titus for vav

engineering guidelines - terminal units

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Engineering Guidelines - Terminals

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overviewEngineering Guidelines Overview ..................................................................................................................................................B4

Table of Contents

Types of Terminals .........................................................................................................................................................................B5Single Duct .............................................................................................................................................................................B5Dual Duct, Non-Mixing ..........................................................................................................................................................B5Dual Duct, Mixing ..................................................................................................................................................................B5Single Duct, with Heating Coil ...............................................................................................................................................B5Fan Powered, Parallel Type (Variable Volume) .......................................................................................................................B5Fan Powered, Series Type (Constant Volume) ........................................................................................................................B6Low Temperature Fan Terminals ............................................................................................................................................B6Fan Powered, Low Profile ......................................................................................................................................................B6Fan Powered, Access Floor Profile (Constant Volume) ..........................................................................................................B6

Types of Controls ...........................................................................................................................................................................B7Reaction to Duct Pressure Controls .......................................................................................................................................B7

Control Operation in Terminals ......................................................................................................................................................B9Damper Operation ..................................................................................................................................................................B9Direct Acting/Reverse Acting Pneumatic Thermostat Action ...............................................................................................B9Direct Reset/Reverse Reset Pneumatic Velocity Controller Action ........................................................................................B9Pneumatic Thermostat-Controller Combinations .................................................................................................................B10Actuator Terminology ..........................................................................................................................................................B10Pneumatic Control/Actuator Combinations .........................................................................................................................B10

Velocity Controller Operation .......................................................................................................................................................B11Definitions of Terms .............................................................................................................................................................B11Thermostat Sensitivity ........................................................................................................................................................B11Hysteresis.............................................................................................................................................................................B11Pneumatic Feedback ...........................................................................................................................................................B12

Fan Terminal Flow Control ...........................................................................................................................................................B12Series Fan Shift ....................................................................................................................................................................B12Mechanical Trimming ..........................................................................................................................................................B12Voltage Adjustment ..............................................................................................................................................................B13

Fan Speed Control ........................................................................................................................................................................B13Catalog Fan Curves ..............................................................................................................................................................B14

ECM Motors - Fan Powered Terminals ........................................................................................................................................B15Pressure Independent - Energy Efficient Analog Speed Settings ........................................................................................B15

Direct Digital Control ..................................................................................................................................................................B16Applying Computers to Control ............................................................................................................................................B16Direct Digital Control ...........................................................................................................................................................B16Advantages of DDC ..............................................................................................................................................................B16DDC Distributed Processing .................................................................................................................................................B18

Sizing Basic Terminals from Capacity Tables ...............................................................................................................................B18Certified Air Terminals .........................................................................................................................................................B18Sizing Single Duct Terminals ...............................................................................................................................................B18Sizing Parallel Fan Powered Terminals ................................................................................................................................B19

terminals, controls & accessories


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ENGINEERING GUIDELINES

Table of Contents (continued)

Acoustical Applications & Factors ...............................................................................................................................................B28Noise Criteria (NC) ...............................................................................................................................................................B28Room Criteria (RC) ...............................................................................................................................................................B30Air Terminal Sound Issues ...................................................................................................................................................B32AHRI Standard 885 ..............................................................................................................................................................B33Environmental Adjustment Factor .......................................................................................................................................B33Discharge Sound Power Levels ............................................................................................................................................B34Acceptable Total Sound in a Space ......................................................................................................................................B35Maximum Sound Power Levels for Manufacturers’ Data ....................................................................................................B37Desired Room Sound Pressure Levels ..................................................................................................................................B37Radiated Sound Power Level Specifications ........................................................................................................................B38Discharge Sound Power Level Specifications ......................................................................................................................B38Diffuser Specifications .........................................................................................................................................................B39Determining Compliance to a Specification .........................................................................................................................B40Standard Attenuations .........................................................................................................................................................B41

acoustical applications & factors

Sizing Series Fan Powered Terminals...................................................................................................................................B20

Typical Problems ..........................................................................................................................................................................B21Oversizing Terminal ..............................................................................................................................................................B21Capacity Concentrated in Too Few Terminals ......................................................................................................................B21Insufficient Space ................................................................................................................................................................B21Improper Discharge Conditions ...........................................................................................................................................B21Improper Inlet Conditions ....................................................................................................................................................B21Incompatibility with Power Source ......................................................................................................................................B22Excessive Air Temperature Rise and Air Change Effectiveness ...........................................................................................B22Excessive Air Leakage ..........................................................................................................................................................B22Improper Support Of Terminal ..............................................................................................................................................B22Wrong Type of Insulation .....................................................................................................................................................B22Non-Compliance with Local Codes ......................................................................................................................................B23Installation Techniques-Duct Connections ..........................................................................................................................B23

Some Basic Concepts-Pressure Measurement ............................................................................................................................B25The Fan Laws ...............................................................................................................................................................................B26Equations and Definitions ............................................................................................................................................................B27

references

glossary

References ...................................................................................................................................................................................B43

Glossary .......................................................................................................................................................................................B44


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The selection and performance data contained in this catalog are the result of extensive studies conducted in the Titus engineering laboratories under professional engineering guidance, with adherence to sound engineering applications. They are intended to be aids to heating and air conditioning engineers and designers with skill and knowledge in the art of air distribution. The data have been obtained in accordance with the principles outlined within the American Society of Heating, Refrigerating and

Air Conditioning Engineers (ASHRAE) Standard 70, Standard 113, Standard 130 and AHRI 880. Although Titus has no control over the system, design and application of these products, a function which rightfully belongs to the designer, this data accurately represents the product performance based on the results of laboratory tests. Furthermore, the recommended methods of applying this information have been shown by field experience to result in optimum space air distribution.

Overview


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TERMINAL CONTROLS AND ACCESSORIES

TYPES OF TERMINALSSINGLE DUCTThis basic terminal consists of casing, a damper, a damper actuator and associated controls. In response to control signals from a thermostat or other source, the terminal varies the airflow through a single duct handling hot or cold air. In some applications the same terminal is used for both heating and cooling; a dual

function thermostat, together with the necessary change-over circuitry, makes this possible. Controls can be pneumatic, electric, analog electronic or direct digital electronic. Accessories such as round outlets, multiple outlets and sound attenuators may be added. The single duct terminal is most often used in an interior zone of the building, for cooling only.

DUAL DUCT, NON-MIXINGEssentially the same as two single duct terminals side-by-side, this terminal modulates the flow of hot and cold air in two separate streams supplied by a dual duct central air handling unit. Because there is no provision for mixing the two airstreams, this terminal should not be used for simultaneous heating and cooling,

which would result in stratification in the discharge duct. (When stratification occurs, the several outlets served by the terminal may deliver air at noticeably different temperatures.) The non-mixing, dual duct terminal is best used in an exterior zone, in which zero-to-low airflow can be tolerated as the temperature requirement shifts from cooling to heating.

DUAL DUCT, MIXINGHere the terminal is designed specifically for mixing hot (or tempered ventilation) and cold air in any proportion. When equipped with pneumatic controls, there is a velocity sensor in the hot air inlet, but none in the cold air inlet. A velocity sensor at the discharge measures the total flow of air and sends the signal to the cold air controller. In the mixing cycle, the

hot airflow changes first, and a change in cold airflow follows to maintain a constant total (mixed) volume. When equipped with DDC controls by Titus, both hot and cold inlets have velocity sensors, with the summation of flows computed by the microprocessor. No discharge velocity sensor is used. This dual duct terminal is often used in an exterior zone of a building or to ensure ventilation rates.

SINGLE DUCT, WITH HEATING COILThis is the single duct terminal described above, with a heating coil added. The coil may be of either the hot water or the electric type. The hot water coil is usually modulated by a proportioning valve controlled by the same thermostat that controls the terminal. Control for the electric coil is either 100% on-off or in steps of capacity, energized by contactors in response to the room thermostat. The single duct terminal with heating coil

is most often used in an exterior zone with moderate heating requirements. Since the terminal normally handles its minimum cfm in the heating mode, a dual minimum cfm or “flip-flop” control can be added for increased heating airflow. Separate minimum cfm setpoints are standard with most DDC controls (available optionally on most other control types) and should be considered in design. A higher minimum cfm in heating mode will improve overhead air distribution performance.

FAN POWERED, PARALLEL TYPE (VARIABLE VOLUME)In this terminal a fan is added to recirculate plenum air, for heating only. The heating cycle occurs generally when the primary air is off or at minimum flow. Heat is picked up as the recirculated air is drawn from the ceiling space and the fan motor. Additional heat can be provided by a hot water or electric coil on the

terminal. Because the fan handles only the heating airflow, which is usually less than that for cooling, the fan can be sized smaller than in the series flow type terminal (see below). During the cooling cycle, the fan is off and cool primary air is supplied from the central system. A backdraft damper prevents reverse flow through the fan. The flow of the primary air is regulated by variable air volume controls. Used in exterior zones.

Figure 47. Elevation - Single Duct

Figure 48. Plan View - Dual Duct, Non-Mixing

Figure 49. Plan View - Duct Duct, Mixing

Figure 50. Elevation - Single Duct, with Heating Coil

Figure 51. Plan View - Fan Powered, Parallel Type (Variable Volume)

Terminals, Controls and Accessories


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Types of Terminals (continued)

FAN POWERED, SERIES TYPE (CONSTANT VOLUME)The fan runs continuously, fed by a mixture of primary and plenum air. The more primary air is forced in, the less plenum air is drawn in. The result is variable volume from the central system, constant volume (and sound) to the room. Because the central system need only deliver air as far

as the fan, the inlet static pressure can be lower than in the parallel flow terminal (above). The fan, however, is sized to handle the total airflow. These are often used in applications where constant background sound and continuous airflow are desired.

Figure 52. Plan View - Fan Powered, Series Type (Constant Volume)

LOW TEMPERATURE FAN TERMINALSThe fan terminal, with its inherent mixing, is well suited to handle the very cold air delivered by systems designed for air much colder than with conventional 55°F supply systems. In order to use standard diffusers, the primary air must be raised to

a conventional supply temperature before it enters the room. A commonly utilized solution is to mix it with recirculated air with a fan powered terminal. The most common application uses a Series Flow unit, but many applications have been utilized with Parallel units with a constant running fan.

FAN POWERED, LOW PROFILEThis series or parallel type terminal has a vertical dimension of only 10.5” for all sizes, to minimize the depth of ceiling space required. Notice in the diagram at the right that the recirculating fan is laid flat on its side, shaft vertical. In localities where building heights are limited, the low profile terminal saves enough space to allow extra floors to be included

in a high-rise structure. Ceiling space can be as little as 12” to 14” deep. The low profile terminal is also useful in buildings constructed with precast concrete channel floors. The terminal can fit into the channel space with no extra depth required (Series type shown).

Figure 53. Plan View - Fan Powered, Low Profile

FAN POWERED, ACCESS FLOOR PROFILE (CONSTANT VOLUME)This series type terminal is designed to fit around the pedestal support grid of access, or raised, floor systems. In a typical access floor the grid is 24” x 24”.

The terminal can fit into the floor plenum without any modifications to the pedestal system.

Figure 54. Plan View - Fan Powered, Access Floor Profile (Constant Volume)


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REACTION TO DUCT PRESSURE CONTROLSPRESSURE INDEPENDENTWith this type of control the terminal maintains the flow rate required to handle the heating or cooling load, regardless of system pressure fluctuations. It is the best choice where the system pressure will vary extensively and where precise control is essential. Key components in pressure independent control are the velocity sensor, which furnishes a continuous reading of the air velocity through the terminal, and the velocity controller, which processes this information along with signals from the thermostat. In the chart (Figure 55), vertical lines AB and EF represent minimum and maximum cfm settings which are adjustable at the controller. Line CD represents any cfm setting maintained by the controller in response to the thermostat. The damper will open and close as needed to hold the cfm constant up and down this vertical line for the full range of pressure drops shown. Notice that the vertical cfm lines are cut off by the diagonal line AE, which represents the pressure drop from inlet to outlet with the damper wide open. This is the minimum DP shown in our data.

PRESSURE DEPENDENTA terminal with this type of control is designed for those applications where neither pressure independence nor cfm limit regulation is required. An example is a variable volume makeup air supply in which the downstream duct pressure is held constant by other controls. The terminal consists essentially of a casing, a damper and a damper actuator. There is no controller and no velocity sensor; the damper moves in direct response to the thermostat or other signal input. The line AB (Figure 56) shows the typical performance characteristic. It represents a given damper setting, with the flow rate varying as the square root of the static pressure drop through the terminal. This, of course, is typical of any damper or fixed orifice. Lines CD and EF represent random additional settings as the damper opens to the full open position line GH. Line GH is the minimum pressure loss of the assembly.

Most of the control types shown here have certain principle elements in common:

ROOM THERMOSTAT OR SENSORThe thermostat contains not only a temperature sensing element, but also a means of changing the setpoint. The room sensor used with the direct digital control system is simply an electronic temperature sensor; setpoint changes are handled along with other signal processing in the digital controller.

VELOCITY SENSORMounted in the inlet of the terminal, this device senses air velocity, which can easily be converted to airflow rate. The sensor’s signal provides feedback to monitor and directs the operation of the controller and damper actuator.

CONTROLLERCommands from the thermostat or room sensor, together with feedback from the velocity sensor, are processed in the controller to regulate the damper actuator. Operation is pressure independent.

DAMPER ACTUATORThe damper actuator opens and closes the damper to change the airflow, or to hold it constant, as dictated by the controller.

Figure 55. Pneumatic Pressure Independent

Figure 56. Pneumatic Pressure Dependent

Note: Excessive airflow may lead to excessive noise. Pressure independent control has less opportunity for variable (and unwanted) sounds in the occupied spaces.

TYPES OF CONTROLS

Minumum SP, W

ide Open D

amper

1000

1000

G

Air Flow, cfm

1.00

Pres

sure

Dro

p in

w.g

.

