Architectural and Spatial Constraints. Air distribution products must fulfill both the functional requirement of conditioning space and the visual aesthetic determined by the architect. Architectural constraints may limit placement of air outlets and ductwork.

Heat Gain and Loss Characteristics. Large sensible heat loads can drive air movement in a space. Warm air is more buoyant and rises; cooler air is denser and descends to the ground. A zone may experience different heat loads depending on the season or time of day. The air distribution system must meet the varying heating and cooling needs throughout the building’s operation. Care should be taken to ensure cool surfaces, such as exterior windows, do not reach temperatures below dew point, or condensation will occur.

Thermal Comfort. Occupant thermal comfort depends on several variables, including air velocity, air temperature, thermal radiation, humidity, occupant metabolic rate, and occupant clothing. Air distribution systems that use higher air velocities and temperature differentials may create a greater risk of draft. Likewise, exterior windows with warm or cold surfaces can produce undesired thermal radiation to nearby occupants. For more information, see ASHRAE Standard 55-2020 and its user’s manual (ASHRAE 2016).

Acoustical Requirements. Sound emitted from inlets and outlets is directly related to the airflow quantity and free area velocity. The airflow sound intensity in a space also depends on the room’s acoustical absorption and the observer’s distance from air distribution devices. For more information, see Chapter 49 of this volume and Chapter 8 in the 2021 ASHRAE Handbook—Fundamentals.

Available Locations for Air Inlets and Outlets. Inlet and outlet characteristics are discussed in Chapter 20 of the 2020 ASHRAE Handbook—HVAC Systems and Equipment. This chapter discusses more specific application considerations for air inlets and outlets.

Code Requirements. Some applications (e.g., operating rooms) may require compliance with special or local codes that limit the selection and placement of certain types of air outlets.

The success of a mixed air distribution system depends primarily on supply diffuser location. Return grille location is far less critical than the location of air outlets except in applications where contaminant control and removal are a primary concern. In fact, the return air intake affects room air motion only in the area immediately around the grille. Measurements of velocity near a return air grille show a rapid decrease in magnitude as the measuring device is moved away from the grille face. Table 1 shows recommended maximum (to avoid excessive noise) return inlet face velocities as a function of grille location. Every enclosed space should have return/transfer inlets of adequate size per this table.

For stratified and partially mixed air distribution systems, it can be advantageous to place the return air inlet in the ceiling. ASHRAE Standard 62.1 allows ventilation effectiveness greater than 1.0 for some stratified and partially mixed air distribution systems in cooling mode if the return air inlet is located in the ceiling.

Supply air short-circuiting is normally not a problem if the outlet is selected to provide adequate throw and directed away from returns or exhausts. The success of this practice is confirmed by the availability and use of combination supply and return diffusers.

Air distribution systems affect not only indoor air quality (IAQ) and thermal comfort, but also energy consumption over the entire life of the project. Choices made early in the design process are important. ASHRAE Standard 90.1 provides energy efficiency requirements that affect supply air characteristics.

U.S. Green Building Council’s (USGBC) Leadership in Energy and Environmental Design (LEED^®) Green Building Rating System™ was originally created in response to indoor air quality concerns, and has evolved to include prerequisites and credits for increasing ventilation effectiveness and improving thermal comfort (new.usgbc.org/leed). These requirements and optional points are relatively easy to achieve if good room air distribution design principles, methods, and standards are followed.

Environmental tobacco smoke (ETS) control is a LEED prerequisite. Banning indoor smoking is a common approach, but if indoor smoking is to be allowed, ASHRAE Standard 62.1 requires that more than the base non-ETS ventilation air be provided where ETS is present in all or part of a building. Rock (2006) provides additional advice on dealing with ETS.

Ventilation effectiveness is affected directly by the room air distribution system’s design, construction, and operation, but is very difficult to predict. Many attempts have been made to quantify ventilation effectiveness, including ASHRAE Standard 129. However, this standard is only for experimental tests in well-controlled laboratories and should not be applied directly to real buildings.

Because of the difficulty in predicting ventilation effectiveness, ASHRAE Standard 62.1 provides a table of typical values that were determined through the experiences of its Standard Project Committee and reviewers or extracted from research literature; for example, well-designed ceiling-based air diffusion systems produce near-perfect air mixing in cooling mode, and yield an air change effectiveness of almost 1.0. More information on ASHRAE Standard 62.1 is available in its user’s manual (ASHRAE 2011).

Displacement and underfloor air distribution (UFAD) systems have the potential for values greater than 1.0. More information on ceiling- and wall-mounted air inlets and outlets can be found in Rock and Zhu (2002). Performance of displacement systems is described by Chen and Glicksman (2003), and UFAD is discussed in detail by ASHRAE (2013).

Air terminals, such as diffusers or grilles, may become unsightly over time because of accumulation of dirt on their faces (smudging). Instead of replacing air terminals, and thus requiring new materials and energy for manufacturing, they can often be cleaned in place to restore their appearance. Those that cannot be cleaned and must be replaced should be recycled, not discarded, to recover the various metals and other desirable construction materials.

The number and location of supply diffusers and return grilles and their operation can have a major effect on room airflow patterns and the distribution of airborne contaminants. As such, the following recommendations apply to all room air distribution systems (Khankari et al. 2020):

Mixed systems generally provide comfort by entraining room air into discharge jets located outside occupied zones, mixing supply and room air. Ideally, these systems generate low-velocity air motion (less than 0.25 m/s) throughout the occupied zone to provide uniform temperature gradients and velocities. Proper selection of an air outlet is critical for proper air distribution; improper selection can result in room air stagnation, unacceptable temperature gradients, and unacceptable velocities in the occupied zone, possibly leading to occupant discomfort or poor air quality.

The location of a discharge jet relative to surrounding surfaces is important. Discharge jets attach to parallel surfaces, given sufficient velocity and proximity. When a jet is attached, the throw increases by about 30% over a jet discharged in an open area. This difference is important when selecting an air outlet. For detailed discussion of the surface effect on discharge jets, see Chapter 20 of the 2021 ASHRAE Handbook—Fundamentals.

These systems are intended to maintain acceptable indoor air quality by mixing supply and room air (dilution ventilation). Supply air is typically a conditioned mixture of ventilation and recirculated air. Outlet type and discharge velocity determine the mixing rate of the space and should be a design consideration. The room’s return or exhaust air carries away diluted air contaminants. Space air ventilation rates are mandated under ASHRAE Standard 62.1, but supply airflow rates are often higher because of thermal loads.

Benefits of fully mixed systems include the following:

Limitations of mixed systems include the following:

Mixed air systems typically use either ceiling or sidewall outlets discharging air horizontally, or floor- or sill-mounted outlets discharging air vertically. They are the most common method of air distribution.

The way an airstream approaches an outlet is important. For good air diffusion, the inlet configuration should create a uniform discharge velocity profile from the outlet, or the outlet may not perform as intended.

The outlet usually cannot correct effects of improper duct approach. Many sidewall outlets are installed either at the end of vertical ducts or in the side of horizontal ducts, and most ceiling outlets are attached either directly to the bottom of horizontal ducts or to special vertical takeoff ducts that connect the outlet with the horizontal duct. In all these cases, devices for directing and equalizing the airflow may be necessary for proper direction and diffusion of the air.

ASHRAE research project RP-1335 (Landsberger et al. 2011) determined that a wide-open damper installed in the neck of a diffuser could add up to 8 NC to the cataloged NC value, depending on diffuser and damper types. Significantly closed balancing dampers can add more than 10 NC, depending on duct pressure and how far upstream it is installed. Table 2 gives forward throw asymmetries for various diffuser types, and Figure 2 compares the total pressure ratio of a diffuser with no damper versus a neck-mounted damper that is increasingly throttled.

Ceiling air outlets are tested using ideal installation conditions described in ASHRAE Standard 70: a minimum of three straight equivalent duct diameters before the diffuser inlet.

Field installations are often not ideal because of ceiling plenum limitations, architectural design, structural supports of the building, etc. ASHRAE research project RP-1335 (Landsberger et al. 2011) developed some useful application data that can be used by design engineers to predict throw, pressure drop, and sound caused by non-ideal installations. Tables 2, 3, and 4 give a general summary of application data from RP-1335. The multipliers provided in these tables may be applied to diffuser performance data measured from tests compliant with ASHRAE Standard 70-2006.

A fully mixed system creates homogeneous thermal conditions throughout the space. As such, thermal gradients should not exist in the occupied zone. Improper selection, sizing, or placement may prevent full mixing and can result in stagnant areas, or having high-velocity air entering the occupied zone.

Supply airflow requirements to satisfy space sensible heat gains or losses are inversely proportional to the temperature difference between supply and return air. The following equation can be used to calculate space airflow requirements (at standard conditions):

where

Q = required supply airflow rate to meet sensible load, L/s

qs = net sensible heat gain in the space, W

tr = return or exhaust air temperature, °C

ts = supply air temperature, °C

For fully mixed systems with conventional ceiling heights, the return (or exhaust) and room air temperatures are the same; for example, a room with a set-point temperature of 24°C has, on average, a 24°C return or exhaust air temperature.

The objective of air diffusion is to create the proper combination of room air temperature, humidity, and air motion to provide thermal comfort and acceptable indoor environmental quality in the occupied zone. There are three recommended methods of selecting outlets for mixed air systems using manufacturers’ data:

These selection methods are not meant to be independent. It is the designer’s choice as to which to start with, but it is recommended that at least two methods be used for any design.

Variation from accepted thermal limits (ASHRAE Standard 55), lack of uniform thermal conditions in the space, or excessive fluctuation of conditions in one part of the space may produce discomfort. Thermal discomfort can also arise from any of the following conditions:

Selection by Appearance, Flow Rate, and Sound Data. For a given appearance, flow rate, pressure drop, and sound level criteria, designers can select outlets from manufacturers’ catalogs, using the following steps:

Although these selections may meet the sound requirements for a project, the results do not fully address occupant comfort. Without evaluating the throw of the outlets or room air mixing, this selection method may result in excessive air velocities in the occupied zone, or limited mixing and resultant stagnation. It is recommended that the designer consider selection by isovel mapping or by comfort criteria in addition to selection by appearance, flow rate, and sound data. Either of these methods addresses resulting air motion in the occupied zone and occupant comfort.

Selection by Isovels and Mapping. Using manufacturers’ catalog throw data, a designer can predict the path of an outlet’s discharge jet. Most manufacturers’ catalogs list the distance a jet travels to reach a terminal velocity of 0.75 to 0.25 m/s. With this information, the designer can map the path of the discharge jet for a given outlet. This evaluation can prevent problems such as excessively high air velocities in the occupied zone, or stagnation in a given area. Note that most manufacturers’ throw data are based on isothermal supply air; the supply jet temperature is equal to the room air temperature. When using this mapping method, consider the positive or negative buoyancy of nonisothermal (heated or cooled) supply air. In both heating and cooling, a discharge jet should travel the distance shown in the catalog to a terminal velocity of 0.75 m/s without much influence from buoyancy. When evaluating a jet at lower terminal velocities (e.g., 0.5 to 0.25 m/s), consider buoyancy’s effect on the distance the jet will travel.

Horizontal Throw. A cooled confined air jet projecting along a horizontal surface travels a shorter distance than the equivalent volumetric isothermal air jet traveling the same path. If an outlet is selected so that the cooled horizontal jet does not have enough velocity to reach a vertical surface, the jet can separate from the horizontal surface and project down into the occupied zone, causing drafts and discomfort.

A heated confined air jet projecting along a horizontal surface travels farther than the equivalent volumetric isothermal jet traveling the same path. If an outlet is selected so that the heated, horizontal jet does not have enough velocity to reach a vertical surface, the jet can pool at the ceiling level, causing stagnation (no mixing) within the occupied space and resulting in high levels of contaminants, thermal stratification, and discomfort.

In a free jet scenario, expect similar results as for a confined jet, but if isothermal performance is based on confined jet installation, a 30% reduction in throw length from isothermal performance will occur, along with buoyancy effects.

Vertical Throw. A cooled air jet projecting vertically upward travels a shorter distance than the equivalent volumetric isothermal air jet traveling the same path. A heated air jet projecting upward travels farther than the equivalent volumetric isothermal air jet traveling the same path. For both cooled and heated air jets, the opposite is true for downward projection.

Mapping with acoustical selection. Combining selection by isovels allows discharge jet location and intensity in a space to be predicted. Outlet selection should be evaluated at the space’s typical operating points (i.e., maximum heating and cooling, and minimum heating and cooling).

The following steps may be used:

Other design considerations include the following:

Selection by Comfort Criteria T_0.25/L. Selection by isovels and mapping is effective at predicting the path of the discharge jet from an outlet and evaluating resultant occupant comfort. However, there is an established method to quantify occupant comfort for both cooling and heating conditions, based on space dimensions and isothermal catalog throw data. This method can be used to predict a space’s resulting air diffusion performance index (ADPI).

The comfort criteria T_0.25/L method was developed to predict occupant comfort using manufacturers’ isothermal catalog throw data (T, usually for 0.25 m/s terminal velocity) and the dimensions available for throw L on the plan view of a mechanical drawing. By using the ratio of T_0.25/L, the designer can predict the level of comfort with a single rating number, ADPI, which can provide further information about the comfort level in a space for results obtained from the NC and mapping selection methods.

Air Distribution Performance Index (ADPI). The air distribution performance index was developed as a way to quantify the comfort level in heating and cooling for a space conditioned by a mixed air system. ADPI uses the effective draft temperature collected at an array of points taken within the occupied zone to predict comfort. ADPI is the percentage of points in a space where the effective draft temperature is between –1.7 and +1.1 K for cooling and –2.2 and +2 K for heating. The acceptable air velocity is less than 0.36 m/s for heating and cooling. In addition, the acceptable vertical temperature gradient should be less than 3 K/m. High ADPI values generally correlate to high space thermal comfort levels with the maximum obtainable value of 100. Selecting outlets to provide a minimum ADPI value of 80 generally results in a well-mixed space.

