VAV System Design - WN Mechwnmech.com/class/VAV Clinic Rev2.pdf · WN Mechanical Systems Variable...

VAV System Design

HVAC Clinic

Table Of Contents

Introduction ............................................................................................................................................ 3 CV vs VAV Systems ............................................................................................................................... 4 VAV System Components ..................................................................................................................... 9 VAV Control .......................................................................................................................................... 17 Design Considerations ........................................................................................................................ 20 Ventilation of VAV Systems ................................................................................................................ 23 Single Zone VAV Systems ................................................................................................................... 33

HVAC Clinics Draft - Not For Distribution

WN Mechanical Systems Variable Air Volume Systems of 35

Introduction

There are two primary ways in which we can heat or cool a space. Those two methods are:

Constant Volume, Variable Air Temperature

Variable Volume, Constant Temperature

The most traditional and widely used method is the constant volume, variable air temperature system (CV). With a CV system, our goal is to condition the space by providing a constant volume of air to the space and modulate the supply air temperature (figure 1). For example, in a cooling application, the higher the load, a lower leaving air temperature would be required to meet the space load requirements.

Figure 1. Constant Volume (CV) System

Conversely, a variable volume, constant temperature system (VAV) is less widely implemented that a traditional CV system. However, a VAV system presents a number of advantages when compared to a traditional VAV systems. In a VAV system, the air volume is modulated, while the air temperature is held constant (figure 2). For example, in a cooling application, at higher loads, a higher volume of air would be delivered to the space. VAV systems present the designer with two primary advantages compared to CV systems; individual zone control and energy savings. Both of these advantages will be discussed in the next section.

Figure 2. Variable Air Volume (VAV) System



CV vs VAV Systems

The goal of any airside system is to remove load (cooling) or add load (heating) to the conditioned space. Removing sensible load from a space is a function of the airflow and the difference between the space dry bulb temperature and the supply air dry bulb temperature. The sensible load in a space follows the relationship (equation 1):

Where: Qs = Sensible Heat CFM = Airflow Tspace = Space Dry Bulb Temperature Tsupply = Supply Dry Bulb Temperature If a constant volume system is employed to condition the space, the airflow will remain constant while the delivery temperature will become the independent variable. Solving equation 1 for the delivery temperature, we find (equation 2):

Assuming an example where the peak space sensible load (Qs) is 12,000 Btu’s/hr, the air handler is delivering a constant 550 CFM, and the space temperature is 75oF; equation 2 reduces to:

Thus, at full load, the dry bulb delivery temperature should be 54.89oF. For the purposes of the example, we are assuming there is no latent removal from the cooling coil.

At part load, the airflow will remain constant and the delivery temperature will be variable. Assuming the space load is 50% of design (6,000 Btu’s/hr); equation 2 reduces to:



Thus, at 50% load, we are supplying warmer air at a constant airflow rate.

Conversely, if we were to compare a VAV system, airflow would become the independent variable. Solving equation 1 for airflow (equation 3):

If we were to use the same example where the peak space load (Qs) is 12,000 Btu’s/hr, and the air handler is delivering a constant 55oF discharge air temperature; equation 3 reduces to:

At full load, the air handler is delivering 553 CFM at a constant 55oF discharge air temperature. Similar to the previous example, this example assumes the coil is not removing any latent energy.

At part load, the discharge air temperature will remain at 55oF and the airflow will modulate. Assuming the space load is 50% of design (6,000 Btu’s/hr); equation 3 reduces to:

Thus, at 50% load, we are delivering half the design air volume. While it is readily apparent that there will be fan energy savings at part load with a VAV system (due to the reduced volume of air being delivered), those savings come at the added advantage of greater humidity control in wetter climates. Generally speaking, delivering air at lower temperatures results in more latent removal at the coil. This is beneficial in humid climates. Being that a VAV system



delivers warmer air at part loads, some or all of this latent removal will be lost, resulting in a loss of humidity control. In drier climates (like Northern Nevada), this can actually be a benefit. In climates like Reno, designers have difficulty maintaining relative humidity levels above 30% in the space. In these climates, less humidity removal at part load becomes a benefit when trying to maintain minimal levels of relative humidity for comfort and static electricity control.

So, why VAV? VAV systems provide designers two main advantages compared to CV systems:

Increased Occupant Comfort

Increased Energy Savings

Decreased design airflow, fan size and trunk duct size

Increased humidity control in Dry Climates CV systems deliver a constant volume of air at variable temperature. When a CV air handler serves several spaces, possibly with variable thermal load profiles, the space temperatures can fluctuate. The fluctuations in temperature can lead to occupant complaints. In a CV system, generally only one space receives the thermostat, which acts as the control input to the cooling coil (or compressor staging). The other spaces may suffer (figure 3).

Figure 3. Single Thermostat

There are two possible solutions to this problem in CV systems. This first solution is to employ temperature averaging among several spaces (figure 4). While this will improve occupant comfort in the averaged spaces, it is still not as good as direct temperature control of the individual space.

Figure 4. Temperature Averaging

The second method is to employ terminal reheat with direct space temperature control of each reheat box (figure 5). While this method solves the problem of not having direct temperature control of each zone, in most locations, the build energy code prohibits the use of terminal reheat unless the energy is produced by means of energy recovery (condenser energy recovery, solar, geothermal, etc).



Figure 5. Reheat

A relatively simple solution to the temperature control problems associated with CV systems is to employ a VAV terminal system. A VAV system delivers constant temperature air to each zone. A VAV terminal unit at each zone modulates the volume of air delivered to the zone based on the zone temperature (figure 6).

Figure 6. VAV System

Unlike a CV system, the zone thermostat controls the zone terminal unit. In a CV system, the zone thermostat (or thermostats is averaging is employed) controls the air handler discharge temperature. A VAV system enables the potential for each zone to have its own individual temperature control. This gives the designer the flexibility to choose a system which will provide much better occupant comfort.

Second, VAV systems provide for increased energy savings compared to similar CV systems. Recall from the fan clinic that the fan law which describes the relationship between speed and fan BHP is:

Recalling that RPM is proportinal to CFM, we see that if we cut the airflow by 50%, we save 87.5%

Thus, by reducing fan CFM and fan speed, we can save a signifigant amount of fan energy. Buildings spend the vast majority of time at part load. Air-Conditioning, Heating and Refrigeration Institute (AHRI) Standard 550/590-2003 estimates that a typical commerical building spend:



Table 1. AHRI Standard 550/590-2003 Part Load of Commercial Buildings

% Load % Hours

100% 1%

75% 42%

50% 45%

25% 12%

Being that airflow (CFM) is proportional to load in a VAV system, we can estimate a typical commercial building fan power energy savings at part load as:

Table 2. Part Load Fan Power Reduction

% Load % Hours Fan BHP Savings

100% 1% 0%

75% 42% 57.8%

50% 45% 87.5%

25% 12% 98.4%

Knowing the percentage of hours at each load point and its associated fan BHP savings, we can estimate the average yearly fan energy savings as:

Average Fan Savings -= .1*(0) + .42*(57.8) + .45*(87.5) + .12*(98.4) = 75.46%

Thus, we see that with a VAV system, it is theoretically possible to save up to 75% of our fan energy for a typical commercial building (as defined by AHRI Standard 550/590).

