48435_08d

CHAPTER 8.2

NOISE CONTROL

Martin HirschornPresident, Industrial Acoustics Company,

Bronx, New York

8.2.1 INTRODUCTION

Is noise control engineering a science or an art? It is a bit of both.Acoustic theory helps explain the acoustic world we live in and enables us to

establish general design parameters for engineered noise control solutions and prod-ucts, but it does not always do so very accurately. For instance, it is impossible tocalculate the noise reduction of barriers, walls, enclosures, rooms, and silencers orthe propagation of sound waves over open surfaces with the degree of accuracyneeded for reliability. There are just too many variables. Consequently, we cannotrely on theory for more than directional indicators.

Optimum noise control solutions must therefore be based on engineered productswith performance characteristics obtained from repeatable laboratory tests and/orextensive field databecause if we overdesign, it costs too much money, and ifwe do not adequately provide for noise control, we may have an unacceptable job.For critical jobs, where there are significant uncertainty factors, model testing isessential. This may include power plants, aviation terminals and test facilities, in-dustrial factories, and air-handling units in high-rise buildings. Furthermore, apartfrom economics considerations, the structural, mechanical, aerodynamic, and ther-modynamic engineering aspects of the solution to a noise control problem are oftenmore complex than its acoustic components, so in many instances a multidiscipli-nary approach is essential.

This chapter is concerned primarily with basic acoustic engineering principlesand how they can be applied to solve noise problems inherent in HVAC systems.The chapter first discusses the theory of sound, with emphasis on the acousticengineering aspects, and then examines the nature of noise in HVAC systems andthe means available for controlling noise.

8.2.2 THENATUREOFSOUND

Sound is essentially the sensation produced through the ear by fluctuations of pres-sure in the adjacent air, and "noise" can be defined as sound that annoys, usually

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Copyright 1997 by The McGraw-Hill Companies Retrieved from: www.knovel.com

because the sound pressure level is too high. High noise levels not only interferewith direct voice communications and electronically transmitted speech; they arealso considered a health hazard in both the working and living environments.

Sound waves are propagated in air as compressional waves. Although compres-sional waves are generally caused by vibrations of solid bodies, they can also becaused by pressure waves generated by the gas discharge of a jet engine or thesubsonic velocities in an air-conditioning duct. When these waves strike solid bod-ies, they cause the bodies to vibrate, or oscillate.

To illustrate what happens, we can think of sound being generated by a pistonoscillating back and forth in an air-filled tube (Fig. 8.2.1). This action of the com-pressor causes the air molecules adjacent to the piston to be alternately crowdedtogether (or compressed) and then moved apart (or rarefied). The oscillation gen-erated by the piston in this manner is referred to as "simple harmonic motion."And as shown in Fig. 8.2.2, a plot of the piston displacement can be presented asa sinusoidal function; that is, the sound wave generated in its purest form for a

\= wavelength or shortestdistance between twosequential pressure crests in aplain wave oscillating in samephase.

f = cycles per second or hertzc= velocity of sound propagation= f\

X = wavelength

FIGURE 8.2.1 Sound wave being propagated through a compressible medium in ;tube.

Amplitude

Amplitude

One complete cycleFIGURE 8.2.2 Sine wave of the simple harmonic motion characterizing apure tone.

Direction of pistonmovement oscillatingat frequency, t

Minimumamplitude

Maximumamplitude

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discrete sound is sinusoidal and has a frequency equal to the number of times persecond that the piston moves back and forth.

8.2.2.1 Displacement Amplitude and Particle VelocitySpecifically, sound is transmitted through individual vibrating air particles. Thevibration causes the particles to move, but they do not change their average posi-tions if the transmitting medium itself is not in motion. The average maximumdistance moved by individual particles is called the "displacement amplitude," andthe speed at which they move is called the "particle velocity."

In air, the displacement amplitude may range from 4 X 10~9 in (10~7 mm) to afew millimeters per second. The smallest amplitude would be the lowest discerniblesound, and the largest amplitude would be the loudest sound the human ear canperceive as a proper sound.

8.2.2.2 FrequencyThe frequency of a sound wave is expressed in hertz (Hz). The range of humanhearing extends from 20 to 20,000 Hz, but 12,000 to 13,000 Hz is the limit formany adults (and the exposure of teenagers to noisy rock music is likely to resultin "old-age deafness" before they reach the age of 30). Figure 8.2.3 plots thethreshold of hearing for young adults with normal hearing.

8.2.2.3 WavelengthThe wavelength of sound is the distance between analogous points of two succes-sive waves. It is denoted by the Greek letter A and can be calculated from therelationship

\ = ~ (8.2.1)

where c is the speed of sound in ft/s (m/s), and / is the frequency in Hz.

Amplitude

Amplitude

FIGURE 8.2.3 Threshold of hearing for young adults with normal hearing.[C. M Harris (ed.\ Handbook of Noise Control, 2d ed., McGraw-Hill, NewYork, 1979, part 1, p. 8-4.}

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8.2.2.4 Sound LevelFor convenience in measuring sound without having to take data at a large numberof discrete frequencies, sound levels are often measured in one-third octave bandsor in full octaves, as per Table 8.2.1. This table shows, for instance, that the one-third octave band with a center frequency of 63 Hz has a range from 56 to 71 Hz,while the corresponding full octave has a range from 45 to 89 Hz.

8.2.3 THESPEEDOFSOUNDINAIR

The speed of sound in air can be calculated from the expression

= [49.03V460 + 0F in English unitsC [20.05V273 + 0C in metric units (*'2'Z)

TABLE 8.2.1 Nominal One-Third Octave Band Center Frequencies and Ranges

Third octave band Center frequency, Frequency range, Corresponding fullno. Hz Hz octave band

g l l S 3 ^gr~16 40 35-45 ^017 50 45-56 j18 -63- 56-71 L19 80 71-89 ^ *20 100 89-112 -21 125 112-141

CQ17C22 160 141-178 89~178

23 200 178-224 .24 -250- 224-282 -_25 315 282-355 "* ^26 400 355-447 .27 -500- 447-563 28 630 562-708 ^

where c is the speed of sound in ft/s (m/s). Note that for all practical purposes,the speed of sound in air is independent of pressure.

For example, if the temperature is 7O0F (210C), the speed of sound isc = 49.03 V530 = 1129 ft/s (344 m/s)

We can then use this value of c to compute the wavelength A at various frequencies/ at 7O0F (210C). At a frequency of 1000 Hz, for instance, we find from Eq. (8.2.1)that A = C / /= 1129/1000 =1.129 ft (0.344 m); likewise, at 20 Hz the wavelengthwould measure about 56.5 ft (17.2 m), and at 20,000 Hz it would measure about% in (17.2 mm). For 7O0F (210C), Table 8.2.2 gives the wavelengths of sound inair at several frequencies.

As a practical matter, because the thickness of walls and the absorptive sectionsof silencers are small in relation to the wavelengths of low frequencies, such struc-tures generally attenuate sound much better in the middle and high frequencies thanin the low ones. Larger and more complex structures are required for reducing low-frequency noise.

8.2.4 THESPEEDOFSOUNDINSOLIDS

The speed of sound in longitudinal waves in a solid bar can be shown to be

c, = J- (8.2.3)\P

where cs is the speed of sound in solids (m/s), E is the bar's modulus of elasticity(N/m2), and p is its density (kg/m3). This obviously means that sound travels fasterthrough media of high modulus of elasticity and of low density. Accordingly, be-cause rubber has a much higher elasticity and lower density than steel does (as oneexample), a rubber insert in a steel pipe will tend to slow down sound transmissionalong the pipe. Table 8.2.3 shows the speed of sound in various media.

One can speculate that since the elasticity and density in an absolute vacuumare zero, theoretically no sound waves should be able to travel through it. Anabsolute vacuum may thus be the ultimate noise barrier. However, no one is yetknown to have been able to come up with a practical earthborn design.

8.2.5 THEDECIBEL

In using the term "decibel" it is important to understand the difference betweensound power levels and sound pressure levels, since both are expressed in decibels.

TABLE 8.2.2 Wavelengths of Sound in Air at 7O0F (210C) c = 1129 ft Is)

/,Hz 63 125 250 500 1000 2000 4000 8000X, ft 17.92 9.03 4.52 2.26 1.129 0.56 0.28 0.14X, m 5.46 2.75 1.38 0.69 0.34 0.17 0.085 0.043

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TABLE 8.2.3 Speed of Sound in Various Media (Shown in Descending Order ofMagnitude)

Speed SpeedMedium ft/s m/s Medium ft/s m/s

Steel 16,500 5029 Concrete 10,600 3231Aluminum 16,000 4877 Water 4,700 1433Brick 13,700 4176 Lead 3,800 1158Wood (hard) 13,000 3962 Cork 1,200-1,700 366-518Glass 13,000 3962 Air 1,129 344Copper 12,800 3901 Rubber 130-490 40-149Brass 11,400 3475

Sound pressure levels, which can readily be measured, quantify in decibels theintensity of given sound sources. Sound pressure levels vary substantially withdistance from the source, and they also diminish as a result of intervening obstaclesand barriers, air absorption, wind, and other factors.

Sound power levels, on the other hand, are constants independent of distance. Itis very difficult to establish the sound power level of any given source because thislevel cannot be measured directly, but must be calculated by means of elaborateprocedures; thus, as a practical matter, sound power levels are converted to soundpressure levels, which form the basis of practically all noise control criteria.

