Thursday, October 18, 2012

Q&A: How Do I Determine Vertical Throw For a Grille When Catalogs Only Show Horizontal Throw?

Vertical throw for supply grilles is not usually required since grilles mounted in a ceiling are not a normal application. But it sometimes happens and diffusers can be set to throw vertically.

The method of obtaining throw data is as follows.

1. Determine the flow rate in CFM and the jet velocity from the published horizontal throw data pages.

2. Refer to the engineering section of the Titus catalog; page B24. Find the sections entitled Estimating Downward Vertical Projection.

3. Using the table below, find the vertical line representing your flow ate in CFM.

4. Follow the vertical line to the jet velocity curve that corresponds to core velocity listed in the published data.

5. Project a horizontal line from where the vertical flow rate line intersects the velocity curve to the column on the right hand side of the figure.

6. Select the appropriate temperature differential (delta T) column and read that vertical distance.

Example: Using the published data from the data page above, determine a vertical projection of a 36 x 18 grille supplied with 2500 cfm at a delta T of 20 degrees.The core velocity at the top of the data page for that grille size and volume is 600fpm.

Using the procedure described above and Figure 24, we can estimate vertical projection for a grille supplied with 2500cfm and a core velocity of 600 fpm to be nearly 12 feet with a 20 degree delta T.

Additionally, the table is also useful for comparing isothermal throw to differential throw when only the isothermal data is available, by simply reading across the vertical projection columns on the right.

For example, when an isothermal vertical throw of 25 feet is known, simply read across the column to determine a vertical throw of 15 feet for vertical throw when the delta T is 20 degrees.

Mark Costello - GRD Product Manager

Q&A: I Need Performance For an Oversize Grille But I Can’t Find a Duct Area That Matches - How Do I Calculate Performance?

1. Divide the duct area by two or more until you finda comparable duct area, and then multiply the throw by 1.4, this is a common constant for the adjustment in total air mass that is used for extrapolating throw performance. This conversion factor is also used in all of the linear diffuser conversion charts. 

2. A good rule of thumb on sound is to increase the sound data by 3 NC to obtain a cumulative value.

Mark Costello - GRD Product Manager

Q&A: How Do I Calculate Performance Data For Grille Sizes That Are Not Published in a Catalog?

Most manufacturers have a selection software program that can help solve this problem. 

The Titus selection software program TEAMS is available as a free download from the Titus website.  TEAMS contains many grille sizes not published in the catalog, and allows the user to input air volume (CFM) in increments other than those published.

If software like TEAMS is not available, the following method should be used.

1. Determine the nominal duct area in square inches by multiplying the nominal length and width of the desired grille size in inches.

 E.g. 64 x 10 = 640 square inches

2. Convert the duct area from square inches to square feet by dividing the total square inches by 144 and rounding to a 2-place decimal.

E.g. 640 square inches / 144 = 4.44 square feet

3. Locate a published grille size in the catalog for which the nominal duct size is approximately the same. The nominal duct area in square feet is the column directly to the right of the listed nominal duct sizes.

E.g. 4.44 square feet approximates to the 4.50 square feet shown for a 36 x 18 supply grille, and is less than is less than a difference of 2%. Duct leakage often accounts for more of a discrepancy than 2% in terms of performance.   

4. Locate the nearest approximate desired volume listed for the approximated grille size.

E.g. A desired volume of 2300 CFM falls approximately midway between the listed volumes of 2110 and 2532, therefore the median throw and NC values between the values for 2110 and 2532 should be used.

The median value of NC 23 and 28 is 25 after rounding to an integer value.

The median values of the 0 degree deflection values 39-59-84 and 47-65-93, are 43-62-88.

Mark Costello - GRD Product Manager

Tuesday, August 21, 2012

Q&A: Compare Series vs. Parallel Fan-Powered Terminal Units

There are two types of fan-powered terminal units - series and parallel. Every manufacturer offers both types and special variations such as low profile and quiet units. Although the type of unit to use is often up to personal preference, there are distinct differences:

Series fan-powered terminals have fans that must run throughout the occupied mode in order to deliver ventilation air to the zone:

  • These units act as boosters for the air handler because their fans move the air the rest of the way to the zone. This allows the air handler to run at system pressure far lower than other types of terminal units require. The typical system pressure supplying series fan boxes is 0.50 IN WG.

  • Since the fan runs continuously during occupied periods, they provide constant air motion and more air changes than other types of terminal units.

