085 1 050 cfm 74°F– Tpa Tpa = 57.3°F () Qs = 8.7– 3.1 kW ()× × []∴ = 1.21 0.5 m³/s 23.3°C– Tpa Tpa= 14.1°C

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Figure 68. Dehumidification performance of a VAV system with separately conditioned outdoor air at peak dew-point condition

Space sensible-cooling load offset by dedicated outdoor-air handler: × × (), Qs,doa = 1.085 450 cfm 74– 52°F = 10 742 Btu/hr ()× × [] Qs,doa = 1.21 0.21 m³/s 23.3– 11.1°C = 3.1 kW

Recirculating air handler at peak dew-point condition: , , Qs = 18 750 – 10 742 Btu/hr × × ()∴ = 1.085 Vpa 74– 57.3°F Vpa = 392 cfm () Qs = 5.2– 3.1 kW ()× × []∴ = 1.21 Vpa 23.3– 14.1°C Vpa= 0.18 m³/s

for the dedicated outdoor-air handler and 1.8 tons (6.2 kW) for the VAV air handler.

Less sensible cooling is required at the part-load, peak dew-point condition (Figure 68). Therefore, while the dual-duct VAV terminal receives the same amount of 52°F DP (11.1°C DP), conditioned outdoor air, it receives only 392 cfm (0.18 m³/s) of primary air from the VAV air handler. The combined airflow, 842 cfm (0.40 m³/s), yields a relative humidity of 55 percent in the classroom. This time, the cooling-coil load is 3.4 tons (12.1 kW) for the dedicated outdoor- air handler and 0.7 tons (2.6 kW) for the VAV air handler.

On a mild, rainy day (Figure 69, p. 72), the VAV terminal mixes 450 cfm (0.21 m³/s) of conditioned outdoor air with only 83 cfm (0.04 m³/s) of primary air, which results in a relative humidity of 58 percent. In this application, the minimum airflow for the VAV terminal is only 450 cfm (0.21 m³/s), which equals the ventilation requirement for the classroom.

Cooling-coil loads are 2.0 tons (7.2 kW) for the dedicated outdoor-air handler and 0.2 tons (0.7 kW) for the VAV air handler.

SYS-APM004-EN 71 Dehumidifying with Variable-Volume Mixed Air

Figure 69. Dehumidification performance of a VAV system with separately conditioned outdoor air on a mild, rainy day

Space sensible-cooling load offset by dedicated outdoor-air handler: × × (), Qs,doa = 1.085 450 cfm 74– 52°F = 10 742 Btu/hr ()× × [] Qs,doa = 1.21 0.21 m³/s 23.3– 11.1°C = 3.1 kW

Recirculating air handler on a mild, rainy day: , , Qs = 12 250 – 10 742 Btu/hr × × ()∴ = 1.085 Vpa 74– 57.3°F Vpa = 83 cfm () Qs = 3.6– 3.1 kW ()× × []∴ = 1.21 Vpa 23.3– 14.1°C Vpa= 0.04 m³/s

Application considerations ■ In a traditional, mixed-air VAV system, the multiple-space equation For more information about the (Equation 6–1) from ASHRAE Standard 62–2001 requires that the VAV air “multiple-space” equation from handler bring in more outdoor air than the sum of the space ventilation ASHRAE Standard 62 and how it requirements. A dedicated-outdoor-air design delivers conditioned outdoor applies to traditional, mixed-air VAV air directly to individual VAV terminals (or spaces) and, therefore, is not systems, see “The Threefold Challenge considered to be a multiple-space, recirculating ventilation system. For this of Ventilating Single-Duct VAV reason, a dedicated outdoor-air system requires less total outdoor airflow Systems” in Trane Engineers Newsletter than a traditional VAV design. ENEWS-27/1. It is archived in the “newsletters” section of www.trane.com/ ■ Using a dual-duct VAV terminal in this manner typically requires the commercial. ■ addition of a reheat coil for tempering at low cooling loads. (Eventually, the space sensible-cooling load decreases to the point that the primary-air damper closes. The cold, conditioned outdoor air then must be tempered to avoid overcooling the space.) ■ As an alternative to separately conditioning the outdoor air in a VAV system, modify the fan-powered VAV terminals by adding a second “ventilation” damper. The modified VAV terminals operate similarly to dual-duct VAV terminals, but the “free” heat from the plenum can be used to temper the conditioned outdoor air when the space cooling load is low.

The next chapter, “Dehumidifying with Dedicated Outdoor Air,” discusses system configurations, design procedures, and application considerations for air distribution systems that separately treat outdoor air.

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Deliver Colder Supply Air

Peak dry-bulb condition: Lowering the leaving-air temperature for the central cooling coil condenses × × () Qs = 1.085 Vsa 74– 50°F more moisture from the supply air and requires less airflow to offset the , ∴ , sensible-cooling load in the space. = 29 750 Btu/hr Vsa = 1 142 cfm

()× × [] At the peak dry-bulb condition (Figure 70), delivering supply air to the example Qs = 1.21 Vsa 23.3– 10°C classroom at 50°F (10°C) instead of 55.7°F (13.1°C) reduces the required airflow ()= 8.7 kW∴ V = 0.54 m³/s sa from 1,500 cfm (0.7 m³/s) to 1,142 cfm (0.54 m³/s). Supplying the classroom with colder, drier supply air also reduces the relative humidity from 52 percent to Peak dew-point condition: 47 percent. The required cooling-coil capacity is 5.1 tons (17.8 kW). × × () Qs = 1.085 Vsa 74– 50°F The classroom needs less of the constant-temperature supply air at the peak = 17,850 Btu/hr∴ V = 685 cfm sa dew-point condition because the sensible-cooling load is smaller. To match the ()× × [] load in the space, the VAV system reduces the supply airflow to 685 cfm Qs = 1.21 Vsa 23.3– 10°C (0.32 m³/s). The resulting relative humidity is 50 percent, and the required ()= 5.2 kW∴ V = 0.32 m³/s sa cooling capacity is 4.3 tons (15.1 kW).

Mild, rainy day: On a mild, rainy day, the amount of supply air needed to offset the classroom’s × × ()° sensible-cooling load is less than the VAV terminal’s minimum airflow setting; Qs = 1.085 550 cfm 74 FT– supply local tempering is required to avoid overcooling. As in the example at the = 12,250 Btu/hr∴ T = 53.5°F supply beginning of this chapter, the minimum airflow setting is 550 cfm (0.26 m³/s). ()× × []° The VAV terminal must temper the air to 53.5°F (11.9°C) before delivering it to Qs = 1.21 0.26 m³/s 23.3 CT– supply the space. The resulting relative humidity in the classroom is 52 percent. The ()= 3.6 kW∴ T = 11.9°C supply cooling-coil load is 2.5 tons (8.8 kW), while the heating-coil at the VAV terminal is 2.1 MBh (0.6 kW).

Figure 70. Effect of “cold” supply air on dehumidification performance of a VAV system

Design Full load Part load condition Peak dry bulb Peak dew point Mild, rainy OA 96.0°F DB, 76.0°F WB 76.0°F DP, 84.0°F DB 70.0°F DB, 69.0°F WB RA 74.0°F DB, 47% RH 74.0°F DB, 50% RH 74.0°F DB, 52% RH MA 82.7°F DB 80.6°F DB 70.7°F DB SA 50.0°F DB 50.0°F DB 53.5°F DB

SYS-APM004-EN 73 Dehumidifying with Variable-Volume Mixed Air

Application considerations ■ In addition to drier spaces, other benefits of cold-air systems include smaller A previous issue of the Trane Engineers air handlers, VAV terminals, and ducts, as well as less supply-fan energy. Newsletter, titled “Cold Air Makes Good ■ Increased reheat energy and fewer hours of airside economizer operation $ense” (ENEWS-29/2), discusses the partially offset the supply-fan energy savings. Intelligent system control is benefits and design considerations associated with cold-air VAV systems. crucial to fully realize the potential savings. To read it, visit the “newsletters” section ■ Some designers raise the space thermostat setpoint to further reduce the of www.trane.com/commercial. supply airflow in cold air systems. This decision increases the indoor

The Cold Air Distribution System dew point, which may negate the benefit of otherwise drier spaces. Design Guide (ISBN 1-883413-37-0), by ■ Applied, chilled water systems typically work best for “cold air” distribution Allan T. Kirkpatrick and James S. because the designer can match the design requirements for airflow and Elleson, is another useful reference. It is cooling capacity. Packaged, direct-expansion (DX) systems defer many available from ASHRAE’s online design decisions to the manufacturer, which reduces the initial cost; bookstore at www.ashrae.org. ■ however, the limitations of a fixed design may make it difficult to achieve the desired “cold coil” temperature. ■

74 SYS-APM004-EN Dehumidifying with Dedicated Outdoor Air

System Configurations

Figure 71. Configurations for dedicated A dedicated outdoor-air handler separately filters, cools, dehumidifies, heats, outdoor-air systems and/or humidifies all ventilation (outdoor) air. It also can enable the direct control of indoor humidity. There are several ways to lay out a constant- or variable-air-volume HVAC system for separate treatment of ventilation air (Figure 71). These configurations are based on two factors:

■ Where the conditioned outdoor air is delivered—directly to occupied spaces, to air terminals, or to other types of HVAC equipment ■ The dry-bulb temperature of the conditioned air (CA)—neutral, which approximates room air, or cold, which usually approximates conventional, 55°F (13°C) supply air.

Table 8 summarizes the traits of each configuration.

Table 8. Configurations for HVAC systems with dedicated outdoor-air treatment

System schematic Typical characteristics Application considerations

Neutral CA to space Dedicated outdoor-air handler: ■ Provides design flexibility by accommodating ■ Offsets latent- and sensible-cooling loads any type of terminal device, plus horizontal resulting from ventilation (outdoor air) and vertical discharge arrangements ■ Cools and dehumidifies the outdoor air so that ■ Well-suited for applications with coincidental its dew point is lower than that of room air heating and cooling needs because it has little ■ Reheats conditioned outdoor air to a neutral effect on space loads dry-bulb temperature ■ Commonly applied with in-room HVAC ■ Delivers conditioned outdoor air (via equipment, such as fan–coils, through-the- ductwork) directly to the space wall air conditioners, and classroom unit ventilators Local HVAC units: ■ Are installed within the occupied space ■ Offset local sensible-cooling and heating loads

Cold CA to space Dedicated outdoor-air handler: ■ Typically applied with in-room HVAC ■ Offsets latent- and sensible-cooling loads equipment; also used with radiant ceiling resulting from ventilation (outdoor air) panels and other non-traditional devices ■ Cools and dehumidifies the outdoor air so that ■ Requires less cooling capacity from the dry-bulb and dew-point temperatures are terminal devices, which permits downsizing, lower than room air but also may create heating loads in spaces ■ Does not reheat the conditioned outdoor air that require little or no cooling ■ Delivers conditioned outdoor air (via ■ To prevent uncomfortable drafts, select high- ductwork) directly to the space performance diffusers that induce room air to mix with (and warm) the cold conditioned air Local HVAC units: before it reaches occupants ■ Are installed within the occupied space ■ Adding even a small amount of heat to the ■ Offset local sensible-cooling and conditioned air after it leaves the cooling coil heating loads will increase the cooling capacity needed from the terminal devices

continued on next page

SYS-APM004-EN 75 Dehumidifying with Dedicated Outdoor Air

Table 8. Configurations for HVAC systems with dedicated outdoor-air treatment (continued)

System schematic Typical characteristics Application considerations

Neutral CA to other equipment… Dedicated outdoor-air handler: ■ Commonly applied with central air handlers ■ Offsets latent- and sensible-cooling loads (chilled water or DX) or local HVAC terminals, resulting from ventilation (outdoor air) such as blower–coils, packaged rooftop air ■ Cools and dehumidifies the outdoor air conditioners, vertical self-contained air conditioners, and dual-duct VAV terminals ■ Reheats conditioned outdoor air to a neutral dry-bulb temperature ■ Well-suited for applications with coincidental heating and cooling needs because it has little ■ Delivers conditioned outdoor air (via effect on space loads ductwork) to outdoor-air inlets of local units

Local HVAC units: ■ Are installed outside the occupied space in ceiling plenums, closets, or equipment rooms ■ Mix recirculated return air with conditioned outdoor air inside the unit casing ■ Offset local sensible-cooling and heating loads

…or to plenum (near other equipment) Dedicated outdoor-air handler: ■ Typically applied with water-source heat ■ Offsets latent- and sensible-cooling loads pumps, horizontal fan–coils, and other local resulting from ventilation (outdoor air) HVAC units without mixing boxes ■ Cools and dehumidifies the outdoor air ■ Requires ducts to deliver conditioned outdoor ■ Reheats conditioned outdoor air to a neutral air within 5 ft (1.5 m) of local HVAC units dry-bulb temperature ■ Requires a local or powered central exhaust ■ Delivers conditioned outdoor air (via system ductwork) near return-air inlets of local units ■ To prevent condensation on beams, ducts, and other surfaces, assure that the CA dry- Local HVAC units: bulb temperature exceeds the plenum-air ■ Are installed in the ceiling plenum dew point ■ Offset local sensible-cooling and heating loads ■ Are without mixing boxes, so conditioned outdoor air and recirculated return air mix in the ceiling plenum

Cold CA to other equipment Dedicated outdoor-air handler: ■ Typically applied with blower–coils, packaged ■ Offsets latent- and sensible-cooling loads rooftop air conditioners, central air handlers, resulting from ventilation (outdoor air) self-contained air conditioners, and dual-duct ■ Cools and dehumidifies outdoor air VAV terminals ■ Does not reheat the conditioned outdoor air ■ Spaces that need little or no cooling may require local reheat ■ Delivers conditioned outdoor air (via ductwork) to outdoor-air inlets of local units ■ Requires less cooling capacity from terminal devices, which permits downsizing Local HVAC units: ■ Avoids reheat at the dedicated outdoor-air ■ Are installed outside the occupied space handler ■ Mix recirculated return air with conditioned outdoor air inside the unit casing ■ Offset sensible-cooling and heating loads in the space

76 SYS-APM004-EN Dehumidifying with Dedicated Outdoor Air

Design Objectives for Conditioned Outdoor Air

Moisture Content

Regardless of where the conditioned outdoor air is delivered, the dedicated Appendix B in this manual describes outdoor-air handler should make the incoming outdoor air drier than the air in how to determine the conditioned-air the space. Delivering drier (low-dew-point) air eliminates the sensible- and dew point that will allow the dedicated latent-cooling loads associated with ventilation. If the dew-point temperature of outdoor-air handler to offset the local the conditioned outdoor air is lower than the dew point in the space, it will also latent cooling loads. ■ offset the local latent-cooling loads. Consequently, the HVAC terminals only need to offset the local sensible-cooling loads.

Delivering low-dew-point, conditioned air can greatly improve the dehumidification performance of constant-volume systems, particularly at part-load conditions. (See “‘Direct’ Control of Humidity,” pp. 44–46.) It also allows the HVAC terminals to “run dry” under most conditions: With the dedicated outdoor-air handler offsetting the local latent loads, the space dew point can be kept below the dew point of the air supplied by the local terminals. Little or no moisture condenses on local cooling coils and collects in drain pans, perhaps eliminating one source of moisture within the space and the HVAC system.

Why not condition the outdoor air to a neutral dry-bulb temperature without overcooling it to obtain a low dew point? To demonstrate the effect of this design choice, let’s revisit the 30-occupant classroom in Jacksonville, Florida, which was introduced on p. 19. For this example, the HVAC system includes a dedicated outdoor-air handler that cools the outdoor air to a neutral dry-bulb temperature of 74°F (23.3°C) without overcooling it; the conditioned air is then delivered directly to the classroom, where a fan–coil provides the local air conditioning.

Although the outdoor-air handler eliminates the sensible-cooling load associated with ventilation, it only offsets part of the latent-cooling load because the dew point of the conditioned outdoor air is still much higher than the dew point in the space. The remaining moisture in the conditioned outdoor air must be removed from the classroom by the fan–coil. The fan–coil also must offset the entire space cooling load, both sensible and latent, because the “neutral” outdoor air does not provide a cooling effect.

At the full-load, peak dry-bulb condition, the fan–coil must supply air at 55.7°F DB (13.1°C DB) to maintain the classroom at 74°F DB (23.3°C DB). The psychrometric analysis in Figure 72 (p. 78) illustrates that delivering the outdoor air at a neutral dry-bulb temperature, but humid with respect to the

SYS-APM004-EN 77 Dehumidifying with Dedicated Outdoor Air

Figure 72. Dehumidification performance of neutral-temperature conditioned air, without overcooling, at peak dry-bulb condition

space, yields a relative humidity of 59 percent. Contrast this performance with the 52 percent relative humidity that the basic constant-volume system provides without a dedicated outdoor-air handler (p. 20).