0.06

0.04

0.02

0.01100

C

0.20

0.40

0.60

0.10 A

E

200

6.00

4.00

2.00

0.01100 200

Setting

Air Flow, cfm

#3

SettingOpen

Damper

300 400

#2SettingDamper

SettingDamper

#1

600500 800700

H

F

Damper

300 400

D

B

600500 800700

Maximum

A

SettingB

Minimumcfm

2.00

Pres

sure

Dro

p in

w.g

.

0.06

0.08

0.04

0.02

0.10

0.80

0.40

0.20

0.60

1.00

6.00

4.00

C

E

Setting

Variable

D

cfmSettingcfm

F

Minumum SP, Wide O

pen Damper

1000

1000

G

Air Flow, cfm

1.00

Pre

ssur

e D

rop

in w

.g.

0.06

0.04

0.02

0.01100

C

0.20

0.400.60

0.10 A

E

200

6.00

4.00

2.00

0.01100 200

Setting

Air Flow, cfm

#3

SettingOpen

Damper

300 400

#2SettingDamper

SettingDamper

#1

600500 800700

H

F

Damper

300 400

D

B

600500 800700

Maximum

A

Setting

B

Minimumcfm

2.00

Pre

ssur

e D

rop

in w

.g.

0.060.08

0.04

0.02

0.10

0.80

0.40

0.20

0.60

1.00

6.00

4.00

C

E

Setting

Variable

D

cfmSettingcfm

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PNEUMATIC SYSTEMSIn a pneumatic control system, the various components are powered by compressed air, usually at 15-25 psi, from a central system. The thermostat receives air at full pressure directly from the main air supply. In response to room temperature, the air pressure is modulated to the controller, which regulates the damper actuator. The sensor and controller compensate for changes in duct pressure so that operation is pressure independent.

ELECTRIC SYSTEMS - FIGURE 58AElectric controls operate at low voltage, usually 24 VAC, supplied by a transformer which is often built into the control box of the terminal. The room thermostat has single-pole-double-throw contacts so that (in the cooling mode) a rise in temperature drives the damper actuator in the opening direction; a fall in temperature reverses the actuator. Since the electric system has no velocity sensor and no controller, there is no compensation for duct pressure fluctuations. Operation of the terminal is pressure dependent, the thermostat and room response time are typically much less than the actuator response time, and excessive room temperature variations are a likely result.

ANALOG ELECTRONIC SYSTEMS - FIGURE 58BLike the electric controls, analog electronic controls operate at low voltage, usually 24 VAC, supplied by a transformer which is often built into the control box of the terminal. These controls, however, also include a velocity sensor of either the thermistor type, or pneumatic velocity sensor with electronic transducer, together with an electronic velocity controller that is pressure independent. The electronic thermostat can control both cooling and heating operations. Because of the pressure independent operation and integrated thermostat, excellent room temperature control can be achieved.

DIRECT DIGITAL ELECTRONIC SYSTEMS - FIGURE 58CHere again the power source is a low voltage supply. Signals from a pneumatic or electronic velocity sensor, together with signals from the room temperature sensor, are converted to digital impulses in the controller, which is a specialized microcomputer. The controller not only performs the reset and pressure independent volume control functions, but it also can be adjusted and programmed either locally or remotely for multiple control strategies, including scheduling. In addition, it can link to other controllers and interface with security, lighting, and other equipment. Control can be centralized in one computer.

Figure 57. Pneumatic System

Figure 58A. Electric Pressure Dependent System

Figure 58B. Electric Pressure Independent System

Figure 58C. Electric Pressure Independent System

Types of Terminals (continued)

T’Stat

Main Air

Pneum.

Sensor

Velocity

Pneumatic

Actuator

Pneumatic

Controller

Damper

Pneum.

Damper

LineTransformer

T’Stat

Electric

Actuator

Electric

Damper

Damper

Analog Electronic

Transformer-Usually 24 VAC

T’Stat

Velocity

Sensor

Line

Analog VAV

Controller with

Velocity Transducer

Secondary

Damper

Electric

Damper

Actuator

Velocity Transducer

Transformer-Usually 24 VAC

Secondary

Direct Digital Electronic

Digital VAV

Controller with

Room

Line

Sensor

Sensor

VelocityDamper

Damper

Actuator

Electric


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DAMPER OPERATIONLinearity (Figure 61) is the ideal characteristic for most damper applications. How nearly linear the operation is depends upon the percentage of the overall system pressure drop contributed by the wide open damper. Pressure independent control operations eliminate the effect of nonlinear dampers, but simulate the effect of a true linear damper to the system. For a linear damper characteristic, the damper is sized to contribute about 10% of the overall system resistance. Also (Figure 62), actuator torque must be sufficient to close the damper under all design conditions. In Titus terminals, the torque is always more than adequate.

DIRECT ACTING/REVERSE ACTING PNEUMATIC THERMOSTAT ACTION In the direct acting pneumatic thermostat (Figure 63), a room temperature increase causes a corresponding increase in thermostat output. In the reverse acting thermostat (Figure 64), the sequence is the opposite. Because of these characteristics, direct acting thermostats are often used for cooling, reverse acting for heating. (With electronic systems, this term has no application.)

DIRECT RESET/REVERSE RESET PNEUMATIC VELOCITY CONTROLLER ACTIONIn the direct reset pneumatic velocity controller (Figure 65), an increase in thermostat output pressure causes a corresponding increase in controller cfm setting. The damper will open and close to maintain this cfm when duct pressures change. In the reverse reset controller (Figure 66) the same action results from a decrease in controller cfm setting.

Figure 61. Linear Damper Operation Figure 62. Damper Torque Requirement

Figure 63. Direct Acting Thermostat Action

Figure 64. Reverse Acting Thermostat Action

Figure 65. Direct Reset Pneumatic Velocity Controller

Figure 66. Reverse Reset Controller

CONTROL OPERATION IN TERMINALS

Control Operation in Terminals

Damper

Damper Opening, %

cfm%

Characteristic

Sized for Damper

Linear

Oversized100

Face Area x Total Pressure

Required

cfm%

Torque

Open Closed

TorqueMaximum

100

(Cooling)cfm

Room Temperature Increase

ThermostatActingDirect

%

100


(Cooling)cfm

%

100

ThermostatActingReverse

(Cooling)cfm%


ResetThermostat

Min

Direct100 Max


cfm(Cooling)

%

100 Max

ResetThermostat

Min

Direct


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PNEUMATIC THERMOSTAT-CONTROLLER COMBINATIONSFor systems supplying cold air when a direct acting pneumatic thermostat signals a direct acting controller (Figure 67), an increase in room temperature produces an increase in cfm setting. A reverse acting thermostat with a reverse reset controller produces the same result. A direct acting thermostat with a reverse reset controller or a reverse acting thermostat with a direct reset controller (Figure 68) will produce a decrease in cfm as the room temperature increases. With warm supply air, the logic is reversed.

ACTUATOR TERMINOLOGY Pneumatic actuators have an internal spring which is overcome by control air pressure. When air pressure is less than the spring tension, the actuator will retract. Depending on how it is connected to a damper, the damper may open or close on increase in control signal. Electronic actuators, however, are typically “fail stopped” unless they have a return spring which is activated by a loss of control signal. These are several times the cost of “fail stopped” actuators. When normally open or normally closed actuators are specified in an electronic control project, the requirement is most often in error.

NORMALLY OPENThis describes a pneumatic operator which is configured so that on loss of air pressure the damper in the unit will open fully. These applications are typically ones where all like units are desired to be open for control purposes such as smoke removal or to prevent excessive pressure on system start-up.

NORMALLY CLOSEDWhen air pressure is removed, the actuator will cause the damper in the unit to go fully closed. This is typically specified when an area is to be isolated.

PNEUMATIC CONTROL/ACTUATOR COMBINATIONSControllers and actuators work in concert to control space temperatures. With pneumatic controls the most common combinations are Direct Acting Normally Open (DANO) and Reverse Acting Normally Closed (RANC). With most pneumatic controls special controllers are used for direct and reverse acting and any combinations other than DANO or RANC require extra components and increase air consumption. (With the Titus II controller, no extra components are required as the unit is switchable.)

Figure 67. DA Pneumatic Thermostat Signaling DA Controller Combination

Figure 68. RA Thermostat with Reverse Reset Controller or RA Thermostat with

Direct Reset Controller Combination

Control Operation in Terminals (continued)


cfmMin

(HOT AIR)

(COLD AIR)

DA ControllerRA Thermostat

DA Controlleror

G

%

100

DA Thermostat

Max

RA Controller(COLD AIR)


cfmMin

(HOT AIR)RA Thermostat

RA ControllerDA Thermostat

%

H 100 Max

or


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B11


DEFINITIONS OF TERMSThe controller setpoint is the cfm setting that the control system is calling for at any given moment. At that setpoint the damper opening may vary widely to compensate for any duct pressure changes reported by the inlet sensor, and thus hold the cfm constant.

With pneumatic systems, the setpoint, 11 psi in the example (Figure 69), can be reset by the action of the thermostat anywhere between the maximum and minimum cfm settings of the controller. The corresponding thermostat output pressures are called the start and stop points. The range of possible setpoints between the start and stop points is called the reset span, 8 to 13 psi in the example shown here.

The thermostat may also control an auxiliary piece of equipment, such as a proportioning valve on a hot water coil, shown here modulating over a range of 3 to 8 psi, in sequence with the reset span of the controller. The overall range over which the thermostat controls these devices is its proportional band or total throttling range, 3 to 13 psi in this example.

THERMOSTAT SENSITIVITY This is the change in output signal caused by a change in room temperature. This rating (Figure 70) is usually 1°F = 2.5 psi for pneumatic systems. Electronic systems have a wide variance in output responses.

HYSTERESISThis is the failure of an object to return to its original position after a force has moved or deflected it. For example, in some velocity controllers (Figure 71) the cfm setting increases along the lower curved line and decreases along the upper curved line. At the setpoint, the cfm may be either A or B.

Figure 70. Thermostat Sensitivity Example

Figure 71. Hysteresis Example

Figure 69. Set Point Example

VELOCITY CONTROLLER OPERATION

Velocity Controller Operation

13

Room Temperature

Point

Set

73 75

8

77

Thermostat Output, psi

PointSet

A

8

Min

cfm

B

13

Max

Max. gpm

% M

ax. F

low

0

100

(Cooling)

Thermostat Output, psi3 8

Modulation

WaterValve

Hot

Min. cfm

1311

SetPoint

(Cooling)

of Controller

of ThermostatTotal Throttling Range

(DDES)

PointStart

PointStop

Reset Span Max. cfm


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PNEUMATIC FEEDBACK Signals from the thermostat determine the cfm setpoint of the controller. The duct velocity acting on the velocity sensor forms a feedback (closed) loop (Figure 72) that allows the controller to monitor the airflow resulting from its settings and make corrections continuously. This is a form of closed loop control and is used on both pneumatic and electronic pressure independent systems.

In the Titus ll pneumatic controller there is also an internal feedback loop that works in conjunction with a positive positioning reset mechanism to eliminate hysteresis (Figure 71, page B11).

FAN TERMINAL FLOW CONTROLEngineers designing air systems try to match the airflow capacity of fan powered terminals to the needs of the space. Exact matches are rare, however. The design may not allow an exact match, a product other than the one which is the subject of the design might be selected, or system balancing might require a different airflow to meet field conditions. The two commonly used methods of trimming fan airflow are:

• Mechanical Trimming• Voltage Adjustment

SERIES FAN SHIFTWith Series fan terminals, the fan output is intended to remain constant over a range of primary inlet damper flow rates. With proper design, this is normally so. With improper design, or with additional inlet attenuators added to a terminal, the fan may see a different external pressure when in full induction mode than when in full cooling. This results in a variation in the quantity of air delivered to the space, or “Fan Shift.” The consequences of fan shift depend on individual zone characteristics and building design. If diffusers are selected such that they may add background masking sound at design flow, variations in flow may be an annoyance to the occupants. If a designed ventilation rate is assumed, this may vary if fan shift happens. (Titus terminals are designed to minimize fan shift.)

MECHANICAL TRIMMINGMechanical trimming involves the use of a mechanical device, such as a damper, to adjust the fan airflow to meet the design requirements. Typically, these are used in conjunction with a multi-tap motor to provide a greater operating range and keep the energy consumption and sound levels as low as possible. Mechanical trimming offers a lower first cost versus a voltage adjustment, but at increased operating costs and increased sound. Multi-tap motors are not always effective in changing flow.

In operation, the mechanical device will raise the static pressure the fan operates against by either restricting the free area downstream of the fan or restricting the free flow of air drawn into the fan. A forward curved fan riding the fan curve will reduce airflow accordingly (Figure 73).

Although the rpm of the fan will increase, less work will be performed. This will result in a reduction of the amp draw of the fan motor. Since voltage remains constant, the overall power consumption of the fan is reduced. The power reduction from mechanical trimming is less, however, than the power reduction from voltage adjustment. When mechanical trimming is used, the sound levels of the fan terminal will increase. When the dampering occurs downstream of the fan, the velocity of the discharge air must rise, thereby increasing the discharge sound power levels. Additional sound contributions are made by the fan. The increased rpm of the fan results in greater tip speed. This occurs with either dampering method, raising the level of both the radiated and discharge sound.

Figure 72. Pneumatics Feedback

Figure 73. Forward Curved Fan Performance Curve

Velocity Controller Operation (continued)

Thermostat

T

Controller

ResetPositioningPositive

Transmitter

Actuator Loop

Velocity SensorCenter AveragingMulti-Point

Damper FeedbackCompletesVelocityDuct

0

SP

100%

cfm 100%

Operating

HorsepowerBrake

EfficiencyStatic

Static Pressure

Point

System

SE

BHP

100%


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VOLTAGE ADJUSTMENTVoltage adjustment of fan powered terminals typically involves the use of a silicon controlled rectifier (SCR). An SCR uses a triac to phase proportion (chop) the electrical sine wave.