  • The continuous operation of the fan results in relatively constant sound levels, unlike other types of terminal units that vary air volumes and/or cycle fans.
Parallel fan-powered terminals have fans that only switch on during the heating mode to pull warm return air from the ceiling plenum:

  • Since the unit fan is off during the cooling mode, the box acts like a single duct VAV and simply varies airflow from the air handler to maintain room temperature. Typical system pressures are between 1.00 and 1.50 IN WG.

  • Some engineers do not specify parallel fan units because the fan cycling is often noticeable to occupants.

  • Parallel fan units must include a backdraft damper to prevent primary air from leaking back through the blower into the ceiling plenum. Leakage around the backdraft damper can be an issue and could be considerable when downstream pressure requirements are greater.
An ASHRAE research project (RP-1292) completed in 2007 was conducted to determine which type of fan-powered terminal used the least energy from a whole building perspective. The report said that either unit could be equally efficient when properly sized and applied. This original report only included units with standard PSC fan motors. A subsequent addendum to the report, paid for by a consortium of interested parties, took the newer ECM technology into account in the same energy model. It gave more of an advantage to the series fan units.

 Randy Zimmerman - Chief Engineer

Monday, August 13, 2012

Q&A: What is End Reflection and How Will Discharge Sound Ratings for Terminal Units Change in 2012

There is an acoustical phenomenon known as end reflection that is regularly encountered in HVAC systems. It occurs whenever air flowing in a duct reaches an outlet and suddenly expands to fill a room. Although it might not be obvious to the casual observer, sound doesn’t necessarily travel in the same direction as airflow. The greater the degree of expansion, the more sound is reflected away from the room.

An acoustician might say, “End reflection is the acoustic energy in an acoustic test duct that is prevented from entering the test space by the impedance mismatch created by the termination of the acoustic test duct”. In layman’s terms, whenever a rapid air expansion occurs, some portion of the sound energy generated by the supply device (i.e. terminal unit, fan system, air handler, room fan coil, etc) travels upstream back towards the source. This is end reflection.

All terminal unit manufacturers test their products in accordance with ASHRAE Standard 130 ‘Methods of Testing Air Terminal Units’. This standard provides testing procedures for both radiated and discharge sound. End reflection has been known to affect discharge sound readings for many years, so the standard was amended in 1994 to specify that discharge ducts in discharge sound tests must terminate flush to the inside wall of the test chamber. This was necessary because the further a discharge duct projects into the test chamber, the more end reflection occurs, effectively lowering the sound levels measured within the test chamber.

Test data measured in accordance with ASHRAE Standard 130 is used to produce catalog data in accordance with AHRI Standard 880 ‘Standard for Performance Rating of Air Terminals’. The latest version of this standard (880-2011) went into effect on January 1, 2012. It requires that manufacturers calculate the end reflection loss (ERL) and add it back to the rated discharge sound power levels of terminal unit products. A formula based on ASHRAE Research Project RP-1314 is used to calculate the ERL for the dimensions of the discharge duct used during the sound test. Although the calculated ERL is most accurately applied to 1/3 octave sound data, manufacturers may apply it to existing full octave sound data through 2014.

Here’s how the ERL is calculated:

First determine De, the equivalent duct diameter (ft). If the discharge duct is round, simply use the duct diameter. In the more likely situation that the discharge duct is rectangular, the equivalent duct diameter must be calculated as:

De = SQRT [(4 x A) / (144 x π)]


A = cross sectional area of duct (in2)

So for a terminal unit with a 15 in by 12 in discharge duct:

De = SQRT [(4 x 180) / (144 x π)] = 1.26 ft2


ERL = 10 log [1 + (0.7 x  Co/π x f x De)2]


Co = Speed of sound in air (use 1128 fps)
f = Octave band center frequency (Hz)
De = Equivalent diameter of the duct (ft)

So the end reflection loss of a 15 in x 12 in discharge duct is:

2nd Octave band (125 Hz) = 5 dB
3rd Octave band (250 Hz) = 2 dB
4th Octave band (500 Hz) = 1 dB
5th Octave band (1000 Hz) = 0 dB
6th Octave band (2000 Hz) = 0 dB
7th Octave band (4000 Hz) = 0 dB
Adding this to the existing discharge sound levels of a fan-powered product would likely raise the NC level by 6 points. The smaller the discharge duct is, the greater the correction will be. Since research has shown that low frequency corrections tend to become overstated as duct sizes get very small, the maximum correction is limited to 14 dB.
So what does all this mean?