At the part-load, peak dew-point condition, the dedicated outdoor-air handler again cools the outdoor air to 74°F DB (23.3°C DB). Due to the smaller sensible- cooling load, the fan–coil supplies the classroom with warmer air, 63°F (17.2°C), to avoid overcooling. Although the dry-bulb target is maintained, the relative humidity increases to 70 percent (Figure 73).

Figure 73. Dehumidification performance of neutral-temperature conditioned air, without overcooling, at part-load, peak dew-point condition

78 SYS-APM004-EN Dehumidifying with Dedicated Outdoor Air

Figure 74. Dehumidification performance of neutral-temperature conditioned air, Dedicated outdoor-air handler: without overcooling, on a mild and rainy day × × () Qs = 1.085 450 cfm 74– 70°F = 1,953 Btu/hr ()× × [] Qs = 1.21 0.21 m³/s 23.3– 21.2°C ()= 0.53 kW

Fan–coil: , , Qs = 12 250 – 1 953 Btu/hr × , × () = 1.085 1 500 cfm 74°F– Tsupply ∴ Tsupply = 67.7°F () Qs = 3.6– 0.53 kW ()× × [] = 1.21 0.7 m³/s 23.3°C– Tsupply ()∴ Tsupply = 19.8°C

Dehumidification performance is even worse on a mild, rainy day. The cooling coil in the dedicated outdoor-air handler shuts off because the 70°F DB (21.2°C DB) outdoor air is cooler than the thermostat setpoint. The outdoor air offsets a small portion of the classroom’s sensible-cooling load, so the fan–coil only cools the recirculated return air to 67.7°F DB (19.8°C DB). As a result, the relative humidity in the classroom increases even more, to 76 percent (Figure 74).

The preceding examples demonstrate that supplying conditioned air at a neutral dry-bulb temperature, without subcooling it to reduce its moisture content, provides less dehumidification than a system without a dedicated outdoor-air handler. In addition to lowering the sensible-heat ratio in the space, eliminating only the sensible-cooling load from the outdoor air makes it difficult for the local units to provide adequate coincidental dehumidification. Avoid designing dedicated outdoor-air systems that operate in this manner if the outdoor dew point is greater than the desired dew point indoors for any significant amount of time.

SYS-APM004-EN 79 Dehumidifying with Dedicated Outdoor Air

Dry-Bulb Temperature

Whether the dedicated outdoor-air handler delivers conditioned, dry air directly to the space, to HVAC terminals, or to other air handlers, it can be designed to condition the outdoor air to a neutral or cold dry-bulb temperature.

Dry, neutral-temperature air. A dedicated outdoor-air handler that supplies dry, neutral-temperature air (Figure 75) initially cools and dehumidifies the outdoor air; the resulting dew point usually ranges from 40°F (4°C) to 60°F (16°C), depending on the latent load in the space. The same unit then reheats the dehumidified outdoor air to approximately , usually between 70°F DB and 75°F DB (21°C DB and 24°C DB), before discharging it.

The appropriate dry-bulb temperature for a particular application best balances the following concerns:

■ Energy required to reheat the conditioned outdoor air ■ Cooling effect of the conditioned air on the sizing and operation of local HVAC terminals ■ Heating energy used by the local HVAC terminals at part-load conditions

Figure 76 shows the peak dry-bulb conditions that exist when the classroom Appendix B in this manual describes receives 450 cfm (0.21 m³/s) of conditioned outdoor air. In this case, the how to determine the conditioned-air dedicated outdoor-air handler cools and dehumidifies the outdoor air to dew point that will allow the dedicated 52°F DP (11.1°C DP), which offsets the latent cooling load in the classroom and outdoor-air handler to offset the local imposes a load of 3.0 tons (10.6 kW) on the cooling coil. This unit then reheats latent cooling loads. This appendix the air to 71°F DB (21.7°C DB) before discharging it directly into the classroom; also explains how to size the local the reheat load is 9.3 MBh (2.7 kW). HVAC terminals. ■ Because the dry-bulb temperature of the conditioned outdoor air approximates that of the classroom—74°F (23.3°C) in this example—the cooling effect is minimal. To offset the classroom’s remaining sensible-cooling load, the fan–coil

Figure 75. Dedicated outdoor-air handler that supplies dry, neutral-temperature air

80 SYS-APM004-EN Dehumidifying with Dedicated Outdoor Air

Space sensible-cooling load offset by Figure 76. Dehumidification performance of dry, neutral-temperature, conditioned outdoor dedicated outdoor-air handler: air at the peak dry-bulb condition × × () Qs = 1.085 450 cfm 74– 71°F = 1,465 Btu/hr ()× × [] Qs = 1.21 0.21 m³/s 23.3– 21.7°C ()= 0.41 kW

Fan–coil at peak dry-bulb condition: , , Qs = 29 750 – 1 465 Btu/hr × , × () = 1.085 1 500 cfm 74°F– Tsupply ∴ Tsupply = 56.6°F () Qs = 8.7– 0.41 kW ()× × [] = 1.21 0.7 m³/s 23.3°C– Tsupply ()∴ Tsupply = 13.5°C

cools 1,500 cfm (0.7 m³/s) of recirculated return air to 56.6°F DB (13.5°C DB); this requires 2.8 tons (9.8 kW) of capacity and yields a 50 percent-relative humidity.

At the part-load, peak dew-point condition (Figure 77), the dedicated outdoor-air handler again delivers 71°F DB, 52°F DP (21.7°C DB, 11.1°C DP) air. Although the reheat load remains 9.3 MBh (2.7 kW), the cooling load increases to 3.4 tons (12.0 kW). Meanwhile, fan–coil capacity modulates to raise the supply-air temperature to 63.9°F DB (17.7°C DB) and maintain the thermostat setpoint. The resulting relative humidity is 56 percent, and the cooling load on the fan–coil is 1.4 tons (4.9 kW).

Space sensible-cooling load offset by Figure 77. Dehumidification performance of dry, neutral-temperature, conditioned outdoor dedicated outdoor-air handler: air at the peak dew-point condition × × () Qs = 1.085 450 cfm 74– 71°F = 1,465 Btu/hr ()× × [] Qs = 1.21 0.21 m³/s 23.3– 21.7°C ()= 0.41 kW

Fan–coil at peak dew-point condition: , , Qs = 17 850 – 1 465 Btu/hr × , × () = 1.085 1 500 cfm 74°F– Tsupply ∴ Tsupply = 63.9°F () Qs = 5.2– 0.41 kW ()× × [] = 1.21 0.7 m³/s 23.3°C– Tsupply ()∴ Tsupply = 17.7°C

SYS-APM004-EN 81 Dehumidifying with Dedicated Outdoor Air

Space sensible-cooling load offset by Figure 78. Dehumidification performance of dry, neutral-temperature, conditioned outdoor dedicated outdoor-air handler: air on a mild, rainy day × × () Qs = 1.085 450 cfm 74– 71°F = 1,465 Btu/hr ()× × [] Qs = 1.21 0.21 m³/s 23.3– 21.7°C ()= 0.41 kW

Fan–coil on a mild, rainy day: , , Qs = 12 250 – 1 465 Btu/hr × , × () = 1.085 1 500 cfm 74°F– Tsupply ∴ Tsupply = 67.4°F () Qs = 3.6– 0.41 kW ()× × [] = 1.21 0.7 m³/s 23.3°C– Tsupply ()∴ Tsupply = 19.7°C

On a mild and rainy day (Figure 78, p. 82), the dedicated outdoor-air handler still produces 71°F DB, 52°F DP (21.7°C DB, 11.1°C DP) air; the cooling-coil load decreases to 2.0 tons (7.0 kW) and the reheat load is still 9.3 MBh (2.7 kW). In the classroom, the fan–coil modulates to supply 67.4°F DB (19.7°C DB) air, which corresponds to a cooling load of 0.9 tons (3.2 kW). The resulting relative humidity is 60 percent.

Dry, cold air. In contrast, some dedicated outdoor-air systems deliver air at a cold temperature, usually 45°F to 55°F DB (7°C to 13°C DB). These systems cool and dehumidify the outdoor air, but do not reheat it. Typically, the conditioned outdoor air is delivered directly to the mixing box of local air-handling equipment, such as blower–coils, unit ventilators, air handlers, packaged rooftop air conditioners, or dual-duct VAV terminals. Some systems, however, are designed to deliver cold ventilation air directly to occupied spaces.

Using the example classroom, Figure 79 shows the psychrometric analysis of the full-load, peak dry-bulb condition. The dedicated outdoor-air handler cools and dehumidifies the entering air to 52°F DP (11.1°C DP), then delivers the conditioned air to the classroom. The dew point of this conditioned outdoor air is low enough to offset the latent cooling load in the classroom. At the peak dry- bulb condition for Jacksonville, Florida, the cooling-coil load for the dedicated outdoor-air handler is 3.0 tons (10.6 kW). Because the ventilation air is delivered cold, there is no reheat load.

The dry-bulb temperature of the conditioned outdoor air is much cooler than the space, so it provides a greater sensible-cooling effect than dry, neutral- temperature air. It also permits the use of a fan–coil with less cooling capacity. At the full-load, peak dry-bulb condition, the fan–coil in the classroom cools

82 SYS-APM004-EN Dehumidifying with Dedicated Outdoor Air

Space sensible cooling load offset by Figure 79. Dehumidification performance of dry, cold, conditioned outdoor air at the dedicated outdoor-air handler: peak dry-bulb condition × × () Qs = 1.085 450 cfm 74– 52°F = 10,742 Btu/hr ()× × [] Qs = 1.21 0.21 m³/s 23.3– 11.1°C ()= 3.1 kW

Fan–coil at peak dry-bulb condition: , , Qs = 29 750 – 10 742 Btu/hr × , × () = 1.085 1 050 cfm 74°F– Tsupply ∴ Tsupply = 57.3°F () Qs = 8.7– 3.1 kW ()× × [] = 1.21 0.5 m³/s 23.3°C– Tsupply ()∴ Tsupply = 14.1°C

only 1,050 cfm (0.5 m³/s) of recirculated return air to 57.3°F DB (14.1°C DB), which offsets the remaining sensible-cooling load. The cooling load on the fan– coil is 1.8 tons (6.3 kW), and the resulting relative humidity is 50 percent.

At the part-load, peak dew-point condition for the classroom (Figure 80), the dedicated outdoor-air handler delivers the conditioned air, CA, at the same temperature as for the peak dry-bulb condition. Because the outdoor enthalpy is higher, however, the cooling load increases to 3.4 tons (12.0 kW). Meanwhile, the fan–coil modulates to match the thermostat setpoint by raising the supply-

Space sensible cooling load offset by Figure 80. Dehumidification performance of dry, cold conditioned outdoor at the dedicated outdoor-air handler: peak dew-point condition × × () Qs = 1.085 450 cfm 74– 52°F = 10,742 Btu/hr ()× × [] Qs = 1.21 0.21 m³/s 23.3– 11.1°C ()= 3.1 kW

Fan–coil at peak dew-point condition: , , Qs = 17 850 – 10 742 Btu/hr × , × () = 1.085 1 050 cfm 74°F– Tsupply ∴ Tsupply = 67.8°F () Qs = 5.2– 3.1 kW ()× × [] = 1.21 0.5 m³/s 23.3°C– Tsupply ()∴ Tsupply = 19.8°C

SYS-APM004-EN 83 Dehumidifying with Dedicated Outdoor Air

Space sensible cooling load offset by Figure 81. Dehumidification performance of dry, cold, conditioned outdoor air on a dedicated outdoor-air handler: mild, rainy day × × () Qs = 1.085 450 cfm 74– 52°F = 10,742 Btu/hr ()× × [] Qs = 1.21 0.21 m³/s 23.3– 11.1°C ()= 3.1 kW

Fan–coil on a mild, rainy day: , , Qs = 12 250 – 10 742 Btu/hr × , × () = 1.085 1 050 cfm 74°F– Tsupply ∴ Tsupply = 72.7°F () Qs = 3.6– 3.1 kW ()× × [] = 1.21 0.5 m³/s 23.3°C– Tsupply ()∴ Tsupply = 22.6°C

air temperature to 67.8°F DB (19.8°C DB). The resulting relative humidity is 56 percent, and the cooling load on the fan–coil is 0.67 tons (2.4 kW).

On a mild and rainy day (Figure 81), the dedicated outdoor-air handler still produces 52°F DP (11.1°C DP) air, but the cooling load decreases to 2.0 tons (7.0 kW). Fan–coil capacity modulates to supply air at 72.7°F DB (22.6°C DB), which results in a 60 percent-relative humidity and a cooling load of 0.12 tons (0.42 kW).

“Neutral” or “cold”?. Either design of a dedicated outdoor-air system can effectively control indoor humidity as long as the conditioned outdoor air is sufficiently dry. Table 9 helps to illustrate the benefits and tradeoffs of these designs by comparing the loads and airflows for the full-load, peak dry-bulb condition.

If the conditioned outdoor air is delivered directly to the space, providing it at a neutral dry-bulb temperature simplifies local comfort control. If delivered to a ceiling plenum, the neutral-temperature air prevents condensate from forming on the plenum surfaces.

Systems that provide dry, cold conditioned air avoid the energy costs of reheating the outdoor air to a neutral dry-bulb temperature. They also require less overall cooling capacity than dedicated outdoor-air systems that deliver dry, neutral-temperature air. The low dry-bulb temperature offsets part of the sensible-cooling load in the space. Similarly, when cold conditioned air is delivered to other HVAC equipment, it mixes with recirculated return air from the space; the resulting mixed-air enthalpy is lower, which in turn reduces the cooling capacity required by the coils in the HVAC terminals.

84 SYS-APM004-EN Dehumidifying with Dedicated Outdoor Air

Table 9. Comparison of coil loads and airflows for dedicated outdoor-air systems1

Dry, neutral CA2 to space Dry, cold CA2 to space

Cooling-coil load: ■ dedicated outdoor-air handler 3.0 tons (10.6 kW) 3.0 tons (10.6 kW) ■ fan–coil 2.8 tons ( 9.8 kW) 1.8 tons ( 6.3 kW) 5.8 tons (20.4 kW) 4.8 tons (16.9 kW) ■ total

Reheat-coil load: ■ dedicated outdoor-air handler 9.3 MBh (2.7 kW) 0 MBh (0.0 kW) ■ fan–coil  

Fan airflow: ■ dedicated outdoor-air handler 450 cfm (0.21 m³/s) 450 cfm (0.21 m³/s) ■ fan–coil 1,500 cfm (0.7 m³/s) 1,050 cfm (0.5 m³/s) 1,950 cfm (0.9 m³/s) 1,500 cfm (0.7 m³/s) ■ total

1 Coil loads and airflows are based on the full-load, peak dry-bulb condition for a classroom in Jacksonville, Florida. The classroom holds 30 occupants, has an area of 10,000 ft³ (283 m³), and a target condition of 74°F DB (23.3°C DB), 50% RH. 2 CA = Conditioned outdoor air provided by a dedicated outdoor-air handler/system

Compared with “neutral-temperature” designs, dedicated outdoor-air systems that deliver cold air provide additional benefits:

■ Less fan airflow and, therefore, lower fan-energy consumption. (Neutral- temperature conditioned air provides the space with less sensible cooling, so the local units must provide more cooling to achieve the same effect.) ■ Smaller HVAC terminals, which can either lower the initial cost and increase usable floor space or provide an acoustical benefit by keeping the same- sized cabinet and operating the fan at a lower speed. (In the examples summarized in Table 9, the “cold air” design represents a 30 percent reduction in both cooling-coil capacity and fan airflow.)

SYS-APM004-EN 85 Dehumidifying with Dedicated Outdoor Air

Application Considerations

Humidity Control during Unoccupied Periods

As discussed in a previous chapter, “Dehumidification Primer” (p. 17), around- the-clock control of humidity can greatly reduce the risk of microbial growth on building surfaces and furnishings.

Application considerations ■ Delivering conditioned outdoor air directly to the space permits after-hours Including an air-to-air energy recovery humidity control without operating the local HVAC terminals. If the device to precondition the outdoor air dedicated outdoor-air unit includes a return-air path, close the outdoor-air can help justify the cost of routing the damper and open the return-air damper to avoid conditioning unneeded building exhaust back to the dedicated outdoor air. (The ventilation requirement is significantly less after-hours, outdoor-air handler (which provides the when there are few or no occupants in the building.) return-air path during after-hours operation). Outdoor-air preconditioning If the outdoor air is delivered to local HVAC terminals, then the fans in the is discussed on pp. 98–99. ■ local units must operate in conjunction with the dedicated outdoor-air unit. ■ Using a packaged, direct-expansion air conditioner to condition the outdoor air eliminates after-hours operation of the central, chilled water plant. ■ To control after-hours operation of the dedicated outdoor-air unit, install a humidity sensor in the space. Use the sensor to enforce the maximum humidity limit for periods when the space is unoccupied.