In effect, the SCR switches power off 120 times a second on a 60 Hertz cycle. This reduces the voltage to the motor, slowing its speed. In operation, the SCR responds to the current but controls voltage. Thus, while an SCR’s triac may be energized at zero current, the current sine wave generally lags the voltage sine wave with an induction motor. This results in the idealized voltage sine wave (Figure 74). As the SCR is used to further reduce fan speed, the true RMS value of the voltage is reduced.

As voltage to the motor is reduced, the motor tries to compensate and the motor’s amp draw rises slightly. The amperes will continue to increase until 50% of the current sine wave is phase proportioned. After this point, the amp draw will decrease. The increased amp draw is small relative to the reduction in voltage. As a result, comparing power consumption of the mechanical trimming method with the voltage adjustment method is analogous to comparing the power consumption of inlet guide vanes on central air handlers with speed inverters (Figures 75 and 76).

FAN SPEED CONTROLThe rpm of the motor is reduced by the SCR, lowering the tip speed of the fan. Since the free area downstream of the fan is not reduced, the velocity either meets design conditions or is lowered if the airflow is reduced below design for balancing purposes. There is no increase in sound from air disturbances.

Figure 74. Idealized Voltage Sine Wave Resulting from an SCR

Figure 75. Watt Reduction Versus cfm

A Note on Nameplate RatingsThe amp draw can increase above the nameplate rating of the motor! The motor’s nameplate specifies the amp draw for one set of design conditions. Since the voltage to the motor is reduced, the nameplate rating is no longer applicable. If proper care is taken in the design, specification and selection of the motor by the terminal manufacturer, the increased amp draw will pose absolutely no problem in operation or longevity. Thousands of fan powered terminals shipped with SCRs over the years serve as confirmation.

Titus accounts for the increased amp draw in the specification and selection of motors used for fan powered terminals. As a result, Titus specifies unit fusing adequate to handle the maximum amp draw possible under all operating conditions. This differs from the motor nameplate; it is essential that electric circuit fuses/overcurrent protection are sized according to the nameplate of the terminal, not the motor nameplate.

Fan Terminal Flow Control (continued)

0

Across MotorVoltage

Wat

ts

0

100%

Air Flow, cfm

50% 100%


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CATALOG FAN CURVESThe fan curves in a catalog represent the operating range of the fan powered terminal. Fan operation is dependent on the static pressure on the fan, so fan curves show airflow vs. static pressure. As the static pressure increases, airflow decreases. A typical fan curve will show maximum and minimum airflow for a fan powered terminal.

In (Figure 77), the top curve represents the maximum airflow that the fan and motor can provide. This corresponds to the recommended maximum operating rpm of the motor. The bottom curve shows the minimum airflow that the fan and motor can provide. This corresponds to either the minimum operating rpm of the motor or the minimum voltage of the SCR fan speed controller.

The SCR minimum is designed to protect the motor from operating below its recommended rpm. Most standard fan powered terminal motors must operate above a manufacturer’s specified rpm to effectively self-lubricate.

However, the relationship between rpm and SCR voltage is dependent of static pressure. At minimum voltage on the SCR, the motor rpm will be different at different static pressures. Because of this, there is a possibility that at minimum SCR voltage, the rpm will be below the motor minimum recommended operating rpm. When this happens, the cataloged fan curve will use minimum rpm to set the minimum fan curve, not minimum SCR voltage.

To ensure proper motor operation, always operate a fan powered terminal with the cataloged fan curve.

A Note on Meter Usage Many Digital Multi-Meters (DMMs) will provide erroneous readings when attempting to measure current or voltage near an SCR. These meters are designed for normal, smooth sine waves. The SCR, by changing the shape of the sine wave, throws off the readings from these meters. To measure the current voltage, a true RMS DMM designed for these conditions must be used.

Figure 77. Typical Fan Curve

Fan Speed Control (continued)

Figure 76. Watt, Volt and rpm Relationships

60

Volts --- 115

0

10

20

30

WA

TT

S

40

50

60

70

80

90

700

110

RMS VOLT TO MOTOR

8070 10090 120

320

RP

M

350

400

500

600

--- (2) 6" dia. x 6.5" wide

Type --- Permanent Split Capacitor

BLOWER --- Squirrel Cage

Rise --- 40 C Amps --- .77 FLA

MOTOR WATTS & RPM

RMS VOLTS TO MOTORVS

1050

900

800

1000

200

400

600

800

1000

1200

1400

1600

cfm

0 0.1 0.2 0.3 0.4 0.5 0.6

Static Pressure - Inches of Water


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PRESSURE INDEPENDENT - ENERGY EFFICIENT ANALOG SPEED SETTINGSECM MOTOR TECHNOLOGYThe ECM motor is an ultra-high efficiency, brushless DC motor with a unique microprocessor based motor controller. Motor efficiencies of 70% or better across the entire operating range of the motor saves considerable electrical energy when compared to conventional induction motors. The motor controller, when tuned to the fan powered terminal, provides a large turn down ratio and constant volume airflow regardless of changes in downstream static pressure operating against the fan. With the introduction of the ECM motor, factory setting of the fan cfm is now possible.

Separate controls are required to enable field adjustment of fan speed. The fan speed control allows adjustments to be made three ways.

• Manually with a screwdriver, similar to the SCR control.

• Remotely (as an option) through the DDC controls using a laptop at the unit.

• Remotely through the Building Management System.

HARMONICSPower for a given motor is drawn through the line in the form of a pure sine wave. This sine wave contains a fundamental frequency, in the US typically 60 Hz. When there exists other pure sine waves, each with individual frequencies, other than the fundamental frequency, they are called harmonics.These waves cause distortion or “noise” in the power line. Therefore, harmonic distortion is a collection of pure sine waves, including the 60 Hz fundamental frequency, which when summed together point by point in time creates distortion in the incoming line.

Due to the way a standard split capacitor motor draws power, they have slightly fewer harmonic frequencies as compared to the ECM motor. The ECM motor, unlike the standard split capacitor motor, draws peak power only when needed, resulting in less electrical noise generation.

As of 2011, the most stringent of limitations for harmonics is published in the CAN/CSA - CEI/IEC 61000-4-3-07 (R2011). These values set the ceiling for allowable harmonic levels. The critical maximum or peak amp values for a given harmonic level occur in the third harmonic closely followed by that of the fifth harmonic. Published data for a 1hp ECM without filtering capability violates these CEI limits. Titus has developed technology to decrease the harmonic frequencies while continuing to deliver peak power as it is requested. The Titus ECM motor meets the criteria, as well as specified national and international harmonic limitations.

ENERGY SAVINGS POTENTIALThe ECM motor, as applied to the Titus TQS fan powered terminal, offers significant energy savings over time to the owner when compared to conventional induction motors. Titus has evaluated an actual field trial and confirmed through bench testing an example of the potential energy savings when using the ECM motor. The following charts show the watt reduction associated with the ½ hp and 1 hp ECM motor when compared to standard TQS units of equivalent application range.

Note: TQS Size 6 with 1 hp ECM motor watt comparison to standard permanent split capacitor motor. The average watt reduction over the above range is 335 watts.

Note: TQS Size 4 with 1/2 hp ECM motor kW comparison to

standard permanent split capacitor motor. The average watt reduction over the above range is 178 watts.

Figure 78. Watt Reduction with 1/2 hp ECM Motor

Figure 79. Watt Reduction with 1 hp ECM Motor

ECM MOTORS

TQS Size 6 - 1hp ECM Motor

0

200

400

600

800

1000

1200

1263 1350 1601 1786 1849 2081 2162 2200cfm

Wat

ts ECMSCR

TQS Size 4 - ½ hp ECM Motor

0

100

200

300

400

500

600

509 752 1000 1415

cfm

Wat

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When evaluating this reduction in watts for energy usage the following table shows, at various usage rates, the annual savings per motor. Annual savings assume a run time of 3000 hours per year (250 days at 12 hours/day).

Also, reduction in demand charges must also be considered. Typically, demand charges are calculated during a 15-minute peak window. Some utilities will qualify the peak demand to only the summer months and use this peak as the monthly charge throughout the remainder of the year while other utilities will calculate demand charges using that months peak kW requirement. The savings associated with reduced demand charges are substantial, as demand charges are usually several dollars per kW. As an example, a typical multi-story office application may require 200 fan terminals.

Each fan terminal equipped with an ECM motor may have approximately 0.4 kW reduction in power. This translates to an 80 kW reduction in demand and with a demand rate of $10.00 per kW equates to a potential $800 per month reduction in the demand charges. While this model is simplistic, it is indicative of the payback potential of the motor. Utilities will vary not only in price but also in calculation methods with contract kW’s versus actual kW usage so actual savings must be calculated according to local market conditions.

Coupling the usage and demand savings associated with the ECM motors can provide a substantial savings throughout the life of the building.

DIRECT DIGITAL CONTROL

APPLYING COMPUTERS TO CONTROLWith many years of experience, design engineers have established the basic principles of temperature control for heating, ventilating, and air conditioning (HVAC) systems. These control strategies have been applied utilizing conventional pneumatic, electric or analog electronic devices.

Recent advances in micro-technology have made it possible to apply the power and precision of computers to HVAC control. Microprocessors, which cost less than ever before and offer superior computing power, are now suitable for application to individual air handlers, packaged heating/cooling units, VAV terminals or the entire HVAC system.

DIRECT DIGITAL CONTROLMicroprocessor-based controllers inherently perform direct digital control (DDC) and typically replace the conventional pneumatic or analog electronic controls. Digital controllers measure signals from sensors (input), process these signals in software (through the microprocessor), and initiate a corrective action to a controlled device (outputs) (Figure 80). A more technical definition is provided in the ASHRAE Applications Handbook.

ADVANTAGES OF DDCDDC systems offer several potential advantages over conventional counterparts.

• DDC systems provide improved comfort and greater energy efficiency through precise and accurate control. Pneumatic and Analog systems utilizing proportional (P) control have the inherent characteristic of offset (Figure 81). Microprocessor based controls can eliminate offset by adding the integral (I) or reset action. Furthermore, addition of the derivative (D) action can result in a faster response and greater stability (Figure 82), but requires significant tuning.

• DDC systems require less maintenance than conventional systems. Since there are no moving

parts, periodic preventive maintenance (PM) tasks such as calibration, lubrication, cleaning and adjustments are seldom required.

• Control strategies can be modified quickly and easily without the need to rewire, repipe or install additional components.

Table 9. Annual Savings per Motor

Usage KW/hr reductionsRate 0.28 0.35 0.40$0.05 $43.08 $52.50 $60.75 $0.06 $51.70 $63.00 $72.90 $0.07 $60.31 $73.50 $85.05 $0.08 $68.93 $84.00 $97.20 $0.10 $86.16 $105.00 $121.50 $0.12 $103.39 $126.00 $145.80 $0.14 $120.62 $147.00 $170.10

A direct digital controller receives electronic signals from the sensors, converts the electronic signals to numbers and performs mathematical operations on these numbers inside the computer. The output from the computer takes the form of a number, and can be converted to a voltage or pneumatic signal to operate the actuator.

ECM Motors - Fan Powered Terminals (continued)


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• Since microprocessor controllers are software based, multiple control sequences can be preprogrammed in memory thus allowing a single controller to be fully interchangeable between different equipment. For example, an application specific VAV controller may be used to control single duct, dual duct or fan powered terminals by simply choosing the appropriate operating sequence from a software library maintained on board every controller (Figure 83).

• While functioning completely independent, digital controllers perform all essential functions necessary to control different pieces of HVAC equipment without interconnecting to other computers. In this way each piece of HVAC equipment has its own digital controller in the same way conventional systems would provide individual control panels.

Figure 80. Direct Digital Controller Figure 81. Inherent Offset - Lost Energy Dollars and Sacrificed Comfort

Figure 82. Offset Completely Eliminated - Improved Comfort and Less Energy Usage

Figure 83. Frequently Used Control Sequences

Direct Digital Control (continued)

Air Flow

Humidity

Sensors

TemperatureOutputsMicroprocessorInputs

DamperActuators

Valve Actuator

ControlledDevices

Setpoint

Co

ntr

olled

Vari

ab

le

Time

InherentOffset

Co

ntr

olled

Vari

ab

le

Setpoint

Time

Primary Air

Variable Volume (Parallel Type)Fan Powered

Recirculated Air

Fan PoweredConstant Volume (Series Type)

Primary Air

Recirculated Air

Total Air FlowHeating / Cooling

Hot / Cold BlendingWith or WithoutDual Duct

Changeover

1F 1F

Cooling Only

Room TemperatureHeating Setpoint

Morning Warmup

Air

Flo

w, c

fm

Heating

Heating Min.

Slope Depends on rate and

Cooling Setpoint

Cooling Max.

CoolingDeadband

Cooling Min.

temperature change.magnitude of space


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DDC DISTRIBUTED PROCESSINGUsing a concept commonly referred to as distributed processing, DDC controllers can function as standalone devices. In this way if one controller fails, others throughout the system can continue to function unaffected. The controllers are connected over a system communication bus or local area network (LAN) for system wide sharing of information. This information is used to perform sophisticated building control strategies not possible with conventional noncommunicating systems. The network also allows system access locally through a personal computer or remotely via modem over telephone lines (Figure 84).

SIZING BASIC TERMINALS FROM CAPACITY TABLES

CERTIFIED AIR TERMINALSTo provide engineers with sound power data which can be compared on an even basis, leading air terminal manufacturers joined together under the Air-Conditioning, Heating and Refrigeration Institute (AHRI) to develop an industry standard for rating air terminals and certifying performance data. The result was AHRI Standard 880, Air Terminals, and the 880 Certification Program. Standard 880 specifies the procedure, using a reverberant chamber, for developing sound power data. The certification program ensures manufacturers’ equipment performance meets their claims.

