Valves: Noise Calculation, Prediction, and Reduction

1213

6.14

V

alv

es:

Noise Calculation,

Prediction,

and Reduction

H. D

. BAUMANN

(1970)

J. B. ARANT

(1985)

B

. G. LIPTÁK

(1995)

F. M. CAIN

(2005)

V

alve Noise Types:

Mechanical vibration (usually below 100 dBA); hydrodynamic caused by liquid turbulence, cavitation, or flashing (usually below 110 dBA); aerodynamic (can reach 150 dBA)

Sizes:

1 to 24 in. (25 to 600 mm) in standard bodies; sizes above 24 in. in special castings or weldment fabrications

Design Pr

essure:

Up to ANSI Class 2500 (PN 420) standard; above Class 2500 in special designs

Materials of Construction:

Any machinable wrought or cast metal for body and trim approved for use in valves or pressure vessels

Special F

eatures:

Balanced plugs, special seal designs, hard facings, piloted inner valve, character-ized flow, dual (high/low) operating conditions, multistage trims

Cost:

Highly variable depending upon type of design, size, metallurgy, special features. Range may be from 2 to 10 times equivalent standard valve

P

artial List of Low-Noise

ABB Control Valves (www.abb.com/controlvalves)

V

alve and Diffuser/Silencer

Control Components Inc. (www.ccivalve.com)

Supplier

s:

Dresser Flow Solutions (www.masoneilan.com)Emerson Process Management (www.emersonprocess.com/home/products)Flo-Dyne Limited (www.flo-dyne.net)Flowserve Corporation (www.flowserve.com/valves)GE-Nuovo Pignone (www.gepower.com/prod_serv/products/valves/en/std_control.htm)Industrial & Marine Silencers Ltd. (www.silencers.co.uk)Koso Hammel Dahl (www.rexa.com/hammel_dahl/hammeldahl_index.htm)McGuffy Systems, Inc. (www.mcguffy.com/products/manufacture/silencers/

silenc.html)Metso Automation (www.neles.com)Samson AG (www.samson.de)SPX Valves and Controls (www.dezurik.com)Tyco Flow Control (www.tycovalves.com)Weir Valves & Controls (www.weirvalve. com)Welland & Tuxhorn (www.welland-tuxhorn.de)

INTRODUCTION

This section be

gins with an overview of general noise prin-ciples, followed by a description of the types of noise pro-duced by fluid flow through control valves. The discussion of control valve noise mitigation includes both the treatment of the noise source (modifying the valve) and the treatment of the noise path (providing downstream insulation or silencers). Other options include protection of the receiver (by personal protective equipment such as earplugs or earmuffs) or the

removal of the receiver (by placing a barrier or distance between the noise source and personnel). The section ends with a discussion about recent improvements in predicting and calculating probable noise levels.

Because most valve noise calculation standards avoid excessive detail, only the SI system of units will be used in this section. Users of U.S. Customary units should refer to Appendix A.1 and A.2 for the proper conversion factors, including gravitational units conversions (i.e.,

g

c

) when

necessary.

© 2006 by Béla Lipták

http://www.abb.com

http://www.ccivalve.com

http://www.masoneilan.com

http://www.emersonprocess.com

http://www.flo-dyne.net

http://www.flowserve.com

http://www.gepower.com

http://www.silencers.co.uk

http://www.rexa.com

http://www.mcguffy.com

http://www.mcguffy.com

http://www.metsoautomation.com

http://www.samson.de

http://www.dezurik.com

http://www.tycovalves.com

http://www.weirvalve.com

http://www.welland-tuxhorn.de

1214

Control Valve Selection and Sizing

SOUND

AND NOISE

A weed has been defi

ned as an unwanted plant or flower. As an environmental analogy, noise may be considered as an unpleasant or unwanted sound. Sound, in the context of this discussion, is defined as pressure fluctuations generated in the air or other medium, which are capable of stimulating the physiological hear-ing response of the human ear and brain. For ease of understand-ing, we will hereafter refer to

sound

and

noise

as equivalent terms. Most common sounds are a complex mixture of many

frequencies at varying magnitudes. Pure tones have discrete frequencies. It is customary to model sound as pressure waves with sinusoidal characteristics such as frequency (

f

), magni-tude (

p

), wavelength (

λ

), and speed (

c

). Of course, sound waves possess other more complex characteristics that are beyond the scope of this topic.

Frequency is expressed in cycles per second (cps) or Hertz (Hz), where cps and Hz are equivalent units. The magnitude of sound pressure is measured in units of pressure (Pascal in the SI system). The range of sound pressures that humans can discern from the threshold of hearing to the threshold of pain spans over 12 orders of magnitude! Therefore, it is more con-venient to use a logarithmic comparison of an actual sound pressure to a standard pressure reference at the threshold of hearing and to define this comparison as a

sound pressure level, L

p

, expressed by Equation 6.14(1) in decibels (dB).

6.14(1)

where

p

is the actual sound pressure, and the reference sound pressure,

p

o

, is defi

ned as 2

×

10

−

5

P

ascal (2

×

10

−

4

microbar or

29

×

10

−

8

psi.). Because the decibel is a log

arithmic function, for every 10 dB increase, there is a tenfold increase in sound intensity. Thus, a 100 dB sound is 10 times as intense as 90 dB and 100 times as intense as 80 dB. However, the human ear perceives each 10 dB increase as an approximate doubling of loudness.

The sound pressure fluctuations must be generated by some energy source that transfers power into the air or other wave-conducting medium. (Sound waves cannot travel in a vacuum.) The total acoustic power created by the noise source is defined as sound power,

W

a

, usually e

xpressed in watts (W). The calculation of sound power will be used in this sec-

tion to predict sound pressure levels in valve applications. It is worth remembering that, while sound is produced by a power source, it is sound pressure that the ear perceives. Sound power can also be presented as a sound power level,

L

w

, in decibels by log

arithmic comparison with the standard reference level,

W

o

, of 10

–12

W.

6.14(2)

W

avelength,

λ

, is the distance required for one complete pressure cycle.

Speed of Sound

The speed of sound,

c

, in any medium is a function of its mass density and elastic properties.

For a solid:

6.14(3)

where

E

is the elastic modulus and

ρ

is the mass density. For carbon steel pipe at 100

°

C,

E

=

198 GPa,

ρ

=

7.86 g/cm

3

, and

c

=

5020 m/s. For CrMo steel alloy pipe at 100

°

C,

E

=

207 GPa,

ρ

=

7.84 g/cm3, and

c

=

5140 m/s. Austenitic stainless steel pipe (UNS S30400) at 100

°

C with

E

=

190 GPa and

ρ

=

8.03 g/cm

3

has

c

=

4860 m/s. So, for purposes of estimating the speed of sound in steel pipe, using 5000 m/s usually produces satisfactory results.

1

F

or a liquid:

6.14(4)

where

E

s

is the isentropic b

ulk modulus. It can be shown that at 20

°

C, speed of sound in fresh water is 1481 m/s, in sea-water 1521 m/s, and in machine oil (sp. gr.

=

0.90) 1297 m/s.For a gas or vapor:

6.14(5)

where

γ

is the ratio of specific heats,

R

is the universal gas constant (8

×

314 J/kmol

×

K),

T

is absolute temperature (Kelvin), and

M

is molecular mass of the fluid. Using this relationship, we find that the speed of sound in air at 0

°

C (273

°

K) is 331 m/s.Wavelength, frequency, and speed of sound are related

as shown in Equation 6.14(6).

6.14(6)

THE HUMAN EAR

The human ear is an intricate acoustic instrument that is

described here in only general terms. The anatomy of the ear is divided into three major regions, each with unique functions: the outer ear, middle ear, and inner ear. The outer ear consists of the pinna, ear canal, and outer layer of the eardrum. It channels sound waves to the eardrum, where sound pressure waves are converted into mechanical energy by vibrating the eardrum.

Lpp

ppp

o o

=

=

10 2010

2

10log log

LW

Wwa

o

=

10 10log

1

The most recent v

alve noise calculation standard and the field of acoustics in general use SI units. Users of U.S. Customary Units are cautioned to use the proper gravitational units conversions (i.e.,

g

c

) when necessary

. To avoid excessive detail, only the SI system of units will be used in this section.

c E= /ρ

c Es= /ρ

cp RT

M= =γ

ργ

c f= λ


6.14

V

alves: Noise Calculation, Pr

ediction, and Reduction

1215

The middle ear is an air

-filled cavity containing the oss-icles (bones) that connect to the oval window to the inner ear. The middle ear cavity is also connected to the Eustachian tube, which equalizes static pressure across the eardrum. The middle ear mechanism acts as an impedance-matching trans-former. It is matching the impedance of the air in the ear canal to the impedance of the liquid of the inner ear.

The inner ear vestibule leads to the semicircular canals (providing sense of balance) and the “snail-shaped” cochlea, where the final energy transformation occurs. In the cochlea, mechanical energy is conducted through a traveling wave pattern on the basilar membrane, causing a shearing of the cilia of the outer and inner hair cells of the Organ of Corti. The design and stiffness gradient of the basilar membrane allow more efficient response to higher frequencies at the basal end, and progressively lower frequencies are detected along the membrane toward its apex.