Finally, VAV systems often allow the designer to decrease the system airflow. The system design airflow for a CV system is determined by the sum of the peak loads. Because a CV air handler delivers a constant volume of airflow whenever the building is occupied, the system supply fan must be able to deliver the peak airflow to each zone. This must happen regardless of the time of the peak. For example, assume three zones experience the following load profile (table 1):

Table 3. Part Load Example

Zone 9 AM Load (Btu/hr) 1 PM Load (Btu/hr) 4 PM Load (Btu/hr)

Office 1 12,000 14,000 11,000

East Office 15,000 12,000 11,000

West Office 6,000 9,000 14,000

Sum 33,000 35,000 36,000

A CV air handler would have to be sized a the sum of the peak airflow (highlighted in red). The system airflow would be based on the sum of the peak load, or 43,000 btu/hr (sum of the red values). In contrast, A VAV air handler is sized based on peak load at any given time. In the example above, the peak load is 36,000 CFM at 4 PM. This reduced airflow will likely result in a smaller system supply fan and reduced size of the trunk ductwork.

The challenges associated with the design of VAV systems include:

Managing ventilation air

Diffuser mixing at minimum cooling and heating airflows

The first challenge, managing ventilation air, is critical to the design of any VAV system. ASHRAE (American Society of Heating, Refrigerating and Air-Conditioning Engineers) Standard 62.1, Ventilation for Acceptable Indoor Air Quality defines the minimum amount of ventilation or outside air required for a space (the Ventilation rate procedure) where rates are determined using per-person and per-unit-area rates prescribed in a table based on occupancy category. The



question becomes, what happens when the air handling system unloads? As the fan speed slows down, the amount of ventilation air drawn through the outside air damper decreases (assuming a fixed position OA damper). It is very likely that this reduced quantity of ventilation air at part load will not meet AHRAE Standard 62.1 requirements. The solution is to reset the OA damper based on the quantity of ventilation air being drawn into the air handling system. This very often requires the use of an airflow monitoring stations at the OA inlet of the air handler. Another solution is to use CO2 to estimate occupancy and reset the OA damper position. Both of these methods will be described in more detail in the application consideration section later in this clinic.

The second problem associated with VAV systems is the proper design and selection of the air distribution system with respect to assuring proper mixing of airflow in the space. Typically, designers will assign minimum damper positions at the air terminal unit to assure proper mixing at both minimum cooling and minimum heating airflows. Proper diffuser selection and placement is critical at these minimum airflows in order to assure proper airflow mixing in the space. If proper mixing at the minimum airflow is not established, both occupant comfort issues and improper ventilation air distribution will not be maintained.

VAV System Components

A typical VAV system consists (figure 7) of the air handler, supply ductwork, terminal unit, diffusers, grilles, return ductwork (or open return plenum).

Figure 7. VAV System Components

The VAV terminal modulates the airflow to the space. This is accomplished with a sheet metal assembly with some type of modulation device, such as a rotating blade damper and actuator, that regulates the airflow in repsonse to a change in load (figure 8). The modulation devices increase or decreases the resistance to airflow in the duct system, changing the air quantity being delivered to the space. The terminal unit may include other components such as a heating coil, filter or fan.

Figure 8. VAV Box Components



The most common and simplest form of terminal device is a single duct, cooling only terminal device (figure 9). It comprises of a sheet metal enclosure which houses an air modulation device, actuator and controller. A single duct, cooling only terminal device can only modulate the primary (air from air handler) airflow to the space. A single duct, cooling only terminal device typically only serves cooling only spaces such as interior zones. If heating is required in the zone, some other method of introducing heat must be incorporated into the design. A good example of this would be to use a single duct, cooling only terminal device to modulate the cooling airflow and hot water baseboard at the perimeter of the space to handle the heating load.

Figure 9. Single Duct VAV Box

The operation of a VAV terminal unit can be described by plotting the airflow to the space as a function of space load (figure 10). The independent variable or space load is plotted on the horizontal axis. The dependent variable, or percentage of airflow to the space is plotted on the vertical axis. At the maximum space load, the terminal unit is delivering the maximum amount of air to the space. As the load decreases, the airflow decreases linearly until it reaches a predetermined flow setpoint. The minimum predetermined flow setpoint is generally a function of the air distribution system (the effectiveness of the diffusers) and the ventilation requirements for the space.

Figure 10. Cooling Only VAV Box

Below the minimum flow setpoint of the VAV terminal unit, some means of conditioning the air or space must be provided. Without some means of heating the primary air (or adding heat in the space), any space load below the minimum flow setpoint would be overcooled. For many cooling only applications (interior spaces for example), heating the primary air may not be required. However, if a cooling only terminal unit is employed in a space that requires heating or tempering of the cold air during cooling, then some means of space heating must be utilized.



If a space requires either tempering of primary air or heating (exterior zones for example), then a VAV reheat terminal unit may be employed (figure 11). A VAV reheat terminal unit has all of the same components as a single duct, cooling only terminal unit with the addition of a heating or electric resistance coil.

The operation of a VAV reheat terminal unit is very similar to that of a cooling only terminal unit. As the space load decreases, the primary airflow is decreased until a minimum pre-determined cooling airflow. The pre-determined cooling airflow is generally a function of the effectiveness of the air distribution system and the ventilation requirements. As the load decreases below the minimum cooling airflow setpoint, the heating coil is enabled. Generally, as soon as the heating coil is enabled, the minimum box setpoint is increased. Like the cooling minimum flow setpoint, the minimum heating airflow setpoint is a function of the air distribution effectiveness. Because heating air is more buoyant the cooling air (hot air is less dense than cool air), the flow setpoint is greater to make up for the reduced effectiveness of the diffusers when handling warmer air. As the heating load increases, the heating or resistance coil modulates in order to adequately condition the load.

Figure 11. VAV Reheat

The reheat energy delivered to VAV reheat boxes can often be substantial in cooler climates. A solution to this quandary is to employ fan powered VAV boxes, especially to those zones with higher heating loads. Fan powered VAV boxes utilize a fan to introduce warm plenum air into the VAV terminal unit. Fan powered VAV boxes come in two configurations; parallel and series configuration.

A parallel fan powered terminal unit utilizes a fan which introduces a constant volume of plenum air into the box (figure 12).

Figure 12. Fan Powered Parallel VAV Box



The warm plenum air is mixed with cool primary air and is delivered to the space. With a parallel fan powered terminal box, the fan is enabled upon the first call for heat, resulting in an increase in airflow to the space. Thus, fan powered boxes not only provide a means of utilizing free energy for heating (plenum heat), but help provide better zone mixing at heating airflows by increasing the airflow to the zone during a call for heat.

Figure 13 demonstrates the operation of a parallel fan powered terminal unit. As the space load decreases, the primary airflow is decreased until a minimum pre-determined cooling airflow. When terminal unit reaches the minimum cooling airflow, the parallel fan is enabled. The total airflow to the space is the sum of the primary air plus the plenum air introduced by the fan. The heat from the plenum is considered free energy. It is generated from the lights, roof load, etc. If the heat introduced from the plenum is not adequate to satisfy the zone heating load, some other means of satisfying the load must be introduced. This could be a heating coil at the terminal unit or some means of heating in the space (radiant heaters, wall fin, etc).