Compliance with 880 is assured through third party testing. If a manufacturer fails to match claimed performance, the manufacturer must immediately rerate the terminal or lose the ability to use the AHRI Standard 880 seal. Another standard, AHRI Standard 885, was developed at the same time to assist the engineer in using certified product data.

Terminal selection involves a series of trade-offs. The designer needs to try to balance all of the constraining factors and select the terminal which meets overall needs best.

SIZING SINGLE DUCT TERMINALSThe starting point for sizing single duct terminals is to identify the type and model of controller. This is necessary because some controllers are more accurate at lower velocities than others.

Once the type of control is identified, the minimum and maximum primary airflows should be considered against the published cfm range. The trade-offs start here. Some engineers will select terminals near the bottom of the cfm range to reduce sound levels since large inlets reduce face velocity. Others select terminals near the top of the cfm range to hold down equipment costs. Still other engineers believe that one should remain comfortably in the middle to avoid potential control problems resulting from low velocities and sound problems occurring at high velocities.

All Titus products operate extremely well within the published cfm ranges. Therefore, low velocity control concerns can be eliminated. This leaves sound and first cost as the key issues. If the terminal is relatively small to begin with and will be located over a kitchen or hallway, sound will probably not be of concern and the designer may choose to slightly undersize the terminal. If, on the other hand, the terminal is located over office space, the designer may slightly oversize the terminal.

The selection of an appropriate water coil should also be considered at this time. In some cases, a terminal may need to be increased in size in order to obtain the desired heat output from the coil. With single duct units, the water coil air pressure drop should be subtracted from the duct pressure when determining sound generation. The sound produced by the damper is proportional to the pressure drop across the damper and discharge water coils may reduce that pressure drop. Other significant downstream pressure drops should be considered, and their pressure drop subtracted as well.

Figure 84. System Access via Network

Engineers who specify AHRI Certified air terminals are assured that the manufacturer’s performance meets the manufacturer’s claims. This is protection for the engineer, the building owner and the building occupant.

Direct Digital Control (continued)

VAV

Local Bus Network

Local Bus Network

Local Bus Network

InterfaceSystem

HVAC EquipmentControlling

Modem

VAVVAV



VAV VAVVAV

VAV VAVVAV

VAV

VAV

VAV


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SIZING PARALLEL FAN POWERED TERMINALSParallel 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).

Ps2 = Ps1 (V1 / V2)2

Where: Ps1 = Primary Air Static Pressure Ps2 = Secondary (Fan) Air Static Pressure V1 = Primary Air Velocity V2 = Secondary (Fan) Air Velocity

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 86, 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 chart on page R46, 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.

o1400

p

2x 0.30” = 0.44” sp

1150

Figure 85. Schematic Diagram of Airflow in Parallel Flow (Variable Volume) Models

Figure 86. Actual Arrangement of Components Shown in the Previous Schematic Diagram

Sizing Terminals from Capacity Tables (continued)

Fan

AirSection

Primary

Recirculated

PrimaryAir

Air

Either / Or Downstream Ductand Diffuser

Backdraft

Secondary

Primary

Damper

OutletsTo


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SIZING SERIES FAN POWERED TERMINALSCompared to single duct terminals, series flow (constant volume) fan powered terminals add the additional factor of fan cfm requirements. The designer must consider both the primary airflow and the fan. Series terminals are selected based on the capacity of their fans. The secondary (or fan) cfm should be equal to or slightly more than the primary air to ensure primary air does not short circuit through the induced air port into the plenum, thereby wasting energy.Before selecting the fan, the static pressure downstream of the terminal must be determined. This is the resistance of the ducts and diffuser(s) at design airflow rates.

Once the downstream static pressure is known, the designer can select the fan based on the fan curves (these are shown throughout the catalog with the performance data for each fan powered terminal). The designer should find the intersection of the static pressure line on the horizontal axis and the fan cfm on the vertical axis. Selecting toward the upper end of the range will ensure that first costs are kept low and the fan motor efficiency is high. Selecting below the indicated minimum flow will result in shortened motor life as the bearings in the motor are centrifugally lubricated.If a water coil is needed, the designer must use the curves provided for a one or two row coil. These curves account for the additional static pressure generated by the coil. The static pressure added for an electric coil is negligible and may be disregarded. Neither has an appreciable effect on sound levels.

Inlet size must also be selected. Fan powered terminals come with varying inlet sizes. In general, inlets should be selected toward the bottom of the range. This reduces the face velocity of the inlet and minimizes the sound generated by the primary air valve.

To select a Titus series fan powered terminal unit, refer to the published fan curves and primary air pressure drop curves together with the application and sound power data. An abbreviated table is shown at the right for use with the example discussed here.

In the series flow type of unit, the fan runs continuously in the standard version. With the optional night shutdown and night setback controls, the fan can be cycled ON and OFF when the primary air is OFF.

As shown in the diagrams below, the primary air is drawn into the fan inlet along with secondary (recirculated) air from the room. The maximum primary airflow must always be equal to, or less than, the total airflow through the fan.When the primary air section reduces its airflow in response to a reduced demand for cooling, the fan makes up the difference by drawing more recirculated air from the room. As a result, the flow rate to the room is constant.

The primary air section discharges into the unit casing near the fan inlet, where the static pressure is slightly below atmospheric. For this reason, the available inlet pressure need only be enough to overcome the internal pressure drop through the primary air damper itself.

Example: Select a Model DTQS for a maximum of 1200 cfm of primary air at 0.50” wg inlet static pressure. The fan airflow is 1200 cfm. The downstream resistance offered by the duct and diffusers is 0.30” at 1200 cfm.

Primary Air: From the table on page R11, a size 4 will handle 1200 cfm of primary air with a minimum static pressure drop of .18” through the primary air section. Since 0.50” static pressure is available at the inlet, the selection will work.

Secondary Air (Fan): From the published fan curves, a size 4 terminal will handle 1200 cfm at 0.30” static pressure, with the proper setting of the standard SCR speed control.

Figure 87. Schematic Diagram of Airflow in Constant Volume (Series Flow) Models

Figure 88. Actual Arrangement of Components Shown in the Previous Schematic Diagram

Sizing Terminals from Capacity Tables (continued)

Secondary

Primary

OutletsTo

Internal ∆Ps

Section

Primary

Air

AirRecirculated

Primary Air

Downstream Duct

Downstream (Fan)

Fanand Diffuser

∆Ps


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TYPICAL PROBLEMS

OVERSIZING TERMINALThe direct result of oversizing is low air velocity. With the velocity too low, the damper must operate in a pinched-down condition most of the time, making control difficult. The inlet velocity can also be too low for effective operation of the sensor and controller. Too low a velocity through an electric heater will cause the safety airflow switch to shut down the heater. Oversizing fan terminals results in low fan motor rpm and the potential for under-lubrication of the motor bearings, resulting in shortened motor life and additional sound from larger motors (Figure 89).

CAPACITY CONCENTRATED IN TOO FEW TERMINALSWhen one large terminal serves a space that should be served by two or more smaller ones, comfort problems can result. There may be noticeable temperature differences between rooms, since the thermostat is located in just one room as at the right. Also, for a given air velocity, the larger the terminal the more sound power it generates (Figure 90).

INSUFFICIENT SPACECarefully planning the locations of the terminals avoids problems with installation, performance, and maintenance.In the example shown at the right, the control side of the terminal is against the wall, making connections difficult and service almost impossible. The cramped location also creates the need for close-coupled duct elbows, which reduce performance (Figure 91).

IMPROPER DISCHARGE CONDITIONSThe duct connections at the discharge end of the terminal have a major effect on pressure drop. A tee close to the discharge is especially to be avoided, along with transition pieces and elbows. Another common error is running too much flex duct, as at the right. It would have been better to continue the rectangular duct to the last diffuser, then install short flex branches (Figure 92).

IMPROPER INLET CONDITIONSThe arrangement of duct at the terminal inlet affects both pressure drop and control accuracy.

The conditions shown at the right will create turbulence at the inlet. This makes it difficult for the sensor to measure airflow accurately. Although Titus velocity sensors correct for a considerable amount of turbulence, the best practice is to use straight duct at the inlet the same size or larger than the inlet (Figure 93).

Figure 89. Low Velocity Effects

Figure 90. Too Few Terminals Effect

Figure 91. Installation Affecting Performance

Figure 92. Improper Discharge Conditions

Figure 93. Improper Inlet Conditions

accurate damperingVelocity too low for

damper travelTotal

accurate damperingVelocity too low for

for sensorVelocity too low

electric heaterVelocity too low for

than one zone.serves more

Terminal

Large terminal

Room 101Too Hot

Room 103

T

Terminal

Controls Coil Connections

ChasePipe

flex ductToo much

TerminalTerminal

Tee atDischarge

smaller than inletSupply ductInlet tapped

into side of duct at inletTight elbow


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INCOMPATIBILITY WITH POWER SOURCEIn fan powered terminals, electrically or electronically controlled terminals, and all terminals with electric heating coils, the order to the factory should be carefully checked against the electrical characteristics of the power source at the point of connection.

Not only must the voltage, phase, and frequency match, but the distinction between 3 phase-3 wire and 4 wire wye must be observed (Figure 94).

EXCESSIVE AIR TEMPERATURE RISE AND AIR CHANGE EFFECTIVENESSThe discharge temperature for terminal units should be selected so the maximum temperature difference between the room and the diffuser discharge is no greater than 15BF. This can be found in the ASHRAE Handbook of Fundamentals. According to ASHRAE 62.1, ceiling diffusers which have ceiling returns and are used for heating, should be mounted as shown to allow the supply air jet of 150 fpm to come down the exposed wall to within 4.5 ft. of the floor level. This reduces the short circuiting of warm air at the ceiling level and can be used to achieve an Ez air change effectiveness value of 1.0 as determined in ASHRAE Standard 129 for all air distribution configurations except unidirectional flow (Figure 95).

EXCESSIVE AIR LEAKAGELeakage from the branch duct upstream and downstream from the terminal, as well as from the terminal itself, can be serious. In some installations it is found to be as much as 10% or more of the total airflow.

Most of this leakage can be avoided by careful fabrication and installation and the use of top quality terminals(Figure 96).

IMPROPER SUPPORT OF TERMINALMany terminals are light enough to need no support other than the duct work itself. However, the larger sizes, units with electric coils and fan powered models, are heavy enough to require additional support. A practical method is to use hanger straps screwed to the sides of the terminal. The bottom should be left clear where there are access panels (Figure 97).

WRONG TYPE OF INSULATIONInstallations in hospitals, clean rooms, and laboratories often require a special insulation liner to prevent air erosion or microbial growth. In the past, Mylar and Tedlar were often specified in these installations. Neither, however, meet current safety codes in many cities. Foil faced insulations, such as foil-faced Eco-Shield and Steri-Loc, provide the required covering, meet all safety codes and actually provide some sound attenuation. Titus Fibre-Free insulation provides both sound attenuation and resistance to erosion and mold growth (Figure 98).

Figure 94. Power Source Compatibility

Figure 95. Overhead Heated Air

Figure 96. Possible Air Leakage

Figure 97. Terminal Support

Figure 98. Anti-Erosion Skin Effects

Typical Problems (continued)

3 Wire3 Phase

L3

3 Phase4 Wire Wye

L2

L1

L3

Neutral

L2

L1

150 100 50

X 150 fpm

4.5 ft.

Branch Take-offTerminal Coil

DiffuserNeck

Joint

Hanger Straps Screwed to TerminalSupported by Duct Work

Large, Heavy Terminal

Skin Addedto Insulation

Insulation LacksAnti-Erosion Skin


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B23


NON-COMPLIANCE WITH LOCAL CODESSome localities have stringent codes of their own, with requirements beyond those of NEC, UL, and CSA. An example is the primary fusing in the control transformer at the right (Figure 99).

INSTALLATION TECHNIQUES-DUCT CONNECTIONSThe inlet duct slips over the inlet collar of the terminal. It should be fastened and sealed according to the job specifications.

The diameter of the inlet duct must be equal to the listed size of the terminal. For example, a duct that measures 8” in diameter must be fitted to a size 8 terminal. The inlet collar of the terminal is made ⅛” smaller than nominal size in order to fit inside the duct (Figure 100).

Note: A duct should never be inserted inside the inlet collar of the terminal.

For optimum control accuracy, a straight section of unrestricted duct at least 1½ diameters long should be installed at the inlet (Figure 101). Where this condition does not exist, field adjustment of the airflow setting on the velocity controller may be required.

If space does not permit using the 1½ diameter length of straight duct, a hard duct elbow up to 90° can be installed at the inlet of the Titus terminal without altering the factory maximum or minimum airflow setting by more than 10% (Figure 102).

The outlet end of the Titus terminal is designed for a slip and drive connection. Unless a round duct adapter is furnished, a rectangular outlet duct should be fitted to match the size of the terminal casing. It should be fastened and sealed according to the job specifications.

If a round outlet adapter is furnished, it should be fastened and sealed by the same method used for the inlet.Close coupling the terminal inlet to the side of the main supply duct is not recommended. Where this condition is unavoidable, a flow straightening device (Figure 103) should be installed between the main supply duct and the inlet to the terminal. Even with the flow straightening device, the terminal may still require some field adjustment of the factory airflow settings at the velocity controller.

Air leakage adds significantly to the operating cost of an HVAC system. Important savings are realized by carefully fitting and sealing all duct joints and specifying tightly constructed Titus terminals. The Titus box has very low damper and casing loss leakage. These values can be found on page Q26.

Figure 99. Primary Fusing in the Control Transformer

Figure 100. Terminal Inlet Collar Fitting Properly

Figure 101. Unrestricted Duct Properly Install at the Inlet

Figure 102. 90 Degree Hard Elbow Duct Installed to Inlet

Figure 103. Flow Straightening Device Placement


277 VAC

277 VAC

TransformerFuse

24 VAC

24 VAC

Straight Duct

Terminal

Inlet

Inlet

Flexible Duct

Minimum1.5 Diameter

Terminal

Terminal

Elbow

Duct

Hard

Inlet

Straightener

Main Supply

Duct

Inlet

Air

Terminal


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A 10” terminal handles 1150 cfm. The central system cools air from 80°F dry bulb / 67°F wet bulb to 53°F dry bulb / 51.5°F wet bulb before sending it to the terminal.