The Organ of Corti is the sense organ that changes vibra-tion energy into neural energy. This conversion takes place as shearing stress on hair cells induces a depolarization that generates neural impulses. The neural impulses are con-ducted by the auditory nerve to the brain, where they are processed and interpreted as sound. Damage to the hair cells connecting the basilar membrane and Organ of Corti usually produces permanent loss of hearing. Damage or deterioration can occur from sudden, loud noise (explosions), excessive exposure to moderately loud noise (industrial environments, loud music), physical injury (head trauma), advancing age, infections, or disease.

Loudness P

er

ception

A health

y, young adult human is able to perceive sound over a wide range of frequencies from approximately 20 to 18,000 Hz, which is generally accepted as the audible range. The human ear, however, does not give equal weight (loudness perception) to the same sound pressure level across the frequency spectrum.

Studies of apparent loudness by many human subjects over the frequency spectrum when compared to a pure tone of 1000 Hz frequency has resulted in mapping the ear response. The loudness level in phons represents the pressure level in dB of a 1 kHz tone that a typical hearer feels is as loud as the sound in question. Figure 6.14a shows the loud-ness level map as function of frequency.

We can see from Figure 6.14a that a sound at 1000 Hz and 50 dB sounds equally loud as 67 dB at 100 Hz or 62 dB at 10 kHz. The resulting correction numbers, which are approximating the response of the human ear, are called “A” weighting. The corresponding decibel level is indicated as dBA, as shown in Figure 6.14b.

There are other weighting schemes for various purposes, but A weighting is used in governmental regulations on noise pollution. Hence, for the discussion of valve noise levels and environmental noise reduction, we will use the dBA scale. Noise levels of some common environmental sounds are compared in Table 6.14c.

Limiting

V

alv

e Noise

There are se

veral important reasons to limit the noise levels emitted by valves and piping. One of them is to prevent the harmful effects of environmental noise pollution, which includes hearing loss in people. As was noted earlier, we can tolerate much louder sounds at low and at very high frequen-cies than we can in the middle of the spectrum. This is represented in the A-weighting curve of Figure 6.14b.

Note that in the 500–7000 Hz range, the human ear is most responsive, and this is the area where high noise level exposure can do the most damage. For this reason, the U.S. government enacted the Occupational Safety and Health Act of 1970 (amended in 1998), establishing the Occupational Safety and Health Administration (OSHA). OSHA regulations limit a

FIG. 6.14a

Appar

ent loudness contours for human hearing.

FIG. 6.14b

The

A-weighting filter characteristic approximates the human ear’s response to different sound frequencies.

120110100

908070605040302010

0

120

100

80

60

40

20

0

20 100 500 1000 5000 10,000

Soundfrequency

(CPS)

Sound level(decibels)

Feeling LoudnessLevels-phons

5

0

−5

−10

−15

−20200 500 1,000 2,000 5,000 10,000 20,000

Ampl

ifica

tion

dB

Frequency Hz

Internationally standardizedA-weighting filter characteristic


1216

Control Valve Selection and Sizing

weighted 90 dBA maximum level exposure to 8 hours per day. Table 6.14d below shows general exposure time limits estab-lished by OSHA.

Figure 6.14e shows a typical frequency octave band noise level contour that will meet this limit. Note that if the predom-inant noise frequency exposure is in the critical middle fre-quency range of 1000–5000 Hz, the allowable weighted noise level over 8 hours would be considerably less than 90 dBA.

VALVE NOISE

While there are many noise sources in industrial and process plants, some of the main contributors can be control valves operating under conditions of high pressure drop. These are one of the few and sometimes the only sources of over 100 dBA sound levels found in process plants. To gain some per-spective of how loud 100 dBA actually is, refer to Table 6.14c for a comparison of common environmental sounds.

However, even if people are removed from areas with high noise levels, other hazards are still created by excessive

noise. High intensity noise can produce vibrations in struc-tures, which become magnified when the natural (resonant) frequencies within the structure are close to the dominant frequency of the noise.

Even without resonance, studies by Fagerlund have shown that the sound power that is produced downstream of valves can cause fatigue failures in valves and piping systems. This can occur when noise levels outside the pipe (1 m downstream of a valve and 1 m away from the pipe wall) exceed 110–115 dBA depending on pipe size.

Table 6.14f provides a summary of the likely causes of noise in valves and of their frequencies.

The five major sources of noise generated by control valves are as follows:

• Mechanical vibration• Control element instability• Resonant vibration• Hydrodynamic noise• Aerodynamic noise

Mechanical Vibration Mechanical vibration of valve inter-nal parts is caused by unsteady flow and turbulence within the valve. It is usually unpredictable and is really a design

TABLE 6.14c Approximate Sound Pressures Levels of Typical Sounds

Source of Sound Lp (dBA)

Near jet engine; artillery fire 140

50 hp victory siren at 30 m; threshold of pain 130

Rock-and-roll band; threshold of feeling 120

Jet flying overhead at 300 m 110

Air chisel; high pressure gas leak 100

Motorcycle at 15 m; subway train at 6 m; symphony 90

Inside sports car (100 km/h) 80

Loud conversation; noisy business office 70

Normal conversation; light traffic at 30 m 60

Private business office; normal conversation 50

Quiet conversation 40

Quiet home at night; still forest; soft whisper 30

Empty theater; rustling leaves 20

Inside a soundproof room; quiet breathing 10

TABLE 6.14dOSHA Exposure Time Limits for Various Noise Levels

Hours per Day dBA

8 90

4 95

2 100

1 1051/2 110

1/4 115 (max.)

Fig. 6.14e Frequency octave band noise level contours, which will result in the weighted average exposure to meet OSHA limits.

Oshalimits

AllowableB HR level

140

130

120

110

100

90

80

70

60

50

40

30

20

10

0Over

all65 125 250 500 1000 2000 4000 8000

Octave band center frequency

Soun

d pr

essu

re le

vel d

ecib

els -

RE 0.

0002

micr

obar

Maximummomentary level

Pipe noise Valve noise


6.14 Valves: Noise Calculation, Prediction, and Reduction 1217

problem for the manufacturer. Noise levels are typically low, usually well under 90 dBA, and in the 50 and 1500 Hz frequency range.

The problem is often not the noise, but the progressively worsening vibration as guides and parts wear. The solution can be in improving the valve design by adding heavy-duty stems and guides. Improvements in design may also include small changes in the flow path geometry of the trim, which can also eliminate some vibration problems.

Control Element Instability

Control element instability is usually due to mass flow turbu-lence impingement on the valve plug. The relationship between velocity and static pressure forces acting across the plug or disc face and the actuator force balance varies over time. With-out sufficient stiffness in the actuator, valve, and mechanical connections, fluid buffeting forces may produce vertical stem oscillations in linear valves and torsional shaft oscillations in rotary valves, resulting in low-level rattle noise usually under 100 Hz. This instability is detrimental to control.

Correction requires changing the damping characteristics of the valve and actuator combination. This is done by pro-viding a stiffer valve actuator or eliminating mechanical backlash. If the actuator is a spring-and-diaphragm type, then one can increase the nominal spring rate from 20–100 kPa (3–15 PSIG) to 40–200 kPa (6–30 PSIG).

For single-acting piston actuators, the cushion air loading can be increased. If these changes do not solve the problem, then either actuator can be replaced with double-acting air piston actuators, which are generally stiffer and allow use of higher air pressures. In extreme cases, a hydraulic snubber, an all-hydraulic actuator, or electromechanical actuator may be required.

Resonant Vibration

Resonant noise is characterized by a discrete tone and pos-sibly a few harmonic multiples. Resonance can involve merely an acoustic interaction within the valve and piping geometry, with certain frequencies of the flow turbulence. Resonant frequencies from 200 Hz to 10 kHz can be excited acoustically by the flow in the same way that tones are produced in musical instruments.

Localized metal fretting or wear is likely on internal valve parts. In some cases, turbulence and acoustic resonance excites mechanical or structural natural frequencies, producing severe vibration capable of causing damage to piping, equipment, and supporting structures. Resonant noise levels exceed calculated predictions based on current prediction standards and may be in the 90–125 dBA range.

Work by Glenn refers to this discrete resonance as screech, and his research shows that screech is possible at pressure drops lower than required to produce sonic flow in gases or cavitation in liquids. It is possible for conditions to exist that produce screech-type resonance in virtually any valve type and brand. Glenn identifies several possible causes of screech in valves:

Higher than expected velocities in the valve, due to uncer-tainties in the pressure recovery characteristics as a function of valve opening. (Refer to the discussion of the pressure recovery factor FL in Section 6.15.)

Excitation of harmonic pipe modes of vibrationFlow instabilities, due to

• Vortex shedding• Tollmien-Schlichting waves in the laminar-to-

turbulent flow transition• Bi-stable flow separation• Unstable shock waves• Unstable vapor-liquid interface

Two approaches to solving these problems include 1) modifying the design of the flow path to change the charac-teristics of the turbulence, or 2) changing the stiffness and resonant frequencies of the valve trim.

Valve trims can often be modified by a change in stem diameter, change in the plug mass, or method of guiding. Flow paths can be modified sometimes by reversal of flow direction through the valve or by minor design changes to seats, plugs, or cages. These changes shift the natural fre-quency of the plug and stem out of the excitation range of the flow turbulence, or vice versa.

