02 acoustics.in.air.conditioning

ApplicationsEngineering Manual

Acoustics inAir ConditioningFundamental Concepts for HVAC System Design

April 2006 ISS-APM001-EN

Acoustics in Air ConditioningFundamental Concepts for HVAC System Design

Dave Guckelberger, applications engineer

Brenda Bradley, information designer

© 2006 American Standard All rights reserved Acoustics in Air Conditioning (Fundamental Concepts) ISS-APM001-EN

Trane, in proposing these system design and application concepts, assumes no responsibility for the performance or desirability of any resulting system design. Design of the HVAC system is the prerogative and responsibility of the engineering professional.

“Trane” and the Trane logo are registered trademarks, and TAP is a trademark, of Trane, which is a business of American Standard Companies.

Preface

As a leading HVAC manufacturer, we believe that it is our responsibility to serve the building industry by regularly disseminating information gathered through laboratory research, testing programs, and practical experience. Trane publishes a variety of educational materials for this purpose. Applications engineering manuals, such as this document, can serve as comprehensive references for professionals who design building comfort systems.

This manual focuses on acoustics. It summarizes basic terminology, the characteristics of sound and how we respond to it, acoustical design goals, and how to predict the acoustical effect of HVAC equipment. A list of references identifies sources for further information.

We encourage you to familiarize yourself with the contents of this manual and to review the appropriate sections when designing comfort-system applications that must meet specific acoustical requirements.

ISS-APM001-EN Acoustics in Air Conditioning (Fundamental Concepts) iii

Contents

Introduction ..................................................................................... 1

Characteristics of Sound ............................................................. 2Wave, frequency, and amplitude ...................................................... 2Magnitude ........................................................................................ 3Pure tones and broadband sound ..................................................... 6Octave bands ................................................................................... 7

Evaluation Methods ...................................................................... 8Human hearing ................................................................................. 8Single-number rating methods ........................................................ 10

A-B-C weighting networks .......................................................... 11Noise criteria (NC) curves ........................................................... 12Room criteria (RC) curves ........................................................... 14Sone scale .................................................................................. 18Phon scale .................................................................................. 18

Octave-band rating method ............................................................. 19

Rating Equipment Sound ......................................................... 20Sound fields ................................................................................... 20

Free field .................................................................................... 20Acoustic nearfield ...................................................................... 21Reverberant field ....................................................................... 22Semireverberant field ................................................................ 22

Rating environments ...................................................................... 23Reverberant room ...................................................................... 23Free field (with flat reflecting plane) .......................................... 23

Rating methods .............................................................................. 24Substitution ................................................................................ 24One-third-octave-band and full-spectrum charts ........................ 25

Industry standards ......................................................................... 26

Environmental Effect of Equipment Sound ...................... 30Setting an acoustical target ............................................................ 31

“High-quality” ambient sound ................................................... 32Predicting ambient sound ............................................................... 33

Acoustical characteristics of materials ....................................... 33

Creating an

Acoustical Model ......................................................................... 39Modeling terminology .................................................................... 40

Attenuation ................................................................................ 40Regenerated sound ................................................................... 40End reflection ............................................................................ 40Receiver sound correction ......................................................... 42

iv Acoustics in Air Conditioning (Fundamental Concepts) ISS-APM001-EN

Contents

Modeling tools ............................................................................... 42Algorithms ................................................................................. 42Analysis software ....................................................................... 43Transfer functions ...................................................................... 43

Identifying sound paths .................................................................. 43An example of sound-path modeling .............................................. 45

Supply airborne path .................................................................. 46Supply breakout path ................................................................. 48Return airborne path .................................................................. 50Wall transmission path .............................................................. 52Combining the path results ....................................................... 53Redesigning a sound path ......................................................... 54

Good Acoustical Design ............................................................ 57Appropriate acoustical target(s) ...................................................... 57Accurate sound data ...................................................................... 59Acoustical modeling ....................................................................... 59Conscientious implementation ....................................................... 60

Acquiring the right equipment ................................................... 60

Appendix A: Working with Decibels ................................... 62Adding decibel values ..................................................................... 62Subtracting decibel values .............................................................. 63Averaging decibel values ................................................................ 63

Appendix B: Fan-Generated Sound ..................................... 65Factors that affect sound generation .............................................. 65

Air velocity ................................................................................. 65Blade-pass frequency ................................................................ 65Operating point .......................................................................... 66Flow disturbances ...................................................................... 66

Shortcomings of acoustical rules of thumb .................................... 66

Glossary ........................................................................................... 68

References ....................................................................................... 74

Index ................................................................................................. 75

ISS-APM001-EN Acoustics in Air Conditioning (Fundamental Concepts) 1

Introduction

The term acoustics describes the “scientific study of sound, especially its generation, transmission, and reception”; it also describes “the total effect of sound, especially as produced in an enclosed space.”1 Acoustics affects how each of us judges and responds to the comfort of the immediate environment. It can enhance or impair our ability to perform and, especially in the case of commercial buildings, can affect the value (marketability and tenancy) of the space.

The sound that we detect at a particular location is really the sum of the sounds emanating from various sources indoors and/or outdoors. The degree to which a given piece of HVAC equipment affects the sound heard in a specific location depends on the strength of the sound source and the environmental effects on the sound as it travels from source to hearer.

In the context of designing an HVAC system, the primary acoustical goal is to achieve a level of background sound that is unobtrusive and that does not interfere with the activity requirements of the space. To accomplish this goal, building professionals must understand how the design, installation, and use of HVAC equipment contributes to the sounds that ultimately will be heard within each occupied space. To aid that understanding, this manual:

• Introduces the fundamentals of sound

• Explains commonly used terminology

• Describes relevant industry standards and methods for obtaining accurate sound data

• Identifies predictive modeling techniques

Given the complex effects of sound within a built environment, this manual focuses on indoor acoustics. Outdoor considerations are addressed where appropriate.

1 The American Heritage Dictionary of the English Language, third edition, 1996 (Boston: Houghton Mifflin Company), 15.

2 Acoustics in Air Conditioning (Fundamental Concepts) ISS-APM001-EN

Characteristics of Sound

For many, the terminology associated with acoustics is daunting as well as unfamiliar. It also is unavoidable if you wish to design building comfort systems that satisfy specific acoustical requirements.

Wave, frequency, and amplitude

Sound describes the audible emissions that result from the vibration of molecules within an elastic medium. In the context of a building, the “elastic medium” is either air or the building structure itself. (To be audible, vibration conveyed by the structure first must become airborne.) Sound is judged to be noise when it is unwanted, obtrusive, and objectionable—for example, when it interferes with speech, concentration, or sleep.

Sound is a phenomenon of pressure. A vibrating body (the sound source) alternately compresses and expands, or “rarefies,” the adjacent air molecules. As energy transfers from one air molecule to the next, airborne sound spreads away from the vibrating body. When represented graphically, the pressure fluctuations resulting from this compression and rarefaction of air molecules (Figure 1) resemble harmonic sine waves. The amplitude of these waves indicates how much the pressure fluctuates from atmospheric pressure (the mean); the higher the amplitude, the louder the sound.

The transfer of energy takes time. Each complete sequence of motion (compression and rarefaction) constitutes a cycle, and the time required to complete one cycle is the cycle period, T. The frequency, ƒ, of the periodic motion—which is measured in hertz (Hz)—is the number of cycles that occurs in one second (s):

Sound transmission is a physical property of the transferring medium; it is not affected by the amplitude, frequency, or wavelength of the sound. Sound travels at about 4,700 ft/s (1,433 m/s) in water, at 16,500 ft/s (5,029 m/s) in steel, and at 13,000 ft/s (3,962 m/s) in wood.

Given the relatively narrow temperature range encountered in HVAC applications, sound transmission through air is considered as a constant 1,127 ft/s at 68°F (344 m/s at 20°C). Notice that this value is based on an air temperature. That’s because the speed of sound transmission through a gaseous medium varies slightly with dry-bulb, wet-bulb, and dew-point temperatures, and with relative humidity, enthalpy, and humidity ratio.

Figure 1. Wave, frequency, and amplitude of airborne sound

frequency cycles1 second--------------------------= 1 Hz 1 cycle

1 s--------------------=∴



The following equation describes the relationship between cycle period and cycle frequency. If the period is 0.005 second (one complete cycle occurs every 0.005 second), then the frequency is 200 Hz:

Wavelength, λ, describes the distance covered by one complete cycle of a sound wave (Figure 2). The following equation illustrates how the speed, ν, and frequency of a sound determine its wavelength:

If the frequency of an airborne sound is 200 Hz, then the wavelength of that sound is calculated by:

Magnitude

Sound occurs at a seemingly infinite range of magnitudes, or levels. The loudest sound that the human ear can hear (without damage due to prolonged exposure) is about 10 million times greater than the quietest perceptible sound. Such an extensive range makes an arithmetic scale cumbersome, so a system of logarithmic ratios—called the decibel scale—is applied instead. (Other examples of logarithmic measurements include the Richter scale, which measures the energy of earthquakes, and the pH scale, which measures the concentration of hydrogen ions in liquids.)

The decibel scale provides a dimensionless measure of the magnitude of sound. A decibel, dB, is one tenth of a bel; and a bel is a calculated value that is ten times the logarithm to the base 10 (log10) of the measured quantity divided by the reference quantity:

Establishing the reference value and placing it in the denominator of the ratio makes it possible to compute the sound level, in decibels, for any value entered as the numerator. If the sample (measured) value equals the reference value, then the ratio is 1 and the log of 1 is 0. Thus, the reference value defines 0 dB for the scale. If the sample value is less than the reference value, then the ratio is less than one and the log is negative.

In acoustics, the decibel scale quantifies three aspects of sound: intensity, power, and pressure. The logarithmic reference value differs with each type

ƒ 1 cycleT

-------------------- ∴=

ƒ 1 cycle0.005 second---------------------------------------- 200 Hz= =

Figure 2. Wavelength

λνsound in air

ƒ---------------------------------=

λ 1,127 ft/s200 Hz 1/s( )--------------------------------------- 5.6 ft== λ 344 m/s

200 Hz 1/s( )--------------------------------------- 1.7 m==⎝ ⎠

⎛ ⎞

where,N = sample or measured valueNref = reference value

dB 10 log10N

N ref------------⎝ ⎠

⎛ ⎞×=



of sound level; therefore, it is important to specify the appropriate reference value to prevent confusion about the magnitude of the ratio.

Sound pressure level, Lp measures the pressure disturbance caused by sound waves as they pass through the atmosphere. Stated more simply, it describes what our ears hear and what sound meters measure. The extent of the pressure disturbance varies with the strength of the source, that is, with the amplitude of the sound waves. Sound pressure also is affected by the distance between the sound source and the listener (receiver), and by the immediate surroundings.

In an existing space, sound pressure is measured directly using a sound meter. The pressure reading is converted from pascals, Pa, to decibels using this equation:

In this case, the ratio is multiplied by 20 rather than 10 because sound power is proportional to the square of sound pressure. The log of a squared value is the same as multiplying the result of the log by 2; for example:

For a space not yet constructed, the sound pressure can be predicted by performing an acoustical analysis.

Sound power level, Lw, describes the acoustical energy—measured in watts, W—emitted by the sound source. It is not affected by distance nor by the

where,Lp = sound pressure level, in dBP = measured sound pressure, in PaPref = threshold of hearing, which is 20 μPa, or 2 × 10-5 Pa or 0.0002 microbar

Lp 20 log10P

Pref----------⎝ ⎠

⎛ ⎞×=

10 log10AB----⎝ ⎠

⎛ ⎞2× 2 10× log10× AB----⎝ ⎠

⎛ ⎞ 20 log10AB----⎝ ⎠⎛ ⎞×= =

“Illuminating” the difference between sound power and sound pressure

Think of sound power as analogous to the wattage rating of a light bulb; both characteristics describe a fixed amount of energy.

Sound pressure corresponds to the brightness in a particular part of the room; both attributes can be measured with meters and are affected by the immediate surroundings. In the case of light, the degree of brightness is more than a matter of bulb wattage. Brightness also is influenced by the distance between the bulb and the observer, the color of the room, the reflectiveness of the wall surface, and whether the bulb is covered

with a shade. All of these factors affect how much light reaches the receiver without affecting the wattage of the light bulb.

Similarly, sound pressure depends on more than the sound power (acoustical energy) emitted by the source. Once again, the distance between the source and the receiver is a factor, as is whether the room is carpeted or tiled, and furnished or bare. These factors (and others) determine how much sound reaches the receiver—but without affecting the wattage (sound power) of the sound source.



environment. If the acoustical energy of the source is known, then sound power is calculated as:

However, acoustical energy isn’t easy to measure, so sound power typically is calculated from measurements of sound pressure. (See “Rating methods: Substitution,” p. 24.)

Sound power ratings usually are determined by the equipment manufacturer, and should account for the modifying influences of the immediate environment as well as the character of the sound source. Industry standards define specific requirements for collecting and converting the necessary sound measurements to encourage uniformity and permit objective comparisons of equipment.

Sound intensity, LI, represents the average rate at which acoustical energy passes through a given area of the medium, perpendicular to the direction of transmission, per unit of time. Although seldom used directly as an acoustical criterion in HVAC applications, sound intensity helps to illustrate the relationship between sound power and sound pressure. The greater the amplitude of the acoustical vibrations (Figure 3), the greater the transmission rate of the acoustical energy and the more intense the sound wave is.

As the sound wave carries its energy through the transferring medium and away from the source, the intensity of the resulting pressure disturbance diminishes. Intensity varies inversely with the square of the distance from the source, assuming that the source radiates sound uniformly in all directions, because the sound wave spreads out over an ever-greater spherical surface area (Figure 4).

The logarithmic equation for determining the sound intensity level is:

where,Lw = sound power level, in dBW = acoustical power, in W, radiated by the sourceWref = threshold of hearing, which is 1 picowatt (pW) or 10-12 W

Lw 10 log10W

Wref-------------⎝ ⎠

⎛ ⎞×=

Figure 3. Amplitude and energy*

Figure 4. Intensity and distance*

*SOURCE: Adapted from “The Physics Classroom,” StudyWorks! Online; available at http://www.physicsclassroom.com/class/sound/u11l2b.html; accessed September 2005

LI 10 log10I

Iref---------⎝ ⎠

⎛ ⎞×=

where,LI = sound intensity level, in dBI = measured sample intensity, in W/m²Iref = threshold of hearing, which is 10-12 W/m²



Pure tones and broadband sound

The waveform in Figure 2 (p. 3) represents sound that occurs at a single frequency—a pure tone. But most of the sound we hear is of a broadband nature, which means that it consists of a mixture of sounds that are generated at the same time but at different frequencies, amplitudes, and phases (Figure 5). Many of the sounds that HVAC systems generate exhibit both broadband and tonal characteristics.

When the sound in a narrow band of frequencies is significantly greater than the sound at adjacent frequencies, it exhibits a tonal quality that makes it stand out from background sounds. (Figure 6 illustrates this characteristic; the graph was created by plotting the amplitudes of the sound waves in Figure 5 for each frequency.) Because tones are obtrusive, they also may be objectionable.

No discrete width or magnitude universally distinguishes a (narrowband) tone from the adjacent frequencies in broadband sound; but special-purpose definitions do exist to help minimize objectionable sound. For example, the sound ordinance for a city may state that a tone is present when “the 1/3-octave-band sound level in a band exceeds the arithmetic average of the sound in the two contiguous 1/3-octave bands by 5 dB or more.”

Don’t confuse pitch with frequency

Although “pitch” and “frequency” are often used interchangeably, these terms are not synonymous. Frequency is a quantifiable property of sound that is measured in cycles per second (Hz) and is independent of the sound pressure level. Pitch, on the other hand, is a subjective quality that is primarily determined by frequency but that also depends on sound pressure level and composition. Two tones of the same frequency but with different sound pressure levels will have different pitches.

Auditory pitch is not directly measurable; instead, various dimensionless scales are used to rank sounds from “low” to “high” based on a reference point. In the case of a piano keyboard, the notes in the musical scale are determined from the reference point of Middle C at 256 Hz.

The mel-scale is another means of ranking sound based on pitch. In this system, the reference point equates a 1,000 Hz tone at a sound pressure level of 60 dB re 0.0002 microbar (40 dB above the listener’s threshold) with a pitch of 1,000 mels.

Figure 5. Broadband sound

Figure 6. Broadband sound and tones



Octave bands

The frequencies of sound pressure waves cover a wide spectrum for humans. Detectable HVAC-related sounds can range from 20 Hz to 20,000 Hz (perhaps beyond), but the range of interest in most HVAC applications is somewhat narrower—45 Hz to 11,200 Hz. To understand how a sound will be perceived in a particular environment, it must be measured at each frequency; for an HVAC system, that represents 11,156 data points!

To make the amount of data more manageable, sound frequencies typically are divided into smaller ranges called octave bands. Each octave band represents a range of sound frequencies in which the highest frequency, ƒ2, is twice the value of the lowest frequency, ƒ1. Each octave band is identified by its center frequency, ƒc, which is calculated by finding the square root of the product of the highest and lowest frequencies:

Dividing the 45-to-11,200 Hz range in this fashion yields eight octave bands (Table 1). Each octave band compresses the range of frequencies between the upper and lower ends of the band into a single value, which is the logarithmic sum of the sound level at each frequency within the band. Although practical, logarithmically summing measurements of sound sacrifices valuable information about the “character” of the sound, such as whether or not it contains a tone.

Note: Depending on the application, it may be advantageous to measure and display the sound in each frequency over the entire range of frequencies studied. The result of such a study, called a “full spectrum analysis,” is plotted on a logarithmic scale similar to the example in Figure 7.

One-third-octave bands provide a useful middle ground between a full-octave-band analysis and a full-spectrum analysis. As the name implies, 1/3-octave bands divide each full octave into three parts. The upper cutoff frequency of each one-third octave is greater than the lower cutoff frequency by a factor of the cube root of 2 (approximately 1.2599). Compressing fewer frequencies into a single value means that measurements based on 1/3- octave bands are more likely to reveal the presence of tones in broadband sound (Figure 8).

The use of octave bands usually is sufficient for rating the acoustical environment in a given space. One-third-octave bands are more useful for product development and for troubleshooting acoustical problems. Although other divisions of octave bands exist and are used, full octaves and 1/3 octaves are most prevalent in HVAC- and building-related acoustics.

ƒc ƒ1 ƒ2×( )0.5 ƒ1 2 ƒ1×( )×[ ]0.5 ƒ1 20.5× ƒ1 1.414×= = = =

Table 1. Octave-band frequenciesa

Octave band

Included frequencies, Hz

Lowest, ƒ1 Center, ƒc Highest, ƒ2

— 11 16 22

— 22 31.5 45

1 45 63 90

2 90 125 180

3 180 250 355

4 355 500 710

5 710 1,000 1,400

6 1,400 2,000 2,800

7 2,800 4,000 5,600

8 5,600 8,000 11,200

a Occasionally, an acoustical analysis will include the two octave bands below the 63 Hz octave. To avoid confusion, refer to an octave band by its center frequency (“63 Hz,” for example) rather than its relative rank (“Octave 1” or “first octave”).