It means that every manufacturer will need to update all published terminal unit discharge sound performance data and selection software to meet the latest standards. Discharge sound levels will increase for all terminal units and smaller units will see the largest increases. The effect on large units could be negligible.
Will the actual discharge be higher?

No. The product will perform exactly as it did before, but now all of the sound energy will be properly accounted for. It would be fair to say that under the previous standard, discharge sound was in many cases being understated.

In order to change the certified performance listings posted on the AHRI website, all participating manufactures were required to resubmit all of their products to the program. It will probably be months before all of the changes are complete and posted on their new website. Although AHRI has agreed to publish a full page announcement in trade magazines to explain why these changes are necessary, it has not yet been sent out for membership approval.

Randy Zimmerman - Chief Engineer

Monday, July 16, 2012

Q&A: How Do You Size Parallel Fan Powered Terminal Units?

Parallel flow (variable volume) fan powered terminals are selected based on their capacity to handle the primary airflow. The same rules which apply to the selection of single duct terminals can be used, except that water coils are not in the primary airstream path, and will not affect sound levels. The pressure drop of the water coils, however, which are on the fan inlet in Titus parallel fan units, must be added to the expected discharge pressure at the fan flow rate when entering the fan curve tables.

The fan is selected based on the minimum airflow requirements for the space or the heating load required. In most cases the fan can be downsized from the cooling flow requirement considerably, reducing both first cost and operating cost. The fan is selected from the fan curves. The downstream static pressure of the secondary air may not be the same as the primary air, however. If the secondary airflow requirements are less than the primary air requirements, the static pressure will be reduced. The following equation can be used to determine the static pressure at reduced airflows. (Do not forget to add water coil pressure drops to the fan requirement).

To select a Titus parallel fan powered terminal, refer to the published fan curves and primary air pressure drop curves, together with the application and sound power data.

 In the parallel flow type of unit, when the primary air is ON, the fan is typically OFF, and vice versa. As shown in the Figure 1, the primary air and the fan discharge air follow parallel paths into a common plenum. Therefore both airflows will encounter the same downstream resistance at a given flow rate.

Since the primary and secondary airflows come from two different sources-and often at two different specified flow rates-the volume vs. pressure relationship in each of these airflows must be checked to ensure adequate flow rates under actual job conditions.

Example: Select a Model DTQP for a maximum of 1400 cfm of primary air with 1.00” wg inlet static pressure. The fan airflow required is 1150 cfm. The downstream resistance offered by the duct and diffusers has been determined to be 0.30” static pressure at 1150 cfm.

Primary Air: From the air inlet pressure table, a size 4 with a 12” inlet will handle 1400 cfm of primary air with a minimum static pressure drop of 0.23” through the primary air section. But since the downstream resistance is 0.30” at 1150 cfm:

The overall primary air static pressure drop is:

0.23”+ 0.44”= 0.67” sp

Since a 1.0” static pressure is available at the inlet, the selection will work. The damper in the primary air section will do some throttling to hold the maximum air flow to 1400 cfm.

Secondary Air (Fan): From the fan curves, a size 4, without coils, terminal will handle 1150 cfm at 0.30” static pressure, with the proper setting of the standard SCR speed control.

Trenton Yarbrough - Director of Engineering

Tuesday, July 10, 2012

Q&A: Why Are Filters for Terminal Units An Option?

Optional filters on fan-powered terminal units are 1” thick disposable filters that are highly recommended if there is any chance that the units will be operating during construction. Construction dust can easily ruin both the motor and blower, if the units are operated without filters in place. Dust tends to deposit unevenly on blower wheels resulting in loss of balance. Any build-up of fine dust or lint on the exterior of a permanently-lubricated fan motor can create a migration path and wick the oil out of the bearings.

 These optional filters should later be removed and discarded during the balancing and commissioning process. Operating units after construction with filters in place is not recommended. Fan-powered units, like all terminal units, are designed for zero maintenance. This is necessary because terminal units, unlike air handlers, are typically located above finished ceilings in tenant spaces. Accessing the units regularly to change filters can be nuisance to tenants, time-consuming for maintenance personnel and can easily result in ceiling damage.

 If filtering of return air is desired, we recommend the installation of filter grilles in the ceiling. This allows the building and the building owner the benefits of less costly standard filter sizes; more filter area for longer service intervals and quicker access for easier replacement.