The effective draft temperature provides a quantifiable indication of comfort at a discrete point in a space by combining the physiological effects of air temperature and air motion on a human body. The effective draft temperature t_ed (the difference in temperature between any point in the occupied zone and the control condition) can be calculated using the following equation for the cooling condition, proposed by Rydberg and Norback (1949) and modified by Straub (Straub and Chen 1957; Straub et al. 1956) in discussion of a paper by Koestel and Tuve (1955):

where

ted =effective draft temperature, K

tx =local airstream dry-bulb temperature, °C

tc =average (control) room dry-bulb temperature, °C

Vx =local airstream centerline velocity, m/s

Liu and Novoselac (2015) developed an effective draft temperature t_ed for the heating condition based on predicted mean vote (PMV) model specified in ASHRAE Standard 55. A high vertical temperature gradient occurs during heating when throws are insufficient, resulting in stagnation of supply air and unacceptable ventilation to the breathing zone.

where

ted =effective draft temperature, K

tx =local airstream dry-bulb temperature, °C

tc =average (control) room dry-bulb temperature, °C

Vx =local airstream centerline velocity, m/s

T_0.25/L Selection Method. This method uses the ratio of cataloged isothermal throw data at 0.25 m/s to the characteristic length for a given device (Table 5).

Each type of diffuser has different performance characteristics and therefore may provide a different ADPI value for the same conditions. Calculating T_0.25/L for a given outlet can predict the level of comfort for a space. Using Tables 6A and 6B, the designer can optimize not only the type of diffuser to select but also the size and capacity for both cooling and heating. The maximum ADPI listed in Tables 6A and 6B reflects the optimal value expected for the given outlet type and location. The maximum ADPI T₅₀/L represents the ratio corresponding with that maximum ADPI value.

Using T_0.25/L helps designers maximize space cooling comfort; however, this method is not meant to, nor may it be practical to, evaluate T_0.25/L values for each outlet on a project. The design guidelines in Tables 6A and 6B were developed from laboratory experiments in chambers lower than 3 m, with test diffusers symmetrically distributed (Liu and Novoselac 2014; Liu et al. 2016; Miller and Nash 1971). Therefore, attention should be paid to ceiling height of buildings (e.g., airport terminals) or highly asymmetric diffuser layouts.

Design Procedures. T_0.25/L can be used as a general tool to evaluate cooling comfort levels in a space, at the beginning of design to optimize outlet selection (as shown in the following steps), or at the end of the process to predict comfort levels in spaces designed using NC and mapping methods:

There are many other considerations to be made when designing a fully mixed system. For more information, see Rock and Zhu (2002).

Horizontal Discharge Cooling with Ceiling-Mounted Outlets. Ceiling-mounted outlets typically use the surface effect to transport supply air in the unoccupied zone. The supply air projects across the ceiling and, with sufficient velocity, can continue down wall surfaces and across floors, as shown in Figure 5. In this application, supply air should remain outside the occupied zone until it is adequately mixed and tempered with room air. Air motion in the occupied zone is generated by room air entrainment into the supply air (Nevins 1976).

Overhead outlets may also be installed on exposed ducts, in which case the surface effect does not apply. Typically, if the outlet is mounted 300 mm or more below a ceiling surface, discharge air will not attach to the surface. The unattached supply air has a shorter throw and can project downward, resulting in high air velocities in the occupied zone. Some outlets are designed for use in exposed duct applications. Typical outlet performance data presented by manufacturers are for outlets with surface effect; consult manufacturers for information on exposed duct applications.

Vertical-Discharge Cooling or Heating with Ceiling-Mounted Outlets. Vertically projected outlets are typically selected for high-ceiling applications that require forcing supply air down to the occupied zone. It is important to keep cooling supply air velocity below 0.25 m/s in the occupied zone.

There are outlets specifically designed for vertical projection, and it is important to review the manufacturer’s performance data notes to understand how to apply catalog data. Throws for heating and cooling differ and also vary depending on the difference between supply and room air temperatures.

Cooling with Sidewall Outlets. Sidewall outlets are usually selected when access to the ceiling plenum is restricted. Sidewall outlets that are within 300 mm of a ceiling and set for horizontal or a slightly upward projection provide a discharge pattern that attaches to the ceiling and travels in the unoccupied zone. This pattern entrains air from the occupied zone to provide mixing.

In some applications, the outlet must be located 0.5 to 1.25 m below the ceiling. When set for horizontal projection, the discharge at some distance from the outlet may drop into the occupied zone. Most devices used for sidewall application can be adjusted to project the air pattern upwards toward the ceiling. This allows the discharge air to attach to the ceiling, increasing throw distance and minimizing drop. This application provides occupant comfort by inducing air from the occupied zone into the supply air.

Some outlets may be more than 1.25 m below the ceiling (e.g., in high-ceiling applications, the outlet may be located closer to the occupied zone to minimize the volume of the conditioned space). Most devices used for sidewall applications can be adjusted to project the air pattern upward or downward, which allows the device’s throw distance to be adjusted to maximize performance.

When selecting sidewall outlets, it is important to understand the manufacturer’s data. Most manufacturers offer data for outlets tested with surface effect, so they only apply if the device is set to direct supply air toward the ceiling. When the device is 1.25 m or more below a ceiling, or supply air is directed horizontally or downward, the actual throw distance of the device is typically 30% shorter. Many sidewall outlets can be adjusted to change the spread of supply air, which can significantly change throw distance. Manufacturers usually publish throw distances based on specific spread angles.

Cooling with Floor-Mounted Air Outlets. Although not typically selected for nonresidential buildings, floor-mounted outlets can be used for fully and partially mixed-system cooling applications. In this configuration, room air from the occupied zone is induced into the supply air, providing mixing. When cooling, the device should be selected to discharge vertically along windows, walls, or other vertical surfaces. Typical nonresidential applications include lobbies, long corridors, and houses of worship. See also the description of underfloor air distribution in the Partially Mixed Air Distribution section of this chapter.

It is important to select a device that is specially designed for floor applications. It must be able to withstand both the required dynamic and static structural loads (e.g., people walking, loaded carts rolling across). Also, many manufacturers offer devices designed to reduce the possibility of objects falling into the device. It is strongly recommended that obstructions are not located above these in-floor air terminals, to avoid restricting their air jets.

Long floor-mounted grilles generally have both functioning and nonfunctioning segments. When selecting air outlets for floor mounting, it is important to note that the throw distance and sound generated depend on the length of the active section. Most manufacturers’ catalog data include correction factors for length’s effects on both throw and sound. These corrections can be significant and should be evaluated. Understanding manufacturers’ performance data and corresponding notes is imperative.

Cooling with Sill-Mounted Air Outlets. Sill-mounted air outlets are commonly used in applications that include unit ventilators and fan-coil units. The outlet should be selected to discharge vertically along windows, walls, or other vertical surfaces, and project supply air above the occupied zone.

As with floor-mounted grilles, when selecting and locating sill grilles, consider selecting devices designed to reduce the nuisance of objects falling inside them. It is also recommended that sills be designed in a way that prevents their use as shelves.

Perimeter Control Techniques. In many cases, it is advantageous to decouple perimeter and interior supply air sources, especially as new buildings trend toward full glazing exteriors. Often, interior spaces require cooling all year round due to personnel, electronic equipment, and lighting load requirements. Perimeter spaces, however, are more susceptible to the outside environmental conditions and can have major effects on interior space loads if not properly managed.

Heating and Cooling with Perimeter Ceiling-Mounted Outlets. When air outlets are used at the perimeter with vertical projection for heating and/or cooling, they should be located near the perimeter surface, and selected so that the published 0.75 m/s isothermal throw extends at least halfway down the surface or 1.4 m above the floor, whichever is lower. In this manner, during heating, warm air mixes with the cool downdraft on the perimeter surface, to reduce or even eliminate drafts in the occupied space.

If a ceiling-mounted air outlet is located away from the perimeter wall, in cooling mode, the high-velocity cool air reduces or overcomes the thermal updrafts on the perimeter surface. To accomplish this, the outlet should be selected for horizontal discharge toward the wall. Outlet selection should be such that isothermal throw with terminal velocity of 0.75 m/s should include the distance from the outlet to the perimeter surface. For heating, the supply air temperature should not exceed 8.5 K above the room air temperature.

Be conscious of room design. If there are shelves or desks against or near the perimeter surface the high velocity air jet could disturb items on the surfaces or result in drafts.

Perimeter Heating with Floor/Sill-Mounted Outlets. The outlets used in this configuration must be located near the perimeter surface. In heating mode, it is important to select a vertical throw such that it overcomes the down draft produced by a cold perimeter surface. The amount of heat generated from the outlet must be enough to overcome or mix with this down draft airstream to avoid ankle-level draft issues in the occupied space.

DV systems (Figure 6) use very low discharge velocities, with diffusers typically sized to provide outlet velocities between 0.20 and 0.35 m/s. In addition to the low velocity discharge, the temperature of the supply air is also different from a fully mixed system, with temperatures generally above 16°C; lower temperatures may be used in industrial applications, exercise or sports facilities, and transient areas where comfort concerns are minimal. This cool supply air is more dense than the ambient air and drops to the floor after discharge, whether from floor, low sidewall, or ceiling mounted locations, spreading across the lower level of the space (typically less than 0.2 m in height).

As convective heat sources (Figure 6) in the space transfer heat to the cooler air around them, natural convection currents form and rise along the heat transfer boundary. Without significant room air movement, these currents rise to form a convective heat plume (thermal plume) around and above the heat source; as the plume rises, it expands by entraining surrounding air. Its growth and velocity are proportional to the heat source’s size and sensible load, as well as the temperature of the ambient air above it. As the plume rises, ambient air from below and around the heat source fills the void. An occupant in a DV system entrains the cool, conditioned air directly into their breathing zone. As the occupant exhales, the spent air, being warmer and more humid than the ambient air, is pulled out of the breathing zone by the rising plume. Convective heat from sources located above the occupied zone has little effect on occupied-zone air temperature.

At a certain height, where plume temperature equals ambient temperature, the plume dissipates and spills horizontally. Two distinct zones are thus formed in the room: a lower occupied zone with little or no recirculation flow (close to displacement flow), and an upper zone with recirculation flow. The boundary between these two zones is often called the shift zone (or stratification height). The height of this boundary layer between the upper and lower zones is determined based on the convection flow rates of the thermal plumes in relation to the airflow rate supplied by the diffusers. In a DV system, increasing the airflow rate reduces stratification and raises the boundary layer height, with a decrease in airflow providing the opposite effect. Actual and simplified representations of the temperature gradient in the space are shown in Figure 6.

DV systems can be modeled as shown in Figure 7. A thin layer of conditioned supply air (typically between 100 and 200 mm) lies adjacent to the floor. Directly above this layer of conditioned air is the lower zone, in which both ambient air temperature and contaminant concentration levels increase with height; this relationship is mostly linear. As the air transitions through the boundary layer, the upper, unoccupied zone contains a pool of warm, used, and/or contaminated air. This upper zone may or may not form, depending on the supplied airflow rate in proportion to the volume of thermal plumes rising through the space.

Thermal plumes created by heat sources, in conjunction with thermal stratification within the space, allow DV to be very effective at removing airborne contaminants that are equal to or lighter than the ambient air (e.g., respiratory-produced contaminants, tobacco smoke). The upward momentum of room air created by thermal plumes from heat sources drive an overall upward momentum of air, which displace contaminants out of the breathing zone. This typically results in a concentration of contaminants above the occupied zone greater than that in the breathing zone. This has been recognized in ASHRAE Standard 62.1-2022, which allows DV systems to have a zone air distribution effectiveness E_Z value of 1.2, compared to maximum values of 1.0 for mixed air systems. Thus, the designer can decrease the required outdoor air by 17% when using a DV system, which may result in energy savings for the system. Care should be taken when selecting DV systems when it is known that there will be hazardous airborne contaminants that are heavier than the ambient air. In these systems, a fully mixed system or dilution ventilation may be recommended.

Figure 8 shows thermal temperature gradients that might be expected for a classroom with a 3 m ceiling, served by DV. If loads are typical to the application and proper space airflow is supplied, Skistad et al. (2002) indicate that approximately 50% of the total temperature difference between supply air and return or exhaust air is dissipated in clear zone(s) next to the outlet(s). The other half of the temperature gradient is the space temperature gradient (STG), assumed to be linear with air temperature, increasing gradually from floor to ceiling. For applications where contaminant control and removal are a primary concern, refer to the Application Guidelines section of this chapter.

For stationary, low-activity occupants, keep supply air temperatures above 16°C. When occupants are very near outlets (e.g., in underseat delivery), it is recommended to keep supply air temperatures at or above 18°C.

Displacement outlets are designed for average outlet velocities between 0.20 to 0.35 m/s, and are typically mounted in a low sidewall or floor location. Returns or exhausts should be located as high as possible in the space to promote upward momentum of the air, and should efficiently remove as much of the stratified zone as possible; they also should be equally distributed throughout the space to encourage equal air movement and discourage cross flow in the stratified zone. Intensive heat sources may benefit from having returns located directly above them to quickly remove generated heat from the room.

Displacement outlets are available in a number of configurations and sizes. Some models are designed to fit in corners or along sidewalls, or stand freely as columns. It is important to consider the degree of flow equalization the outlet achieves, because use of the entire outlet surface for air discharge is paramount to minimizing clear zones and maintaining acceptable temperatures at the lower levels of the space.

Stationary occupants should not be subjected to discharge velocities exceeding around 0.25 m/s because air at the ankle level within this velocity envelope tends to be quite cool. As such, most outlet manufacturers define an adjacent zone (also called a clear zone) in which locating stationary, low-activity occupants is strongly discouraged, but transient occupancy, such as in corridors or aisles, is allowed. Occupants with high activity levels may also find the clear zone acceptable.

For a typical space, a single displacement diffuser can penetrate around 8 to 9 m. Spaces that are wider than this or have high load densities should look at providing displacement diffusers on multiple walls, or free standing (e.g., around columns) for even airflow distribution. A computational fluid analysis (CFD) may be warranted for more complex spaces to validate the design. Contact the diffuser manufacturer for further diffuser selection assistance.