Figure 13. Fan Powered VAV Box

A series fan powered terminal unit, similar to a parallel fan powered terminal unit, employs a fan to introduce plenum heat into the terminal unit. However, the fan in a series fan powered terminal unit is positioned at the outlet of the terminal box. Thus, the airflow being supplied, both primary air and plenum air, is being supplied in series with the fan. This results in a constant delivery of air flow to the space at all load conditions.

Figure 14. Series Fan Powered VAV Box



Because the airflow being delivered to the space is constant at all load conditions, the plenum air being drawn into the terminal unit is the difference between the total airflow (at 100%) and the primary airflow. As the primary air is decreased, the fan supplements the loss in airflow by drawing in an equal amount of warm plenum air. As the terminal unit reaches the minimum primary airflow, any additional tempering or heating demand must be met through the addition of a terminal reheat coil or some means of heat in the space.

Figure 15. Fan Powered Series Terminal Unit

Fan powered series VAV systems provide several advantages compared to parallel fan powered systems. Those advantages are:

Constant airflow to space Improved acoustics

A series fan powered VAV terminal unit provides a constant volume of air to the space. This dramatically simplifies diffuser selection and generally leads to improved air mixing in the space. VAV diffuser selection can be difficult and often leads to stratification or short circuiting at lower air flows. In addition, the fan in a series fan powered terminal unit runs continuously when the space is occupied. While the fan generates noise, this increase in noise is often more tolerable than the cyclic nature of a parallel fan powered terminal unit. Occupants will generally prefer to listen to a constant noise source as opposed to a noise source than cycles.

The downsides to a series fan powered terminal unit compared to a similar parallel fan powered unit are:

Fan Size Cost

Generally speaking, the fan in a series fan powered terminal unit is larger than a comparable parallel fan powered terminal unit. However, systems utilizing series fan powered terminal units can often utilize smaller air handler fans. The air handler in a system utilizing series fan powered terminal units only has to overcome the friction to the terminal unit itself. The terminal unit fan overcomes the friction between the terminal unit fan and the diffuser. While the fans employed in terminal units are generally less efficient than a comparable air handler fan, utilizing a smaller fan in the air handler in series systems will offset some of the fan brake horse power differences between the two system types.

In addition, series fan powered series boxes are generally more expensive than a comparable fan powered parallel terminal unit. Much of this increase in cost is due to the larger size fan.

Thus far, we have discussed the use of VAV terminal units with respect to the design of single duct air handling systems. Another type of air handling system utilizes two ducts to distribute the air to the terminal unit; one duct dedicated to transporting cold air and another duct is dedicated to transporting hot air. These type of systems are called dual duct systems (figure 16).



Figure 16. Dual Duct VAV Box

A dual duct terminal unit consists of two air modulation devices and controllers housed within a sheet metal enclosure (figure 17). One modulation device controls the air within the hot deck while the other modulation device controls the air within the hot deck. These two modulation devices my act independently of one another or may be linked depending on the type of air handler utilized in the system. Often, some type of mixing baffle is employed to ensure proper mixing of the two airstreams.

Figure 17. Dual Duct VAV Box

A dual duct box may be mixing or non-mixing. A non-mixing dual duct box has cold duct and hot duct paths both upstream and downstream of the terminal unit. A mixing dual duct box had independent cold duct and hot duct paths upstream and a single duct system which transports the blended air downstream of the terminal unit (figure 18).

Figure 18. Non-Mixing & Mixing Terminal Units



If the system utilizes a two fan air handling system, one fan dedicated to the hot deck and another dedicated to the cold deck (figure 19), then the dual duct terminal unit dampers may act independently of one another. Each fan can run independently, responding to changes in duct static as the hot deck and cold deck terminal unit dampers modulate. In addition, each fan must have some means of modulation control (VFD’s, ECM motor, inlet guide vanes, etc).

Figure 19. Two Fan Dual Duct VAV System

A system employing independent control of the air modulation devices behaves in a manner similar to that of figure 20. Ventilation air is drawn in through the cold deck fan. The modulation device serving the cold duct within the terminal unit is allowed to modulate down to its minimum primary airflow. This minimum primary airflow is a function of the amount of ventilation air required for the zone and the effectiveness of the air distribution system. After the cold deck damper reaches this minimum predetermined level, the hot deck damper begins to open. The cold deck damper remains at its minimum position while the hot deck damper modulates in response to the tempering or zone heating load.

Figure 20. Two fan Dual duct VAV System

Conversely, a system which employs a single fan serving both the cold and hot deck (figure 21) will generally utilize a terminal unit with linked (or slaved) cold and hot deck modulation devices. In single fan dual duct systems, the fan delivers a constant volume of air at all load conditions.



Figure 21. Single Fan Dual Duct System

As the cooling load decreases, the air modulation device within the terminal unit decreases the cold deck airflow. However, in this instance, the cold deck damper is linked to the hot deck damper. The two dampers are linked such that they control in opposite directions. As the cold deck damper closes, the hot deck damper opens an equal amount. Ideally, these two dampers modulate to produce a constant flow of air downstream of the terminal unit (figure 22).

Figure 22. Single Fan Dual Duct System

Single fan double duct systems are generally less expensive, easier to control and simplify the air distribution and diffuser design (due to the constant volume nature of the system downstream of the terminal unit). However, single fan dual duct systems do not experience the fan energy savings realized with dual fan double duct systems. In addition, slaved dual duct VAV boxes generally do not provide a purely constant volume of air. Most VAV boxes do not provide linear changes in airflow with damper position. Thus, as the cold deck opens an incremental amount, the same change in the hot deck damper position will likely not provide the same incremental change in airflow. More on this discussion later in the clinic.

Dual duct systems, due to their increased installation cost, have rarely been employed in recent years. However, many older buildings utilized dual duct systems, making their understanding by designers equally as relevant today.



VAV Control

The modulation device within a terminal device either directly or indirectly responds to a change in space load or temperature. Two types control schemes are used for VAV systems. With the first, pressure dependent VAV control, the controller varies the modulation device in direct response to a change in temperature in the space (figure 23).

Figure 23. Pressure Dependent VAV Box

However, there are several problems associated with controlling the modulation device directly as a function of space temperature. The first problem is that most modulation devices do not linearly modulate airflow as a function of position. Most VAV terminal units today employ the use of a round damper mounted to a shaft to modulate airflow. Changes in load in the space respond in a linear fashion to airflow. Recall that:

Assumming the supply air temperature is constant and the space temperature is constant, space load changes in direct repsonse to airflow. However, when round dampers are modulated, position is not linear with respect to damper positon (figure 24).

Figure 24. Airflow versus Damper Position

The second problem associated with direct control of damper position with respect to space temperature variations is that the upstream pressure at a VAV terminal unit is rarely constant. Airflow through a terminal unit changes in response to upstream static pressure at constant damper position (figure 25).