Building Pressurization

Along with ventilation, the dedicated outdoor-air handler can provide makeup air for local exhaust systems (such as restroom exhaust fans and kitchen exhaust hoods) and combustion processes. In some buildings—either due to oversight or by design—the dedicated outdoor-air unit is turned off during unoccupied periods, while local exhaust fans and processes continue to operate. This creates negative pressure within the building, and causes unconditioned outdoor air to infiltrate the building envelope. Infiltration of humid air can raise the humidity indoors; it can also lead to condensation within building walls and on cold surfaces in the occupied space.

While the dedicated outdoor-air unit operates, a central relief fan directly controls building pressure. Based on input from a building-pressure sensor, the capacity of the relief fan modulates to maintain the desired difference between the indoor and outdoor static pressures.

Application considerations ■ If the system control scheme disables ventilation when the building is unoccupied, make sure that it also turns off all local exhaust fans and

86 SYS-APM004-EN Dehumidifying with Dedicated Outdoor Air

combustion processes. Provide a manual override to permit operation during after-hours cleaning. ■ Monitor indoor humidity at all times and operate the dedicated outdoor-air unit whenever humidity rises to an unacceptable level. ■ Wind, variable operation of local exhaust fans, and “stack effect” can cause pressure fluctuations inside the building, even when the HVAC system is Refer to Building Pressurization properly balanced. To prevent infiltration and avoid overpressurization Control, Trane applications engineering during economizer operation, consider implementing a system design that manual AM-CON-17, to learn more directly controls building pressure. about the design and control of HVAC ■ systems that regulate building If the dedicated outdoor-air unit serves multiple spaces with different pressure. ■ occupancy schedules, consider providing a means of varying the intake airflow to meet changing ventilation requirements and to prevent overpressurization when some of the local exhaust fans are turned off.

Economizer Cooling

Designers usually size dedicated outdoor-air handlers to condition only the minimum amount of outdoor air required for ventilation. When this is the case, the supply fan in the dedicated outdoor-air handler and the ductwork are too small to take advantage of economizer-cooling opportunities by bringing in more outdoor air. There are other ways to provide an economizer cycle, however, depending on where the conditioned ventilation air is delivered and on the other equipment in the system.

Separate economizer-air path

When ventilation air is delivered directly to the occupied spaces, economizer cooling can be provided via outdoor-air intakes at the local HVAC terminals. Classroom unit ventilators and fan–coils, which are installed along perimeter walls, simplify the addition of a separate outdoor-air path for economizer cooling. When outdoor conditions are suitable, the local outdoor-air damper modulates between fully closed and fully open to control space temperature. The local outdoor-air damper remains closed whenever the outdoor air is too warm or too cold for economizer operation.

The dedicated outdoor-air unit typically is operated to provide conditioned air during both “economizer” and “non-economizer” modes. This practice permits the minimum outdoor airflow to be monitored and recorded during all occupied hours.

When ventilation air is delivered to the mixing boxes of other air handlers rather than to each space, each local unit can be configured to include two outdoor-air paths (Figure 82, p. 88). One path, from the dedicated outdoor-air unit, provides minimum outdoor airflow for ventilation. The second path

SYS-APM004-EN 87 Dehumidifying with Dedicated Outdoor Air

Figure 82. Dedicated outdoor-air system with two outdoor-air paths for local economizing

permits economizer cooling. For example, a dedicated outdoor-air unit can deliver conditioned outdoor air to the mixing boxes of several packaged rooftop air conditioners. Each mixing box includes two outdoor-air dampers: one for conditioned ventilation air and the other for untreated “economizer” air.

During normal (mechanical) cooling or heating operation, the economizer damper is closed. During economizer cooling, the economizer damper modulates as needed between fully closed and fully open to offset as much of the cooling load as possible. If the system includes a device that measures outdoor airflow, the damper in the ventilation path typically remains open, regardless of the economizer damper position. This makes it possible to monitor and record the minimum outdoor airflow for ventilation during all occupied hours.

Application considerations ■ Local economizer operation requires some form of building-pressure control. Bringing more outdoor air into the building for economizer cooling requires more relief airflow. ■ Providing two outdoor-air paths to each local air handler may require additional ductwork and controls.

88 SYS-APM004-EN Dehumidifying with Dedicated Outdoor Air

Implications of ASHRAE Standard 90.1. Section 6.3.1 in the 2001 version Weather permitting, outdoor air is a of the standard defines requirements for HVAC systems with economizers. In source of for the building. the case of airside economizers, the section specifies damper characteristics, But as the air warms, it will eventually how to control the economizer dampers, and how to relieve air from the increase energy usage by creating a building to prevent overpressurization. It also prescribes how and when to shut cooling load. To avoid the unwanted off the economizer cycle when the weather is no longer suitable for free load, the high-limit shutoff disables the cooling; high-limit shutoff requirements vary by climate and control method. economizer and the HVAC system provides the minimum outdoor airflow Although compliance with these requirements minimizes energy use, it may for ventilation. not provide acceptable humidity control at all operating conditions or in all Standard 90.1 identifies five common climates. For instance, Standard 90.1 permits the use of the fixed-dry-bulb- types of high-limit shutoff for the temperature high limit in any climate. When applied in a humid locale, the high control of airside economizers: limit can only disable economizer cooling when the outdoor temperature is 65°F DB (18.4°C DB) or higher. A previous example (p. 31) demonstrated what ■ fixed dry-bulb temperature can happen in a constant-volume system when the shutoff setting is too high; in ■ differential dry-bulb temperature that case, the indoor humidity reached 75 percent! ■ fixed enthalpy Understand that Standard 90.1 does not mandate an economizer if the cooling ■ differential enthalpy capacity of each local HVAC unit is less than either 65,000 Btu/hr (19 kW) or ■ electronic hybrid enthalpy/ 135,000 Btu/hr (38 kW), depending on climate—which is typically true of temperature systems with dedicated outdoor-air handlers and local HVAC terminals. If you Not all control methods are appropriate include an economizer anyway, the control requirements in Section 6.3.1 will in all climates. The right choice will not apply. Therefore, you can devise a control strategy that more effectively depend on climate, hours of occupancy, controls humidity. method of air distribution (constant- or variable-volume), and operating-cost When using fixed dry-bulb temperature for economizer control in a constant- savings ■ volume system, choose a shutoff setting that is low enough to avoid bringing mild-but-humid outdoor air into the building. For better dehumidification, however, consider a more sophisticated control method that uses an indoor humidity sensor to disable the economizer whenever the relative humidity in the space exceeds 60 percent.

In constant-volume applications for which Standard 90.1 does require an economizer, investigate control methods such as fixed enthalpy or electronic hybrid enthalpy.

Waterside economizing

Waterside economizers can satisfy many of the economizer requirements For more information about using a defined in Section 6.3.1 of ASHRAE Standard 90.1–2001 and in local codes. plate-and-frame heat exchanger to provide free cooling in a chilled water One way to provide “free” cooling in a chilled water system is to install a plate- system, refer to the Trane Engineers and-frame heat exchanger between the chilled water and condenser-water Newsletter titled “A New Era of Free loops, upstream of the chillers. When the leaving-tower water is cold enough, Cooling” (ENEWS-20/3). It is archived circulating it through the heat exchanger will satisfy at least part of the cooling in the “newsletters” section of load and reduce or eliminate the need to operate the chillers. www.trane.com/commercial. ■

SYS-APM004-EN 89 Dehumidifying with Dedicated Outdoor Air

Figure 83. Waterside-economizer coil in a water-source heat pump

Water-source heat pump (WSHP) systems are candidates for waterside economizing, too. For much of the year, some of the heat pumps provide cooling while others simultaneously provide heating. During the winter, the perimeter heat pumps extract heat from the water loop. The cool water then can be used by the interior heat pumps to offset the heating loads generated by lights, people, and office equipment. Diverting the cool loop water through a separate economizer coil (Figure 83) allows the heat pump to cool the entering air without operating the compressor. Because the economizer coil is situated upstream of the main coil, it can also supplement mechanical cooling when possible.

Reset Control Strategies

Many dedicated outdoor-air systems are designed to deliver conditioned outdoor air at a constant dry-bulb temperature and at a dew point that does not exceed the setpoint. This control approach is simple because it allows the dedicated outdoor-air unit to operate independently of the local HVAC terminals. It also maintains the indoor humidity at or below the upper limit.

Implementing a reset control strategy can help minimize the additional energy cost of separately treating outdoor air. The reset strategy can be based on dew point, dry-bulb temperature, or airflow.

Reset the supply-air dew point

When sensible-cooling loads in the space are high, the local HVAC terminals may offset part of the dehumidification load by performing latent cooling. When latent-cooling loads in the space are low, the dehumidification load is less, too. In either situation, the resulting indoor humidity will be lower than the upper limit, which means that the dew-point temperature for the conditioned outdoor air can be reset upward.

90 SYS-APM004-EN Dehumidifying with Dedicated Outdoor Air

Dedicated outdoor-air handler serves a single space Install the humidity sensor either in the space or in the return-air path. When the humidity in the space drops below the upper limit, the sensor resets the dew point of the conditioned air upward only enough to enforce the maximum humidity and, in doing so, reduces cooling-energy consumption. To enforce the maximum humidity limit when the space humidity rises, the sensor resets the dew point of the conditioned air downward.

Dedicated outdoor-air handler serves multiple spaces One option is to install a single humidity sensor in a “representative” space or in a common return-air path. When the humidity of the monitored air is low enough, the sensor resets the conditioned-air dew point upward to reduce cooling-energy use. Although economical to install, using a single sensor to represent several spaces necessarily reflects the “average” humidity. Depending on the load characteristics of the spaces, the actual humidity may exceed the limit in one space and be less than the limit in another.

(Figure 84) To resolve this multiple-space problem, install a humidity sensor in each space that is served by the dedicated outdoor-air handler, or at least in those spaces where high humidity is expected. A system (BAS) monitors all of the spaces to determine the critical space with the highest humidity. Then, the BAS sends a signal to the dedicated outdoor-air unit to reset the leaving-air dew point only enough to enforce the maximum humidity limit in the critical space. This approach optimizes system performance: By responding to actual conditions, the system maintains the humidity at or below the limit in all spaces while consuming no more cooling energy than necessary.

Figure 84. Control arrangement for fan–coil system with dedicated outdoor-air

SYS-APM004-EN 91 Dehumidifying with Dedicated Outdoor Air

Reset the supply-air dry-bulb temperature

When a dedicated outdoor-air handler delivers air that is cooler than the space Use particular care if the system setpoint, it reduces the sensible-cooling load on the local HVAC terminals. If the discharges conditioned outdoor air into space load is already low, the terminal may need to add heat to avoid the ceiling plenum, near the terminals. overcooling the space. Resetting the dry-bulb temperature of the conditioned To prevent condensation on cold outdoor air by modulating the reheat capacity of the dedicated outdoor-air unit surfaces, the dry-bulb temperature of minimizes the reheat energy used by the local terminals. the conditioned outdoor air should be well above the dew point of the air in Dedicated outdoor-air handler serves a single space the plenum. ■ To reset the dry-bulb temperature of the conditioned outdoor air, install a thermostat in the space to modulate the reheat capacity of the dedicated outdoor-air handler. Commonly used in hotel–hallway applications (discussed on p. 93), the dedicated outdoor-air handler enforces the humidity limit in the space without overcooling it.

Dedicated outdoor-air handler serves multiple spaces To most effectively implement a reset strategy, use a building automation system to monitor all of the HVAC terminals and to identify the “critical space”—that is, the space with the lowest sensible-cooling load. If the terminals contain chilled water coils, the BAS monitors the position of the chilled-water control valves. The space with the lowest load is served by the terminal whose control valve is the most closed. (If the terminals contain refrigerant coils, the BAS monitors compressor operation; in that case, the critical space is the one with the shortest compressor run-time.)

Based on a signal from the BAS, the dedicated outdoor-air handler then modulates its reheat capacity, resetting the dry-bulb temperature of the conditioned air upward only enough so that the terminal with the lowest sensible-cooling load operates at near-zero cooling capacity. Notice that reset is accomplished by modulating reheat capacity, not cooling capacity. The premise is that the dew point of the conditioned air is controlled independently to meet the humidity-control requirements for the spaces. This optimization strategy provides conditioned outdoor air that offsets as much of the sensible-cooling load in each space as possible without requiring any of the terminals to activate their heating coils. This can be beneficial if the dedicated outdoor-air handler recovers reheat energy from some other part of the system, thereby avoiding the use of “new” energy for reheat, both at the dedicated outdoor-air unit and at the terminals.

Dynamically resetting the dry-bulb temperature of the conditioned outdoor air is particularly beneficial in a two-pipe system…especially if a packaged, DX air conditioner serves as the dedicated outdoor-air unit. The packaged DX unit can delay the changeover from cooling to heating because it is not connected to the chiller or ; in doing so, it saves energy and improves occupant comfort. When the two-pipe system is in cooling mode (boiler off), the dry-bulb temperature supplied by the dedicated outdoor-air unit is reset upward to maintain a heating load of zero on the terminal in the critical space. When the

92 SYS-APM004-EN Dehumidifying with Dedicated Outdoor Air

Figure 85. Conditioned outdoor air delivered to a common hallway

two-pipe system is in heating mode (chiller off), the dry-bulb temperature is reset downward to maintain a cooling load of zero on the terminal in the critical space.

If one dedicated outdoor-air handler serves multiple spaces with widely differing loads—for example, interior spaces with nearly constant full-load cooling and perimeter spaces that require heating—then for the interior spaces, it may be necessary to size the terminals as if the dedicated outdoor-air unit delivers neutral-temperature air. This practice may prevent downsizing of some of the terminals, so the benefits of this strategy must be weighed against the first-cost impact of terminal-unit sizing.

Dedicated outdoor-air handler serves a common hallway In some applications, such as hotels, the dedicated outdoor-air handler ventilates an entire floor of guest rooms by delivering conditioned outdoor air to a central hallway (Figure 85). Local exhaust then draws the ventilation air into each guest room from the hallway. In addition to providing outdoor air for ventilation, the dedicated outdoor-air handler also controls the temperature and humidity in the hallway.

In this case, size the dedicated outdoor-air unit based on either the ventilation airflow requirement of the spaces it serves, or the sensible load in the hallway, whichever is the highest. A humidity sensor installed in the hallway resets the dew point of the conditioned outdoor air, maintaining the humidity in the hallway at or below the desired limit. A thermostat, also installed in the hallway, resets the dry-bulb temperature of the conditioned outdoor air as needed to maintain the setpoint.

SYS-APM004-EN 93 Dehumidifying with Dedicated Outdoor Air

Reset outdoor-air quantity

Most dedicated outdoor-air handlers are designed to deliver a constant flow of

“Using CO2 for Demand-Controlled conditioned outdoor air; however, applications in which the population varies Ventilation,” Trane Engineers widely during normal building operation may benefit from variable intake Newsletter (volume 31, number 3), airflow. Automatically adjusting the quantity of conditioned outdoor air to describes the use of carbon-dioxide match the system’s current ventilation requirements can reduce operating costs sensors to estimate the population in the without sacrificing indoor air quality. building and reset the amount of outdoor air introduced for ventilation. With separate dampers and a method for fan-capacity control, a dedicated You can find it in the “newsletters” outdoor-air system can be designed to vary the outdoor airflow to each space ■ section of www.trane.com/commercial. based on the estimated population. Population estimates can originate from

occupancy sensors (occupied or unoccupied), (CO2) sensors, or time-of-day schedules in a building automation system (BAS).

Reheating Conditioned Air with Recovered Heat

Dedicated outdoor-air handlers that reheat the outdoor air, after cooling it to a low dew point for dehumidification, can use either “new” energy (gas, electricity, hot water, or steam) or sensible heat recovered from another part of the system.

Section 6.3.2.3 of the U.S. energy standard, ASHRAE 90.1–2001 (see excerpt on This exception is the subject of p. 8) defines the circumstances, or “exceptions,” in which new-energy reheat is Example 6-YY, found in the User’s permitted. For example, Exception A permits the use of new-energy reheat if Manual that accompanies ASHRAE the system first reduces the supply-air volume to the minimum ventilation rate Standard 90.1. Both publications are specified by ASHRAE Standard 62. Because the quantity of air typically supplied available from the bookstore on the by the dedicated outdoor-air unit already equals the minimum ventilation ■ ASHRAE Web site (www.ashrae.org). airflow required by Standard 62, its airflow cannot be reduced further. Therefore, in most cases, Standard 90.1 will not prohibit new-energy reheat in dedicated outdoor-air units that deliver neutral-temperature air for ventilation.