Cost of RefrigerationThe total heat removed from the system is: 31.65 Btu/# at 80/67 minus 21.10 Btu/# at 53/51. = 10.45 Btu/# of dry air.

The amount of leakage given in this example is 58 cfm.

The loss of refrigeration energy through leakage is: 58 cfm x 10.45 Btu/# x 4.5 = 2727 Btu/hr.

Assuming a cooling system EER of 7.5 overall (reference ASHRAE Standard 90), in a space where the system operates 24 hours a day, 365 days a year, (worst case)At a power cost of $0.06 per kwh, 0.3636 x $0.06 x 24 x 365 = $191.11

If the system operates at 40% capacity, averaged over:one year, $191.11 x 0.40 = $76.44, the cost of wasted refrigeration power alone, again worst case, assuming continuous operation.

The amount of leakage in the branch duct serving the terminal, the connections to the terminal, the terminal

itself, and the duct downstream from the terminal is 5% of the 1150 cfm being handled, or about 58 cfm.

Cost of Fan OperationIf the static pressure across the fan is 5” wg and the fan static efficiency averages 75%, the leakage converts to:

Assuming that the motor efficiency multiplied by the power factor averages 0.80, 0.0569 x $0.06 x 24 x 365 x 0.40 = $11.96, the cost of wasted fan power.

Combined cost equals:

per year for one terminal.

The example in (Figure 104) shows how many dollars can be lost in the leakage from just one terminal together with its connected duct work. Multiply that amount by the hundreds or thousands of terminals that may be in one building, and the seriousness of the loss is apparent.

This is a conservative example, in that the leakage is only 5%; a much higher percentage is found in many installations. Also, the compressor, pumps, and fans may not run as efficiently as indicated here, and the cost of electric power in many parts of the country is greater than $0.06 per kilowatt hour.Figure 104. Possible Air Leakage

Example of Leakage Costs

2727 = .0.3636 kWh input

7.5 x 1000

58 cfm x 5 = .0.061 bhp

6356 x 0.75

0.061 x 746 = .0.0569 kw

0.080 x 1000

$76.44

$11.96

$88.40


Branch Take-offTerminal Coil

DiffuserNeck

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Three categories of pressure are connected with air handling:

1. Static pressure may be thought of as the pressure in a tire or storage tank. It is exerted in all directions equally.

2. Velocity pressure, as its name implies, is entirely a function of air velocity and its direction. It is the pressure you feel against your hand if you hold it outside the window of a moving car.

3. Total pressure is the sum of static pressure and velocity pressure. It and static pressure are the pressures actually sampled by velocity sensors in terminals and by commonly used measuring devices, as described next.

The interaction of static, velocity, and total pressures is illustrated by (Figure 105). The Pitot tube, which is used to measure velocities and pressures, is really a tube within a tube. The inner, or impact, tube senses both the velocity pressure and static pressure combined (total pressure). The outer tube, which communicates with the airstream through small holes in its wall, avoids the impact of the air movement and senses only static pressure.

The U-tube manometer, connected to both parts of the Pitot tube, has the effect of subtracting static pressure from total pressure to give a reading of velocity pressure.

Once the velocity pressure is known, the velocity can be calculated easily:

where V = Air Velocity and Pv = Velocity Pressure

Knowing both the velocity and the cross-sectional area of the duct, the flow rate is then:

cfm = Area x Velocity Figure 105. Static, Velocity and Total Pressures Interaction

PRESSURE MEASUREMENT

Pv = oV

p

2 or V = 4005√Pv4005

PV Difference in Liquid Level

PT minus P

S = P

V

Glass Manometer Tube

Connecting

Tubing

Pitot Tube

PT

PS

PT

PS


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The Fan Laws are basic tools in air handling. Three of the most common relationships are illustrated as follows.

Example:

A fan handles 40,000 cfm at 2” static pressure. It runs at 760 rpm and draws 18 brake horsepower. The fan is increased to 800 rpm. What are the new cfm, sp, and bhp?

1. Airflow rate varies directly with shaft speed.

2. cfm2 = (cfm1 x rpm2) / rpm1 = (40,000 x 800) / 760 = 42,105

3. Pressure varies as the square of shaft speed. P2 = P1 x rpm22 / rpm12 = 2 x (800)2 / (760)2 = 2.22”

4. Horsepower varies as the cube of shaft speed. bhp1 = rpm13 bhp2 = ((bhp1) x (rpm2)3 / (rpm1)3 = ((18) x (800)3) / (760)3 = 21.0

The relationships stated here apply when the air density remains constant and when there is no change in the fan or the system. They are based on Fan Laws 1, 2 and 3. For a complete presentation of the Fan Laws, see the ASHRAE Handbook, Systems and Equipment.

Each fan design has its characteristic set of performance curves. Those shown in (Figure 109) are typical of a centrifugal fan with forward curved blades in the wheel, as commonly used in fan powered terminals. For a full discussion of the characteristics of the various types of fans, see the ASHRAE Handbook, Systems and Equipment.

The solid curve represents a fan running at constant speed, as it is throttled from free delivery to close-off. The broken line square curve represents the pressure drop through the complete air handling system in which the fan operates. Intersection (A) is the operating point of the fan.

The dashed line represents another system pressure curve which intersects at point B. This point is a poor operation point as instability will likely reset.

Figure 106. Fan Law - Airflow

Figure 107. Fan Law - Pressure

Figure 108. Fan Law - Brake Horsepower

Figure 109. Centrifuge Fan Performance Curves

cfm1 rpm1

cfm2 rpm2=

P1 rpm12

P2 rpm22=

THE FAN LAWS

Fan Law 1

0

500

600

700

400

800

900

1000

0 10000 20000 30000 40000 50000

cfm

rpm

100

200

300

Fan Law 3

0

100

200

300

400

500600

700

800

900

1000

0 2 4 6 8 10 12 14 16 18 20

bhp

rpm

Fan Law 3

0

100

200

300

400

500600

700

800

900

1000

0 2 4 6 8 10 12 14 16 18 20

bhp

rpm

System Curve

0

0.5

1

1.5

2

2.5

3

3.5

1000 3000 5000 7000 9000 11000 13000

Volume Flow Rate

Tota

l Pre

ssur

e

A

B


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REHEAT COILS:

Several types of terminal devices are available with reheat coils, both hot water and electric. When determining the heat requirement for a terminal, the engineer will often start with the known zone heating demand, typically expressed in BTUH, or more conveniently, MBH (thousands of Btu’s). The room load requirements for heating are then used to determine the Room Entering Air temperature (EATr) by the equation:

Btuh (room) = 1.085 * (EATr - Tr) * Q

Where;

EATr = Temperature (°F) entering the roomTr = Room setpoint temperature or average temperatureQ = Flowrate ( cfm) (typically 30 - 50% of the cooling cfm)

By solving for the EATr, the coil Btuh requirements can then be determined. The room entering air temperature (EATr) now becomes the required LAT of the VAV box (ignoring any duct heat losses). The coil can now be sized according to:

Btuh (coil) = 1.085 * (LAT - EATc) * Q

Where;

LAT = Coil leaving air temperatureEATc = Coil entering air temperature (primary or mixed air)Q = Flowrate (cfm)

Now that the coil requirements are known, published catalog data may be used to select the proper hot water or electric coil.

Formulas and Definitions PowerVP = (fpm / 4,005)2 W = Watts

(Q) cfm = Cubic Feet per Minute A = AmpsTP = Total Pressure hp = HorsepowerSP = Static Pressure V = VoltsVP = Velocity Pressure E1 = Efficiency

(V) fpm = Feet per Minute PF = Power FactorDP = Differential PressureDPs = Static Differential Pressure Power DC CircuitsDPT = Total Differential Pressure W = V x A

(A) Area Factor = Dimension in Square Feet A = W / VVP = TP - SP hp = V x A x E / 746TP = SP + VP E = 746 x HP / WSP = TP - VP

cfm = fpm x Area Factor Power AC Circuits (Single Phase)DPT = TP1 - TP2 PF = W / (V x A)DPs = SP1 - SP2 A = 746 x HP / (V x E x PF)DP = (cfm / K)2 E = 746 x HP / (V x A x PF)

fpm = cfm / Area Factor kW = V x A x PF / 1,000K = cfm/ √ (DP) hp = V x A x E x PF / 746

Water CoilsMBH = 1,000s of Btus per Hour Power AC Circuits (3 Phase)Btu = British Thermal Unit PF = W / (V x A x 1.732)

gpm = Gallons per Minute A = 746 x HP / (1.732 x V x E x PF)DT = Temperature Differential E = 746 x HP / (V x A x PF x 1.732)

Air DT = 927 x MBH / cfm kW = V x A x PF x 1.732 / 1,000H20∆T = 2.04 x MBH / gpm hp = V x A x 1.732 x E x PF / 746

Electric CoilskW = Kilowatts

Air DT = Temperature Differential, Leaving Air - (minus) Entering Air Temperature

kW = cfm x DT / 3,160DT = kW x 3160 / cfm

Table 10. Units of Measurements

EQUATIONS AND DEFINITIONS


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Air terminals are the most noise sensitive of all HVAC prod-ucts since they are almost always mounted in or directly over occupied spaces. They usually determine the residual background noise level from 125 Hz to 2,000 Hz. The term “Air Terminals” has historically been used to describe a num-ber of devices which control airflows into occupied spaces at the zone (or individual temperature control area) level. There are two types: those that control the amount of air-flow to a temperature zone (Air Control Units, ACUs, or more commonly “Boxes”), and those that distribute or collect the flow of air (Grilles & Diffusers, GRDs). On some occasions, the two functions are combined. As these two elements are the final components in many built-up air delivery systems and those closest to the building occupants, both are critical components in the acoustical design of a space. There is also a critical interplay between acoustics and the primary function of these devices; providing a proper quantity of well mixed air to the building occupants. Before discussing types of devices, we must have an understanding of some issues regarding sound levels in occupied spaces.

The sound level in an occupied space can be measured directly with a sound level meter, or estimated from published sound power after accounting for room volume and other acoustical factors. Sound level meters measure the sound pressure level at the microphone location. Estimation techniques calculate sound pressure level at a specified point in an occupied space. Measured sound pressure levels in frequency bands can then be plotted and analyzed, and compared with established criteria for room sound levels.

Sound power cannot be measured directly, (except using special Acoustic Intensity techniques), and is a measure of the acoustical energy created by a source. It is normally determined in special facilities and reported for devices under stated conditions. Sound Power Level (Lw) values

for Air Terminal devices are usually reported as the sound power level in each of several octave bands with center frequencies as shown in Table 11. Sound Power Levels are given in decibels (dB) referenced to a base power in watts, typically 10-12 watts. Sound power levels can also be reported for full or 1/3 octave bands, but usually as full octave bands, unless pure tones (narrow bands significantly louder than adjacent adjacent bands) are present.

NOISE CRITERIA (NC)Sound Pressure Levels (Lp) are measured directly by sound level meters at one or more points in a room. They reference a pressure rather than a power. A product’s estimated Sound Pressure Level (Lp) performance curve is obtained by subtracting space (or other appropriate) sound attenuation effects from the unit sound power (Lw). Currently, most Air Outlet and Inlets (GRDs) sound performance is reported by subtracting a 10 dB attenuation from all octave band sound power levels, and determining the NC rating. This room effect approximates a 3,000 cu. ft. room, 10 ft. from the source for VAV boxes, which peak in lower frequencies, and a 2,500 cu. ft. room 7 ft. from a diffuser, which typically peaks @ 1000 Hz. (This is defined in the AHRI Standard 885 space effect calculation described later in this section). NC curves were developed to represent lines of equal hearing perception in all bands and at varying sound levels. Most air terminal products are currently specified and reported as single number NC ratings.

Table 11. Octave Band Designations

Center Frequency 63 125 250 500 1000 2000 4000 8000

Band Designation 1 2 3 4 5 6 7 8

Acoustical Applications and Factors


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ACOUSTICAL APPLICATIONS AND FACTORS

The use of a 10 dB “room effect” as used in this example, while in common practice and accepted for many years, is not as accurate a prediction as is possible using newer techniques. The ASHRAE Handbook and AHRI Standard 885 present an equation for determining the “space effect” based on both room volume and the distance from the observer to a point sound source.

Space Effect = (25) - 10 Log (ft.) - 5 Log (cu. ft.) - 3 Log (Hz)

Where: ft. = Distance from observer to source cu. ft. = Room volume Hz = Octave band center frequency

This yields a range of deductions which differ in each octave band, as shown in Table 12.

The 10 dB room effect which has been traditionally used for diffuser sound ratings, which typically peak in the 5th band, can be considered to be equivalent to a room about 2,500 cu. ft. in size, with the observer located about 7 ft. from the source.

With VAV terminals, which peak in lower bands, the “10 dB” room size is larger, or the distance is greater.

NC ratings have been common in specifications for a number of years, with an NC-35 being the most common requirement. While NC is a great improvement over previous single number ratings, including Sones, Bels, and dBA requirements, it gives little indication of the “quality” of the sound. A more comprehensive method, RC, has been proposed; while a good analysis tool, RC is a very poor design tool.