Figure 25. Upstream Pressure Variations

Both of these problems combine to make the prospect of controlling the damper position in response to changes in load difficult. Ideally, we would like to be able to directly control airflow in response to load changes.

Pressure independent terminal unit controllers directly modulate airflow in response to changes in load. A pressure independent terminal unit employs the use of an airflow measuring device to directly measure the airflow through the terminal unit (figure 26). Utilizing the flow measuring device, a pressure independent controller modulates damper position in order to achieve an airflow response. It is that airflow response that is controlled in response to a change in load. Being that a pressure independent terminal unit is modulating airflow in response to load; tighter temperature control of the zone can be maintained.

Figure 26. Airflow Measuring Device

Pressure independent controllers have the added advantage of being able to maintain minimum and maximum airflows at the terminal unit. This type of control would be impossible without the airflow ring utilized in pressure independent VAV terminal units.

Most modern controllers employ the use of direct digital control (DDC) controllers to control pressure dependent and pressure independent controller. Older designs utilized both pneumatic and electronic controllers. Because these type of controllers are becoming exceedingly rare, their use will not be discussed.

Beyond the controls of the terminal unit, some method of controlling the system supply fan must be made available. Otherwise, as the VAV terminal units modulate, the supply fan may over pressurize the ductwork. Not only does this waste valuable fan energy, but the static pressures reached in the system may exceed the design rating of the ductwork, leading to a catastrophic failure of the duct system.



Two methods of VAV supply fan system control are used predominantly; the two thirds method and critical zone reset method. Both methods require some type of means to unload the fan, such as a VFD or inlet guide vanes. For further information, please see the clinic entitled “Fans”.

The most common method, known as the two thirds method of fan capacity control, involves mounting the sensor two-thirds of the way down the longest duct run (figure 27). The controller is set to maintain the pressure corresponding to that location in the duct system at design airflow conditions. The pressure is determined as the pressure at that point that would be required in a typical constant volume system. In larger systems, where the location of the sensor may not be obvious, multiple sensors may be required.

Figure 27. Two Thirds Method

However, this method is arbitrary and often leads to duct over pressurization. For example, it is impossible to tell if we are providing enough static pressure to ensure adequate airflow to all the zones. Conversely, we may be providing too much pressure, over pressurizing the ductwork and wasting valuable fan energy. The latter scenario is the more common. One clever solution which enables designers to overcome the limitations imposed with the “two-thirds” method is a controls strategy called critical zone reset. The critical zone reset controls strategy is employed by polling all of the VAV terminal units in the system at regular intervals. The VAV terminal with the highest damper position (% wide open) is tagged as the critical zone. The critical zone damper position becomes the controlled variable. A static pressure sensor, located at the discharge of the supply fan, is reset until the critical zone damper position reaches a setpoint of 95% (figure 43). In order to implement this control scheme, the building automation system must be able to poll and read the damper positions at all the VAV terminal units. This method is far less arbitrary than the two-thirds method. The control scheme ensures that none of the terminal units ever reach 100% wide open. If they did, we know that we are starving the zones for airflow and that the static pressure in the system needs to be increased. In addition, we are optimizing the energy consumption of the fan. We are never over pressurizing the system, as is common problem associated with the two-thirds method. The critical zone reset method can save as much as 50% of the total system fan energy consumption.

Figure 28. Critical Zone Reset



Design Considerations

Most zones can be divided into two types of zones; internal and perimeter zones. Perimeter zones are those zones that make contact with exterior components of the building (wall, roof, etc). Interior zones are those zones that do not make any contact with exterior components. Interior zones are completely surrounded by conditioned zones or space (figure 29).

Figure 29. Interior versus Perimiter Zones

Interior zones, by their nature, are spaces that only experience a cooling load. Because all of the spaces surrounding an interior zone are conditioned, heat loss cannot occur. Only heat gain, and thus cooling, can arise in an interior zone. This is not to say that tempering of the conditioned air may be required, depending on the minimum cooling load in the space and the minimum airflow setpoint at the terminal unit. If the minimum cooling load is less than the airflow that can be delivered at the minimum airflow setpoint, then some type of tempering will be required. This may take the form of a reheat coil in the terminal unit or direct heating in the space. If the minimum cooling load is greater than the airflow that can be delivered at the minimum flow setpoint, a cooling only VAV terminal unit may be used.

Perimeter spaces generally experience both cooling and heating loads. Thus, the VAV terminal unit must be designed to meet both of these load conditions. Table 1 summarizes the type of terminal unit, diffuser type and diffuser location as a function of the heating load condition per square foot.

Table 4. Perimeter Heating Loads For Comfort @ 4500’

Heating Load Per Square foot (Btu/hr/ft2)

Terminal Type Diffuser Location Diffuser Type

<15 VAV W/ Reheat Center of Room (figure 30) High Aspiration Rate

>15,<30 VAV W/ Reheat Diffuser blanket perimeter wall

Linear Slot at Perimeter

>30 VAV W/ Baseboard heat on perimeter

Center of Room Linear Slot or High Aspiration Rate

If the heating load is less than 15 Btu/hr/ft2, a VAV terminal unit with reheat may be employed with high aspiration rate diffusers placed in the center of the room. While hot air is buoyant, the heating load is low enough that high aspiration rate diffusers will likely assure proper mixing in the zone even with the heat being provided in the terminal unit. Cold drafts at the perimeter exposed wall should not be a concern. The discharge air temperature should be kept within 150F of the room setpoint. This generally results in a maximum heating air delivery of about one cfm per square foot.



If the heating load is between 15 Btu/hr/ft2 and 30 Btu/hr/ft2, a VAV terminal reheat unit should still be able to properly mix the warm supply air with the room air assuming linear slot diffusers are employed. Linear slot diffuser not only have a very high aspiration rate (mixing of room air with supply air), but they utilize a principle known as the coanda effect. The coanda effect occurs when air is discharged at a relatively high velocity along the surface of the ceiling. This creates an area of low pressure that causes the supply air to hug the ceiling. As it travels along the ceiling, air from the space is drawn into and mixed with the supply air stream. When the air settles to the occupied levels of the space, it has reached an average temperature (during cooling). Linear slot diffusers, when placed on perimeter walls, can also act as a thermal barrier. The high velocity stream of air acts very much like an air curtain. When the load is between 15 Btu/hr/ft2 and 30 Btu/hr/ft2 linear slot diffusers should be placed on the perimeter wall, utilizing the air curtain effect. This prevents cold drafts at the perimeter wall during high heating load conditions.