Recovering heat for reheat may be desirable even if it is not required by Standard 90.1 because it reduces system operating costs. Sources of recoverable heat include:

■ Hot refrigerant in a refrigeration system ■ Condenser water in a water-cooled, chilled water system ■ Another air stream, or another location within the same air stream

Refrigerant heat recovery

Systems that include a packaged, DX-type air conditioner as the dedicated outdoor-air unit are excellent candidates for refrigerant-heat recovery (Figure 86). Because reheat is required while the cooling coil dehumidifies the outdoor air, recovering heat from the refrigeration system saves reheat energy. A condenser-reheat coil transfers sensible heat from the hot refrigerant to the

94 SYS-APM004-EN Dehumidifying with Dedicated Outdoor Air

Figure 86. Packaged DX, dedicated outdoor-air unit with condenser-reheat coil

air downstream of the cooling coil. Refrigerant heat recovery is discussed further on pp. 57–58.

Condenser-water heat recovery

Sensible heat also can be recovered from the warm water leaving the condenser of a water-cooled chiller. Instead of pumping the condenser water to the cooling tower, it can be circulated to one or more dedicated outdoor-air handlers and used for reheat. Any water-cooled chiller can provide sensible heat for reheat. Also, water is more easily distributed to multiple units than refrigerant. For examples of common system configurations, see pp. 53–56.

Air-to-air heat recovery/transfer

Sensible heat for reheat can be recovered from another air stream—or another Consult Air-to-Air Energy Recovery in location in the same air stream—using an air-to-air heat exchanger, such as a HVAC Systems, Trane applications coil loop, a fixed-plate heat exchanger, a heat pipe, or a rotary sensible-heat engineering manual SYS-APM003-EN, exchanger (heat wheel). When the air-to-air heat exchanger provides reheat in a for more information about using an dedicated outdoor-air system, it can be applied in one of two configurations: air-to-air heat exchanger for reheat in a series or parallel. dedicated outdoor-air system. Topics include series and parallel Series configuration configurations and a coil loop with Figure 87 (p. 96) shows a dedicated outdoor-air handler that includes a fixed- three coils. ■ plate heat exchanger. The heat exchanger, which could be any type of sensible- energy recovery device, is applied in a series, or “wrap-around,” configuration. The heat exchanger removes sensible heat from the air upstream of the cooling/dehumidifying coil and transfers it downstream to reheat the air leaving the coil.

Applying the air-to-air heat exchanger in a series arrangement conserves “new” energy that would otherwise be used to reheat the outdoor air. It also reduces the cooling load by precooling the outdoor air upstream of the

SYS-APM004-EN 95 Dehumidifying with Dedicated Outdoor Air

Figure 87. Series heat-recovery arrangement for reheating conditioned outdoor air

In the strict sense of the term, “heat recovery” does not include the series arrangement of an air-to-air heat exchanger. Rather than capturing heat from another source (exhaust air, for example), the series arrangement moves heat from one location to another within the same air stream. ■

dehumidifying coil. In this application, reheat is required whenever the cooling coil dehumidifies, even at design conditions, because the conditioned outdoor air is delivered at a “neutral” dry-bulb temperature. The precooling effect provided by the heat exchanger reduces the design load on the cooling coil; it also may permit downsizing of the cooling plant.

During cold weather, when the dedicated outdoor-air unit must heat the entering air, the cooling coil is turned off. Because the air-to-air heat exchanger is arranged in series with the cooling coil, it provides no benefit. An arrangement that bypasses the cooling coil and air-to-air heat exchanger during cold weather would save fan energy by reducing the airside pressure drop.

Parallel configuration Figure 88 shows the same dedicated outdoor-air handler, this time with the air- to-air heat exchanger applied in a parallel arrangement. The heat exchanger recovers sensible heat from the exhaust air and releases it downstream of the cooling/dehumidifying coil to reheat the leaving air. (This example shows a fixed-plate heat exchanger, but any type of air-to-air, sensible-heat exchanger will work.)

Like the series arrangement, applying the air-to-air heat exchanger in a parallel arrangement saves the “new” energy that would otherwise be needed to reheat the air. Unlike the series arrangement, however, the parallel arrangement does not reduce the load on the cooling coil, nor does it enable downsizing of the cooling plant. This is because the heat recovered from the exhaust air would be rejected from the building anyway.

96 SYS-APM004-EN Dehumidifying with Dedicated Outdoor Air

Figure 88. Parallel heat-recovery arrangement for reheating conditioned outdoor air

Another important advantage characterizes the parallel arrangement: When it is cold outside and the dedicated outdoor-air handler must heat the entering outdoor air, the parallel-arranged, air-to-air heat exchanger can recover heat from the exhaust air to warm the entering outdoor air. This lowers the operating costs for heating and may allow downsizing of the heating plant.

Cold climates may require a means of frost prevention for the air-to-air heat exchanger during full-recovery, heating operation. If the dedicated outdoor-air handler includes a chilled-water cooling coil, it may also require freeze protection because the heat exchanger adds heat downstream of the coil.

Series or parallel? In dedicated outdoor-air handlers that supply neutral-temperature air, both A coil loop that consists of three coils arrangements reduce the “new” energy required for reheat. The series combines both arrangements of air-to- arrangement can transfer more heat whenever the outdoor air is warmer than air energy recovery. The series the exhaust air. It also saves cooling energy, which may allow downsizing of the arrangement is active when the cooling cooling plant. By comparison, the parallel configuration depends less on coil operates, saving cooling energy and outdoor conditions and transfers more heat whenever the outdoor air is cooler reheat energy. When the cooling coil than the exhaust air. (This is because the temperature of the exhaust air is turns off and the heating coil operates, relatively constant.) The parallel configuration also can operate during the the parallel arrangement becomes heating mode, which saves heating energy and may allow downsizing of the active, saving heating energy. ■ heating plant.

Implementing the parallel configuration requires additional ductwork to bring most of the building’s exhaust air back to a central location; the series configuration does not. Given the difference in initial cost, air-to-air heat recovery is commonly implemented using the series arrangement unless the system includes an additional energy-recovery device (for example, a total-

SYS-APM004-EN 97 Dehumidifying with Dedicated Outdoor Air

Figure 89. “Dual-function” air-to-air energy recovery

energy wheel) for outdoor-air preconditioning. Figure 89 illustrates a dedicated outdoor-air handler that includes “dual-function” energy recovery.

Ultimately, the right choice for a given project will depend on the balance of initial cost, cooling- and heating-energy savings, and the increased fan energy imposed by the additional static-pressure drop through the heat exchanger.

Preconditioning Outdoor Air with Recovered Energy

The previous section discussed the use of air-to-air energy recovery for Consult Air-to-Air Energy Recovery in reheating conditioned outdoor air. This type of energy recovery also is suited HVAC Systems, Trane applications for preconditioning outdoor air as it enters the dedicated outdoor-air handler. engineering manual SYS-APM003-EN, to learn more about using an air-to-air Figure 90 illustrates a dedicated outdoor-air handler that includes a total-energy heat exchanger to precondition outdoor wheel to transfer energy between the outdoor air and exhaust air. During the air. The manual discusses compliance summer, when the outdoor air is hot and humid, the total-energy recovery with ASHRAE Standard 90.1 and device can precool and predry the outdoor air by transferring both sensible heat various energy-recovery technologies; it and moisture to the exhaust air. During the winter, when the outdoor air is cold also outlines control modes for and dry, the same device can preheat and prehumidify the outdoor air by dedicated outdoor-air systems. ■ recovering both sensible heat and moisture from the exhaust air and transferring it to the entering outdoor air.

Substituting a sensible-energy recovery device (such as a coil loop, fixed-plate heat exchanger, heat pipe, or sensible-energy wheel) may reduce the initial cost of the device, but it precludes the opportunity to predry or prehumidify the entering outdoor air.

Using air-to-air energy recovery to precondition the outdoor air can reduce the operating cost of the dedicated outdoor-air handler; it may even allow downsizing of the mechanical cooling, heating, and humidification equipment.

98 SYS-APM004-EN Dehumidifying with Dedicated Outdoor Air

Figure 90. Parallel heat-recovery arrangement for preconditioning outdoor air

As demonstrated on pp. 27–29, however, a total-energy recovery device does little to control space humidity. A cold coil (or an active desiccant system) is needed to dehumidify the air to a dew point that is drier than the air in the space.

Any sensible- or total-energy recovery device can precondition outdoor air. The choice between them usually is based on the initial cost and energy-savings potential in the specific application. In either case, the parallel arrangement required for preconditioning also requires additional ductwork to bring most of the building-exhaust air back to a central location.

Control of the energy-recovery device Integrating the operation of the energy-recovery device into the control sequences for the dedicated outdoor-air handler (and the rest of the HVAC system) is critical to realize the anticipated cost savings and to reduce the payback period. Although the simplest strategy cycles the device with the fan, this approach does not maximize energy savings.

At many outdoor conditions, the capacity of the energy-recovery device must modulate or the device must be turned off to avoid transferring unwanted moisture or sensible heat to the entering outdoor air. This is especially true for dedicated outdoor-air units that deliver dry, cold air rather than reheating the conditioned air to a neutral temperature.

Operating the energy-recovery device at full capacity under these conditions, which could represent 15 percent to 25 percent of the annual operating hours, will waste energy by operating the wheel and the dedicated outdoor-air handler in opposition to each other. ■

SYS-APM004-EN 99 Afterword

The dehumidification performance of a cold-coil HVAC system hinges on the ability of the cooling coil to reduce the dew point of the passing air below that of the air in the space. Single-zone, constant-volume, mixed-air systems are most often used to condition entering outdoor air; they also can be the most problematic when it comes to dehumidification, particularly at part-load conditions. VAV systems generally provide effective, coincidental dehumidification over a range of indoor load conditions.

Contrary to popular belief, high indoor humidity can be an issue in nearly all geographic locations, not just in areas where hot, humid conditions prevail. Indoor humidity typically depends as much on the sensible- and latent-cooling loads in the space, the type of HVAC system, and the method of controlling that system as it does on outdoor conditions. Whenever high relative humidity exists at or near a cold, porous surface, moisture adsorption increases and moisture-related problems (such as increased health risks from mold growth and premature replacement of equipment and furnishings) become likely.

When properly designed and controlled, however, the HVAC system can significantly reduce the moisture content of indoor air. Increased humidity in the space can also result from higher-than-necessary supply airflow…either due to an overly conservative estimate of the space sensible-cooling load or the use of oversized, packaged DX equipment. When selecting cooling equipment, especially for constant-volume applications, exercise particular care to avoid oversizing the design airflow. Analyze system dehumidification performance at both full- and part-load conditions, and consider the advantages and disadvantages of each system enhancement. The right choice for a given project depends on the climate, building use, available budget, and operating- cost goals. ■

100 SYS-APM004-EN Appendix A: Psychrometric Analysis

An imaginary, elementary-school classroom in Jacksonville, Florida, provides the context for the examples of dehumidification performance that appear throughout this manual. In the chapter titled “Dehumidifying with Constant-Volume Mixed Air” (pp. 19–60), the classroom is air conditioned by the basic single-zone, constant-volume HVAC system shown in Figure 91.

A thermostat compares the dry-bulb temperature in the classroom to the setpoint. The system then modulates the capacity of the cooling coil accordingly, which adjusts the supply-air temperature until the temperature in the classroom matches the setpoint.

As the coil cools the passing air stream, it reduces the air’s moisture content via condensation on the coil surface. The thermostat determines the coil- surface temperature and, therefore, the supply-air temperature. Modulating the coil to increase its sensible-cooling capacity makes the coil surface colder, which causes more condensation and (therefore) more dehumidification. Reducing coil capacity results in a warmer coil that provides less dehumidification.

Predicting the dehumidification performance of an HVAC system entails a The peak dew-point condition does not psychrometric analysis of its operation, both at full-load and at the range of necessarily represent the worst-case part-load conditions. This section demonstrates the analysis process by condition for humidity control. It simply examining the performance of this basic system at two operating points… is an easy “test case” for analyzing part- load dehumidification performance. ■ ■ Full load, peak dry bulb: 96°F DB, 76°F WB (35.7°C DB, 24.5°C) ■ Part-load, peak dew point: 76°F DP, 84°F DB (24.6°C DP, 28.8°C DB)

The outcome of this two-part analysis is summarized on pp. 20–21.

Figure 91. Basic, constant-volume HVAC system

SYS-APM004-EN 101 Appendix A: Psychrometric Analysis

Full-Load, Peak Dry-Bulb Condition

Step 1: Plot the outdoor condition (OA) and the desired indoor The examples throughout this manual condition (RA) on a psychrometric chart. The ASHRAE Handbook– are based on “0.4 percent” data from Fundamentals lists the 0.4 percent peak-dry-bulb condition for Jacksonville, the ASHRAE climatic data tables. Florida, as 96°F DB, 76°F WB (35.7°C DB, 24.5°C WB). For this example, the This percentage indicates that the desired condition in the classroom is 74°F DB (23.3°C DB) with a relative temperature is likely to equal or exceed humidity of 50 percent. See Figure 92. the design value for 35 hours each year. Some design engineers choose to use Step 2: Calculate the sensible- and latent-cooling loads in the space. more extreme conditions; others base At the peak dry-bulb condition, the sensible and latent loads calculated for the their designs on the “1 percent” or “2 percent” values, which represent classroom are 29,750 Btu/hr (8.7 kW) and 5,250 Btu/hr (1.5 kW), respectively. more hours. ■ These loads yield a space sensible-heat ratio (SHR) of 0.85, which means that 85 percent of the classroom’s cooling load is sensible and 15 percent is latent.

29, 750 Btu/hr SHR = ------= 0.85 29, 750 Btu/hr + 5, 250 Btu/hr

§·8.7 kW SHR = ------= 0.85 ©¹8.7 kW + 1.5 kW

Step 3: Calculate the temperature and quantity of air required to offset the sensible-cooling load in the space. Having already determined the space sensible-cooling load and the desired indoor temperature, either: define the supply airflow and calculate the supply-air

Figure 92. Outdoor (peak dry bulb) and indoor conditions for Jacksonville classroom

102 SYS-APM004-EN Appendix A: Psychrometric Analysis

temperature, or choose the supply-air temperature and calculate the required

In the Qs equations at right, 1.085 and airflow. In either case, solve following equation: 1.21 are derived from the properties of air; they are not constants. At the × × () Qs = 1.085 Vsa Tsp – Tsa “standard” air condition, which is 69°F ()× × [] (21°C) dry air at sea level, the product of Qs = 1.21 Vsa Tsp – Tsa density, specific heat, and the conversion for time (I-P units only) where,

equals 1.085 (1.21). A different air Qs = sensible-cooling load in the space, Btu/hr (kW)

condition or elevation will result in a Tsa = dry-bulb temperature of the supply air, °F (°C) ■ different value. Tsp = desired dry-bulb temperature for the space, °F (°C)

Vsa = supply airflow delivered to the space, cfm (m³/s)

For this example, assume that the supply airflow for the classroom is nine air changes per hour. The volume of the classroom is 10,000 ft³ (283 m³), so the required supply airflow is 1,500 cfm (0.7 m³/s):

10, 000 ft³ ⁄ air change × 9 air changes⁄ hr V = ------= 1, 500 cfm sa 60 min⁄ hr

§·283 m³⁄ air change × 9 air changes⁄ hr V = ------= 0.7 m³/s ©¹sa 3, 600 sec⁄ hr

To offset the sensible-cooling load and satisfy the thermostat setpoint by maintaining 74°F DB (23.3°C DB) in the classroom, the supply airflow must be delivered at 55.7°F DB (13.1°C DB):

× , × (), ∴ Qs = 1.085 1 500 cfm 74°F– Tsa = 29 750 Btu/hr Tsa = 55.7°F ()× × []∴ Qs = 1.21 0.7 m³/s 23.3°C– Tsa = 8.7 kW Tsa = 13.1°C

Space- and return-air temperatures are essentially the same in systems that Step 4: Calculate the entering-air condition for the cooling coil. The use HVAC units installed within the air entering the cooling coil is a mixture of outdoor air and recirculated air occupied space (fan–coils or classroom returning from the space. For this example, assume that the condition of the unit ventilators, for example). In other return air is the same as the air in the classroom—74°F DB (23.3°C DB) and a system configurations, the air returning relative humidity of 50 percent. from the space picks up sensible heat and enters the air handler at a warmer According to ASHRAE Standard 62–2001, properly ventilating the classroom temperature than that in the occupied requires 15 cfm (8 L/s) of outdoor air per person. Given the design occupancy space. This is common in systems where of 30 people, and because the system only serves this classroom, 450 cfm the return air travels through an open ceiling plenum or passes through light fixtures.