Figure 127. Typical NC Graph for a Diffuser

Table 12. Space Effect (AHRI 885 and ASHRAE)

Room Hz 63 125 250 500 1000 2000 4000 8000

Volume Band 1 2 3 4 5 6 7 8

2000 CuFt

@ 5ft -4 -5 -6 -7 -7 -8 -9 -10

@ 10ft -7 -8 -9 -10 -11 -11 -12 -13

@ 15ft -9 -10 -10 -11 -12 -13 -14 -15

2500 CuFt

@ 5ft -4 -5 -6 -7 -8 -9 -10 -11

@ 10ft -7 -8 -9 -10 -11 -12 -13 -14

@ 15ft -9 -10 -11 -12 -13 -14 -15 -15

3000 CuFt

@ 5ft -5 -6 -7 -7 -8 -9 -10 -11

@ 10ft -8 -9 -10 -10 -11 -12 -13 -14

@ 15ft -10 -10 -11 -12 -13 -14 -15 -16

5000 CuFt

@ 5ft -6 -7 -8 -9 -9 -10 -11 -12

@ 10ft -9 -10 -11 -12 -12 -13 -14 -15

@ 15ft -11 -12 -12 -13 -14 -15 -16 -17

In this example, the outlet Lp spectrum does not exceed the NC curve of 30 in any of the eight octave bands and is thus referred to as meeting an NC-30 criteria and specification.

It should be noted that while this spectra meets NC-30, if the critical band resulted in an NC-33, most building occupants would not be able to discern the difference.

NC - Noise Criteria (continued)

10

20

30

40

50

60

70

80

63 125 250 500 1K 2K 4K 8K

Mid - Frequency, Hz

dB

ApproximateThresholdof humanhearing

NC rating given is NC-30 since this is the highestpoint tangent to an NC curve

Sound Power

NC-70

NC-20

NC-60

NC-50

NC-30

NC-40

Sound Power less 10 dB in each band


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Room Criteria (RC) is based on ASHRAE sponsored studies of preference and requirements for speech privacy, along with ratings for “Acoustical Quality.” RC ratings contain both a numerical value and a letter “Quality” rating. The RC numerical rating is simply the arithmetic average of the sound pressure level in the 500, 1,000 and 2,000 Hz octave bands, which is the speech interference level (SIL).These are the frequencies that affect speech communication privacy and impairment. Studies show that an RC between 35 and 45 will usually provide speech privacy in open-plan offices, while a value below 35 does not. Above RC-45, the sound is likely to interfere with speech communication.

In addition to the numerical SIL portion of the RC method, there is a “Quality” portion of the RC rating which involves an analysis of potential low and high frequency annoyance. The goals of acoustical quality are described in Table 13.

Recommended NC and RC goals for various space applications, given in the current ASHRAE Handbook, are shown in the table to the right.

Table 13. ASHRAE Defined Acoustic Quality

Not too quiet Don’t destroy acoustic privacy

Not too loud Avoid hearing damage Don’t interfere with speech

Not to annoyingNo rumble, No hiss

No identifiable machinery sounds, No time modulation

Not to be felt No feel able wall vibration

Table 14. NC/RC Guidelines

ROOM CRITERIA (RC)

Space RC NCOccupancyPrivate residence RC 25-30(N) NC 25-30

Apartments RC 30-35(N) NC 30-35

Hotels/Motels

Individual rooms or suites RC 30-35(N) NC 30-35

Halls, corridors, lobbies RC 35-40(N) NC 35-40

Service / support areas RC 40-45(N) NC 40-45

Offices

Executive RC 25-30(N) NC 25-30

Conference Rooms RC 25-30(N) NC 25-30

Private RC 30-35(N) NC 30-35

Open plan areas RC 35-40(N) NC 35-40

Public circulation RC 40-45(N) NC 40-45

Hospitals and clinics

Private rooms RC 25-30(N) NC 25-30

Wards RC 30-35(N) NC 30-35

Operating rooms RC 25-30(N) NC 25-30

Laboratories RC 35-40(N) NC 35-40

Corridors RC 30-35(N) NC 30-35

Public areas RC 35-40(N) NC 35-40

Churches RC 30-35(N) NC 30-35

Schools

Lecture and classrooms RC 25-30(N) NC 25-30

Open-plan classrooms RC 35-40(N) NC 35-40


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The Four “Quality” letter designations currently in use are:

R Rumble H Hiss V Vibration (acoustically induced) N Neutral

These letters are determined by analyzing the low and high frequency spectra compared to a line drawn with a -5 dB slope per band through the numerical RC point @ 1,000 Hz. This establishes the SIL (Speech Interference Level) line. Lines of -5 dB slope create the RC chart, shown in (Figure 128).

Also shown in (Figure 128) are areas of rumble (B) and vibration (A), as well as a threshold of audibility. Note that the RC chart goes well below the 63 Hz lower cutoff of the NC chart, as these low frequency sound levels have been discovered to be a major source of discomfort to occupants.

If the sound spectrum being analyzed exceeds a line drawn parallel to the SIL line plus 3 dB in the higher frequencies (> 2,000 Hz), the “hiss roof,” then it is declared to be “Hissy” and gets an “H” designation.

Figure 128. Room Criteria (RC) Curves

Figure 129. Hissy Spectrum

Room Criteria (continued)

RC50454035

3025

A

B

C

Adapted from 1989 ASHRAE Fundementals Handbook - Atlanta, GA

10

30

40

50

60

70

20

80

90

16 63 250 1K 4KOctave Band Center Frequency, Hz

Mic

ropa

scal

s

Threshold ofAudibility

OctaveBand

Sound Pressure

Level,dB re 20

Region A: High probabilitythat noise induced vibrationlevels in light wall andceiling structures will benoticeable. Rattling of lightweight light fixtures, doors, and windows should be anticipated.

Region B: moderate probability that noise induced vibration will be noticeable in lightweight fixtures, doors, and windows.

10

20

30

40

50

60

70

80

90


OctaveBand

Sound Pressure

Level

Adapted from 1989 ASHRAE Fundementals Handbook - Atlanta, GA

C

Measured data is outside the referenceregion by >3 dB, above the 1000 Hz octave band,therefore the noise is likely to be interpreted as "hissy."

PSIL = (35+36+34) / 3 = 35

RC-35(H)


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If the plotted spectra exceed a +5 dB Rumble roof in the lower frequencies, it gets an “R” (Rumble) Rating. Finally, if the sound spectrum enters the “A” or “B” zones in the very low frequencies shown on the RC graphs, it warrants a “V” for possible wall or furniture vibration induced by acoustical energy in low frequencies. The B region may be characterized as “Be Careful” where one may have complaints, but the “A” region is “Awful”, and complaints should be expected. This is an example of an RV spectrum.

When room sound levels are analyzed using RC curves, air diffusers tend to give the same ratings as they do in an NC analysis. Boxes, on the other hand, which are typically predominant in lower frequency sound and often characterized as “Roar” (250-500 Hz), often yield RC values lower than NC, but with an “R” classification.

AIR TERMINAL SOUND ISSUESSound is an important design criterion in the application of air terminals. In the context of the total building environment, comfort cannot be achieved with excessive sound or noise levels. By definition, sound is a change in pressure for a medium, such as air. This change in pressure involves a radiation of energy. Energy is used in the generation of sound and this energy is radiated from a source.

All sound has a source and travels down a path to a receiver. Air terminals are one source of sound in a mechanical system. The path for sound emanating from the air terminal is through the plenum or down the duct into the conditioned space where it reaches the occupant or receiver.

Mechanical system designers should not be concerned so much with sound, but rather with noise. Noise can be thought of as unwanted or excessive sound. Good design practice dictates that a designer establish the acceptable noise for the occupied space and then determine the selection criteria for the mechanical system components.

In any application, both radiated and discharge sound should be considered. Radiated sound “breaks out” from the terminal casing or induction port and travels through the plenum and ceiling to enter the occupied space. Discharge sound travels out the discharge of the terminal through the duct work and outlet to enter the occupied space.

Figure 130. Rumble and Induced Vibration (RV) Spectrum

Figure 131. Typical Sound Sources for Fan Terminal System

Room Criteria (continued)

10

20

30

40

60

50

70

8080

8090


Octave Band Sound

PressureLevel

A

B

C

Even though the PSILis only 33 dB, thenoise spectrum falls within regionsA & B indicating ahigh probability ofmoise-induced vibration in lights,ceilings, air diffusers,and return air grilles.

PSIL= (38+32+29) / 3 = 33 RC-33(RV)

C

O

C

D

O

= Casing Radiated and Induction Inlet

= Discharge Sound

= Outlet Generated Sound

Sound Power L w

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AHRI Standard 885, Procedure for Estimating Occupied Space Sound Levels in the Application of Air Terminal and Air Outlets, provides the most current application factors for converting rated sound power to a predicted room sound pressure level. This standard is the basis by which most air terminal manufacturers convert sound power, as measured in reverberant rooms per ASHRAE Standard 130 and rated in accordance with AHRI Standard 880, to a predicted room sound pressure level. The standard provides a number of equations and tables available elsewhere, but puts them all in one document, and includes some unique tables as well. It also includes examples and diagrams to make the process easier to use. The most important of those are included here.

ENVIRONMENTAL ADJUSTMENT FACTORIn order to use the AHRI 885 Standard, sound power must be corrected for differences between reverberant room and free field calibrations when AHRI Standard 880 sound power is the base. This Environmental Adjustment Factor is listed in AHRI Standard 885.

According to AHRI Standard 885, an Environmental Adjustment Factor must be applied to manufacturers’ data if the sound power data has been taken under a free field RSS (reference sound source). According to AHRI, this is “necessary because at low frequencies, all real occupied spaces behave acoustically more like reverberant rooms than open spaces (free field).” In other words, manufacturers’ sound power data which is based on ILG/RSS with a free field calibration must be adjusted to match actual operating conditions found in the field. This applies to Titus and other participants in the AHRI Standard 880 Certification program. For data rated per AHRI Standard 880, the environmental adjustment factor must be subtracted from the manufacturers’ sound power level data in order to use the adjustments provided in AHRI Standard 885.

RADIATED SOUND POWER LEVELSTo determine the maximum allowable radiated sound power levels for a project, the attenuation from the ceiling/space effect must be added to the desired room sound pressure for each octave band.

CEILING/SPACE EFFECTAHRI Standard 885 combines the effect of the absorption of the ceiling tile, plenum absorption and room absorption into the Ceiling/Space Effect. Experience has shown that the Sound Transmission Class (STC) rating for ceiling tiles, which is based on a two room pair test, is not well correlated with observed data for a noise source located above a ceiling. The AHRI Standard 885 Ceiling/Space Effect table D14 Table 16 is derived from a number of manufacturers’ observations and is only found in the AHRI Standard. This table assumes that the plenum space is at least 3 ft. deep, is over 30 ft. wide or lined with insulation and that there are no penetrations directly under the unit.

From the AHRI Standard, the following attenuation values, or transfer functions, should be used for the Ceiling/Space Effect:

LW RAD = LP + S + P/C + Env

where: LW RAD = Radiated Sound Power Level LP = Sound Pressure Level S = Space Effect P/C = Plenum/Ceiling Effect Env = Environmental Effect

Once the Ceiling/Space Effect has been determined, they are added to the sound pressure level to determine the maximum acceptable sound power levels. This must be done for each octave band.

Table 15. Environmental Adjustment Factor

Octave Band 2 3 4 5 6 7 8

Env Factor 2 1 0 0 0 0 0

Table 16. Ceiling/Space Effect (Table D14, AHRI Standard 885)

Frequency 125 250 500 1K 2K 4KOctave Band 2 3 4 5 6 7Mineral Fiber Tile Ceiling ⅝” 16 18 20 26 31 36

Glass Fiber Tile Ceiling ⅝” 16 15 17 17 18 19

Solid Gypsum Board ⅝” 23 27 27 29 29 30

AHRI STANDARD 885


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Discharge sound (sometimes called airborne sound) is the sound that travels down the duct and discharges into the room along with the conditioned air. The procedure for determining the maximum acceptable discharge sound power levels requires the addition of the space effect, end reflection, duct insertion, flow division (or branch power division) and elbow and tees to the maximum acceptable room sound pressure levels. If more than one outlet supplies air to a room, separate evaluations should occur for each discharge path. This is done for each octave band.

Space Effect. The discharge sound space effect is determined in the same manner as the radiated sound space effect. The sound source in this case might be the outlet (i.e., grille or diffuser) supplying air to the space, or may be sound from an upstream noise source (damper or fan) which passes through the outlets, or a sum of both Table 12.

End Reflection. When the area across the airstream expands suddenly as the duct work terminates or ends at the outlet to the occupied space, a significant amount of low frequency sound is reflected back into the duct work. This is called end reflection. The amount of end reflection varies based on the inlet size and type of duct.

Duct Insertion Loss. The addition of lined duct work results in significant attenuation of higher frequency sound. The amount of attenuation varies with duct size and lining thickness. AHRI Standard 885 contains several tables helpful in determining the appropriate attenuation values. Each section of duct work inserted downstream of the terminal must be evaluated. For example, one might have separate duct insertion attenuation values for straight lined discharge duct, branch duct (if lined) and flex duct from the branch to the outlet. The AHRI Standard, as well as the ASHRAE Handbook, provide tables of insertion loss per foot of duct based on inside duct dimensions.

While lined duct factors are available in both ASHRAE and AHRI documents, flexible duct insertion loss data is available only from manufacturers or as found in the AHRI Standard 885. Table 20 is the flexible duct insertion loss data from AHRI Standard 885.

Two tables are provided here for rectangular and round lined duct from the AHRI Standard.