(As one example, Sec. 8.2.24 illustrates how the sound power level of a fan inan HVAC system is a critical element in the silencer selection procedure to meetspecified sound pressure level criteria in an office or space.)

8.2.5.1 Sound Power LevelThe lowest sound level that people of excellent hearing can discern has an acousticpower, or sound power, of about 10~12 W. On the other hand, the loudest soundgenerally encountered is that of a jet aircraft, with a sound power of about 105 W.Thus the ratio of loudest to softest sounds generally encountered is 1017:1.

A tenfold increase is called a "bel," so the intensity of the jet aircraft's noisecan also be referred to as "17 bels." This cuts the expression of immense rangesof intensity down to manageable size. However, since the bel is still a rather largeunit, it is divided into 10 subunits called "decibels" (dB). Thus the jet noise is 170dB, and to avoid confusion with any other reference intensity, we can say that it is170 dB with reference to 10~12 W.

Sound power level Lw in decibels is therefore defined as

WLw = 10 log ^ pn dB re 10~12 W (8.2.4)

where W is the sound power in watts. The sound power level in decibels can alsobe computed from

Lw = 10 log W + 120 (8.2.5)Since 10~12 as a power ratio corresponds to -120 dB, we can see that by definitionC

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1 W is equivalent to a 120-dB power level. Table 8.2.4 shows the sound powerlevels of typical sources.

Note that certain older literature may contain sound power level data referencedto 10~13 W, an absolete standard. Where this is the case, deduct 10 dB to convertto the current standard of 10~12 W.

The question now is, How does one measure sound power Wl This is whereanother way of looking at sound power helps. As shown in Fig. 8.2.4, consider a

TABLE 8.2.4 Sound Power Level Lw of Typical Sources

Source Sound power W9 W Lw, dB re 10"12 WSaturn rocket 100,000,000 200Afterburning jet engine 100,000 170Large centrifugal fan at 500,000 ft3/min 100 140(849,500 m3/h)

Seventy-five-piece orchestra/vaneaxial 10 130fan at 100,000 ft3/min (169,900 m3/h)

Large chipping hammer 1 120Blaring radio 0.1 110Centrifugal fan at 13,000 ft3/min (22,087 0.1 110m3/h)

Automobile on highway 0.01 100Food blendersupper range 0.001 90Dishwashersupper range 0.0001 80Voiceconversational level 0.00001 70Quite-Duct silencer, self-noise at +1000 0.00000001 40

ft/min (5.1 m/s)Voicevery soft whisper 0.000000001 30Quietest audible sound for persons with 0.000000000001 Oexcellent hearing

Spherical surfaceswith radius r

Intensity Ipower perunit area

Power W

FIGURE 8.2.4 Ideal nondirectional sound source radiating Wwatts and producing a sound intensity / in watts per unit.

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simple nondirectional source located at the center of a spherical surface (or at thecenter of a number of expanding spherical surfaces). Here the total sound powerin watts is equal to sound intensity / (W/m2) times the surface area S (m2):

W = IS (8.2.6)where S for a spherical surface is 4nr2. Of course, as the sound waves move fartherfrom their source, the surrounding spherical surface will become larger, and lesspower will pass through any unit element of the surface.

If the sound source is directional, the intensity will vary over the surface andthe radiated power must be found by integration:

W = I ISds (8.2.7)JsSince sound intensity / is rather difficult to measure, we measure sound pressure

p instead. The relationship between sound pressure and intensity is

/ = W/m2 (8.2.8)pc

where p = root-mean-square (rms) sound pressure, N/m2p = density of air, kg/m3c = speed of sound in air, m/s

The form of this equation will be familiar to many since it is analogous to theformula relating to electric power, voltage, and resistance:

'-Iwhere P = power, W

E = voltage, VR = resistance, U

Sound intensity level L1 is defined as

L1 = 10 log / dB re /ref (8.2.9)'ref

where /ref is 1(T12 W/m2.

8.2.5.2 Sound Pressure LevelSince sound-measuring instruments respond to sound pressure, the word "decibel"is generally associated with sound pressure level, but it is also a unit of soundpower level. The square of sound pressure is proportional to, though not equal to,sound power.

Assuming a point source of sound radiating spherically in all directions, Eq.(8.2.6) tells us that W = IS. Accordingly, 10 log W= 10 log / + 10 log S, whereS is the surface of the radiating sphere in ft2 (m2). Equation (8.2.5), however, tellsus that 10 log W = Lw - 120. It can also be shown that 10 log / = L1 - 120. Asa result, we get

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Lw - 120 - L1 - 120 + 10 log S

or

Lw = L1 + 10 log S

Since L1f = 10 log.(///ref) and 7 = P2Ipc, and since L1 can also be expressed as 10log (p2//?ref) which is also referred to as sound pressure level Lp, then

/ \2Lp = 10 log {-} = 20 log -^- (8.2.10)VPref/ Pref

Accordingly,

Lw = Lp + 10 log S (8.2.11)

A2.0 DETERMINATIONOFSOUNDPOWER LEVELS

The concept of the imaginary radiating sphere emanating from the sound sourcewill be referred to again in Sec. 8.2.8, Propagation of Sound Outdoors. Here, onthe other hand, without considering imaginary spheres, we are concerned with mea-suring the sound power of a source that is confined within a structurally rigid space;for very large pieces of equipment and operating machinery in a plant, this approachmay be the most practical way to estimate sound power.

The best method for determining the sound power level of a source is to measureit inside a good reverberant room with a truly diffuse sound field. With the soundpower thus contained within the room, and with its intensity evenly distributedthroughout the room, often only one sound pressure level measurement has to betaken. Then the sound power level can be calculated from Lw = Lp + K, where Kis a constant dependent on the room volume, on the reverberation time at a givenfrequency or frequency band, and on the humidity.

Another method consists of containing the sound within the rigid walls of a pipeor duct equipped with an anechoic termination to minimize end reflections. Hereall the sound energy must travel through the duct, and its sound field can be mea-sured at a suitable measuring plane by averaging the sound pressure level acrossit. Equation (8.2.11) can then be used for calculating the power level of the noise-maker. Figure 8.2.5 illustrates such an arrangement for a ducted fan, which is alsothe basis of U.S. and British standards. (The 1986 U.S. standard was publishedjointly by ASHRAE, ANSI, and AMCA: ANSI/ASHRAE 68-1986 and ANSI/AMCA 330-86.) Although the anechoic duct method must overcome some practicaldifficulties, such as allowing for aerodynamically induced noise at the microphone,it clearly illustrates the relationship between measurements of sound pressure leveland sound power level.

A test code of the U.S. Air Movement and Control Association (AMCA) requiresthe use of a reverberant or semireverberant room for determination of fan soundpower levels. In such an arrangement the microphone would not be affected byaerodynamic flow. These two methods can yield comparable results, but relativefan sound power levels are likely to be most comparable if they have been deter-mined under identical conditions.

In the British Standard 848, Methods of Testing Fans, 1966, part 2, the soundpower level Lw in each frequency band would be calculated after averaging the

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Test on fan outlet with open inletFIGURE 8.2.5 Anechoic duct method for fan sound power level determination. (British Stan-dard 848, Methods of Testing Fans, 1966, Part 2; ASHRAE/AMCA, Laboratory Method of TestingIn-Duct Sound Power Measurement Procedure for Fans, 1986.)

sound pressure level Lp across the duct area according to Lw = Lp + 10 log A,where A is the cross-sectional area in ft2 (m2) at the plane of measurement. TheU.S. standard (as in Fig. 8.2.5) uses Lw = Lp + 20 log D - 1.1, where D is thediameter in ft (m) of the test duct.

8.2.7 CALCULATINGCHANGESINSOUNDPOWER AND SOUND PRESSURE LEVELS

8.2.7.1 Sound Power LevelLet Lwl be the sound power level corresponding to sound power W, and let LW2be the sound power level twice as great, or 2W. Then from Eq. (8.2.4)

WLwl = 10 log ^ref

2W Wand LW2 = 10 log = 10 log + 10 log 2 = Lwl + 3 dB

Nref "refNote: In eq. (8.2.4), Wref = 10~12 W = 1 pW (picowatt).

8.2.7.2 Sound Pressure LevelAssume L 1 to correspond to sound pressure /?, and L 2 to sound pressure 2p. Thenfrom Eq. (8.2.10),

Lpl = 20 log -^ -Pref

and

Measuringplane

Samplingtube Air flow

Fan Intermediateduct Test duct Anechoictermination

Throttlesection

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Lp2 = 20 log ^ - = 20 log -- + 20 log 2 = L771 + 6 dBPref PrefThe addition of two equal sound pressures results in an increase of 6 dB, and

the addition of two equal sound powers results in an increase of 3 dB. However,when two equal sound pressure levels are added, we are adding in effect two equalsound power levels, therefore:

Lpl + Lpl = 10 log (^) + 10 log (^YV7ICf/ V7ICf/

= 10 log I-?-} x 2VW/ \2

= lOlog (--) + 3 dBVW

Similarly, it can be said that when W identical sound sources are added,

Lp (total) = Lp(single source) + 10 log W (8.2.12)where Af is the number of sources; 10 log N is plotted as a function of N in Fig.8.2.6.

Table 8.2.5 shows how to add two unequal decibel levels, and Fig. 8.2.7 presentsTable 8.2.5 graphically. Examples:

1. Two fans produce an Lp of 95 dB each in the fourth octave band at a givenlocation. The combined Lp in that band would then be 98 dB.