To simplify this example, we ignored return-air heat gain in our analysis. ■

SYS-APM004-EN 103 Appendix A: Psychrometric Analysis

Figure 93. Plotting the entering mixed-air condition for the cooling coil

(0.21 m³/s) of the 1,500 cfm (0.7 m³/s) supply airflow must be outdoor air. The resulting dry-bulb temperature of this mixture is 80.6°F (26.8°C):

()450 cfm × 96°F + ()1, 050 cfm × 74°F T = ------= 80.6°F ma 450 cfm + 1, 050 cfm

§·[]0.21 m³/s × 35.7°C + []0.5 m³/s × 23.3°C T = ------= 26.8°C ©¹ma 0.21 m³/s + 0.5 m³/s

When plotted on a psychrometric chart (Figure 93), the mixed-air (MA) condition lies on the straight line that connects the outdoor-air (OA) and return-air (RA) conditions. In this case, the condition of the mixed air, which will enter the cooling coil, lies at the intersection of OA–RA and the line that corresponds to the calculated dry-bulb temperature, 80.6°F (26.8°C).

Step 5: Determine the leaving-coil condition of the supply air. Using the curvature of the adjacent “coil curves” as a guide, draw a curve from the mixed-air condition (MA) until it intersects the supply-air temperature (SA) calculated in Step 3—in this case, 55.7°F (13.1°C).

Note: The Trane psychrometric chart includes a series of curves depicting the changes in dry-bulb and dew-point temperatures that occur as air passes through a “typical” cooling coil. Derived from laboratory tests, these “coil curves” represent various coil geometries at different air and coolant temperatures. For a more precise prediction of performance, use software that accurately models the actual coil geometry.

104 SYS-APM004-EN Appendix A: Psychrometric Analysis

Figure 94. Using supply-air temperature and the sensible-heat ratio to find the resulting air condition in the space

Step 6: Determine the air condition of the space. The space sensible- heat ratio (SHR) determines the resulting air condition in the space for a given supply-air temperature:

a Establish the slope of the SHR line by drawing a line that connects the appropriate value on the SHR scale—0.85, in this case—with the index point in the center of the chart. b Using the SHR line as a reference, draw a parallel line from the supply- air condition (SA) to the line representing the desired space dry-bulb temperature—that is, from 55.7°F DB (13.1°C) to 74°F (23.3°C).

(Figure 94) Although the target humidity for the classroom is 50 percent, the 0.85 SHR and 55.7°F DB (13.1°C) supply air will yield a relative humidity of roughly 52 percent. To “close the loop,” we must repeat the psychrometric analysis (Steps 1–6) until successive iterations produce the same value for RA.

Repeat Step 1. The outdoor-air condition (OA) remains the same; but in this iteration, RA' represents the indoor condition.

Repeat Steps 2 and 3. Because the sensible- and latent-cooling loads in the classroom are unchanged, the supply airflow and supply-air dry-bulb temperature also remain the same.

Repeat Step 4. Plot the new mixed-air condition (MA') by drawing a line that connects the outdoor-air condition (OA) to the new return-air condition (RA'). MA' lies at the intersection of OA–RA' and the dry-bulb temperature that was calculated for the mixed air, 80.6°F DB (26.8°C DB).

SYS-APM004-EN 105 Appendix A: Psychrometric Analysis

Figure 95. “Closing the psychrometric loop” for the peak dry-bulb condition

Repeat Step 5. To find the new supply-air condition (SA'), draw a curve from the new mixed-air condition (MA') to the supply-air temperature of 55.7°F (13.1°C). Use the curvature of the adjacent “coil curves” as a guide.

Repeat Step 6. Using the 0.85 SHR line as a reference, draw a parallel line from the supply-air condition (SA') to the line representing the desired space temperature, 74°F (23.3°C) in this case. Figure 95 shows that the resulting air condition (RA'') in the space now coincides with the starting point.

At the full-load, peak dry-bulb condition for this example, controlling the basic constant-volume system to 74°F DB (23.3°C DB) results in a relative humidity of about 52 percent.

Step 7: Calculate the required cooling-coil capacity. Solving the following equation shows that the cooling coil must provide 4.78 tons (16.8 kW) of cooling at the full-load, peak dry-bulb condition:

× × () In the QT equations at right, 4.5 and QT = 4.5 Vsa hma – hsa 1.2 are derived from properties of the × , × (), air; they are not constants. At the = 4.5 1 500 cfm 31.4– 22.9 Btu/lb = 57 375 Btu/hr or 4.78 tons “standard” air condition, which is 69°F (21°C) dry air at sea level, the product of ()× × [] QT = 1.2 Vsa hma – hsa density and the conversion for time (I-P units only) equals 4.5 (1.2). A different ()= 1.2× 0.7 m³/s × []73.1– 53.1 kJ/kg = 16.8 kW air condition or elevation will result in a different value. ■ where,

hma = enthalpy of mixed air entering the cooling coil, Btu/lb (kJ/kg)

106 SYS-APM004-EN Appendix A: Psychrometric Analysis

hsa = enthalpy of supply air leaving the cooling coil, Btu/lb (kJ/kg)

QT = total cooling load on the coil, Btu/hr (kW)

Vsa = supply airflow passing through the coil, cfm (m³/s)

Part-Load, Peak Dew-Point Condition

Note: The steps involved in a psychrometric analysis remain the same, regardless of the operating conditions. This section details these steps again by way of illustration.

Step 1: Plot the outdoor (OA) and target indoor (RA) conditions on a psychrometric chart. The ASHRAE Handbook–Fundamentals lists the peak dew-point condition for Jacksonville, Florida, as 76°F DP, 84°F DB (24.6°C DP, 28.8°C DB). For this example, the desired condition in the classroom remains at 74°F DB (23.3°C DB) with a relative humidity of 50 percent.

Step 2: Calculate the sensible- and latent-cooling loads in the space. The outdoor condition at the peak dew point is comparatively cooler than at the peak dry bulb. Because less heat conduction occurs, the sensible-cooling load in the space drops to 17,850 Btu/hr (5.2 kW). However, the latent-cooling load remains the same because the occupants are still in the classroom. These loads represent a space sensible-heat ratio (SHR) of 0.77:

17, 850 Btu/hr SHR = ------= 0.77 17, 850 Btu/hr + 5, 250 Btu/hr

§·5.2 kW SHR = ------= 0.77 ©¹5.2 kW + 1.5 kW

Space sensible and latent loads, and therefore the space SHR, vary with outdoor conditions, the time of year, and the time of day. Unlike the peak dry- bulb condition, where you can assume (for most North American locations) that the highest dry-bulb temperature occurs during the midafternoon hours of July or August, such an assumption is not valid for the peak dew-point condition.

Hourly weather data for Jacksonville indicates that the 0.4 percent, peak The occurrences of the peak dew-point dew-point condition occurs on the occasions listed below. Notice that the time in Jacksonville appear to be concentrated in late August/early of each occurrence differs from one date to the next. Depending on the September. By contrast, a similar analysis for St. Louis, Missouri, reveals Outdoor condition that this condition occurs in late June/ early July. Clearly, it is unwise to Date Time of day Dry bulb Dew point assume that the peak dew-point August 23 8:00 a.m. condition occurs during the same September 1 1:00 p.m. 84.0°F DB 76.1°F DP ■ month, regardless of location. September 2 4:00 p.m.

SYS-APM004-EN 107 Appendix A: Psychrometric Analysis

orientation of the exterior surfaces of a given space, as well as the occupancy schedule, the space SHR can vary dramatically between these occasions.

The classroom in this example has one, west-facing exterior wall with windows. Of the three, peak dew-point occurrences, the lowest space SHR occurs at 8:00 a.m. because the solar and conduction loads are smaller. A psychrometric analysis will yield substantially different results, depending on which peak dew- point occurrence is used to calculate the cooling loads in the space.

Cooling loads in classroom, Btu/hr Date Time of day Sensible Latent Space SHR August 23 8:00 a.m. 17,850 5,250 0.77 September 1 1:00 p.m. 29,975 5,250 0.85 September 2 4:00 p.m. 29,250 875 0.97

Step 3: Calculate the temperature and quantity of air required to offset the space sensible-cooling load. Because a constant-volume system serves the classroom in this example, the supply airflow must remain at 1,500 cfm (0.7 m³/s). To compensate for the smaller space sensible-cooling load and still maintain the 74°F DB (23.3°C) setpoint, the system must deliver the supply airflow at 63°F (17.2°C):

× , × (), ∴ Qs = 1.085 1 500 cfm 74°F– Tsa = 17 850 Btu/hr Tsa = 63.0°F ()× × []∴ Qs = 1.21 0.7 m³/s 23.3°C– Tsa = 5.2 kW Tsa = 17.2°C

Step 4: Calculate the entering-air condition for the cooling coil. The coil cools a mixture of outdoor air and recirculated return air. Assuming that the return-air condition is identical to the condition in the classroom—74°F DB (23.3°C) DB and 50 percent relative humidity—the resulting mixed-air dry-bulb temperature is 77°F (24.9°C):

()450 cfm × 84°F + ()1, 050 cfm × 74°F T = ------= 77.0°F ma 450 cfm + 1, 050 cfm

§·[]0.21 m³/s × 28.8°C + []0.5 m³/s × 23.3°C T = ------= 24.9°C ©¹ma 0.21 m³/s + 0.5 m³/s

On the psychrometric chart (Figure 96), mixed-air condition MA lies on the straight line that connects the outdoor-air and return-air conditions, OA and RA.

108 SYS-APM004-EN Appendix A: Psychrometric Analysis

Figure 96. Part-load, peak dew-point conditions for the Jacksonville, Florida, classroom

Step 5: Determine the leaving-coil, supply-air condition. Using the curvature of the adjacent “coil curves,” draw a curve from the mixed-air condition (MA) until it intersects the supply-air temperature (SA) calculated in Step 3—in this case, 63°F (17.2°C).

Step 6: Determine the air condition in the space. (Figure 97) Draw a 0.77 SHR line that intersects both the supply-air condition and the target dry-bulb

Figure 97. Resulting air condition (RA') in the space does not match the target (RA)

SYS-APM004-EN 109 Appendix A: Psychrometric Analysis

Figure 98. “Closing the psychrometric loop” for the peak dew-point condition

temperature for the classroom, 74°F (23.3°C). This time, the space sensible-heat ratio and supply-air temperature yield a relative humidity of 65 percent.

Repeat Steps 1–6: Because the starting and ending conditions are not the same, repeat the psychrometric analysis as often as necessary to “close the loop.” The outdoor-air condition (OA) remains the same, but RA' now represents the indoor condition.

(Figure 98) For this example, completing the third iteration finally yields a space condition that matches the starting point. At the part-load, peak dew-point condition, controlling the classroom’s basic constant-volume system to 74°F DB (23.3°C DB) results in a relative humidity of approximately 67 percent (RA'').

Step 7: Calculate the required cooling-coil capacity. Solving the following equation shows that the total cooling-coil load to maintain the target dry-bulb temperature is 3.66 tons (12.7 kW)…but without adequately dehumidifying the space.

× , × (), QT = 4.5 1 500 cfm 34.1– 27.6 Btu/lb = 43 875 Btu/hr or 3.66 tons ()× × []■ QT = 1.2 0.7 m³/s 79.3– 64.2 kJ/kg = 12.7 kW

110 SYS-APM004-EN Appendix B: Designing a Dedicated Outdoor-Air System

Two fundamental decisions underlie the design of a dedicated outdoor-air system and affect the selection of the dedicated outdoor-air handler and local HVAC terminals:

■ Whether to deliver the conditioned outdoor air directly to the space or to local HVAC terminals ■ Whether to condition the outdoor air to a neutral or cold dry-bulb temperature

(Figure 99) To provide the desired dehumidification, the dedicated outdoor-air handler commonly is sized to offset the local latent loads in the space, plus the sensible and latent loads of the outdoor air brought into the building for ventilation. The local HVAC terminals may also help to dehumidify the space when the sensible load is high, yielding an indoor humidity that is drier than the target. As a rule of thumb:

■ Size the dedicated outdoor-air handler so that it also offsets the space latent loads and the total ventilation load at the peak outdoor-enthalpy condition ( p p . 11 2 – 11 5 ) . ■ Size each local HVAC terminal so that it offsets the sensible-cooling load in the space at the peak dry-bulb condition. Design capacity will vary with the system configuration, that is…

❚ neutral to space (pp. 117–118) ❚ neutral to unit (pp. 120–122) ❚ cold to space (pp. 119–120) ❚ cold to unit (pp. 122–124)

The rest of this chapter describes the selection process for each case.

Figure 99. Typical division of loads in dedicated outdoor-air systems

SYS-APM004-EN 111 Appendix B: Designing a Dedicated Outdoor-Air System

Selecting the Dedicated Outdoor-Air Handler

The following steps establish the required airflow, dew point, and dry-bulb temperature for the conditioned air.

Step 1: Determine the entering-air condition. Three factors dictate the cooling capacity required from the dedicated outdoor-air handler: airflow, the enthalpy of the entering outdoor air, and the enthalpy of the conditioned air leaving the cooling coil. If the outdoor airflow is constant, then the basis of design is the condition resulting in the greatest difference in enthalpy across the cooling coil.

Indoor latent loads fluctuate with occupancy and processes, as well as with ambient conditions and wind through infiltration. These variables can make it difficult to discover when the greatest enthalpy difference occurs. However, if the latent loads within the space are relatively constant and infiltration is minimal, assume that the greatest enthalpy difference occurs at the highest outdoor-air enthalpy.

In most climates, the peak latent ventilation load occurs at a lower dry-bulb temperature and higher dew point than the outdoor-air condition that produces the peak sensible ventilation load. The ASHRAE Handbook—Fundamentals is a popular source for climatic data representing the outdoor design conditions for many locations. To aid the design of cooling and dehumidifying systems, the handbook includes:

■ Peak dry-bulb and mean-coincident wet-bulb temperatures (sensible-design condition) ■ Peak dew-point and mean-coincident dry-bulb temperatures (latent-design condition) ■ Peak wet-bulb and mean-coincident dry-bulb temperatures (enthalpy-design condition)

Table 10 lists the 0.4 percent, cooling-design data for Jacksonville, Florida. The examples throughout this manual Plotting these values on the psychrometric chart (Figure 100) illustrates that the are based on “0.4 percent” data from highest outdoor enthalpy exists at the peak wet-bulb condition. In this case, the the ASHRAE climatic data tables. This enthalpy of the outdoor air is 8 percent higher than it is at the peak dry-bulb percentage indicates that the (sensible-cooling design) condition. temperature is likely to equal or exceed the design value for 35 hours each year. Note: Using the peak dry-bulb condition as the basis of design will undersize Some design engineers choose to use the dedicated outdoor-air handler, making it unable to properly dehumidify the more extreme conditions; others base outdoor air at certain part-load conditions. Remember that the primary purpose their designs on the “1 percent” or of the dedicated outdoor-air system is to properly control space humidity at all “2 percent” values, which represent load conditions. more hours. ■

11 2 SYS-APM004-EN Appendix B: Designing a Dedicated Outdoor-Air System

Table 10. Design weather conditions for cooling/dehumidifying in Jacksonville, Florida 1 Design condition Enthalpy Peak dry bulb, mean-coincident wet bulb 96°F (35.7°C) DB, 39.3 Btu/lb 76°F (24.5°C) WB (91.4 kJ/kg) Peak dew point, mean-coincident dry bulb 76°F (24.6°C) DP, 41.5 Btu/lb 84°F (28.8°C) DB (96.5 kJ/kg) Peak wet bulb, mean-coincident dry bulb 79°F (26.1°C) WB, 42.4 Btu/lb 91°F (32.8°C) DB (98.6 kJ/kg)

1 Source: 2001 ASHRAE Handbook–Fundamentals, Chapter 27, Table 1B (0.4% condition)

Figure 100. Comparison of outdoor-air at 0.4% cooling-design conditions for Jacksonville, Florida (see Table 10)

Step 2: Choose the maximum limit for space humidity. ASHRAE Standard 62–2001 recommends maintaining the indoor humidity at less than 60 percent-relative humidity (RH) to prevent high relative humidity at surfaces and thereby minimize microbial growth on building materials and furnishings. Therefore, select the dedicated outdoor-air handler with enough capacity to prevent the humidity in any space from exceeding 60 percent at the worst-case condition. The indoor humidity may be drier than 60 percent at less extreme conditions, depending on how the system is controlled.

In this example, combining the 74°F (23.3°C) setpoint for the space with a maximum relative humidity of 60 percent corresponds to a humidity ratio of 75.2 grains/lb (10.8 g/kg) or a dew point of approximately 59°F (15°C).

SYS-APM004-EN 113 Appendix B: Designing a Dedicated Outdoor-Air System

Note: Selecting the unit to enforce a lower humidity limit (50 percent, for example) is possible, but requires larger equipment and consumes more energy.