Benefits of displacement ventilation systems include the following:

Some applications do not favor the use of DV systems. Small offices, especially with perimeter exposures, often do not have room for the large outlets that may be required. The following types of areas may be better served by a mixed system:

Unlike mixed systems, outlets in thermal displacement systems discharge air at very low velocities, resulting in very little mixing. As such, design of these systems primarily involves determining a supply airflow rate to manage the thermal gradients in the space in accordance with ASHRAE comfort guidelines. ASHRAE Standard 55 recommends that the vertical temperature difference between the ankle and head levels of seated occupants be no more than 3 K and 4 K for standing occupants to maintain an acceptable degree of occupant satisfaction.

Similar to mixing system diffusers, inlet conditions (including presence of dampers, elbows, flex duct, etc.) can have a negative effect on performance, primarily noise (NC) and pressure drop. Inlet conditions should match ASHRAE Standard 70 testing conditions as closely as possible to mimic catalogue performance data. DV diffusers are typically designed with internal baffles or nozzles, which should not be affected by inlet conditions at relatively low plenum velocities.

DV system design is somewhat different than for mixing ventilation. For mixing ventilation systems, where air is mixed relatively evenly throughout the space, the return/exhaust air temperature is assumed to equal the space temperature. In displacement ventilation systems, the space is divided into two vertical zones. The desired space air temperature is maintained only in the lower zone and is always higher in the upper zone because of the temperature stratification created by natural convection.

ASHRAE research project RP-949 (Chen et al. 1999) developed a calculation method for determining supply air volume, air temperature, and other design parameters specifically for DV systems. Using these calculations, the maximum temperature stratification, as outlined in ASHRAE Standard 55, is not exceeded. Example 4 uses the calculations outlined in RP-949 and the procedure presented by Chen & Glicksman (2003). This research project is intended for use with typical office and classroom spaces; for larger spaces such as atriums and theaters, careful consideration is necessary, and a computational fluid dynamic (CFD) analysis is recommended to verify design. The Federation of European Heating and Air Conditioning Associations (REHVA) also developed two procedures for calculation of air volume in DV systems; see the REHVA design guide for further information on these methods.

Space temperature gradient (STG) is affected by the strength and location of heat sources in the space, heat exchange by radiation between surfaces in the space, and supply airflow. The design procedure presented in this section is based on Skistad et al.’s (2002) simplified method of estimating temperature gradient (Figure 8). This method is applicable for typical spaces with a ceiling height up to 3.7 m, such as classrooms, office spaces, and meeting rooms. When designing more complex spaces, computational fluid dynamics (CFD) software programs may be used (see Chapter 13 of the 2021 ASHRAE Handbook—Fundamentals for more information).

The thermal gradient relationships illustrated in Figure 8 can be used to establish an acceptable supply-to-return air temperature differential Δt_SR from which the supply airflow rate is calculated. Because the space temperature gradient is assumed to be linear, the occupied gradient in the occupied zone is proportional to the volume of the space it represents. For example, if return height is located at height 3 m and the occupied zone is 1.2 m high, its gradient comprises 40% of the space temperature gradient. The temperature difference between room air at the top of the occupied zone and the supply air is therefore 70% of Δt_SR.

Determining an acceptable Δt_SR should consider both the room-to-supply temperature differential and the occupied zone temperature gradient (as limited by ASHRAE Standard 55).

In general, high-ceiling applications allow larger supply-to-return air temperature differentials, because the occupied zone is a smaller percentage of total room air volume. However, the differential may be reduced by limitations on supply air temperature, as shown in Example 5.

The supply airflow rate Q to achieve a given Δt_SR is calculated from Equation (1).

Thermal displacement ventilation systems typically have higher return air temperatures than mixed systems. Thus, they may allow extended periods of air- or water-side economizer operation, especially in mild, relatively dry climates.

Thermal displacement ventilation systems are commonly used in applications such as

DV systems rely on buoyancy of the cool supply air and thermal plumes of the heat sources to drive stratification in the space. Due to this, DV is typically a cooling-only method of room air distribution. When heating is required, the warmer (more buoyant) air, in conjunction with the low discharge velocity, can mean that the air may not effectively condition the occupied zone, or may completely bypass the occupied zone to the returns. This can have a negative effect on both the thermal comfort and indoor air quality in the space. Limited heating differentials may be acceptable depending on a number of factors. ASHRAE Standard 62.1-2019 prov ides effectiveness guidance, and ASHRAE research project RP-1373 (Jiang and Chen 2009) explores the ventilation effectiveness of various DV systems in heating and cooling scenarios.

For heating in cold climates, a separate decoupled system is generally recommended. Displacement ventilation can be used successfully in combination with perimeter fan coils, hydronic systems, or radiators and convectors installed at exterior walls to offset space heat losses. Radiant heating panels and heated floors can also be used. Using this hybrid approach, the DV diffusers can provide adequately cool ventilation air, with the auxiliary devices handling the heating load.

Displacement diffusers with integrated heat typically have a separate heating plenum featuring a heating coil and are often used for perimeter applications. Heat/cool changeover diffusers can also be used. For this design, the heating portion of the diffuser will not provide a displacement pattern, so in this case ventilation effectiveness is compromised.

Care is also needed when using DV systems in high perimeter cooling load applications. DV diffuser mounting locations should be carefully considered to minimize the risk of conditioned air being drawn toward the strong thermal plume of the exterior surface. It may be better to place diffusers along an interior surface to ensure that the conditioned air reaches interior occupants before conditioning the exterior envelope loads.

Thermal displacement ventilation systems can be either constant (CV) or variable air volume (VAV). A thermostat in a representative location in the space or return plenum should determine the delivered air volume or temperature. If the time-averaged requirements of ASHRAE Standard 62.1-2019 are met, intermittent on/off airflow control can be used.

Avoid using DV and mixed air systems in the same space, because mixing destroys the natural stratification that drives the thermal displacement ventilation system. DV systems can be complemented by hydronic systems such as chilled floors. Use caution when combining chilled ceilings, beams, or panels with fully stratified systems, because cold surfaces in the upper zone of the space may recirculate contaminants stratified in the upper zone back into the occupied zone.

It is often beneficial to couple DV systems with hydronic products. Displacement chilled beams are available that integrate sensible heating and cooling coils. for more information about chilled beams, see the Chilled Beams section of this chapter.

Another common application is to use DV systems to provide only the space ventilation and cooling, while a complementary hydronic system satisfies the majority of the cooling load. In this case, passive (not active) sensible cooling devices (e.g., passive beams, radiant ceiling panels) should be used to minimize the disruption of the natural stratification that drives the DV system.

When thermal displacement systems are used in humid climates, it may be necessary to dehumidify and reheat supply air to a temperature that will sufficiently maintain the desired space conditions. As with all HVAC air systems’ design, a psychrometric analysis is advised.

Chen and Glicksman (2003) provide additional information on fully stratified air distribution systems.

Supply air is discharged, usually vertically, at relatively high velocities and entrains room air in a similar fashion to outlets used in mixed air systems. This entrainment, as shown in Figure 9, reduces the temperature and velocity differentials between supply and ambient room air. This discharge results in a vertical plume that rises until its velocity is reduced to about 0.25 m/s. At this point, its kinetic energy is insufficient to entrain much more room air, so mixing stops. Because air in the plume is still cooler than the surrounding air, the supply air spreads horizontally across the space, where it is entrained by rising thermal plumes generated by nearby heat sources.

Research and experience have shown that the amount of room air stratification varies depending on design, commissioning, and operation. Control of stratification includes the following considerations:

In practice, successful installation requires an optimal balance of these issues (Webster and Bauman 2006).

Figure 9 shows one example of the resulting room air distribution in which the room air is mixed in the lower mixed zone, which is bounded by the floor and the elevation (throw height) at which the 0.25 m/s terminal velocity occurs. At this elevation, stratification begins to occur and a linear temperature gradient, similar to that found in thermal displacement systems, forms and extends through the stratified zone. As with thermal displacement ventilation, convective heat plumes from space heat sources draw conditioned air from the lower (mixed) level through the stratified zone and to the overhead return location. A third zone, referred to as the upper mixed zone, may exist where the volume of rising heat plumes terminate. Although velocities in this area are quite low, the air tends to be mixed.

Partially mixed systems’ ventilation and contaminant removal efficiencies vary considerably. Restricting mixed conditions to below the breathing level results in most respiratory-associated contaminants being conveyed directly to the overhead return by heat plumes rising from occupants. If the lower mixed zone extends above the breathing level, contaminants are entrained and horizontally transmitted across occupied levels of the space, as occurs in mixed air (dilution ventilation) systems.

According to ASHRAE Standard 62.1, these systems may have zone air distribution effectiveness E_Z values that exceed those of fully mixed systems. For applications where contaminant control and removal are a primary concern, see the section on Application Guidelines.

One outlet type is a swirl diffuser with a high-induction core, which induces large amounts of room air to quickly reduce supply to ambient air velocity and temperature differentials. Supply air is injected into the room as a swirling vertical plume close to the outlet. Properly selected, these outlets produce a limited vertical projection of the supply air plume, restricting mixing to the lower portions of the space. Most of these outlets allow occupants to adjust the outlet airflow rate easily. Other versions incorporate automatically controlled dampers that are repositioned by a signal from the space thermostat and/or central control system.

Another category includes more conventional floor grilles designed for directional discharge of supplied airflow. These grilles may be either linear or modular in design, and may allow occupants to adjust the discharge air pattern by repositioning the core of the outlet. Most floor grilles include an integral actuated damper or other means to automatically throttle the volume of air in response to the zone conditioning requirements.

Room air induction allows UFAD diffusers to comfortably deliver supply air a few degrees cooler than possible with outlets used for thermal displacement ventilation outlets. Keeping clear or adjacent zones above and around the diffusers, where stationary occupants should not reside, is recommended. Outlet manufacturers typically identify such restrictive areas in their product literature.

Partially mixed systems are commonly used in applications such as the following:

Many UFAD systems can be classified as partially mixed systems. These systems are popular because of their relocation flexibility when used in conjunction with raised-access flooring systems. Outlet accessibility also allows easy occupant adjustment of space airflow delivery. The cavity beneath the access floor tiles is generally pressurized and used as a supply air plenum. Supply outlets placed in access floor tiles are commonly tapped directly into the pressurized plenum, but may be ducted from a fan-assisted terminal unit mounted beneath the floor.

Benefits of UFAD systems include the following:

Limitations to consider are

As with thermal displacement systems, design involves determining a supply airflow rate that limits thermal gradients in the occupied zone in accordance with ASHRAE Standard 55 guidelines; that is, the vertical temperature difference between the ankle and head level of space occupants should be no more than 3 K if a high degree (>95%) of occupant comfort is to be maintained.

UFAD systems usually pressurize the open plenum space under the access floor, this way the diffusers do not need to have an inlet. If linear floor diffusers with plenum boxes are utilized, they are usually made with very large rectangular inlets that are the same size as the linear length of those plenums. For ducted installations, use the same inlet criteria as the other air distribution systems.

The design of partially mixed air distribution systems requires identifying both thermal and contaminant removal objectives:

There are several ways to provide perimeter air distribution using a partially mixed air system. Air distribution in the perimeter zone is highly influenced by weather and solar loads at the building envelope; this means that different amounts of cooling or heating may be needed at different times, independent of interior zone load needs. The most common scenario is when cooling is needed at the interior zone and heating is needed at the perimeter zone; both internal and perimeter loads can be met using a zone divider or a dedicated plenum system. See ASHRAE (2013) for more information on how to handle perimeter zone applications.

The objective of partially mixed systems is to condition the air in the occupied zone while allowing stratification to naturally occur. By allowing this stratification, some of the space heat gain can be removed by return or exhaust instead of by supply air delivery to the space. If the supply airflow rate and sensible heat gains affecting the lower zone are balanced, an acceptable temperature gradient (<3 K) can be achieved in the occupied zone. Supply airflow beyond that required by these heat gains reduces the degree of stratification shown in Figure 9. If the supply airflow rate is insufficient, excessive vertical space temperature gradients may occur.

Accurate calculation of the space design supply airflow rate requires analysis of all space sensible heat gains to determine their contribution to the lower zone. Although there is not yet a single recognized procedure for calculating these airflow rates, most UFAD equipment manufacturers offer guidance.

The ASHRAE UFAD Guide (2013) includes a thorough discussion of issues involved in the design, application, and commissioning of UFAD systems. Some considerations include the following:

The design methodologies provided in this section cover textile air dispersion systems, though the same principles can be applied to systems constructed with metal. Textile air dispersion systems are low pressure extended plenum systems with pressurized tubing and air distributed along the path of least resistance. Suggestions are provided for material selection and dispersion style, support and structure of the textile air dispersion system, and sizing/venting locations of the system. Consult with the manufacturer and specific product data and performance information when designing and specifying textile air dispersion systems.

These systems can also be found designed into the plenum space of raised floor applications. In these applications, the designer must consider that the system is exposed in the space that it is conditioning and delivering air to. Air outlet throws of the textile air dispersion system must be designed for the plenum space in which they are located.

A version of these systems is also being used for return air (negative pressure). This requires a sturdy framework that holds the outer walls of the product open to prevent the restriction from becoming too great and restricting airflow beyond design intent. Maximum internal velocity of 5 m/s and –31 Pa static pressure

Various supply air outlet styles are available: porous fabric weave used as an outlet, microperforations, linear vents, orifices, and nozzles. Many of these air outlet styles can be used together to achieve specific results.

Porous Fabric Weave (Figure 10). Air is delivered through the weave of the fabric. This can result in air velocities of less than 0.15 m/s at the material surface. The ability to achieve such a low face velocity makes this the preferred dispersion style where displacement ventilation is needed. This dispersion style is ideal for food processing, cleanrooms, and laboratory environments where elimination of drafts and uniform air distribution is required, and can be used in combination with other types of air outlets to achieve the specified room airflow. Cool air drops to the floor and then spreads across the lower level of the space.