Figure 7. Full-octave bands

Figure 8. One-third-octave bands


EvaluationMethods

How we hear sound involves physiology and psychology as well as acoustics. Although the mechanics of the hearing process remain the same from one person to the next, individual responses vary.

Independently judging whether or not the background sound in an environment is acceptable is straightforward: it’s simply a matter of listening. Creating an acoustically acceptable environment is much more difficult. It entails anticipating the desired acoustical traits of the finished space, accurately predicting the acoustical effect of the HVAC system, and then devising appropriate specifications.

Quantitative methods of rating sound compensate for the subjectivity of hearing by providing an effective, unbiased system for predicting and evaluating sound. Several of the most common, single-number rating methods—A-B-C weighting networks, noise criteria (NC) curves, room criteria (RC) curves, sones, and phons—are discussed later in this section. All of them are based on how we perceive sound.

Human hearing

Electronic sound-measuring equipment provides an objective, repeatable evaluation of sound pressure and frequency. By contrast, the human ear responds to sensations of pitch (differences in frequency) and loudness (sound wave intensity). Not only does sensitivity vary by individual, but a highly arbitrary evaluation device is involved, too: the human brain.

The human ear (Figure 9) acts as an antenna; it converts sound energy to mechanical energy, then to a nerve impulse that is transmitted to the brain. The outer ear collects sound waves and channels them to the middle ear, where they cause the ear drum to vibrate. These vibrations pass to the fluid in the inner ear, which converts them to electrical impulses that travel to the

Figure 9. Human ear

Evaluation Methods


brain along the auditory nerve. The brain perceives the signal as sound, which it then analyzes and evaluates. Part of that evaluation is an assessment of loudness.

Although loudness is largely determined by the intensity (sound pressure disturbance) of the sound wave, it is influenced by various physiological and psychological factors. For example, our ears respond to frequency as well as to pressure; therefore, we will not perceive two sounds of the same intensity but different frequencies to be of equal loudness. Also, our ears tend to amplify sounds at frequencies ranging from 1000 Hz to 5000 Hz; sounds in this range will therefore seem louder than they actually are.

The equal loudness contours graphed in Figure 10 compare the sensitivity of the human ear to loudness based on sound pressure level (dB) and frequency (Hz). For example, the perceived loudness of a 60 dB sound at 100 Hz equals the perceived loudness of a 50 dB sound at 1000 Hz. Figure 10 also illustrates that the human ear does not respond linearly to changes in pressure and frequency. Each contour slants downward, indicating greater sensitivity as the frequency increases from 20 Hz to 200 Hz. The contours become flatter at sound pressures greater than 90 dB, indicating a more uniform response to “loud” sounds across this range of frequencies.

Perception is another important aspect of human hearing. We perceive a sound spectrum that contains a tone (p. 6) as “louder” than an equally loud spectrum without a tone; that’s because we involuntarily tend to focus on a sound after we notice it. If the distraction is unwanted, the sound becomes an annoyance.

These aural characteristics pose a significant challenge for those who are asked to correct a noisy environment. It is much easier to address acoustical considerations during the design and construction phases of a project than after the space is occupied.

Figure 10. Loudness contours


Evaluation Methods

Single-number rating methods

While the human ear interprets sound in terms of loudness and pitch, electronic sound-measuring equipment interprets it in terms of pressure and frequency. The effort to resolve the disparity between these interpretations has resulted in considerable research to develop a system of single-number descriptors that express both the intensity and quality of a sound. Such a system makes it possible to establish sound targets for different environments. These targets, in turn, enable building designers to specify appropriate acoustical requirements that can be verified through measurement. For example, a designer can specify that “the background sound level in the theater shall be X,” where X is a single-number descriptor that communicates the desired ambient quality.

The most frequently used single-number descriptors are A-B-C weighting networks, noise criteria (NC), and room criteria (RC); see Table 2. All three systems share a common problem, however: They unavoidably lose valuable information about the character or quality of sound. Each of these systems is based on a linear full-octave-band analysis, which can mask tones and does not provide information about fluctuating or pulsating sound. The process of converting from eight octave bands to a single number omits still more data. Sound meters average sound pressure over a defined period of time (10 seconds, for example); only the averages are used to determine the single-number rating.

Note: Pulsating sound can interfere with voice communication and is likely to be judged as annoying or objectionable. Unfortunately, averaged sound level readings are unlikely to reveal the presence of pulsating sound, which

Table 2. Comparison of single-number sound rating methodsa

a Adapted from Table 31 in Chapter 46, “Sound and Vibration Control,” of the 1999 ASHRAE Handbook–HVAC Applications

Method Overview

Considers speech interference?

Evaluates sound quality?

Components rated by this method

dBA • Determined using a sound level meter

• Does not assess sound quality

• Commonly used in noise ordinances for outdoor environments

Yes No • Cooling towers

• Water chillers

• Condensing units

NC • Evaluates HVAC components

• Does not evaluate low-frequency rumble

Yes No • Air terminals

• Diffusers

NCB • Evaluates HVAC components

• Provides limited assessment of sound quality

Yes Yes —

RC • Evaluates entire HVAC systems

• Unsuitable for evaluating HVAC components

• Provides limited diagnostic capability

Yes Yes —

RC Mark II • Improves diagnostic capability Yes Yes —

Consult the ASHRAE Handbook series—specifically, Fundamentals and HVAC Applications—for more information about the rating methods described here, or about rating methods developed since this writing.

Evaluation Methods


typically results from improperly operated equipment. Although it is difficult to predict, be sure to note the presence of pulsations when measuring sound in an existing space.

Despite these drawbacks, single-number rating methods are valuable tools for defining sound levels in a space and are widely used to specify acoustical design goals.

A-B-C weighting networks

One simple method for combining octave-band readings into a single-number descriptor is the A-, B-, or C-weighting network. The weighting curves (Figure 11) compensate for the ear’s varying sensitivity at different frequencies:

• “A” weighting, dBA, is appropriate for low-volume (quiet) sound pressures. It best approximates human hearing in the comfort range, where protection is unnecessary.

• “B” weighting, dBB, is applied to medium-volume sound levels.

• “C” weighting, dBC, is applied to high-volume (loud) sound levels where the ear’s sensitivity is relatively uniform.

To calculate an A-weighted value:

1 Obtain full-octave-band measurements of the actual sound pressure levels in the target space.

2 Using the A-weighting curve (Figure 11) as a reference, add or subtract the appropriate weighting value, in decibels, to the measured sound pressure for each octave band. Table 3 on page 12 provides an example.

3 Using logarithmic addition (see “Appendix A: Working with Decibels”), sum the adjusted sound pressure levels for all eight octave bands. The resulting value represents the overall sound pressure level in A-weighted decibels … in this case, 42 dBA.

Figure 11. A-B-C weighting curves


Evaluation Methods

Most sound meters can automatically calculate and display A-weighted sound values, which makes it easy to objectively verify acoustical performance. But the calculation process (Step 3, p. 11) obscures the relative magnitude of each octave band. Therefore, even if the target dBA level is achieved, an objectionable tonal quality or spectrum imbalance may exist.

“A” weighting often is used to define sound targets for outdoor environments. For example, local sound ordinances typically use dBA readings to regulate sound pressure levels at property lines. Hearing-related safety standards written by entities, such as the Occupational Safety and Health Organization (OSHA), commonly refer to A-weighted sound pressure levels when determining whether hearing protection is required in a certain environment. “A” weighting also is the basis for the classroom acoustics criteria for schools, which is set by ANSI/ASA Standard S12.60.

Note: Used correctly, “A” weighting always results in a single numeric value. Do not display sound levels by octave band as A-weighted values.

Noise criteria (NC) curves

Noise criteria, or NC, curves may be the most common single-number descriptor for rating sound pressure in indoor environments. Like the equal-loudness contours (Figure 10, p. 9) on which they’re based, each NC curve slopes downward to reflect the ear’s increasing sensitivity at higher frequencies.

Determining the NC value that corresponds to a full-octave-band set of sound pressure readings is straightforward:

1 On an NC chart, plot the sound pressure level for each octave band.

2 Identify the highest NC curve that intersects with the plotted sound pressure level. That curve identifies the NC rating.

Although NC curves are popular and easy to use, they do not account for the tonal nature and relative magnitude of each octave band. An example illustrates the effect of this shortcoming. The NC chart in Figure 12 plots the full-octave-band analysis (from Table 3) for an open-plan office. Notice that

Table 3. Correction factors to demonstrate the A-weighting network

Octave-bandcenter frequency, Hz

Actual sound pressure level, dB

A-weightingfactor, dB

Adjusted soundpressure level, dB

63 63 –26 37

125 52 –16 36

250 45 –9 36

500 38 –3 35

1000 31 +0 31

2000 24 +1 25

4000 16 +1 17

8000 10 +0 10

Overall A-weighted value: 42 dBA

“A” weighting is usually applied to sound pressure data; however, some ARI standards that govern equipment sound ratings permit the use of A-weighted sound power data. To avoid confusion with A-weighted sound pressure values, A-weighted sound power is sometimes expressed in bels rather than decibels. (One bel equals 10 decibels.)

Evaluation Methods


the 63 Hz octave band determines the NC 39 rating, and that the sound in the higher frequency ranges quickly drops off. According to the generally accepted design guidelines for background sound, NC 39 is acceptable for this type of environment. But will the tenants agree?

In this particular example, sound generated by the air handler travels through the ductwork and radiates, or “breaks out,” through the duct wall into the office area. The NC 39 level was achieved by adding two layers of plasterboard to the duct exterior, which sufficiently blocked the low-frequency sound. It also overattenuated the more easily quieted sound at higher frequencies. Although an objective analysis deems the resulting NC 39 sound level as acceptable, most listeners would (unhappily) detect the rumble produced by the unbalanced spectrum.

Before leaving this example, it is interesting to note that the sound quality in the open-plan office could be improved by adding sound to the space using speakers (installed in the space or above the suspended ceiling). Balancing the spectrum would eliminate the perceptible rumble; the subjective analysis of the office occupants would then agree with the objective acoustical data.

Finally, note that NC charts do not include the 31.5 Hz and 16 Hz octave bands. Although manufacturers of HVAC equipment seldom provide sound power ratings for these bands (because it is difficult to obtain reliable data), these frequencies do affect acoustical comfort. Sound pressure levels at the 16 Hz and 31.5 Hz octave bands are measurable in an existing space; such readings may provide useful diagnostic information.

Note: The “balanced noise criteria” (NCB) method, which is described in the 1999 ASHRAE Handbook–HVAC Applications, includes the 16 Hz and 31.5 Hz

Figure 12. Plotting sound pressure levels on an NC chart (example)


Evaluation Methods

octave bands, and it lowers the permissible noise levels at high frequencies. Although the NCB method is more complex than the NC rating procedure that it is intended to replace, it also is better at indicating whether a sound spectrum requires corrective action.

Room criteria (RC) curves

Like NC curves, room criteria, or RC, curves are used to rate the sound pressure levels in indoor environments. What distinguishes the RC rating system from its NC counterpart is that it also indicates the character of the overall sound.

Sound spectrums can be unbalanced in ways that result in poor acoustical quality. Too much low-frequency sound results in a rumble; too much high-frequency sound produces a hiss. RC curves, which are based on sound pressure levels at center frequencies of 31.5 Hz to 4000 Hz, provide a means of identifying these imbalances via a two-part descriptor:

• Speech interference level (SIL), a numeric value that represents the arithmetic average of the sound pressure levels for the 500 Hz, 1000 Hz, and 2000 Hz octave bands.

• Spectrum quality, a letter that characterizes the sound as a subjective observer might—

As in the NC rating system, an RC descriptor is derived from a full-octave-band set of sound pressure readings. Although the rating process2 involves more steps, it also is relatively simple and provides more information about the ambience of the space:

1 On an RC chart, plot the sound pressure level for each octave band.

2 Calculate the SIL by arithmetically averaging the sound pressure levels at the octave-band center frequencies of 500 Hz, 1000 Hz, and 2000 Hz.

In this example, the arithmetic average of 38 dB, 31 dB, and 24 dB is 31 dB.

3 Draw the reference line, C, that will serve as a reference for evaluating the quality of the sound spectrum (Figure 13). The line has a slope of –5 dB per octave and passes through the calculated SIL at the 1000 Hz octave band.

4 Draw the boundaries (lines D and E) that represent the maximum allowable deviation for a “neutral” rating. Line D extends from 31.5 Hz to 500 Hz, and is 5 dB above reference line C, while line E extends from 1000 Hz to 4000 Hz, and is 3 dB above reference line C.

N = neutral (or balanced; unnoticeable)R = rumbleH = hissRV = perceptible vibration

2 The RC rating process described here paraphrases the instructions provided in Chapter 42, “Sound and Vibration Control,” of the 1991 ASHRAE Handbook–Applications.

Accurate determination of sound power levels for the 16 Hz and 31.5 Hz octave bands requires a large, open reverberant room. Construction costs and the difficulty of qualifying such a facility deter most HVAC equipment manufacturers from providing sound data in these two octave bands … which, in turn, makes it difficult to predict low-frequency sound for HVAC equipment.

Evaluation Methods


5 Judge the quality of the sound spectrum by observing how it deviates from the boundary lines drawn in Step 4. Based on the following descriptions, choose the letter descriptor that best characterizes the subjective quality of the background noise:

– Neutral, N. The sound level in each of the octave bands between 31.5 Hz and 500 Hz is below line D, and the sound level in each of the octave bands between 1000 Hz and 4000 Hz is below line E. (Most observers would judge this sound spectrum as “acceptable.”)

– Rumble, R. The sound level in any octave band between 31.5 Hz and 500 Hz is above line D.

– Hiss, H. The sound level in any octave band between 1000 Hz and 4000 Hz is above line E.

– Perceptible vibration, RV. The sound level in the octave bands between 16 Hz and 63 Hz lies in the shaded region (A or B) of the RC chart. Region A represents the high probability of noticeable, noise-induced vibration in lightweight wall and ceiling constructions; in other words, expect audible rattles in light fixtures, doors, and wall hangings. Region B indicates the possibility of such vibration.

The resulting RC rating is a composite of the SIL number calculated in Step 2 and the letter descriptor determined in Step 5. If we plot the acoustical data for our example office space (Figure 12, p. 13), the result is a rating of RC 31(R). The SIL is 31 and the sound pressure levels in the 63 Hz and 125 Hz octave bands are above line D, indicating that the sound spectrum will include a perceptible rumble (therefore, warranting the R descriptor).

Note: The three-part RC Mark II method revises the RC method. It consists of a family of criterion curves and two procedures: one for determining the RC

Figure 13. Plotting sound pressure levels on an RC chart (example)


Evaluation Methods

numerical rating (a number plus a letter, which denotes sound quality), and one for estimating occupant satisfaction if the spectrum doesn’t match the shape of an RC curve (the quality assessment index, or QAI). For details on how to determine the RC Mark II rating, refer to Chapter 7 of the 2005 ASHRAE Handbook—Fundamentals.

To aid building/HVAC professionals, ASHRAE recommends RC targets for various types of spaces (Table 4).

Effects of noise on speech

The underlying objective of an acoustical design target is to achieve a favorable listening environment. Of the myriad factors that hamper the intelligibility of speech, background noise often predominates. The speech interference level predicts the extent to which background noise will interfere with oral communication. It’s determined by calculating the arithmetic average of the sound pressure levels measured at the 500 Hz, 1000 Hz, and 2000 Hz octave-band center frequencies.

The following chart shows how speech interference varies with the level of the speaker’s voice and with the distance from the speaker to the ear. In this case, the speakers were males with average vocal strengths, they faced the listener, there were no reflecting surfaces nearby, and the spoken material was unfamiliar to the listener.

The masking effect of a particular sound is greatest on the sounds that are close to it in frequency.

SOURCE:A. Peterson & E. Gross, Handbook of Noise Measurement (Concord, MA: GenRad, Inc.), 1972

Evaluation Methods


Table 4. Design guidelines for HVAC-related background sound in unoccupied spaces

Space typeRC(N) target a,b

(QAI ≤ 5 dB)

Residences Private dwellings, apartments, condominiums 25 to 35

Hotels, motels Individual rooms or suites 25 to 35

Meeting/banquet rooms 25 to 35

Corridors, lobbies, service/support areas 35 to 45

Office buildings Executive/private offices 25 to 35

Conference rooms 25 to 35

Teleconference rooms 25 (max)

Open-plan offices 30 to 40

Corridors and public lobbies 40 to 45

Hospitals, clinics Private rooms, operating rooms 25 to 35

Wards, corridors, public areas 30 to 40

Performing arts spaces

Drama theaters 25

Concert and recital halls —c

Music-teaching studios 25

Music practice rooms 30 to 35

Laboratories with fume hoods

Testing/research, minimal speech communication 45 to 55

Research, extensive telephone use, speech communication

40 to 50

Group teaching 35 to 45

Churches, mosques, synagogues

General assembly with critical music programs 25 to 35c

Schoolsd Classrooms 25 to 30

Large lecture rooms 25 to 30

(without speech amplification) 25

Courtrooms Unamplified speech 25 to 35

Amplified speech 30 to 40

Libraries 30 to 40

Indoor stadiums, gymnasiums

School/college gyms, natatoriums 40 to 50e

Large-seating-capacity spaces with amplified speech

45 to 55e

a Values and ranges are based on judgment and experience, not on quantitative evaluations of human reactions. They represent general limits of acceptability for typical building occupancies. Higher or lower values may be appropriate and should be based on a careful analysis of economics, space usage, and user needs.

b When quality of sound in the space is important, specify criteria in terms of RC(N). If quality of sound in the space is of secondary concern, the criteria may be specified in terms of NC or NCB levels of similar magnitude.

c An experienced acoustical consultant should be retained for guidance on acoustically critical spaces (below RC 30) and for all performing arts spaces.

d Some educators and others believe that HVAC-related sound criteria for schools, as listed in previous editions of the ASHRAE Handbook–HVAC Applications, are too high and impede learning for affected groups of all ages. See ANSI/ASA Standard S12.60–2002 for classroom acoustics and a justification for lower sound criteria in schools. The HVAC component of total noise meets the background noise requirement of that standard if HVAC-related background sound is ≤ RC 25(N).

e RC or NC criteria for these spaces need only be selected for the desired speech and hearing conditions.

SOURCE: Adapted from Table 34 in Chapter 47, “Sound and Vibration Control,” of the 2003 ASHRAE Handbook–HVAC Applications

The RC(N) values in Table 4 are based on a quality assessment index, QAI, of ≤ 5 dB. The index can help you predict how occupants are likely to evaluate less-than-optimal sound quality. A QAI of ≤ 5 dB indicates acceptable quality, provided that the perceived level of sound falls within the recommended range and that the sound pressure level for the 16 Hz or 31 Hz octave band does not exceed 65 dB.

If the QAI exceeds 5 dB but is ≤ 10 dB, the likelihood of occupant satisfaction is questionable. The average occupant is unlikely to accept the sound quality if the QAI exceeds 10 dB.

For more information about using the QAI to predict occupant satisfaction, see Chapter 7 in the 2005 ASHRAE Handbook–Fundamentals.