Valve manufacturers and users should collaborate to implement effective solutions in valves and piping when these problems arise. For example, one investigation identified the source of serious screech noise as a gap between valve and pipe flanges created by an oversized inside diameter of the gasket. Filling the gap with a properly sized gasket eliminated the problem.

TABLE 6.14f Sound Frequencies and Sources in Valves

Frequency(Hz)

Octave BandNumber

Sound Description

Typical Noise Source in Valves

20–75 1 Rumble Vertical plug oscillation

75–150 2 Cavitation*

150–300 3 Rattle

300–600 4 Howl Horizontal plug vibration

600–1200 5

1200–2400 6 Hiss Flowing gas

2400–4800 7 Whistle

4800–7000 8 Squeal Natural frequencyvibration

20,000 and up Ultrasonic

* Cavitation frequencies vary widely from about 100 Hz to 15 kHz depending on valve and trim design.


1218 Control Valve Selection and Sizing

Hydrodynamic Noise

Hydrodynamic noise is generally less troublesome and less severe than aerodynamic noise and usually only becomes excessive when accompanied by cavitation or flashing (dis-cussed in Sections 6.1 and 6.15). Severe cavitation can pro-duce noise in the range of 90–100 dBA or higher.

Problems with cavitation or flashing are usually avoided by use of a suitable trim or valve type with low-pressure recovery characteristics (high FL). Noise caused by cavitation is the result of imploding vapor bubbles in the liquid stream. This noise can vary from a low-frequency rumble or rattling to a high-frequency squeal. This latter condition is due to acoustic or pipe resonance with cavitating fluid.

In most cases, the problem is not so much with noise as it is the destruction of the valve trim and piping from erosion and pitting by the imploding vapor bubbles. Reducing or elim-inating the cavitation and its damage also eliminates the noise. Single-stage multiorifice valves (see Figure 6.1z) and multi-stage valves (see Figures 6.1y and 6.1aa) are typical solutions to cavitation erosion and noise.

The sizing of any liquid service control valve should include an evaluation of the cavitation potential, with emphasis on eliminating or mitigating the cavitation. Section 6.15 out-lines methods for predicting the onset of cavitation. The stan-dards VDMA 24422 and IEC 60534-8-4 include methods for calculating hydrodynamic noise, but these methods have been shown by Kiesbauer and Baumann to predict lower than actual noise in many cases. At the time of this writing, work is under way in the International Electrotechnical Commission (IEC) to improve the accuracy of hydrodynamic noise prediction.

Flashing is rarely a significant source of valve noise, although it can cause valve trim erosion damage in some cases. Flashing produces increasing valve exit velocity and downstream piping velocity as a result of the higher specific volume of the two-phase flow. In cases where sonic flow and shock cells develop in downstream piping, excessive noise can result. Expanded outlet valves and larger down-stream piping will be required under conditions where a large percentage of the liquid undergoes flashing. At this time there is not a standard method for predicting noise from flashing.

Aerodynamic Noise

In control valve design, aerodynamic noise can be a major prob-lem. It is a category of valve noise capable of generating noise levels of 120 dBA or greater. Noise produced by fluid turbulence in liquids is almost negligible as compared to the noise generated by the turbulence and shock cells due to the high velocity of gases and vapors passing through the valve orifice.

The mechanisms of noise generation in valves and trans-mission through pipe walls are highly complex and are still not completely predictable. As a result of the many variables influencing noise generation and the need for simplifying assumptions in calculations, predicting the noise levels from

valves or atmospheric exhaust vents is an inexact science. However, universities, manufacturers, and interested techni-cal societies have made much progress, which has resulted in better noise prediction methods based on scientific funda-mentals, which will be discussed in the section on Noise Calculations.

Aerodynamic noise generation, in general, is a function of mass flow rate and the pressure ratio (p1/p2) across the valve. The point at which sonic speed is reached in the valve vena contracta is a function of the valve design and its pres-sure recovery coefficient, FL, combined with the ratio of upstream to downstream absolute pressure (p1/p2). For exam-ple, valves with FL values of 0.5 and 0.95 require pressure ratios of 1.15 and 1.80, respectively, to generate sonic flow in the valve.

When sonic velocity is reached at the vena contracta, the valves are said to be choked, because their capacity does not increase if the pressure ratio is increased while the upstream pressure is kept constant. Generally, choked valves are the sources of the highest noise levels, but subsonic flows can also generate high noise levels. Valves that are not choked operate in a subsonic flow regime. For a given mass flow, they are less noisy than choked valves, but the noise level will increase as the pressure ratio approaches the sonic level.

Velocity of the flow in downstream pipe can also gen-erate significant noise starting at pipe velocities of about Mach 0.4 to Mach 1.0 (sonic). Noisy gas or vapor control valves can have acoustically induced and turbulence-induced vibration damage, trim wear, and control instabilities. High-intensity noise can produce vibration-related stresses at very high cycle rates (1,000–10,000 cps). Hence, noise-induced damage can drastically reduce valve service life, and in some cases, it can cause valve or piping failures in a matter of minutes or hours.

CONTROLLING NOISE

The transmission of a noise requires a source of sound, a medium through which the sound is transmitted, and a receiver. Each of these can be changed to reduce the noise level. In cases when the noise is from vibrating control valve components, the vibrations must be eliminated or they might result in valve failure. In cases when the source of noise is the hiss of a gas-reducing station, the acoustical treatment of the noise is sufficient.

Depending upon the magnitude of the aerodynamic noise and assuming that massive valve damage is not a factor, valve noise treatment can be accomplished either by path treatment or source treatment. Valve damage can only be reduced or eliminated by source treatment, which minimizes or elimi-nates the damage mechanism.

There is no absolute rule that will enable one to choose between path or source treatment. However, in general, if the noise is under 100 dBA, then either a path or source treatment is a possible solution. Noise above 100 dBA almost always



requires source treatment to successfully solve the noise problem.

The proper choice of noise treatment method is not always easy to select, but with the help of improved noise predictions through frequency spectrum evaluation and with expertise based on experience it can be obtained. Conserva-tive solutions are preferred, because reworking or retrofitting in case of poor design is often very expensive.

Path Treatment

Path treatment, as its name implies, does not focus on chang-ing the noise source. The intent of path treatment is to atten-uate the noise transmission from the source to the receiver (ear). There are several common path treatments: the use of heavy wall pipe; installation of diffusers, mufflers, or silenc-ers; and application of acoustical insulation.

Path treatment is not always a more economical solution than source treatment, and economics must be evaluated for individual applications. For existing installations, path treat-ment may be used, not because it is the best solution, but because it may be the only feasible one.

Pipe Wall Thickness Heavy wall pipe reduces noise by increasing the transmission loss through the pipe wall. The amount of attenuation depends on the stiffness and mass of the pipe. The mechanisms are complex and beyond the scope of this text. However, as a simple rule for rough estimation, each doubling of pipe wall thickness results in approximately 6 dBA more attenuation depending upon pipe size (attenuation increases with pipe size).

Refer to the works of Fagerlund and Chow (1981) and Reethof and Ward (1986) for important foundation in cal-culating transmission losses through pipe walls. Sample calculations will be introduced in the Noise Prediction section.

Insulation and Absorption Another method of increasing transmission loss at the pipe wall is the use of acoustic insu-lation. Even thermal insulation can add 3–5 dBA attenuation. Proper selection and application of 1–2 in. (25–50 mm) of a good acoustic insulation can reduce the noise level by roughly 10 dBA.

Certain types of insulation are more effective at specific frequency bands, so this information is important for proper selection. Because sound travels down the pipeline with very little attenuation over long distances, increasing the pipe wall thickness or applying acoustical insulation can be a very expensive solution. This approach is most useful when down-stream piping runs are short.

The higher the frequency of vibration, the more effective are the commercially available sound absorption materials. Figure 6.14g gives an example of acoustical treatment for the outside of a pipe.

It is often beneficial to cover the inside walls of the build-ing with sound-absorbing materials to prevent the reflection

and radiation of the sound waves from process equipment. If a valve is installed close to a single reflective surface (e.g., a hard floor or wall), the apparent noise increases by 3 dBA; with two reflective surfaces, noise increases by 6 dBA; and for three nearby reflective surfaces, noise increases by 9 dBA. A valve installed in a small room with all reflective surfaces can elevate noise levels by 30–40 dBA. When using walls as sound barriers, it is important to seal all openings.

Isolation Locating a potentially noisy valve installation at a substantial distance from normal working areas may be effective and economical. If the valve can be located on top of a structure or pipe bridge, the distance attenuation can minimize the noise treatment required at the valve source and downstream pipe.

For example, instead of a control valve with a noise specification of 85 dBA, it may be possible to relax this to 90 or 95 dBA, which can considerably reduce the cost of the valve or noise treatment system. It is important to note that very little noise actually radiates from the valve itself, due to generally heavy wall thickness and rigidity of most valve bodies; downstream piping radiates the great majority of noise produced in the valve to the surroundings.

As a general rule, each doubling of a person’s distance from the piping downstream of a valve will reduce the sound level by 3 dBA in a nonreflective environment. For example, the sound level that a person hears at 8 m from a valve will be 9 dBA quieter than the sound level at 1 m, and at 16 m it will be 12 dBA quieter than at 1 m.