Table 17. End Reflection dB(Table D13, AHRI Standard 885)

Eq. Dia. or Duct Width

Octave Band

1 2 3 4 5 6 7 8

6” 18 12 7 3 1 0 0 0

8” 16 10 5 2 1 0 0 0

10” 14 8 4 1 0 0 0 0

12” 12 7 3 1 0 0 0 0

16” 10 5 2 1 0 0 0 0

24” 7 3 1 0 0 0 0 0

Table 18. Round 1-inch Lined Spiral Duct, dB / ft. (Table D7, AHRI Standard 885)

Duct Diameter

Octave Band2 3 4 5 6 7

6” 0.59 0.93 1.53 2.17 2.31 2.04

12” 0.46 0.81 1.45 2.18 1.91 1.48

24” 0.25 0.57 1.28 1.71 1.24 0.85

48” 0 0.18 0.63 0.26 0.34 0.45

Table 19. Rectangular, 1-inch Lined Duct, dB / ft. (Table D8, AHRI Standard 885)

Duct Dimension

Octave Band2 3 4 5 6 7

6” x 6” 0.6 1.5 2.7 5.8 7.4 4.3

12” x 12” 0.4 0.8 1.9 4 4.1 2.8

24” x 24” 0.2 0.5 1.4 2.8 2.2 1.8

48” x 48” 0.1 0.3 1 2 1.2 1.2

Table 20. Flexible Duct Insertion Loss, dB(Table D9, AHRI Standard 885)

Duct Diameter Inches

Duct Length Feet

Insertion Loss, dB

Octave Bands

2 3 4 5 6 7

4

10 9 9 27 32 38 24

5 6 5 16 23 27 18

3 4 4 12 19 23 15

5

10 9 12 28 32 37 23

5 5 7 17 22 25 16

3 4 5 13 18 21 13

6

10 9 15 28 32 35 22

5 5 9 18 21 24 15

3 4 6 13 16 19 11

8

10 9 18 29 31 32 20

5 5 10 18 19 21 12

3 3 7 14 14 16 8

10

10 8 19 28 30 29 18

5 4 11 18 17 18 9

3 3 7 14 11 13 6

12

10 7 17 26 28 26 15

5 3 9 16 15 15 7

3 2 6 12 9 11 4

14

10 5 13 23 25 23 12

5 2 7 14 13 13 6

3 1 4 10 8 9 4

16

10 3 7 19 23 20 8

5 1 2 11 11 11 5

3 0 0 8 7 8 4

DISCHARGE SOUND POWER LEVELS


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Flow Division. When the airstream is divided, the sound carried in each downstream branch is less than the sound upstream of the branch take-off. This shows the percent of total airflow carried by the branch. The appropriate level of attenuation can then be determined from Table 21.

Elbows and Tees. A certain amount of attenuation of higher frequency sound is gained when an airstream enters an elbow or tee duct connection. If the elbow is round and unlined, the attenuation is considered by AHRI Standard 885 to be negligible. Attenuation of rectangular tees is determined by treating the tee as two elbows placed side by side.

ACCEPTABLE TOTAL SOUND IN A SPACEOnce the Radiated and Discharge sound pressure paths and effects are known, the resulting room sound level can be evaluated. Other factors may play a part in determining the final room sound levels. All these factors must be included to achieve an accurate prediction or analysis. The results may be very complex.

In the example here, many paths are shown. In practice, only a couple are significant, but changes in designs may make a one time insignificant path become predominant. For example, should duct lining be eliminated and no flex duct employed, discharge sound may be much more important than radiated, the usual acoustical problem. Poor duct design may cause duct breakout to be the highest sound heard in the space.

The AHRI Standard 885 Standard provides guidance on all the possible paths. Not shown here is background sound, which is often at an NC-30 or greater in occupied spaces.

Once all of the attenuation factors have been determined, they are added to the sound pressure level to determine the maximum acceptable sound power levels. This must be done for each octave band, and again the Environmental Adjustment factor must be added.

LW DIS = LP + S + ER + I + D + T/E + Env

where: LW DIS = Discharge Sound Power Level LP = Sound Pressure Level S = Space Effect ER = End Reflection I = Duct Insertion D = Flow Division T/E = Tee/Elbow Env = Environmental Factor

Approximate Attenuation of 90° Elbows without Turning Vanes

Table 21. Duct Splits, dB

% of Total Air Flow 5 10 15 20 30 40 50 60

Attenuation 31 10 8 7 5 4 3 1

Table 22. Lined Rectangular, dB (Table D12, AHRI Standard 885)

Duct Width

Octave Band

2 3 4 5 6 7

5-10 0 0 1 6 11 10

11-20 1 6 6 11 10 10

21-40 6 6 11 10 10 10

41-80 6 11 10 10 10 10

Table 23. Circular with Lining Ahead or Behind Elbow, dB (Table D10, AHRI Standard 885)

Duct WidthOctave Band

2 3 4 5 6 7

5-10 0 0 1 2 3 3

11-20 1 2 2 3 3 3

21-40 2 2 3 3 3 3

41-80 2 3 3 3 3 3

Figure 132. Fan Powered Terminal or Induction Terminal - Summary Calculation, Sound Sources and Paths

Discharge Sound Power Levels (continued)

= Flex Duct Breakout

= Duct Breakout

2 34

5

6

= Casing Radiated & Inlet

= Distribution Duct Breakout

= Discharge1

= Outlet Generated Sound

1

2

3

4

5

6

Sound Pressure L p


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All the sound paths must be combined to predict the room sound level.

When combining path elements, the math is done using log addition, not algebraically. Logarithmic (log) addition requires taking the antilog of the dB in each band, adding them together, then taking the log of the answer. While this sounds complicated, (Figure 133) here shows an easier way of estimating the result.

More importantly, most people cannot differentiate between two sources which differ by less than 3 dB. If the background sound is an NC-35, and the device in question is predicted at an NC-35, it is likely that the space will be at an NC-38 (although this is dependent on which octave bands are critical), but most people cannot hear the difference.

In a similar manner, sound from VAV boxes and diffusers combine to create the room sound pressure level. Since they peak in different bands, however, they often complement each other. In many cases, a Series Fan Terminal will be predicted to have an NC-30 in a space, but when combined with an NC-40 diffuser, will result in a room sound pressure level of NC-40, which is optimum for providing speech privacy in open plan spaces.

Figure 133. Decibel Addition Example (Incoherent Sound)

Figure 134. Quiet VAV and Fan Terminal Recommended Installation

Acceptable Total Sound in a Space (continued)

To Add Two Decibel Values:

80 dB+ 74 dB

154 dB (Incorrect)

Difference in Values: 6 dB

From Chart: Add 1.0 dB to higher Value

80 dB+ 1 dB

81 dBDifference In Decibels

Between Two Values BeingAdded (dB)

Cor

rect

ion

To B

e A

dded

To

Hig

her V

alue

(dB

)

0

0.5

1

1.5

2

2.5

3

0 2 4 6 8 10

(Correct)

VAVUnit

Lined Sheet Metal Plenum(Max velocity 1,000 fpm)

Flexible DuctsTo Diffusers

Flexible ConnectorsFor Fan Powered Units

D

> 3 D

Ceiling

Maximize HeightAbove Ceiling

4' Min.


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A proper specification for acoustical performance in a space will limit the maximum sound generation for a product. This should be based on the desired resultant sound in the space and accepted and clearly stated sound path attenuations/reductions. AHRI Standard 885 provides a consistent method for accomplishing this task.

Example:A designer wants to achieve RC 40N for an open office to achieve an acceptable level of speech privacy. The office has a 9 ft. ceiling and a volume of 3,000 ft.3. The terminal will be located over the occupied space with 5 ft. of lined duct on the discharge.The lined discharge duct is 14” x 14” outside with one inch of insulation (12” x 12” internal cross section). This duct branches into an unlined trunk duct with two runs of lined flex duct taking off to the outlet. The flex duct is 6 ft. long and 8” in diameter. The terminal will supply 600 cfm with each diffuser taking 300 cfm. The ceiling is made of ⅝ inch mineral fiber tile with a 35 lb. ¼ ft.3 density. The diffuser is selected to provide an NC 40(N) at design flow.

DESIRED ROOM SOUND PRESSURE LEVELSFrom AHRI Standard 885 the appropriate sound pressure levels for an RC 40 can be determined. Since a Neutral Spectrum is desired, the Rumble Roof and Hiss Roof can be added to the spectra and still result in a neutral designation. The resultant maximum room sound pressure spectra are:

Figure 135. Sound Design Guidelines

Engineers can minimize the sound contribution of air terminals to an occupied space through good design practice.

• Whenever possible, terminals should be located over areas less sensitive to noise. This includes corridors, copy rooms, storage rooms, etc. Quiet air terminals facilitate the location of terminals over unoccupied space as with these units larger zones are possible resulting in fewer terminals. This also reduces first cost and improves energy efficiency.

• The use of lined duct work or manufacturers’ attenuators downstream of air terminals can help attenuate higher frequency discharge sound. Flexible duct (used with moderation) is also an excellent attenuation element.

• Sound will be reduced when appropriate fan speed controllers are used to reduce fan rpm rather than using mechanical devices to restrict airflow. This form of motor control is often more energy efficient.

• The air terminal and the return air grille location should be separated as far as possible. Radiated sound can travel directly from the terminal through the return air grille without the benefit of ceiling attenuation.

• Designing systems to operate at low supply air static pressure will reduce the generated sound level. This will also provide more energy efficient operation and allow the central fan to be downsized.

• Sharp edges and transitions in the duct design should be minimized to reduce turbulent airflow and its resulting sound contribution.

Table 24. Maximum LP for (RC 40N), dB

Octave Band 2 3 4 5 6 7

LP (RC 40N) 55 50 45 40 35 30

Rumble Roof 5 5 0 0 0 0

Hiss Roof 0 0 0 0 0 3Max Room Sound Pressure

60 55 45 40 35 33

MAXIMUM SOUND POWER LEVELS FOR MANUFACTURERS’ DATA

Maximum Sound Power Levels


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The ceiling/space effect may be determined from Table 16, page B33.

The maximum acceptable radiated sound power levels are determined by adding all the factors, including the environmental factor. For example, in the 2nd band:

LW RAD 2 = 55 (LP) + 16 (C/S) + 2 (Env) = 73 (Octave band 2. LW in bands 3-7 is calculated in a similar manner.)

For all bands, the following table results in a maximum allowed sound power, per AHRI Standard 880, to achieve an RC 40N:

For this specification to be compared evenly against all manufacturers, the environmental adjustment factor for manufacturers using a free field calibration Reference Sound Source (RSS), as required in AHRI Standard 880, has been subtracted from the appropriate manufacturer’s data or added to the maximum acceptable sound power levels. If data is tested in another method, the appropriateness of the environmental factor must be understood and properly applied.

DISCHARGE SOUND POWER LEVEL SPECIFICATIONSThe room absorption is determined by the space effect table. With a 10 ft. ceiling and the terminal located a few feet away from the receiver and a room volume of 3,000 ft.3. The effect varies with the octave band. The space effect will be obtained from Table 12, page B29.

End reflection is based on Table 17, page B34, and an 8-inch duct connection.

Duct insertion for 5 ft. of 8-inch lined flex duct can be taken from Table 20, page B34.

No insertion value will be gained from the unlined trunk duct.

Flow division based on a 50 percent split (300 cfm / 600 cfm) can be taken from Table 21, page B35.

The rectangular tee attenuation can be taken from Table 23, page B35.

Duct insertion for the 5 ft. of rectangular discharge duct can be taken from the AHRI Standard 885, or from Table 19, page B34 in this case.

As with radiated sound, the environmental adjustment factor for manufacturers using a free field calibration RSS, as required in AHRI Standard 880, has been subtracted from the appropriate manufacturer’s data or added to the maximum acceptable sound power levels. If data is tested in another method, the appropriateness of the environmental factor must be understood and properly applied.

Octave Band Mineral Fiber 2 3 4 5 6 7

Tile ⅝” 35#/Ft3 16 18 20 26 31 36

Table 25. Allowed Sound Power Maximums (AHRI Standard 880)

Octave Band 2 3 4 5 6 7Lp (RC 40N) 60 55 45 40 35 33C/S (Table 16) 16 18 20 26 31 36

Env Effect (Table 15) 2 1 0 0 0 0

Lw 78 74 65 66 66 69

Octave Band 2 3 4 5 6 7Space Effect 9 10 10 11 12 13

Octave Band 2 3 4 5 6 7End Reflection 10 5 2 1 0 0

Octave Band 2 3 4 5 6 7Flex Insertion Loss 5 10 18 19 21 12


Flow Division 3 3 3 3 3 3


Tee Attenuation 0 0 1 2 3 3


Duct Ins. Loss 2 4 10 20 21 14


LP (RC 40N) 60 55 45 40 35 33


Space (Table 12) 9 10 10 11 12 13

End Ref (Table 17) 10 5 2 1 0 0

FLEX (Table 20) 5 10 18 19 21 12

Flow Div (Table 21) 3 3 3 3 3 3

Elbow & Tee (Table 23) 0 0 1 2 3 3

Rect Duct (Table 19) 2 4 10 20 21 14

Lw 91 88 89 96 95 78

RADIATED SOUND POWER LEVEL SPECIFICATIONS

Radiated Sound Power Levels


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It can be seen that discharge sound is not likely to be a problem, especially in the mid-frequencies. If duct lining is eliminated, however, the maximum allowable power is reduced by the duct insertion loss data on the previous page. If flex duct is disallowed, the maximum sound power allowed is further decreased by flex insertion loss. This can result in the following requirement for the unit:

In many cases, this is a borderline acceptable case for many unit sizes and flow rates, especially in smaller rooms with RC 35N requirements. Flexible duct can be included as a solution to discharge noise in these cases.

DIFFUSER SPECIFICATIONSDiffusers are commonly specified and reported in NC, rather than RC. In most cases, there is no difference between NC and RC for diffusers as they usually peak in the 500-2,000 Hz region, and the resultant numerical specification is the same for both NC and RC. Diffuser NC ratings commonly subtract 10 dB from measured sound power levels in all bands to account for room attenuation. As described earlier, this will be a valid assumption for a number of combinations of room volume and distance to the source. While an ideal specification will be based on octave band sound levels, these are seldom available for diffusers, and so the NC rating must be used. For a close approximation of diffuser sound power when only NC is known, one can assume that the sound power for the diffuser in the 5th octave band (1,000 Hz) is equal to the reported NC plus 10 dB, the 4th band (500 Hz) is 3 greater than this, and the 6th band (2000 Hz) is 5 less. This will be suitable for most applications.

The room sound pressure level requirements should be based on the resultant desired acoustical environment. As the only attenuation element for diffusers is the room effect, this should be the primary attenuation path.