Evaluation Methods

Sone scale

The sone scale is based on the rule of thumb for loudness, which is that the intensity of a particular sound must increase by a factor of 10 for it to be perceived as twice as loud. (For example, “it takes 10 violins to sound twice as loud as one violin”; or, “loudness doubles for every 10 phon increase in the sound loudness level.”) As such, it provides a single-number rating scheme that establishes a direct, proportional relationship between sound intensity and loudness as perceived binaurally—that is, with both ears—by a person with normal hearing.

One sone is defined as the loudness of a 1000 Hz tone at a level of 40 dB (0.02 microbar, reference 0.0002 microbar; Figure 14). Because the scale is linear, 2 sones are twice as loud as 1 sone and half as loud as 4 sones.

AMCA Standard 301, Methods for Calculating Fan Sound Ratings from Laboratory Test Data, provides a method for calculating the sone rating from octave-band data. With the exception of unducted fans and power ventilators, the sone scale is seldom used to rate equipment sound.

Note: Use particular discretion when comparing equipment based on sone ratings. Several calculation methods exist and their results differ. One further caveat: Sone ratings do not reveal the character of sound, which limits the scale’s usefulness for predicting how equipment will sound after it’s installed.

Phon scale

Equal loudness curves (Figure 14) illustrate how the sensitivity of our hearing varies based on sound frequency. The decibel level at a standard frequency (1000 Hz) on each of these curves becomes the reference for measuring loudness in phons. For example, a sound pressure level of 40 dB at 1000 Hz

The sone scale is not appropriate for rating extreme sound levels, whether low or high, because the subjective nature of human hearing does not follow the 10 dB “rule of thumb.”

Figure 14. Phon scale

Evaluation Methods


has a loudness of 40 phons. Therefore, any sound pressure level lying on the 40 phon contour will be perceived as having a loudness of 40 phons, regardless of its frequency.

The sone scale, which is a comparative measure of loudness, is the linear equivalent of the logarithmic phon scale (Figure 15); doubling the sone level increases the corresponding phon level by 10. For example: 1 sone = 40 phon, 2 sone = 50 phon, 4 sone = 60 phon, and so on.

Octave-band rating method

Sound levels rated by octave band are more useful than single-number ratings to acousticians and building professionals who are charged with providing an acoustically comfortable environment. Although octave-band data is not as simple to interpret, it provides considerably more information about the character of the sound (which affects how the sound will be judged by occupants).

Both sound power levels and sound pressure levels can be presented in octave-band format. Ratings of acoustical output in terms of sound power level by octave band enable an “apples to apples” comparison of various equipment types and models. Such ratings also can be converted to sound pressure levels (Table 5), allowing predictions of background sound when the details of the environment are known. (See “Predicting ambient sound,” pp. 33–38.)

Sound pressure levels in each octave band, whether predicted from sound power data or measured in an existing environment, will reveal more about the character of sound than any of the single-number rating methods.

Note: Any of the single-number ratings described in this chapter can be calculated from octave-band sound pressure data; however, octave-band data cannot be derived from any of the single-number ratings.

Table 5. Example of octave-band ratings

Octave-bandcenter frequency, Hz

Sound power of equipment,dB ref 10-12 W

Sound pressure level in the space,dB ref 20 μPa

63 103 63

125 104 52

250 100 45

500 101 38

1,000 98 31

2,000 93 24

4,000 88 16

8,000 85 10

Figure 15. Sone–phon relationship


RatingEquipment Sound

Sound fields

Understanding how sound behaves in various environments, or “fields,” is fundamental to accurate measurements of sound pressure levels, uniform sound power ratings, and appropriate, enforceable acoustical criteria for HVAC system designs.

Free field

The academic definition of a free field is a “homogeneous, isotropic medium that is free from boundaries,” or more simply, an environment that is uniform throughout and identical in all directions. In practice, most sound sources rest on a reflecting planar surface rather than float in midair. Given this reality, a large open area without obstructions (a parking lot or meadow, for example) most closely represents a free field.

An ideal sound source—one that radiates sound equally in all directions—within a free field emits sound pressure waves in a spherical pattern (Figure 16). Therefore, the sound pressure at equal distances from the source is the same in all directions. As the sound waves travel farther from the source, they cover a larger spherical area: doubling the distance from the source disperses the sound over four times as much surface area (Figure 17). (Recall that the area of a sphere increases with the square of the distance from its center, that is, 4π × r².)

The following mathematical model uses the relationship between distance and area to predict how sound will change with distance from the source:

From this model, we know that doubling the distance from the sound source will reduce the sound pressure level by 6 dB. This fact is useful when estimating outdoor sound levels.

Within the context of HVAC systems, this equation is commonly used to determine how loud a piece of equipment will be at a given distance. For example, it may be important to know how much of the sound produced by an air-cooled chiller will reach the property line situated 120 ft (37 m) away (Figure 18). If the cataloged acoustical output of the chiller is 95 dB at a distance of 30 ft (9 m), we can predict that the sound pressure level 120 ft (37 m) from the chiller will be 83 dB:

Figure 16. Free field

Lp2 Lp1 20 log10 r2 r1⁄( )×–=

where,Lp2 = unknown sound pressure, in dBLp1 = known sound pressure, in dBr1 = distance between source and site of Lp1 measurementr2 = distance between source and site of predicted measurementFigure 17. Increased distance spreads

sound over more volumetric area

Lp2 95 95 log10 120 30⁄( )×[ ] 83 dB=–=

Rating Equipment Sound


This mathematical model is only valid for predicting sound pressure from a point source in a free-field environment. It cannot convert a sound power level (Lw) to a sound pressure level (Lp), nor does it account for the effect of reflective surfaces or obstacles in the sound path.

Note: Large air-cooled equipment can resemble a line source of sound rather than a point source, which means that the drop in sound intensity may be other than 6 dB per doubling of distance from the source.

Acoustic nearfield

Acoustic nearfield describes an area, adjacent to the source, where sound does not behave as it would in a free field (Figure 19). Most sound sources, including all HVAC equipment, do not radiate sound in a perfectly spherical pattern. Equipment consists of different materials; as a result, sound radiates from different surfaces at different intensities. Also, the irregular shape of equipment causes pressure-wave interactions that make sound wave behavior virtually unpredictable. For these reasons, it is unwise to measure or predict sound pressure for “nearfield” locations around an equipment sound source.

The extent of the acoustic nearfield depends on the overall dimensions of the equipment and the type(s) of sound source(s) it contains. For HVAC equipment, the practical definition of “nearfield” is twice (2×) the largest unit dimension for a suspended unit or quadruple (4×) the largest unit dimension for a unit placed on a hard, reflective surface.

Figure 18. Distance correction for free-field sound measurements

Figure 19. Acoustic nearfield



Reverberant field

A reverberant field is nearly the opposite of a free field. Reverberant fields typically exist in rooms with many sound-reflecting surfaces (including walls, floors, and ceilings; Figure 20), but they also can occur in streets lined by high buildings.

When a sound source is placed within an enclosed space, the sound waves it generates bounce back and forth between the reflective surfaces many times. The resulting effect is a nearly uniform, or diffuse, sound field. If the enclosed space is perfectly reverberant, then the sound pressure at all points in the room is equal.

Because the characteristics of a reverberant field make it easier to measure equipment-generated sound, special reverberant rooms are designed, constructed, and qualified specifically for that purpose. (See “Rating environments,” p. 23.)

Semireverberant field

Most interior spaces exhibit both free-field and reverberant-field acoustical characteristics (Figure 21). The walls and ceiling prevent sound from behaving in a free-field manner, but they are not perfectly reflective. When a sound wave reaches one of these boundaries, part of the wave is reflected and part of it is absorbed (or transmitted).

In the semireverberant field of an interior space, sound behaves much as it does in a free field: that is, the sound level decreases as the distance from the source increases—but by less than 6 dB per doubling of distance, as it would in a perfect free field. Close to the unit, in the acoustic nearfield, measurement is unpredictable. Near the wall, in the reverberant field, the additive effect of the reflected sound begins to cancel out the sound level reduction that occurs with distance from the source.

Two factors determine where the reverberant and semireverberant fields exist in a room: room construction, which determines the absorption effect, and room volume relative to the size of the sound source. (Together, these factors define the “room constant.”) Small rooms with hard, reflective surfaces behave similarly to reverberant rooms. This description fits many mechanical equipment rooms, which typically are constructed of concrete and are small with respect to the size of the sound source.

Figure 20. Reverberant field (plan view)

Figure 21. Semireverberant field (plan view)



Rating environments

Sound power most accurately represents the acoustical output of equipment, but it cannot be measured directly. Instead, it must be determined from sound pressure measurements. Because sound pressure is influenced by the surroundings, it is measured in an environment with known acoustical characteristics—an acoustical laboratory; the measured sound pressure levels then are adjusted to account for environmental effects. Two testing environments commonly are used to determine sound power ratings for HVAC equipment: a reverberant room and a free field (with flat reflecting plane).

Reverberant room

The primary test environment for developing equipment sound power ratings is the hard-surfaced reverberant room (Figure 22), which creates a diffuse sound field by reflecting and mixing sound waves. The hard surfaces of the walls, floor, and ceiling cause multiple reflections of sound waves, which result in nearly uniform sound pressure levels throughout the reverberant room.

Sound power levels for the tested equipment are calculated from the sound pressure levels measured in the reverberant room. The reverberant-room method is commonly used to determine the sound power of fans, air handlers, compressors, in-room air conditioners, terminal equipment (such as fan-coils and VAV boxes), and diffusers.

Free field (with flat reflecting plane)

A free field rating environment provides acoustical surroundings in which sound from the tested equipment radiates equally in all directions (except through the planar surface on which the equipment rests) and diminishes at a predictable rate with distance. Sound power levels are determined by measuring the sound pressure levels at various points on an imaginary, hemispherical surface surrounding the equipment. (See Figure 16, p. 20.) Any physical boundaries, which may be present, have little or no effect on the sound pressure readings for the frequency range of interest.

Typically, a building constructed to house an acoustical lab is situated apart from other structures and at a considerable distance from heavy manufacturing activities. This minimizes the conduction of sound into the building through the foundation.

The outer walls, which usually are constructed of thick masonry, are totally isolated from the reinforced concrete test rooms inside the building. Each test room is supported by spring-and-rubber-pad vibration isolators, which sit atop concrete piers. This construction prevents building, ground, and cross-room sound transmissions from reaching the testing environment.

Some testing facilities accommodate acoustical mockups, which simulate the acoustical characteristics of a proposed design by duplicating the construction and configuration of the equipment room and the adjacent occupied space. The mockup can then be used to verify the projected sound levels generated by the proposed equipment.

When a room dimension equals (or is a multiple of) a sound wavelength, the reflected wave is nearly in phase with the original. This phenomenon produces a standing wave, which can produce large pressure variations that distort sound-rating measurements. To prevent standing waves from forming, the reverberant room shown here contains a large, rotating vane that continuously changes the reflective dimensions of the room. Using a moving microphone to sample equipment-generated sound can further guard against erroneous measurements caused by standing waves.

Figure 22. Reverberant room



Large HVAC equipment (packaged rooftop air conditioners, air-cooled chillers, and cooling towers, for example), usually are tested outdoors in a parking-lot setting. Smaller equipment, such as fan-coils and VAV boxes, typically are tested indoors, in an anechoic, or “no echo,” room that simulates a free field. To help assure that sound pressure waves travel away from the equipment in an even, hemispherical pattern, the anechoic room in Figure 23 combines a hard-surfaced floor with a 3-ft-deep layer of glass-fiber wedges on the walls and ceiling. (The glass-fiber layer absorbs virtually all sound at frequencies higher than 100 Hz.)

An anechoic room’s acoustical characteristics make it particularly useful for identifying the dominant sources and directions of radiated equipment sound. Because sound travels in a direct path from the equipment, it’s possible to locate a particular sound-radiating source that otherwise would be masked by sound emitted from elsewhere in the equipment. Armed with this information, manufacturers can design their equipment with appropriate sound attentuation and provide application suggestions (regarding equipment placement, for example) that will contribute to an acoustically successful installation.

Rating methods

Substitution

Although sound power can be determined from electronic measurements of sound intensity, sound power ratings for HVAC equipment typically are derived from measurements of sound pressure and a technique called substitution.

The substitution method is based on a reference source that has a known sound power level. Most acoustical laboratories use a specially designed centrifugal fan. The fan has no housing and is driven by a motor that maintains a constant speed of rotation. These characteristics help to assure that fan emits uniform sound power in each octave. (Reference sources for sound power ratings must be calibrated by an accredited acoustical lab and must meet the requirements of ISO 3741, ISO 3747, ISO6926, and ANSI S1.31.)

The reference source and the equipment to be rated (the “test unit”) are placed in a reverberant room, side by side and at equal distances from a microphone. Sound pressure is measured twice for each octave band: once while only the reference source operates, Lp1, and once while only the test unit operates, Lp2. The sound power for the test unit is then calculated by substituting the measured values in this equation:

Figure 23. Anechoic room

A sound intensity probe makes it possible to detect sound intensity, which is the amount of sound power that passes through 1 m² of surface. The probe is an acoustical sensor, usually consisting of two closely spaced microphones. A computer calculates the sound velocity as the difference between the two pressure signals (corrected for air density and microphone spacing), and then multiplies this velocity by the average sound pressure from the two microphones. The resulting sound intensity is expressed in watts per square meter (W/m²).

Lw2 Lp2 Lw1 Lp1–( )+=

where,Lp1 = measured sound pressure for the reference sound source, in dBLp2 = measured sound pressure for the test unit, in dBLw1 = known sound power for the reference sound source, in dBLw2 = calculated sound power for the test unit, in dB

55= 95 88–( ) 62 dB=+



If the known sound power for the reference source is 95 dB, and the measured sound pressure is 88 dB for the reference source and 55 dB for the test unit, then the calculated sound power for the test unit is 62 dB.

Repeating this calculation for the measured sound pressures at each octave band establishes a sound power profile for the test unit.

One-third-octave-band and full-spectrum charts

Cataloged sound ratings for HVAC equipment commonly consist of full-octave-band values that range from 63 Hz to 8000 Hz. However, the equipment usually is tested across a broader range of frequencies, and at narrower increments—either at 1/3-octave bands (Table 6 and Figure 24), or at each frequency for a full-spectrum analysis. Such detailed studies make it possible to isolate and reduce or eliminate tones through equipment redesign and/or attenuation.

Table 6. Center frequencies for octave vs. 1/3-octave bandsa

a Bold values denote the center frequencies for full-octave bands.

Octaves, Hz 1/3-Octaves, Hz Octaves, Hz 1/3-Octaves, Hz

1612.51620

500400500630

31.525

31.540

1000800

10001250

63506380

2000160020002500

125100125160

4000315040005000

250200250315

80006300800010000

Figure 24. Example of sound plotted on a 1/3-octave-band chart



Industry standards

Published sound ratings for HVAC equipment can be derived from laboratory testing or estimated from generic sound projection equations. Understandably, ratings based on actual test results are preferable to abstract calculations; field testing to confirm published data is difficult and seldom conclusive. However, collecting accurate sound data for an entire line of products is both time-consuming and expensive. A single product line may consist of many models offered in a range of sizes and with various options, all of which can affect the sound generated by each unique combination of variables.

Not only does the basis for sound ratings vary, but so do the terms in which the ratings are expressed. (See the previous chapter, “Evaluation Methods.”) This lack of consistency between manufacturers, equipment types, and rating schemes can be problematic for building professionals because it hampers equitable comparisons of similar equipment.

Although the HVAC industry lacks a single, mandatory, all-inclusive acoustical standard, voluntary standards do exist to encourage uniformity in data collection and determination of ratings. These standards are authored by different organizations with different goals; consequently, they address different types of equipment and define different methods for testing, measuring, and rating sound sources. Table 7 identifies the relevant industry standards for the major types of HVAC equipment, while Table 8 (pp. 28–29) summarizes the scope of each standard and comments on its application.

When using sound ratings to establish acoustical design targets and select equipment, keep these considerations in mind:

• A standard is not a static document; take care to review the most current version.

• When comparing equipment based on sound ratings, find out how the ratings were determined. Significant differences can exist between generalized “averages” and data derived from actual test results.

• Critical applications may warrant witness tests, with an acoustical engineer or consultant present, to verify that equipment performs as predicted.

• Be wary of single-number sound ratings: They enable quick comparisons of equipment, but the actual ambient sound represented by identical ratings can seem very different to the ear.

The Air-Conditioning & Refrigeration Institute (ARI), the Air Movement and Control Association (AMCA) International, and the American Society of Heating, Refrigerating, and Air-Conditioning Engineers (ASHRAE) promulgate the standards referenced in this manual. Through voluntary compliance, manufacturers facilitate equitable comparisons of similar equipment, regardless of brand and model.



Table 7. Sound rating standards for commercial HVAC equipment

Equipment type Relevant standard(s)a

Air terminals (VAV boxes) Radiated and discharge sound ARI 880

Conditioned, occupied space ARI 885b

Stand-alone fans AMCA 300

Central-station air handlers ARI 260

Fan–coils Ducted ARI 260

Unducted in-room ARI 350

Packaged terminal air conditioners and heat pumps (PTAC, PTHP)

ARI 300

Heat pumps (water-, air-, and ground-source)

Ducted ARI 260

Unducted ARI 350

Packaged (self-contained) air conditioners

ARI 260

Packaged rooftop air conditioners

≤ 135 MBtu/hr (40 kW), indoor ARI 260

≤ 135 MBtu/hr (40 kW), outdoor ARI 270c

> 135 MBtu/hr (40 kW), indoor ARI 260

> 135 MBtu/hr (40 kW), outdoor ARI 370

Air-cooled condensing units < 135 MBtu/hr (40 kW) ARI 270c

≥ 135 MBtu/hr (40 kW) ARI 370

Air-cooled chillers < 135 MBtu/hr (40 kW) ARI 270c

≥ 135 MBtu/hr (40 kW) ARI 370

Water-cooled chillers ARI 575

a Table 8 (pp. 28–29) summarizes the sound rating parameters for these standards. For more information or to obtain copies, visit these Web sites: www.ari.org and www.amca.org.

b ARI 885, Procedure for Estimating Occupied Space Sound Levels in the Application of Air Terminals and Air Outlets, provides “industry agreed-upon methods to use ARI Standard 880 sound ratings to estimate the sound levels [that] will occur in the conditioned, occupied space.” In scope, ARI 885 only addresses the sound level contributions of air terminals, air outlets, and the low-pressure ductwork that connects these devices. It does not account for contributions by the central fan system, ductwork upstream of the air terminal, equipment room machinery, nor exterior ambient sound.

c Also ARI 275, Application of Sound Rating Levels of Outdoor Unitary Equipment, to estimate A-weighted sound pressure levels for factory-built air-conditioning and heat-pump equipment rated in accordance with ARI 270.