Diffusers, Mufflers Diffusers located downstream of the valves (Figure 6.14h) can be helpful in both original instal-lations and retrofit situations. These devices can aid in reduc-ing exit flow turbulence or shock. Another important function of the multiple-hole design of diffusers depends on the fact that sound frequency increases as the size of the flow passage decreases. Using many small holes forces the dominant fre-quency of the turbulence into a higher range to which human hearing is less sensitive.

Diffusers can be designed to serve as pressure drop devices to reduce the pressure drop across the control valve

FIG. 6.14g Acoustical treatment of pipe walls.

Approximately 1''

Space necessaryfor flange bolting

Gustin bacon snap on insulationseal wrapped with glass cloth

impregnated with resin



and thus reduce its noise generation. The valve and diffuser system works best in a situation when the flow rate is sub-stantially constant or at least does not vary over a wide range. As a restrictor, the diffuser’s effectiveness in generating back-pressure on the valve decreases substantially as the flow rate drops. However, this shift of pressure drop back to the control valve does not necessarily increase noise, because when this occurs, the lower mass flow produces less noise.

Figure 6.14i illustrates a silencer design that can be installed downstream of a gas-regulating valve. Due to the resulting acoustical attenuation, it can reduce the sound pres-sure by a factor of five (e.g., from 96 to 82 dBA).

Mufflers or silencers (Figure 6.14j) can be used for in-line path treatment or for atmospheric vents. These are usu-ally expensive devices, with the cost escalating dramatically with size.

Dissipative or dissipative/reactive silencers are most com-monly used, but a comprehensive discussion of these devices is beyond the scope of this text. However, there are some rough

guidelines for application. Inlet velocity must be subsonic, and the silencer cannot be sized to serve as a pressure reducer. An inlet diffuser (as shown in Figure 6.14j) can be helpful, because it breaks up turbulence or shock cell oscillations that often occur in downstream sound fields and reduce the effectiveness of the unit. The outer shell should have a thick enough wall to prevent resonance. Materials of construction are selected to meet process conditions and to retain absorptive materials.

Source Treatment

Source treatments reduce noise by limiting sound power gen-erated at the source. In most cases, treatment consists of a special control valve and trim design, sometimes combined with a special diffuser or back-pressure element. While they may differ in concept, design, and manufacturing technolo-gies, these special systems are designed for one or more of the following objectives:

• Reduce pressure drop in stages • Limit fluid velocity to subsonic levels • Reduce or eliminate the formation of high turbulence

and shock cells • Shift as much sound power as possible into higher fre-

quency bands that have greater transmission losses in the pipe wall and have reduced response by human hearing

Depending upon the particular design, noise can be reduced with relatively inexpensive, simple elements by 7–10 dBA, whereas the more sophisticated valve designs or multielement systems can accomplish as much as 30–40 dBA attenuation from an untreated configuration.

Specification of source treatment valves or systems is not a simple matter. There are a number of design considerations, including the following:

• Application: in-line or vent• Noise reduction required or maximum SPL allowed

(dBA)

FIG. 6.14h Acoustical diffusers are used to reduce the exit turbulence down-stream of the valve. (Courtesy of Emerson Process Management.)

Flange to customerspecifications

Optional flange

FIG. 6.14i Silencer for gas regulating stations. (From King, C. F., “Control Valve Noise,” Emerson Process Management.)

8'' butterflyregulator

Gustin baconfiberglass

Perforatedmetal screen

10'' plugvalve

Flow

10'' specialflange

Retainerplate

10'' weld neck flange10'' pipe

Retainerplate



• Valve absolute pressure ratio: p1/p2 or ∆ p/p1

• Pressure drop, ∆ p• Fluid properties• Temperature operating level and range• Mass flow rate and turndown• Metallurgy and mechanical design considerations• Other potential velocity-induced problems• Valve shut-off requirements• Valve service life• Valve location and orientation, piping arrangement,

valve support, and maintenance access• Actuation and control requirements• Economics, including purchase, installation, and

maintenance costs

The importance of each factor is a matter of judgment and experience, understanding all aspects of the application and plant operation. Thus it is up to the user to carefully weigh and evaluate all vendor proposals. If a vendor is using one of the standardized noise prediction methods, the user should verify that accurate input values were specified and used for the calculations. Vendors that offer proprietary noise prediction calculations should also offer empirical justifica-tion to support their noise prediction.

While initial cost is one factor, this is not the only con-sideration; it may be that the more expensive equipment is the most economical solution in the long run. Plant downtime and retrofit costs for deficient valve noise solutions are usu-ally very expensive. With these caveats, there is much good equipment available, and vendor expertise and experience can be very valuable to the user with limited experience in controlling valve noise.

Basically, source treatment control valve designs fall into three categories: multipath, multistage, and combination of multipath/multistage. These designs are listed in order of sophistication and capability for noise reduction under severe

operating conditions. As might be expected, cost also in-creases. The cost of these special designs tends to range from 2–20 times the cost of a standard valve with the same flow capacity.

Multipath Valves The multipath valve (Figure 6.14k) pro-vides multiple orifices in parallel. A cylindrical plug to vary the flow rate through the valve uncovers these orifices. Although the path shape may vary by manufacturer, the principle consists of splitting the single-path flow into a large number of small paths. (See also the designs in Figures 6.1y and 6.1z.)

The noise radiated outside the pipe from the combined flow through multiple small paths is much less than that from the same flow rate through a single path restriction. Typical attenuation levels are 7–10 dBA but may reach 12–15 dBA in some applications. Variations of the multipath design are used for both hydrodynamic and aerodynamic valve noise with damage potential of low to moderate severity. Typically, compressible fluid pressure ratios (p1/p2) of 1.5–5 with valve exit velocities below Mach 0.33 are good candidates for this design.

FIG. 6.14j Silencers are installed in the flow path to dissipate the sound energy by absorbing it in an acoustical pack. They are designed to cause less than 1 PSI pressure drop. In-line silencers are often the most economical means of noise control in applications where the mass flow rate is high and the pressure drop is low. These units are normally installed immediately downstream of control valves, but in some cases they may also be required upstream of the valves. (Courtesy of Emerson Process Management.)

Shell closureAcoustical absorption material

Silencer core assemblyInlet

diffuser

ShellNozzle

FIG. 6.14k A multipath valve design, which can provide moderate noise reduc-tion. (Courtesy of Emerson Process Management.)

A A

Flow

Section A−A



Multistage Valves Pure multistage valves force the flow through a single path of two or more restrictions in series. (Figure 6.1y gives some examples.) An example of this design is shown in Figure 6.14l. The multiple orifices in series divide the total valve system pressure drop over several stages (typ-ically three to nine). Thus, the reduced pressure drop per stage results in greater friction loss, reduced local velocities, and reduced noise.

The shape of the trim element allows an increasing effec-tive flow area between the inlet and the outlet to compensate for the change in gas density and increase in specific volume. Thus, the outlet flange size is often larger than the inlet to limit the exit velocity to a level that will not regenerate exces-sive noise. Typically, these valves can provide noise attenua-tion up to 25 dBA, depending on pressure ratio and exit Mach number.

Resistor Elements In addition to diffusers, other special designs of multiple orifice restrictors are available (Figure 6.14m). These devices are built in a wafer design for install-ing between flanges and can be used in single- or in multi-stage configurations.

Such resistor elements can be installed in series, as shown in Figure 6.14n. These resistor plates are designed to work in series with the control valve to share the total pressure drop in a way that reduces pressure ratios on each element, thereby reducing the potential to generate noise.

The design of some of these devices forces the fluid through multiple changes in direction, acting like friction

elements with noise attenuation capability of several dBA. These multiple orifice restrictors are very useful in valve noise control work, but like diffusers, they lose effectiveness with flow turndown.

Combination Valves Combination multipath and multi-stage valves are usually required for the more severe and high noise producing services. These are the workhorse designs for the really tough applications, especially those that can cause extensive valve trim damage due to erosion or cavitation or noise levels in excess of 100 dBA. Various manufacturers have taken different approaches to the design of this type of valve.

Many are based upon multiple orifices in series and par-allel with the flow controlled by a close-fitting cylindrical plug inside the cage for throttling (Figure 6.14o).

A variation of this design also incorporates a secondary diffuser element built into the valve (Figure 6.14p).

One manufacturer has designed a valve for moderate pres-sure drops using a standard valve trim assembly that elimi-nates the close-fitting cylindrical plug and uses a noise atten-uator element ranging from one to seven stages (Figures 6.14q and 6.14r).

Another design utilizes the pressure loss producing effects of a fluid passing through a series of sharp turns

FIG. 6.14l Multistage step trim valve for use on compressible fluids. Outlet is expanded to compensate for volume change. (Courtesy of Dresser Flow Solutions.)

FIG. 6.14m Resistor element used for valve back-pressure and noise reduction. (Courtesy of Dresser Flow Solutions.)

FIG. 6.14n Noise can be reduced by resistor elements that are installed in series. (Courtesy of Dresser Flow Solutions.)

Flow



machined into a set of stacked disks (Figure 6.14s). Other similar designs are illustrated in Figure 6.1y and 6.1aa.