Figure 30. Diffuser near Center of Room

Finally, if the heating load is greater than 30 Btu/hr/ft2, perimeter heating should be utilized at floor level. Due to the large heating load, if terminal reheat is employed, the buoyancy of hot air would make it virtually impossible to properly mix the supply air with the room air. Drafts would be present at the perimeter of the zone. Perimeter heating, preferably using something similar to baseboard type systems, overcome the obstacle associated with trying to distribute hot air through a ceiling mounted diffuser. Baseboard heating systems should be placed on the floor at the perimeter wall. The buoyant air created by the baseboard system will prevent cold drafts from negatively influencing the occupied space. In this scenario, high aspiration rate diffusers could be placed in the center of the zone. The function of the diffusers is to provide proper mixing at all airflows during cooling and mixing at minimum position for ventilation during heating. A few guidelines should be followed when designing the ductwork for VAV systems. Those guidelines are:



Avoid elbows 3-5 duct diameters upstream of terminal boxes Keep the duct layout as simple and symmetrical as possible Apply a duct lining or duct silencer to the first section of the duct system in order to attenuate supply fan noise Place terminal units above hallways and other “unoccupied” areas to ease installation and maintenance and to

help attenuate the sound radiated to the occupied spaces Avoid the use of flexible ductwork upstream of the terminal unit Use high aspiration ratio diffusers Use fan powered boxes on perimeter in areas with moderate to high heating loads Select smallest diffuser for acceptable noise in occupied space If needed, reducing transitions should be located several duct diameters upstream of terminal units

Generally, two types of duct design methods are used for VAV systems:

Equal Friction Method Static Regain Method

The equal friction method assumes a constant pressure drop per linear foot of duct. A system utilizing the equal friction duct design method can easily be designed using hand calculation procedures. While the equal friction duct design method is acceptable for VAV systems, balancing dampers must be used throughout the system.

The static regain method strives to size the ductwork such that the static pressure at each takeoff is equal. The static regain method generally requires the use of a computer analysis tool in order to properly size the ductwork. The static regain method is the preferred method for the design of VAV systems. Because the static pressure is relatively constant throughout the system, employing the method simplifies the selection and control of the terminal units.

The 2009 ASHRAE Handbook of Fundamentals, Chapter 20, provides two basic rules for overhead heating and cooling:

In cooling mode, diffuser selection should be based on the ratio of the diffuser’s throw to the length of the zone/area being supplied, at all design air flow rates, to achieve an acceptable Air Diffusion Performance Index (ADPI).

In heating mode, the diffuser to room temperature difference (delta-t) should not exceed 15o F, in order to avoid excessive temperature stratification. ASHRAE Standard 55-2004 defines the level of acceptable vertical temperature gradation at not to exceed 5oF.

These requirements stem from the need to provide proper comfort to the space (as defined by ASHRAE Standard 55) and the need to provide adequate ventilation air to the space. ADPI is:

ADPI is the percentage of points within the occupied zone having a range of effective draft temperatures of -3oF to 2oF of the average room temperature at a coincident air velocity of less than 70 FPM.

ADPI is essentially a measure of the degree of mixing in zones served by overhead cooling systems

When air distribution is designed with a minimum ADPI of 80% the probability of vertical temperature stratification or horizontal temperature non-uniformity is low and conformance with ASHRAE Standard 55 (Thermal Comfort) is high.

ADPI does not apply to heating situations or ventilation related mixing. During heating, perimeter zones in areas with moderate to high heating loads will require as much a one cfm per square foot in order to meet the heating load, maintain proper airflow for mixing and maintain a supply temperature of less than 15o F of the room air temperature. This often necessitates the use of fan powered boxes in order to achieve the required heating airflow. ASHRAE 90.1 generally does not allow more than 0.4 cfm/ft2 of primary airflow while permitting reheat. Thus, in order to achieve higher heating airflows (1.0 cfm/ft2) fan powered boxes may be utilized. More on this subject will be touched on later in this clinic.



Ventilation of VAV Systems

VAV systems can be difficult to ventilate, especially at part loads. ASHRAE Standard 62.1, Ventilation for Acceptable Indoor Air Quality, details the calculations necessary to determine the amount of outdoor air that must be delivered to each zone. Based on the zone level outdoor airflow, the standard goes on to describe how to calculate the amount of airflow needed at the system level intake. Standard 62.1-2004, the standard referenced in most codes today, prescribes to ventilation rates for each occupancy category; one for people related sources and another for building related sources. These values are given in table 6-1 of the standard.

For calculating the ventilation air at the zone level, ASHRAE 62.1 follows a three step procedure.

1. Calculate the outdoor airflow that must be delivered to the breathing zone of the space (using table 6-1 of the standard).

2. Determine the zone air distribution effectiveness (Ez) using table 6-2 of the standard. The zone air distribution effectiveness depends on the location of the supply diffusers, return air grilles and supply air temperature relative to the zone temperature.

3. Calculate the outdoor airflow required for the zone, using the equation Voz = Vbz/Ez; where Voz = zone outdoor airflow, Ez is the zone air distribution effectiveness and Vbz = breathing zone outdoor airflow.

The breathing zone air distribution effectiveness is a measure of how well ventilation air is distributed within the zone. This is a function of the location of the diffusers and the temperature of the air. The location of the diffusers can dramatically impact how well the air mixes within the space. For example, when the supply diffusers are located on the floor and the return grilles are located in the ceiling, the zone air distribution effectiveness (Ez) is 1.2. This is a system indicative of a displacement ventilation system. Note that because the air distribution effectiveness is greater than one, it actually reduces the amount of outdoor air required to be delivered to the zone. Temperature is also is large determining factor in zone air distribution effectiveness. If air is being delivered and returned overhead and the air is warmer than the zone temperature, the buoyant air will resist mixing down at the occupant level. The buoyant air will want to rise and short circuit back to the return grille.

Table 5. Sample Table 6-2. Air Distribution Effectiveness

Location Of Supply Air Diffusers

Location of Return Air Grilles

Supply Air Temperature Ez

Ceiling Floor Cooler than zone 1.0

Ceiling Ceiling Warmer than zone ≥ Tzone + 15oF 0.8

Ceiling Ceiling Warmer than zone ≤ Tzone + 15oF provided that the 150 fpm supply air jet reaches to within 4.5 ft of floor level

1.0

Ceiling Floor Warmer than zone 1.0

Floor (thermal displacement ventilation)

Ceiling Cooler than zone 1.2

Floor (underfloor air distribution)

Ceiling Cooler than zone 1.0

Floor Ceiling Warmer than zone .7

Table 6-2 of the standard demonstrates the difficulty of designing systems that require larger heating loads with regards to providing proper ventilation air. For VAV systems and zones that require heating, employ one of the three strategies:



Supply the design air temperature such that the supply air temperature during the heating mode is within 15oF of the space temperature. Select diffusers such that velocity is 150 fpm within 4.5 feet of the floor (figure 31) on perimeter walls.

If the supply air temperature must be greater than 15oF above the space temperature, increase the outdoor airflow delivered to the space to compensate for the zone air distribution effectiveness.

Use baseboard radiant heat (creating an effective zone air distribution effectiveness of 1.0).

Figure 31. Perimeter Heating Example

In a typical multi-zone VAV system, the supply fan delivers a mixture of outdoor air and return air to multiple zones; of which will each have VAV terminal units. While the supply air delivers a constant ratio of outdoor air to return air, the amount of outdoor air required at each zone will most likely vary. For example (figure 30), if a three zone system contains three zone containing differing outdoor air fractions (Zp), what mixture of outdoor air to return air should be delivered at the air handler? For the example, if we assume 40% (zone 2, the worst case zone), we will be over ventilating zones 1 and 3.

