Noise from Steam Valves in Power Plants – Calculation and Practice
by Dipl.-Phys. Stephan Heim
Vo lu me 90, Is sue 5/2010, Pa ge 72 to 77
International Journal for Electricity and Heat Generation
Valve Noise in Power Plants
VGB PowerTech 5/2010 3
Noise from Steam Valves in Power Plants – Calculation and Practice Stephan Heim
Author
Dipl.-Phys. Stephan HeimMüller-BBM GmbH Planegg/München/Germany
Kurzfassung
Lärm in Kraftwerken durch Dampfventile – Berechnung und
Praxis
Ventile zur Druck- und Mengenregelung von Wasserdampf werden an vielen Stellen in Kraftwerken eingesetzt. Anhand der schall-technisch oft kritischen Dampfumformstatio-nen wird beispielhaft gezeigt, wo die Ursachen liegen können, wenn in Planung angestrebte Schallgrenzwerte in der Praxis überschritten werden.
Dazu werden zunächst die grundsätzlichen Geräuschentstehungs- und Schallübertra-gungsmechanismen bei gas- und dampfdurch-strömten Ventilen vorgestellt und die aktuellen Richtlinien zur Berechnung von Ventilge-räuschen diskutiert. Die Einhaltung schalltech-nischer Vorgaben für die betriebsfertige Arma-tur im Kraftwerk hängt nicht alleine von den akustischen Eigenschaften des Ventils ab. Sie ist vielmehr das Ergebnis der Abstimmung und Zusammenarbeit zwischen Anlagenbauer, Ar-maturenhersteller, Rohrleitungsplaner und Iso-lierer, also einer Gewerke übergreifenden schalltechnischen Planung.
Abschließend wird ein Messverfahren unter Verwendung von Laservibrometern vorgestellt, mit denen berührungslos Schwingungen von heißen Dampfleitungen messtechnisch erfasst werden können. Die Ergebnisse können so-wohl zum schalltechnischen Garantienachweis als auch der Lösung von luftschall- und schwingungstechnischen Problemen im Zu-sammenhang mit Ventilen dienen.
Introduction
Valves for control of steam pressure and flow are used in various applications in power plants. From an acoustics point-of-view, steam conditioning valves are of particular impor-tance. These valves serve, for example, as tur-bine bypass valves or for conditioning of process steam. High pressure drop and flow rate generate high levels of noise emission which can pose a particular challenge in view of rather low maximum permissible sound levels that are often explicitly asked for in contractual specifications.
Existing guidelines for valve noise calculation can be helpful in the planning phase, but can-not take all site and application specific pa-rameters into account that have an effect on the actual noise emissions of valves in opera-tion at the power plant. As a result, the pre-dicted theoretical emission levels may be lower than the real noise levels encountered so that noise limits are not complied with. Be-fore looking at possible reasons for this, some basics of the noise generation mechanisms and sound propagation paths in steam valves
are explained and an overview of calculation methods in current guidelines is given.
Noise Generation in Valves and Sound Propagation Paths
Steam conditioning valves reduce steam from a high pressure level to a lower pressure level and cool it down by the injection of cooling water.
F i g u r e 1 , left hand side, shows a steam conditioning valve, with parts of valve body removed. The steam enters the valve at the in-let, which, in the figure, is covered by the body. Inside the valve body, the steam flow turns through 90 ° and leaves it via the outlet at the bottom of the figure. The steam cooling system (desuperheater) is located in the valve outlet area, which will not be further dis-cussed. The steam flow rate depends on the degree of opening of the plug by the stem, which is moved in vertical direction by the valve actuator. The plug – a cage with drilled holes in this case – slides up and down the seat, thus uncovering a greater or smaller number of the throttling holes. The yoke con-
Actuator
Yoke
Valve body
Stem
Plug
Inlet
Seat
Outlet
Figure 1. Example of a low-noise steam conditioning valve with seven pressure reducing stages (four controlled and three non-controlled stages) and cooling system at the outlet (figure and drawing: with permission of Welland & Tuxhorn AG, labels and flow paths by the author).
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nects the valve actuator with the body. Further details of the valve are shown in the sectional drawing on the right hand side of Figure 1. The blue line illustrates the flow through the valve. The pressure reduction in the valve takes place across seven stages. Four of these pressure-reducing stages are controlled (green coloured) and three are non-controlled (red coloured). The cross sectional flow area of a controlled stage depends on the position of the plug, whereas the area of a non-controlled stage is independent from the stem/plug posi-tion. The left hand half of the drawing in Fig-ure 1 shows the valve closed; the right hand side fully open.