2. If one of these fans is slowed down to produce an Lp of 90 dB, the combinationLp would then be 96 dB.

Number of sources, NFIGURE 8.2.6 Predicting the combined noise level ofidentical sources.C

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TABLE 8.2.5 Addition of Sound Levels

Difference between the Add to the higher level,two levels, dB dB

0 31 2.52 23 24 1.55 16 17 18 0.59 0.5

10 or more O

Add t

o hig

her le

vel

Difference between two levels

Example: 8OdB + 74 dB = 81 dBFIGURE 8.2.7 Decibel addition.

8.2.8 PROPAGATION OF SOUND OUTDOORS

Section 8.2.5.1 (and Fig. 8.2.4) introduced the concept of sound propagatingthrough a series of spheres increasing in size as the distance r from its sourceincreases. We now need to differentiate between hemispherical and spherical soundsources. If the source is considered hemispherical, the surface area S = 27rr2; ifthe source is spherical, S = 47rr2.

A fully spherical source would not be encountered frequently in a practicalsituation (examples would be an aircraft flying overhead, a rocket in flight, or noiseemanating from the top of a tall building or a vertical stack or from a bird flyingthrough air). If the source radiates hemispherically, as most sources do when closeto the ground or to other reflective surfaces, then for a uniformly directional source(such as a siren), the relationship between sound pressure and sound power wouldbe Co

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Lp = Lw - 10 lOg 27TT2

= Lw - 20 log r - 10 log 27T= [lw - 20 log r + 2.3 if r is in feet

[L w - 20 log r - 8 if r is in meters l.^.i^

For a spherical source, the relationship would be

Lp = Lw~ 10 lOg 47TT2

= Lw - 20 log r - 20 log 4w= \LW - 20 log r + 0.7 if r is in feet

[L^ 20 log r 11 if r is in meters

It will be noted that the sound pressure level for a hemispherical source is 3 dBhigher than for a spherical source because the same sound intensity is consideredto pass through an area half the size of a full sphere.

Not all sound sources radiate uniformly. If a sound source has a marked direc-tional characteristic, this characteristic has to be taken into account; it is called the"directivity index" (DI). Figure 8.2.8 illustrates how noise emanating from an open-ing, stack, or pipe will vary with the directivity angle.

Other factors affecting the radiation of sound might be barriers and the attenu-ation of sound due to atmospheric conditions (such as molecular absorption in theair, wind, and rain) and ground conditions (including grass, trees, shrubbery, snow,paving, and water). Attenuation due to such factors is generally significant in the

Directivity angle

Direc

tivity

ind

ex, dB

Octave band center frequency, Hz

The noise emanating from an opening, stack or pipe, will vary withdirectivity angle between the point of measurement and the conduitcenterline. Data shown for stack or pipe diameter of approximately10-ft (3.05-m) equivalent diameter.FIGURE 8.2.8 Directivity index from openings, stacks, or pipes.(NEMA Standards Publication SM 33, Directivity in Openings, Stacksor Pipes, 1964.)Copy

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high frequencies and over long distances and makes reliable and reputable outdoormeasurements very difficult to obtain.

For a directional noise source, we can therefore estimate sound pressure levelsby modifying Eq. (8.2.13) as follows:

= \LW- 20 log r + DI - Aa - Ab + 2.3 if r is in feet / o 9 1 c xp \LW- 20 log r 4- DI - Aa - Ab - 8 if r is in meters ^'z'1^

where DI = directivity indexAa = attenuation due to atmospheric conditionsAb = attenuation due to barriers

r = distance from source, ft (m)For instance, if we know or estimate the Lw of a fan (now provided by manymanufacturers), we can also estimate the Lp at a distance r by taking into accountdirectivity and the other factors indicated in Eq. (8.2.15).

8.2.9 THEINVERSE-SQUARELAW

From Eq. (8.2.15) we can see that if the sound pressure level of a source is mea-sured at two different distances from the source, the difference in sound pressurelevels at those locations is

Ir Y rLp2 - Lpl = 10 log M) =20 log ^ (8.2.16)

Vi/ riwhere Lpl = sound pressure level at location 1, dB

Lp2 = sound pressure level at location 2, dBT1 = distance from source to location 1, ft (m)r2 = distance from source to location 2, ft (m)

The relationship between (Lp2 - Lpl) and T2Ir1 is shown in Fig. 8.2.9.

R1 = distance from source to location 1R2 = distance from source to location 2Lpi = sound pressure level, location 1L

1 = sound pressure level, location 2FIGURE 8.2.9 Inverse-square law.Copy

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Equation (8.2.16) shows that the sound pressure level varies inversely with thesquare of the distance from the source, with Lp decreasing by 6 dB for each dou-bling of distance from the source. This relationship is known as the "inverse-squarelaw."

At locations very close to a sound source, a measurement point will be in whatis known as the "near field" or the source. In the near field, neither Eq. (8.2.15)nor Eq. (8.2.16) applies, and Lp will vary substantially with small changes in po-sition. As the distance increases, however, Lp will decrease according to the inverse-square law; Eqs. (8.2.15) and (8.2.16) will apply, and a measurement point can besaid to be in the "far field" of the source.

For all practical purposes, the inverse-square law functions only in a "free field,"which is defined as a space with no reflective boundaries or surfaces. Outdoors,such conditions are likely to exist only in an open field. In a reverberant field, suchas might exist in the courtyard of a building or in a narrow street, the sound pressurelevel may decrease by a factor of less than 6 dB for each doubling of the distance.On the other hand, in a field of freshly fallen snow the decrease may be more thanthat predicted by the inverse-square law.

8.2. W PARTIALBARRIERS

Unobstructed sound propagates directly along a straight-line path from the source.If a barrier is interposed between that source and a receiver, some of the sound willbe reflected back toward the source. These reflections can, of course, be attenuatedby placing sound-absorptive surfaces on the barrier side facing the source.

Another portion of the sound emanating from the source is transmitted throughthe barrier (Fig. 8.2.10). To meet structural and wind loading criteria, however,most barrier designs significantly inhibit noise transmission to the extent that soundreaches the receiver primarily by diffracting over and around the barrier. As shownin Fig. 8.2.11, the presence of the barrier creates a "shadow zone" in which dif-fraction attenuates the noise reaching the receiver; the extent of this attenuation isthe angle S between the straight and diffracted sound paths. Angle (and therebybarrier attenuation) increases if the receiver or source is placed closer to the barrieror (assuming that the barrier is long enough to prevent sound from diffractingaround the ends) if the barrier height is increased.

The theoretical relationship between barrier height, source and receiver position,and barrier attenuation from diffraction can be mathematically expressed as a func-tion of Fresnel number N9 as shown in Fig. 8.2.12.

Reflected

FIGURE 8.2.10 Barrier reflection, diffraction, and trans-mission.

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Shadow zoneFIGURE 8.2.11 The shadow zone behind a barrier.

Barrie

r atte

nuati

on, dB Stationarysource

Movingvehiclesource

Fresnel number = N = A>where X = wavelength of sound, ft or m

8 = A + B - d , f t o rm

Note: N is a dimensionless number and can be used in Englishor metric units on a consistent basis.

FIGURE 8.2.12 Barrier attenuation as a function of Fresnel number. (Z. Maekawa,"Shielding Highway Noise,"Noise Control Engineering, vol. 9, no. 1, July-Aug, 1977.)Cop

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8.2.11 PROPAGATION OF SOUND INDOORS

Assume that a sound source is on the floor of an enclosed space and that there areno partitions or barriers between the source and the receiver, and assume furtherthat none of the sound leaves the space and reaches the receiver by a flanking path.Under these conditions, the sound in the space will reach the receiver by two paths:a direct sound path and a reverberant sound path.

8.2.11.1 Direct Sound PathIn the far field of the source, sound from a source on or near the center of a wallor floor in a room will propagate to the receiver according to the inverse-squarelaw:

j = [LW - 20 log r + 2.3 if r is in feet .R 9 ._,L^ [L

w - 20 log r - 8 if r is in meters (*'Z'L /}

where Lpd is the sound pressure level from direct sound.

8.2.11.2 Reverberant Sound PathReverberant sound will reach the receiver after reflecting off surfaces in the space.If the sound in the space is diffuse (essentially equal at all locations), Eq. (8.2.18)applies:

T = I^w ~ 10 log A + 16.3 if A is in sabins r 9 isnpr

~ \LW - 10 log A + 6.0 if A is in metric sabins t*-2-15'

where Lpr is the sound pressure level from reverberant sound, and A is the totalabsorption. The "total absorption" of a surface is the product of the surface area Sand the absorption coefficient a of that surface:

A = Sa (8.2.19)where the units of A are sabins if S is in ft2, and metric sabins if S is in m2; a isthe sound absorption coefficient, the dimensionless ratio of sound energy absorbedby a given surface to that incident upon the surface (see Sec. 8.2.22 and Table8.2.43).

Total room absorption can be calculated as follows:

A = ^Sa = S1Ct1 + S2a2 + S3Ot3 + H- Snan (8.2.20)where A = total absorption in room, sabins (metric sabins)

S = total surface area in room, ft2 (m2)a = average room absorption coefficient

S1, S2, S3, . . . , Sn = surface area of different segments of wall, ceiling, andfloor surfaces in room

Qf1, OJ2, a3, . . . , Oin = corresponding sound absorption coefficientsReverberant sound may be reduced by adding sound-absorptive materials to

reflective room surfaces. The theoretical reduction in reverberant sound due to add-

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ing sound-absorptive treatment to the surfaces of a room containing a diffuse soundfield is equal to

^Reduction in reverberant sound =10 log (8.2.21)^i

where A1 is the total room absorption after adding sound-absorptive treatment, andA2 is the total room absorption before adding treatment. This is illustrated in Fig.8.2.13.