Porous fabric ducts can be used in combination with most linear vents and orifice applications. Air throw distance is generally lower than with linear vents, and may not be appropriate for heated supply air applications.

Microperforation Outlets (Figure 11). Air is delivered through laser-cut microperforations, generally smaller than 0.4 mm. As with porous fabric, air velocities of less than 0.15 m/s can be realized 0.61 m from the duct. Microperforations offer low-velocity distribution in the occupied zone without forcing the majority of the air through the weave of the material. Further, the microperforations can be uniformly located along 360°, or within a specific area or side of the air dispersion device, allowing designers to disperse air exactly where it is needed. This method can be used for displacement ventilation and isothermal/makeup air, but is limited for heating applications (when using directed microperforations for heating, consult manufacturer for throw data). This dispersion style is also ideal for food processing, cleanrooms, and laboratory environments where elimination of drafts and lower-velocity air is required. Cool air can be more evenly distributed in the space using directed microperforations rather than dropping or “dumping” due to buoyancy, directly below the diffuser. Microperforations can be used in combination with most other flow models to modify throw and velocity to meet design requirements.

Linear Vent Outlet (Figure 12). Air is delivered through a linear vent outlet, which generally consists of many small outlets in a linear pattern. This provides uniform airflow with throw suitable for commercial and retail spaces, schools, and theaters. Generally, medium air throw distance can be achieved.

Orifice and Nozzle Outlets (Figure 13). Air is delivered through orifices or venturi-shaped nozzles providing jet-type air distribution for applications such as gymnasiums, pools and manufacturing facilities. This style generally has a higher air throw distance than other outlet styles. Nozzles can direct air perpendicularly away from the surface of the duct. Adjustable nozzles allow for changing the direction and/or flow rate and throw. Nozzles generally have the least entrainment of longer-throw outlets. Manufacturers have many specific options for this style of air outlet, and the manufacturer’s data must be consulted.

Three shapes are commonly available: cylindrical, half circle (D-shape), and the quarter circle, as shown in Figure 14. Cylindrical systems are typical for open ceiling spaces and are mounted using a tension cable or suspended aluminum track suspension system. The half circle can be installed against a ceiling or wall. Manufacturers also have many custom shapes that can be used to solve installation challenges.

Air dispersion system ducts, classified by Underwriters Laboratories as an air distribution device, should have a Class 1 rating per UL Standard 723. The maximum flame spread/smoke developed index is currently 25/50.

Additional important properties in selecting a material for air dispersion systems are durability and aesthetics. Durability includes not only the environment, but the design. Supports and static pressure must hold fabric air dispersion system materials in tension to prevent fabric wear and tear. The maximum velocity in a system should not exceed the manufacturer’s rated velocity; excessive air velocities cause fabrics to fail prematurely. The International Mechanical Code^® (ICC 2018) and the Uniform Mechanical Code (IAPMO 2018) require that air dispersion systems be listed and labeled in compliance to UL Standard 2518.

The use of a slightly permeable material allows dispersion of air below dew point without the risk of condensation forming on the fabric air dispersion system. The duct is pressurized, and air is forced through the surface of the material, forming an insulating barrier of cooler air around the duct, preventing warm moist air from contacting the cold surface of the air dispersion system and thereby preventing water droplets from forming on the surface. The permeated air is often induced by outlets on the system and entrained into the supply air exiting the outlets.

Nonpermeable materials can be used where risk of condensation is low and applied in a way similar to uninsulated single wall metal ducts. These are often used in dryer climates or for extensive systems.

Polyester is typically the preferred material for fabric air dispersion systems because it has very low hygroscopic properties, is a synthetic material that is not a food source for organisms, and is easy to launder. Several manufacturers offer materials with complementary properties, such as antimicrobial treatment (to prevent microbial growth), antistatic for use around sensitive electronics and in explosion proof facilities, and fabrics that do not “shed” filament particles for use in cleanrooms and critical laboratories.

Textile air dispersion system manufacturers offer many different types of suspension systems to address the various applications in which they are used (Figure 15). Typical types include

Hold-Open and Fabric Retension Systems (Figure 16). Textile air dispersion systems can also be held open (restricting their potential to collapse onto themselves when the HVAC unit is not providing airflow) with rings, hoops, and arcs.

Fabric Tension Systems (Figure 17). Fabric tensioning systems hold the fabric in place regardless of whether the system is pressurized. These systems increase fabric longevity by reducing movement of the fabric. Some hold the fabric in such a way that air velocities in the product can be increased to a higher maximum (which results in smaller diameter sizes) without the fabric wall being unstable and fluttering, which could lead to premature failure.

A fabric air dispersion system performs as both a duct and a diffuser. When designing a textile air dispersion system, keep the layout as simple as possible, preferably with straight runs. Because porous fabric duct, linear vents, and orifice outlets can be integrated into all sections, system design may vary significantly while providing adequate air dispersion.

Metal ductwork before the fabric inlet should have the same air velocity as the fabric duct, and, when possible, it is recommended to use 3× the diameter of straight metal duct. If there is a transition preceding the fabric inlet, it is recommended to use 1.5× the diameter radius elbows and turning vanes in sheet metal transitions before fabric. High velocity (8 m/s) and turbulent conditions immediately before the system can cause excessive fabric movement and premature wear.

There is little need to reduce diameters because air dispersion systems are essentially extended plenums that can maintain static pressure due to continuously occurring static regain. Custom fittings should be coordinated with the manufacturer. It is recommended that end caps be one diameter from a wall to maintain clearance for systems and allow for movement and ease of installation.

Fittings. Straight duct lengths and fittings are connected together using a circumference zipper, which is affixed with its start/stop typically located at the top center, and includes a circumferential fabric overlap to conceal the zipper. Typical lengths of zippered sections are sized for easy laundering and installation. Longer sections are broken into multiple lengths.

Elbows. The typical centerline radius of an elbow is figured by multiplying the cross-sectional diameter by 1.5. For example, a 610 mm diameter air dispersion system elbow would have a centerline radius of 914 mm. The number of gores depends on the angle of the turn, as shown by Figure 18.

Transitions. Reducing transitions are available in concentric, top flat, or bottom flat configurations (Figure 19). Transition length varies and is based on the change in diameter, such that the total angle does not exceed 30°.

Tees. Tees are saddle type, and the branch requires a zipper for attachment. Typical tee arrangement is shown in Figure 20. It is recommended that tees be located at least 1.5× the outlet diameter from end caps (Figure 21).

Cross, Capped. See Figure 22 for the capped cross, which is SD5-20 in the ASHRAE (2017a) Duct Fitting Database (DFDB).

Dampers and Static Regain Devices. These fittings are used to address multiple objectives. The fittings offer engineered resistance to balance airflow in multiple runs and branches, reduce turbulence, reduce inflation pop, reduce noise, reduce movement from abrupt start-ups, and balance static regain. Locate flow devices at (1) inlet collars, (2) one-third point of a straight section, and (3) after take-offs and elbows. A flow device at the inlet collar reduces fluttering due to turbulent airflow entering the air dispersion system. At the first zipper, the flow device reduces the force of initial inflation and reduces “pop;” a fabric tensioning system can also eliminate any movement, deflation, or noise due to the HVAC system cycling on and off. Straightening out velocity profiles also reduces wear in a fabric duct system. Examples of dampers and static regain devices are shown in Figure 23.

Cylindrical Air Dispersion Systems. The recommended design velocity is 8.1 m/s for a system without fittings, and 7.1 m/s for a system with fittings. Use lower velocity (5.1 to 6.1 m/s) for noise sensitive areas. If the diameter is too large, design the system for multiple runs.

Half Circle (D-Shape) Air Dispersion Systems.

D-shape air dispersion systems should be sized so that the fabric duct velocity does not exceed the rigid supply duct velocity. The maximum velocity for a straight run is 5.1 m/s. For a system with a fitting, the recommended velocity is 4.1 m/s. A lower inlet velocity to the fabric duct reduces stress and noise.

Flow can be any split (e.g., for a top inlet in the middle, split is 50/50; for a top inlet a third of the distance from one end, split is 33/67), including flow from the end. The D-shape diameters and the duct diameters are good for a split to 38 to 62%. For a greater split, the inlet duct diameter needs to be reduced so the branch duct velocity does not exceed the design inlet duct velocity (4.1 or 5.1 m/s).

To design air dispersion systems, follow the steps below:

Filtration. It is essential to filter incoming air before it reaches the textile air dispersion system. An efficient filtration system means less cleaning, resulting in a longer product life. Higher filtration (MERV 8) is recommended for low-velocity outlets (porous fabric and microperforations). Textile air distribution systems are launderable.

Pressure Required for Inflation. For proper inflation, the static pressure at the inlet to a fabric air distribution system should be no less than 1.3× the inlet velocity pressure.

Single-Duct ATUs. Single-duct ATUs are used to regulate airflow to a conditioned space. Primary air is ducted from an air handler. The basic single-duct unit consists of an airflow regulator and may also include an actuator, an airflow-measuring device, selected controls, and heating coils.

Single-duct ATUs can be applied to pressure-dependent or -independent systems. Pressure-dependent units consist, at minimum, of an airflow regulator, thermostat, and actuator. The thermostat dictates the damper position based on room demand, which is independent of system pressure. Using pressure-dependent ATUs can cause system pressure fluctuation, which must be considered. Current ventilation requirements may preclude the use of pressure-dependent ATUs.

Pressure-independent units consist of, at minimum, an airflow regulator, airflow measuring device, controller, room sensor, and actuator. This is a closed-loop control system that measures and regulates airflow based on room demand. Pressure-independent ATUs allow a minimum airflow set point to be established to enable proper ventilation to the zone.

Pressure-independent single-duct ATUs can be used for constant- or variable-volume applications. Variable-volume applications often include controls that limit maximum and minimum airflow set points determined by load, acoustical, and zone ventilation requirements.

Dual-Duct ATUs. Dual-duct ATUs function similarly to single-duct terminals but are composed of two sets of airflow regulators merging two airstreams into a single mixed airstream. These are typically used to mix cold and hot primary air ducts to a single warm discharge air duct.

Series Fan-Powered ATUs. These fan-powered ATUs have a continuously operating fan during occupied mode, supplying either a constant or variable volume of air to the space, and they are typically installed in the ceiling plenum. Primary air is ducted from an air handler. Induction air is either from the ceiling plenum or ducted from the space. The fan and damper are aligned so that all primary air and all induced air independently enter the mixing section and go through the fan to be delivered to the space. The mixing space is between the VAV air valve and the fan.

Parallel Fan-Powered ATUs. These ATUs have an intermittently operating fan and are typically installed in the ceiling plenum. Primary air is ducted from an air handler and flows directly to the zone in a variable-volume cooling mode. Fan air is induced from the ceiling plenum or may be ducted from the space. The fan should run in heating and dead-band mode. In dead band and heating mode, the supply air is a combination of primary and induced air. The fan and damper are arranged so that all induced air enters the fan, and primary air bypasses the fan. Any mixing of primary air with the induced plenum air occurs on the discharge side of the fan. A backdraft damper inhibits air from exiting the unit through the fan when it is not running.

Fan-powered ATUs applied with underfloor air distribution systems differ from conventional overhead systems in that the ATUs are mounted in a supply air plenum rather than a return air plenum. The ASHRAE UFAD Guide (2013) addresses these ATUs in detail.

DOAS ATUs. ATUs may be used in conjunction with dedicated outdoor air systems (DOAS) to provide ventilation air (and in many cases latent cooling) to individual zones. The ATUs could be single-duct VAV, dual-duct, or fan-powered. Single ATUs can work in parallel with sensible cooling devices (e.g., passive chilled beams, radiant cooling ceilings) to handle latent loads and meet ventilation requirements. Dual-duct ATUs can blend ventilation air with conditioned building air to meet ventilation requirements. Series-fan-powered ATUs designed with chilled-water sensible cooling coils on the induction inlets are specifically designed to work with DOAS applications. The chilled-water coils are intended to perform sensible cooling only, thus the entering chilled-water temperature should be warm enough to prevent condensate formation. The entire zone latent load is satisfied by an outdoor air handler providing preconditioned air to the fan-powered chilled-water ATU.

Single-Duct ATUs. Benefits: Typically produce very low sound levels and can be used for both VAV and CV applications. With the addition of an inlet-mounted thermistor, these units can also provide VAV heating and cooling from a central air handler.

Limitations: The central air handler must provide adequate inlet static pressure to supply air to the box, overcome any pressure drop of the unit and any integral coils, and move air to the zone. Any heating coils integral to the unit must be treated as reheat. Only remote heating devices located in the occupied space such as finned-tube elements or radiant panels that could be operated by unit controls would be treated as heat.

Dual-Duct ATUs. Benefits: Allow centralized heating and cooling in the facility. They can provide very high comfort at relatively low sound levels.

Limitations: Dual-duct systems generally have higher initial costs and operating costs.

Series Fan-Powered ATUs. Benefits: Make use of free heat in the form of return air for reduced heating requirement. Typically provide constant air motion and low noise levels in the zone for improved air circulation and occupant comfort. The constant fan operation recirculates return air from the plenum, returning unvitiated air and reducing outdoor air requirements. Low inlet static pressure requirements can reduce fan energy at the air handler. Energy savings can be maximized with electronically commutated motor (ECM). Fans with ECMs can modulate the supply air based on zone sensible load, reducing the amount of plenum air sent to the occupied space during cooling and nearly eliminating motor heat in the supply air stream (John et al. 2018). Unit fan allows night setback heating capability on a zone-by-zone basis without need to operate the air handler during unoccupied mode.

Limitations: Some states still allow permanent split capacitor (PSC) motors to be used. Such applications send excess plenum heat to the space in cooling. PSC motors are not efficiently modulated and are difficult to control when attempting to track part-load conditions.

Parallel Fan-Powered ATUs. Benefits. Make use of free heat in the form of return air for reduced heating requirement. Fan only runs during heating and dead band modes, and fan is sized to only handle heating airflow. Unit fan allows night setback heating capability on a zone-by-zone basis without need to operate the air handler during unoccupied mode.