Diffusers, moreover, have typically been tested in the same facilities as VAV terminals, with the same reference sound source, and therefore the AHRI Standard 885 Environmental Effect must be included as well. The following is a proposed procedure for determining the Diffuser NC requirement based on an RC analysis:

Steps:1. Determine the desired RC level for the space. This is

the sound pressure level requirement in the 5th band.

2. Determine the room effect in the 5th (1,000 Hz) band, based on room volume and distance to the diffuser from the observer. Add this to the RC number.

3. Subtract 10 dB from the result in Step 2. This is the required diffuser NC.


No Lining/Flex 83 74 61 56 53 52

Figure 136. Fan Powered Terminal or Induction Terminal - Summary Calculation, Sound Sources and Paths

Discharge Attenuation Elements

Octave Band Center Frequency, Hz

dB

RC 40

Lp (RC 40N)Rumble Roof

Hiss Roof

Env Effect

Space

End Ref

Flex

Flow Div

Elbow & Tee

Rect Duct

125 250 500 1K 2K 4K

Discharge Sound Power Level (continued)


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Example: The open office from the previous example is used here. The room is large and speech privacy is desired, requiring an RC 40N specification. Steps:

1. From AHRI Standard 885 (or ASHRAE) the appropriate sound pressure levels for an RC 40 can be determined. In the 5th band the RC is = to the sound pressure level requirement, 40 dB.

2. The room absorption is determined from the space effect equation. With a 10 ft. ceiling and the terminal located a few feet away from the receiver, the distance variable will equal 10 ft. The volume of the room is 3,000 ft.3. The frequency varies with the octave band. The space effect in the 5th band will be obtained from Table 12, page B29, and is = 11 dB. 40 + 11 = 51 dB.

3. Subtract 10dB = the diffuser’s NC requirement, or NC 41.

DETERMINING COMPLIANCE TO A SPECIFICATION When determining if a unit will meet a specification, it may be necessary to conduct a total room sound evaluation with multiple sound sources and multiple paths. These are added using logarithmic addition to determine total sound level. Once all path elements are identified, the noncritical paths can be determined using (Figure 133), page B36. Paths which are 10 dB or more below the loudest, in any given band, can usually be ignored.

VAV terminals, if evaluated by themselves, often result in an “R” classification because of the high mid-frequency absorption provided by lined and flexible duct. The diffuser, however, can overcome this apparent “Rumble” spectra by filling in the resultant sound with its high frequency sound generation. This results in an “N” rating, as required. Using the estimated sound power procedure from the Diffuser Specification section above, the diffuser’s contribution can be added (using log addition) to the VAV boxes sound pressure level, and a resultant sound pressure level classification developed. Example: A project engineer desires a space sound pressure level of RC 35(N) for a private office. He has selected a Titus TQS Fan Terminal, size 5-12, at 1500 cfm, with a design inlet pressure of 1 in. static pressure. From the sound tables for the product, the sound power levels for this unit and the reduction factors as in the previous example are shown above:

As the estimated room sound pressure level exceeds the rumble roof in the 2nd band, this unit must be classified RC 29 (R). If an RC 35 (N) diffuser is also supplied, however, the sound from it must be added to the terminals to get the room total sound level. Using the procedure described under “Diffuser Specification”, we can estimate the sound power level of an NC-35 diffuser:

When these are added, the resulting spectra is:

The RC = (41 + 36 + 29)/3 = 35. The rumble roof for this spectra is therefore:

Therefore the sound pressure level in the space is an RC 35 (N). This works because diffusers and VAV terminals seldom peak in the same frequencies, with diffusers being critical in the speech bands (500-2,000 Hz) and boxes producing the most sound in the 125-250 Hz region. When these two sounds combine in the space, they often complement each other, producing a full spectrum of sound and resulting in an “N” rating.


Unit Pwl 71 64 59 57 50 45

C/S (Table 16) 16 18 20 26 31 36


Estimated Room Spl 53 45 39 31 19 9


Diffuser Pwl 48 45 40

Space Effect (Table 12) 9 10 10 11 12 13


Estimated Diffuser Spl 38 34 28


Diffuser Pwl 38 34 28

Terminal Spl 53 45 39 31 19 9

Log Sum 53 45 41 36 29 9

Octave Band 2 3 4

Rumble Roof 55 50 40

Diffuser Specifications (continued)


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STANDARD ATTENUATIONS FOR TERMINAL UNIT APPLICATION DATA

All NC levels are estimated in accordance with AHRI Standard 885-2008 Appendix E. This standard recommends that all manufacturers use the same default attenuation factors when publishing application data.

RADIATED SOUND ATTENUATION

• Environmental Adjustment Factor per AHRI 885-2008, Table C1.• Ceiling/Space Effect for Ceiling Type 1 (5/8 in, 20 lb/ft3 mineral fiber tile) per AHRI 885-2008, Table D14.• Assumes 3 ft deep ceiling plenum with non-bounded sides, per AHRI 885-2008, Table E1.

DISCHARGE SOUND ATTENUATION (< 300 CFM)

• Environmental Adjustment Factor per AHRI 885-2008, Table C1.• Duct Lining for 5 ft of 8 x 8 in lined duct per AHRI 885-2008, Table D8.• End Reflection for 8 in termination per AHRI 885-2008, Table D13.• Flex Duct for 5 ft of 8 in vinyl core flex duct per AHRI 885-2008, Table D9.• Space Effect for a 2500 ft3 room, 5 ft from source per AHRI 885-2008, Table D16.• Sound Power Division based on 10*log of the number of rooms served (1).

DISCHARGE SOUND ATTENUATION (300-700 CFM)

• Environmental Adjustment Factor per AHRI 885-2008, Table C1.• Duct Lining for 5 ft of 12 x 12 in lined duct per AHRI 885-2008, Table D8.• End Reflection for 8 in termination per AHRI 885-2008, Table D13.

Octave Bands

125 Hz 250 Hz 500 Hz 1000 Hz 2000 Hz 4000 HzEnvironmental Effect 2 1 0 0 0 0Duct Lining 2 6 12 25 29 18End Reflection 10 5 2 1 0 0Flex Duct 5 10 18 19 21 12Space Effect 5 6 7 8 9 10Sound Power Division 0 0 0 0 0 0Total Attenuation, dB 24 28 39 53 59 40

Octave Bands


Octave Bands

125 Hz 250 Hz 500 Hz 1000 Hz 2000 Hz 4000 HzEnvironmental Effect 2 1 0 0 0 0Ceiling/Space Effect 16 18 20 26 31 36Total Attenuation, dB 18 19 20 26 31 36

Standard Attenuations

STANDARD ATTENTUATIONS


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Octave Bands


• Flex Duct for 5 ft of 8 in vinyl core flex duct per AHRI 885-2008, Table D9.• Space Effect for a 2500 ft3 room, 5 ft from source per AHRI 885-2008, Table D16.• Sound Power Division based on 10*log of the number of rooms served (2).

DISCHARGE SOUND ATTENUATION (>700 CFM)

• Environmental Adjustment Factor per AHRI 885-2008, Table C1.• Duct Lining for 5 ft of 15 x 15 in lined duct per AHRI 885-2008, Table D8.• End Reflection for 8 in termination per AHRI 885-2008, Table D13.• Flex Duct for 5 ft of 8 in vinyl core flex duct per AHRI 885-2008, Table D9.• Space Effect for a 2500 ft3 room, 5 ft from source per AHRI 885-2008, Table D16.• Sound Power Division based on 10*log of the number of rooms served (3).

Standard Attenuations (continued)

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REFERENCES

TERMINAL UNITS

AHRI 220-2007 “Reverberation Room Qualification and Testing Procedures for Determining Sound Power of HVAC Equipment”This standard provides the methodology for the determination of sound power levels of noise sources that emit broadband sound and/or discrete frequency sounds/Tones in reverberation rooms.

AHRI 880-2011 “Performance Rating of Air Terminals”The purpose of this standard is to establish for air terminals: definitions; classifications; test requirements; rating requirements; minimum data requirements for published ratings; marking and nameplate data and conformance conditions. This standard applies to air control devices used in air distribution systems.

AHRI 885-2008 “Procedure for Estimating Occupied Space Sound Levels in the Application of Air Terminals and Air Outlets”This standard provides a consistent industry-accepted method for estimating sound pressure levels in a conditioned space for the application of air terminals and air outlets. Air terminals, air outlets, and the low pressure ductwork which connects them are considered sound sources and are the subjects of this standard. The method described in this standard can be used to identify acoustically critical paths in the system design. The design effects of inserting alternative components and changes in the system can be evaluated.

ANSI/AHRI 250-2008 “Performance and Calibration of Reference Sound Sources”This standard applies to all reference sound sources (RSS’s) used in conjunction with AHRI sound rating standards and covers the one-third octave band frequency range from 50 to 10,000 Hz.

ANSI/AHRI 280-2008 “Requirements for the Qualification of Reverberation Rooms in the 63Hz Octave Band”This standard applies to products rated in the 63 Hz octave band (50, 63 and 80 Hz one-third octave bands) where the sound power is determined from measurements made in a reverberation room by using the comparison method as specified per ANSI Standard S12.51/ISO: 3741.

ASHRAE Standard 130-2008 “Methods of Testing Air Terminal Units”First published in 1996 and reaffirmed in 2006, Standard 130 specifies instrumentation and facilities, test installation methods, and procedures for determining the capacity and related performance of constant-volume and variable-volume air terminal units. The standard is classified as an ASHRAE standard method of measurement. This revision of the standard includes updates and revisions to all parts of the standard, including its title, purpose, and scope. It updates definitions, adds modulating diffusers, redefines airflow sensor performance testing, and adds a method to determine the power factor. New appendices contain some material that was formerly in the body of the standard and some new reference material. This standard is required for compliance with AHRI Standard 880.

References


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GLOS

SARY

Breakout - Sound which passes through the walls of a duct or device and is passed directly to an observer.

Casing Radiated Sound - A type of Breakout Sound which passes through the walls of a device. In some cases, it includes induction port radiated sound as well.

dBA - A single number rating of a broadband spectrum. Typically used to rate outdoor noise levels. Not practical for use in rating indoor sound levels as it bears little relevance to occupants’ needs.

Discharge Sound - Airborne sound which is transmitted through ductwork from a noise source to an observer.

End Reflection - The reduction in sound, typically in low frequencies only, resulting from a rapid change in the shape or size of a duct or a duct termination.

Environmental Adjustment Factor (Environmental Effect) - A correction required to accurately use data obtained in accordance with AHRI Standard 880, it corrects for a calibration difference between the Free Field calibration of the Reference Sound Source and the reverberant field in which it is used.

Incoherent Sound - Sound which is broadband and contains no repeating fluctuations.

Induction Port Radiated Sound - That sound which passes out from the induction port of a VAV device. In practice it is impossible to differentiate from Casing Radiated sound, and is reported as a combined value under Casing Radiated sound levels.

Insertion Loss - The reduction in sound resulting from inserting an attenuation device, such as a section of lined duct. The difference before and after the insertion of such a device is the insertion loss.

Multiple Outlet Effect - When an airstream is split, the sound traveling in the duct is also reduced, typically in proportion to the percent of airflow in each duct. The amount of reduction must be calculated logarithmically, not arithmetically, however.

Pure Tone - A sound spectrum which is very concentrated in a narrow band.

Radiated Sound - Sound which travels from the source to the observer in a direct path, outside ductwork.

Room Effect - Typically a 10 dB reduction in all bands, this is the assumed value for attenuation of a room. In practice, it is reasonable for diffusers which peak in the mid-frequencies, but not necessarily for VAV terminals, which peak in lower frequencies.

Sound Power - The energy released as acoustic energy by a device. It is measured indirectly by one of several methods. It is reported as dB (the log base 10 of the value) referenced to a base power level, typically 10-12 watts. It is reported by frequency, typically in octave bands, although sometimes in 1/3 octave bands.

Sound Pressure - The directly measurable fluctuation in pressure, heard as sound. The sound pressure is reported in dB (the log base 10 of the value), referenced to a pressure 0.0001 microbars. It is reported by frequency, typically in octave bands, although sometimes in 1/3 octave bands.

Space Effect - The calculated attenuation of a space which is different

in each frequency band, and is a function of room volume and distance from the source.

Terminals (Boxes) - Devices which vary the flow through a duct with a moveable damper. They typically have a control device to vary the flow in response to a control signal. In some cases, the term terminals can also mean boxes and GRDs. In those cases, “Boxes” are referred to as Air Control Devices (ACDs). The terminology is inconsistent throughout the industry.

Glossary


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INDEX

AAir Leakage - Terminal Units ............................................. B24Air Velocity ........................................................................ B25 ANSI Standards ................................................................. B41AHRI Standards ................................................................. B41 ASHARE 70-2006 ........................................................ B4, B41ASHRAE Standards............................................................ B41

BBasic Terminal Sizing (by capacity tables) ........................ B18

CControls Classified by Power Source ................................... B7Controls Reaction to Duct Pressure .................................... B7Control Operation ................................................................ B9

DDirect Digital Control ......................................................... B16

EECM Motors - Fan Powered Terminals .............................. B15Equations and Definitions - Terminal Units ....................... B27Errors in Terminal Application ........................................... B21 FFan Laws, Fan Performance - Terminal Units .................... B26Fan Speed Control ..................................................... B13, B26

IISO ..................................................................................... B42

MMultiple Outlets and NC .................................................... B32

OOutlet NC Level and Space NC ........................................... B28

SSizing Fan Terminals .................................................. B18–B20

TTerminals, Controls, and Accessories .................................. B6Terminal Installation Techniques ....................................... B23

Engineering Guidelines Index

Engineering Guidelines - TerminalsNotes

titus for vav

Documents

series type constant

contentstypes of terminals

parallel type variable

pneumatic thermostat

temperature fan terminals

low profile

pressure controls

heating coil