Randy Zimmerman ~ Chief Engineer

Tuesday, June 26, 2012

NC Ratings for Diffusers

Designers should always take into account acoustical performance when selecting grilles, registers and diffusers, but it is important to understand how manufacturers’ noise criteria (NC) ratings are determined, what they mean and how they should be applied.

All grilles, registers and diffusers should be sound tested in a reverberant chamber per ASHRAE Standard 70 Method of Testing the Performance of Air Outlets and Air Inlets. This provides the proper test procedure needed to determine the raw sound energy or sound power level (Lw) of a device in terms of decibels (dB). The tests are conducted in reverberant chambers in order to have an environment with low sound absorption, no directionality from sound sources and adequate isolation from background noise interference. Measured sound pressure levels (Lp) are corrected to calculated sound power levels by adding room absorption determined by the use of a reference sound source (RSS).

Few manufacturers publish catalog sound power levels but rather provide noise criteria ratings. According to AHRI Standard 885 Procedure for Estimating Occupied Space Sound Levels in the Application of Air Terminals and Air Outlets, the room sound level for any grille, register or diffuser can be estimated by deducting 10 dB of room absorption from the sound power level in each octave band prior to determining the NC level.

In order to meet a particular NC level, the sound pressure level in each octave band cannot exceed a maximum level. The NC curves were determined based on the human response to sound levels of various frequencies by adults with average hearing in the 1940’s. So long as the sound levels in each octave band are less the maximum allowable levels for a given NC level, the overall room sound level is said to be in compliance.

Now it’s time to start asking questions:

  • Where does the 10 dB room absorption come from?
  • Is every room the same?
  • Does this same 10 dB deduction apply to other devices?

First off, the 10 dB room absorption deduction has been around for a long time and it only applies to grilles, registers and diffusers. These products tend to generate their highest sound levels in the 4th (500 Hz), 5th (1000 Hz) and 6th (2000 Hz) octave bands. This 10 dB deduction is meant to approximate the room absorption in these three critical octave bands for a typical office space. This typical space is a medium-sized room with some office furniture, commercial carpet, a lay-in ceiling and sheetrock walls.

No two rooms are exactly the same, but when selecting diffusers for typical office space, the 10 dB deduction is adequate to estimate the actual room sound level of these devices within a reasonably level of accuracy. The 10 dB deduction should not be applied to other devices such as terminal units and room fan coils, because these products tend to create maximum sound levels in other octave bands that could have more or less room absorption in a typical office environment.

Diffusers should never be selected in such a way that they will be heard. This is due to the fact that they produce their maximum sound levels in octave bands 4 thru 6. These octave bands are known as the ‘speech interference bands’ because this portion of the audible sound spectrum is also used for vocal communication. Diffuser selections that result in audible noise would likely create poor environments for speech communication. In order to avoid such issues, it is a good idea to select diffusers for room sound levels at least 10 NC points lower than the desired room sound level.

When selecting grilles, registers and diffusers for atypical spaces such as performance halls, laboratories or even office spaces with hard-surfaced floors, it is necessary to use the sound power levels (Lw) to estimate the room sound level. There’s no sure way to work backward from manufacturers’ published NC levels to determine sound power levels. Sound power levels for grilles, registers and diffusers can be easily obtained from the Titus TEAMS selection program by making a performance selection for a specific flow rate. With this information, any qualified acoustical consultant should be able to estimate sound performance for an atypical space.

More could be said regarding the acoustics of these products, but I’ll leave that for another time and another article.

Randy Zimmerman ~ Chief Engineer

Tuesday, May 22, 2012

Q&A: How Do You Determine the Hand of a Titus Terminal Unit?

Generally speaking, the hand of a Titus terminal unit is defined as the location of the control package. This is determined by looking into the primary inlet duct. If the controls or damper shaft are on your left, it’s a left hand box. If the connections of an optional water coil are on the right, the unit has a right hand coil.

There are a few additional things to know:

  • Single duct and series fan-powered terminal units can be ordered with water coil connections on the same side or the opposite side from the controls. It is more common to locate the water coil valve on the same side of the unit as the damper controls to simplify the valve wiring and provide service access from a single location. Presumably, if there is good access on one side of the unit for the damper controls there should also be sufficient space for the plumbing.

  • On Titus terminal units with electric coils, the electric coil controls are always located on the same side of the unit as the damper controls.

  • The hand of a Titus dual duct terminal unit is defined as the location of the cold duct controller. This is determined by looking at the inlet panel. If the cold duct is on the right, it’s a right hand box. It is assumed that the control package on the cold duct is in the master control box and that any supply power would be pulled to this location. The hot duct controls are assumed to be controlled and powered by the cold duct controls.