Limitations: VAV turndown must be considered when selecting diffusers. Sudden changes in airflow cause noticeable changes in sound levels in the space, which is generally annoying to occupants. Casing and backdraft damper air leakage during cooling severely affects overall efficiency of these units (Sardoueinasab et al. 2018) and measurably limits the amount of heat to be reclaimed in the ceiling plenum.

Fan-Powered Chilled-Water ATUs with Sensible Cooling Coils. Benefits: Make use of free heat in the form of return air for reduced heating requirement. Typically provides constant air motion and sound level in the zone for improved air circulation and occupant comfort. Low inlet static pressure requirements can reduce fan energy at the air-handling unit. DOAS allows for reduced ventilation air volumes. Energy savings can be maximized with ECMs. Unit fan allows night setback heating capability on a zone-by-zone basis without need to operate the air handler during unoccupied mode.

Limitations: Units must be selected and operated to keep local cooling 100% sensible. Units should not be allowed to condense water on the coils.

Sizing ATUs. For primary air inlets, select an inlet size that meets the minimum and maximum airflows desired from the recommended primary airflow range table in the manufacturer’s catalog. Selecting terminals near the top of their range increases velocity, sound, and air handler fan energy. Oversizing single-duct terminals reduces control accuracy and stability, and may result in higher sound levels. Transducer fidelity at the low end varies by manufacturer and must be considered in combination with the airflow sensor signal at the minimum airflow rate when selecting an inlet size. To maximize performance, size the terminal’s maximum airflow limit for 70 to 85% of its rated capacity (approximately 10 m/s) in accordance with the catalog recommendations. For accurate control, the minimum setting guideline should not be lower than 2 m/s inlet neck velocity for units using inlet velocity sensors. Other minimum guidelines may apply for units with specialty controls.

Select terminals based on recommended air volume ranges. A pressure-independent terminal’s main feature is its ability to accept factory-recommended minimum and maximum airflow limits that correspond to the designer’s space load and ventilation requirements for a given zone. A common misconception is that oversizing a terminal makes the unit’s operation quieter. In reality, the oversized terminal damper must operate in a near-closed condition most of the time, which may actually increase noise levels to the space. Control accuracy may suffer because the terminal is only using a fraction of its total damper travel or stroke. In addition, the low inlet velocities may be insufficient to produce a reliable signal for the velocity pressure measuring device and reset controller. This means minimum settings may not hold, with a resultant loss of control accuracy and undesirable hunting.

Oversizing the discharge duct may create low static conditions, requiring the fan to operate outside its recommended operating range.

Oversizing terminals with electric heat can lead to insufficient total pressure, which can occasionally trip the airflow safety switch.

Environmental Factors. Environmental factors play an important role in system selection. They include the climate and air conditions both indoors and outdoors. They also include legislative requirements such as outdoor air ventilation rates and local building codes. When high ventilation rates are required, such as in critical hospital spaces and clean rooms, reheat is often required to maintain human comfort. Fan-powered ATUs are usually used in noncritical spaces where the thermal load changes significantly and heating is required. Single-duct ATUs are frequently applied in the interior, where the thermal load is normally stable.

Building Use. Before specifying equipment types, the designer must consider the building’s intended use. Office buildings with daily operational schedules frequently use fan-powered ATUs. Usually, fan-powered ATUs with auxiliary heaters (supplementary heat) are used in perimeter zones; these units allow the greatest flexibility for individual zones while also allowing the central system to be turned off during unoccupied periods. During unoccupied periods, the fan-powered ATUs maintain the minimum or setback temperature levels without the help of the central air-conditioning equipment.

In institutional, medical, or campus buildings, systems that provide pressurization differences between interior areas may be required.

Buildings with centralized heating and cooling plants sometimes use dual-duct ATUs.

Building Size. In large buildings, central air handlers deliver large quantities of air to many zones with different needs. Interior zones may not require heat; therefore, they may be served by single-duct or fan-powered ATUs with no heat. Unless the building is located in a mild heating climate, the perimeter zones require heat, typically electric or hot water. These are usually included with the ATUs, but sometimes separate heating systems are used (such as baseboard heat). The static pressure in the ducts should be lowered to the minimum pressure in accordance with ASHRAE Standard 90.1, which allows at least one VAV damper to remain near full open. Interior zones in these buildings can use fan-powered ATUs to keep the static pressure low. Buildings with parallel-type fan-powered ATUs usually use single-duct ATUs in the interior zones and require higher system static pressures.

In small buildings, such as shopping malls and other low-rise buildings where each tenant area is small, it is common to use small packaged air conditioners. If ATUs are used on these systems, single-duct or bypass units are usually selected. A variation of this system, variable-volume variable-temperature (VVT), uses pressure-dependent single-duct units with a main bypass air valve in the supply duct. The bypass damper is regulated by static pressure in the supply duct. A near constant pressure can be maintained, allowing the packaged units to operate at constant volume and the individual zones to be pressure-dependent VAV.

Building Controls. The type of control system depends somewhat on the size and type of building.

Cost Factors. Consider costs before finalizing system selection. Installation, operation, and maintenance all contribute to total cost. Often, one of these costs overrides the others. Electric heaters usually have a lower installed cost than hot-water coils, but they may have a higher operating cost. Research local utility rates and building codes to arrive at the correct decision before making the final selection.

Acoustical Considerations. ATUs and room air distribution devices are typical sound sources. However, they are not the only equipment affecting room acoustics. See Chapter 49 of this volume, Chapter 4 of ASHRAE (2017b), AHRI Standard 885, and other standards for guidance on space acoustics. Broadcast studios, theaters, libraries, and other acoustically sensitive applications require careful consideration, because equipment selection and location are important. Radio-frequency interference (RFI) and electromagnetic interference (EMI) should also be considered when designing broadcast studios.

Sound levels are affected by primary air valve and/or fan-generated sound. The maximum sound generated by a given primary air valve size is determined by the difference between the highest inlet static pressure and external static pressure at the design cooling airflow for a non-fan-powered ATU and parallel fan-powered ATUs. For series fan-powered ATUs, the highest air valve sound level usually occurs at the highest inlet static pressure; however, the fan usually sets the room noise levels for series ATUs. To determine fan noise levels, fan airflow (adjusted within its range by the speed controller) and external static pressure conditions are required.

Acoustical performance data are presented in formats for both the parallel and series ATUs, because their sequence of operation differs. With a parallel ATU, air valve and fan operation are evaluated separately. With a series ATU, air valve and fan are evaluated together. Series fan-powered ATU sound levels are more consistent compared to ambient background sound levels than those of the parallel fan-powered ATU, which has a cycling fan.

From the performance data, determine the sound power levels and predicted room noise criteria for discharge and radiated paths under the appropriate operating conditions. Use care, because some published room noise criteria are based on certain path attenuation assumptions that may not correspond to specific applications. To minimize these differences, AHRI Standard 885, Appendix E, provides specific values to apply to sound power levels for NC calculations. These may not match any specific space, but yield comparable NC values between different manufacturers. For a complete description of these processes, see AHRI Standard 885 and Chapter 4 in ASHRAE (2017b). Both of these sources provide guides to apply specific attenuation values for a specific space to predict actual room noise criteria. However, the attenuation values depend on the room furnishing and finish, as well as the sound power levels generated by the ATU.

It is necessary for the architect and the engineer to recognize all factors in the building specifications that affect sound attenuation. This is necessary to properly evaluate the room noise criteria and ensure the finished levels do not exceed the design goal in the occupied space. An ideal specification identifies maximum allowed discharge and radiated sound power by octave band, rather than just catalog based NC values.

A sound-sensitive occupied space (e.g., conference room, private office, music studio, concert hall, classroom) may require more discharge sound attenuation than less sound-sensitive spaces. Usually some internal lining in the discharge duct can accomplish this. Occasionally, a discharge attenuator or silencer may be attached to the ATU. Care must be taken to not increase the discharge pressure enough to increase fan speed, generating additional noise that may be greater than the attenuated value.

Space Restrictions. During design, try to ensure that terminals are located for ease of installation, optimum performance, and maintenance accessibility.

Optimizing Inlet Conditions.

Zoning Requirements. Correctly sizing terminals with regard to the physical conditions of the occupied space is vital to ensure acceptable performance. One large terminal serving a space with divided work areas may result in the single thermostat only providing acceptable temperature control for the area where the thermostat is located. The other area(s) served may be too cold or too hot if they have differing space load requirements.

Optimizing Discharge Conditions. Poor discharge duct connections may have an adverse effect on pressure drop. Try to avoid installing tees, transitions, and elbows close to the unit discharge. Avoid long runs of flex, and keep short flex runs as straight as possible. Make curves as shallow as possible, and ensure that the entrance condition to diffuser outlet is straight. Discharge ducts should be designed for a maximum velocity of 5 m/s. Flex duct runouts at the diffuser inlets are generally a good attenuator.

Compliance with Applicable Building Codes. Some local jurisdictions have more exacting codes than the minimum requirements of national codes and standards such as the International Code Council’s (ICC) International Building Code^® (IBC). One example is the primary fusing required of the power circuit in some areas.

Power Source Compatibility. Terminals with an electrical power supply (e.g., fan-powered terminals, single-duct terminals with electric heat) should be checked for compatibility with source. Voltage, phase, and frequency must match.

Fan Interlocks. Typically, series fan-powered ATUs are designed to run continuously. Usually, they are energized only during occupied periods or when needed for emergency heating during unoccupied periods. Use care to interlock the unit fan with building’s air handlers to ensure that the ATU fans start during occupied periods. Series-unit fans should be started prior to the air handler to prevent backflow into the plenum and backward rotation of the fan.

Fan Shift in Fan-Powered ATUs. Before adjusting the fan, the possibility of fan shift must be considered. This occurs when the blower is subjected to variations in pressure or airflow patterns. As the primary airflow changes, pressure drop and changes in local jets may cause the fan to shift its performance as it rides the fan curve. Consequences vary from building to building and zone to zone. Noise levels may change greatly as the volume changes, and this may be annoying. Design ventilation rates can also vary, sometimes by more than 20%. This can be aggravated by undersizing the ATU.

Avoiding Excessive Air Temperature Rise. ATUs with electric or hot-water reheat coils should be designed to satisfy load conditions, but attention should be paid to the temperature differential Δt between the supply air temperature and room air temperature. Hart and Int-Hout (1980) and Lorch and Straub (1983) recommend a maximum Δt of 8.5 K to avoid possible stratification when heating from overhead caused by the excessive buoyancy of the warm air. This ensures good room mixing and temperature equalization. Exceeding a Δt of 8.5 K requires an increase of 25% in the ventilation air per ASHRAE Standard 62.1. Absolute maximum discharge air temperature is 49°C. Although this temperature will probably keep the equipment on line, it will not provide comfortable temperatures in the space.

Correctly Supporting Terminals. Although the basic single-duct terminal is light enough that it usually can be supported by the ductwork in which it is installed, these units should be independently supported. When accessory modules such as heating coils, attenuators, or multiple-outlet plenums are included, the assembly must be supported independently. Larger terminals such as fan-powered ATUs should always be independently supported, secured to building structure, and may require isolation mounting. Be careful not to block access panels with straps, thread rods, or trapeze supports. Be sure to comply with all building and local codes regarding seismic restraints (see Chapter 56).

Minimizing Duct Leakage. To prevent excess air leakage and minimize energy waste, all joints should be sealed with a UL-approved duct sealer. Most leakage can be avoided by practicing good fabrication and installation techniques, particularly upstream of the terminal, which may be required to hold significantly higher pressures than downstream of the terminal.

Acoustic Design and Installation. To help ensure an acceptable room noise level in the occupied space, engineers can minimize the sound contribution of air terminals by taking into account the following precautions:

Typical Applications. Terminal units are typically not easily accessible after building occupation; they should be selected and located with consideration for required maintenance. Review the applicable building codes (e.g., ICC [2009]) for required access. Fan-powered ATUs do not require filters, per section 307.2 of International Code Council’s (ICC) International Mechanical Code^® (ICC 2018) and ASHRAE Standard 62.1. Many fan-powered terminals are manufactured with construction filters that must be removed after construction is complete and are not suggested to be replaced.

Critical Environments. Some applications, such as cleanrooms and operating theaters, require high levels of reliability from ATUs because of the difficulty and cost associated with servicing or maintaining the equipment. In a cleanroom, for example, if the ceiling must be opened, the space may require disinfection before it can be used again. Associated costs might include lost production time as well as the cost for wipedown and/or disinfecting the room and/or equipment. In cases like these, consider locating the equipment outside of the clean space or using highly reliable, very-low-maintenance, very basic equipment and controls. ATUs may require access to internal components for cleaning in the case of contamination.

Fan-powered ATUs nearly always use single-phase motors, either ECMs, PSC motors with electronic fan speed control (sometimes called wave choppers, thyristor controllers, or silicon-controlled rectifiers [SCRs]), or PSC motors with three-speed switches combined with electronic fan speed control.

Electronically Commutated Motor (ECM). Currently, ASHRAE Standard 90.1 and ICC’s 2021 International Energy Conservation Code^® (IECC) require ECMs in fan-powered ATUs. ECMs provide superior controllability. They can be pressure-independent devices that calculate airflow, and can respond to a signal from a local controller or a remote signal from the building management system (BMS).

Electronic Fan Speed Control (PSC Motors). Electronic fan speed controls use a thyristor to adjust the fan’s electrical input ac voltage. This is called phase proportioning or wave chopping. Some units may suffer from large changes in amp draw that significantly affect the motor efficiency and operating characteristics. The PSC motor is a pressure-dependent device with a single setting.

ECMs provide significant energy savings over PSC motors, especially in part-load conditions (Edmondson et al. 2011). The part-load savings are primarily achieved by reducing airflow and taking advantage of the increased fan efficiency at below-design airflows. PSC motors have a limited amount of turndown available, but they also lose efficiency at or sometimes above the savings that can be achieved by the blowers. ECMs have a very small decrease in efficiency when they are operating at reduced speeds, and consequently they allow for nearly all of the improved blower efficiency to be captured in the operating system costs. PSC motor inefficiency manifests itself in the motor heat that is added to the supply air. Generally, PSC motors in fan-powered VAV ATUs can add 0.6 to 1.7 K to the airstream (Davis et al. 2007).