Step 3: Determine the latent loads in the space. The dedicated outdoor- Managing Building Moisture, Trane air handler will offset the local latent loads in the space it serves, as well as the applications engineering manual total ventilation load. Common sources of latent load include respiration from SYS-AM-15, helps designers identify people, processes (such as cooking), and the infiltration of humid outdoor air and quantify moisture sources. It also through cracks and other openings in the building structure; see “Sources and presents moisture-management Effects of Indoor Moisture,” p. 2. techniques for the building envelope, occupied spaces, and mechanical For this example, the dedicated outdoor-air handler serves four classrooms of equipment rooms. ■ a school in Jacksonville, Florida. Table 11 lists the latent load for each space; in this case, the latent loads presumedly remain constant whenever the building is occupied.

Step 4: Determine the total airflow that the dedicated outdoor-air handler must deliver. According to ASHRAE Standard 62–2001, a classroom requires 15 cfm (0.007 m³/s) of outdoor air per person for ventilation. Given the design occupancies and associated outdoor-air requirements listed in Table 11, the dedicated outdoor-air handler in this example must deliver a total outdoor airflow of 1,815 cfm (0.86 m³/s).

Table 11. Design criteria for a dedicated outdoor-air system that serves four classrooms in Jacksonville, Florida (example) Space characteristics Classroom 101 Classroom 102 Classroom 103 Classroom 104 Sensible load 29,750 Btu/hr 26,775 Btu/hr 26,927 Btu/hr 28,262 Btu/hr (8.7 kW) (7.8 kW) (7.9 kW) (8.3 kW) Latent load 5,250 Btu/hr 5,465 Btu/hr 5,697 Btu/hr 5,250 Btu/hr (1.5 kW) (1.6 kW) (1.7 kW) (1.5 kW) Sensible-heat ratio (SHR) 0.85 0.83 0.83 0.84 Occupants 30 30 32 29 Required outdoor airflow 450 cfm 450 cfm 480 cfm 435 cfm (0.21 m³/s) (0.21 m³/s) (0.23 m³/s) (0.20 m³/s) Humidity ratio of 58.3 grains/lb 57.6 grains/lb 58.0 grains/lb 57.7 grains/lb

conditioned air, Wca (8.34 g/kg) (8.24 g/kg) (8.29 g/kg) (8.25 g/kg)

Step 5: Determine which zone requires the driest conditioned outdoor air. Because the dedicated outdoor-air handler will offset the latent loads in each space (as well as the total ventilation load), the conditioned outdoor air must be dry enough to enforce the maximum humidity limit in the

114 SYS-APM004-EN Appendix B: Designing a Dedicated Outdoor-Air System

worst-case space. Use the following equation to calculate the required

conditioned-air humidity ratio, Wca , for each space: In the QL equations at right, 0.69 and × × () 3.0 are derived from the properties of QL = 0.69 Voa Wsp – Wca air; they are not constants. At the ()× × [] “standard” air condition, which is 69°F QL = 3.0 Voa Wsp – Wca (21°C) dry air at sea level, the product of density, the of water vapor, where, and a conversion factor for units— QL = latent load in the space, Btu/hr (kW) 7,000 grains/lb (1,000 grams/kg) and Voa = conditioned outdoor airflow, cfm (m³/s), which is supplied to the 60 min/hr—equals 0.69 (3.0). A space by the dedicated outdoor-air handler different air condition or elevation will Wca = humidity ratio of the conditioned outdoor air, grains/lb (grams/kg) result in a different value. ■ Wsp = maximum limit for the humidity ratio in the space, grains/lb (grams/kg)

For example, to assure that the humidity (Wsp ) in Classroom 101 does not Step 5 (Classroom 101): exceed the maximum limit of 75.2 grains/lb (10.8 g/kg), the humidity ratio of the

, conditioned outdoor air, Wca, must be 58.3 grains/lb (8.34 g/kg). QL = 5 250 Btu/hr × × () = 0.69 450 cfm 75.2 gr/lb– Wca Table 11 shows the results of this calculation for all four classrooms. Although ∴ Wca = 58.3 gr/lb the highest latent load exists in Classroom 103, the “critical space” is

Classroom 102 because it requires the driest air (lowest humidity ratio, Wca ). () QL = 1.5 kW Supplying the conditioned outdoor air at a humidity ratio of 57.6 gr/lb ()× × [] = 3.0 0.21 m³/s 10.8 g/kg– Wca (8.24 g/kg) will offset the latent load in each classroom and assure that the ()∴ Wca = 8.34 g/kg humidity in Classroom 102 does not exceed the maximum limit; lower humidities will result in the other classrooms.

Step 6: Determine the required dew point for the conditioned outdoor air. With the help of a psychrometric chart (Figure 101, p. 116), we Step 6: find that a humidity ratio of 57.6 grains/lb (8.24 g/kg) is equivalent to a dew- × , × ()point temperature of 52°F (11.1°C). QT = 4.5 1 815 cfm 42.4– 21.4 Btu/lb , = 171 518 Btu/hr = 14.3 tons Dehumidifying 1,815 cfm (0.86 m³/s) of outdoor air from the peak wet-bulb condition to a 52°F (11.1°C) dew point requires 14.3 tons (50.4 kW) of ()Q = 1.2× 0.86 m³/s × []98.6– 49.8 kJ/kg T cooling capacity. ()= 50.4 kW Incidentally, lowering the maximum humidity limit to 50 percent would require 16.9 tons (59.4 kW)…an 18 percent increase in capacity.

Step 7: Determine the supply-air dry-bulb temperature for the dedicated outdoor-air handler. If the system design requires neutral- temperature conditioned air, then the air leaving the dedicated outdoor-air handler must be reheated to the desired dry-bulb temperature. This is typically between 70°F and 75°F (21°C and 24°C). For our example, assume that the air is reheated to 71°F (21.7°C).

If the system design is based on cold conditioned air rather than neutral- temperature air, then the dry-bulb temperature from the dedicated outdoor-air

SYS-APM004-EN 115 Appendix B: Designing a Dedicated Outdoor-Air System

Figure 101. Plotting the humidity ratio to determine the equivalent dew-point temperature

handler depends on the supply-air dew point. In our example, assuming that saturated air leaves the cooling coil, then the leaving-air dry-bulb temperature is 52°F (11.1°C).

Note: For simplicity, our example does not include the effect of fan heat. A draw-through fan arrangement will increase the dry-bulb temperature of the conditioned outdoor air. The slightly warmer air offsets less of the sensible load in the space, which will affect the selection criteria for the local HVAC terminals.

Selecting the Local HVAC Terminals

Table 12. Sensible-cooling-design criteria for Classroom 101 Using the dedicated outdoor-air handler to offset the latent loads in the space Sensible load 29,750 Btu/hr reduces the total cooling capacity that the local HVAC terminals must provide (8.7 kW) and, therefore, affects the selection of these units. For this discussion, a “local Latent load 5,250 Btu/hr HVAC terminal” represents any type of air-handling equipment that controls the (1.5 kW) sensible-cooling load in the space, including fan–coils, unit ventilators, water- Sensible-heat ratio (SHR) 0.85 source heat pumps, blower–coils, air handlers (chilled water or DX cooling), Required outdoor airflow 450 cfm packaged rooftop air conditioners, and vertical self-contained air conditioners. (0.21 m³/s) The following instructions describe how to determine the required cooling Target condition 74°F (23.3°C) DB, 60% RH max capacity, supply airflow, and supply-air dry-bulb temperature for the local “Neutral” conditioned 71°F (21.7°C) DB, HVAC terminals when they are used in conjunction with a dedicated outdoor-air outdoor air 52°F (11.1°C) DP handler. Design criteria for the local HVAC terminals depend on where the “Cold” conditioned 52°F (11.1°C) DB, conditioned outdoor air is delivered (to the space or to the unit), and at what outdoor air 52°F (11.1°C) DP dry-bulb temperature (neutral or cold). To illustrate how these factors affect the selection of HVAC terminals, the process is repeated for neutral-to-space,

11 6 SYS-APM004-EN Appendix B: Designing a Dedicated Outdoor-Air System

cold-to-space, neutral-to-unit, and cold-to-unit systems using Classroom 101 as an example. (Classroom 101 was introduced on p. 114; Table 12 summarizes its design criteria.)

Neutral CA to space

(Figure 102) When the conditioned outdoor air (CA) is delivered directly to the space at a dry-bulb temperature that approximately equals room air, it has little effect on the sensible-cooling load in the space.

Step 1: Determine the entering-air condition for the HVAC terminal. Because the outdoor air is delivered directly to the space, all of the air that enters the HVAC terminal’s cooling coil is recirculated return air; therefore, the room-air condition becomes the entering-air condition for the local HVAC terminal. For this example, the dry-bulb temperature of the recirculated return air in Classroom 101 is 74°F (23.3°C).

Step 2: Calculate the sensible-cooling load on the HVAC terminal. If it is cooler than room air, the conditioned outdoor air reduces the sensible- cooling load on the HVAC terminal. To determine the required cooling-coil capacity, subtract the amount of sensible cooling provided by the conditioned

outdoor air (Qs, ca ) from the sensible-cooling load in the space (Qs, space ) at the peak dry-bulb condition:

Qs, coil = Qs, space – Qs, ca

where,

Qs, ca = sensible cooling provided by conditioned outdoor air, Btu/hr (kW)

Qs, coil = sensible-cooling load on the terminal coil, Btu/hr (kW)

Qs, space = sensible-cooling load in the space, Btu/hr (kW)

Figure 102. Typical HVAC system with “neutral-to-space” conditioned outdoor air

SYS-APM004-EN 117 Appendix B: Designing a Dedicated Outdoor-Air System

Qs, ca can be determined using the following equation: Step 2 (Classroom 101): × × () × × () Qs, ca = 1.085 450 cfm 74– 71°F Qs, ca = 1.085 Voa Tsp – Tca , = 1 465 Btu/hr ()× × [] Qs, ca = 1.21 Voa Tsp – Tca ()Q = 1.21× 0.21 m³/s × []23.3– 21.7°C s, ca where, ()= 0.41 kW Tca = dry-bulb temperature of the conditioned outdoor air, °F (°C)

Tsp = desired dry-bulb temperature for the space, °F (°C)

, , , Voa = conditioned outdoor airflow delivered to the space, Qs, coil ==29 750 – 1 465 28 285 Btu/hr cfm (m³/s) () Qs, coil = 8.7– 0.41= 8.29 kW The conditioned outdoor air in this example is only 3°F (1.6°C) cooler than the air in Classroom 101, and only slightly reduces the sensible-cooling load in the space. The HVAC terminal must offset the remaining 28,285 Btu/hr (8.29 kW).

Step 3: Determine the airflow and dry-bulb temperature for supply air from the HVAC terminal. After determining both the sensible-cooling load and the target dry-bulb temperature for the space, either:

Step 3 (Classroom 101): ■ Define the supply airflow and use that value to calculate the supply-air , temperature; or, Qs, coil = 28 285 Btu/hr × , × ()■ = 1.085 1 500 cfm 74°F– Tsa Choose the supply-air temperature and use that value to calculate the ∴ required airflow. Tsa = 56.6°F If we assume that the HVAC terminal for Classroom 101 must deliver nine air ()Q = 8.29 kW s, coil changes per hour, then the required supply airflow is 1,500 cfm (0.7 m³/s), and ()= 1.21× 0.7 m³/s × []23.3°C– T sa the resulting supply-air temperature is 56.6°F DB (13.5°C DB). ()∴ Tsa = 13.5°C Alternatively, a designer may choose to select the HVAC terminal so that the total airflow circulated in the space—that is, conditioned outdoor air delivered , Qs, coil = 28 285 Btu/hr directly to the space plus return air recirculated by the HVAC terminal—equals × , × ()nine air changes per hour. When we apply this strategy to Classroom 101, the = 1.085 1 050 cfm 74°F– Tsa ∴ HVAC terminal must provide only 1,050 cfm (0.5 m³/s) of supply air, but at the Tsa = 49.2°F much lower supply-air temperature of 49.2°F (9.6°C) DB. ()Q = 8.29 kW s, coil Note: Delivering cold supply air through an HVAC terminal that is located in the ()× × [] = 1.21 0.5 m³/s 23.3°C– Tsa occupied space may create uncomfortable drafts. To mitigate this risk, either ()∴ Tsa = 9.6°C deliver the supply air through overhead, high-aspiration diffusers, which will promote better mixing and increase air throw, or select the HVAC terminals based on a supply-air temperature that is closer to 55°F (12.8°C). The design of the HVAC terminal can make it difficult to achieve low supply-air temperatures, too; the available coil configurations may dictate the combination of supply airflow and supply-air temperature.

For this example, the HVAC terminal that was selected for Classroom 101 will cool 1,500 cfm (0.7 m3/s) of recirculated return air to 56.6°F (13.5°C).

11 8 SYS-APM004-EN Appendix B: Designing a Dedicated Outdoor-Air System

Cold CA to space

(Figure 103) When the conditioned outdoor air is delivered directly to the space at a dry-bulb temperature that is significantly cooler than room air, it offsets a much larger portion of the sensible-cooling load in the space.

Step 1: Determine the entering-air condition for the HVAC terminal. Because the outdoor air is delivered directly to the space, all of the air that enters the HVAC terminal’s cooling coil is recirculated return air; therefore, the room-air condition becomes the entering-air condition for the local HVAC terminal. For this example, the dry-bulb temperature of the recirculated return air in Classroom 101 is 74°F (23.3°C).

Step 2: Determine the sensible-cooling load on the HVAC terminal. If it is cooler than room air, the conditioned outdoor air reduces the sensible- cooling load on the HVAC terminal. To determine the required cooling-coil capacity, subtract the amount of sensible cooling provided by the conditioned

outdoor air (Qs, ca ) from the sensible-cooling load in the space (Qs, space ) at the peak dry-bulb condition:

Qs, coil = Qs, space – Qs, ca

where,

Qs, ca = sensible cooling provided by conditioned outdoor air, Btu/hr (kW)

Qs, coil = sensible-cooling load on the terminal coil, Btu/hr (kW)

Qs, space = sensible-cooling load in the space, Btu/hr (kW)

Figure 103. Typical HVAC system with “cold-to-space” conditioned outdoor air

SYS-APM004-EN 119 Appendix B: Designing a Dedicated Outdoor-Air System

Qs, ca can be determined using the following equation: Step 2 (Classroom 101): × × () × × () Qs, ca = 1.085 450 cfm 74– 52°F Qs, ca = 1.085 Voa Tsp – Tca , = 10 742 Btu/hr ()× × [] Qs, ca = 1.21 Voa Tsp – Tca ()Q = 1.21× 0.21 m³/s × []23.3– 11.1°C s, ca where, ()= 3.13 kW Tca = dry-bulb temperature of the conditioned outdoor air, °F (°C)

Tsp = desired dry-bulb temperature for the space, °F (°C)

, , , Voa = conditioned outdoor airflow delivered to the space, Qs, coil ==29 750 – 10 742 19 008 Btu/hr cfm (m³/s) () Qs, coil = 8.7– 3.13= 5.57 kW The conditioned outdoor air in this example is 22°F (12.2°C) colder than the air in Classroom 101, which offsets 36 percent of the sensible-cooling load in that space. The HVAC terminal must offset the remaining 19,008 Btu/hr (5.57 kW).

Note: This example omits the effect of fan heat. A draw-through fan arrangement will increase the dry-bulb temperature of the conditioned outdoor air, which will affect the selection of the HVAC terminal.

Step 3: Determine the airflow and dry-bulb temperature for supply air from the HVAC terminal. After you know both the sensible-cooling load and the target dry-bulb temperature for the space, you can either: Step 3 (Classroom 101): , ■ Qs, coil = 19 008 Btu/hr Define the supply airflow and use that value to calculate the supply-air × , × ()temperature; or, = 1.085 1 050 cfm 74°F– Tsa ∴ ■ Tsa = 57.3°F Choose the supply-air temperature and use that value to calculate the required airflow. () Qs, coil = 5.57 kW ()× × []For our example, we assumed that Classroom 101 requires nine air changes per = 1.21 0.5 m³/s 23.3°C– Tsa hour, or 1,500 cfm (0.7 m³/s). Because the dedicated outdoor-air handler ()∴ T = 14.1°C sa delivers 450 cfm (0.21 m³/s) directly to the space, the local HVAC terminal must condition only 1,050 cfm (0.5 m³/s) of recirculated return air. The required supply-air temperature is 57.3°F DB (14.1°C DB).

Neutral CA to unit

(Figure 104) When the conditioned outdoor air is delivered to HVAC terminals instead of directly to the space, it affects the condition of the mixed air that enters the HVAC terminal’s cooling coil.