Figure 32. Ventilation Example

Much of the unused outdoor air from zone 1 and zone 3 will be recirculated through the air handler during the next pass, further over ventilating the system. ASHRAE standard 62.1 allows the designer to take credit for this unused outdoor air in multiple zone recirculating systems. The standard defines a system ventilation efficiency (Ev) which is a measurement that takes into account the unused outdoor air that recirculates in the system. The system designer can then correct the sum of the zone level outdoor airflows, corrected for diversity, by the system ventilation efficiency in order to calculate the amount of outdoor air required at the air handler.



Two methods are described in the standard for calculating the outdoor airflow rate at the air handler. The first is the default Ev method. The second, the calculated Ev method, is very similar but will not be discussed in this clinic. For more information on the calculated Ev method, please see Appendix A of ASHRAE standard 62.1 2004.

The default Ev method involves seven steps. Those steps are:

1. Calculate the breathing zone outdoor airflow (Vbz) from Table 6-3 in ASHRAE standard 62.1 2004 2. Determine the zone air distribution effectiveness (Ez) 3. Calculate zone outdoor airflow (Voz) 4. Determine the “uncorrected” outdoor air intake (Vou) 5. Calculate the zone primary outdoor air fraction (Zp) 6. Determine the system ventilation efficiency (Ev) 7. Calculate the system outdoor air intake (Vot)

Figure 31 shows an example of a six zone VAV system serving an office building. Supply air and return air is delivered through ceiling mounted diffusers and grilles. Each zone has its own individual VAV terminal unit. The minimum cooling VAV terminal air setpoint is 30% of design primary airflow. The building occupant diversity is 75%.

North Office (1200)

Office 1 (600)

West Conf East Office (600) Office

(2000) Office 2 (2000) (800)

South Office (1200)

Figure 33. VAV System Ventilation Example

Table 6 summarizes steps 1, 2, and 3. Rp and Ra are determined from Table 6-3 in ASHRAE standard 62.1 2004. The person breathing zone and building breathing zone outdoor airflows are calculated as:

Where:

Vbz,p = people related zone ventilation airflow Vbz,a = building related zone ventilation airflow Rp = people related ventilation rate Ra = building related ventilation rate Pz = quantity people/zone Az = zone floor area

For this example, the system is a cooling only system with the diffusers and grilles being located overhead. Thus, the zone air distribution effectiveness (Ez) is 1.0 for all of the zones.



The third step is to find the zone outdoor airflows. The zone outdoor airflow are

Table 6. VAV System Ventilation Example

Rp (cfm/p) Pz (qty)

Vbz,p (cfm)

Ra (cfm/ft2) Az (ft2)

Vbz,a (cfm)

Vbz (cfm) Ez

Voz (cfm)

Office 1 5 4 20 0.06 600 36 56 1.0 56

Office 2 5 6 30 0.06 800 48 78 1.0 78

South Office 5 8 40 0.06 1200 72 112 1.0 112

East Office 5 18 90 0.06 2000 120 210 1.0 210

West Office 5 10 50 0.06 2000 120 170 1.0 170

North Office 5 12 60 0.06 1200 72 132 1.0 132

Conf Room 5 20 100 0.06 600 36 136 1.0 136

System Totals 78 390 504

The fourth step is to determine the “uncorrected” outdoor air intake (Vou). The uncorrected outdoor airflow would be the amount of air required at the air handler if the system was 100% efficient. That is to say, all of the outdoor air introduced at the air handler is completely utilized by the occupants during its first pass through the occupied zone. The equation for calculating the uncorrected outdoor air intake is:

Where: D = occupant diversity The equation for calculating the uncorrected outdoor air intake does allow the designer to adjust the people related zone ventilation airflow by the occupant diversity. If the occupant diversity is known, this allows the designer to reduce the amount of outdoor air required at the air handler. The equation for calculating the occupant diversity is:

Where: Ps = Expected peak system population Sum(Pz) = Sum of the peak populations Note that the occupant diversity can only be applied to the people component of the ventilation requirement. The building component, according to the standard, should always remain constant regardless of occupancy. For our example, the uncorrected outdoor air intake would be:



The fifth step is to calculate the fraction of outdoor air to primary air that is required in each zone (Zp).

Where:

Zp= Zone Ventilation Fraction Voz = zone outdoor airflow, cfm Vpz = minimum expected primary airflow, cfm In a VAV system, the primary airflow changes in proportion to the load. ASHRAE 62.1 recommends that the minimum expected primary airflow be the minimum airflow setting for the terminal unit. If the primary airflow is expected to remain above the minimum cooling flow setpoint, that flow may be used. Using airflow higher than the minimum cooling flow setpoint, if such a flow is a reasonable expectation, will result in higher system ventilation efficiency. Calculating Zp for the example system (table 7) based on 30% minimum cooling airflow (Vpzm), we see that the zone outdoor air fraction varies between .23 and .50. One question that should be asked when analyzing the zone outdoor air fractions is: should each zone have the same minimum airflow setpoint? That question will be examined with more scrutiny later in this clinic.

Table 7. Zone Ventilation Fraction

design Vpz (cfm)

Vpzm

(cfm) Voz (cfm)

Zp (Voz/Vpzm)

Office 1 400 120 56 0.47

Office 2 600 180 78 0.43

South Office 1600 480 112 0.23

East Office 2500 750 210 0.28

West Office 2450 735 170 0.23

North Office 1250 375 132 0.35

Conf Room 900 270 136 0.50 The sixth step is to determine the ventilation efficiency (Ev). The ventilation efficiency is determined based on the zone with the largest outdoor air fraction, Zp. Table 8 is an excerpt from ASHRAE 62.1.

Table 8. Exerpt From ASHRAE 62.1

Maximum outdoor air fraction, Zp Ventilation Efficiency, Ev

≤0.25 0.90

≤0.35 0.80

≤0.45 0.70

≤0.55 0.60

>.55 Use Appendix A

Note that if the largest outdoor air fraction is greater than .55, the calculated Ev method (from appendix A) must be used. Interpolation between values is permitted. Examining the outdoor air fractions from the example, we see that the largest outdoor air fraction occurs at the conference room. That outdoor air fraction is .5. Interpolating between values, we see that the ventilation efficiency for the example is 0.65. The final step is to calculate the required outdoor air intake for the system. This is determined using the equation:



Where: Vot = Outdoor intake for system Ev = Ventilation Efficiency For our example, the required outdoor air intake for the system is:

For this particular example, the fraction of outdoor air is 12.6% (1225 cfm / 9700 cfm). Note that the conference room is driving outdoor air intake for the system. If we were to reduce the minimum airflow setpoints on the conference room and office 1 and office such that the outdoor air fraction is .35, the new outdoor air intake for the system would be:

This would result in a 19% reduction in the outdoor air intake for the system. The question becomes, can the minimum airflow setting be increased from 30% for those zones? In order to achieve a maximum zone outdoor air fraction of .35; office 1, office 2 and the conference room would have to have the following minimum airflows:

Table 9. Revised Minimum Airflow

design Vpz (cfm)

Vpzm (cfm)

Voz (cfm)

Zp (Voz/Vpzm)

% Min (Vpzm /Vpz)

Office 1 400 160 56 0.35 40%

Office 2 600 223 78 0.35 37%

Conf Room 900 389 136 0.35 43%

Where the new minimum expected primary airflow (Vpzm) is:



Based on achieving a 0.35 air fraction (Zp), we see that the minimum cooling airflow for Office 1, Office 2 and the Conference room are 40%, 37% and 43% respectively. However, can we increase the minimum primary airflow in these zones without adversely effecting occupant comfort (if reheat is not employed) and without vastly increasing heating energy consumption (if reheat is employed)? In this example, all three zones are interior zones. Interior zones are generally cooling only zones. Thus, it is very likely we could increase the minimum primary airflow setpoint to these zones without adversely affecting space comfort or negatively impacting heating energy consumption. What we have done is reduced the outdoor air intake for the system by 19%, saving both cooling and heating energy for the system. The main point to this example is that the minimum primary airflow should be an interpolated value based ventilation fraction to the zone. Rarely should the minimum primary airflow setting be set to a fixed percentage. A calculation spreadsheet (like 62MZCalc) can be very helpful when determining the minimum airflow setting at each terminal unit.