  • On Titus parallel fan-powered terminal units, the standard location for a hot water coil is on the induction port. To prevent possible interference between the primary ductwork and coil plumbing, these coils should be ordered in the opposite hand from the damper controls. A left hand unit should therefore be ordered with a right hand coil.

  • Titus parallel fan-powered terminal units can be special ordered with hot water coils located on the unit discharge. In this location the coils can be specified to be either right or left hand, just like series fan-powered terminal units.
Hopefully this explanation will help answer questions many people have regarding the handed orientation of Titus terminal units.

Randy Zimmerman - Chief Engineer

Friday, April 20, 2012

Designing for Comfort per ASHRAE Standards 55 and 62.1

The goal of a room air distribution system is to provide thermal comfort and a healthy living environment for occupants in the space. ASHRAE Standard 55-2010 Thermal Environmental Conditions for Human Occupancy and ASHRAE Standard 62.1-2010 Ventilation for Acceptable Indoor Air Quality provide designers with the guidance to optimize health and comfort in building spaces. Many codes including LEED 2009 require compliance with these ASHRAE Standards. This article will outline the goals of these standards and illustrate how to comply with these requirements.

Dsplacement ventilation diffuser shown in the
Willard Elementary School in Concord, Massachusetts.

The occupied zone as defined by Standard 55-2010 reads: “the region normally occupied by people within a space, generally considered to be between the floor and 6 ft. level above the floor and more than 3.3 ft. from outside walls/ windows or fixed heating, ventilation, or air-conditioning equipment and 1 ft. from internal walls.” The space from the interior walls inward 1 ft. serves as a mixing zone where room air is entrained into the supply air jet and mixes to provide thermal comfort in the occupied space. When designing UnderFloor Air Distribution (UFAD) systems or Thermal Displacement Ventilation (TDV) systems, the occupied area around the outlets may be excluded to a boundary where the total air jet from the outlet contains velocities greater than 50 feet per minute (fpm). These areas may also be known as the “clear zone”, “adjacent zone”, or “near zone”.
Any design must also include an adequate supply of Ventilation Air to the breathing zone of the space. ASHRAE 62.1-2010 defines ventilation air as “that portion of supply air that is outdoor air plus any re-circulated air that has been treated for the purpose of maintaining acceptable indoor air quality”. The breathing zone is “the region within the occupied space between planes 3 and 72 inches above the floor…”. We will discuss additional requirements for ventilation air later in this article. The primary factors to be considered when determining conditions for thermal comfort in the occupied space are: 1) Temperature, 2) Air Velocity, 3) Humidity, 4) Clothing insulation, and 5) Activity level of the occupants. All of these factors are inter-connected when determining the general occupant comfort of a space. The ideal temperature in a space (operative temperature) is where the occupant will feel neutral to their surrounding neither feeling any heat loss to the space or heat gain from the space. While the range of acceptable operative temperature may vary depending on other conditions, ASHRAE 55 requires the “Allowable Vertical Air Temperature Difference-Between Head (67”) and Ankles (4”) is limited to 5.4 F (3.0 C)”. Ideal air velocity in the space can vary with other factors but in general the goal is to keep spatial velocities less than 50 fpm during the cooling mode and less than 30 fpm during the heating mode. For many years, Titus has recommended maintaining the relative humidity level in the space between 25-60%. ASHRAE 55 does not define a lower limit and requires the dew point temperature be less than 62.2 degrees (F). Another factor affecting comfort is the clothing insulation level of the occupant. In most office environments, occupants clo level is between 0.5 and 1.1, where 0.5 would be a person wearing no socks, sandals, short sleeve shirt or blouse, and shorts or skirt. The 1.1 clo level would include long pants, socks, long sleeve shirt and dress coat or sweater. The range of operative temperature where both a 0.5 and 1.1 occupants are in the same space is very narrow. The final item of consideration for design comfort is the intended activity level of the occupant in the space. In most office environments the metabolic (met. Rate) is between 1.0 and 1.3. This includes occupants who are sedentary to casual movement about the space.
The three common methods of room air distribution used in commercial buildings in the United States are fully mixed, (e.g. overhead distribution); fully stratified (e.g. displacement ventilation); and partially mixed, (e.g. most underfloor air distribution systems). Since interior zones usually have adequate heat loads from occupants and equipment and few heat losses, the discussion for interior spaces will be cooling only. For the perimeter spaces, discussion will be how to meet the requirements for heating and cooling from the same overhead outlet. Design methods for cooling an interior zone and heating a perimeter zone vary with each method.
For fully mixed systems, the pattern of the air delivered to the space must be considered when selecting an air outlet. Ceiling diffusers typically exhibit flow in a circular (radial) or cross flow (directional) discharge air pattern. The circular pattern usually provides shorter throw, higher mixing and tends to maintain ceiling effect to low velocity before turning back on itself. This pattern is ideal for variable air volume (VAV) cooling by providing less drop and more uniform temperatures in the space. The cross flow (directional) air pattern has longer throw but with less induction may lose ceiling effect creating drafts in the occupied zone. Plenum slot diffusers typically discharge air in a directional air pattern but some are available with “spreaders” to produce a more radial discharge air pattern. Sidewall grilles equipped with vertical deflectors can be adjusted from zero degree (directional pattern) to 45 degree spread (radial pattern). So, regardless of the desired type of outlet, the air pattern can be either radial or directional to best meet the comfort requirements of the space. Proper selection for comfort can be insured by using the ADPI selection program in TEAMS.