Similar to other sources of flow-induced noise, the noise caused by valves is generated mainly in those areas with a high flow veloci-ty. To outline the associated noise generation mechanisms, for reasons of simplicity, a valve with a single pressure-reducing stage, as shown in F i g u r e 2 , is taken as an example in the following. At the pressure-reducing stage where the cross sectional area of the flow has its minimum (the so-called vena con-tracta), the lowest pressure occurs and thus the highest flow velocity. Noise generation mainly takes place in the flow field down-stream the vena contracta, by
vortex shedding and interaction of turbulent −flow with internal parts of the valve,
turbulent jets, −
shock waves and shock-turbulence interac- −tion when the flow becomes critical, i.e. velocity in the valve reaches or exceeds the speed of sound (sonic flow, Mach number > = 1).
These mechanisms of noise generation by aerodynamic processes can, in principle, be taken into account, with sufficient accuracy, in valve noise calculations although signifi-cant simplifications have to be made in mod-elling these effects. However, in many cases
additional flow-induced tonal noise – which is particularly annoying to the human ear or can lead to valve or piping damage – may, for ex-ample, be caused by
resonant excitation of valve cavities caused −by the flow (happens when the frequency generated by turbulence at a pressure-re-ducing stage acoustically matches resonant frequencies of cavities inside the valve or the piping),
flow over and into valve cavities (excitation −of acoustic modes of the cavity by the un-stable shear layer across the cavity open-ing),
pulsations in the flow due to resonant vibra- −tions of the plug-stem system.
This additional flow-induced noise, in gener-al, is unpredictable. Only the probability of these effects occurring can be predicted by a valve noise expert.
In multi-stage valves (Figure 1, for example) the situation is considerably more complicat-ed as noise is generated at each individual pressure-reducing stage and is attenuated again by the subsequent stages ( F i g u r e 3 ). The total noise downstream of a multi-stage valve is the sum of the noise contribution from each stage. However, in spite of this con-tribution of many stages, the total noise is typically lower than for a corresponding sin-gle stage valve.
Although having its origin in the pressure stages, the noise generated by the valve is mostly radiated via the connected pipes. The pipe wall is mainly excited by the acoustic field in the pipe and, to a much lesser extent, by the turbulent flow field behind the valve exit.
For the high-pressure drop in steam condition-ing valves (typically much greater than 2 : 1), the noise mostly propagates downstream and only a smaller portion of it travels upstream
Stem
Actuator
Yoke
Plug
Acousticfield
Acousticfield
Turbulentfield
Pressure
Flow speed
P1
P2
P3
V3
V2V1
Figure 2. Typical static pressure and flow velocity distribution in a single-stage valve.
Stage 1 Stage 2 Stage n
LWi,1 LWi,1-∆L2 LWi,1-∆L2....-∆Ln
LWi,2....-∆Ln
LWi,n
LWi,total
LWi,2
Figure 3. Noise generation at multi-stage pressure reduction.
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and is radiated to the outside via the upstream piping. Some sound is also radiated by the valve body and from connected construc-tion parts such as the stem, the yoke and the valve actuator, although this portion is clearly lower.
An index for the aerodynamic noise radiated from the valve into the pipe is the internal sound power level LWi, which depends on
the valve construction (for example: the −lower the highest flow velocity in the valve due to a proper design of the pressure re-
ducing stages, the lower the noise genera-tion),
the flow rate (the higher the flow rate, the −higher the noise generation),
the speed of sound of the flow medium (the −higher the speed of sound in the flow me-dium, the higher the internal sound power level) and
the pressure up- and downstream of the re- −ducing stage (the higher the pressure differ-ence across the stage, the higher the noise generation).
The frequency spectrum of the noise radiated into the pipe is a broadband spectrum with a maximum at the so-called peak frequency fpeak ( F i g u r e 4 , top). This peak frequency depends on the flow velocity and the jet diam-eter which is given by the characteristic di-mension of the smallest cross sectional flow area, e.g., for perforated discs or cages, the port diameter. The higher the flow velocity and the smaller the jet diameter, the higher the peak frequency.