Depending upon the manufacturer and design, multi-stage, multiorifice valves will typically have 2–7 stages, although some might use 20 or more. However, the number of stages required by any specific design for a given appli-cation depends on the design principles employed and their effectiveness in the application. In other words, having more stages does not necessarily make a valve quieter than another design with fewer stages. Inexperienced users are advised to

require some validation of manufacturers’ claims of noise reduction for these critical service valves in addition to their noise calculations.

AERODYNAMIC NOISE PREDICTION

Valve noise prediction is an inexact science because of the complex nature of noise generation by the control valve and the transmission of this noise through the pipe wall. So it is not surprising that a number of different prediction methods are used by manufacturers and others. What is surprising is that the various methods can give answers for the same appli-cation that differ by up to 20 dBA.

The subject of valve noise prediction is still subject to continuing research and evaluation. So what should we do?

FIG. 6.14o Multipath and multistage valve with shaped first stage holes. (Cour-tesy of Emerson Process Management.)

FIG. 6.14p Multipath and multistage valve with integral secondary diffuser element. (Courtesy of ABB Control Valves.)

FIG. 6.14q Two-stage noise attenuator in a valve, which was designed for use with a standard inner valve trim assembly. (Courtesy of Flowserve Corporation.)

FIG. 6.14r Multistage noise attenuator (and detail of that attenuator) in a valve, which was designed for use with a standard trim assembly with pressure balance. (Courtesy of Flowserve Corporation.)

Bonnet gasketBody

Seat ringgasket Seat ring Plug

Bonnet Bonnet flangebolting

Bonnetflange

Attenuator



Each manufacturer claims to be able to predict the valve noise level and provide a valve design solution. It finally falls upon the user to obtain the best possible process data, carefully evaluate all proposals, ask questions and resolve marked dif-ferences, and finally use good engineering judgment and experience in selecting the vendor for each application.

It is wise to err on the conservative side when making a final selection, because the cost of mistakes and of the required retrofit may far outweigh valve cost differentials. Fortunately, the noise prediction standards of the various standards organi-zations have helped to make comparisons of noise predictions somewhat easier for users and manufacturers alike.

Standards

The past quarter-century has seen continuous improvements in the standardized methods and in the industrial acceptance of noise prediction standards for valves. In 1979 the Verband Deutcher Maschinen- und Anlagenbau e.V. (VDMA) published the first standardized method of calculating the sound level for valves as Standard VDMA 24422, which addressed both hydro-dynamic and aerodynamic noise. VDMA revised 24422 in 1989 to include calculations of the frequency domain.

The weakness of the VDMA method was that key valve noise parameters had to be determined experimentally; when testing was not practical, prediction accuracy was unsatisfac-tory. Meanwhile, other organizations were developing pre-diction methods based on free jet turbulence theories.

The Instrumentation, Systems, and Automation Society (ISA) published standard ISA-75.17 in 1989, and the

International Electrotechnical Commission published the first edition of IEC 534-8-3 in 1995 and the second edition IEC 60534-8-3 in 2000. The basic methods in these standards are essentially the same. They both are based on the published works by Lighthill, Powell, Fowcs and Hawkins, Reethof and Ward, Shea, Fagerlund, Baumann, and the contributions of many others. These organizations update their respective standards when new information is validated.

The author has selected the more recent IEC Standard 60534-8-3 (2000) to demonstrate the basic calculation process. This standard and the field of acoustics in general use SI units. For conversion factors to U.S. Customary units, the reader should refer to Appendices A.1 and A.2 and is cautioned to use the proper gravitational unit conversions (i.e., gc ). To avoid excessive detail, only the SI system of units will be shown in this section.

Calculations

Acknowledgment: The author thanks the International Elec-trotechnical Commission for permission to reproduce informa-tion from its International Standard IEC 60534-8-3. All such extracts are copyright of IEC, Geneva, Switzerland. All rights reserved. Further information on the IEC is available from www.iec.ch. IEC has no responsibility for the placement and context in which the extracts and contents are reproduced by the author; nor is IEC in any way responsible for the other content or accuracy therein.

FIG. 6.14s Special noise element design using labyrinth passages incorporated on plates. (Courtesy of Control Components Inc.)

EDMdiskstack Punched

diskstack

Disk stack configurations


http://www.iec.ch


IEC 60534-8-3 The intent of this section is to familiarize the reader with the nomenclature and illustrate the basic procedure of the IEC 60534-8-3 aerodynamic noise predic-tion method. However, the standard covers additional spe-cial cases and details that are too extensive for coverage here. Like all theoretical methods, it is based on assump-tions and limitations that must be applied with appropriate engineering skill and judgment to practical applications. Some of the stated assumptions and limitations of IEC 60534-8-3 include

• The valve is installed with steel or alloy steel piping upstream and downstream, possibly with pipe expand-ers, and that “the downstream piping is straight for a length of at least 2 m from the point where the noise measurement is made.”

• The method assumes that the fluid properties can be modeled on the perfect gas laws.

• The method can be used for most valve types. How-ever, it is not applicable for full-bore ball valves where the product FpC (for the operating condition) exceeds 50% of the valve’s rated flow coefficient.

• The method shown for standard single-stage trims in this section applies only for valve outlet and down-stream pipe velocities with Mach numbers up to 0.3. Refer to IEC 60534-8-3 for multistage trims and higher velocities.

• Transmission loss of noise through the pipe wall is based on a simplified method due to the wide toler-ances in pipe wall thickness for commercial steel pipe.

• The calculated sound pressure level assumes a location 1 m downstream from the valve or expander and 1 m from the outside of the pipe wall in an acoustic free field.

• The prediction does not guarantee actual results in the field. Validation tests were conducted with low pres-sure air and steam in laboratory tests, and the majority of test results were within 5 dBA of the predicted sound pressure level.

Nomenclature The following list of symbols, definitions, and units is a partial set of those used in IEC 60534-8-3.

Symbol Description Unit

A Area of a single flow passage m2

An Total flow area of last stage of multistage trim with n stages at given travel

m2

C Flow coefficient (Kv and Cv) (see IEC 60534-2-1)

Various(see IEC 60534-1)

cvc Speed of sound in the vena contracta at subsonic flow conditions

m/s

cvcc Speed of sound in the vena contacta at critical flow conditions

m/s

c2 Speed of sound at downstream conditions m/s

D Valve outlet diameter m

d Diameter of a circular flow passage m

dH Hydraulic diameter of a single flow passage

m

di Smaller of valve outlet or expander inlet internal diameters

m

Di Internal downstream pipe diameter m

Dj Jet diameter at the vena contracta m

do Diameter of a circular orifice, the area of which equals the sum of areas of all flow passages at a given travel

m

Fd Valve style modifier = dH /do Dimensionless

FL Liquid pressure recovery factor of a valve without attached fittings (see note 4)

Dimensionless

FLP Combined liquid pressure recovery factor and piping geometry factor of a control valve with attached fittings (see note 4)

Dimensionless

FP Piping geometry factor Dimensionless

fg External coincidence frequency Hz

fo Internal coincidence pipe frequency Hz

fp Generated peak frequency Hz

fpR Generated peak frequency in valve outlet or reduced diameter of expander

Hz

fr Ring frequency Hz

l Length of a radial flow passage m

lw Wetted perimeter of a single flow passage m

Lg Correction for Mach number dB (ref po)

LpAe A-weighted sound pressure level external of pipe

dBA (ref po)

LpAe,1m A-weighted sound pressure level 1 m from pipe wall

dBA (ref po)

Lpi Internal sound pressure level at pipe wall dB (ref po)

Lwi Total internal sound power level dB (ref Wo)

M Molecular mass of flowing fluid kg/kmol

Mj Freely expanded jet Mach number in regimes II to IV

Dimensionless

Mj5 Freely expanded jet Mach number in regime V

Dimensionless

Mo Mach number at valve outlet Dimensionless

Mvc Mach number at the vena contracta Dimensionless

M2 Mach number in downstream pipe Dimensionless

Mass flow rate kg/s

Mass flow rate at sonic velocity kg/s

N Numerical constants (see Table 6.14v) Various

No Number of independent and identical flow passages in valve trim

Dimensionless

pa Actual atmospheric pressure outside pipe Pa (see note 3)

po Reference sound pressure = 2 × 10–5 (see note 5)

Pa

m

ms



Method Outline Numerous equations are involved in the calculation procedure, but a brief outline of the five general steps will ensure continuity in the procedures.

1. Gather the necessary input data• Valve sizing data and dimensions of trim and body

ports• Configuration and dimensions of adjacent piping• Service conditions and fluid properties

2. Calculate key pressures and pressure ratios, and deter-mine the noise regime

3. Calculate the effective jet diameter4. Calculate jet conditions, acoustic efficiency, sound

power, peak frequency for the noise, and internal sound pressure level

5. Calculate pipe natural frequencies, pipe transmission loss, and external sound pressure level

Noise Regimes There are five key noise regimes identified in IEC 60534-8-3, which have different mechanisms governing the generation and transmission of sound. In order to determine which regime applies to a given set of conditions, several important pressures and pressure ratios must be calculated and compared with the actual downstream pressure, p2.

The vena contracta is the region of flow constriction with maximum velocity and minimum pressure, pvc, given in Equation 6.14(7).