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how

the

unit w

as

sound-t

este

d

(entire

unit o

r fa

n o

nly

)

•Fo

r fa

n-o

nly

ratings,

fin

dout

how

se

condary

sourc

es

wer

e addre

ssed

ARI

270:

Sound R

ating o

f O

utd

oor

Unitary

Equip

men

tO

utd

oor, s

ingle

-num

ber

, A-w

eighte

d

sound p

ow

er, in b

els

… F

ull-

oct

ave

sound

pow

er

(63

Hz

to 8

000

Hz)

may

als

o b

e a

vaila

ble

•Te

st s

etu

ps

for

reve

rber

ant-

room

and f

ree-

field

en

viro

nm

ents

•C

alc

ula

ting 1

/3-o

ctav

e-b

and d

ata

fro

m t

est

m

easu

rem

ents

; co

nve

rtin

g t

o a s

ingle

-num

ber

, A-w

eighte

d s

ound p

ow

er v

alu

e

•Acc

ounting f

or

spec

trum

irr

egula

rities

cause

d b

y th

epre

sence

of

tones

•Equip

men

t: P

ack

aged

rooft

op a

ir c

onditio

ners

≤135

MBtu

/hr

(≤ 4

0kW

), a

ir-c

oole

d c

onden

sing u

nits

<135

MBtu

/hr, a

ir-c

oole

d c

hill

ers

<135 M

Btu

/hr

If o

ctav

e-band d

ata

is

unava

ilable

, use

ARI

275,

Applic

ation o

f Sound R

ating L

evels

of

Outd

oor

Unitary

Equip

men

t, t

o est

imate

th

e ef

fect

of

job-s

ite

conditio

ns

and c

onve

rt

from

the

single

-num

ber

, A-w

eighte

d s

ound

pow

er t

o s

ound p

ress

ure

for

a p

art

icula

r lo

cation

ARI

275:

Applic

ation

of

Sound

Rating L

eve

ls o

f O

utd

oor

Unitary

Equip

men

t

Outd

oor, s

ingle

-num

ber

, A-w

eighte

d

sound p

ress

ure

lev

els,

in d

ecib

els

•Acc

ounts

for

applic

ation f

act

ors

:Equip

men

t lo

cation,

barr

ier

shie

ldin

g,

sound p

ath

, and d

ista

nce

•Equip

men

t:O

utd

oor

sec

tions

of

fact

ory

-made a

ir

conditio

ners

and h

eat

pum

ps

ARI

275 p

redic

ts s

ound p

ress

ure

leve

ls u

sing

tone-

adju

sted

sound p

ower

leve

ls d

ete

rmin

ed

in a

ccord

ance

with A

RI

270 …

As

a r

esu

lt,

pre

dic

tions

base

d o

n A

RI

275 m

ay b

e s

lightly

hig

her

than m

easu

red v

alu

es

ARI

300:

Sound R

ating a

nd

Sound T

ransm

issi

on L

oss

of

Pack

aged

Term

inal Equip

men

t

•In

doo

r so

und p

ow

er lev

els

, by

full

oct

ave,

from

63 H

z to

8000

Hz

•A-w

eighte

d s

ound p

ower

lev

el

•To

ne-

adju

sted

, A-w

eighte

d s

ound

pow

er lev

el

•Pro

vides

a m

ethod f

or

dete

rmin

ing s

ound

transm

issi

on

loss

•D

escr

ibes

test

req

uirem

ents

for

det

erm

inin

g o

r ve

rify

ing indoor

and o

utd

oor

sound r

atings

•Equip

men

t: P

ack

aged

ter

min

al air c

onditio

ner

s and

pack

aged

ter

min

al hea

t pum

ps

inte

nded

for

use

in

resi

den

tial, c

omm

erci

al, a

nd indust

rial hea

ting a

nd

coolin

g s

yste

ms

•W

hen c

om

paring e

quip

men

t base

d o

n

publis

hed

sound r

atings,

make

sure

that

the

rating c

onditio

ns

are

identica

l in

all

case

s

•If

AM

CA

300 is

cite

d f

or

air-h

andle

r so

und

ratings,

dis

cove

r if t

he

dat

a re

flec

ts t

he

entire

unit o

r only

the

fan

ARI

350:

Sound R

ating o

f N

on-D

uct

ed I

ndoor

Air-

Con

ditio

nin

g E

quip

men

t

•In

doo

r so

und p

ow

er lev

els

, by

full

oct

ave,

from

63 H

z to

8000

Hz

•A-w

eighte

d s

ound p

ower

lev

el

•To

ne-

adju

sted

, A-w

eighte

d s

ound

pow

er lev

el

•Te

st s

etu

ps

for

reve

rber

ant-

room

envi

ronm

ent

•C

alc

ula

ting 1

/3-o

ctav

e-b

and d

ata

fro

m t

est

m

easu

rem

ents

; co

nve

rtin

g t

o a s

ingle

-num

ber

, A-w

eighte

d s

ound p

ow

er v

alu

e (b

el)

•Acc

ounting f

or

spec

trum

irr

egula

rities

•Equip

men

t: f

an–co

ils,

unduct

ed h

eat

pum

ps

For

aco

ust

ically

critica

l applic

ations,

use

so

und p

ow

er b

y oct

ave b

and r

ath

er t

han a

si

ngle

-num

ber

ave

rage

(con

tinued

on n

ext

page)



ARI

370:

Sou

nd R

ating o

f La

rge

Outd

oor Ref

riger

ating a

nd

Air-C

onditio

nin

g E

quip

men

t

•O

utd

oor

sound p

ow

er lev

els

, by

full

octa

ve,

from

63 H

z to

8000

Hz

•A-w

eighte

d s

ound p

ow

er

leve

l

•To

ne-

adju

sted

, A-w

eighte

d s

ound

pow

er lev

el

•Req

uires

two

sets

of

ratings

at

ARI

ther

mal lo

ad

conditio

ns:

one

set

with t

he

entire

unit o

per

ating

and

one

set

with o

nly

the

fans

oper

ating

•M

anufa

cture

r ca

lcula

tes

1/3

-oct

ave-

band s

ound p

ower

data

fro

m r

everb

erant-

room

or

outd

oor

free

-fie

ld tes

ts,

then

acc

ounts

for

spec

trum

irre

gula

rities

, and c

onve

rts

to a

sin

gle

-num

ber

, A-w

eighte

d s

ound p

ow

er

valu

e (b

el)

•Exc

ludes

rem

ote

air-c

oole

d a

nd e

vapora

tive

co

nden

sers

•If

oct

ave-b

and d

ata

is

unav

aila

ble

, use

ARI

275 t

o e

stim

ate

the

eff

ect

of

job-s

ite

conditio

ns

•Fo

r aco

ust

ically

cri

tica

l applic

ations,

use

so

und p

ow

er lev

els

by

oct

ave

band r

ath

er

than a

sin

gle

-num

ber

ave

rage

ARI

575:

Meth

od o

f M

eas

uri

ng

Mach

inery

Sound w

ithin

an

Equip

men

t Room

Space

•In

door

sound p

ress

ure

lev

els

, by

full

octa

ve,

from

63

Hz

to 8

000

Hz

•Sin

gle

-num

ber, A

-wei

ghte

d s

ound

pre

ssure

lev

el

•D

escr

ibes

how

to d

eter

min

e th

e av

erage

sound

pre

ssure

for

a lis

tener

loca

ted 1

m f

rom

the u

nit

•In

tended a

s an in s

itu m

easu

rem

ent

pro

cedure

—not

as

a r

ating s

tandard

Use

ful fo

r det

erm

inin

g:

•W

het

her

hearing p

rote

ctio

n s

hould

be w

orn

w

hen

opera

ting o

r se

rvic

ing t

he u

nit

•Tr

ansm

issi

on-l

oss

req

uirem

ents

for

equip

men

t-ro

om

walls

How

eve

r, b

eca

use

ARI

575 d

oes

not

def

ine

the

test

ing e

nvi

ronm

ent,

rea

din

gs

reco

rded

at

the

job s

ite

can v

ary

consi

der

ably

fro

m t

he

manufa

cture

r’s

test

data

b

ARI

880:

Air T

erm

inals

Indoor

sound p

ow

er

leve

ls,

by

full

oct

ave,

from

125

Hz

to 4

000

Hz

•Rate

s both

radia

ted a

nd d

isch

arg

e so

und

•Pa

ralle

l-flow

, fa

n-p

ow

ered

ter

min

als

req

uire

two

sets

ofra

tings:

with a

nd w

ithout

fan o

pera

tion

•Exc

ludes

regis

ters

, diffu

sers

, and g

rille

s

•O

mits

63

Hz

oct

ave

band b

eca

use

air

ter

min

als

do

not

contr

ibute

sig

nific

ant

sound a

t th

is f

requen

cy

•Equip

men

t: c

om

mer

cial air

ter

min

als

and a

ir o

utlet

s,

plu

s th

e in

terc

onnec

ting low

-pre

ssure

duct

work

Est

ablis

hes

a c

ertifica

tion p

rogra

m f

or

sound-

rating a

nd -

test

ing a

ir t

erm

inal

s (e

xclu

din

g

retr

ofit

units)

at

standard

rating c

onditio

ns

ARI

885:

Pro

cedure

for

Est

imating O

ccupie

d S

pace

Sou

nd L

evel

s in

the

Applic

atio

n

of

Air

Term

inals

and A

ir O

utlet

s

(Not

applic

able

)•

Pre

dic

ts s

ound (

room

NC o

r RC)

in a

conditio

ned,

occ

upie

d s

pace

base

d o

n s

ound p

ow

er r

atings

for

air

term

inals

and a

ir-d

istr

ibution d

evic

esc

•Acc

ounts

for

low

-pre

ssure

duct

work

, ce

ilings,

and

room

eff

ect

•Exc

ludes

sound c

ontr

ibute

d t

o th

e occ

upie

d s

pace

by:

centr

al fan

, duct

work

upst

ream

of th

e ai

r te

rmin

al,

equip

men

t-ro

om

mach

iner

y, a

nd o

ther

sound s

ourc

es

outs

ide

the

occ

upie

d s

pace

•N

ot

all

manufa

cture

rs a

pply

the

sam

e ass

um

ptions

when

usi

ng A

RI

885 t

o

calc

ula

te N

C v

alu

es.

As

a re

sult,

com

pari

sons

of

equip

men

t by

diffe

rent

manufa

cture

rs m

ay b

e m

isle

adin

g

•Fo

r aco

ust

ically

cri

tica

l applic

ations,

use

ARI 885 to e

stim

ate

the u

niq

ue

transf

er

funct

ions

of t

he a

pplic

ation

; th

enapply

the

transf

er f

unct

ions

to a

ny

manufa

cture

r’s

ARI

880 s

ound p

ow

er d

ata

aVis

it w

ww

.ari.o

rg (

or

ww

w.a

mca

.org

for

AM

CA 3

00)

to fin

d o

ut

more

about

a par

ticu

lar

stan

dar

d.

Most

are

ava

ilable

onlin

e at

no c

ost

.b

To c

olle

ct s

ound d

ata, m

anufa

cture

rs c

om

monly

use

the

larg

est

pos

sible

tes

ting e

nvi

ronm

ent

(rel

ative

to t

he

phys

ical si

ze o

f th

e eq

uip

men

t) t

o m

inim

ize

reve

rber

ant

sound fro

m w

alls

, floo

r, a

nd c

eilin

g. In

a

typic

al eq

uip

men

t ro

om

, re

verb

eration w

ill s

ubst

antial

ly incr

ease

equip

men

t-gen

erat

ed s

ound.

cThe

estim

atin

g p

roce

dure

outlin

ed b

y ARI

885 p

resu

mes

that

the

sound p

ow

er r

atin

gs

of

the

air

term

inal

s co

nfo

rm w

ith A

RI

880.

Tab

le 8

.S

tan

dard

ized

so

un

d r

ati

ng

para

mete

rs

Indust

ry s

tandard

aSou

nd r

atings

Sco

pe

Applic

ation c

onsi

der

ation

s

(continued

fro

m p

revi

ous

page)


Environmental Effect of Equipment Sound

The annual surveys conducted by the Building Owners and Managers Association (BOMA) demonstrate how closely tenants equate sound/noise to overall comfort. Year after year, respondents consistently identify three factors as the primary motivation for relocating from one rented space to another: poor indoor air quality (IAQ), uncomfortable temperatures, and noise. The fact that the respective ranking of these factors changes annually suggests that they are of approximately equal importance.

Despite the subjective nature of sound, the acoustical quality of a commercial or industrial space can be measured, predicted, and objectively evaluated on the basis of two characteristics:

• the extent to which the background sound level disrupts or interferes with the functional use(s) of the space; and …

• speech privacy, which enables conversation without being overheard or overhearing the conversations of others.

Controlling background sound always has been a requirement for acoustically sensitive environments, such as music halls and recording studios. But the adverse effect of conversational distractions and uncontrolled noise on concentration and productivity has made acoustics an increasingly important design element for commercial buildings. Participants in a survey of 400 business managers estimated, on average, that reducing office noise would improve productivity or organizational effectiveness by 26 percent.3 The prospect of increased productivity provides an economic motivation to design and build quieter working environments.

The body of research substantiating the link between learning and the acoustical character of classrooms provided a similar incentive to improve learning environments in schools. Adding to that incentive is ANSI/ASA S12.60, Acoustical Performance Criteria, Design Requirements, and Guidelines for Schools, a standard of care that was co-developed by the American National Standards Institute (ANSI) and the Acoustical Society of America (ASA). First published in 2002, ANSI/ASA S12.60 establishes sound level targets based on studies, which showed that background sound—often generated primarily by the HVAC system—dramatically reduced the ability of students to understand the teacher. Because schools rely heavily on the spoken word for instruction, proper acoustical levels become a critical performance factor.

Many factors determine the acoustical environment of a space. Designing an acoustically appropriate environment requires careful attention to:

• Where the building is situated

• Size and shape of the room, and its placement relative to other interior spaces

3 The survey was conducted by BOMA and the University of Maryland in College Park.

Figure 25. Factors of ambient sound

• Occupant activities

• HVAC equipment

• Lighting ballasts

• Computers and other electrical appliances

• Pass-through noise from adjacent spaces

• Room construction, surface treatments, and furnishings

• Sound-masking systems



• Construction and surface treatments of the ceiling, walls, and floor

• Number, type, location, and intensity of contributing sound sources

Although HVAC equipment is not the sole contributor to background sound, it is often the dominant source. That’s why it is important to use accurate sound data, by octave band, for the HVAC system when acoustically analyzing the space.

Setting an acoustical target

Because vocal communication, which is carried primarily by the 500 Hz, 1000 Hz, and 2000 Hz octave bands, is central to most human activities, considerable research has been done to determine sound levels that interfere with speech. Table 9 compares average voice levels that yield barely acceptable speech at various distances and vocal strengths. On the surface, the data indicates that average background sound levels of less than 51 dB will permit a normal conversation between two people standing within 4 ft of each other. Bear in mind, however, that at the levels in Table 9, speech is just loud enough to be understandable. Lowering the acoustical target by 5 dB to 10 dB may improve the listening environment … but ironically, depending on the space and activity, larger reductions may make the space too quiet. The interference caused by background sound can provide privacy; in a cafeteria, for example, background sound can mask the conversations at neighboring tables.

Another way to set acoustical targets is to use the recommended room criteria (RC) ratings for various types of occupancy. For example, the 2003 ASHRAE Handbook suggested RC 35(N) to RC 45(N) for a hotel lobby. The average decibel value for the sound pressures lying on the RC 45 curve at 500 Hz, 1000 Hz, and 2000 Hz is 45 dB. Using the speech interference data in Table 9, we can expect that if the background sound level is 45 dB, a woman speaking in a normal voice will be barely audible at a distance of 8 ft and a man at 12 ft.

In hotel lobbies and similar environments where multiple conversations are the norm, it’s important to provide background sound that is loud enough to

Table 9. Average sound levels, dB, for barely audible speecha,b

a Decibel values in this table represent the average sound levels for the 500 Hz, 1000 Hz, and 2000 Hz octave bands.

b See “Effects of noise on speech,” p. 16, for a graphical representation of speech interference levels.

Speaker-to-listener distance, ft

Normal voice Raised voice Very loud

Man Woman Man Woman Man Woman

2 62 57 68 63 74 69

4 56 51 62 57 68 63

6 53 48 59 54 65 60

8 50 45 56 51 62 57

10 48 43 54 49 60 55

12 46 41 52 47 58 53

16 44 39 50 45 56 51



provide privacy without annoying or interfering with communication. By contrast, lower levels of background sound are required for environments in which hearing or concentration are important to the occupants’ activities. For example, the appropriate sound level for conference rooms, churches, and courtrooms (without speech amplification) ranges from RC 25(N) to RC 35(N).

“High-quality” ambient sound

Although speech interference is important, it is not a sufficient basis for designing an occupied space that provides high-quality ambient sound. (Ambient sound describes the many small sounds that we hear in aggregate as background sound.)

High-quality ambient sound is characterized by a balanced sound spectrum. To be “balanced,” sound pressure levels must be distributed across the range of audible frequencies. In other words, the sound pressure levels (plotted by octave band) will approximate the shape of an NC curve. Figure 26 illustrates the sound spectrums of four ambient environments. Only curve A represents a balanced sound spectrum. The sound pressure levels that make up curve B only occupy the low frequencies of the spectrum and will be perceived as a rumble. By contrast, the sound pressure levels of curve C only occupy the high frequencies of the spectrum and will be perceived as a hiss. Although the sound pressure levels of curve D are distributed across the frequency range, the “spike” at 500 Hz indicates that the background sound contains a tone that’s likely to annoy or distract occupants.

This example highlights a shortcoming of the NC rating system: Despite the apparent differences in the spectra represented by curves A, B, C, and D, each of them results in an NC 45 rating—which meets the industry’s recommendation for background sound in a cafeteria. However, except for

Figure 26. Spectral variations between identical NC ratings



curve A, it’s unlikely that anyone hearing these spectra will consider them to be acceptable. ASHRAE currently recommends use of the RC Mark II room criteria method because it better conveys the subjective quality of a sound spectrum; see “Room criteria (RC) curves,” (pp. 14–15). But the most accurate means for describing or predicting the sound quality of an occupied space is an octave-band analysis.

Predicting ambient sound

An accurate acoustical model makes it possible to predict the ambient sound of individual rooms early in the design phase. The model consists of:

• the receiver, which represents the specific location in the room where we want to predict the sound level

• the sources of the sound that will be heard at the receiver, which requires accurate octave-band sound data for each source

• the path, which represents an analysis of how the sound travels from each source to the receiver

A path can be very simple: For example, the path from a fan–coil to a receiver in the same room consists only of the correction for distance and the acoustical quality of the room. By contrast, the path from a remotely located air handler includes the connecting ductwork (straight sections, transitions, elbows, junctions), diffusers, and room characteristics.

Sound may travel from a single source to the receiver by several paths. The sound level at the receiver is the sum of all of the sound from the contributing sources via every path.

Acoustical characteristics of materials

Anticipating how the path will affect the sound that reaches the receiver requires a basic understanding of the acoustical characteristics of materials. These characteristics fall into three categories (Figure 27): absorption, transmission, and reflection.

Figure 27. Acoustical effects of a material on sound energy



Absorption

Absorptive materials convert acoustical energy into heat energy. The absorbed energy, Wa, is the portion of the incident sound energy, Wi , that is neither transmitted through nor reflected off the material (Figure 27, p. 33).