Nameplate Ratings. UL Standard 1995 covers fan-powered ATU nameplate ratings. This standard relates to equipment manufacturers and not field issues (which are covered in international and local codes). Nameplate ratings on the unit usually do not match the nameplate ratings on the motor. Amperage can be above or below the motor nameplate. Differences between the motor label and the unit label may be significant in some cases. Refer to the unit nameplate ratings and not the motor nameplate ratings when determining supply circuit requirements. These ratings are set at the safest possible condition. Static pressure and set points vary on each unit, so performance may not match what is on the unit nameplate.

Series. At typical design airflows, the ECM uses about 70% less energy than the PSC (Edmondson et al. 2011). At reduced airflows during part-load conditions, the overall energy savings is greater. Additionally, the reduced plenum air during cooling saves as much or more energy than the motor due to reduced cooling requirements.

Parallel. Default programming in digital controllers run the fan in the parallel ATU during dead band and heating. Because both conditions can suffer from overcooling in the occupied zone due to cold air supplying the required outdoor airflows, the PSC motor heat is used to augment the heat in the induced air to the zone. If ECMs are used, heat may need to be supplied from some other source to offset the overcooling potential during dead band and short cycling during heating. The heat source may be more efficient than the PSC motor heat generation, but overall energy savings is nearly negated by the motor heat loss.

Series. Most direct digital controllers (DDCs) provide an optional output that may be used for controlling fan airflow by the BMS if an ECM is used. This allows dynamic fan volume control, which may be either modulating or multiple-speed operation from a single-speed motor. The fan must be sized to match the maximum airflow to be supplied to the zone.

The sequence for series constant volume is as follows: the fan runs constantly at maximum design fan airflow during all occupied periods. On a call for cooling, the controls modulate the primary air valve toward maximum airflow, delivering primary air to the mixing chamber. If the fan is set at the same airflow as the primary air at maximum cooling, no air is induced from the plenum. If the fan is at a higher airflow than the primary air (e.g., as in a low-temperature application), air is constantly induced from the plenum.

As cooling demand decreases, the primary air valve modulates toward minimum airflow, reducing the flow of primary air into the mixing chamber. This increases the volume of warmer induced air into the mixing chamber. The increased percentage of induced plenum air causes the discharge temperature to rise to approach the plenum temperature, taking advantage of recaptured heat. Figure 28 demonstrates operation of a typical series constant-volume ATU (Reid et al. 2016).

On a call for heating, the controls automatically energize the supplemental heat (optional equipment), which can be either electric or hot-water coils. The discharge temperature increases as heat is applied. As the temperature increases in the zone, the sequence reverses.

For series variable volume, the sequence is the same, except that the fan modulates.

As cooling demand decreases, the primary air valve and the fan modulate toward minimum airflow with the primary air valve reducing airflow more quickly than the fan.

On a call for heating, the fan modulates toward maximum heating airflow set point. On a further increase in heating demand, the controls energize the supplemental heat (optional equipment), which can be either electric or hot-water coils. Figure 29 illustrates the operation of a typical series variable-volume ATU (Yin et al. 2018); see ASHRAE Guideline 36 for control sequence.

Parallel. In the heating and dead-band modes, the fan supplies a relatively constant volume of induced air to the space. The fan must be sized to supply the required heating airflow to the zone, which requires overcoming the pressure created in the mixing chamber caused by the inclusion of primary air. Figure 30 shows the operation of a typical parallel constant-volume ATU (Yin et al. 2016).

Parallel Variable-Volume Sequence. On a call for cooling, the controls modulate the air valve toward maximum airflow while the fan is off. Variable-volume, constant-temperature air is then discharged into the space. On a decreasing call for cooling, the sequence reverses.

In dead band, the controls energize the fan. Fan air and primary air are blended in the mixing chamber on the fan’s discharge side. The increased plenum air causes the discharge temperature to rise. Constant-volume, constant-temperature air is delivered to the space.

On a call for heating, the controls automatically energize the supplemental heat (optional equipment), which can be either electric or hot-water coils. Figure 31 shows the operation of a typical parallel variable-volume ATU (Yin and O’Neal 2018).

Fan-powered VAV ATUs take advantage of typical VAV savings at the air handler and chiller during cooling periods, and even more savings are realized when heating is required. Fan-powered terminals induce warm plenum air from the ceiling and blend it with the primary air at minimum ventilation requirements. This recaptures much of the heat created in the zone and plenum. If additional heating is required, supplemental heat is added to the sequence, thereby complying with ASHRAE Standard 90.1 during the heating sequence. The unit saves energy by warming blended air, for example, at 22°C rather than reheating primary cooled air at 13°C, saving the cost of 9 K at the heating airflow. According to ASHRAE research project RP-1292 (Davis et al. 2007; Furr et al. 2007), there is very little difference in total building energy use between series and parallel units when both are equipped with permanent split-capacitor (PSC) motors. However, each type of unit consumes that energy in different ways. ASHRAE (2017b) explains these differences in greater detail.

Series. Series units are designed for very low inlet static pressures. This saves energy at the air handler compared to a parallel unit for a similar zone. RP-1292 identified the PSC motor as the biggest energy user in a series unit (Davis et al. 2007; Furr et al. 2007). The fan energy raises the air temperature across the motor by 0.6 to 1.7 K. When modeling energy consumption for this unit, it is important to model the energy consumption and heat generated by the fan motor for both unoccupied and occupied periods (Davis et al. 2007; Furr et al. 2007). This means that total energy use can be reduced if the fan energy is reduced. Using ECMs can significantly increase motor lifetimes and provide significant energy savings. Warm air induced from the plenum at part load conditions increases the cooling requirement because of the increased air temperature in the mixing chamber of the ATU. The combination of motor and plenum heat can equal or exceed the energy consumed by the PSC motor.

Because the fan cabinet is neutral or negative during operation, the casing leakage does not affect energy consumption of series units. All the primary air injected into the unit is delivered to the occupied zone. This causes ceiling plenum temperatures to be warmer for installations with series units than those with parallel units, and more heat can be reclaimed from the ceiling plenum when the unit is in heating mode.

Parallel. More energy is required at the air handler in a parallel ATU system due to the higher inlet static pressure requirement. When just one duct path is designated as the critical path for static pressure set point, system pressure (and thus energy usage) can increase compared to a building with series units. However, ASHRAE Standard 90.1 procedures reset the system static pressure to always drive at least one VAV air valve to its nearly full-open position, making the increased static pressure required for parallel units lesser.

According to RP-1292, the largest single energy usage is leakage in the parallel unit even with dynamic reset at the air handler (Davis et al. 2007; Furr et al. 2007). Leakage is typically between 5 and 12% of the total primary airflow through the parallel unit, and it is highest at full cooling. This does not account for other casing leakage like seams and penetrations. Thus, parallel units may need to be oversized to cover the total load to the occupied zone. Due to the leakage, ceiling plenums are cooler with parallel units than with series units, reducing or eliminating the amount of free heat that can be reclaimed in heating mode. Per ASHRAE (2017b), casing leakage was calculated at $3.91 per L/s per year. When modeling energy consumption for this unit, it is important to model casing and backdraft damper leakage as well as fan energy consumption for both the ATU and the air handler during operation. Motor heat should also be included in heating mode calculation.

Since the motor runs in dead band and heating modes, any heat generated by the motor is delivered to the occupied space. Because of this, using more efficient motors (which run cooler) will not measurably improve the unit’s energy consumption.

Series. The inlet static pressure is the pressure required to push the conditioned air across the air valve into the mixing section.

Parallel. The inlet static pressure is the pressure required to push the conditioned air across the damper, through the mixing section, through the discharge duct, and across the diffuser(s) into the space.

Selection of fan-powered ATUs involves three elements: primary inlet, fan size, and heating coil. Selection of these elements and their interactive effects determine the overall performance of the units.

Primary Air Inlet Selection. See the section on Single-Duct Primary Air Inlet Selection. Note: casing leakage is an important issue for parallel fan-powered ATUs and must be taken into account when determining actual airflow supplied to and from the ATU.

Fan Size. Series fan-powered ATUs require the fan to be sized to handle the maximum design airflow. Fan airflow must be at least equal to primary airflow to ensure the mixing chamber in the terminal does not become positively pressurized, resulting in primary air short-circuiting into the ceiling plenum through the induction port(s). The external static pressure requirements are the sum of the ductwork and diffusers at design airflow plus an applicable hot-water coil or electric heater and optional filter, if required. When fan airflow and external static pressure have been determined, select the fan size from the fan curves provided by the manufacturer. Upsizing the fan and operating it at a reduced speed can result in quieter operation, but pay attention to fan stability at very low airflows.

Parallel fan airflow is determined by calculating the difference between the total discharge design heating airflow and minimum primary airflow. If minimum airflow is zero, then fan airflow is the heating airflow. In most applications of a parallel ATU, a minimum primary airflow is required to meet ventilation requirements. This primary airflow contributes to the total external static pressure experienced by the fan and should be accounted for along with all components, such as heaters, ductwork, diffusers, and optional filters.

Heating coils on ATUs are generally either electric or hot water coils. Sometimes steam may be used. Electric heaters are always encased for safety. Water coils may or may not be encased.

Heating coils on single-duct ATUs are located on the discharge where the duct is attached. They are almost always not encased and are generally sized to be the same size as the discharge duct. If the ATU is internally lined, the internal diameter of the ATU will be smaller than the coil face area. Multiple outlet plenums or silencers may be mounted downstream of the coil. Sometimes the coils are mounted at the discharge end of the silencers. Access doors may be located upstream and/or downstream of the coils. Heating coils on single-duct ATUs are always reheat coils and must be designed to heat the air from the leaving air temperature at the air handler to the desired entering air temperature at the occupied space. Heating coils are usually either electric or hot water finned tube type.

Heating coils on series fan-powered ATUs are mounted on the discharge where the duct is attached. If the ATU is properly controlled, these coils provide supplemental heat and not reheat. Water coils may or may not be encased. Discharge ducts may have access doors to allow access to the coil. Coils should be designed to heat the mixed air in the ATU mixing chamber. The mixed air is a blend of the minimum primary airflow and the induced plenum air at the point where the heater is energized.

Heating coils on parallel fan-powered ATUs may be mounted on the discharge of the unit or at the induction inlet upstream of the fan. Heating coils on the induction port must be sized to heat the fan airflow to a temperature high enough to mix with colder primary air, achieving the required entering air temperature for the occupied space. The resulting air temperature must not exceed that of the motor(s) rated operating temperature. Coils mounted on the ATU discharge must be sized to heat the blended air from the mixing chamber to the required leaving air temperature. On ATUs with coils on their discharge, discharge ducts may have access doors to allow access to the coil. Coils may or may not be encased.

To avoid compromising zone ventilation effectiveness (ASHRAE Standard 62.1), the discharge temperature should not be more than 8 K warmer than the room temperature it serves.

Use manufacturers’ catalog data or ASHRAE (2017b) for details on sizing coils.

When selecting a unit for a particular set of conditions, ensure that the air delivery is designed to meet the room’s sound criteria and the system’s static requirements. Specific sound data are provided by manufacturers for various airflow deliveries for each unit and should be the guiding factor in selecting unit sizes. Avoid selecting equipment near the maximum or minimum of the fan curves. Selecting fans at these points may limit flexibility for future changes. When designing air systems and using fan-powered ATUs, it is important to match the fan air and primary air capacities to the space requirements. Series units require precise adjustment of fan airflow in relation to the primary air to ensure proper discharge air temperatures and to protect against short circuiting of primary air into the plenum.

ATUs for Cold-Air Systems. These systems often involve ice storage and can provide supply air typically colder than 9°C, below that supplied by conventional systems. To reduce the chances of condensation on outer surfaces, keep the following in mind:

Several other design and operational considerations can also minimize the possibility of condensation and related issues:

ATU Critical Environment Considerations. Hospitals, cleanrooms, and laboratories pose special challenges. Protective isolation spaces such as operating rooms, bone marrow transplant patient rooms, immunocompromised patient areas, and cleanrooms require positively pressurized environments. Infectious isolation spaces such as tuberculosis patient rooms require negatively pressurized environments. See ASHRAE Standard 170 for details on room pressurization, Chapter 9 for specifics on health care requirements, and Chapter 19 for details on clean spaces. Hospital rooms and cleanrooms frequently require constant high ventilation rates, which tend to favor single- or dual-duct ATUs. Pressure-independent, variable-speed motor technology has led to the development of fan-powered pressurization units.

To minimize entrainment of fibers into the airstream, either do not use internal insulation or use nonfibrous liners in the ATUs and duct systems. Insulations can be isolated from the airstream by metal, foil, or polymer liners inside silencers and ATUs. All of these liners have different thermal, acoustic, and other physical properties and should be evaluated for each application.

See Chapter 12 of ASHRAE (2017b) for more information on health care facilities.

Fan-Coil Types. Vertical Stack: A vertical stack fan-coil uses a fan and a water coil to condition air within a space by regulating tempered water flow through the coil. Designed for free-blow or ducted, concealed, or painted cabinet applications. This fan-coil model usually uses a riser piping system to deliver the chilled and hot water to the coil. This type of unit is typically used in high-rise hotels, condominiums, dormitories, and residential buildings.

Vertical: A standard vertical fan-coil uses a fan and a water coil to condition air within a space by regulating tempered water flow through the coil. These units either are concealed in the wall or have a painted metal cabinet built around them. This type of fan-coil is commonly used in hallways, dormitories, small apartments, etc.

Horizontal (Blow-Through): A horizontal unit is normally mounted overhead and contains a fan and a water coil to condition air in a space by regulating tempered water flow through the coil. The fan blows air into the coil and then discharges into the space. These units may be built in or have decorative housings installed.

Horizontal (Draw-Through): A horizontal unit is normally mounted overhead and contains a fan and a water coil to condition air in a space by regulating tempered water flow through the coil. The fan draws air across the coil and then discharges into the space. These units may be built in or have decorative housings installed.