6.14(7)2

If the valve has attached fittings, FL is replaced by the value of FLP / FP . At critical flow conditions, the vena con-tracta pressure is p vcc.

6.14(8)

The downstream pressure at the critical pressure drop where sonic flow begins at the vena contracta is P2C.

6.14(9)

At the break point pressure, P2B , shock cell turbulent interaction begins to dominate noise generation. For down-stream pressures greater than p2B , turbulent shear flow gen-erates most of the sound power.

6.14(10)

where α is a correction factor defined as

ps Standard atmospheric pressure (see note 1) Pa

pvc Absolute vena contracta pressure at subsonic flow conditions

Pa

pvcc Absolute vena contracta pressure at critical flow conditions

Pa

p1 Valve inlet absolute pressure Pa

p2 Valve outlet absolute pressure Pa

p2B Valve outlet absolute pressure at break point

Pa

p2C Valve outlet absolute pressure at critical flow conditions

Pa

p2CE Valve outlet absolute pressure where region of constant acoustical efficiency begins

Pa

R Universal gas constant = 8314 J/kmol × K

rw Acoustic power ratio (see Table 6.14w) Dimensionless

Tvc Vena contracta absolute temperature at subsonic flow conditions

K

Tvcc Vena contracta absolute temperature at critical flow conditions

K

T1 Inlet absolute temperature K

T2 Outlet absolute temperature K

TL Transmission loss dB

tp Pipe wall thickness m

Up Gas velocity in downstream pipe m/s

Uvc Vena contracta velocity at subsonic flow conditions

m/s

Wa Sound power W

Wm Stream power of mass flow W

Wms Stream power of mass flow rate at sonic velocity

W

Wo Reference sound power = 10–12 (see note 5) W

α Recovery correction factor Dimensionless

β Contraction coefficient for valve outlet or expander inlet

Dimensionless

γ Specific heat ratio Dimensionless

η Acoustical efficiency factor (see note 2) Dimensionless

ρ1 Density of fluid at p1 and T1 kg/m3

ρ2 Density of fluid at p2 and T2 kg/m3

Φ Relative flow coefficient Dimensionless

Note 1. Standard atmospheric pressure is 101.325 kPa or 1.01325 bar.

Note 2. Subscripts 1, 2, 3, 4, and 5 denote regimes I, II, III, IV, and V, respectively.

Note 3. 1 bar = 102 kPa = 105 Pa.

Note 4. For the purpose of calculating the vena contracta pressure, and therefore velocity, in this standard, pressure recovery for gases is assumed to be identical to that of liquids.

Note 5. Sound power and sound pressure are customarily expressed using the logarithmic scale known as the decibel scale. This scale relates the quantity logarithmically to some standard reference. This reference is 2 × 10−5 Pa for sound pressure and 10−12 W for sound power. 2 Equations 6.14(7)–(46) and 6.14(48)–(53) are used here by permission of

IEC. Copyright © 2000, IEC, Geneva, Switzerland.www.iec.ch.

p pp p

FLvc = −

−1

1 22

p pvcc =+

−

1

12

1γ

γ γ( )

p p F p pC L2 12

1= − −( )vcc

pp

B21

11=

−

α γ

γ γ( )


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6.14(11)

Regime V begins when downstream pressure drops to P2CE and is where acoustical efficiency becomes constant. Further reductions in downstream pressure will not increase the sound pressure level.

6.14(12)

Table 6.14t summarizes the boundaries and characteris-tics of Regimes I through V.

Calculate Jet Diameter Determining the jet diameter requires information about the valve trim dimensions in order to calculate the valve-style modifier Fd , which is the ratio of hydraulic diameter, dH, of a single flow passage to the circle diameter, do , corresponding to the total flow area.

6.14(13)

6.14(14)

6.14(15)

Lacking specific dimensions, approximate values of Fd

are given in Table 6.14u.The jet diameter, Dj, is calculated from Equation

6.14(16).

6.14(16)

where units conversion factor N14 is found in Table 6.14v and depends on whether the required flow coefficient, C, is given as Cv or Kv . (Refer to Section 6.15 in this chapter for infor-mation about flow coefficients.)

α ≡

=

p

p

p

p

C p

p C

1

2

1 2

vcc

vcc

pp

21

22CE =α

Fd

ddH

o

=

dA

lHw

= 4

dN A

oo=

4

π

TABLE 6.14t Characteristics of IEC Noise Regimes

RegimeDownstream

Pressure Description

I p1 > p2 ≥ p2C Subsonic flow; isentropic recompression; turbulent shear noise

II p2C > p2 ≥ pvcc

Sonic flow at vena contracta; isentropic recompression; turbulent shear noise

III pvcc > p2 ≥ p2B

Supersonic flow past the vena contracta; no recompression; noise from shock turbulence and shear turbulence

IV p2B > p2 ≥ p2CE

Sonic flow at vena contracta; supersonic Mach cone terminates in Mach disk at outlet; shock interaction dominates noise

V p2CE > p2 Supersonic Mach cone reaches maximum Mach number; acoustical efficiency and noise level are constant

D N F CFj d L= 14

TABLE 6.14u Typical Values of Valve-Style Modifier Fd (Full-Size Trim)

Relative Flow Coefficient Φ

Valve Type Flow Direction 0.10 0.20 0.40 0.60 0.80 1.00

Globe, parabolic plug To openTo close

0.100.20

0.150.30

0.250.50

0.310.60

0.390.80

0.461.00

Globe, 3 V-port plug Either* 0.29 0.40 0.42 0.43 0.45 0.48



Globe, 60 equal diameter hole drilled cage Either* 0.40 0.29 0.20 0.17 0.14 0.13

Globe, 120 equal diameter hole drilled cage Either* 0.29 0.20 0.14 0.12 0.10 0.09

Butterfly, swing-through (centered shaft), to 70° Either 0.26 0.34 0.42 0.50 0.53 0.57

Butterfly, fluted vane to 70° Either 0.08 0.10 0.15 0.20 0.24 0.30

Butterfly, 60° flat disk Either 0.50

Eccentric rotary plug Either 0.12 0.18 0.22 0.30 0.36 0.42

Segmented ball 90° Either 0.60 0.65 0.70 0.75 0.78 0.98

NOTE: These values are typical only; actual values are stated by the manufacturer.

* Limited p1 – p2 may apply in flow-to-close direction.

Copyright © 2000, IEC, Geneva, Switzerland. www.iec.ch.


http://www.iec.ch


Regime I Calculations Calculate the following subsonic parameters for the vena contracta.

Gas velocity:

6.14(17)

Stream power:

6.14(18)

Absolute temperature:

6.14(19)

Speed of sound:

6.14(20)

Mach number:

6.14(21)

With this information, calculate the acoustical efficiency factor, η1, sound power, Wa, and peak frequency, fp.

6.14(22)

6.14(23)

where rw is the acoustic power ratio taken from Table 6.14w.

Peak frequencies in Regimes I and II are based on Strou-hal’s equation with the Strouhal number = 0.2.

6.14(24)

Common Calculations for Regimes II–V For sonic condi-tions in the vena contracta, calculate the following parameters.

Vena contracta temperature:

6.14(25)

Velocity of sound:

6.14(26)

Stream power:

6.14(27)

Mach number in the freely expanding jet:

6.14(28)

TABLE 6.14v Numerical Constants N

Constant

Flow Coefficient

Kv Cv

N14 4.9 × 10–3 4.6 × 10–3

N16 4.23 × 104 4.89 × 104

Note: Unlisted numerical constants are not used in this standard


Up

pvcvc=

−

−

−

21

11

1γ

γ

γ γ( )

p1

1ρ

Wm U

m =( )vc

2

2

T Tp

pvcvc=

−

11

1( )γ γ

cRT

Mvcvc=

γ

MU

cvcvc

vc

=

η141 10

3 6

= × −( ).

Mvc

W r W Fa w m L= η12

TABLE 6.14w Acoustic Power Ratio rw

Valve or Fitting rw

Globe, parabolic plug 0.25

Globe, 3 V-port plug 0.25



Globe, 60 equal diameter hole drilled cage 0.25

Globe, 120 equal diameter hole drilled cage 0.25

Butterfly, swing-through (centered shaft), to 70° 0.5

Butterfly, fluted vane, to 70° 0.5

Butterfly, 60° flat disk 0.5

Eccentric rotary plug 0.25

Segmented ball 90° 0.25

Expanders 1


fU

Dpj

=0 2. vc

TT

vcc =+

2

11

γ

cRT

Mvccvcc=

γ

Wmc

msvcc2

=2

Mp

pj =−

−

−2

111

2

1

γ α

γ γ( )


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Next, the acoustical efficiency factors, sound power, and peak frequency are calculated for the regime in question.

Regime II Acoustical efficiency factor:

6.14(29)

Sound power:

6.14(30)

Peak frequency:

6.14(31)

Regime III Acoustical efficiency factor:

6.14(32)

Sound power:

6.14(33)

Peak frequency is calculated from Equation 6.14(31).

Regime IV Acoustical efficiency factor:

6.14(34)

Sound power:

6.14(35)

Peak frequency:

6.14(36)

Regime V Jet Mach number reaches it maximum:

6.14(37)

The acoustical efficiency factor becomes constant:

6.14(38)

Sound power generated in this regime that radiates into downstream pipe is

6.14(39)

Peak frequency:

6.14(40)

Noise Calculations The following calculations are used for all regimes.