The sound level outside of the valve or pipe – the external sound level – depends on how much noise is transmitted through the walls of valve body and the pipes. Because of the rela-tively thin walls of the pipes compared to those of the valve body the transmission loss of the pipes is the most important factor. The external sound power level of the noise radi-ated from the pipe is a result of the difference between the internal sound power level and the transmission loss of the pipe and some frequency-independent corrections (Figure 4, bottom). The transmission loss of the pipe de-pends very strongly on the frequency and is very difficult to calculate accurately. For brev-ity, only some basics are discussed in the fol-lowing1.
One fundamental minimum of the transmis-sion loss versus frequency occurs at the so-called ring frequency fring (Figure 4, centre). At this frequency, the longitudinal wavelength
1 In particular, the so-called coincidence effect is neglected here, which results in further local minima of the transmission loss of the pipe.
in the pipe wall is equal to the pipe perimeter. For many steam-conditioning valves in power plants, typical peak frequencies of the internal sound power level are in the range of or above the ring frequencies of the connected pipes.
The frequency-dependence of the internal sound power and the characteristics of the transmission loss of the pipe result in a possi-bility to reduce the external noise. By splitting the flow at the pressure reducing stage into smaller independent jets (e.g. change from parabolic plug to perforated cage plug or use of smaller drillings in perforated cages or disks) the peak frequency of the internal noise is shifted to higher frequencies. If the shifted peak frequency is well above the ring frequen-cy the transmission loss of the pipe is higher and, as a result, the external sound power is lower.
Ta b l e 1 gives typical values for the contri-butions of a steam conditioning valve’s sound transmission paths (body, pipes, yoke etc.) to the total A-weighted external sound power level and the sound pressure level at 1 m dis-tance from the valve (assuming free-field conditions). Values are calculated with Müller-BBM proprietary software and have been confirmed by field measurements. The valve in this example has six pressure-reduc-ing stages. The nominal diameter of the pipe at the valve inlet is DN 250 and at the outlet DN 800. The flow rate is 150 t/h. The steam parameters are 500 °C, 80 bar absolute at the inlet and 160 °C, 5 bar absolute at the outlet. A length of 50 m each for the pipes at the inlet and outlet is assumed. The A-weighted sound power level of the noise radiated into the pipe by the valve (internal sound power level) in this case is LWi = 145 dB(A).
Sound power levels characterise the radiated sound energies. They are independent of the local installation situation of the valve. In con-trast to this, sound pressure levels describe the effect at a specified distance from the sound source. Among other things, they depend on the distance from the sound source and the lo-cal conditions (free field conditions vs. sound reflections, diffuse sound fields etc.).
Inte
rnal
sou
nd p
ower
leve
lE
xter
nal s
ound
pow
er le
vel
Pip
e tr
ansm
issi
on lo
ss
Frequency
Frequency
Frequency
f peak
f ring
Figure 4. Top: sound power level spectrum of the noise radiated into the pipe at the valve outlet. Centre: typical pipe transmission loss Bottom: resulting external sound power level spectrum Table 1. Typical values for the contributions of a steam conditioning valve‘s sound transmission
paths (body, pipes, yoke etc.) to the total A-weighted external sound power level and the sound pressure level at 1 m distance from the valve. From calculations for a steam conditioning valve with six pressure reducing stages.
Sound transmission path (radiation to outside)
Sound power level [dB(A)]
without insulation
Sound pressure level at 1 m [dB(A)] without insulation
Sound pressure level at 1 m [dB(A)]
with insulation
Pipe at valve outlet 132 105 76
Pipe at valve inlet 122 97 69
Valve body 110 94 73
Yoke, stem and actuator
105 89 89
Total 132 106 89
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The total levels given in the last row of Table 1 are the result of the energetic addition of the contributions from the individual transmission paths in the rows above. It is clearly seen that the total levels are dominated by the noise contribution from the pipe at the valve outlet for a valve and pipes without insulation (“with-out insulation”). If valve and connected pipes are thermally or acoustically insulated, the sit-uation changes. This will be discussed later.
Valve Noise Calculation Guidelines
There are a number of guidelines on noise calculation for gas and steam valves that pro-vide calculation methods to predict the A-weighted sound pressure level 1 m down-stream from the valve and at 1 m distance from the pipe wall, assuming acoustic free field conditions. The methods only consider noise which is generated by aerodynamic processes. Level reductions from additional noise control measures like acoustic insula-tions are not taken into account.
The most important current guidelines are:
VDMA 24422: Armaturen. Richtlinien für −die Geräuschberechnung Regel- und Ab-sperrarmaturen [1].
IEC 60534-8-3: Industrial process control −valves. Part 8-3: Noise considerations con-trol valve, aerodynamic noise prediction method [2]2.