8.2.11.3 Effects of Direct and Reverberant SoundThe effects of direct and reverberant sound are shown in Fig. 8.2.14. Direct soundpredominates close to the source, but direct sound diminishes with distance. Thus,farther from the source, reverberant sound predominates; under ideal conditions thisoccurs when the Lp in the room levels off with increasing distance from the source.

The quantitative relationship between Lw and Lp from both direct and reflectedsound paths is shown in Fig. 8.2.15 as a function of distance from the source andtotal room absorption. Add 3 dB to Lp if the source is on the wall or floor of theroom, add 6 dB if the source is at the intersection of two walls (or a wall andceiling), and add 9 dB if the source is in a corner (Ref. 1).

Note in Fig. 8.2.15 [and also Eq. (8.2.17)] that increasing the total room ab-sorption has no effect on direct sound; accordingly, adding sound-absorptive ma-terials to room surfaces will show maximum reduction in Lp in areas where rever-berant sound predominates. Also, note that small increases in total room absorptionwill not produce significant decreases in sound pressure level; even in a locationdominated by reverberant sound, doubling the room absorption will decrease thesound pressure level by only 3 dB.

8.2.12 SOUNDTRANSMISSIONLOSS

Figure 8.2.16 shows that when a sound path is broken by a partition, part of thesound is reflected, part is absorbed, and part is transmitted through the partition.

Redu

ction

inre

verbe

rant s

ound

, dB

A2 = Total room absorption after adding soundabsorptive treatment

A2 = Total room absorption before adding treatmentFIGURE 8.2.13 Effect of increasing room absorption.

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Diffuser inceiling

Acousticalceiling

Direct sound

Reverberant sound

Near field More distant field

Noise level

Without sound absorption

With sound absorption

Distance from sound sourceFIGURE 8.2.14 Effects of direct and reverberant sound on listeners in the source's nearand far fields. Close to the source, direct sound predominates; at a distance, reverberantsound predominates.

L^ -

L p, dB

; r, m

eters

Distance from source, r, feet or meters ERSFIGURE 8.2.15 Effects of direct and reverberant sound in rooms. [C M. Harris (ed.\Handbook of Noise Control, 2d ed., McGraw-Hill, New York, 1979, part 1, p. 8-4.]

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Partition

Incidentsound

Transmittedsound

Reflectedsound

FIGURE 8.2.16 Effect of partitions on in-cident sound. (Noise Control: A Guide forWorkers and Employees, U.S. Department ofLabor, Occupational Safety and Health Ad-ministration, 1980.)

Absorbedsound

Ten times the logarithmic ratio of incident sound power to transmitted soundpower is defined as "sound transmission loss" (TL). As shown in Fig. 8.2.17,

TL = 10 log =: = 10 log - (8.2.22)Wt Twhere r = sound transmission coefficient

W1 = incident sound power, WWt = transmitted sound power, W

8.2.12.1 The Mass LawThe mass law provides a theoretical relationship between the sound transmissionloss of a single-wall (solid) partition, its weight, and the frequency of sound beingtransmitted through it. For normal incidence (NI), the relationship is

TT = I10 10S ^ + 20 log / - 33.5 if w is in lb/ft2[10 log w + 20 log / - 47.5 if w is in kg/m2 ^-^

where w is the weight (or mass density), and / is the frequency in hertz.

Incident sound powerTransmitted sound power

Reflected sound power

FIGURE 8.2.17 Incident sound power versus reflected and transmittedsound power.

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Equation (8.2.23) tells us that for each doubling of the barrier's weight, thetransmission loss increases by 6 dB. Equally, by doubling or halving the frequency,a 6-dB shift in TL occurs.

Equation (8.2.23) is commonly known as "the mass law," but more accuratelyit is an approximation. Actual data can deviate from mass law predictions by 10dB or more, and the law generally does not apply to nonhomogeneous structures.As will be shown in Sec. 8.2.21.1, for example, multilayer walls or double wallsseparated by an air space generally provide greater TL than that predicted by themass law.

8.2.12.1 The Effect of Openings on Partition TLWindows, access ports, door seals, wall-to-ceiling joints, cutouts for wiring orplumbing, and other openings can significantly diminish the TL capabilities of astructure. As an example, if a 100-ft2 (9.3-m2) partition has a TL rating of 40 dBat a given frequency, a 1 percent [or 1-ft2 (0.093-m2)] opening in that partition willreduce the overall TL to 20 dB unless noise control measures are applied to theopening. The theoretical effect of openings in partitions or complete enclosures isshown in Fig. 8.2.18.

8.2.12.3 Single-Number TL Ratings: STC RatingsFor engineering rating purposes, the TL of partitions is frequently defined in termsof a single-number decibel rating known as "sound transmission class" (STC). STCratings are determined by plotting contours of TL versus frequency in one-thirdoctave bands from 125 to 4000 Hz and comparing the results with standard contoursdefined in ASTM E413 (Fig. 8.2.19). The TL data and STC ratings of typicalstructures are listed in Table 8.2.3a. The total deficiencies must not be greater than32 dB, but any single band's deficiency cannot be greater than 8 dB.

Trans

miss

ion los

s w

ith lea

ks, dB

Potential transmission losswithout leaks, dB

FIGURE 8.2.18 Effect of openings on parti-tion TL.

% Opening

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Trans

miss

ion los

s, dB

MeasuredTL data

STC 40contour perASTM E413

1/3 octave band center frequency, HzFIGURE 8.2.19 ASTM E413 contours for sound trans-mission class (STC) and noise isolation class (NIC).(ASTM E413, Standard Classification for Determination ofSound Transmission Class, 1973.}

8.2.13 NOISE REDUCTION AND INSERTIONLOSS

As shown in Fig. 8.2.20, "noise reduction" (NR) is simply the difference in soundpressure level between any two points along the sound path from a noise source:

NR = Lpl - Lp2 (8.2.24)"Insertion loss" (IL), on the other hand is the before-versus-after difference at thesame measurement point, brought about by interposing a means of noise controlbetween the source and the receiver (Fig. 8.2.21):

IL = L770 - Lp2 (8.2.25)

Enclosingthe source

BarrierBARRIER

Enclosingthe receiver

FIGURE 8.2.20 Illustration of noise reduction:NR = Lpl - Lp2. (Lawrence G. Copley, "Control ofNoise by Partitions and Enclosures," Tutorial Papers onNoise Control for Inter-Noise, Institute of Noise ControlEngineers, 1972.)

Source Receiver

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Source SourceBefore After

FIGURE 8.2.21 Illustration of insertionloss: IL = LPQ Lp2. (Lawrence G. Copley,"Control of Noise by Partitions and Enclo-sures," Tutorial Papers on Noise Control forInter-Noise, 1972.)

Like TL, NR and IL are typically rates as a function of full octave bands orone-third octave bands. The NR ratings of several types of soundproof room arelisted in Table 8.2.42. A single-number NR rating system called "noise isolationclass" (NIC) is often used for such rooms. Similar to the STC ratings described inSec. 8.2.12.3, NIC ratings are established by plotting NR as a function of frequencyand comparing the results against standard contours defined in ASTM E413.

8.2. U THE EFFECTS OF SOUND ABSORPTIONON RECEIVING-ROOM NR CHARACTERISTICS

Figure 8.2.22 shows a receiver located within a room outside of which is a noisesource. The relationship between the NR and TL characteristics of such a roomcan be shown to be represented by

NR - TL + 10 log 0^ (8.2.26)o

where NR =Lpl - Lp2

Lpl = sound pressure level in source room, dBLp2 = sound pressure level in receiving room, dB

FIGURE 8.2.22 Noise source in outer room, and receiver in inner room.

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TL = transmission loss of receiving-room walls, dB~a.2 = average sound absorption coefficient in receiving roomA2 = total wall area in receiving room, ft2 (m2)

S = surface area separating the two rooms, ft2 (m2)If the source room is highly reverberant and if the receiving room is highly

absorptive such that ~a2 is close to unity, then NR = TL. In the event that thereceiving room is highly reflective, however, a2 will be very low; for instance, ifS = A2 and if a2 = 0.01, then NR = TL - 20 dB.

Accordingly, a highly absorptive receiving room can be seen to have a potentialof 20-dB more noise reduction than a reflective receiving room with the same TL.This effect is illustrated in Fig. 8.2.23, which shows the NR of a 6-ft 4-in by 6-ft0-in by 6-ft 6-in (1930- by 1829- by 1981-mm) room, which could be a fan plenum,tested with and without 2 in (51 mm) of sound-absorptive materials on the otherwisehighly reflective steel inside walls. The sound absorption coefficient of a 2-in (51-mm) liner is relatively low at low frequencies, so the liner has little effect on NR.At the higher frequencies, however, NR is approximately 20 dB higher with theabsorptive liner in place.