Water Piping Distribution Systems. For fan-coil units requiring chilled and/or hot water, the piping arrangement determines the performance quality, ease of operation, operating cost, and initial cost of the system.

Two-Pipe Changeover Without Electric Heat: In this system, either hot or cold water is supplied through the same piping. The fan-coil unit has a single coil. The most common system change-over scheme with the lowest initial cost is the two-pipe changeover scheduled by outdoor temperatures. The outdoor changeover temperature is set at some predetermined set point. If a thermostat is used to control water flow, it must reverse its action depending on whether hot or cold water is available.

The two-pipe system cannot simultaneously heat and cool, which is required for most projects during intermediate seasons when the morning hours may need heat and the afternoon hours likely require cooling. In intermediate seasons when the boiler is not on, two-pipe systems without supplemental electric heat cannot meet heating demands.

Two-Pipe Changeover with Partial Electric Strip Heat: This arrangement can provide electric heating in some zones and cooling in other zones in intermediate seasons by using a small electric strip heater in the fan-coil unit. The unit can handle heating requirements in mild weather while continuing to circulate chilled water to handle any cooling requirements. When the outdoor temperature drops sufficiently to require heating beyond the electric strip heater capacity, the water system must be changed over to hot water. The designer should consider the disadvantages of the two-pipe system carefully; many installations of this type increase operational cost, and can be unsatisfactory if climates temperatures vary widely from morning to afternoon.

Two-Pipe Nonchangeover with Full Electric Strip Heat: This system should be closely evaluated due to the escalated operational cost of electric heat compared to hot-water heat. It may be practical in areas with small heating requirements.

Three-Pipe Distribution: Three-pipe distribution uses separate hot- and cold-water supply pipes. A common return pipe carries both hot and cold water back to the central plant. The fan-coil unit control introduces hot or cold water to the common unit coil based on the need for heating or cooling. This type of distribution is not recommended because of its energy inefficiency and does not comply with most recognized energy codes.

Four-Pipe Distribution of Chilled and Hot Water: This system has dedicated supply and return pipes for chilled and hot water. It generally has a high initial cost compared to a two-pipe system. Four-pipe systems have better performance because of all-season availability of heating and cooling at each unit, no summer/winter changeover requirement, and simpler operation. It can be controlled to maintain a dead band between heating and cooling during shoulder season.

Variable-Refrigerant-Flow (VRF) Systems. The zone side of these systems is a fan-coil unit. Further information is available in Chapter 18 of the 2020 ASHRAE Handbook—HVAC Systems and Equipment.

Vertical Stack. Benefits: These units take advantage of the multi-story feature to plumb the water piping vertically floor to floor to conserve space.

Limitations: These units occupy floor space in the room. Because the units are traditionally located in the space, noise can be a problem if the fan and motor are not sized properly or if noise is not considered during design.

Vertical. Benefits: These units traditionally condition air directly in the space. Return air is typically drawn in the bottom and discharged out the top, directly back into the space.

Limitations: These units occupy floor space in the room. Because the units are traditionally located in the space, noise can be a problem if the fan and motor are not sized properly or if noise is not considered during design.

Horizontal (Blow-Through). Benefits: These units do not occupy floor space in the room, and can be manufactured with a lower profile.

Limitations: The velocity profile may not be evenly distributed across the cooling coil.

Horizontal (Draw-Through). Benefits: These units do not occupy floor space in the room, and generally have evenly distributed airflow across the cooling coil, resulting in greater heat transfer and lower noise.

Limitations: Any condensate blow-off caused by airflow over the coil has the potential to land on electrical components, including the fan motor.

Sizing Fan-Coil Units. How fan-coil units are selected and how they interact with the rest of the system determine the overall performance of the units.

General Sizing Guidelines: Select fan-coil based on recommended air volume ranges. A common misconception is that oversizing a fan-coil makes the unit operate more efficiently. On the contrary, an oversized fan-coil reduces run times, making the room air stratified and humidity more difficult to maintain at comfortable levels.

The recommended selection for maximizing performance is to size the fan-coil maximum airflow limit for 70 to 85% of its rated capacity in accordance with the catalog recommendations.

Do not exceed 2.5 m/s air velocity through the coil section of a chilled-water fan-coil. Exceeding this limit could cause condensate blow-off from the fin material, potentially causing damage to surrounding materials and reduces the latent cooling capacity of the unit.

Select water flow rates that will not exceed a velocity of 2.4 m/s in the coil tubing, because this velocity can erode the copper and cause pinholes, which could result in water leakage and potential damage to material and property.

Fan Size Selection: Fan-coil units require the fan to be sized to handle the maximum design airflow. The external static pressure requirements are the sum of the ductwork, diffusers, and filters at design airflow. When fan airflow and external static pressure have been determined, select the fan size from the fan curves or the selection software provided by the manufacturer. Vertical fan-coils with one return and one supply grille integral to the cabinet have no appreciable external static pressure requirements. This style of vertical fan-coil must deliver the design airflow for the application between 0 and 124 Pa. Oversizing the fan and operating it at a reduced speed can result in quieter operation, but pay attention to fan stability at very low airflows.

ASHRAE research project RP-1741, Understanding Fan Coil Components and How They Relate to Energy Consumption and Energy Modeling, analyzed energy modeling of fan-coils. RP-1741 benchmarked the energy savings from air and water modulation control relative to conventional on-off control of hydronic fan-coil units. Similar to other HVAC equipment, hydronic fan-coil units are generally oversized to be able to meet peak cooling and heating loads. Consequently, fan-coil units rarely operate at their full load for an extended time. Running fans and pumps at part-load conditions could allow energy savings. The modulation control of air and water flow based on thermal loads can take advantage of oversized equipment and operate them at part-load conditions for energy savings. Building energy simulation conducted in RP-1741 showed that modulation control could save over 50% of the fan-coil units’ fan energy compared with cycling on and off. Modulation control can also maintain an higher chilled-water ΔT at part-load conditions, leading to over 50% savings in pump energy and more efficient chiller staging. The total HVAC system savings ranges from 5 to 15% compared with the conventional on-off control.

Environmental Factors: Environmental factors include the climate and air conditions, indoors as well as outdoors. They also include legislative requirements such as outdoor air ventilation rates and local building codes. Fan-coil units receive ventilation air from a penetration in the outer wall or from a central air handler. Units that have outdoor air ducted to them from an aperture in the building envelope are not suitable for commercial buildings because wind pressure allows no control over the amount of outdoor air admitted. Ventilation rates can be affected by stack effect and by wind direction and speed. Also, freeze protection may be required in cold climates.

Fan-coils are, however, often used in residential construction because of their simple operation and low first cost, and because residential rooms are often ventilated by opening windows or by outer wall apertures, if not handled by a central system. Operable windows can cause imbalance in a ducted ventilation air system.

When outdoor air is introduced from a central ventilation system, it may be connected to the inlet plenum of the fan-coil or introduced directly into the space. If introduced directly, ensure that this air is pretreated, dehumidified, filtered, and held at the room’s temperature so as not to cause occupant discomfort when the fan-coil unit is off. One way to prevent air leakage is to provide a spring-loaded motorized damper that closes off ventilation air when the unit’s fan is off.

Building Size: In large buildings, where a central air handler(s) is not practical and where individual power costs are the responsibility of each space owner, a fan-coil is generally the system of choice for temperature and humidity control. These systems require a central water supply and a water transport system. If the initial cost and cost of operation of central chilled- and hot-water-generating equipment make more economical sense, the hydronic fan-coil comfort system is a good choice.

Cost Factors: Costs should be considered before the final system selection is made. Installation, operation, and maintenance all contribute to total cost. Sometimes one of these costs is more important than others. Electric heaters usually have a lower installed cost than hot-water coils, but may have a higher operating cost. Local rates and codes should be researched to arrive at the correct decision before making the final selection. See Chapters 13 and 46 of the 2020 ASHRAE Handbook—HVAC Systems and Equipment, Chapter 22 of the 2021 ASHRAE Handbook—Fundamentals, and Chapter 51 of this volume for more information on piping systems.

Acoustical Considerations. Since fan coils are often located within the occupied space, extra consideration should be given to avoid generating noise near occupants. For fan coils located outside the occupied space, the acoustical considerations are very similar to those of terminal units. Refer to the Acoustical Considerations portion of the section on Terminal Units for more information.

Installation and Operational Considerations. Space Restrictions. During the design phase, ensure the fan-coils are located for ease of installation, optimum performance, and maintenance accessibility.

Optimizing Inlet Conditions. Fan-coils should be located to optimize the inlet return air and discharge grilles or duct work. Proper location and size of grille and duct contribute to the overall effectiveness of air mixing and efficiency.

Locating a return grille in an area that does not properly serve the entire conditioned space can lead to poor performance of the coil and hot or cold spots within the occupied space.

Zoning Requirements. Correctly sizing fan-coils with regard to the physical conditions of the occupied space is vital to ensure acceptable performance. One large fan-coil serving a space with divided work areas may result in the single thermostat only providing acceptable temperature control where the thermostat is located. The other space(s) served may be too cold or too hot if they have differing space load requirements.

Optimizing Discharge Conditions. Poor discharge duct connections may have an adverse effect on pressure drop. Avoid installing tees, transitions, or elbows close to the unit discharge. Avoid long runs of flex duct, and keep short flex runs as straight as possible. Make curves as shallow as possible, and ensure entrance condition to diffuser outlet is straight. Discharge ducts should be designed for a maximum velocity of 5 m/s.

Noncompliance with Local Electric Codes. Some local jurisdictions have more exacting codes than the minimum requirements of national codes and standards such as IBC, the National Fire Protection Association’s National Electrical Code^® (NEC^®; NFPA Standard 70), UL, and Canadian Standards Association (CSA) standards. One example is the primary fusing required of the power circuit in some areas.

Power Source Compatibility. Fan-coils with electric heat should be checked for compatibility with source. Voltage, phase, and frequency must match. Where motor voltage differs, the single-phase voltage requirement may have to be tapped from a three-phase (4 wire wye) power source.

Fan Interlocks. When an electric heat source is used, a fan interlock is required to ensure the fan is on during the call for heat.

Avoiding Excessive Air Temperature Rise. Fan-coils with electric heaters or hot-water coils should be designed to satisfy load conditions, but attention should be paid to the temperature differential Δt between the supply air and room air. Exceeding Δt of 8.5 K requires an increase of 25% in the ventilation air per ASHRAE Standard 62.1. Absolute maximum discharge air temperature is 48.9°C per UL standards. Although this temperature will probably keep the equipment on line, it will probably not provide comfortable temperatures in the space.

Correctly Supporting Fan Coils. Fan-coil units should always be independently supported and secured to building structure, using the proper isolation mounting. Be careful not to block access panels with straps, thread rods, or trapeze supports. Where codes require seismic restraints, be sure to comply with all building and local codes.

Minimize Duct Leakage. To prevent excess air leakage and minimize energy waste, all joints should be sealed with a UL approved duct sealer. Most leakage can be avoided by practicing good fabrication and installation techniques

Acoustic Design and Installation. To help ensure an acceptable room noise level in the occupied space, engineers can minimize the sound contribution of a fan-coil by taking into account several design considerations and by using the following guidelines for good design practice:

Fan Airflow Control of Fan-Coil Units. Three-speed permanent split capacitor (PSC) motors are the standard motor option for most fan-coils. These motors have high, medium, and low speeds. These motors are likely be phased out because of minimum efficiency requirements in ASHRAE Standard 90.1.

The availability of electronically commutated motors (ECMs) for fan-coil units is rapidly growing, and most manufacturers offer a model. These motors often provide significant energy savings, improved controllability, and wider operating range in fan-coil units.

ECM versus PSC Motors in Fan-Coil Units. Three-Speed PSC: When a PSC motor is used in a fan-coil, the fan motor is switched on and off by a thermostat. This control scheme does not allow precise control of temperature and humidity levels in the conditioned space, and can cause comfort issues and increased energy cost.

Multispeed ECM: When a multispeed ECM is used in a fan-coil, the fan motor is switched on and off by a thermostat. The control scheme can be identical to that of a PSC but with the added benefit of a higher-efficiency motor, thus lower operating cost. Generally, an ECM is a 24 V control, and the motor can be energized directly from the thermostat or through a specially designed ECM fan control board, which could offer additional features.

Fully Modulating ECM: When used in the fan-coil with a more sophisticated control (e.g., analog control or DDC), this sequence of operation can provide a higher level of comfort and energy savings. Typically, an analog signal (0 to 10 V DC) from a controller is used to remotely vary the speed of the ECM. The dead-band differential between no call for heating/cooling and the energized state is generally much smaller in this type of control scheme. The water flow can also be modulated along with airflow modulation to avoid overcooling or overheating. This approach provides minimal swings in temperature and humidity levels in the conditioned space, which can provide superior comfort and reduced energy costs.

Basic Control Types. Electric Thermostat: Traditionally, the most common control type for fan-coils is a room thermostat. These thermostats come in both line voltage (i.e., same voltage as motor) and 24 V, which require a Class 2 step-down transformer (typically provided with the fan-coil).

When using a line voltage thermostat, the blower motor speed taps (three-speed) are connected directly to the thermostat. In this configuration, the water control valve is also line voltage. This type of thermostat is limited, and is generally used only on smaller motors because of ampacity limitations and thermostat ratings. In much of the United States, these thermostat leads must be run in a conduit between the fan-coil and thermostat.

When using a 24 V thermostat, fan relays, rated for the current of the motor, are used to switch each speed of the motor on and off. In this configuration, the water control valve is also typically 24 V. In most of the United States, these thermostat leads are not required to be run in a conduit between the fan-coil and thermostat.

Analog Stand-Alone: Another control type for fan-coils is an analog control combined with a room sensor. This control type is powered with 24 V thermostats, which require a Class 2 step-down transformer (typically provided with the fan-coil).

An analog controller is used when the fan-coil has modulating components such as an ECM and modulating water valves. The analog controller provides the appropriate outputs, based on room demand, to modulate the fan and water valves, if equipped.