The Titus DynaFuser installed in the
Virginia Tech Solarhouse.

Typically for perimeter applications where the same outlet is being used for both heating and cooling, a linear slot diffuser or plenum slot diffuser is employed. When a fixed air pattern diffuser is used, it is typical to supply half of the air across the ceiling for cooling and half down the glass for heating. For perimeter heating, the requirements for table 6-2 of ASHRAE Standard 62.1-2010 must be considered. The intent of table 6-2 is to insure that the ventilation air supplied to the space be delivered to the breathing zone as well. For ceiling supply of warm air with a ceiling return, the requirements for heated air are to reach a terminal air velocity of 150 feet per minute (fpm) to within 4.5 ft. of the floor. To a terminal velocity of 150 fpm or more, air is temperature independent which means the distance air will travel will be the same for isothermal air (catalog values), warm air and cool air. This means that during heating, ventilation air will be pushed down into the breathing zone with enough heat energy to meet Standard 55’s requirement for a temperature gradient of less than 5.4 degrees. In addition, the differential temperature between warm supply air and space temperature with a ceiling return must be 15 degrees or less. Thus the maximum supply air temperature for a 75 degree room would be 90 degrees. When the heating supply air temperature exceeds the 15 degree limit, the ventilation air volume for heating must be increased by 25%.
Choosing an auto-changeover diffuser like Dynafuser or EOS does not change the Standard 62.1 requirements, but will lower energy cost and improve comfort in the space. Delivering all the warm air down the glass during heating will save energy. With a fixed pattern diffuser, half of the warm air will be discharged across the ceiling and with a ceiling return can be short circuited without reaching the occupied space level. Additionally, higher comfort will be realized in the space as the heated air can be designed to deliver warm air all the way to the floor. Comfort may be increased during cooling as well as the cool air will be projected across the ceiling eliminating the potential for drafts from the jet projected down the glass with a fixed pattern diffuser.
For partially mixed air distribution systems (typically UFAD), the core area usually experiences even loading throughout the occupied area. The goal of partially mixed systems is to save energy by comfort conditioning the lower occupied level in the space and allowing the upper level of the space to stratify. Occupant comfort is achieved by delivering cool conditioned air from the plenum under the floor through swirl diffusers or rectangular shaped outlets near the occupants work area. Individuals can enhance their personal comfort by adjusting the damper at the outlet near their workspace. For common areas such as hallways and break rooms, outlets can be equipped with actuators that are controlled by a common thermostat located in the space.

UFAD Perimeter System installed in the Visteon
Village project in Van Buren Township, Michigan.