8.2.15 FANNOISE

Fans are the primary source of noise generation in HVAC systems. It is always bestto use fan Lw data provided by the fan manufacturer. However, if these data arenot available, Eq. (8.2.27) can be used to predict the estimated fan Lw (dB re1/V); see note in section 7.1 (Ref. 2):

Noise

red

uction

, dB

Octave band center frequency, Hz

1. 2-in (51 -mm) sound absorptive materialson inside wall surface

2. Reflective steel wallsFIGURE 8.2.23 Additional NR dem-onstrated by adding sound absorption tothe inside surface of a reflective receiv-ing room. (Martin Hirschorn, NoiseControl Reference Handbook, Indus-trial Acoustics Company, 1982.)C

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Lw = K w + 10 log -^ - + 20 log ^ - + C + BFI (8.2.27)Qi PIwhere Kw = specific sound power level, dB re lpw (from Table 8.2.6)

Q = flow rate, fWmin (L/s)Q1 = 1 when Q is in fWmin, 0.472 when Q is in L/sP = fan pressure head, in WG (Pa) [in WG is "inch water gauge"]

P1 = 1 when P is in inches WG, 249 when P is in PaC = correction factor for point of operation, dB

BFI = blade frequency increment to be added only to octave band contain-ing blade pass frequency

The values of Kw and BFI are shown in Table 8.2.6, and Table 8.2.7 shows theoctave band in which the BFI is likely to occur for different fan types. Values forC are given in Table 8.2.8.

Fans can generate high-intensity noise levels of a discrete tone at the blade passfrequency (BPF). The noise level's intensity will vary with the type of fan. TheBPF can be established if the rpm and number of blades of the fan are known; thefollowing equation an then be used:

BPF = -Pm x number of blades ^ ^^OU

For example, if the rpm is 1200 and the number of blades is 8, BPF =160 Hz.These discrete tones at the BPF are usually the most predominant noises ema-

nating from large fans, but such discrete frequencies may not show up in an octaveband analysis. To find them, narrower frequency ranges may have to be measured,such as one-third octave bands or even one-tenth octave bands.

BFIs vary from 2 dB for centrifugal radial-blade fans to 8 dB for centrifugalpressure-blower fans (see Table 8.2.6). BFIs and second harmonic frequencies gen-erally occur in the 63- to 500-Hz region (see Table 8.2.7).

Once a decision has been made as to the type of fan to be used, it is best toselect one that operates close to the peak of its efficiency curve. Such a fan willtypically generate the lowest noise level. The correction factor C for off-peak op-eration is shown in Table 8.2.8.EXAMPLE 8.2.1 A 35.5-in-diameter vaneaxial fan with eight blades has a 20,000-fWmin flow rate, develops a 4-in WG head at a speed of 1765 r/min, and operatesat 95 percent of peak efficiency. Determine the fan's Lw and BPF.Solution Calculate the fan's total Lw from Eq. (8.2.27):

For Kw, Table 8.2.6 gives a range of octave band center frequencies for a va-neaxial fan with a diameter (or wheel size) under 40 in.

Flow rate Q = 20,000 ft3/min, and Q1 = 1. Thus 10 log (QIQ1) = 43. Fan pressure head P = 4 in WG, and P1 = I . Thus 20 log (PfP1) = 12. Correction factor C comes from Table 8.2.8; at 95 percent peak efficiency, C =

O. From Table 8.2.6, for vaneaxial fans, BFI = 6. Furthermore, Table 8.2.7 and its

note show that this BFI occurs in the 250-Hz octave band.

These data are tabulated in Table 8.2.9, which shows the total Lw.

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TABLE 8.2.6 Specific Sound Power Levels Kw (dB re lpw) and Blade Frequency Increment (BFI) for Various Types of Fans

Octave band center frequency, HzBFI40002000100050025012563Wheel sizeFan type

3

2

8

6

5

5

1520233038

3832283437

46

2328253240

3934413743

52

2833283745

4437433944

55

2934333945

5039434146

56

3136394248

5838434347

58

3238433948

5736393941

51

3236474555

6339374140

48

Over 36 in (900 mm)Under 36 in (900 mm)AllOver 40 in (1000 mm)From 40 in (1000 mm) to

20 in (500 mm)Under 20 in (500 mm)Over 40 in (1000 mm)Under 40 in (1000 mm)Over 40 in (1000 mm)Under 40 in (1000 mm)

All

CentrifugalAirfoil, backward-curved,backward-inclined

Forward-curvedRadial blade and pressureblower

Vaneaxial

Tubeaxial

PropellerCooling towerNote: These values are the specific sound power levels radiated from either the inlet or the outlet of the fan. If

the total sound power level being radiated is desired, add 3 dB to the above values.Source: 1987 ASHRAE Handbook, Systems, ASHRAE, Atlanta, 1987, chap. 52, "Sound and Vibration Control."

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TABLE 8.2.7 Octave Band in Which Blade Frequency Increment (BFI) Is Likely to Occur

Octave band in which BFIFan type occurs*

CentrifugalAirfoil, backward-curved, backward-inclined 250 HzForward-curved 500 HzRadial blade and pressure blower 125 Hz

Vaneaxial 125 HzTubeaxial 63 HzPropeller

Cooling Tower 63 Hz*Use for estimating purposes. For speeds of 1750 r/min (29 r/s) or more, move the BFI to

the next higher octave band. Where the actual fan is known, use the manufacturer's data.Source: 1987 ASHRAE Handbook, Systems, ASHRAE, Atlanta, 1987, chap. 52,

"Sound and Vibration Control."

TABLE 8.2.8 Correction Factor C for Off-Peak Operation

Static efficiency, % of peak Correction factor, dB90 to 100 O85 to 89 375 to 84 665 to 74 955 to 64 1250 to 54 15

Source: 1984 ASHRAE Handbook, Systems, ASHRAE, Atlanta,GA, 1984, chap. 32, "Sound and Vibration Control."

To calculate the fan's BPF, use Eq. (8.2.28). Given 1765 r/min and eight blades,BPF = (1765 X 8)/60 = 235 Hz, and Table 8.2.9 shows that the nearest octaveband to 235 Hz in this example is 250 Hz.

TABLE 8.2.9 Calculation of Total Fan Lw in Example 49.1

Octave band center frequency, HzCalculation 63 125 250 500 1000 2000 4000

Specific fan Kw 37 39 43 43 43 41 2810 log fQ/ej +20 log (PTP1) 55 55 55 55 55 55 55C 0 0 0 0 0 0 0BFI 6-1-3 dB to get total Lw (see notebelow Table 49.6) 3 3 3 3 3 3 3

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8.2.76 COOLINGTOWERNOISE

In the typical mechanically induced-draft cooling tower (Fig. 8.2.24), noise is gen-erated by fan noise and water impact; at most locations of interest, however, fannoise predominates. For evaluation and control of cooling tower noise, see Refs. 3to 6. A typical cooling tower noise control installation, consisting of air-intake and-discharge silencers, is shown in Fig. 8.2.25.

Cooling tower fan noise, if not available from the manufacturer, can be estimatedfrom Eq. (8.2.27) and Tables 8.2.6 to 8.2.8. It should be noted, however, that theintake noise must propagate upstream against the air flow, make a 90 turn, divideas it disperses through the side of the tower, and pass through the louvers. Thistortuous path results in the cooling tower fan's intake noise being less than itsdischarge noise. Typical fan attenuation at the air intake can amount to as much as3, 7, 11, and 9 dB in the first four octave bands, respectively; however, in the lastfour bands water noise predominates. Clearly, wherever possible, data based onactual measurements and provided by the cooling tower manufacturer should beused.

8.2.17 DUCT SILENCERSTERMINOLOGYAND TYPES

Duct silencers reduce the air-flow noise inside air-handling systems that is causedby the following:

The fanthe air's prime mover The passage of air through straight ducts

Airflow

Motor

Sheathing

Airflow

Propeller fan

Spraynozzles

FillLouvers

Airflow

Collecting basinFIGURE 8.2.24 Mechanically induced-draft coolingtower.Copy

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IAC Quiet-DUCTdischargesilencers

IAC Quiet-DUCTintake

silencers

FIGURE 8.2.25 Silencers for cooling towers.(Application Manual for Duct Silencers, Bulletin1.0301 A, Industrial Acoustics Company, 1989.)

The impact of air flowing through duct components, such as elbows, branches,mixing boxes, rods, and orifices

We can generalize that any form of air movement will generate noise. If V isthe velocity of air flow in a straight duct, the sound power level may be a functionof V5 to V7, depending on the frequency and the duct component. This means thatthe noise generated by air flow inside a duct may increase or decrease by 15 to 21dB every time the velocity is doubled or halved.

Six principal parameters are generally used to describe the aeroacoustic char-acteristics of silencers:

1. Dynamic insertion loss (DIL): The DIL is the difference between two soundpower levels or intensity levels when measured at the same point in space beforeand after a silencer has been inserted between the measuring point and the noisesource.

2. Self-noise (SN): The SN is the sound power level in decibels generated by agiven volume of air flowing through a silencer of stated cross-sectional area.

3. Air flow: Accurate aerodynamic measurements are essential in describing anycomponent of an air-handling system. DIL and SN data are always reported asa function of silencer face air-flow velocity.

4. Static pressure drop: This is generally related to silencer face velocity and vol-umetric air-flow capacity for a given silencer face area. For energy conservationconsiderations, it can also be related to the horsepower (kilowatts) required toovercome the pressure drop.

5. Forward flow: This applies to DIL and SN ratings with the air flow moving inthe same direction as the noise propagation, such as in a fan discharge system.

6. Reverse flow: This applies to DIL and SN ratings with the air flow and noisepropagation moving in opposite directions, such as in a fan inlet system.Co

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There are many types of silencers, including the following:

Reactive Silencers. These have tuned cavities and/or membranes and are de-signed mainly to attenuate low-frequency noise in diesel, gasoline, and similarengines. Such silencers, however, are rarely used in HVAC systems.