Downstream mass density:

6.14(41)

Downstream sonic velocity:

6.14(42)

where T2 may be found from thermodynamic isenthalpic relationships. If fluid properties are not known, reasonable results can be obtained by assuming T2 is approximately equal to T1.

Mach number at valve outlet:

6.14(43)

If the outlet Mach number Mo is above 0.3, accuracy of this method diminishes. IEC 60534-8-3 clause 7 provides further procedures for high Mach number applications, which is outside the discussion of this basic process.

The internal sound pressure level, Lpi, referenced to 2 ×10–5 Pa is calculated in dB from the following:

6.14(44)

η24 6 61 10

2

= × −( ) .M jFL

W r Wp p

p pa w=−−

η2

1 2

1ms

vcc

fM c

Dpj

j

=0 2. vcc

η34 6 61 10

2

= × −( ) .M jFL

W r Wa w= η2 ms

η44

2 6 61 10

22

2

= ×

( )−( )

.M j FL

W r Wa w= η4 ms

fc

D Mp

j j

=−

0 35

1 25 12

.

.vcc

M j512

122 1=

−−−

γγ γ[( ) ]( )

η54 5

2 6 61 10

22

2

= ×

( )−( )

.M j FL

W r Wa w= η5 ms

fc

D Mp

j j

=−

0 35

1 25 15

.

.vcc

2

ρ ρ2 12

1

=

p

p

cRT

M22=

γ

Mm

D co = 42

2 2π ρ

LW c

Dpia

i

=×

103 2 10

10

92 2

2log

( . ) ρ



Transmission through the Pipe The pipe wall must be made to vibrate in order for noise inside the pipe to radiate into the air outside the pipe. The mode of pipe vibration, for the purpose of this prediction method, is determined from the peak frequency of the noise source and the natural fre-quencies of the pipe.

The assumption is made that the shape of the sound frequency spectrum is an arc or “haystack”-shaped curve that reaches a pronounced maximum level at peak frequency, fp. Although this is true for most valves, some configurations can possess “flatter” broadband spectra that could radiate more noise than the simplified model predicts.

Pipe natural frequencies are functions of the pipe diam-eter, wall thickness, and density. The transmission loss model used by the IEC standard is based on the work of Fagerlund and Chow. The important characteristic frequen-cies are explained in detail by Singleton and are summarized below.

Ring frequency, fr , has a wavelength exactly equal to the circumference of the pipe, which produces a resonant stress wave around the circumference.

6.14(45)

External coincidence frequency, fg, corresponds to the external acoustic wave speed that matches the speed of a flexural wave in pipe wall. Assuming the speed of sound in steel is 5000 m/s and 343 m/s in air,

6.14(46)

First internal coincidence frequency, fo, is the lowest natural frequency of the pipe wall and produces a longitudinal flexural wave that spirals along the length of the pipe.

6.14(47)

Cutoff frequency, fc, though not part of the IEC standard, is significant because at the cutoff frequency and below, the wavelengths are too long to reflect off the internal pipe wall, making them incapable of vibrating the pipe.

6.14(48)

The relationship of the peak frequency in the flow stream to the pipe natural frequencies is used to calculate the fre-quency factors used in the transmission loss calculation. Table 6.14x, taken from IEC 60534-8-3, shows how fre-quency factors Gx and Gy are determined.

The transmission loss across the pipe wall is calculated from Equation 6.14(49).

6.14(49)

Next, calculate the downstream pipe velocity correction factor, Lg.

6.14(50)

where M2 should not exceed 0.3 and is calculated by

6.14(51)

The A-weighted sound pressure level radiated from the outside surface of the pipe is given a 5 dB correction to account for all frequency peaks and is calculated below.

6.14(52)

Finally, a distance adjustment is made to calculate the sound pressure level in dBA at 1 m from the pipe wall.

6.14(53)

Noise Calculation Example

These calculations are typically carried out with computer software and presented as part of the sizing calculations done by valve manufacturers. For a thorough understanding, a simple calculation example is tabulated below.

fDr

i

= 5000π

ftg

p

= 3 3435000

2( )( )π

fD

c

c

f co

i o

r=

=

12504 343

2 2

π

fc

Dci

= 0 586 2.

TABLE 6.14x Frequency Factors Gx and Gy

fp < fo fp ≥ fo

for fp < fr

Gx = 1 for fp ≥ fr

for fo < fg

Gy = 1 for fo ≥ fg

for fp < fg

Gy = 1 for fp ≥ fg

Copyright © 2000, IEC, Geneva, Switzerland, www.iec.ch

Gf

f

f

fxo

r

p

o

=

2 3 4

Gf

fxp

r

=

2 3

Gf

fyo

g

=

G

f

fyp

g

=

TLc

t f

G

p p

x

c= ×

−10 7 6 10107 2

2

412 2

log ( . )ρ

55 1G

a

sy

p

p+( )

LMg =

−

161

1102

log

Mm

D ci2 2

2 2

4=π ρ

L L TL Lpi gpAe = + + +5

L LD t

D ti p

i ppAe,1m pAe= −

+ ++

102 2

210log


http://www.iec.ch


Example 1. Steam Valve

Inputs

Valve: NPS 4-in. (DN 100)Trim: Parabolic plugC = Cv = 152 required (from sizing calculations)FL = 0.90 Φ = 0.60Fd = 0.31Maximum allowable noise level: 90 dBA

Pipe: NPS 4-in. (DN 100) Schedule 40 carbon steelDi = 0.102 mtp = 0.00602 m

Fluid: SteamM = 18.02 kg/kmolT1 = 260 °C = 533 Kp1 = 2.5 MPap2 = 1.7 MPa

γ = 1.32From steam tables: ρ1 = 11.17 kg/m3

T2 = 247 °C = 520 Kρ2 = 7.58 kg/m3

Preliminary Calculations

Variable Equation Results

pvc 6.14(7) 2.5 − (2.5 − 1.7)/(0.9)2 = 1.512 MPa

pvcc 6.14(8) 2.5[2/(1.32 + 1)]1.32/0.32 = 1.355 MPa

P2C 6.14(9) 2.5 − (0.9)2 (2.5 −1.355) = 1.573 MPa

α 6.14(11) 1.355/1.573 = 0.861

P2B 6.14(10) (2.5/0.861) (1/1.32)1.32/0.32 = 0.924 MPa

P2CE 6.14(12) 2.5/[(22) (0.861)] = 0.132 MPa

Regime? Table 6.14t p2 ≥ p2C : 1.7 ≥ 1.573 ∴ Regime I

Dj 6.14(16) (4.6 × 10−3)(0.31)(152 × 0.9)1/2 = 0.0167 m

Regime I Calculations

Uvc 6.14(17) {2(1.32/0.32)[1(1.512 / 2.5)0.32/1.32](2.5 × 106 /11.17)}1/2 = 460 m/s

Wm 6.14(18) (9.1)(460)2/2 = 9.63 × 105 W

Tvc 6.14(19) (533)(1.512/2.5)0.32/1.32 = 472 K

cvc 6.14(20) [(1.32)(8314)(472)/(18.02)]1/2 = 536 m/s

Mvc 6.14(21) 460/536 = 0.858

Determine Internal Noise

η1 6.14(22) (1 × 10–4)(0.858)3.6 = 5.76 × 10-−5

rw Table 6.14w 0.25

Wa 6.14(23) (5.76 × 10–5)(0.25)(9.63 × 105) = 13.9 W

fp 6.14(24) (0.2)(460)/(0.0167) = 5,509 Hz

ρ2

6.14(41) or steam tables

7.58 kg/m3

c2 6.14(42) [(1.32)(8314)(520)/(18.02)]1/2 = 563 m/s

Mo 6.14(43) (4)(9.1)/[π(0.1016)2(7.58)(563)] = 0.263

Lpi 6.14(44) 10log[(3.2 × 109)(13.9)(7.58)(563)/(0.102)2] = 162.6 dB

Determine Radiated Noise

fr 6.14(45) (5000)/[π(0.102)] = 15.6 kHz

fg 6.14(46) (3)1/2(343)2/[π (0.00602)(5000)] = 2155 Hz

fo 6.14(47) (15600/4)(563/343) = 6401

Gx Table 6.14x fp< fo: (6401/15,600)2/3(5509/6401)4 = 0.303

Gy Table 6.14x fo ≥ fg: 1.0

m = 9.10 kg/s



Applying Distance Corrections

Placing extra distance between noisy equipment and people is sometimes a viable alternative to expensive noise reduction treatment, if there are no other detrimental effects of the noise at the source.

If a noise source can be treated as a point in the acoustic far field, the sound radiates in a spherical pattern. Atmo-spheric vents can be treated this way. The reduced sound pressure level at some distance, r, from the center of a point source can be determined from a measured or calculated sound pressure level taken at a reference distance of ro from the center (typically ro = 1 m + pipe OD/2) using Equation 6.14(54) below.