Perimeter zones for partially mixed systems create a greater challenge as the loads are dynamically changing due to outdoor solar and air temperature changes. A common method for perimeter zone control is locating a low profile fan powered terminal unit under the floor near the perimeter supplying air to linear bar grilles. The fan powered terminal can be equipped with an electric or hydronic coil. Cool plenum air can be supplied to the outlets when cooling is required and the coil can be employed to warm the air as required during heating conditions. The design challenge is selecting outlets that will limit the throw of the air pattern so that air will not bounce off the ceiling and create drafts in the adjacent occupied area.
Energy to operate the fan terminals can be eliminated and higher comfort can be achieved on the perimeter by using the TAF-L perimeter distribution outlets. With a 6” wide custom design TAF-L bar grille located along the perimeter of the space, the modular 4’ long TAF-L-V (cooling), can be attached to provide up to 225 cfm (at 0.07” plenum pressure) per 4’ unit of cooling. The TAF-L-V damper is controlled by a space thermostat to provide cooling as required. The special arrangement of bars in the grille is designed to limit the throw from the outlet during cooling. The 4’ long TAF-L-W or TAF-L-E heating module can be attached to the TAF-L grille to supply up to 3000 Btu heat to the perimeter. The heating units operate by combining the cool convection currents from the glass with the warm currents on the floor. The mixture is induced through the heat exchanger with warm air being discharged through the grille and up the glass. Space temperature is controlled by a room thermostat controlling the water flow or electric current flow to the electric heating element. The modular design allows the system to be custom designed for use in multiple climate regions.
Fully stratified design (typically TDV), conditions a space by discharging cool supply air through an outlet located at floor level near or in a wall or may be centrally located in the open space. Low velocity air (< 80 fpm) is discharged horizontally across the floor. Air moves with little mixing across the floor until it contacts a heat source such as an occupant or piece of warm equipment in the space. Cool air will mix with the radiant heat from the source and stratify toward the ceiling. The return is usually located at or near the ceiling. The area between the outlet and where the air speed reaches 40 fpm is the “clear zone” and should not be included in the occupied area. Titus provides units with adjustable air patterns so the clear zone can be controlled to meet project requirements for space occupancy. ASHRAE Standard 62.1 table 6-2 provides a 20% bonus for TDV systems. This means that ventilation air can be reduced by 20% or the 20% can be used toward the 30% required for an additional LEED IEQ credit 2.
While TDV systems typically require a separate system for heating, Titus has introduced the Plexicon heating/cooling diffuser. A standard rectangular outlet is located near or mounted in a wall discharging cool air from the upper chamber. When heating is required, an internal baffle is moved to change the flow of air from the upper chamber to the lower chamber where it flows through a linear bar grille to satisfy heating requirements.
Regardless of which type of system you are using on your project, studies have shown that occupants who are comfortable are more productive. Designing for comfort, keeps paying back dividends forever.
Jim Aswegan - Chief Engineer

Friday, February 17, 2012

What is the Adjacent Zone?

ASHRAE Standards 55 and 62.1 summarizes ASHRAE’s standards for thermal comfort and minimum ventilation and explains how to apply air outlets to comply with each of these standards. Even though the article also references the definition of the ‘adjacent zone’ that exists in close proximity to Thermal Displacement Ventilation (TDV) diffusers, it’s also important to understand how displacement ventilation products can differ with respect to occupant comfort.

To briefly recap, displacement diffusers used in fully-stratified systems deliver low velocity cooling directly to the occupied zone. Since the supply air is cooler than the room air, it cascades down the face of the diffuser and travels across the floor in a thin layer generally no more than 4 inches deep.

Adjacent Zone

The ‘adjacent zone’ is defined as any portion of the room where discharge velocities exceed 40 fpm. This area is not recommended for stationary occupants who would likely feel a chill around their ankles.

Although displacement ventilation diffusers are available from most manufacturers in a myriad of shapes and sizes, there are really two basic types of designs being manufactured:

• Fixed Air Pattern
• Adjustable Air Pattern

Although all displacement diffusers include a perforated face plate and a rear supply plenum, fixed air pattern diffusers are characterized by a perforated central baffle. The purpose of the central baffle is to further slow and spread the supply air evenly over the face plate. Diffusers of this design are less costly to produce but they must be selected very carefully to ensure that the symmetrical adjacent zones they create will not result in thermal discomfort for stationary occupants.

Fixed air pattern diffusers also lack versatility in situations where spaces are being reconfigured or re-purposed. Rather than used a fixed perforated central baffle, adjustable air pattern diffusers include a baffle fitted with air pattern controllers. These sturdy steel pattern controllers are easy to remove and reset and do not raise concerns about plastic materials in the air stream. Although these pattern controllers ship out of the factory in a default arrangement to create a standard symmetrical discharge pattern, they provide adjustability for enhanced versatility and improved occupant comfort.

Left: Fixed air pattern displacement diffusers in standard operation. Right: Adjustable air pattern controllers allow for adjustability of the adjacent zone for maximum occupant comfort.