Diffuser-Type Silencers. These are used primarily for jet engine test facilities andpneumatic cleaning nozzles in manufacturing operations. They often employ per-forated "pepper pots" that slow down the flow velocities and/or prevent the gen-eration of low-frequency noise.

Active Attenuators. Much work has been done during the last 10 years on "ac-tive" silencers. These attenuate noise by means of electronic cancellation techniquesinvolving microphones, speakers, synchronizing sensors, and microprocessors.

Such silencers are effective at low frequencies under 300 Hz but are not suitablefor broadband noise reduction without the addition of a dissipative silencer.

Moreover, this cost and maintenance requirements do not make such silencersa practical proposition. However, they might constitute an answer in unusual situ-ations where there is no room for conventional silencers and where very low fre-quency noise must be controlled.

Packless Silencers. These can be used where the acoustic infill of conventionalsilencers could become a breeding ground for disease-carrying bacteria or whereparticulate matter from fiber erosion can contaminate streams of air or gas. Thismakes packless silencers particularly suitable for microchip manufacturers, foodprocessing plants, hospitals, and pharmaceutical and other manufacturing plantsrequiring clean-room environments.

The absence of acoustic materials also reduces fire hazards where flammablematerials could saturate the infill. Other applications therefore include engine testcell, kitchen exhausts, and facilities in general where fuels, grease, acids, and sol-vents might be carried in streams of air or gas.

Packless silencers could well become more important for general use if it be-comes established that fiberglass causes lung illnesses.

Dissipative Silencers. These are widely used in HVAC duct systems. Figures8.2.26 and 8.2.27 show the general configuration of rectangular splitter silencers.The splitter, consisting of a strong, perforated-steel envelope containing sound-absorptive materials, divides the air or gas flow into smaller sound-attenuating pas-sages. Rectangular silencers are used in rectangular ducts and are sometimes set upin very large tiers, or banks, on the intakes and exhausts of fans.

Figure 8.2.28 shows a tubular, or cylindrical, silencer. At first sight it lookssimilar in cross section to the rectangular silencer, but it consists of an outer cylin-drical shell and an inner concentric bullet. Cylindrical silencers are often used incircular duct systems in conjunction with vaneaxial fans.

Dissipative silencers are available in a variety of executions, lengths, and crosssections to meet almost any noise-reduction and pressure-drop requirement of anHVAC system. The use of dissipative silencers is further discussed and illustratedin Sees. 8.2.23 to 8.2.29 in terms of applications. For discussions of the principlesof silencer performance and duct break-out noise, respectively, see Sec. 8.2.18 and8.2.20.Co

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Sound-absorptive materialPerforated splitter liner

Splitters

Air passage

(a) (b)FIGURE 8.2.26 "Round-nosed" rectangular silencer, (a) Cross section; (b) external view.(Application Manual for Duct Silencers, Bulletin 1.0301.4, Industrial Acoustics Company,1989.)

(Frequently used in Europe. Constitutes poor aerodynamicand self-noise design. See Sect. 49.23.3 and Fig. 49.59.)FIGURE 8.2.27 "Flat-nosed" rectangular silencer. (MartinHirschorn, "The Aero-Acoustic Rating of Silencers for 'For-ward' and 'Reverse' Flow of Air and Sound," Noise ControlEngineering, vol. 2, no. 1, Winter 1974.)

8.2.18 EFFECTSOFFORWARDANDREVERSEFLOW ON SILENCER SN AND DIL

The self-noise (SN) of a silencer varies by 7 to 26 dB for each doubling and halvingof flow velocity, depending on the frequency, on the silencer's configuration, andon whether the noise and air flow are traveling in the same direction (i.e., forwardor reverse flow).

As explained in Sec. 8.2.17, forward flow occurs if the air flow is traveling inthe same direction as the sound propagation, as on the supply side of an HVACsystem, and reverse flow occurs when air is traveling in a direction opposite to thedirection of sound propagation, such as in a duct's return-air system. Both areillustrated in Fig. 8.2.29.Co

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Perforated jacket linerSound-absorptive jacket

Cylindrical \center body(sound absorptive)

FIGURE 8.2.28 Cylindrical silencer, (a) Cross section; (b) external view. (Application Manual for DuctSilencers, Bulletin 1.0301.3, Industrial Acoustics Company.)

Airpassage

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Sound waves Sound waves

Forward flow noise field propagatesin the same direction as airflow. Reverse flow noise field propagatesopposite to air flow.

Note: If velocity of air through silencer is 70 ft/s (21.3 m/s), the speed of sound in the forward-flowdirection would be 11OO + 70 = 1170 ft/s (335.3 + 21.3 = 356.6 m/s). Similarly in the reverse-flow direction, the speed of sound through the silencer would be 11OO - 70 = 103OfVs (335.3- 21/3 = 314 m/s). Approximate velocity of sound at sea level = 110O ft/s (335.3 m/s).

FIGURE 8.2.29 Schematic of reverse flow versus forward flow. (Application Manual for DuctSilencers, Bulletin 1.0301.4, Industrial Acoustics Company, 1989.}

Self-n

oise

soun

d pow

er lev

els,

dB,

re:

10~12

W

Forward flow

Reverse flow

Frequency, HzFIGURE 8.2.30 Characteristic self-noise spectra for rectangular silencerswith 30 percent free area. (Af. Hirschorn, "Acoustic and AerodynamicCharacteristics of Duct Silencers for Airhandling Systems," ASHRAE PaperCH-81-6, 1981.)

Figure 8.2.30 illustrates the effects of forward and reverse flow on silencer SN.Low-frequency SN is the greatest in the forward-flow mode, while high-frequencySN is the greatest in the reverse-flow mode.

Because of the forward- and reverse-flow phenomena, silencer performance isbest rated with air flow in terms of dynamic insertion loss (DIL) determined inaccordance with ASTM E477 (Ref. 7) in a reverberant room in the reverse andforward modes. The test arrangement is shown in Fig. 8.2.31 and 8.2.32. SeeRef. 8.Co

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1. Air flow measurements station2. System fan3. System silencer4. Signal source chamber5. Upstream pressure test station6. Silencer under test7. Downstream pressure test station8. Reverberation room

FIGURE 8.2.31 Typical facility for rating duct silencers with or without air flow.(ASTM E477, Standard Method of Testing Duct Liner Materials and PrefabricatedSilencers for Acoustical and Airflow Performance, American Society for Testing andMaterials, 1973.)

Air flow

Fan Plenum Sound source

System silencer Test silencer

Reverberantreceiving room

FIGURE 8.2.32 Schematic of the facility shown in Fig. 8.2.31; forward flow illustrated.(M. Hirschorn, "Acoustic and Aerodynamic Characteristics of Duct Silencers for Airhan-dling Systems," ASHRAE Paper CH-81-6, 198L)Co

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8.2.18.1 Brief Theory of the Effects of Air-Flow Direction on SilencerPerformanceIn examining the influence of air flow on the acoustic DIL, observers have foundthat air flow affects sound transmission in three major ways: (1) convection, (2)refraction, and (3) flow modification of the acoustic impedance of the duct walls.Since the third effect is rather insignificant for silencers using absorptive materials,it will not be discussed here.

Convection. The term "convection" signifies that the speed of sound in the for-ward direction is greater than in the reverse direction. As a result, the sound waves(previously referred to as the "noise field") maintain longer contact with the ab-sorptive boundary in the silencer in the reverse direction than in the forward-flowmode. This results in higher attenuation in the reverse direction than in the forwarddirection. Quantitatively, this difference between reverse-and forward-flow attenu-ation depends on the Mach number M in the duct, which is defined as

M = - (8.2.29)c

where V is the velocity of air, and c is the velocity of sound. At sea level, c isapproximately 1100 ft/s (335.3 m/s), and V in an air-conditioning silencer mighttypically be on the order of 70-ft/s (21.3-m/s) throat velocity, or a Mach numberof about 0.064.

This dependence on the Mach number is modified by whether the air-flow pat-tern in the flow sublayer close to the boundary is streamlined or turbulent. If thepattern is streamlined, the ratio between reverse- and forward-flow attenuation canbe shown to be (1 + M)I(I - Af)1; if the pattern is turbulent, the ratio is expectedto be (1 + M2)/(l - Af2)2. If the Mach number is about 0.064 and if the turbulentsublayer is streamlined, this would correspond to a theoretical ratio between re-verse- and forward-flow attenuation of about 14 percent; however, much widerfluctuations have been measured under actual test conditions.

Where turbulent flow conditions control, the ratio between reverse- and forward-flow attenuation might then be on the order of 30 percent of more; consequently,it follows that shape and construction can have a major effect on silencer attenuationvalues and that it cannot be concluded that all silencers will necessarily behavealike. There is only one way to be sure that silencers will provide the performancespecified, and that is on the basis of actual test data.

