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

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

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 D_e, 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:

D_e = SQRT [(4 x A) / (144 x π)]

where:

A = cross sectional area of duct (in²)

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

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

Now

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

where:

C_o = Speed of sound in air (use 1128 fps)

f = Octave band center frequency (Hz)

D_e = Equivalent diameter of the duct (ft)

 

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

2^nd Octave band (125 Hz) = 5 dB

3^rd Octave band (250 Hz) = 2 dB

4^th Octave band (500 Hz) = 1 dB

5^th Octave band (1000 Hz) = 0 dB

6^th Octave band (2000 Hz) = 0 dB

7^th 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

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

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