Absorptivity depends on several factors, including the thickness of the material, the frequency of the sound, and whether a reflective surface is located behind the absorptive material. Materials that are porous (such as open-cell foam) or fibrous (such as fiberglass insulation) absorb more energy than materials that are smooth and dense (such as sheet metal or gypsum board).

Increasing the thickness of an absorptive material increases the amount of absorption it provides. But the amount of energy that the material will absorb also depends on the frequency of the sound. High-frequency sound is more easily absorbed than low-frequency sound. The following rule of thumb illustrates this phenomenon: One inch of fiberglass insulation effectively absorbs sound at frequencies of 1000 Hz and higher, but to remain effective, the thickness must be doubled with each drop in octave—which means two inches of insulation to effectively absorb sound at 500 Hz and four inches at 250 Hz.

An absorptive material works best when the sound wave reflects back upon itself through the material. Because sound passes through absorptive materials with little difficulty, adding a reflective surface directly behind the absorptive material will improve its effectiveness.

The absorptive quality of a material is expressed as an absorption coefficient, α, which is the ratio of the sound energy absorbed by the material to the sound energy striking the surface. Generally (and preferably), an absorption coefficient is reported for each octave band, but absorptivity also may be expressed as a noise reduction coefficient (NRC). The NRC is simply the average of the absorption coefficients for the 250 Hz, 500 Hz, 1000 Hz, and 2000 Hz octave bands.

Transmission

A material acts as a barrier to incident sound energy, Wi , when it reduces the amount of sound energy that is conveyed through the material, Wt (Figure 27, p. 33). Like absorption, the transmission property is affected by the type and thickness of the material, the frequency of the sound, and the quality of construction.

Materials that are dense (masonry block or wallboard) or stiff (glass) generally are better at reducing transmitted sound than materials that are lightweight or flexible. Increasing the thickness of a material reduces the amount of sound that transmits through it. Just as high-frequency sound is more readily absorbed, it also is more easily blocked.

With homogeneous materials, the reduction in transmitted sound is primarily a function of mass as defined by the mass law.



Figure 28 plots the mass law equation for a broad range of frequencies. In reality, materials deviate from mass law as a result of less-than-perfect homogeneity, cracks, stiffness, resonance, and coincidence. Coincidence describes the phenomenon in which the natural frequency of the structure (a wall, for example) exactly matches the harmonics of the sound source, making the structure virtually invisible, acoustically.

Figure 29 demonstrates how various factors affect the mass law. Cracks (any opening) in the material dramatically reduce the sound attenuation of a panel, wall, or enclosure. Figure 29 also illustrates the effect of leaks on the performance of a barrier. Obviously, it is critical to carefully seal all leaks to maintain the rated performance of a material.

The ability of a material to reduce transmitted sound is described in terms of its insertion loss, noise reduction, or transmission loss:

• Insertion loss, IL, is the difference in sound pressure measured in a single location with and without a noise-control device located between the source and receiver. For example, the IL of the door in Figure 30 (p. 36) is the difference in sound pressure measured in the occupied space with the door open versus with the door closed.

Figure 28. Mass law

Figure 29. Effects of construction on transmission loss



• Noise reduction, NR, is the difference between the sound pressure levels measured on each side of a barrier. The NR of the door in Figure 30, for example, would be determined by measuring the sound pressure level in the office with the door closed, and in the equipment room (also with the door closed). The difference between these measurements is the NR of the door.

Both insertion loss and noise reduction are based on actual sound pressure measurements in decibels.

• Transmission loss, TL, is proportional to the ratio of the sound power incident on the source side of a wall to the sound power radiated by the wall on the receiver side. TL is expressed in decibels; mathematically, it can be expressed as the logarithmic form of the transmission coefficient, τ, because τ is the fraction of the incident energy, Wi, that’s transmitted through the material, Wt:

Alternatively, TL can be expressed in terms of sound power level:

The sound isolation (typically speech privacy) afforded by a structure, such as a partition, can be rated in terms of its sound transmission class (STC). This single-number rating is determined by plotting the decibel reduction for each 1/3-octave band from 125 Hz through 4000 Hz on an STC chart (Figure 31). The following conditions are then applied to fit an STC “contour” to the plotted data:

• The STL value at any frequency cannot be more than 8 dB below the STC contour; and,

• The sum of all “deficiencies” (STL less than contour value) at all frequencies cannot exceed 32 dB.

Figure 30. Insertion loss (IL)

TL 10 log10 1 τ⁄( ) 10log10 Wi Wt⁄( )= =

TL LWiLWt

–=



The STC value is obtained from the STL value of the highest-valued contour at 500 Hz. Figure 31 shows the STC “contours” that best fit the transmission loss data for three common construction materials: 4-in. concrete (STC 52), 1/2-in. gypsum board (STC 28), and 5/8-in. plywood (STC 21).

Architects use STC ratings to compare the ability of various panels to block sound that interferes with speech—that is, sound in the 500 Hz, 1000 Hz, and 2000 Hz frequencies. When it comes to acoustical analysis, however, the STC rating for a structure does not provide sufficient information about the acoustical profile to accurately model its effect on the sound that reaches the receiver. Contact the material manufacturer for full-octave or 1/3-octave transmission loss data.

Reflection

Some of the incident sound energy, Wi, bounces off—reflects from—the material. Reflected sound becomes an especially important consideration when the sound source and the receiver are located in the same room.

Consider a mechanical equipment room that contains a water chiller, pumps, and other sound-emitting sources. Typically, the walls are constructed of masonry, either cement block or poured concrete. Neither of these materials absorbs or transmits a significant portion of the incident sound energy, so most of it reflects back into the room. The reflected sound adds to the sound emanating from the sound source(s), greatly increasing the sound level in the room.

The best way to reduce reflected sound is to add an absorptive material to the walls, floor, and ceiling. However, it may not be practical to treat all surfaces—particularly the floor—because of the probable damage to

Figure 31. Examples of typical sound transmission losses



absorbent material. In such cases, adding an absorptive material to even one wall will help … adding it to two adjacent surfaces (the ceiling and a wall, or two adjacent walls) will significantly reduce the reflected sound.

Note: Even if the surface area is the same, treating two opposite (rather than adjacent) walls is less effective at reducing reflected sound because it will not reduce standing waves.

Occasionally, reducing reflected sound also may reduce the sound levels in adjacent spaces. Reducing the reflected sound energy in an equipment room, for example, lowers the overall sound level in that space … which means that less sound energy passes through the walls to the adjacent spaces, assuming that the transmission loss of the walls is unchanged. In other words, if it is quieter in the equipment room, then it will be quieter in the adjacent spaces.

Diffusion

Although the diffusion, or spreading, of sound is not a characteristic of materials or structures, it is a fundamental element of acoustical analysis. Diffusion refers to the comparative distribution of sound pressure variations throughout the space. It is a factor in room acoustics and is the primary consideration when calculating outdoor sound. (See “Sound fields,”pp. 20–22.)


Creating anAcoustical Model

Acoustical modeling looks at how sound from the source (a fan or compressor, for example) changes on its way to the receiver (the occupant). An earlier chapter, “Evaluation Methods,” described the single-number rating methods that are used to define the desired sound pressure levels in the space. The “Rating Equipment Sound” chapter then discussed the rating methods for defining the acoustical qualities of equipment, usually in terms of sound power levels.

Converting a sound power level into a sound pressure level requires definition, in acoustical terms, of the environment between the sound source and the receiver location. An acoustical analysis provides this definition by modeling the route that the sound takes from the source to the receiver. Everything that affects the sound along that journey constitutes the path (Figure 32). Sound from an individual source typically reaches the receiver via more than one path.

Anything that affects the sound, between the points of origin and reception, is considered an element of the sound path. For example, one of the sound paths from an air-handling unit (source) to someone (receiver) in a conference room is the supply air stream. Elements of this sound path include the ductwork components (straight duct, silencers, elbows, junctions, diffusers, etc.) and the acoustical characteristics of the space (its size, surface treatments, furnishings, and construction), as well as the source and receiver.

Another potential path carries radiated sound from the source and into the adjacent space through the intervening wall. The primary elements in this path are the air-handling unit, the acoustical characteristics of the source room, the wall, and the acoustical characteristics of the receiver location.

Other sources may contribute to the background sound at the receiver’s location. The important point to remember when performing acoustical modeling is that the sound heard by the receiver is the composite of individual sounds traveling along multiple paths from multiple sources. Each

Figure 32. Acoustical analysis predicts the contribution of HVAC sound in the space

Ductwork directs an airstream from one point to another. It also acts as a conduit for sound, which can travel in the opposite direction of airflow.


Creating an Acoustical Model

path’s contribution must be determined separately; the individual results are then summed to complete the model.

Modeling terminology

Several terms are commonly used in sound-path modeling in addition to the ones already introduced (source, path, receiver, element): attenuation, regenerated sound, end reflection, and receiver sound correction.

Attenuation

Attenuation refers to the reduction in sound level that occurs as the sound travels along the path from source to receiver (Figure 33). Most often, this term is used to describe the reduction of sound—resulting from absorption, transmission loss, or diffusion—as it travels through a duct system. Sound-attenuating elements include straight ducts, elbows, junctions, and silencers.

Regenerated sound

Regenerated sound originates from the airflow turbulence created by duct components, such as elbows, junctions, diffusers, silencers, and dampers. Turbulence results from an abrupt change in the direction or velocity of airflow and produces a corresponding drop in static pressure. Regenerated sound increases with air velocity or when the air is forced to make sharp turns (for example, through transitions with angles greater than 15 degrees inclusive).

Some elements both attenuate and regenerate sound. For example, as air makes a 90° turn in a rectangular duct elbow, some of the sound reflects back toward the source, attenuating the airborne sound downstream of the elbow. At the same time, however, the turbulence created as the air passes across the sharp corner produces regenerated sound.

Note: Properly designed duct transitions with reasonable flow velocities will not generate troublesome sound levels.

End reflection

When airborne sound passes from a restricted area, such as a duct, into a large area, such as a room, the sound-power levels decrease. End reflection causes part of the low frequency sound to be reflected back the way it came, reducing the sound that leaves the opening. The smaller the opening, the greater the amount of reflected sound (Figure 34).

Complete end reflection requires nonturbulent airflow at the end of the duct. Generally, this translates into approximately four equivalent duct diameters of unobstructed duct (that is, no dampers, junctions, or turns) from the opening. The opening of the duct also must be unobstructed by a grille or louver.

Figure 33. Duct elements affect sound

• Attenuators: Straight duct, elbows, junctions, silencers, terminations

• Regenerators: Elbows, junctions, dampers, silencers, diffusers, internal rods/bars

Figure 34. All duct changes reduce sound transmission



• Variation C adds a tee to the straight duct in Variation B. The branches downstream of the tee consist of straight sections of 10-by-10-inch duct, which provide the following end reflection:

The reduction in duct size increases the attenuating effect from end reflection. Not included in the end reflection reductions itemized above are the additional reductions that will result from the fittings, lined duct, and separation of return duct openings.

For larger ducts, it may be helpful to further increase end-reflection attenuation by adding two more tees (Figure 36).

Receiver sound correction

Sometimes described as “room effect,” receiver sound correction represents the relationship between the sound energy (sound power) entering the room and the sound pressure at a given point in the room (where the receiver hears the sound). The “correction”—which is nearly always a reduction in each octave band—results from a combination of effects, including diffusion (spreading), and the absorptive and reflective properties of the surrounding surfaces.

In an outdoor environment, such as a field or parking lot, the reflection of sound is nearly zero. Sound leaves the source in all directions and diminishes as it travels away from the source. The only sound that reaches the receiver is the portion that travels in a direct line from the source to the receiver. Therefore, outdoor receiver sound correction is mainly a function of the distance between the source and the receiver.

By contrast, sound entering a room bounces off walls and other surfaces (Figure 37). The receiver not only hears the sound that originates directly from the source but also the sound that reflects from room surfaces. The amount of sound that reaches the receiver depends on the size of the room and the absorptivity of the surfaces within the room.

Modeling tools

Algorithms

Prediction equations based on test data and experience aid the analysis of sound-path elements. The algorithms collected and developed by the American Society of Heating, Refrigerating, and Air-Conditioning Engineers (ASHRAE) are widely used and generally provide good results.4

Most of these algorithms are based on measurements recorded for isolated elements. For example, the equations that predict the acoustical effect of duct elbows resulted from a regression analysis of data collected in an acoustical lab. The collected data represents a range of test configurations, each of

63 Hz 125 Hz 250 Hz 500 Hz 1000 Hz 2000+ Hz

–15 dB –10 dB –5 dB –2 dB 1 dB 0 dB

Figure 36. Multiple tees to reduce end reflection

Figure 37. Receiver sound correction



which was specifically designed for measuring the attenuated and regenerated sound of a particular elbow design at various airflow rates.

Note: Algorithms are based primarily on test data collected in a laboratory setting. Be aware that the accuracy of algorithms diminishes when they are used to extrapolate acoustical effects for conditions that deviate from the laboratory test setup.

Analysis software

Solving prediction algorithms manually can be tedious, time-consuming, and iterative—especially when one or more paths need further attenuation. Fortunately, software applications are available to perform these calculations quickly and accurately. Using acoustical analysis software also makes it easier to refine the source–path–receiver model, which can be a tremendous benefit when designing a new system or troubleshooting an existing building. For example, with the help of software, you can quickly determine the effects of using a duct silencer, changing the construction of the equipment-room wall, adding absorptive materials to a ceiling, or placing an acoustical barrier between an outdoor sound source and the property line.

Transfer functions

There are cases where the acoustical model for an equipment type is simple and varies little from one application to the next; as a result, the acoustical analysis for that equipment type is similar from job to job. In these situations, the acoustical effects represented by the path calculations can be summed into a single line of octave-band data: a transfer function.

Transfer functions permit a quick, general comparison between units and estimated sound-pressure levels at the receiver. Acoustical analysis is simply a matter of subtracting the transfer function from the sound-power data of the unit. Often the results of the transfer function are further simplified as a single-number NC or RC rating.

Note: Never use a transfer function without knowing how it was derived. A transfer function will not provide an accurate estimate of the sound-pressure levels at the receiving location if the proposed space differs from the model on which the transfer function was based.

Identifying sound paths

Sound can travel between a single source and the receiver along one or more paths. In the case of a receiver who is standing across the parking lot from a roof-mounted air-cooled chiller, only one path carries the chiller-generated

4 ASHRAE collected and developed numerous equations to predict the acoustical effects of sound-path components for HVAC systems and subsequently published them in their Algorithms for HVAC Acoustics handbook (ISBN 090110751, out of print). Similar information can be found in the National Environmental Balancing Bureau (NEBB) publication titled Sound and Vibration Design and Analysis. The 1999 ASHRAE Handbook–HVAC Applications replaced most of the prediction equations for sound and vibration with related tables and simplified design procedures.



sound. Similarly, a single path carries the sound from the fan–coil mounted under a window to the receiver sitting at a desk in the same room.

Multiple paths carry sound from most centralized air-handling equipment, such as rooftop and self-contained air conditioners (Figure 38). In the case of an air handler situated in an equipment room, sound from the fan reaches the receiver in the adjacent room along five discrete paths:

• Supply airborne, which carries fan sound through the supply ductwork and diffusers, and into the space

• Supply breakout, which carries fan sound through the walls of the supply duct, then through the ceiling tile, and into the space

• Return airborne, which carries fan sound through the air-handler intake, return ductwork and grilles, and into the space

• Return breakout, which carries fan sound through the walls of the return duct, then through the ceiling tile, and into the space

• Wall transmission, which carries fan sound through the air-handler casing, then through the adjoining wall, and into the space

The initial step in the modeling process is to identify all of the sound paths for each source of sound. Keep in mind that one piece of equipment may contain several sound sources. For example, a packaged rooftop air conditioner contains supply fans, exhaust or return fans, compressors, and condenser fans.

The sound at the receiver’s location is the sum of all of the sound from all sources along all paths. When all of the paths have been identified, each path can be modeled to determine its contribution to the total sound-pressure level at the receiver’s location. Identifying individual sound paths not only permits the creation of a source–path–receiver model, but it also allows comparison of the magnitudes of the paths. If the total sound-pressure level at the receiver’s location is too high, the designer can readily identify and refine the dominant sound path(s).

Figure 38. Common paths for HVAC sound

Mechanical vibration

Sound and vibration are closely related. Although this manual focuses on airborne sound, vibration must not be ignored. In some cases, structure-borne vibration becomes airborne when it reaches the right types of materials; in other cases, the vibration never becomes audible but irritates occupants and interferes with the operation of sensitive equipment.

Most problems caused by structure-borne vibration can be avoided through proper isolation, which in turn requires:

• analysis of the driving frequency and magnitude of the source;

• review of the building structure; and,

• selection of proper isolators.

It also is imperative to review all of the mechanical connections (hangers, piping, duct, and electrical conduit) to the unit. Such connections can “short circuit” the vibration isolators under the unit by transmitting vibration from the unit into the building structure.

Information about vibration and isolation is available from a number of sources. The ASHRAE Handbook series is a good place to start. For specific recommendations, check with acoustical (or vibration) consultants, equipment manufacturers, or vibration isolator suppliers.



Individually modeling the five paths in Figure 38 reveals that the supply airborne path is largely responsible for the total sound-pressure level in the space (Figure 39).

An example of sound-path modeling

To demonstrate the procedure and calculations for modeling the acoustical effect of an HVAC system, consider the mechanical equipment room and adjacent occupied space in Figure 40 (p. 46). Light fixtures with integral return air openings permit air from the office to pass through the drop ceiling into a common plenum; from there, the return air travels through a short section of ductwork and into the air handler. The main supply duct passes over the occupied space, which is served by a branch duct and two diffusers. We omitted the return-duct-breakout path from our model because the return duct is so short—leaving four paths conveying fan sound from the air handler into the occupied space:

• Supply airborne, which carries sound from the fan discharge into the occupied space with the supply air stream

• Supply breakout, which carries fan sound through the wall of the supply duct and into the occupied space

• Return airborne, which carries fan sound through the return-air duct, plenum, and grille (in the opposite direction of return airflow) and into the occupied space

• Wall transmission, which carries sound radiating out of the air-handler casing through the equipment-room wall and into the occupied space

Figure 39. Example of a multiple-path acoustical model



When modeling a sound path, always start with accurate sound-power values for the source and work progressively through each element of the path to the receiver. For this example, the sound-power data for the air handler is based on 7000 cfm (3.3 m³/s) of supply airflow and a static pressure of 2.5 in. wg (623 Pa).

Note: All of the sound calculations in this example were performed using Trane Acoustics Program™ (TAP) analysis software. TAP applies the algorithms and data published in the 1991 ASHRAE Handbook and in the NEBB Sound and Vibration Design and Analysis manual.

Supply airborne path

Figure 41 illustrates the elements (En) and corresponding acoustical data for the supply airborne path. In the table, “subsum” values separate logarithmic additions of sound power/pressure from reductions, which are subtracted arithmetically.

• E1 represents the discharge sound power for the air handler (the sound source), rated in accordance with ARI Standard 260.

• E2 and E3 represent the supply-duct elbow closest to the unit. An elbow both attenuates and regenerates sound, so each of these characteristics is modeled as a separate element.