(It is interesting to note that if the velocity of air through a duct equals Mach1, then theoretically no noise at all should be transmitted in the reverse-flow direc-tion. In fact, experimental jet engine intake silencers have been constructed on thisprinciple.)Refraction. At higher frequencies, refraction begins to be significant, and it worksin opposition to the effect of convection. That is, refraction tends to increase high-frequency attenuation in the forward-flow direction and decrease it in the reverse-flow direction. This situation is illustrated schematically in Fig. 8.2.33. As a soundwave travels in the forward-flow direction, there is a tendency for it to be refractedtoward the boundary, which leads to smaller attenuation in the reverse-flow direc-tion. This effect is significant only at higher frequencies when the wavelength issmaller than the cross-sectional dimensions of the duct.Co

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Velocityprofile

SoundAir

Sound

Under forward-flow conditions, high-frequencysound is refracted into the duct-silencer walls. Under reverse-flow conditions, sound is refracted awayfrom the walls and toward the center of the duct silencer.FIGURE 8.2.33 The refraction of sound under forward- and reverse-flow conditions. (Appli-cation Manual for Duct Silencers, Bulletin 1.0301.3, Industrial Acoustics Company.)

It will be noted from the data in Fig. 8.2.55 that in the reverse-flow mode,silencer attenuation falls off markedly from the sixth octave band upward and in-creases for the forward-flow mode (Refs. 9 to 11).

8.2.19 COMBINING ACTIVE AND DISSIPATIVESILENCERS:

Active noise control presently is not a broadly used method for achieving HVACnoise control because of relatively higher costs compared with dissipative silencers.However, for selected applications, there may be significant benefits in combiningthe active technology with the broad band performance of dissipative sound ab-sorptive silencers.

In active noise cancellation, sound is cancelled by destructive interference. Thebasis of all active attenuation systems is that the noise from a secondary source isgenerated with a mirror image wave form of the primary sound field to cancelunwanted sound downstream of the attenuator. The secondary noise source mustalso be controlled.

The secondary source must be of the same order of magnitude as the noise tobe cancelled and must also be controlled. Fig. 8.3.34 shows how noise and anti-noise sources cancel each other out.

NOISE

RESULT

ANTI-NOISE

FIGURE 8.2.34 Noise Cancellation The-ory. (Ref. 20)Copy

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A simple active noise control system, Fig. 8.2.35, utilizes two microphones; onefor input and one for error corrections, a loudspeaker and a controller. Unlikedissipative silencers, active silencers will consume small amounts of electricalpower and will require equipment maintenance from time to time. This may entailreplacement of loudspeakers, which must often operate continuously in some timesrugged environments.

Hybrid Active Dissipative Silencers: The most effective application of the activesilencer principle in HVAC Systems is a "hybrid" combination of active and dis-sipative silencers. Table 8.2.10 shows performance of one combination. Active ductsilencers for frequencies in excess of 500 Hz are generally not considered practicaldue to increasingly complex "cross modes" at the higher frequencies.

The active silencer performance shown in Table 8.2.10 provides attenuation upto 500 Hz. The acoustical characteristics of dissipative silencers for 3 m and 900mm long silencers respectively, provide additional low frequency attenuation as wellas greater amounts of mid and high frequency attenuation. Other selections of dis-sipative silencers can be combined with the active silencer where the dissipativesilencer provides a larger amount of low frequency but most of the attenuationabove 500 Hz. Depending on space considerations, these can also be designed withvery minimal pressure drop.

For most active silencer systems performance can be limited by the presence ofexcessive turbulence in the airflow detected by the microphones. Manufacturersrecommend using active silencers only where duct velocities are less than 1500fpm and where the duct configuration is conducive to smooth, evenly distributedairflow. (Ref: 1995 ASHRAE Handbook, HVAC Applications).

InputMicrophone

ErrorMicrophoneLoudspeaker

FIGURE 8.2.35 Active Duct Silencer. (Ref. 21)

TABLE 8.2.10 DIL of Dissipative Silencers and TCM (Tight Coupled Monopole) ActiveAttenuator.

63 125 205 500 IK 2K 4K 8K3 m (10 ft) 13 26 42 52 55 53 51 42900 mm (3 ft) 2 7 11 15 21 26 18 11TCM 10 11 16 13 1 O O OCo

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8.2.20 SOUNDTRANSMISSIONTHROUGHDUCT WALLSDUCT BREAK-OUT ANDBREAK-IN NOISE

The break-out phenomenon in particular illustrates the importance of reducing fannoise by means of silencers directly after the fan. Otherwise, duct runs that lackan adequate acoustic design may radiate unacceptably high noise levels into oc-cupied spaces.

Air ducts are commonly manufactured from light-gauge sheet materials, whichprovide only partial containment of the sound field within the duct. Internal noisecan be transmitted into the surrounding space (break-out), and in some cases ex-ternal noise can pass into the duct (break-in), which then becomes a path for noiseto travel into other occupied areas.

The phenomena of break-out and break-in sound transmission are illustrated inFig. 8.2.35 and 8.2.36.

The magnitude of the sound transmission loss (TL) of a duct wall differs fromthat of a plenum wall panel due to the frequency-dependent nature of the soundpropagation within the duct. If the cross-sectional dimensions of the duct are smallerthan one-half of the wavelength, only plane waves can propagate within the duct.The vibration response of the duct walls and the pattern of radiation of sound fromthe duct are governed by the directional characteristics of the internal sound field.The forced response of the duct wall is proportional to the local sound pressure,which propagates in an axial direction at a speed that is equal to or greater thanthe speed of sound.

Practical TL curves have been developed that are divided into two regions (Ref.2): one where plane-mode transmission within the duct predominates, and anotherwhere cross-modes prevail.

For break-out, the limiting frequency fl between these curves is given by

/, = ^ (8-2.30)

where a and b are the duct cross-sectional dimensions in inches; when working

Wall Wall

Sound power W.entering duct

Duct

Duct outersurfacearea = A0

Ductcross-sectional

area = A1

FIGURE 8.2.36 Break-out sound transmissionthrough duct walls. (ASHRAE Handbook 1987 Sys-tems, chap. 52, "Sound and Vibration Control,"Amer-ican Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc., Atlanta, 1987.}C

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with metric units, convert millimeters to inches before using Eq. (8.2.30) (mmdivided by 25.4).

For break-in, the lowest acoustic cross-mode frequency is used as the limitingfrequency:

/, = ^ (8.2.31)

where a is the larger duct dimension in inches; when working with metric units,convert millimeters to inches before using Eq. (8.2.31) (mm divided by 25.4).

Below the limiting frequencies, break-out TL is given by

TLout = 10 log (^^) + 17 (8.2.32)

and break-in TL is given by the larger of

TL1n = TLout - 4 - 10 log J + 20 log f (8.2.33)D Jl

or

TL i n=101og[l2Z0 + i)j (8.2.34)

where / = frequency, Hzq = mass per unit area of duct wall, lbm/ft2 (kg/m2 X 0.2048)/ = duct length, ft (m)

Above the limiting frequencies, break-out TL is given by

TLout = 201og

comprehensive listings of break-out and break-in TL are shown in Tables 8.2.11 to8.2.17.

Lagging on the outside of ductwork is often used to increase the TL values. Theincrease in performance due to the lagging will depend on the type and rigidity ofthe lagging material. A hard outer layer of sheet metal or gypsum board may notbe a very effective means of reducing low-frequency noise caused by resonanceeffects in rectangular ducts. Limp covering materials that effectively add mass tothe duct wall may improve the TL values by reducing wall response without addingstiffness. In critical situations, it may be necessary to apply panels with air spacesto the duct surfaces for maximum noise reduction.

Undoubtedly, more correlation between field and empirical data is required onbreak-out and break-in noise. Some acoustic consultants and engineers consider

TABLE 8.2.11 Examples of Duct Break-out and Break-in TL versus Frequency

Octave band center frequency, HzDuct and TL type 63 125 250 500 1000 2000 4000 8000

Rectangular,* TLout 19 22 25 28 31 37 43 45Rectangular,* TL1n 14 14 22 25 28 34 40 42Circular,t TL01n 45 50 26 26 25 22 36 43

'Duct size: 44 by 12 in (1118 by 305 mm), 22 ga [0.034 in (0.85 mm)].tDuct size: 26-in (660-mm) diameter, 24 ga [0.028 in (0.7 mm)].

TABLE 8.2.12 TLout versus Frequency for Various Rectangular Ducts

Duct size*Gauge Octave band center frequency, Hz

in (mm) in (mm) 63 125 250 500 1000 2000 4000 800024 ga

12 x 12 (300 x 300) 0.028 (0.7) 21 24 27 30 33 36 41 4524 ga

12 x 24 (300 x 600) 0.028 (0.7) 19 22 25 28 31 35 41 4522 ga

12 x 48 (300 x 1200) 0.034 (0.85) 19 22 25 28 31 37 43 4522 ga

24 x 24 (600 x 600) 0.034 (0.85) 20 23 26 29 32 37 43 4520 ga

24 x 48 (600 x 1200) 0.04 (1.0) 20 23 26 29 31 39 45 4518 ga

48x48(1200x1200) 0.052(1.3) 21 24 27 30 35 41 45 4518 ga

48 x 96 (1200 x 2400) 0.052 (1.3) 19 22 25 29 35 41 45 45*Ali duct lengths are 20 ft (6 m).Cop

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TABLE 8.2.13 TLin versus Frequency for Various Rectangular Ducts

Duct size*Gauge Octave band center frequency, Hz

in (mm) in (mm) 63 125 250 500 1000 2000 4000 800024 ga

12 x 12 (300 x 300) 0.028 (0.7) 16 16 16 25 30 33 38 4224 ga

12 x 24 (300 x 600) 0.028 (0.7) 15 15 17 25 28 32 38 4222 ga

12 x 48 (300 x 1200) 0.034 (0.85) 14 14 22 25 28 34 40 4222 ga

24 x 24 (600 x 600) 0.034 (0.85) 13 13 21 26 29 34 40 4220 ga

24 x 48 (600 x 1200) 0.04 (1.0) 12 15 23 26 28 36 42 4218 ga

48 x 48 (1200 x 1200) 0.052 (1.3) 10 19 24 27 32 38 42 4218 ga

48 x 96 (1200 x 2400) 0.052 (1.3) 11 19 22 26 32 38 42 42*A11 duct lengths are 20 ft (6 m).

that the TL data presented here (from Ref. 2) may be overstated when translatedto field installations; for instance, the break-out noise sound levels are likely to behigher than would be arrived at by using the TL figures in Tables 8.2.10 to 8.2.16.However, in the meantime, the above procedures (including Tables 8.2.10 to 8.2.16)can be used, bearing in mind that the introduction of safety factors might be inorder.