6.14(54)

Because noise produced by valves radiates to the envi-ronment largely through the pipe for great distances down-stream of a valve, this type of noise source is generally treated as a line source. Line sources radiate noise in a cylindrical pattern. The reduced sound-pressure level at distance, r, from a line source is

6.14(55)

Example 2. Valve Noise at a Distance Problem: Using the same valve and conditions from Example 1, what would be the sound pressure level for a worker 30 m away from the downstream pipe (centerline)?

Solution: From Example 1, LpAe,1m = 105 dBA, and pipe OD = 0.114 m. Use Equation 6.14(55).

HYDRODYNAMIC NOISE PREDICTION

Noise prediction for liquid flow through valves should con-sider three major flow regimes. 1) Turbulent flow, which, without cavitation, rarely produces noise levels high enough to create dangerous structural vibration or noise pollution. 2) Cavitating flow, which produces noise from vapor cavity formation and collapse as well as from turbulence, and it frequently causes excessive vibration and noise in addition to erosion of valve and piping materials. 3) Flashing of liquid into vapor across a valve sometimes causes high levels of noise and vibration, if vapor velocities in the downstream piping approach sonic velocities. Piping systems should be so sized as to avoid vapor or two-phase velocities, which are high enough to cause noise and erosion.

Hydrodynamic noise prediction is currently in a state of development. Noise prediction Standards VDMA 24422 (1989) and IEC 60534-8-4 (1994) have been shown by Kiesbauer and Baumann to predict lower than actual noise in many cases. A more accurate method of hydrodynamic noise prediction has been proposed (Kiesbauer, J. and Baumann, H. D., “Recent Developments in the Prediction of Hydrodynamic Noise of Control Valves,” Valve World, February 2004), which is being considered by the IEC as a revision to Standard 60534-8-4 at the time of this writing. This method includes calculations for turbulent flow and cavitating flow regimes. There are no stan-dards or generally accepted methods at this time for predicting noise under flashing conditions. For calculation of noise, the reader is advised to study the Kiesbauer-Baumann method or later revisions of IEC Standard 60534-8-4.

Hydrodynamic noise predictions use a differential pres-sure ratio, xF , to identify noise regimes.

6.14(56)

The incipient cavitation index, xFz, corresponds to the differential pressure ratio at which cavitation in a valve begins and should be determined from cavitation tests, although some of the methods include ways of estimating xFz. This is

TL 6.14(49)

M2 6.14(51) 4(9.1)/[π (0.102)2(7.58)(563)] = 0.261

Lg 6.14(50) 16log[1/(1 − 0.261)] = 2.10 dB

LpAe 6.14(52) 5 + 162.6 − 52.3 + 2.1 = 117.4 dBA

LpAe,1m 6.14(53) 117.4 − 10log{[0.102 + 2(0.00602) + 2]/[0.102 + 2(0.00602)] = 105 dBA

Conclusion: Noise level exceeds desired maximum; consider noise reduction trim; consult manufacturer.

10 7 6 10563

0 00602 55090 37

2

log ( . )( . )( )

.×

− 003

11

7 58 563

415 1

( . )( )

( )( )

( )+

= −52 3. dB

L Lrro

pAe,r pAe,1m= −

20 10log

L Lrro

pAe,r pAe,1m= −

10 10log

Lm

m mpAe,30m = −+

=

105 1030

1 0 114 2

10

10log. /

55 15 90− = dBA

xp p

p pFv

=−−

1 2

1



the index that separates the turbulent flow regime from the cavitating flow regime. (If xF ≥ 1.0, the liquid is flashing.) Each of the methods discussed above follows a general pro-cess similar to that for aerodynamic noise prediction:

1. Gather the necessary input data• Valve sizing data and dimensions of trim and body

ports• Configuration and dimensions of adjacent piping• Service conditions and fluid properties

2. Calculate key pressures and pressure ratios, and deter-mine the noise regime.

3. Calculate the effective jet diameter and stream power.4. Calculate acoustic efficiency and internal sound-pres-

sure level.5. Calculate pipe natural frequencies, pipe transmission

loss, and external sound-pressure level.

Bibliography

Arant, J. B., “Special Control Valves Reduce Noise and Vibration,” Chemical Engineering, March 6, 1972.

Arant, J. B., “How to Cope with Control Valve Noise,” Instrument Technol-ogy, March 1973.

Arant, J. B., “Control Valve Noise,” in Control Valves — Practical Guides for Measurement and Control, G. Borden and P. G. Friedmann (eds.), Research Triangle Park, NC: Instrumentation, Systems, and Automa-tion Society, 1998.

Ball, K. E., “Final Elements: Final Frontier,” InTech, November 1986.Barnes, R. W., “Understanding Automatic Control Valve Noise,” ISA Trans-

actions, Vol. 22, No. 1, Research Triangle Park, NC: Instrumentation, Systems, and Automation Society, 1983.

Baumann, H. D., “On the Prediction of Aerodynamically Created Sound Pressure Level of Control Valves,” ASME Paper 70 WA/FE-28, December 1970.

Baumann, H. D., “How to Estimate Aerodynamic Valve Throttling Noise: A Fresh Look,” ISA Paper #82-902, 1982.

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Baumann, H. D., “A Firm New Handle on Valve Noise,” Chemical Engi-neering, December 1996.

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Baumann, H. D. and Hoffmann, H., “Method for the estimation of frequency-dependent sound pressures at the pipe exterior of throttling valves,” Noise Control Engineering Journal, Vol. 47, No. 2, March–April 1999.

Baumann, H. D. and Page, G. W. Jr., “A Method to Predict Sound Levels from Hydrodynamic Sources Associated with Flow Through Throttling Valves,” Noise Control Engineering Journal, Vol. 43, No. 5, September–October 1995.

Boyle, S. J., “Acoustical Design Considerations for Low-Noise Control Valves,” Proceedings of Noise-Con 83, Institute of Noise Control Engi-neering, March 1983.

Brockbank, G. and Glenn, A., “Compressible Flow Noise,” Valve Magazine, Fall 1992.

CEI/IEC 60534-8-4 (1994-05), ‘‘Prediction of Noise Generated by Hydro-dynamic Flow,” Geneva, Switzerland: International Electrotechnical Commission, 1994.

Chow, G. C. and Reethof, G., “A Study of Valve Noise Generation Processes for Compressible Fluids,” ASME Paper 80 WA/NC-15, 1980.

CEI/IEC 60534-8-3 (2000-07), “Control Valve Aerodynamic Noise Predic-tion Method,” 2nd edition, Geneva, Switzerland: International Electro-technical Commission, 2000.

Fagerlund, A. C, “Use of Pipe Wall Vibrations to Measure Valve Noise,” ISA Paper #74-829.

Fagerlund, A. C., “Recommended Maximum Valve Noise Levels, InTech,November 1986.

Fagerlund, A. C. and Chow, D. C., “Sound Transmission through a Cylin-drical Pipe,” ASME Journal of Engineering for Industry, Vol. 103, No. 4, November 1981.

Glenn, A. H., Eliminating Screech-Type Noise in Control Valves, doctoral dissertation presented to the Department of Mechanical Engineering, Brigham Young University, Provo, UT, July 1987.

Hynes, K. M., “The Development of a Low Noise Constant Area Throttling Device,” ISA #839-70, October 1970.

ISA-75.17-1989, “Control Valve Aerodynamic Noise Prediction,” Research Triangle Park, NC: Instrumentation, Systems, and Automation Society, 1989.

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Kiesbauer, J. and Baumann, H. D., “Recent Developments in the Prediction of Hydrodynamic Noise of Control Valves,” Valve World, February 2004.

Kinsler, L. E., Frey, A. R., Coppens, A. B., and Sanders, J. V., Fundamentals of Acoustics, 3rd edition, New York: John Wiley & Sons, Inc., 1982.

Lighthill, M. J., “On Sound Generated Aerodynamically II, Turbulence as a Source of Sound,” Proceedings of The Royal Society of London,Vol. 222, Series A, 1954.

Lodge, J. M. and Turner, J. T., “Aerodynamic Noise Generation in Control Valves — An Experimental Study,” Paper F1, International Conference on Developments in Valves and Actuators for Fluid Control, sponsored by BHRA and BVAMA, Oxford, England, 10–13 September 1985.

Ng, K. W., “Control Valve Aerodynamic Noise Generation and Prediction,” Proceedings of the National Noise and Vibration Control Conference,Chicago, IL: INCE, 1980.

Ng, K. W., “Control Valve Noise,” ISA Transactions 33, 1994, pp. 275–286.Page, G. W. Jr., “Predict Control Valve Noise–A Graphical Method for Liquid-

Flow Systems Eases the Task,” Chemical Engineering, July 1997.Rahmeyer, W., “The Critical Flow Limit and Pressure Recovery Factor for

Flow Control,” InTech, November 1986.Reed, C. M., “Optimizing Valve Jet Size and Spacing Reduces Valve Noise,”

Control Engineering, September 1976.Reethof, G., “Some Recent Developments in Valve Noise Prediction Meth-

ods,” ASME Transactions of Winter Annual Meeting, November 1982.Reethof, G. and Ward, W. C., “A Theoretically Based Valve Noise Prediction

Method for Compressible Fluids,” Transactions of the ASME — Journal of Vibration, Acoustics, Stress, and Reliability in Design, Vol.

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Valves: Noise Calculation, Prediction, and Reduction

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