ASHRAE 90.1 provides several guidelines for determining the minimum airflow at which simultaneous heating and cooling may occur. Generally speaking, the minimum allowable primary airflow at the terminal unit will likely be limited by the airflow at which tempering may occur. If that airflow is above the larger of:

1. ASHRAE Standard 62’s requirement for outdoor air 2. 0.4 cfm/ft2 3. 30% of Supply Air 4. 300 CFM 5. Site recovered or site solar provides >=75% of reheat energy

In the example above, both the offices fall below the 0.4 cfm/ft2 requirement at 160 cfm and 223 cfm respectively. Thus, we can increase the minimum setpoint at the terminal unit for those zones. However, the conference room cannot be lowered below the 30% minimum of supply air per the requirements above while providing reheat. However, being that the conference room as an interior zone and will likely only be in cooling, we may be able to provide a cooling only terminal unit and still satisfy the temperature requirements in the space at the higher minimum setpoint.

Note that example assumed a cooling only application. In most cases, the outdoor air intake for the system will need to be re-calculated during heating load conditions. Under heating conditions, often the breathing zone air distribution effectiveness will be something less than the cooling breathing zone air distribution effectiveness. However, the minimum expected primary airflow will be greater at the terminal unit when in the heating mode. The overall result may or may not lead to a value greater than the calculated outdoor intake air for the system under cooling. Whichever value is greater should be used as the design value for the system.

ASHRAE 90.1 provides a means to provide a greater amount of heated or tempered air while the system is in cooling mode. ASHRAE allows the designer to increase the heating airflow at the terminal unit in order to maintain a discharge air setpoint of equal to or less that 15oF above the room temperature (figure 32) so long as the maximum heating airflow does not exceed 50% of the maximum primary airflow. This allows the designer more flexibility to meet the requirements of ASHRAE standard 55 while at the same time providing additional heating capacity at the terminal unit.

Figure 34. Increasing Tempering/Heating Minimum Airflow



Additionally, the calculated Ev method may result in lower system outdoor airflow rates. As such, it would be prudent for the designer to analyze the system using both methods, determining which method results in lower over outdoor airflows for the system.

Fan powered VAV terminal units have two potential sources for which the box may introduce outdoor air into the zone. The first is the primary air stream which is supplied from the system air handler. The second is the plenum fan air, which draws in recirculation air return from the space, which may have some amount of unused outdoor air. Dual duct systems potentially provide multiple air paths for ventilation air. For more discussion on the calculations required for systems with multiple paths, refer to the May 2005 ASHRAE Journal article titled “standard 62.1: Designing Dual-Path, Multiple-Zone systems.”

Thus far, this discussion has focused on how much outdoor air must be introduced into a VAV system at design loads. However, for the vast majority of the occupied hours, the building will be at part load and part occupancy. ASHRAE 62.1 does allow the designer to adjust the amount of outdoor air introduced into the system as occupancy varies.

What is commonly done, and rarely if ever works, is using a fixed air damper position at the air handler. The system is balanced by a balancing contractor at full load airflow and the outside air damper position is fixed at a position which introduces the design outside airflow at the maximum primary design airflow setpoint. While this method will properly ventilate the system at full design airflow, as the fan modulates at part load, the system will likely be under ventilated. For example, assume a system is designed to deliver 2000 CFM at full load and the system is balanced such that the outside air damper position is set to 25% in order to achieve 500 CFM of outdoor air (the design outdoor air intake). What happens as the fan modulates in response to the VAV terminal units shutting down at part load? How much outdoor air would be introduced into the system if at 50% load, the system supply fan modulates down to 1000 CFM? Will the system maintain the required 500 CFM of outdoor air at the air handler? The answer is, absolutely not. As the supply fan modulates, less outdoor air will be drawn across the outdoor air damper, likely under ventilating the system.

Two methods, both of which assume constant occupancy and thus a constant system outside airflow rate, are generally employed. The first method is to proportionally vary the outside air damper position in response to fan capacity (figure 32). The outside air damper position required to maintain Vot is noted at design airflow and at the minimum fan primary airflow. The controls system then proportionally varies the damper position between these two points as the fan modulates.

Figure 35. OA Damper Reset

While this method generally results in bringing in enough outdoor air into the system, it generally slightly over ventilates the system. A better system utilizes multiple points (generally five points total at 20% intervals) and creates a curve relating damper position to fan airflow. The control system then interpolates between the five points to create an accurate non-linear curve that tracks damper position to fan airflow.

The second method utilizes an airflow monitoring station at the system air handler (figure 33). The control system measures the amount of outside air being introduced at the outside air intake as the system supply fan modulates. The control system then modulates the outside air damper in order to maintain the design outdoor air ventilation rate.



Figure 36. Airfow Monitoring Station

ASHRAE 62.1, Section 6.2.7 permits dynamic reset of the system outdoor airflow as operating conditions change. This is allowed so long as the system provides, at a minimum, the required breathing zone outdoor airflow (Vbz) whenever the zone is occupied. Three methods are permitted. Those methods are:

Resetting the intake outside airflow in response to variations in zone population.

Resetting the intake outside airflow in response to variations in ventilation efficiency

Resetting VAV minimums in response to variations in intake airflow

The first method, commonly referred to as demand controlled ventilation, resets the intake outside airflow (Vot) based on changes in the zone population. Some methods of estimating the zone population include occupancy schedules, occupancy sensors and carbon dioxide (C02) monitoring. C02 monitoring is becoming increasingly popular among designers. Appendix C of ASHRAE 62.1-2004 provides a mass balance equation to predict the difference between indoor (Cs) and outdoor (Co) concentrations of carbon dioxide at steady-state conditions:

Where: Vo = outdoor airflow rate, cfm/person N = CO2 generation rate, cfm/person CS = CO2 concentration in the space, ppm CO = CO2 concentration in the outdoor air, ppm

Recalling that the zone outdoor airflow (Voz) is:

Where:

Vbz,p = people related zone ventilation airflow Vbz,a = building related zone ventilation airflow Rp = people related ventilation rate



Ra = building related ventilation rate Pz = quantity people/zone Az = zone floor area Combining equation X and equation Y, Voz is (equation Z):

A typical value of N for seated, light work activity is 0.0105 CFM of C02/person. Being able to estimate occupancy using

equation Z (assuming CO2 sensors in each zone), the control system can determine the zone with the largest outdoor air fraction (Zp), determine that zones associated ventilation efficiency (Ev), and then calculate the system outdoor airflow (Vot).