The attenuation effect (E2) is calculated first and is subtracted directly from the unit sound-power values; the resulting reduction in fan sound is represented as a “sum” (line 3 of the table in Figure 41).

E3 (line 4) represents the sound levels generated by the elbow. These values are logarithmically added to the preceding “sum.” Notice that the decibel levels of the regenerated sound are more than 10 dB quieter than (and therefore are masked by) the initial decibel levels of the fan discharge.

• E4 represents the attenuating effect provided by the 15 ft section of straight duct.

Figure 40. Setting for sound-path modeling example

Programs, such as TAP, save time by automating complex acoustical calculations … but accurate results require accurate sound power data for each source component.



• The next elements in the path represent the X junction, or “takeoff,” which diverts supply air from the main trunk to the office. Like an elbow, a junction both attenuates (E5) and regenerates (E6) sound.

• E7 represents the attenuating effect of a 6 ft section of 10 in.-diameter, lined duct.

Figure 41. Supply airborne path (example of sound-path modeling)

Line

Path element Octave-band center frequency, Hz

ID Description 63 125 250 500 1000 2000 4000

1 E1 Supply-fan discharge: 7000 cfm, 2.5 in. wg TSP 94 91 81 80 79 77 71

2 E2 Elbow: 14H×56W (in.), 20-in. inside radius; unlined; 7000 cfm 0 0 –1 –2 –3 –3 –3

3 Subsum (arithmetic subtraction: Line 1 – Line 2) 94 91 80 78 76 74 68

4 E3 Regenerated sound from elbow E2 34 30 26 21 15 9 0

5 Subsum (logarithmic addition: Line 3 + Line 4) 94 91 80 78 76 74 68

6 E4 Duct: rectangular, 14H×56W (in.); straight 15-ft run; unlined –4 –3 –2 –1 –1 –1 –1

7 E5 Junction: X-type, follow 8-in. dia branch, 250 cfm –12 –12 –12 –12 –12 –12 –12

8 Subsum (arithmetic subtraction: Line 5 – Lines 6,7) 78 76 66 65 63 61 55

9 E6 Regenerated sound from junction E5 41 38 33 29 22 16 8


11 E7 Duct: circular, 8-in. dia; straight 6-ft run; 2-in. lining –3 –5 –8 –13 –13 –13 –11


13 E8 Diffusers: 8-in. round inlet; 250 cfm; NC 35 27 35 40 44 45 43 40


15 E9 Room correction: 8×20×40-ft room; 1991 ASHRAE equation; 2 diffusers, 8-ft avg source-to-receiver distance

–4 –5 –6 –7 –8 –9 –10

16 E10 Environmental adjustment factor –4 –2 –1 0 0 0 0

17 SUM (arithmetic subtraction: Line 14 – Lines 15,16) 67 64 51 46 43 40 35

NC 35, RC 43(R)



• E8 represents the sound generated by the diffuser. (It’s at this point that the supply-airborne sound enters the occupied space.) The diffuser manufacturer should provide this data, but if it is unavailable, use ASHRAE algorithms to estimate the sound levels contributed by the diffuser.

• E9 represents receiver room correction, which accounts for the decibel reduction that occurs between the sound power entering the room from a particular source and the sound pressure at a given point in the room.

For this example, Schultz’s equation was used to calculate the receiver room correction because it works well for predicting distributed sound from diffusers. The equation converts the sound power leaving the diffuser to sound pressure by accounting for the acoustical character of the room, the number of sound sources (two diffusers, in this case), and the distance from the sound source to the receiver:

where,Lp(5 ft)= sound-pressure level at a distance of 5 ft above the floor, dBLw(s) = sound-power level of the sound source, dBh = ceiling height, ftN = number of ceiling diffusersX = ratio of the floor area served by each diffuser divided by the ceiling height

squared (X = 1 if the area served = h²)

• And lastly, E10 applies an environmental adjustment factor, EAF, to Schultz’s equation to account for the behavioral difference between equipment sound-power data and the acoustical behavior of an actual occupied space.5

Line 17 predicts the sound-pressure levels that will result in the occupied space due to the supply airborne path.

Supply breakout path

(Figure 42) The supply breakout path shares the same source, elbow, and straight-duct elements as the supply airborne path (E1–E4). Together, these elements determine the sound-power levels inside the main trunk of the supply duct, which passes over the occupied space.

E5 (line 7) predicts how much of the sound inside the main supply trunk will resonate, or “break out” through the duct wall into the ceiling plenum. Duct-breakout sound is loudest near the sound source—the fan, in this case—because the sound levels in the duct typically diminish with distance. The distance at which duct-breakout sound may be a concern depends on the sound power of the fan, as well as the type of duct and the attenuation effects applied to it.

5 Acoustically, an actual occupied space behaves more like a reverberant room than an open space (free field). However, sound-power levels for air terminals tested in accordance with ARI Standards 880 and 890 are equivalent to free-field sound at low frequencies. The EAF, which corrects this discrepancy, is defined in ARI Standard 885–1998, Procedure for Estimating Occupied Space Sound Levels in the Application of Air Terminals and Air Outlets.

Determining receiver room correction

Receiver room correction varies with the furnishings, surface treatments, and geometry of the room, as well as with the distance between the sound source and the receiver. Several equations have been developed to predict the magnitude of the correction, depending on the nature of the sound source:

• The “regression” equation (Reynolds and Zeng) accounts not only for sound from a point source located anywhere in a room but also for a wide variation in the acoustical absorption of the room.

• The 1991 ASHRAE equation (Schultz) accounts for the sound from a distributed array of ceiling diffusers in typical office environments.

• The “diffuse field theory” equation (Thompson) accounts for sound from duct breakout in typical office environments.

Lp 5 ft( ) Lw s( ) 27.6 log10 h( ) 5 log10 X( ) 3 log10 f( ) 1.3 log10 N( ) 30+ +–––=



E6 (line 8) represents the transmission loss that occurs when the sound passes through the ceiling tile and into the occupied space.

The final element in the path is receiver room correction, (E7 in this case). Because duct-breakout sound emanates along the length of the duct rather than from distributed ceiling diffusers, the “diffuse field theory” equation is used to convert the sound power leaving the duct to sound pressure in the receiver room:

where,Lp = sound-pressure level at a specific point in the room, dBLw = sound-power level of the sound source, dBr = distance between the sound source and receiver, ftL = length of the line source, ftR = room constant, ft², which is defined as S = total surface area of the room, ft²avg alpha T= total average absorption coefficient of the receiver room, accounting for

room surfaces and air

Figure 42. Supply breakout path (example of sound-path modeling)

Line


ID Description 63 125 250 500 1000 2000 4000

1 E1 Supply-fan discharge: 7000 cfm, 2.5 in. wg TSP 94 91 81 80 79 77 71

2 E2 Elbow: rectangular, 14H×56W (in.), 20-in. inside radius; unlined; 7000 cfm

0 0 –1 –2 –3 –3 –3




6 E4 Duct: rectangular, 14H×56W (in.); straight 15-ft run; unlined –4 –3 –2 –1 –1 –1 –1

7 E5 Duct breakout: 14H×56W (in.), 20-in. inside radius; 20 ga. –4 –7 –10 –13 –16 –22 –28

8 E6 Ceiling-system transmission loss: 2×4-ft lay-in acoustical tile, 0.7 lb/ft² surface weight

–4 –8 –8 –12 –14 –15 –15

9 E7 Room correction: 8×20×40-ft room; “diffuse field theory” equation; 20-ft duct; 8-ft avg source-to-receiver distance

–9 –9 –8 –9 –10 –10 –10

10 SUM (arithmetic subtraction: Line 5 – Lines 6,7,8,9) 73 64 52 43 35 26 14

NC 53, RC 35(R)

Lp Lw 10log101

π r× L×--------------------- 4

R----+⎝ ⎠

⎛ ⎞ 10.5+ +=

S avg alpha T×( ) 1 avg alpha T–( )⁄[ ]



Note: Sound also emanates from the short lengths of round duct between the main trunk of the supply duct and the diffusers. We omitted these branch ducts from the path analysis because they contribute comparatively little breakout sound to our example room. The sound levels inside the branches are less than in the levels inside the main trunk; also, round duct is characterized by its high transmission loss, which makes it a good choice to attenuate a duct-breakout path.

Line 10 predicts the sound-pressure levels that the supply breakout path will contribute to the occupied space.

Return airborne path

(Figure 43) This path accounts for the sound that reaches the room by traveling in the opposite direction of airflow. Sound radiates from the inlet of the unit into the return duct, through which it travels to the ceiling plenum and then passes through the ceiling tile into the receiver room.

• E1 represents the fan sound power for the ducted inlet of the air handler, which is the sound source.

• E2 and E3, respectively, represent the attenuation and regeneration effects of the elbow closest to the air handler’s return-air opening.

• E4 accounts for the straight duct between the elbows.

• E5 and E6 represent the attenuation and regeneration effects of the elbow near the ceiling.

• E7 accounts for the straight duct that extends from the elbow, through the equipment-room wall, and into the plenum space.

• E8 represents the end reflection that occurs due to the abrupt termination of the return duct in the plenum. (Recall that end reflection, p. 40, reduces sound-power levels at low frequencies.)

• E9 accounts for the transmission loss that results when the sound passes through the ceiling tile into the room.

• E10 accounts for the acoustical characteristics of a room with sound-absorbing surfaces, such as an acoustical tile ceiling. For rooms smaller than 20,000 ft³, the Reynolds and Zeng “regression” equation predicts the decrease in sound-pressure levels from a point source located anywhere in the room:

where,Lp = sound-pressure level at a specific point in the room, dBLw = sound-power level of the sound source, dBN = number of sound sourcesr = distance between the sound source and receiver, ftb = . If b ≥ 2, then b = 2.mfp= mean-free-path of the room, ft, which = 4V/SV = room volume, ft³ (The equation shown here applies when V < 20,000 ft³.)S = total surface area of the room, ft²

Some building designs terminate the return-air duct at an opening in the equipment room. To create an acoustical model of this arrangement, use the “inlet plus casing” sound-power data for the unit. Add a path element called “sound transmission through a hole in the wall” to represent how much of the sound (radiating from the air handler into the equipment room) passes into the return-air duct.

Lp Lw 10 log10 rb( ) 10 log10 N( ) 3.6–+–=

0.009 mfp1.7 T⁄( )



Figure 43. Return airborne path (example of sound-path modeling)

Line


ID Description 63 125 250 500 1000 2000 4000

1 E1 Supply-fan ducted inlet: 7000 cfm, 2.5 in. wg TSP 88 81 67 73 71 69 64

2 E2 Elbow: square, 20H×60W (in.); lined; 7000 cfm 0 –1 –6 –11 –10 –10 –10




6 E4 Duct: rectangular, 20H×60W (in.); straight 8-ft run; 1-in. lining –2 –2 –4 –11 –20 –16 –15

7 E5 Elbow: square, 20H×60W (in.); lined; 7000 cfm 0 –1 –6 –11 –10 –10 –10

8 Subsum (arithmetic subtraction: Line 5 – Lines 6,7) 86 77 51 40 31 33 29



11 E7 Duct: rectangular, 20H×60W (in.); straight 10-ft run; 1-in. lining –2 –3 –5 –14 –25 –21 –19

12 E8 End reflection: 20H×60W (in.) duct terminates in free space –6 –3 –1 0 0 0 0

13 E9 Ceiling: 2×4-ft lay-in acoustical tiles, 0.7 lb/ft² surface weight; transmission loss

–4 –8 –8 –12 –14 –15 –15

14 E10 Room correction: 8×20×40-ft room; “regression” equation; avg source-to-receiver distance = 8 ft

–10 –10 –9 –10 –12 –12 –14

15 SUM (arithmetic subtraction: Line 10 – Lines 11,12,13,14) 64 53 28 < 5 < 5 < 5 < 5

NC 40, RC 5(RH)



Wall transmission path

The wall-transmission path (Figure 44) accounts for the sound that passes through the wall separating the adjacent equipment room and occupied space in this example.

• E1 represents the sound power radiated through the casing of the air handler.

In this example, the return-air duct connects directly to the inlet of the air handler. If the air-handler inlet is unducted, use the “inlet plus casing” sound-power levels.

• E2 accounts for the change in sound that occurs between the air handler and the equipment-room wall. This change is not the same as a room correction, but it, too, depends on the size of the room, the distance between the sound source and the wall, the sound-radiating characteristics of the source, and the amount of absorption in the room.

• E3 represents the transmission loss of the shared wall between the two rooms.

• E4 accounts for the acoustical character of the room. In this case, the calculation treats the entire surface of the wall as the sound source.

Figure 44. Wall transmission path (example of sound-path modeling)

Line


ID Description 63 125 250 500 1000 2000 4000

1 E1 AHU-casing-radiated sound power: 7000 cfm, 2.5 in. wg TSP 90 79 69 66 63 56 51

2 E2 Change in sound power from AHU to wall: 12×20×20-ft equipment room

1 2 5 4 4 4 4

3 Subsum (arithmetic addition: Line 1 + Line 2) 91 81 74 70 67 60 55

4 E3 Wall transmission loss: gypsum wallboard; 2 layers, each with fiberglass fill

–22 –30 –39 –40 –40 –39 –40

5 E4 Room correction: 8×20×40-ft room; 5-ft avg source-to-receiver distance; average absorption

–8 –8 –8 –9 –9 –9 –9

6 SUM (arithmetic subtraction: Line 3 – Lines 4,5) 61 43 27 21 18 12 6

NC 36, RC 17(R)



Combining the path results

Remember that the sound that you hear in a room is the sum of all of the sounds that enter it. Therefore, the contributions of the individual sound paths for each sound source must be summed (Table 10). The resulting octave-band data can be plotted on an RC chart (Figure 45) or NC chart (Figure 46, p. 54). From the plots for our example, we can see that the predicted background sound in the occupied space is RC 44(R) and NC 55.

The next step is to compare this prediction with the desired sound level for the space to determine whether the acoustical performance is acceptable.

Table 10. Total sound-pressure level by octave band (example: AHU with FC fan)

Sound path

Octave-band center frequency, Hz

63 125 250 500 1000 2000 4000

Supply airborne 67 64 51 46 43 40 35

Supply breakout 73 64 52 43 35 26 14

Return airborne 64 53 28 5 5 5 5

Wall transmission 61 43 27 21 18 12 6

SUM 74 67 55 48 44 40 35

NC 55, RC 44(R)

Figure 45. Resulting RC rating for combined sound paths(modeling example)



Redesigning a sound path

When the sound level predicted by the acoustical model compares unfavorably with the target for the space, then the design must be modified in an iterative process that changes the equipment, the path(s), or both. In this context, the model now serves two purposes: It can be used to examine the contribution of each sound path and to predict the effect of each path modification.

Let’s continue our example. Suppose that the desired sound level for the occupied space is RC 45(N). Our model predicted a sound level of RC 44(R), which meets the numerical requirement but indicates excessive sound in the low-frequency bands.

An acoustical redesign begins with an examination of the RC plot in order to determine which octave bands are responsible for the (R) designation. In this case (Figure 45, p. 53), the sound levels in the 63 Hz and 125 Hz octave bands exceed the upper boundary of the RC 45 curve; they must be reduced to achieve the desired “neutral” (N) spectrum.

The next step is to review the summary data for each sound path and note which sounds paths are contributing to the sound levels in the offending bands. From the path summary for our example (Table 10, p. 53), it’s apparent that the supply breakout path is the primary contributor; but the supply airborne path also requires consideration. How can we reduce the sound levels contributed by these paths?

Generally speaking, there are two ways to reduce HVAC sound: Attenuate the source or attenuate the sound path.

Figure 46. Resulting NC rating for combined sound paths(modeling example)



Attenuate the sound source

For the air handler in our example, replacing the forward-curved (FC) supply fan with a vaneaxial fan substantially lowers the sound in the 63 Hz and 125 Hz octave bands. Substituting the vaneaxial fan’s acoustical data into the model changes the resulting overall sound level to RC 45(N); see Table 11. Figure 46 and Figure 47 compare the original sound levels with the those of the redesigned system.

In this case, by changing only the supply fan and retaining the rest of the original system design, the acoustical model now predicts that the sound level in the space will meet the RC goal.

Note: Switching to a vaneaxial fan eliminated the “rumbly” (R) quality of the original design but increased the RC numerical value. Much of the sound produced by vaneaxial fans occupies the middle frequencies rather than the low frequencies that are typical of other fan types.

Table 11. Total sound-pressure level by octave band(example: AHU with vaneaxial fan)

Sound path

Octave-band center frequency, Hz

63 125 250 500 1000 2000 4000

Supply airborne 60 58 54 47 46 40 36

Supply breakout 66 58 55 45 38 26 15

Return airborne 50 59 42 18 5 5 5

Wall transmission 48 44 31 28 28 29 27

SUM 67 63 58 49 47 40 37

NC 50, RC 45(N)

Figure 47. Resulting RC rating for combined sound paths(modeling example)



Attenuate the sound path

The options for altering the acoustics of a sound path are virtually limitless, and can include relocating the source, adding acoustical treatments, rerouting ductwork, or some combination of these changes plus others. The other option for reducing sound at the receiver is to alter the system design in one or more ways. Determining which changes are appropriate for a particular application will depend on the type of sound path. Sometimes a single simple change can produce dramatic results; in other cases, achieving the target sound level will require several alterations.

Deciding whether to attenuate the path, the source, or both requires a “value/benefit” analysis. The acoustical model plays an important role in this analysis because it shows the acoustical benefit for each change. Determining what changes are appropriate will depend, in part, on the system type and application. For guidance, seek the advice of an acoustical consultant. A Practical Guide to Noise and Vibration Control for HVAC Systems, an ASHRAE publication authored by Mark E. Schaffer, is another helpful resource.

Figure 48. Resulting NC rating for combined sound paths(modeling example)


GoodAcoustical Design

Good acoustical design entails four critical elements:

• An appropriate target for background sound

• Accurate acoustical data for sound-generating equipment

• An acoustical source–path–receiver model

• Conscientious implementation of the design during construction

Let’s take a closer look at each of these elements.

Appropriate acoustical target(s)

The number of variables in an acoustically appropriate design precludes a one-size-fits-all formula. But a “good” listening environment is achievable with careful attention to the intended use and construction of each space in the building. Early collaboration of all interested parties—building planners and occupants, as well as architects, contractors, and suppliers—makes it possible to clarify expectations and to set an acceptable acoustical target that best balances utility and budget. An acoustical analysis then can determine whether the proposed system and its layout will meet that goal.

What’s considered “acceptable” sound varies dramatically with the intended use of the finished space. Obviously, a factory requires a less stringent acoustical target than a church, while an office has a different set of requirements altogether. What may be less apparent is that none of these buildings—the factory, the church, the office—is entirely homogenous. The acoustical target for a lobby will differ significantly from the target for a sanctuary or conference room.

The activities that will occur in a particular space dictate the acceptable level of background sound. In an open-plan office (for example, in a bank where credit officials are seated in rows of desks), a certain level of background sound is necessary to provide conversational privacy. If such spaces are too quiet, masking sound or physical barriers must be added to provide the desired level of acoustical isolation.