8.2.27 NOISECRITERIA

Noise is unwanted or objectionable sound, and numerous standards define its limitsfor specific types of noisemakers, specify how the sound is to be measured, and incertain instances specify when. These standards are published by local and nationalgovernment agencies, national and international standards organizations, the mili-tary, professional societies, and others. A few of these standards (or criteria) aregiven below.

8.2.21.1 dBA CriteriaOne way of rating sounds is by means of the A scale, a sound-level meter weighingnetwork that approximates the response of the human ear to sound. Both the humanear and the A-weighing network are more sensitive to high-frequency than low-frequency sound. In decibels, A-scale levels are expressed as dBA. Typical noisesource dBA levels are shown in Table 8.2.18.

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Front MatterTable of ContentsPart A. System ConsiderationsPart B. Systems and ComponentsPart C. General Considerations8.1 Automatic Temperature, Pressure, Flow Control Systems8.1.1 Control Basics8.1.1.1 Control Systems8.1.1.2 Modes of Feedback Control8.1.1.3 Flow-Control Characteristics

8.1.2 Control Equipment Types8.1.2.1 Sensors8.1.2.2 Controllers8.1.2.3 Final-Control Elements8.1.2.4 Auxiliary Equipment8.1.2.5 Pneumatic, Electric, Electronic Comparisons

8.1.3 Control Applications8.1.3.1 Boiler Control8.1.3.2 Control of Excess Air8.1.3.3 HVAC Fan Systems8.1.3.4 Refrigeration Control8.1.3.5 Central Heating and Cooling Plants8.1.3.6 Water-Distribution Control

8.1.4 Building Management Systems8.1.4.1 Building Management System Types8.1.4.2 Management System Applications

8.1.5 Selection8.1.6 Total Building Function8.1.6.1 Type of Building and System Zoning8.1.6.2 Types of Occupancy and Use8.1.6.3 Accuracy Requirements8.1.6.4 Economic Justification

8.2 Noise Control8.2.1 Introduction8.2.2 The Nature of Sound8.2.2.1 Displacement Amplitude and Particle Velocity8.2.2.2 Frequency8.2.2.3 Wavelength8.2.2.4 Sound Level

8.2.3 The Speed of Sound in Air8.2.4 The Speed of Sound in Solids8.2.5 The Decibel8.2.5.1 Sound Power Level8.2.5.2 Sound Pressure Level

8.2.6 Determination of Sound Power Levels8.2.7 Calculating Changes in Sound Power and Sound Pressure Levels8.2.7.1 Sound Power Level8.2.7.2 Sound Pressure Level

8.2.8 Propagation of Sound Outdoors8.2.9 The Inverse-Square Law8.2.10 Partial Barriers8.2.11 Propagation of Sound Indoors8.2.11.1 Direct Sound Path8.2.11.2 Reverberant Sound Path8.2.11.3 Effects of Direct and Reverberant Sound

8.2.12 Sound Transmission Loss8.2.12.1 The Mass Law8.2.12.1 The Effect of Openings on Partition TL8.2.12.3 Single-Number TL Ratings: STC Ratings

8.2.13 Noise Reduction and Insertion Loss8.2.14 The Effects of Sound Absorption on Receiving-Room NR Characteristics8.2.15 Fan Noise8.2.76 Cooling Tower Noise8.2.17 Duct Silencers-Terminology and Types8.2.18 Effects of Forward and Reverse Flow on Silencer SN and DIL8.2.18.1 Brief Theory of the Effects of Air-Flow Direction on Silencer Performance

8.2.19 Combining Active and Dissipative Silencers8.2.20 Sound Transmission Through Duct Walls-Duct Break-out and Break-in Noise8.2.21 Noise Criteria8.2.21.1 dBA Criteria8.2.21.2 Community and Workplace Noise Regulations8.2.21.3 Noise Criteria (NC) Curves8.2.21.4 Speech Interference Levels8.2.21.5 Ambient Noise Levels as Criteria

8.2.22 Enclosure and Noise Partition Design Considerations8.2.22.1 Actual Versus Predicted Sound Transmission Losses 8.2.598.2.22.2 Joints8.2.22.3 Windows and Seals8.2.22.4 Doors and Seals8.2.22.5 Transmission Loss of Composite Structures8.2.22.6 Flanking Paths8.2.22.7 Room Performance

8.2.23 Sound Absorption in Rooms8.2.24 Silencer Application8.2.24.1 Specific Effects of Flow Velocity on Silencer Attenuation8.2.24.2 Interaction of DIL with Self-Noise8.2.24.3 Pressure Drop8.2.24.4 Energy Consumption8.2.24.5 Effects of Silencer Length and Cross Section8.2.24.6 Impact on Silencer p of Proximity to Other Elements in an HVAC Duct System8.2.24.7 Duct Rumble and Silencer Location8.2.24.8 Effect of Silencer Location on Residual Noise Levels

8.2.25 Systemic Noise Analysis Procedure for Ducted Systems8.2.25.1 Procedure8.2.25.2 Silencer Selection8.2.25.3 Calculating the Attenuation Effects of Lined Ducts

8.2.26 Acoustic Louvers8.2.27 HVAC Silencing Applications8.2.28 Self-Noise of Room Terminal Units8.2.29 The Use of Individual Air-Handling Units in High-Rise Buildings8.2.30 Built-Up Acoustic Plenums8.2.31 Fiberglass and Noise Control-Is It Safe?8.2.32 References

8.3 Vibration Control8.3.1 Introduction8.3.2 Theory8.3.3 Application8.3.3.1 Basic Considerations8.3.3.2 Isolation Materials

8.3.4 Selection8.3.5 Seismic Protection of Resiliently Mounted Equipment8.3.5.1 Theory8.3.5.2 Seismic Specifications

8.3.6 Acoustical Isolation by Means of Vibration-Isolated Floating Floors8.3.6.1 Theory and Methods8.3.6.2 Specification

8.4 Energy Conservation Practice8.4.1. Introduction8.4.2 General8.4.3 Design Parameters8.4.3.1 Energy Audit8.4.3.2 Design8.4.3.3 Types of Systems8.4.3.4 Chillers8.4.3.5 Boilers8.4.3.6 Waste Heat and Heat Recovery8.4.3.7 Automatic Temperature Controls (See Also Chapter 8.1)

8.4.4 Life-Cycle Costing8.4.4.1 General8.4.4.2 Discounting, Taxes, and Inflation8.4.4.3 Related Methods of Evaluation

8.4.5 Energy Management Systems8.4.5.1 Components8.4.5.2 Software Programs8.4.5.3 Functions8.4.5.4 Optional Security and Fire Alarm System8.4.5.5 Selecting an EMS

8.4.6 References

8.5 Water Conditioning8.5.1 Introduction8.5.2 Why Water Treatment?8.5.2.1 Cost of Corrosion8.5.2.2 Cost of Scale and Deposits

8.5.3 Water Chemistry8.5.3.1 Hydrologic Cycle8.5.3.2 Water Impurities8.5.3.3 Dissolved Gases8.5.3.4 Dissolved Minerals

8.5.4 Corrosion8.5.4.1 General Corrosion8.5.4.2 Oxygen Pitting8.5.4.3 Galvanic Corrosion8.5.4.4 Concentration Cell Corrosion8.5.4.5 Stress Corrosion8.5.4.6 Erosion-Corrosion8.5.4.7 Condensate Grooving8.5.4.8 Microbiologically Influenced Corrosion (MIC)

8.5.5 Scale and Sludge Deposits8.5.5.1 Mineral Scale and Pipe Scale8.5.5.2 Langelier Index8.5.5.3 Ryznar Index8.5.5.4 Boiler Scale8.5.5.5 Condensate Scale

8.5.6 Foulants8.5.6.1 Mud, Dirt, and Clay8.5.6.2 Black Mud and Mill Scale8.5.6.3 Boiler Foulants8.5.6.4 Construction Debris8.5.6.5 Organic Growths8.5.6.6 Algae8.5.6.7 Fungi8.5.6.8 Bacteria

8.5.7 Pretreatment Equipment8.5.7.1 Water Softeners8.5.7.2 Dealkalizer8.5.7.3 Deaerators8.5.7.4 Abrasive Separators8.5.7.5 Strainers and Filters8.5.7.6 Free Cooling8.5.7.7 Gadgets

8.5.8 Treatment of Systems8.5.8.1 General8.5.8.2 Boiler Water Systems8.5.8.3 Treatment for Open Recirculating Water Systems8.5.8.4 Treatment of Closed Recirculating Water Systems

8.5.9 References8.5.10 Bibliography

Appendices

Index

48435_08d

Documents

nature of noise

introductionis noise

noise control problem

high noise levels

noise reduction of barriers

noise problems inherent

theory of sound

pressure waves