 Rather than settle for a fixed discharge pattern, Titus displacement ventilation products can be easily adjusted to direct air away from occupants and areas that may result in comfort issues (now or in the future). Titus displacement ventilation products include adjustable pattern controllers so that you can control your ‘adjacent zone’.

Randy Zimmerman – Titus Chief Engineer

Thursday, February 9, 2012

RP-1335 Test Effects of Diffuser Performance

ASHRAE Study 1335 on the effects of typical inlet conditions on ceiling diffusers and their performance began in 2009 at the University of Nevada, Las Vegas (UNLV) laboratories.

The goal of this ASHRAE study was to test and determine the impact on manufactures catalog performance data for ceiling diffusers created by typical field-installed inlet conditions.

ASHRAE Standard 70-2006, is the current “Method of Testing the Performance of Air Outlets and Air Inlets” used by manufactures to obtain catalog performance data. Standard 70 requires diffusers to be tested with a minimum of 3 diameter equivalents of straight duct ahead of the inlet with even flow throughout the duct. 

So the question is, “What happens when diffusers are mounted in buildings and how much variation in performance will we see with typical field installed conditions?”

When building air distribution systems, designers and system installers require accurate quantitative information on how the installed system will perform to achieve optimum efficiency and comfort, and they rely on performance data from manufactures catalogs. But catalog performance reflects perfect inlet conditions. If field installations adjustments are required to the manufactures data due to typical field installation procedures, the extent of these adjustments to the performance data of throw, pressure loss, and sound will be shown in the results of this study.

This study incorporated the performance data of multiple installations using six different types of typical ceiling diffusers. The data in this report compares the performance results of these various field installations to the performance data collected per ASHRAE 70-2006. All diffusers were first tested per ASHRAE 70-2006 to determine the base performance data and then various modifications to the installation were conducted.

Full scale testing was done in the UNLV laboratory with diffusers mounted in various inlet conditions over a large range of flow rates with various duct approach angles, as well as:
  • hard duct vs. flex duct
  • various straight duct heights above the diffuser
  • using an elbow attached to flexible duct
  • close coupling duct installations
In addition, the effect of various dampers connected directly to the diffusers with multiple inlet conditions was also studied.



Installations with flexible duct connections and tight bends affect diffuser performance significantly due to uneven air flow through the diffusers.


The throw from the diffuser is affected with elbows directly attached to the diffuser inlets. The throw from the a diffuser with an 90-degree elbow directly attached to diffusers of all types tested showed an average throw increase in the (forward) direction and a decrease in the (backward) direction. Dampers, depending upon the design, can reduce this asymmetry problem a significant amount. These dampers, however, also add a significant sound increase.


The data shows that a flexible duct elbow to a diffuser as compared to an all metal elbow has a greater pressure loss. This data also shows that the higher the pressure loss of the diffuser, the less the effect of inlet conditions and damper conditions. The average pressure loss increase due to dampers on the face of diffusers is about 50%. This data shows that dampers directly mounted on the diffuser inlet should be avoided.


A very interesting group of information deals with the amount of straight duct required after an elbow to obtain the same performance data as shown in manufactures catalogs.  This data shows that three diameters of straight duct down-stream from the elbow typically resulted in the same sound levels as cataloged and tested by ASHRAE 70-2006. However, elbows directly connected to diffusers typically increase the NC. Dampers also contribute to sound increases. Dampers in the full open condition can add to NC levels. Flex duct elbows averaged higher than rigid duct. The sound data from plaque diffusers was interesting. This data showed that with a fixed diffuser face free area as we have with plaque diffusers, that the increase in sound is not as critical as compared to diffusers of this type with inlets having a greater free area than the diffuser face.       

Close coupling is when diffusers are connected directly below a ducting system. The length of duct between the supply duct and the diffuser was studied. The results were very similar to those seen with elbows to diffusers. Another variable was also seen which deals with the main duct velocity and the pressure loss of the diffuser as related to the velocity pressure in the main supply duct. In general, as the main duct velocity increased greater than the diffuser inlet velocity, the sound from the diffuser increased. Once again, open dampers also increased the sound.


In general, the data show that dampers should be mounted as far up stream form the diffusers as possible. Dampers directly mounted on diffusers cause high velocity air streams on the diffuser cones and can cause significant sound increases.


As part of the report, a set of tables were developed to easily predict how various installation configurations will affect diffuser performance. A set of three reports were given at the Winter ASHRAE 2012 meetings in Chicago, and copies of this data are now available through ASHRAE.

Leon Kloostra - Titus Senior Chief Engineer