While placing CO2 sensors in every zone in order to determine occupancy will maximize energy savings at part occupancy, the associated first cost may not yield an acceptable return on investment. A better option may be to place CO2 sensors in the zone with the highest occupant diversity (conference rooms for example), assume all other zones maintain a constant occupancy, and then calculate the system outdoor airflow (Vot). In many instances, this option provides the designer with sufficient energy savings while reducing the first cost investment (figure 34).

Figure 37. CO2 Sensor Example

Demand control ventilation is required by ASHRAE 90.1 when:

Spaces larger than 500 sq-ft and with a design occupancy for ventilation of greater than 40 people per 1000 sq-ft of floor area and served by systems with one or more of the following

An airside economizer

Automatic modulating control of the outdoor air damper or,

A design outdoor airflow greater than 3000 CFM

The exceptions being when:

Systems with Energy Recovery complying with Section 6.5.6.1

Multiple Zones without DDC control in individual zones communicating to a central control panel.

Systems < 1200 cfm of outdoor airflow.

Thus, in many VAV applications, demand control ventilation will be required.



The second method resets the intake outside airflow (Vot) in response to variations in ventilation efficiency (Ev). This method, also known as ventilation reset, resets the amount of outside air based on the actual ventilation efficiency of each zone. Recall that the ventilation efficiency (Ev) is a function of the fraction of outdoor air to primary air for each zone given as a function of the maximum outdoor zone fraction (Zp):

If the system served by the air handler is largely cooling load dominated, then it is very likely the maximum zone Zp is much larger than that if it was assumed all of the terminal units were at their minimum primary airflow. This increased primary airflow at the VAV terminal unit will decrease Zp, thus decreasing Ev. Recall that as the ventilation efficiency increases, less outdoor air is required for the system serving those zones. In order to utilize ventilation reset, the control system must be able to monitor and poll the primary airflow at the VAV boxes, calculate new ventilation efficiency, and reset the outdoor air damper position based on an airflow monitoring station at the outdoor air intake.

The last method, resetting VAV minimums in response to variations in intake airflow, involves resetting the VAV minimum damper position to a lower value when the outdoor air intake brings in more air than required at design. This typically happens when the system economizes. If a system serves zones that require simultaneous heating and cooling and the building is located in an area with a significant number of economizer cooling hours, this method may save a sizeable amount of reheat energy. This generally occurs when a system serves both perimeter and interior zones. In this scenario, it may behoove the designer to utilize fan powered boxes on the perimeter zones in order to decrease the amount of reheat energy. This may negate the need reset the VAV minimums while economizing.

All three of the methods described involve relatively complex control scenarios and often lead to over ventilation of the zone and thus wasted energy. Often, a designer will incorporate control systems that incorporate all three methods of resetting ventilation air. However, the designer may choose to employ a dedicated outdoor air system in order to simplify the process of maintaining adequate ventilation and reduce energy cost associated with single duct VAV ventilation systems. With a dedicated outside air system, it becomes very easy to maintain precise amounts of outside air to each zone.

Single Zone VAV Systems

Single zone VAV systems, similar to CV systems, use a single room temperature sensor to control a central air handing unit. However, the zone temperature sensor for a constant volume system would be used to modulate the delivery temperature. The air handler would deliver a constant volume of air. With a single zone VAV system, the room sensor would control the capacity of the supply fan. In this instance, the delivery temperature would remain constant (figure 35).

Figure 38. Single Zone VAV System



Single zone systems are becoming increasingly popular due to their energy savings potential compared to similar constant volume systems. The coefficient of performance for a fan (COP) is always less than one. That is to say, it will always produce less power out than the input power applied. However, the COP for most cooling systems, like centrifugal chillers for example can achieve COP’s of very near 7.0. In addition, as cooling systems unload, their COP’s only increase marginally. For example, a centrifugal chiller may see a 50% reduction in energy at part load, under ideal circumstances. A fan however, when unloaded to 50%, will see an 87.5% reduction in energy (refer to the discussion of fan laws at the beginning of this clinic).

Thus, systems with larger fan brake horse power requirements, lend themselves well to single zone VAV designs. In addition, single zone VAV systems are soon to be the law of the land. ASHRAE 90.1 2010 requires a fan to capable of unloading to at least 50% if the fan motor horsepower is great than 5 hp. Section 6.4.3.10, “Single Zone Variable Air Volume Controls, reads:

a. Air handling and fan-coil units with chilled-water cooling coils and supply fans with motors greater than or equal to 5 hp shall have their supply fans controlled by two-speed motors or variable-speed drives. At cooling demands less than or equal to 50%, the supply fan controls shall be able to reduce the airflow to no greater than the larger of the following:

1. One half of the full fans speed, or 2. The volume of outdoor air required to meet the ventilation requirements of Standard 62.1.

b. Effective January 1, 2012, all air-conditioning equipment and air-handling units with direct expansion

cooling and a cooling capacity at AHRI conditions greater than or equal to 110,000 Btu/h that serve single zones shall have their supply fans controlled by two-speed motors or variable-speed drives. At cooling demands less than or equal to 50%, the supply fan controls shall be able to reduce the airflow to no greater than the larger of the following:

1. Two-thirds of the full fan speed, or 2. The volume of outdoor air required to meet the ventilation requirements of Standard 62.1

While ASHRAE 90.1 2010 has not been adopted as code in many areas, it will likely be adopted in the years to come.

Single zone VAV system, while offering a tremendous potential for energy savings, present a slightly more complex control scheme.

Referring to figure 38, the room temperature sensor directly controls the delivered air volume in a single zone VAV system. However, the dynamics of both building pressurization and ventilation will change as the supply fan modulates. With constant volume systems, it is easy to maintain the amount of outside air to the space at all operating loads. This is because the supply fan airflow remains constant. This is not the case with single zone VAV systems. As the supply fan unloads, less outside air will be drawn though the outside air damper (assuming a fixed position). Two methods of ensuring proper ventilation are generally employed with single zone VAV systems. Those methods are:

Proportional or curve generating damper control

Airflow monitoring station

Proportional damper or curve generating damper control can generally only be used with a fixed volume of outside air. Demand control ventilation or ventilation reset cannot be used. With fixed damper or curve generating damper control, the damper position which delivers the design outside air volume at full load airflow is determined using the calculation methods discussed earlier in this clinic.

Alternately, an airflow monitoring station may be used. Airflow monitoring stations allow the designer to implement more advanced energy savings strategies, such as demand control ventilation or ventilation reset. For more information on how to control a system utilizing demand control ventilation or ventilation reset, please see the section of this clinic entitled “Ventilation of VAV Systems.”

Pressurization can be accomplished with a room pressure sensor that is used to control either barometric dampers or an exhaust fan (depending on the amount of return static pressure experienced by the system). If the system has much more than 0.1” static pressure on the return system, exhaust fans should be employed in order to maintain proper building pressurization.

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