To help building owners and designers establish appropriate targets for HVAC-related background sound, the ASHRAE Handbook–HVAC Applications outlines typical RC ratings for various types of occupancies (see Table 4, p. 17, in this manual). Of course, it can be difficult for a building owner or tenant to relate an abstract RC or NC value to an actual level of background sound. In such cases, it may be helpful to find and survey a similar existing space with an acceptable acoustical environment; the octave-band levels or single-number rating from the survey can be used as the design target.


Good Acoustical Design

It is not enough to state that the spaces within the building should be “quiet.” One of two outcomes typically results from such an ambiguous design target:

• A costly retrofit to attenuate background sound in the noisy areas of the building; or …

• An “over-designed” building that is quieter, and more costly, than necessary.

Classroom acoustics:Implementing a new standard*

Acoustical performance is an important consideration in the design of classrooms. Research indicates that levels of background noise and reverberation little noticed by adults, who are mature and skillful listeners, adversely affect learning environments for young children, who require optimal conditions for hearing and comprehension. Poor classroom acoustics are an additional educational barrier for children who have hearing loss and those who use cochlear implants, since assistive technologies amplify both wanted and unwanted sound. Children who have temporary hearing loss, who may comprise up to 15% of the school age population according to the Centers for Disease Control (CDC), are also significantly affected, as are children who have speech impairments or learning disabilities. Kids whose home language is different than the teaching language are also at additional risk of educational delay and failure in classrooms that have poor acoustics.

In 1998, the U.S. Access Board joined with the Acoustical Society of America (ASA) to support the development of a classroom acoustics standard. Stakeholders from both public and private sectors, including industry, were involved. Their work has now been approved as ANSI/ASA S12.60-2002, Acoustical Performance Criteria, Design Requirements and Guidelines for Schools. Consistent with long-standing recommendations for good practice in educational settings, the new standard sets maximum limits for background noise (35 decibels) and reverberation time (0.6 to 0.7 seconds) for unoccupied classrooms.

Taken by itself, the standard is voluntary unless referenced by a State code, ordinance, or regulation. The Access Board’s initial proposal to reference the standard in the 2003 edition of the International Building Code (IBC)—which already includes acoustical requirements for multi-family housing—was tabled because of concerns from the modular

classroom and air-conditioning industries. Some school systems now require compliance with the standard as part of their construction documents for new schools, thus making the design team responsible for addressing the issues. Parents may also find the standard useful as a guide to classroom accommodations under IDEA (the Individuals with Disabilities Education Act). States, local jurisdictions, and boards of education that have taken action on classroom acoustics [as of July 2005] are listed below.

Adopted ANSI/ASA S12.60-2002

• New Hampshire State Board of Education

• New Jersey School Construction Board

• Ohio School Facility Commission

• Connecticut

Other classroom acoustics standards/directives in use

• New York State Department of Education

• Los Angeles Unified School District

• Minneapolis Public Schools

• Washington State Board of Health

• Washington, DC Public Schools

• California Collaborative for High-Performance Schools (CHPS)

Standards/guidelines in development

• Maryland State Department of Education

• Minnesota

International standards/guidelines

• UK

• Sweden

• Italy

• Switzerland

• World Health Organization (WHO)

* Excerpt from the U.S. Access Board’s “Classroom Acoustic Fact Sheet” [online]. July 2005 [cited 19 September 2005]. <http://www.access-board.gov/acoustic/>.



Resolving complaints about a noisy HVAC system is particularly difficult. Unhappy occupants become sensitized to HVAC-related sound and listen for it, consciously or not. In such situations, the existing sound level typically must be reduced far below what normally would be acceptable.

Post-construction remediation of background noise is costly, too. Replacing components in a finished space is more costly than installing the right materials during construction. Nor is the expense limited to labor and materials; lost productivity during remediation may impose additional costs.

Accurate sound data

Acoustical prediction depends on accurate sound data. It is the very foundation of acoustical modeling. Therefore, it is important to discover the “pedigree” of the sound data for prospective equipment and to determine if it can be used as is, or if it must be adjusted to account for potential inaccuracy. ASHRAE’s Application of Manufacturers’ Sound Data (Charles Ebbing and Warren Blazier, 1998) can be a useful guide.

Acoustical modeling

The purpose of acoustical modeling is to design a system that meets the acoustical target … but to create the model, the design must be sufficiently far along to enable modeling. For that reason, acoustical modeling is best performed when the building design is near completion but not yet finalized.

There are many variations on the timing and approach to acoustical modeling. One approach determines the maximum equipment-generated sound levels by calculating the contribution of each sound path and “adding” these contributions to the acoustical target for the room. This method is commonly used in retrofit projects because it is difficult to alter existing sound paths, but success depends on selecting equipment that meets the acoustical objective. Such equipment may not be available, affordable, or practical for the application.

Another approach to acoustical modeling treats the sound-generating equipment and sound-transmission paths as a “system.” Modeling starts with the equipment sound data and applies the appurtenance effects along each sound path to determine the resulting sound levels in the space. If the design incorporates proper acoustical practices and treatments that are known to work, then the resulting sound levels in the space should be close to the acoustical target.

Ideally, the system is modeled after the initial decisions are made concerning equipment and design. If the predicted sound levels are unacceptable, the timing of this modeling approach makes it easy to identify the critical sound path(s) and octave bands that require attenuation.

Early coordination with the architect is instrumental to creating an appropriate acoustical environment. Spaces located farthest from the HVAC equipment provide the most path attenuation and, therefore, will be quietest;



consequently, these spaces are well-suited for executive offices, conference rooms, and other sound-critical areas. Conversely, spaces with lenient acoustical targets—bathrooms, stairways, and utility closets—are best situated around mechanical rooms or under roof-mounted equipment.

Having identified the critical sound paths and octaves, alternate equipment selections or path designs can be modeled to determine the most cost-effective means of meeting the acoustical target for each space.

A large project or an especially critical acoustical goal may warrant the construction of a sample room and an operational equipment test. Although the mockup test is expensive, it confirms equipment performance and the resulting sound levels in the finished space before construction begins, and can save significant time and expense.

Conscientious implementation

Attaining the acoustical target in a given space requires more than accurate acoustical modeling of a well-thought-out building design. Every aspect of that design also must be implemented as it was modeled.

Acoustically critical details may seem insignificant from the standpoint of construction. However, the transmission loss provided by a wall, for example, depends largely on how well the penetrations through the wall are sealed.

Cost-cutting during the “value engineering” phase of the project also threatens the acoustical design of the building. The entire design-and-construction team must realize that although there may be less costly ways to achieve the same mechanical characteristics, the alternatives seldom result in the same acoustical characteristics. For example, wrapping the exterior of the duct with insulation is less costly than lining the duct: Although the external insulation provides the same thermal function, it does nothing to attenuate sound.

Constructing a sample room with functioning HVAC equipment not only is useful for acoustical modeling, but also can save the installing contractor considerable time and money on large projects. Construction errors detected in the sample room can be corrected before they are repeated throughout the building.

Acquiring the right equipment

“Adhering to construction details” includes specifying (and purchasing) equipment that meets the acoustical goals of the design. Developing a specification for equipment sound presents two difficulties:

• Writing an unambiguous equipment specification

• Qualifying the equipment manufacturers



Specifying equipment sound. Sound-related design goals are described in terms of what is required in the space, not what the equipment produces. As a further complication, goals for space sound are expressed in sound pressure while equipment is rated in sound power.

Fortunately, acoustical modeling “bridges the gap” between the sound power produced by an HVAC unit and the resulting sound pressure in the space. By the conclusion of the acoustical analysis, we will know the desired equipment sound-power levels.

Equipment sound-power levels must be correctly specified to prevent misunderstandings between what is required and what is supplied. A properly written specification will include:

• Required sound-power levels by octave

• Separate octave-band listing for each aspect of the generated sound

• Reference sound-power level

• Industry standard(s) to be used for the equipment sound ratings

The following example demonstrates a properly written specification for an air-handling unit:

Sound power levels for the unit shall be determined in accordance with ARI 260–2001, and shall not exceed the values listed in the table below at design conditions. Sound power data shall be listed, unweighted, in decibels at a reference value of 1 picowatt (dB re 1 pW):

Qualifying prospective manufacturers. “Qualification” is a matter of assessing the accuracy of the equipment sound ratings. Accurate sound data can only be obtained in a properly equipped acoustical facility with a well-trained staff. Although the process can be time-consuming, the best way to qualify prospective manufacturers is to visit their manufacturing and testing facilities. When sound is a critical aspect of the design, either limit the candidates to reputable firms with good acoustical facilities or arrange for the equipment to be tested by a qualified, third-party laboratory before making the purchase.

Equipment sound-power ratings

Octave band, Hz

63 125 250 500 1000 2000 4000 8000

Ducted discharge 99 97 97 95 90 88 84 80

Inlet plus casing 101 99 92 92 90 89 81 75

Contracting for sound

When the contractual requirements of a project include sound levels, the specification must be sufficiently detailed to permit enforcement. The 1999 ASHRAE Handbook–HVAC Applications advises that the specification clearly stipulate:

• What sound levels—equivalent (Leq) or maximum (Lmax) continuous sound pressure levels*—will be measured in each octave band

• Where and how the sound levels will be measured (for example, a space average over a defined area or at specific points)

• What instruments will be used to make the sound measurements (ASHRAE recommends the use of ANSI Type 1 or Type 2 sound-level meters with octave-band filters.)

• How the sound-measuring instruments will be calibrated

• How the sound-measurement results will be interpreted

* See Chapter 7 of the 2005 ASHRAE Handbook–Fundamentals for further details.


Appendix A:Working with Decibels

Using a logarithmic scale to measure sound levels means that the resulting decibel values cannot be combined arithmetically. Instead, logarithmic addition, subtraction, or averaging must be used to combine two or more sound levels. This process involves converting the decibel values into ratios of sound intensity, adding these ratios, and then reconverting the sum into decibels.

Adding decibel values

It is often necessary to combine decibel sound levels of common derivation. Adding decibel values is a three-step process:

1 Determine the antilogarithm for each decibel value.

2 Add the antilog values together.

3 Determine the logarithm of the sum.

The following example demonstrates the addition of four decibel values (76, 82, 90, and 89):

This method can be somewhat cumbersome, especially if a quick estimate is sufficient. The graph in Figure 49 simplifies the addition of two decibel values. To use the graph, plot the difference between the two values on the x axis; the y axis shows how much to add to the larger of the two values. Only two decibel values can be combined at a time; to add more than two decibel values, add the sum of the first two values to the next value.

If we apply this method to the previous example, in which we added 76, 82, 89, and 90:

Note: Decibel levels are usually presented as integers. Although it is appropriate to use decimal fractions during addition, round the final sum to the nearest whole number.

dBsum 10 log10× 10dB1 10⁄( )

10dB2 10⁄( )

10dB3 10⁄( )

…+ + +=

dBsum 10 log10× 10 7.6( ) 10 8.2( ) 10 9.0( ) 10 8.9( )+ + +[ ]=

10 log10×= 1.99 109×[ ]

93 dB=

Figure 49. Adding decibels logarithmically

82 76– 6 1 82 83=+∴=

89 83– 6 1 89 90=+∴=

90 90– 0 3 90 93 dB=+∴=

Appendix A: Working with Decibels


Table 12 further simplifies the addition of decibel pairs. Both shortcuts (Figure 49 and Table 12) are reasonably accurate, particularly when adding only a few decibel values.

Subtracting decibel values

To determine how the HVAC system will affect a space acoustically, it is sometimes necessary to subtract one decibel level from another—for example, to find out how much of the background sound results from operating the HVAC system. The following equation illustrates the process:

Background sound refers to the sound-pressure levels in a particular environment; it includes all sound sources except for the one being investigated or measured.

Note: Many industry standards for measuring equipment sound require a background sound level that is at least 10 dB less than the combined sound level (background plus equipment). For this reason, do not use subtraction to determine equipment sound unless the combined sound level exceeds the background-sound level by at least 3 dB.

Averaging decibel values

When sound is measured several times at the same location, it can be useful to average the readings into a single value:

Table 12. Adding pairs of decibel values

Difference between two values

Add to highest value

0a or 1 dB

aThe logarithmic sum of two equal sound levels (a difference of 0 dB) will increase the combined sound level by 3 dB.

3 dB

2 or 3 dB 2 dB

4 to 9 dB 1 dB

10 dB or more 0 dB

where,dBT = total sound level, which is a combination of background sound dB2 and

equipment sound dB1, measured in the spacedB2 = measured background sound onlydB1 = equipment sound

dB1 10 log10 10dBT 10⁄( )

10dB2 10⁄( )

–×=

dBavg log10 10dB1 10⁄( )

10dB2 10⁄( )

…10dBn 10⁄( )

++ 10 log10 n×–=


Appendix A: Working with Decibels

Either of two shortcuts can simplify the averaging process if reasonable accuracy is acceptable:

• If the difference between the decibel levels is less than 5 dB, then use an arithmetic average.

• If the difference between the decibel levels is between 5 dB and 10 dB, then add 1 dB to the arithmetic average.


Appendix B:Fan-Generated Sound

A significant portion of the background sound in a conditioned space results from operation of the HVAC system—particularly from air-moving components (fans, air terminals, diffusers, grilles, and ductwork) due to their placement within or adjacent to the space.

Obtaining useful sound-power ratings for air-handling equipment can be difficult, given the myriad of variables: fan types and arrangements, housing configurations, operating speeds, and application characteristics. Not surprisingly, the lack of reliable data has given rise to rules of thumb for estimating noise generation by fans and the air handlers housing them. Some of these rules of thumb give erroneous results, so it’s important to understand the factors that influence noise generation when selecting and applying air-moving equipment.

Factors that affect sound generation

Generally, the factors that affect a fan’s operating efficiency also affect the amount of sound that the fan produces. (Much less energy is lost to sound generation, however, than to inefficient performance.)

Air velocity

Most sound generated by fans results from the high-velocity air passing through the fan wheel and housing. Rule of thumb: Selecting a larger fan for a given set of conditions will lower the required air velocity and, therefore, reduce the amount of fan-generated sound.

Blade-pass frequency

As the fan rotates, air leaving each fan blade creates an “impact” when it strikes the cutoff at the fan discharge. If the cutoff is parallel to the fan blades, these “impacts” may become audible as tonal sound. The frequency at which the tone (or near-tone) occurs depends on the number of fan blades and the rotational speed of the fan wheel:

Depending on its magnitude, the tone may be apparent at the octave-band center frequency that coincides with the blade-pass frequency; however, tones usually are more detectable at 1/3 octaves.

ƒ number of fan blades fan speed, in rpm60 minutes

-----------------------------------------------------×=


Appendix B: Fan-Generated Sound

Operating point

Where the fan operates on its fan curve (Figure 50) also influences sound generation. As the operating point approaches the “block tight” condition (outside the cataloged range of normal performance), the amount of airflow through the fan is relatively small compared to the static pressure. The difference between the pressures at the fan’s inlet and discharge creates an unstable operating condition (called “stall”) in which the air rushes back and forth across the blades. The resulting pressure fluctuations produce audible sound.

The magnitude of stall—and the noise that accompanies it—varies with the design, type, and size of the fan. Forward-curved fans tend to enter stall gradually, whereas backward-inclined, airfoil, axial, and vaneaxial fans enter stall abruptly.

Application tips

• Oversized fans operating in stall (that is, well left of the peak on the performance curve) tend to be noisy, especially at low frequencies. Make sure that the expected operating conditions lie within the stable portion of the fan’s performance range.

• Undersized fans must operate at nearly full speed to deliver design airflow; high-velocity air generates more sound.

• Using sound-power data derived from actual measurements provides the best basis for selecting the “right” fan for a given application.

Flow disturbances

The unavoidable turbulence that occurs in the air stream as the air enters and leaves the fan also influences fan-generated sound. The difficulty for fan manufacturers (and, ultimately, system designers) is providing sound ratings that accurately represent how the fan will sound in a specific setting.

Fans may be tested in a stand-alone configuration (as defined by AMCA Standard 300) or in one of several test setups that more accurately represents the effects of appurtenances, such as the unit casing and attached ductwork (as defined by ARI Standard 260).

Shortcomings ofacoustical rules of thumb

An example can help to illustrate the limitations of rules of thumb as a means of selecting a fan based on sound. Table 13 outlines nine prospective fans, each of which satisfies the design condition. The table also summarizes their octave-band sound-power data, which was collected in accordance with ARI Standard 260, and numerically ranks the prospective selections based on individual comparisons of tip speed, outlet velocity, brake horsepower, static

Figure 50. Fan performance curve (typical)

Appendix B: Fan-Generated Sound


efficiency, and octave-band analysis. From the rankings that result from these comparisons, we can conclude that:

• Tip speed is not a good indicator of fan-generated sound. In this example, the fan with the slowest tip speed (17A) is one of the noisiest selections (fifth out of nine).

• Outlet velocity, brake horsepower, and static efficiency can more reliably predict fan-generated sound than tip speed.

• An octave-band analysis provides the most accurate estimate of equipment-generated sound, making it easier to identify the quietest selection with greater assurance of success.

Put simply, rules of thumb can be useful for preliminary comparisons of equipment, but assuring that the air-handling system contributes to an acoustically appropriate environment requires both accurate sound-power data and acoustical modeling.

Table 13. Prospective fan selections for example of acoustical comparisons

Air-handler size, arrangement 14A 14E 17A 17D 21A 25A 25D 17Q 21HP

Wheel size, in.; fan typea

a FC = forward-curved, Q = vaneaxial, BI = backward-inclined

20FC 16FC 20FC 18FC 22FC 25FC 22FC 27Q 20BI

Operating characteristics at design condition

Tip speed, ft/min 3,812 4,023 3,780 3,856 3,789 3,821 3,818 11,350 8,552

Revolutions/minb

b Rotational speed at 10,000 ft³/min and a total static pressure (TSP) of 2.0 in. wg.

727 959 721 819 657 583 662 1,605 1,631

Outlet velocity 1,994 2,942 1,994 2,411 1,939 1,480 1,939 2,202 1,975

Brake horsepower 6.0 8.2 5.8 6.7 5.3 5.2 5.2 5.5 7.1

Static efficiency 52 38 53 47 59 60 60 58 44

Sound-power level by octave-band center frequency, ref 10-12 watt

63 Hz 94 98 94 96 87 87 88 85 104

125 Hz 92 95 91 93 87 87 88 85 97

250 Hz 86 89 86 87 79 78 79 83 95

500 Hz 83 86 83 84 78 76 78 81 86

1,000 Hz 79 86 79 82 74 72 74 80 86

2,000 Hz 78 85 78 81 70 68 70 76 81

4,000 Hz 73 81 73 77 65 63 66 71 77

8,000 Hz 66 75 66 70 60 58 60 64 72

Basis of comparison Ranking of sound level (1 = quietest, 9 = noisiest)

Tip speed 3 7 1 6 2 5 4 9 8

Outlet velocity 5 9 6 7 2 1 3 8 4

Brake horsepower 6 9 5 7 3 1 2 4 8

Static efficiency 6 9 5 7 3 1 2 4 8

NC by full-octavec

c Example octave-band analysis for a typical ducted installation. After applying the appropriate resulting transfer function to each fan’s sound-power levels, the sound-pressure values were plotted on an NC chart to obtain a comparative ranking based on the noise criteria (NC) curves.