Step 1: Determine the sensible-cooling load on the HVAC terminal. Because the conditioned outdoor air is not delivered directly to the space, it has no direct effect on the sensible-cooling load in the space. Therefore, the sensible-cooling load for Classroom 101 is 29,750 Btu/hr (8.7 kW) at the peak dry-bulb condition.

120 SYS-APM004-EN Appendix B: Designing a Dedicated Outdoor-Air System

Figure 104. Typical HVAC system with “neutral-to-unit” conditioned outdoor air

Step 2: Determine the airflow and dry-bulb temperature for supply air from the HVAC terminal. After you know both the sensible-cooling load and Step 2 (Classroom 101): the target dry-bulb temperature for the space, you can either: Q = 29, 750 Btu/hr s ■ Define the supply airflow and use that value to calculate the supply-air × , × () = 1.085 1 500 cfm 74°F– Tsa temperature; or, ∴ T = 55.7°F sa ■ Choose the supply-air temperature and use that value to calculate the () required airflow. Qs = 8.7 kW ()× × []If we assume that the HVAC terminal for Classroom 101 must deliver nine air = 1.21 0.7 m³/s 23.3°C– Tsa ()∴ changes per hour, then the required supply airflow for the classroom is Tsa = 13.1°C 1,500 cfm (0.7 m3/s) and the resulting supply-air temperature is 55.7°F DB (13.1°C DB).

Step 3: Determine the entering mixed-air condition (Tma ) for the HVAC terminal. Before the conditioned outdoor air passes through the cooling coil of the HVAC terminal, it mixes with recirculated return air from the space. For this example, we assumed that the recirculated return air did not pick up heat as it traveled from the space to the HVAC terminal. Therefore, 73.1°F DB (22.8°C DB) mixed air for Classroom 101 results from combining:

■ 450 cfm (0.21 m³/s) of conditioned outdoor air at 71°F DB and 52°F DP To predict the relative humidity for (21.7°C DB, 11.1°C DP) recirculated return air from the space, perform a psychrometric analysis of the ■ 1,050 cfm (0.5 m³/s) of recirculated return air at 74°F DB (23.3°C DB) and peak dry-bulb condition. ■ 49 percent-relative humidity

A psychrometric analysis (Figure 105, p. 122) of the peak dry-bulb condition shows that the mixed-air condition (MA) lies on the straight line that would connect the conditions of the conditioned outdoor air (CA) and the return air

SYS-APM004-EN 121 Appendix B: Designing a Dedicated Outdoor-Air System

Figure 105. Peak dry-bulb condition for HVAC terminals that receive neutral-temperature conditioned outdoor air (example)

(RA) from the space. For Classroom 101, the entering mixed-air condition at the HVAC terminal is 73.1°F DB and 60.9°F WB (22.8°C DB, 16.1°C WB).

()450 cfm × 71°F + ()1, 050 cfm × 74°F T ==------73.1°F ma 450 cfm + 1, 050 cfm

§·[]0.21 m³/s × 21.7°C + []0.5 m³/s × 23.3°C T ==------22.8°C ©¹ma 0.21 m³/s + 0.5 m³/s

Therefore, when the HVAC terminal for Classroom 101 is part of a “neutral-to- unit,” dedicated outdoor-air system, it must be selected to cool 1,500 cfm (0.7 m³/s) of mixed air from 73.1°F DB, 60.9°F WB (22.8°C DB, 16.1°C WB) to a supply-air temperature of 55.7°F (13.1°C).

Cold CA to unit

(Figure 106) When the conditioned outdoor air is delivered to HVAC terminals instead of directly to the space, it affects the condition of the mixed air entering the terminal’s cooling coil.

Step 1: Determine the sensible-cooling load on the HVAC terminal. Because the conditioned outdoor air is not delivered directly to the space, it has no direct effect on the sensible-cooling load in the space. Therefore, the sensible-cooling load for Classroom 101 is 29,750 Btu/hr (8.7 kW) at the peak dry-bulb condition.

122 SYS-APM004-EN Appendix B: Designing a Dedicated Outdoor-Air System

Figure 106. Typical HVAC system with “cold-to-unit” conditioned outdoor air

Step 2: Determine the airflow and dry-bulb temperature for supply air Step 2 (Classroom 101): from the HVAC terminal. After you know both the sensible-cooling load and , the target dry-bulb temperature for the space, you can either: Qs = 29 750 Btu/hr × , × () = 1.085 1 500 cfm 74°F– Tsa ■ Define the supply airflow and use that value to calculate the supply-air ∴ Tsa = 55.7°F temperature; or, ■ Choose the supply-air temperature and use that value to calculate the ()Q = 8.7 kW s required airflow. ()× × [] = 1.21 0.7 m³/s 23.3°C– Tsa ()∴ Tsa = 13.1°C If we assume that the HVAC terminal must provide Classroom 101 with nine air changes per hour, then the supply airflow for the classroom is 1,500 cfm (0.7 m³/s) and the resulting supply-air temperature is 55.7°F DB (13.1°C DB).

Step 3: Determine the entering mixed-air condition (Tma ) for the HVAC terminal. Before the conditioned outdoor air passes through the cooling coil of the HVAC terminal, it mixes with recirculated return air from the space. If we assume that the recirculated return air from Classroom 101 does not pick up heat en route to the HVAC terminal, then:

■ 450 cfm (0.21 m³/s) of conditioned outdoor air at 52°F DB and 52°F DP To predict the relative humidity for (11.1°C DB, 11.1°C DP) mixes with… recirculated return air from the space, perform a psychrometric analysis of the ■ 1,050 cfm (0.5 m³/s) of recirculated return air at 74°F DB (23.3°C DB) and peak dry-bulb condition. ■ 49 percent-relative humidity, resulting in 67.4°F DB (19.8°C DB) mixed air.

()450 cfm × 52°F + ()1, 050 cfm × 74°F T ==------67.4°F ma 450 cfm + 1, 050 cfm

§·[]0.21 m³/s × 11.1°C + []0.5 m³/s × 23.3°C T ==------19.8°C ©¹ma 0.21 m³/s + 0.5 m³/s

SYS-APM004-EN 123 Appendix B: Designing a Dedicated Outdoor-Air System

Figure 107. Peak dry-bulb condition for HVAC terminals receiving cold, conditioned outdoor air

A psychrometric analysis (Figure 107) of the peak dry-bulb condition shows that the mixed-air condition (MA) lies on the straight line that connects the points representing the conditioned outdoor air (CA) and the return air (RA) from the space. For Classroom 101, the entering mixed-air condition at the HVAC terminal is 67.4°F DB and 58.8°F WB (19.8°C DB, 14.9°C WB).

Therefore, when the HVAC terminal for Classroom 101 is part of a “cold-to- unit,” dedicated outdoor-air system, it must be selected to cool 1,500 cfm (0.7 m3/s) of mixed air from 67.4°F DB, 58.8°F WB (19.8°C DB, 14.9°C WB) to a supply-air temperature of 55.7°F (13.1°C).

Note: This example omits the effect of fan heat. A draw-through fan arrangement will increase the dry-bulb temperature of the conditioned outdoor air, as well as the condition of the mixed air entering the HVAC terminal. ■

124 SYS-APM004-EN Glossary

active desiccant. See desiccant.

adsorption. The process by which a material accumulates moisture from the air without undergoing a chemical change.

air handler (AHU). A piece of equipment that moves and/or conditions air to maintain the thermal comfort of an indoor environment.

air-to-air energy recovery. The transfer of sensible heat, or sensible heat plus moisture, between two or more air streams, or between two locations within the same air stream.

airside economizer. A device and/or control strategy that conserves energy by introducing cool outdoor air to reduce the mechanical cooling requirement. Operation is based on one of the following: a fixed outdoor-air temperature; a fixed outdoor-air enthalpy; a comparison of return- and outdoor-air temperatures; or a comparison return- and outdoor-air enthalpies.

ANSI. American National Standards Institute (www.ansi.org)

ANSI/ASHRAE Standard 55. Titled Thermal Environmental Conditions for Human Occupancy, this standard specifies the combinations of indoor environments and personal factors that will produce thermal conditions acceptable to 80 percent or more of the occupants.

ANSI/ASHRAE Standard 62. Titled Ventilation for Acceptable Indoor Air Quality, this standard specifies minimum ventilation rates necessary to minimize the potential for adverse, building-related health effects and to maintain odor-related comfort.

ANSI/ASHRAE/IESNA Standard 90.1. Titled Energy Standard for Buildings Except Low-Rise Residential Buildings, this standard sets minimum design requirements that encourage energy efficiency throughout the building—that is, for the envelope, lighting, motors, HVAC, and service-water heating systems.

ASHRAE. American Society of Heating, Refrigerating, and Air-Conditioning Engineers (www.ashrae.org)

coil loop. An airside, energy-recovery device consisting of two or more finned-tube coils that are piped together in a closed loop to transfer sensible heat. A small pump circulates the working fluid through the two coils. Also called coil runaround loop or runaround loop.

conditioned outdoor air (CA). Outdoor air that has been heated, cooled, humidified, and/or dehumidified by a dedicated outdoor-air unit before being delivered directly to the conditioned space or to the mixing boxes of other local units. Also called conditioned air.

SYS-APM004-EN 125 Glossary

constant-volume system. Type of air-conditioning system that meets the changing load conditions in the space by varying the temperature of an unchanging volume of supply air.

dedicated outdoor-air unit. An air handler or packaged air conditioner that exclusively cools, heats, dehumidifies, and/or humidifies all of the outdoor air brought into the building for ventilation. The conditioned outdoor air may be delivered directly to the conditioned space or to other air handlers or terminal equipment. Also called makeup-air unit, 100 percent-outdoor-air unit, or fresh-air unit.

desiccant. An absorbent or adsorbent substance, liquid or solid, that removes water or water vapor from a material. A passive desiccant system naturally releases collected moisture to a drier air stream. An active desiccant system releases collected moisture with the help of an external source of heat, thereby improving the moisture-collecting ability of the desiccant.

direct expansion (DX). A type of refrigerating system in which the heat- transfer medium is in direct contact with the air or fluid to be cooled or heated. For example, refrigerant inside an evaporator coil absorbs heat directly from the air passing through the coil.

dual-path air handler. An air handler that uses separate cooling coils to condition the outdoor air and the recirculated return air before the mixture is delivered to the space.

enthalpy (h). The total amount of “heat” energy, both sensible and latent, in the air at a given condition.

exhaust air (EA). Air removed from the conditioned space and expelled from the building by mechanical ventilation. See also recirculated return air (RRA) and return air (RA).

face-split coil circuiting. A type of direct-expansion, evaporator arrangement that divides the coil into parallel, independently controlled sections. Sections are activated or deactivated to match the coil capacity with the cooling load. Also called horizontal-split coil circuiting and parallel-flow coil circuiting.

fixed-plate heat exchanger. An airside, energy-recovery device that consists of thermally conductive plates. It transfers sensible heat between two air streams that flow perpendicular to each other.

heat pipe. An airside, energy-recovery device consisting of refrigerant-filled, finned tubes. Heating one end of the pipe causes the refrigerant to vaporize and migrate to the other end, where it condenses and releases sensible heat. Gravity or a capillary wick returns the liquid to the hot end of the pipe.

126 SYS-APM004-EN Glossary

intertwined coil circuiting. A type of direct-expansion, evaporator arrangement that uses one of two distributors to feed every other tube in each coil row, dividing the coil into overlapping, independently controlled sections. Sections are activated or deactivated to match the coil capacity with the cooling load.

mixed air (MA). Combination of outdoor air (OA) and recirculated return air (RRA).

mixed-air bypass. A method of part-load coil control that changes the temperature of the supply air downstream of the coil by varying the amount of air that passes through and around the coil. Airflow is controlled by the linked face dampers and bypass dampers positioned on the entering-air side of the coil. See also return-air bypass.

neutral air. Supply-air condition characterized by a dry-bulb temperature and humidity ratio that do not impose a heating or cooling load in the space.

outdoor air (OA). Air brought into the building—either by a ventilation system or through openings provided for natural ventilation—from outside the building. Also called fresh air, but distinguished from outside air, which ASHRAE defines as air surrounding a refrigerated or conditioned space.

outdoor-air preconditioning. The process of cooling, dehumidifying, heating, and/or humidifying outdoor air by transferring energy (either sensible heat or sensible heat plus moisture) to and from the exhaust air stream.

passive desiccant. See desiccant.

peak dew-point condition. Design dew-point and mean-coincident dry-bulb temperatures for a specific geographic location, as tabulated in the ASHRAE Handbook—Fundamentals.

peak dry-bulb condition. Design dry-bulb and mean-coincident wet-bulb temperatures for a specific geographic location, as tabulated in the ASHRAE Handbook—Fundamentals.

peak wet-bulb condition. Design wet-bulb and mean-coincident dry-bulb temperatures, as tabulated in the ASHRAE Handbook—Fundamentals.

primary air (PA). Treated air that is delivered to a terminal device, where it mixes with recirculated return air via induction or fan power.

recirculated return air (RRA). Air that is taken from, and then delivered to, the conditioned space(s)—usually after passing through an air-conditioning system. See also return air (RA).

SYS-APM004-EN 127 Glossary

return air (RA). Air removed from the conditioned space(s) and either recirculated or exhausted. See also recirculated return air (RRA) and exhaust air (EA).

return-air bypass. A method of part-load coil control that changes the temperature of the supply air downstream of the coil by mixing the air that passes through the coil with return air that bypasses the coil. Both airflows are controlled by linked face dampers and bypass dampers positioned on the entering-air side of the coil. See also mixed-air bypass.

sensible-design condition. Design dry-bulb and mean-coincident wet-bulb temperatures for a specific geographic location, as tabulated in the ASHRAE Handbook—Fundamentals. See also peak dry-bulb condition.

sensible-heat ratio (SHR). Ratio of sensible-heat gain to total (sensible plus latent) heat gain.

sensible-energy wheel. An energy-recovery device that rotates through two air streams, transferring sensible heat (temperature) from one air stream to the other. The heat-transfer media consists of a matrix of channels, which direct the two air streams in opposite directions, approximately parallel to each other. Also called a heat wheel.

supply air (SA). Air mechanically delivered to the conditioned space for ventilation, heating, cooling, humidification, and/or dehumidification.

supply-air tempering. The process of adding sensible heat to the air downstream of the dehumidifying coil, which allows independent control of both latent and sensible loads in the space. Only systems that directly control supply-air dew point or humidity in the space are candidates for supply-air tempering.

total-energy recovery. The transfer of sensible heat (temperature) and latent heat (moisture) between two or more air streams or between two locations within the same air stream.

total-energy wheel. A desiccant-coated, energy-recovery device that rotates through two air streams, transferring sensible heat (temperature) and latent heat (moisture) from one air stream to the other. Structured as a matrix of channels, the heat-transfer media directs the air streams through the wheel in a counterflow arrangement. The desiccant typically regenerates at room temperature. Also known as enthalpy wheel or passive desiccant wheel.

variable-air-volume (VAV) system. Type of air-conditioning system that varies the volume of constant-temperature air supplied to meet the changing load conditions in the space. ■

128 SYS-APM004-EN References

American National Standards Institute (ANSI) and American Society of Heating, Refrigerating, and Air-Conditioning Engineers, Inc. (ASHRAE). 2001. Ventilation for Acceptable Indoor Air Quality, ANSI/ASHRAE Standard 62. Atlanta, GA: ASHRAE.

ANSI and ASHRAE. 1992. Thermal Environmental Conditions for Human Occupancy, ANSI/ASHRAE Standard 55. Atlanta, GA: ASHRAE.

ANSI, ASHRAE, and Illuminating Engineering Society of North America (IESNA). 2001. Energy Standard for Buildings Except Low-Rise Residential Buildings, ANSI/ASHRAE/IESNA Standard 90.1. Atlanta, GA: ASHRAE.

Dorgan, C.B., R.J. Linder, and C.E. Dorgan. 1999. Application Guide: Chiller Heat Recovery. Atlanta, GA: ASHRAE.

Harriman, L., G. Brundett, and R. Kittler. 2001. Humidity Control Design Guide for Commercial and Institutional Buildings. Atlanta, GA: ASHRAE.

Odom, D. and G. DuBose. 1996. Preventing Indoor Air Quality Problems in Hot, Humid Climates: Problem Avoidance Guidelines, CH2M Hill (www.ch2m.com), in cooperation with the Disney Development Company. Orlando, FL: CH2M Hill.

Trane. 1983. Refrigerant Heat Recovery, SYS-AM-5. La Crosse, WI: Inland Printing Company. Available from http://trane.com/bookstore/; accessed May 2002.

. 1991. “Two GOOD Ideas Combine to Form One New GREAT Idea,” Engineers Newsletter, volume 20, number 1. Available at http://www.trane.com/ commercial/library/EN20-1.pdf; accessed May 2002.