This type of control can provide improved comfort over a simple electric thermostat. Depending on the features of the analog controller, the cost is generally higher than a room thermostat without set back features.

Direct Digital Control (DDC): This control type, paired with a room sensor, is powered with 24 V; these thermostats require a Class 2 stepdown transformer (typically provided with the fan-coil).

A DDC controller can be used for both types of fan-coil sequences described previously.

A DDC has both digital and analog inputs and outputs. The controller is programmable to provide many operation sequences, depending on the application. Generally, these controls can provide anything from a simple control sequence to a more sophisticated application. In some cases, DDC can control features and components outside the fan-coil, such as humidifiers, baseboard heat, bathroom ventilators, or lights, and can be programmed for occupied and unoccupied modes for energy savings.

Some DDCs can communicate through building automation systems (BASs) in several different protocols, such as BACnet™, Lonworks™, and Modbus^®.

This type of control can provide improved comfort over a simple electric thermostat or stand-alone analog controls. Depending on the features, cost for a DDC is typically higher than for a room thermostat or analog controls.

Basic Control Sequences. Three-Speed Motor with Two-Position Valves: Thermostats are available with many features beyond the purview of this explanation, but the simplest sequence is as follows.

On a call for cooling, the fan is energized at the speed (high, medium, or low) selected by the occupant or by the thermostat. The chilled-water control valve opens fully. On reaching the room set point, the fan de-energizes and the chilled-water control valve closes.

On a call for heating, the fan is energized at the speed (high, medium, or low) selected by the occupant or by the thermostat. The hot-water control valve opens fully. On reaching the room set-point, the fan de-energizes and the hot-water control valve closes. For a fan-coil with electric resistance heat, a heater control, like a contactor or heater sequencer, is energized on and off by the thermostat based on room demand for heating.

Consult the manufacturers’ catalog or installation and operating instructions for specific applications.

Modulating Motor with Modulating Valves: Controllers for fan-coils with modulating components are available with many additional features, but a basic sequence is as follows.

On a call for cooling, the chilled-water valve opens to a position consistent with the controls algorithm based on room demand, and the fan modulates toward the maximum cooling airflow set point. At a full call for cooling, the chilled-water valve is wide open and the fan is at the maximum cooling airflow set point. When approaching the room set point, the fan airflow reduces and the chilled-water control valve modulates toward the closed position. On reaching the room temperature setting, the valve closes completely and the fan modulates to the dead-band fan airflow set point.

Dead Band: Fan is off or running at a minimum airflow to circulate air through the space. On a call for heat, the hot-water valve opens to a position based on room demand, and the fan modulates toward the maximum heating airflow set point. At a full call for heating, the hot-water valve is wide open and the fan is at the maximum heating airflow set point. When approaching the room set point, the fan airflow reduces and the hot-water control valve modulates toward the closed position. Similar to the case of a fan-coil with electric resistance heat, a heater control, like a contactor or heater sequencer, is energized on and off by the room demand for heating. An SCR heater control can modulate the resistance heater by an analog output from the controller similar to the hot-water valve operation described above. Figure 32 shows the operation of a typical fan-coil unit with hydronic cooling and electric heating in modulation control.

Consult the manufacturers’ catalog or installation and operating instructions for specific needs.

The designer must consider the intended building use when determining the type of fan-coil equipment to be used. Fan-coil systems are best applied where individual space temperature control or cross-contamination prevention is needed. Suitable applications are shown in Table 10.

First, determine the air temperature rise by calculation using the heat transfer equation:

where

q = sensible coil capacity, kJ

Q = coil airflow rate, L/s

Δt = supply air temperature (SAT) – coil entering air temperature (EAT), °C

where

EAT = entering air temperature of coil, °C

T1 = return air temperature, °C

T2 = primary or ventilation air temperature, °C

Q1 = return airflow rate, L/s

Q2 = primary or ventilation airflow rate, L/s

QT = total airflow rate moved by fan, m/s

where

LAT = leaving air temperature of unit, °C

T1 = coil leaving air temperature, °C

T2 = primary or ventilation air temperature, °C

Q1 = coil airflow rate, L/s

Q2 = primary or ventilation airflow rate, L/s

QT = total airflow rate moved by fan, m/s

The supply air temperature (SAT) to the space equals the leaving air temperature (LAT) for the terminal unit or fan-coil. The terminal unit or fan-coil LAT and airflow rate should be known based on the load calculation. The coil EAT can be calculated using Equation (17) for all units mixing air upstream of the coil, or it should be equal to the return air temperature for all units mixing air downstream of the coil.

Once the coil EAT has been determined, the coil capacity can be calculated using Equation (16).

In applications for units mixing air downstream of the coil, the LAT of the unit can be calculated using Equation (18).

The required kilowatts and number of steps/stages desired should be checked with the manufacturer.

For hot-water coils, refer to the capacity charts in the manufacturer’s performance data or selection software to select the appropriate coil.

In heating applications, it is good practice to limit supply air temperature to 9.4°C above room temperature.

In applications in which mixing occurs downstream of the coil, heat generated by the water coil shortens the motor life and may cause nuisance tripping of the motor thermal overload.

At part-load conditions, it may be desirable to modulate airflow through the terminal unit or fan-coil as well as the heat output to maintain an acceptable discharge air temperature. This can be done with modulating valves on coils or proportional control on electric heaters. Staging the electric heaters can create similar results at a lower equipment cost. Modulating the heat causes the heaters to run longer, but at lower energy consumption. This can make the room more comfortable without increasing energy costs.

Most manufacturers provide coil performance data and/or selection software. Coil capacity depends on coil design characteristics and the properties of the entering air and water. The enthalpy increase or decrease of the water through the coil multiplied by the airflow is the resultant capacity. The water temperature and the flow rate determine the inlet enthalpy; the airflow and coil characteristics determine the outlet enthalpy. The sensible to latent ratio is important to maintain occupant comfort. In the cooling mode, the resultant capacity should be the appropriate combination of sensible capacity for dry-bulb temperature control and latent capacity for humidity control. When these two components of capacity are met, the space being conditioned can stay in control and provide comfort for the occupants.

Do not exceed 2.5 m/s through the coil section. Exceeding this limit could cause condensate blow-off, potentially causing damage to surrounding materials.

Select water flow rates that will not exceed a velocity of 2.4 m/s within the coil tubing; this velocity can erode the copper and cause pin holes, resulting in water leakage and potential damage to material and property.

An active chilled beam is an air diffusion device that introduces conditioned air to the space for ventilation and temperature control purposes. Primary air, conditioned at the air handling unit to meet the ventilation and latent requirements of the space, is delivered through a series of nozzles, creating induction of room air through a unit-mounted sensible heat transfer coil. This primary air often also contributes to sensible cooling of the space and drives the induction function through the coil. Depending on their nozzle size and configuration, active beams typically induce two to five parts of room air for every part of primary air they deliver to the space. When heating is required, warm water can be circulated through the coil.

Passive chilled beams rely on the natural buoyancy of air currents associated with convective heat sources to transport warm air to the upper portion of the space. On contact with the beam’s integral sensible heat transfer coil, this air is cooled and falls back into the space. Primary air must be delivered to the space via a separate system for the purposes of ventilation and dehumidification.

Chilled-beam systems must be designed to treat sensible and latent space heat loads, provide adequate space ventilation, conform to space acoustical requirements, and maintain occupant comfort in conformance with ASHRAE Standard 55 and other applicable codes.

In general, chilled beams offer the opportunity to capitalize on the benefits of decoupled ventilation systems with beam coils being responsible for most of the sensible load in the zone and the primary air satisfying ventilation and latent load.

Benefits. Heat extraction or addition by the coil often allows for significant reduction in primary airflow requirements over all-air systems. Energy to transport cooling and/or heating media can be significantly reduced because of water’s high specific heat and density. As a result, chilled-beam systems require less space for the mechanical services, because of smaller ductwork and air-handling unit sizes. This reduction in mechanical service space requirements may make it possible to reduce the floor-to-floor height of a multistory building.

Water-side economizer opportunities may be extended as a result of the higher beam chilled-water temperatures, and provide an improved selection of available system options (e.g., geothermal, dry coolers, closed-circuit fluid coolers). Chilled-beam systems may offer opportunities to enhance chiller efficiencies and provide broader evaporator ranges.

Chilled beams require minimal maintenance. Vacuuming the beam coils is occasionally required and is typically guided by the needs of the space. Often, it is expected that service intervals could extend to three to five years. The lack of moving parts in chilled beams produces an inherently highly reliable system.

Chilled beams may be considered beneficial for the following applications/spaces:

Limitations. Space humidity levels must be managed closely because of the limited dehumidification capability of the primary air. This is particularly important where operable windows are used or high infiltration levels are encountered in a humid climate.

Chilled beams may be inappropriate for the following applications/spaces:

The objective of chilled-beam system design is to minimize primary airflow rates. When chilled-beam systems are applied, the minimum primary airflow rate is typically the greater of that required for ventilation and for space dehumidification. In cases where these values are similar (differing by no more than about 25%), consider using a dedicated outdoor air system (DOAS). Laboratories and health care may require or benefit from the use of a DOAS, which uses chilled beams efficiently.

In cases where the sensible cooling requirements require significantly more than the minimum primary airflow rate, air-handling units that mix return and outdoor air volumes may be considered for use with chilled beams.

Chilled beams are generally intended to operate without condensation. Consequently, active chilled-beam supply water temperatures should be maintained at or above the room dew-point temperature to prevent condensation on the coil and its supply water piping. Passive chilled-beam water supply temperatures should be kept slightly (1.2 to 1.7 K) above the room dew-point temperature. In both cases, the chilled-water piping must be adequately insulated to prevent condensation on the pipe itself. Where adequate control of space humidity levels cannot be ensured, higher supply water temperatures and/or condensation controls should be considered. This is discussed in the following sections.

Terminal filtration and condensate pans are not required with a properly designed and operated primary air system with chilled-water temperatures maintained above the room dew point. Heating coils provide sensible heat only, and thus filtration and condensate capture devices are not necessary. Chilled-beam systems designed with noncondensing (dry) coils should be treated similarly. Beams with sensible-only cooling coils do not require filters per ASHRAE Standard 170 and section 307.2 of the International Code Council’s International Mechanical Code^® (ICC 2018).

Heating is limited to active chilled-beam systems; heating with overhead passive chilled beams is not effective.

The hot water serving the active beam’s coil must be chosen to limit the discharge air temperature to less than 8.5 K above the room design set point. Additionally, to ensure proper room air distribution, the discharge velocity should be selected in accordance with guidance presented previously.

Alternatively, resetting the primary air temperature with a duct-mounted heating coil allows the primary air serving the interior spaces to continue to provide cooling, while the perimeter duct adds heating capability through this reset. Assuming the active chilled beams are the primary heating system, the beams should be located parallel to the curtain wall, to ensure air movement across these surfaces in order to promote a comfortable environment.

Six-port zone control valves may be used to eliminate a significant amount of zone runout piping and allow more efficient heating across active chilled beam coils.

When ceiling-based active beams are the primary space heating source, their primary air supply must be maintained to project the warm air into the space.

Chilled-beam systems are designed to optimize delivery of cooling to the space, but the paramount consideration in sizing and locating beams in the room should focus on occupant thermal comfort. ASHRAE Standard 55 defines limits on local air temperatures and velocities that maintain acceptable levels of occupant thermal comfort.

Properly applied passive chilled beams have a limited effect on occupant thermal comfort; however, their complementary primary air supply system often does. Stratified or partially mixed air diffusion strategies are commonly used with passive beams because of their minimal influence on the natural buoyancy-driven air patterns associated with the chilled-beam operation. The secondary air circulation through the passive beam transports upper-level air back to the occupied zone, possibly altering the level of stratification in the space.

Active chilled beams directly supply a mixture of primary and secondary air to the space and should therefore be treated like the other air distribution devices used in fully mixed air distribution systems. Because the temperature of the chilled water supplying the coil must be at (or above) the space dew-point temperature, it is typically 13 to 16°C; thus, reconditioned air leaving the coil is typically several degrees warmer than the primary air with which it is subsequently mixed. This results in beam design discharge air temperatures that are above 16°C, thus warmer than those normally used by conventional all-air systems. Because of warmer discharge temperatures, larger active beam discharge air volumes are required.

Chilled-beam system primary airflow rates are much closer to the space ventilation rates than those of all-air systems, so primary control of the space temperature is normally accomplished by throttling the chilled-water flow. Simple on/off operation of two-position water valves provides adequate control of active chilled beams. Proportional valves are recommended for passive beams and active beams in applications where more precise space temperature control is required.

In applications where primary air is supplied at conventional temperatures (13 to 16°C) to spaces with significant sensible load variations, it may also be necessary to reset the primary airflow rate or temperature during low-load conditions. One approach is to vary the primary airflow rate in reaction to thermal demands and/or occupancy of the space.

Although many chilled-beam applications involve a constant-volume supply of primary air, chilled beams can also be served by varying primary airflow rates. For example, classrooms and conference rooms where occupancy levels may vary considerably can be fitted with demand control ventilation (DCV) provisions. In such cases, the primary air supply to the beams can be varied according to occupancy, while a space dew point override ensures that the primary air volume reduction does not compromise room humidity levels. Water flow through the coil remains controlled by the space thermostat, resulting in the ability to control space temperature levels independent of the primary airflow rate.

When chilled-beam systems are used to condition spaces with widely varying sensible cooling requirements (e.g. perimeter spaces), consider using a variable-air-volume (VAV) terminal to vary the primary airflow rate to beams in the zone. This allows the primary airflow rate to be throttled in response to the zone’s sensible cooling requirements. The minimum flow limit on the VAV terminal is set at the minimum primary airflow rate required to ensure proper space ventilation, adequate room air induction, and humidity control.

Thermal zoning of chilled-beam systems should be performed in a manner generally consistent with other HVAC systems. Each thermal zone consists of a space thermostat, a chilled- (and, where applicable, hot-) water control valve, and multiple chilled beams.