6 8 5 7 3 2 4 1 9


Glossary

absorbed sound. Sound energy that strikes a material and is converted from sound energy to heat energy within the material.

absorption coefficient. Fraction of the incident sound that is absorbed by the material, usually given in octave bands.

acoustic nearfield. The area adjacent to the sound source, where sound waves do not behave predictably (as they would in a “free field”) because they do not radiate equally from the source in all directions.

ambient sound. The background sound of an environment in relation to which all foreground sounds are heard, such as the “silence” of an empty room, conversation in a restaurant, or the stillness of a forest. “Ambient sound” typically consists of many small sounds from near and far sources, none of which is particularly dominant.

AMCA. Air Movement and Control Association (www.amca.org)

amplitude. Maximum amount by which the instantaneous sound pressure differs from ambient pressure.

appurtenance. An accessory. In the case of fans, such accessories include inlet boxes, inlet box dampers, variable inlet vanes, outlet dampers, vibration isolation bases, inlet screens, belt guards, diffusers, and sound attenuators.

ARI. Air-Conditioning and Refrigeration Institute (www.ari.org)

ASHRAE. American Society of Heating, Refrigerating, and Air-Conditioning Engineers (www.ashrae.org)

attenuation. The reduction of sound level, per unit of distance, due to divergence, diffusion, absorption, and/or scattering.

audible frequency range. Sound frequencies normally heard by the human ear, which span 20 Hz to 20,000 Hz. In the context of HVAC systems and buildings, most acoustical investigations only consider frequencies ranging from 40 Hz to 11,000 Hz.

A-weighting. A single-number rating method, expressed as dBA, that is used to describe the intensity of sound. It uses weighting factors, by octave band, to approximate human response to sound in the range where no hearing protection is needed. It is most appropriately used for low-volume (quiet) sound levels.

background sound. Sound-pressure levels in a given environment from all sources except the specific sound source under investigation.

balanced noise criteria (NCB). A single-number rating method for evaluating or specifying room sound, including noise resulting from occupant activities.

Glossary


Compared to the NC method (which it was meant to replace), it: adds the low-frequency 16-Hz and 31.5-Hz octave bands; lowers the permissible sound levels in the 4000-Hz and 8000-Hz octave bands; and evaluates rumble and hiss. The NCB rating procedure is based on the speech interference level (SIL), which considers the average of the sound pressure levels centered in four octave bands—500 Hz, 1000 Hz, 2000 Hz, and 4000 Hz. See also noise criteria (NC).

balanced spectrum. Balanced distribution of sound energy across all frequencies of the audio range, resulting in an unobtrusive “neutral” character; also called “spectral balance.” See also unbalanced spectrum.

blade pass frequency (BPF). Mathematical product obtained by multiplying the number of blades or vanes mounted on a fan wheel by the speed of rotation.

blade passage tone. Tone produced by the rotation of the blades on a fan wheel or pump impeller within an equipment housing; sometimes erroneously called a “pure tone.”

broadband sound. Spectrum that consists of acoustical energy at many frequencies, none of which is individually dominant.

center frequency (ƒc). The single frequency that is used to identify a specific octave band. It is calculated by finding the square root of the product of the lowest and highest frequencies in that octave band.

compression. Increase in density and pressure in a medium, such as air, caused by the passage of a sound wave. See also rarefaction.

decibel (dB). A dimensionless ratio of two quantities that is used to describe both sound power and sound pressure. It is defined as 10 times the logarithm to the base ten (log10) of the measured quantity divided by the reference quantity.

dynamic insertion loss. The insertion loss provided by a duct silencer when air flows through it. See also insertion loss.

free field. A homogeneous, isotropic medium, free from boundaries. A practical example of a free field (situated over a reflecting planar surface) is a large open area, such as a parking lot or meadow, without obstructions. See also reverberant field.

free-field method. A common method for acoustically testing HVAC equipment that is too large to be tested in a reverberant room (water chillers and cooling towers, for example). The equipment is placed outdoors on a hard surface in a parking lot, which approximates a free field above a reflecting plane. Sound-pressure waves travel evenly away from the equipment in a hemispherical pattern. Sound-power levels are determined by measuring the sound-pressure levels on an imaginary hemispherical surface surrounding the equipment.


Glossary

frequency. The number of complete back-and-forth vibrations (compressions or rarefactions) of a particle of the medium per unit of time; commonly expressed in hertz (Hz).

insertion loss (IL). The difference in sound pressure measured at a single location, with and without a noise-control device installed between the sound source and the receiver. See also dynamic insertion loss.

line source. A sound source that generates equal sound-power levels per unit of length, resulting in a cylindrical spreading of sound; usually associated with duct breakout. Sound level decreases by 3 dB with each doubling of distance from the line source. See also point source.

loudness. A subjective quality that is related to sound intensity, but which is affected by sound frequency and the ability of the listener’s ears to respond. Simplistically, sounds of greater intensity will be perceived as louder.

nearfield. See acoustic nearfield.

NCB. See balanced noise criteria.

NEBB. National Environmental Balancing Bureau (www.nebb.org)

noise. Sound that is obtrusive, unwanted, and/or objectionable because it interferes with communication, performance, or sleep, or adversely affects health.

noise criteria (NC). A single-number rating method that describes the intensity of sound in a room. Sound pressure, by octave band, is plotted on a series of curves, which can then be used to determine the NC value. See also balanced noise criteria (NCB).

noise reduction (NR). Measurement of the extent to which a barrier reduces the amount of transmitted sound. It is the difference between the sound pressures measured on each side of the barrier.

noise reduction coefficient (NRC). A single-number descriptor that represents the sound-absorption characteristics of a material. It is determined by the mathematical average of the absorption coefficients for the 250 Hz, 500 Hz, 1000 Hz, and 2000 Hz octave bands.

octave band. Frequency range in which the upper limit is twice the value of the lower limit, and which is usually identified by its center frequency (geometric mean).

path. The environment through which sound travels as it propagates from the source to the receiver, including everything that a sound wave encounters on its journey from the source to the receiver. There are three types of sound paths: airborne, breakout, and transmission.

phon. A logarithmic unit of measure that describes the loudness of a sound. See also sone.

Glossary


pitch. A subjective quality that is primarily determined by frequency, but that also is dependent on sound-pressure level and composition. Pitch is not measured; instead, it is described by subjective terms, such as bass, tenor, alto, and soprano.

point source. A sound source that radiates sound equally in all directions, resulting in a spherical spreading of sound. Sound level decreases by 6 dB with each doubling of distance from the point source. In acoustical analysis, typical point sources of sound include grilles, registers, and diffusers; air valves, fan-powered air terminals, and fan–coils installed in ceiling plenums; and return-air openings. See also line source.

propagation. Process by which a disturbance, such as sound waves, is transmitted through a medium, such as air or water.

pure tone. Sound that is emitted at a single frequency (pure tone). See also tone.

rarefaction. Decrease in density and pressure in a medium, such as air, caused by the passage of a sound wave. See also compression.

receiver. One or more observation points at which sound is evaluated (usually near the ear) or measured.

receiver sound correction. The relationship between the sound energy (sound power) entering the room and the sound pressure at a given point in the room, where the receiver detects the sound.

reflected sound. Sound that bounces off a boundary, such as a wall.

regenerated sound. Sound produced by the turbulence that results from placing a device in a stream of moving air or fluid. The magnitude of the regenerated sound depends on the size and shape of the device, and on the velocity of the stream.

reverberant field. An environment that produces uniform, or diffuse, sound. In a perfectly reverberant field, sound pressure is equal at all points. See also free field.

reverberant room. Specially constructed room with hard-surfaced walls, ceiling, and floor to reflect sound. When a sound source is placed in the room, the sound waves bounce back and forth multiple times. In a perfectly reverberant room, the resulting sound pressure is equal throughout the space.

reverberant-room method. Common method for testing HVAC equipment within a specially constructed room that is designed to reflect as much sound as possible. The hard surfaces of the walls, floor, and ceiling create a uniform, or diffuse, sound field by inducing and mixing multiple reflections of sound waves. Sound pressure is essentially the same at all locations in the room. Sound-pressure measurements for HVAC equipment tested in a


Glossary

reverberant room are used to calculate the sound-power levels for that equipment.

room criteria (RC). Single-number rating method that describes the sound exposure in a room. Sound pressure, by octave band, is plotted on a series of curves and reference lines, which then can be used to determine the RC value and a subjective descriptor of the sound quality (neutral, rumble, hiss, perceptible vibration). See also room criteria Mark II.

room criteria Mark II. Single-number rating method that describes the sound performance of the HVAC system in its entirety. It was developed to replace the room criteria method and consists of three parts: criteria curves, a procedure for determining the RC numerical rating and spectral balance; and a procedure for estimating occupant satisfaction (Quality Assessment Index). See also room criteria (RC).

room effect. See receiver sound correction.

semireverberant field. An environment with some of the characteristics of both a free field and a reverberant field; that is, barriers (walls, ceiling) will prevent the sound from behaving in a free-field manner, but the barrier surfaces are not perfectly reflective. Instead, some of the sound is absorbed while the rest is reflected. See also free field and reverberant field.

SIL. See speech interference level (SIL).

sone. Unit of linear measure that describes the loudness of a sound. One sone is the linear equivalent to one phon. See also phon.

sound. Audible emissions resulting from the vibration of molecules within an elastic medium. It is generated either by a vibrating surface or the movement of a fluid. In the context of HVAC systems, the transferring medium can be either air or the building structure. For structurally transmitted sound to become audible, however, it must first become airborne.

sound path. See path.

sound power level (Lw). Acoustical energy that is emitted by the source, and that is neither affected by distance nor by the environment. Sound power cannot be measured directly; instead, it must be calculated from sound-pressure measurements.

sound pressure level (Lp). An audible disturbance in the atmosphere that can be measured directly. Its magnitude is influenced not only by the strength of the source, but also by the environment and the distance between the source and the receiver. Sound pressure is what our ears hear and what sound meters measure.

sound propagation. See propagation.

sound receiver. See receiver.

Glossary


source. Equipment or phenomenon causing the vibratory disturbance that results in sound waves.

source–path–receiver model. Systematic approach to analyzing sound in a space. The model traces sound from the source to a specific location where it will be heard by the receiver, and is used to predict how the sound will be affected by everything it encounters along that path.

spectral balance. See balanced spectrum.

spectral imbalance. See unbalanced spectrum.

speech interference level (SIL). Average of the sound pressure levels measured in the 500 Hz, 1000 Hz, and 2000 Hz octave bands.

tone. Sound that occurs within a narrow band of frequencies and that is significantly more noticeable than sound at adjacent frequencies. See also pure tone.

transmitted sound. Sound that passes through a barrier, for example, between different parts of a building or between adjacent buildings.

transmission loss (TL). Measurement that describes the extent to which a barrier reduces the amount of transmitted sound. It is the ratio of sound power on the source side of a barrier to the sound power on the receiver side of the same barrier.

unbalanced spectrum. Distribution of sound energy that results in an undesirable hiss, rumble, roar, or “whoosh”; also called “spectral balance.” See also balanced spectrum.

wavelength. The distance that a (sound) pressure-wave disturbance travels along the medium during one complete wave cycle.


References

American Society of Heating, Refrigeration and Air-Conditioning Engineers, Inc. (ASHRAE). 2005. ASHRAE Handbook–Fundamentals, chapter 7. Atlanta, GA: ASHRAE.

⎯⎯⎯. 2003. ASHRAE Handbook–Applications, chapter 47. Atlanta, GA: ASHRAE.

⎯⎯⎯. 1998. Application of Manufacturers’ Sound Data. Atlanta, GA: ASHRAE.

Beranek, L. 1954. Acoustics. New York: McGraw-Hill Book Company, Inc.

Diehl, G. 1973. Machinery Acoustics. New York: John Wiley & Sons, Inc.

Lord, H., Gatley, W. and Evensen, H. 1980. Noise Control for Engineers. New York: McGraw-Hill, Inc.

National Environmental Balancing Bureau (NEBB). 1994. Sound and Vibration Design and Analysis. Rockville, MD: NEBB.

Schaffer, M. 1991. A Practical Guide to Noise and Vibration Control for HVAC Systems. Atlanta, GA: ASHRAE.

Trane: Guckelberger, D. and Bradley, B. 1996. “Specifying ‘Quality Sound’,” Trane Engineers Newsletter, volume 25-3.

Trane. 2001. Fundamentals of HVAC Acoustics, TRG-TRC007-EN. Air Conditioning Clinic series. La Crosse, WI: Inland Label & Marketing Services.


Index

aA-B-C weightings of sound 11–12absorption 34absorption coefficient 34acceptable sound 57accuracy of sound data 59acoustic nearfield 21, 22acoustical analysis software 43, 46acoustical design

ASHRAE guidelines 17classrooms 58implementation 60–61rules of thumb for fan selection 66–67setting a target for background sound 1, 31–32, 57–59

acoustical environment 30–31acoustical lab 23acoustical mockup 23, 60acoustical modeling

algorithms 42–43analysis software 43creating a model 33, 39–56timing and strategies 59–60transfer functions 43

adding decibel values 62–63airflow and fan sound 65, 66algorithms as acoustical modeling tools 42–43ambient sound 32–33amplitude 2anechoic room 24ANSI/ASA S12.60–2002 17, 30, 58

See also classroom acousticsApplication of Manufacturers’ Sound Data 59ASHRAE guidelines for HVAC background sound 17attenuation 40, 55, 56averaging decibel values 63–64A-weighted sound data 11–12

bbackground sound 30, 63

ASHRAE design guidelines 17maximum limit for classrooms 58privacy 30, 31subjective quality 15

balanced noise criteria (NCB) method 13–14See also noise criteria (NC) curves

balanced sound spectrum 32bel 3blade-pass frequency and fan sound 65broadband sound 6

ccharacteristics of sound 2–7classroom acoustics 17, 30, 58

See also ANSI/ASA S12.60–2002common paths for HVAC sound 44comparison of single-number sound ratings 10contracting acoustical performance 61correcting for receiver room sound 42, 48, 49, 50creating an acoustical model 39–56


Index

ddB. See decibel (dB)dBA. See A-weighted sound datadecibel (dB) 3

adding 62–63averaging 63–64subtracting 63

designing for good acoustics 57–61diffuse field theory equation 48, 49

See also receiver room correctiondiffusion 38distance effect on sound 20–21duct breakout 48duct elbows 46duct size and attenuation 42duct takeoffs 47

eend reflection 40–42environmental adjustment factor (EAF) 48equal loudness contours 9, 18–19equipment manufacturers, qualifying 61equipment sound

Application of Manufacturers’ Sound Data 59environmental effect 30–38measuring 63rating 20–25specifying 61standards for rating 26–29

ffan selection based on sound 65–67fan sound and duct breakout 48free field 20–21

with flat reflecting plane 23–24frequency 2–3, 6full spectrum analysis 7, 25

hharmonic coincidence 35hiss, background 15human hearing 8–9

maximum sensitivity 18HVAC equipment

acoustical effect 30–38Application of Manufacturers’ Sound Data 59industry standards for rating sound 26–29sound power ratings 24–25

iideal sound source 20identifying sound paths 43–45implementing an acoustical design 60–61incident sound energy 34, 36, 37

See also reflectionindustry standards for rating equipment sound 26–29insertion loss (IL) 35

lline source of sound 21loudness 8, 9

See also sound intensity level

Index


mmagnitude of sound 3–5mass law 35

See also transmission of soundmaterial characteristics affecting sound 33–38mechanical vibration 44mel-scale 6mockup testing 23, 60modeling acoustical performance

examples of sound paths 45–53timing and strategies 59–60tools 42–43, 46

nNC curves. See noise criteria (NC) curvesnearfield, acoustic 21, 22neutral background sound 15noise criteria (NC) curves 12–14, 32

See also balanced noise criteria (NCB) methodnoise reduction (NR) 36noise reduction coefficient (NRC) 34noise remediation 59

ooctave bands 7, 19one-third-octave bands 7, 25

ppaths of sound. See sound pathsperceptible vibration 15phon scale 18pitch 6, 8plotting sound pressure levels 12–15point source of sound 21post-construction noise remediation 59predicting ambient sound 33–38productivity 30properties of sound 2–7pulsating sound 10–11

qqualifying prospective equipment manufacturers 61quality assessment index (QAI) 17

rrating methods for equipment sound 24–25RC curves. See room criteria (RC) curvesRC Mark II method 15

See also room criteria (RC) curvesreceiver room correction 42, 48, 49, 50redesigning a sound path 54–56reducing wall-reflected sound 38reflection 37–38, 40–42regeneration 40, 46, 47regression equation 48, 50

See also receiver room correctionremediating background noise 59return airborne path 44, 45, 50–51

See also sound pathsreturn breakout path 44

See also sound pathsreturn-air duct terminations 41, 50reverberant field 22reverberant room 23reverberation limit for classrooms 58Reynolds and Zeng regression equation. See regression equationroom criteria (RC) curves 14–16


Index

room effect. See receiver room correctionround duct and transmission loss 50rules of thumb for fan selection 66–67rumble, background 15

sSchultz equation 48

See also receiver room correctionselecting fans based on sound 65–67semireverberant field 22setting an acoustical target 1, 31–32, 57–59SIL. See speech interference level (SIL)sone scale 18sound

ambient 32–33broadband 6characteristics 2–7fan-generated 65–67magnitude 3–5regenerated 40, 46, 47wave, frequency, and amplitude 2–3

sound data accuracy 59sound energy 33, 34sound fields 20–22sound intensity level 5, 24

See also loudnesssound paths

combining analysis results 53elements 33, 39identifying 43–45modeling example 45–53redesigning 54–56

sound power level 4–5ratings for HVAC equipment 24–25versus sound pressure 4

sound pressure level 4sound ratings comparison 10sound transmission class (STC) 36–37sound waves 2–3

standing waves 23, 38source attenuation 55source–path–receiver model 33, 39, 44specifying equipment sound 61spectrum quality 14, 32–33speech interference level (SIL) 14, 16speech privacy 30, 31standards for rating equipment sound 26–29standing waves 23, 38strategies for acoustical modeling 59–60structural vibration 2, 15, 44substitution method for rating equipment sound 24–25subtracting decibel values 63supply airborne path 44, 45, 46–48

See also sound pathssupply breakout path 44, 45, 48–50

See also sound paths

tThompson diffuse field theory equation 48, 49

See also receiver room correctiontones 6tools for acoustical modeling 42–43, 46Trane Acoustics Program (TAP) 46transfer functions for acoustical modeling 43transmission coefficient 36transmission loss (TL) 36, 50transmission of sound 2, 34–37turbulence and fan sound 66

Index


vvaneaxial fans, characteristic sound 55vibration 2, 44

wwall transmission path 44, 45, 52

See also sound pathswall treatment to reduce reflected sound 38wavelength of sound 3weightings, A-B-C 11–12when to create an acoustical model 59–60writing an equipment specification 60–61

xX junctions and sound 47

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02 acoustics.in.air.conditioning

Engineering

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