. 1999. Air Conditioning Clinic: Psychrometry, TRG-TRC001-EN. La Crosse, WI: Inland Printing Company. Available from http://trane.com/bookstore/; accessed May 2002.

. 2000. “Building Moisture and Humidity Management,” Engineers Newsletter Live satellite broadcast, APP-APV005-EN (August 30, videocassette). La Crosse, WI: AVS Group. Available from http://trane.com/bookstore/; accessed May 2002.

. 2000. Split Dehumidification Unit for Modular Climate Changer™ Air Handlers, CLCH-PRB005-EN. La Crosse, WI: Inland Printing Company.

. 2001. “Dedicated Outdoor-Air Ventilation Systems,” Engineers Newsletter Live satellite broadcast, APP-APV008-EN (September 19, videocassette). La Crosse, WI: AVS Group. Available from http://trane.com/ bookstore/; accessed May 2002.

SYS-APM004-EN 129 References

: Hallstrom, A., PE, T. Robeson, D. Stanke, and B. Bradley. 1994. Designing an IAQ-Ready Air Handler System, SYS-AM-14. La Crosse, WI: Inland Printing Company. Available from http://trane.com/bookstore/; accessed May 2002.

: Stanke, D., B. Bradway, A. Hallstrom, and N. Bailey. 1998. Managing Building Moisture, SYS-AM-15. La Crosse, WI: Inland Printing Company. Available from http://trane.com/bookstore/; accessed May 2002.

: Stanke, D., J. Murphy, and B. Bradley. 2000. “Air-to-Air Energy Recovery,” Engineers Newsletter, volume 29, number 5. Available at http://www.trane.com/commercial/library/vol29_5/EN29-05.pdf; accessed May 2002.

: Stanke, D. and B. Bradley. 2000. “Dehumidify with Constant-Volume Systems,” Engineers Newsletter, volume 29, number 4. Available at http:// www.trane.com/commercial/library/vol29_4/ENews_29_04_082800.pdf; accessed May 2002.

: Stanke, D. and B. Bradley. 2001. “Dedicated Ventilation Systems,” Engineers Newsletter, volume 30, number 3. Available at http://www.trane.com/ commercial/library/vol30_3/enews_30_03.pdf; accessed May 2002.

: Murphy, J. 2002. Air-to-Air Energy Recovery in HVAC Systems, SYS-APM003-EN. La Crosse, WI: Inland Printing Company. Available from http://trane.com/bookstore/; accessed May 2002.

U.S. Environmental Protection Agency (EPA). 2001. Mold Remediation in Schools and Commercial Buildings, EPA 402-K-01-001 (March). Available from http://www.epa.gov/iaq/molds; accessed May 2002. ■

130 SYS-APM004-EN Index

a active adsorption 14–16 adjustable fan speed constant-volume systems 32–34 with mixed-air bypass 36–37 adsorption dehumidification 13–16 see also cold-coil dehumidification after-hours humidity control 17, 30, 86 airside economizing constant-volume systems 31–32 control methods 31 dedicated outdoor-air systems 87–88 effect on building pressurization 30, 69 overview 18 requirements of ANSI/ASHRAE/IESNA Standard 90.1 31–32, 89 variable-volume systems 69 air-to-air heat recovery coil loop with three coils 97 outdoor-air preconditioning 97, 98–99 parallel configuration 96–97 reheating conditioned outdoor air 95–98 series configuration 95–96 supply-air tempering 58–60 ANSI/ASHRAE Standard 55 4 ANSI/ASHRAE Standard 62 4 ANSI/ASHRAE/IESNA Standard 90.1 heat recovery for reheat 94 implications for dual-path air handlers 49–50 overview 7 reheat at VAV terminals 66 requirements for economizers 31–32, 89 restrictions for humidistatic controls (Section 6.3.2.3 excerpt) 8

b building maintenance, effect of moisture on 5 building pressurization constant-volume systems 30 dedicated outdoor-air systems 86–87 infiltration 3 overview 18 variable-volume systems 69

SYS-APM004-EN 131 Index

c central air-conditioning systems chilled water 12 packaged DX (direct-expansion) 12 remote dehumidification 10 chilled water systems “cold air” distribution 74 condenser-water heat recovery 53–56 design flexibility 12 total-energy wheels 27–28 chilled-water-temperature reset 35 chillers with heat-recovery condensers 55–56 circuiting configurations of finned-tube refrigerant coils 41–43 climate considerations 5–7 coil curves 11, 26, 104 cold supply air constant-volume systems 29–30 variable-volume systems 73–74 cold-coil dehumidification 10–12 see also adsorption dehumidification comfort zone 4–5 compressor cycling 26 condensate management 12 condensation causes of unwanted 1 cold-coil 10–12 source of moisture indoors 2 condenser-water heat recovery constant-volume systems 53–56 reheating conditioned outdoor air 95 supply-air tempering at VAV terminals 66 conditioned outdoor air (CA) cold 75, 76 delivered to a common hallway 93 determining the required dew point 115 determining total airflow 114 moisture content 77–79 neutral temperature 75, 76 neutral-temperature vs. cold 84–85 reheating with recovered heat 94–98 constant-volume systems adjustable fan speed 32–34 airside economizing 31–32 building pressurization 30 cold supply air 29–30 conditioning outdoor air separately 44–50 DX (direct-expansion) coil circuiting 41–43 humidity control during unoccupied periods 30 mixed-air bypass 34–37 overview 19 return-air bypass 37–41 supply-air tempering 50–60 see also full-load dehumidification performance and part-load dehumidification performance

132 SYS-APM004-EN Index

control strategies affecting dehumidification performance 17 chilled-water-temperature reset 35 conditioned-outdoor-air dry-bulb-temperature reset in dedicated outdoor-air systems 92–93 energy recovery in dedicated outdoor-air systems 99 outdoor-airflow reset 94 supply-air dew-point reset 90–91 supply-air dry-bulb-temperature reset in variable-volume systems 64–65 cooling coils used for dehumidification 10–12 cooling-load division for dedicated outdoor-air systems 111 cooling-load equation 103

d damage caused by moisture 5 dampness. See moisture sources inside buildings dedicated outdoor-air systems building pressurization 86–87 common hallway 93 configurations 75–76 division of cooling loads 111 dry, cold air 82–84, 119–120, 122–124 dry, neutral-temperature air 80–82, 117–118, 120–122 economizer cooling 87–90 humidity control during unoccupied periods 86 maximum space humidity 113–114 moisture content 77–79 multiple-space humidity control 92–93 neutral-temperature vs. cold conditioned outdoor air 84–85 preconditioning outdoor air with recovered energy 98–99 reheating conditioned air with recovered heat 94–98 required dew point for conditioned outdoor air 115 reset control strategies 90–94 rules of thumb for sizing equipment 111 selecting local HVAC terminals 116–124 separate air paths for constant-volume systems 44–46 single-space humidity control 92 variable-volume (VAV) systems 70–72 dehumidification adsorption using a desiccant 13–16 local 9 remote 10 restrictions imposed by ANSI/ASHRAE/IESNA Standard 90.1 8 via condensation on a cold coil 10–12 desiccant-based adsorption 13–16 desiccants 13 see also process air and regeneration air design weather data 5, 102, 112 desuperheaters 57 dew-point temperature and space humidity 23 direct control of humidity 44–60 dual-duct air distribution 67–68 dual-path air handler 47–50

SYS-APM004-EN 133 Index

DX (direct-expansion) systems coil circuiting 41–43 dehumidification in constant-volume applications 24–26 design flexibility 12 refrigerant heat recovery 57–58

e economizers. See airside economizing or waterside economizing enthalpy wheels. See total-energy wheels EPA. See U.S. Environmental Protection Agency equations cooling-coil capacity 106 mixed-air condition 122 mixed-air dry-bulb temperature entering the coil 104 sensible-cooling loads on local HVAC terminals 117 sensible-heat ratio (SHR) 102 supply airflow to space 103 equipment selection dedicated outdoor-air handlers 112–116 local HVAC terminals 116–124 evaporation as a source of moisture indoors 2 exhaust air for desiccant regeneration 13

f face-and-bypass dampers mixed-air bypass 34 refrigerant heat-recovery application 57 return-air bypass 37, 39 face-split refrigerant coils 41 fan-powered VAV terminals 66–67 finned-tube refrigerant coils 41 full-load dehumidification performance at peak dry bulb constant-volume systems 20–21 dedicated outdoor-air systems 77–78, 80–81, 82–83 psychrometric analysis demonstrated 102–107 variable-volume systems 62, 70–71, 73

h heat gain by return air 103 heat recovery ANSI/ASHRAE/IESNA Standard 90.1 94 outdoor-air preconditioning 98–99 reheating conditioned outdoor air 94–98 supply-air tempering 53–60 “heating” chiller for condenser heat recovery 54–55 hot-gas reheat coils 57

134 SYS-APM004-EN Index

humidity control “critical” space 114–115 multiple-space, dedicated-outdoor-air systems 92–93 separate path for outdoor air 44–50 single-space, dedicated-outdoor-air systems 92 supply-air tempering 50–60 worst-case condition 21, 101 humidity control during unoccupied periods 17, 68, 86 humidity indoors choosing a maximum limit 113–114 contributing factors 7

i indoor air quality (IAQ) effect of humidity 3–4 indoor humidity choosing a maximum limit 113–114 dependencies 7 reasons to control 3–5, 100 see also relative humidity infiltration 3, 18, 20 in-space air conditioners 9 intertwined refrigerant coils 41

l latent capacity degradation model 26 local dehumidification 9 local HVAC terminals 116 calculating the entering-mixed-air condition 121–122 sizing for dedicated outdoor-air systems 111 supply airflow 118, 120, 121, 123 supply-air dry-bulb temperature 118, 120, 121, 123

m maximum relative humidity ASHRAE recommendation 17 choosing a limit 113–114 finding the “worst-case” space 114–115 revision proposed by ASHRAE 4 microbial growth causes 1, 3 defending against 3, 17 minimizing unwanted moisture 3–4 minimum airflow settings for VAV terminals 63–64 mixed-air bypass application considerations 35 constant-volume systems 34–37 with adjustable fan speed 36–37 see also return-air bypass

SYS-APM004-EN 135 Index

mixed-air condition, plotting on a psychrometric chart 104, 121–122 mixed-air dry-bulb-temperature equation 104 mixed-air systems 19 see also constant-volume systems and variable-volume systems moisture sources inside buildings 2 moisture-related damage 5 mold indoors building maintenance 5 occupant health 3 remediation 3, 4, 17

n nighttime humidity control constant-volume systems 30 dedicated outdoor-air systems 86 overview 17 variable-volume systems 68

o outdoor air treated separately 44–50, 70–72, 75–99 outdoor-air preconditioning adsorption dehumidification 16 benefits 16 constant-volume systems 27–29 dedicated outdoor-air systems with energy recovery 98–99 outdoor-airflow reset 94 oversizing supply airflow, results of 25–26

p packaged DX (direct-expansion) systems dedicated outdoor-air systems 92–93 dehumidification in constant-volume applications 24–26 design flexibility 12 refrigerant heat recovery 57–58, 94–95 total-energy wheels 28–29 parallel, fan-powered VAV terminals 66 part-load dehumidification performance at peak dew point adjustable fan speed 33 comparison of DX (direct-expansion) coils 42 constant-volume systems 21, 24 dedicated outdoor-air systems 44–49, 70–72, 78, 81, 83–84 mixed-air bypass 34–35 psychrometric analysis demonstrated 107–110 return-air bypass 38 supply-air tempering 51 variable-volume systems 62, 71, 73

136 SYS-APM004-EN Index

part-load dehumidification performance on a mild, rainy day comparison of DX (direct-expansion) coils 43 constant-volume systems 21–22, 24, 33 dedicated outdoor-air systems 44–49, 70–72, 79, 82, 84 return-air bypass 38–39 supply-air tempering 51 variable-volume systems 63, 71, 73 passive adsorption 13–14 passive energy-recovery devices 27 peak dew-point condition analyzing dehumidification performance 5–6 constant-volume dehumidification 21, 24 dedicated outdoor-air dehumidification 78, 81, 83–84 variable-volume dehumidification 62, 71, 73 peak dry-bulb condition 5 constant-volume dehumidification 20–21 dedicated outdoor-air dehumidification 77–78, 80–81, 82–83 variable-volume dehumidification 62, 70–71, 73 plate-and-frame heat exchangers condenser heat recovery 54 waterside economizer cooling 89 portable dehumidifiers 9 preconditioning outdoor air adsorption dehumidification 16 benefits 16 constant-volume systems 27–29 dedicated outdoor-air systems with energy recovery 98–99 process air 13 productivity 4 psychrometric analysis demonstrated 101–110 psychrometric chart coil curves 11, 26, 104 described 10 mixed-air condition 104 return-air condition 105 sensible-heat ratio (SHR) 105 supply-air condition 104

r recovered heat ANSI/ASHRAE/IESNA Standard 90.1 94 outdoor-air preconditioning 27–29, 98–99 reheating conditioned outdoor air 94–98 supply-air tempering 53–60, 66 refrigerant coils 41 refrigerant heat recovery 57–58, 94–95 regeneration air 13 “reheat” at VAV terminals 66 reheating conditioned outdoor air (CA) 94–98 relative humidity ASHRAE-recommended range 4 contributing factors 20 limit recommended by ASHRAE 17

SYS-APM004-EN 137 Index

remote dehumidification 10 reset control strategies chilled water temperature 35 conditioned-outdoor-air dry-bulb temperature 92–93 dedicated outdoor-air systems 90–94 outdoor airflow 94 supply-air dew point 90–91 supply-air dry-bulb temperature in variable-volume systems 64–65 return-air bypass constant-volume systems 37–41 full coil face area 37–39 reduced coil face area 39–40 see also mixed-air bypass return-air condition, plotting on a psychrometric chart 105 return-air heat gain 103 room air conditioners 9

s selecting local HVAC terminals for dedicated outdoor-air systems 116–124 sensible-heat ratio (SHR) equation 102 plotting on a psychrometric chart 105 series, fan-powered VAV terminals 67 sidestream “heating” chiller for condenser heat recovery 54–55 simultaneous heating and cooling, restrictions on 8 sizing equipment in dedicated outdoor-air systems cold conditioned outdoor air to local HVAC terminals 122–124 cold conditioned outdoor air to space 119–120 dedicated outdoor-air handlers 112–116 neutral-temperature conditioned outdoor air to local HVAC terminals 120–122 neutral-temperature conditioned outdoor air to space 117–118 Standard 55. See ANSI/ASHRAE Standard 55 Standard 62. See ANSI/ASHRAE Standard 62 Standard 90.1. See ANSI/ASHRAE/IESNA Standard 90.1 supply airflow effects of oversizing 25–26, 100 equation 103 local HVAC terminals 118, 121–122 supply-air condition, plotting on a psychrometric chart 104 supply-air dew-point reset 90–91 supply-air dry-bulb temperature dedicated outdoor-air handlers 115–116 local HVAC terminals 118, 121 supply-air dry-bulb-temperature reset dedicated outdoor-air systems 92–93 variable-volume systems 64–65 supply-air tempering constant-volume systems 50–60 dual-duct air distribution 67–68 fan-powered VAV terminals 66–67 heating coils at VAV terminals 66 variable-volume systems 65–66

138 SYS-APM004-EN Index

t tempering supply air constant-volume systems 50–60 dual-duct air distribution 67–68 fan-powered VAV terminals 66–67 heating coils at VAV terminals 66 variable-volume systems 65–68 terminal units. See local HVAC terminals thermal comfort 4–5 total-energy wheels chilled water applications 27–28 desiccant regeneration for passive adsorption 13–14 packaged DX (direct-expansion) applications 28–29 preconditioning outdoor air 27, 98–99 treating outdoor air separately constant-volume systems 44–50 variable-volume systems 70–72

u U.S. Environmental Protection Agency (EPA) indoor air quality 3 mold remediation 4 underventilation and space humidity 22–23 unoccupied humidity control constant-volume systems 30 dedicated outdoor-air systems 86 overview 17 variable-volume systems 68

v vapor pressure desiccant performance 13 relative humidity 20 variable-volume systems airside economizing 69 building pressurization 69 conditioning outdoor air separately 70–72 delivering colder supply air 73–74 humidity control during unoccupied periods 68 minimum airflow settings 63–64 overview 61 supply-air tempering at VAV terminals 65–68 supply-air-temperature reset 64–65 VAV (variable-air-volume) terminals minimum airflow settings 63–64 see also local HVAC terminals ventilation and loads 22

SYS-APM004-EN 139 Index

w waterside economizing 89–90 water-source heat pumps refrigerant heat recovery 57–58 waterside economizer cooling 90 weather barrier 3 weather data 5, 102, 112

140 SYS-APM004-EN

Literature Order Number SYS-APM004-EN

Date December 2002

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