ANSI/ISA-S75.17: Control valve aerody- −namic noise prediction [3].
The theory behind the individual calculation methods in the guidelines or the methods themselves are outside of the scope of this ar-ticle, however, the basic calculation steps in-volved are:
Calculation of the mechanical stream power −of the valve on the basis of flow data.
Determination of the acoustic efficiency, −i.e. the ratio between the mechanical stream power which is converted into internal sound power and the total mechanical stream power.
Calculation of the sound power generated −at the pipe outlet (internal sound power level LWi), from the mechanical stream power and the acoustic efficiency.
Calculation of the sound power level L − Wa of the noise radiated to the outside (external sound power) from a pipe with a length of 2 m at the valve outlet. For this the internal sound power level LWi and the transmission loss of the pipe is used.
Conversion of the external sound power −level into the A-weighted sound pressure level at 1 m distance LA,1 m.
2 IEC 60534-8-3 is being revised at the moment. Edition 3 is planned to be completed by the end of 2012.
Calculations according to VDMA 24422 are carried out fully frequency-dependent for the octave bands with centre frequencies of 500 Hz up to 8,000 Hz. The acoustic efficien-cy is calculated using one single equation for the whole flow range with the Mach number as a main parameter. The gradient of the acoustic efficiency versus Mach number curve is controlled by two valve specific exponents.
According to IEC 60534-8-3, the acoustic ef-ficiency is calculated in a comparatively so-phisticated way taking five flow regimes into account. The most important valve specific parameters are the flow coefficient, the valve style modifier and the liquid pressure recov-ery factor. However, the acoustic calculation is carried out in a simplified way using over-all levels, without considering the spectral distribution of the internal sound level or the pipe transmission loss in detail3. To a certain degree the “frequency dependency” is taken into account indirectly via a comparison of the peak frequency of the internal sound level, e.g. with the ring frequency of the pipe.
The calculation method according to ANSI/ISA-S75.17 is similar to that of the IEC 60534-8-3, but slightly simpler.
The outlined calculation methods detailed in the guidelines provide fairly good estimates for the noise emissions of standard valves un-der well defined standard conditions. Howev-er, they have a number of shortcomings de-pending on the specific valve type in question, especially with respect to noise predictions for valves under on-site conditions. For exam-ple, with the multi-stage steam conditioning valves with drilled cages or plates, as in Figure 1, the following problems may arise:
Calculation for such valves is usually only carried out for the last (most downstream) pressure reducing stage. If there is a signifi-cant noise contribution from the upstream stages (see also Figure 3), the predicted inter-nal sound power level is invariably too low. In particular, this happens for valves with non-controlled stages at operating conditions be-low maximum load.
In some cases the IEC 60534-8-3 method overestimates the frequency shifting effect of small port diameters for steam conditioning valves with drilled cages or plates as pressure reducing stages and underpredicts the sound pressure levels at 1 m distance from the valve. This happens if the dimension that is acousti-cally relevant for the peak frequency is greater than the drilling diameter due to jet blasting or jet interaction.
In summary, the calculation tools from the mentioned guidelines are useful to compare different valve concepts and the valves of dif-ferent manufacturers from an acoustical point-
3 In revision 3 (Ed .3) the introduction of a frequen-cy-dependent calculation method is planned.
of-view. However, as the calculation methods are based on flow characteristics and thermo-dynamic processes of normal standard valves only, they usually cannot provide reliable re-sults for special valves in which the actual processes are far more complicated.
Acoustic Insulation for Pipes an Valves
The guidelines discussed in the previous sec-tions provide the sound pressure level at 1 m distance of a non-insulated valve/pipe under free field conditions. However, this value does not describe the acoustic situation in a power plant environment properly and cannot be di-rectly verified by measurements on site as the above conditions are not fulfilled: Steam valves and piping in power plants are usually located in power or boiler houses (non-free field conditions) and, typically, are provided with a thermal insulation.
As thermal insulation also reduces the noise radiated from a valve or pipe, it can be suita-bly upgraded to become an efficient acoustic insulation.
Acoustic insulations of pipes and valve bodies usually consist of an outer cladding, which encloses the pipe/valve and a layer of mineral fibre which fills the gap between the pipe/valve and the outer cladding. Depending on the specific acoustic requirements, spacers between the pipe and the outer cladding, if any, are designed more or less soft elastic. For high-performance acoustic insulations, the outer cladding is also equipped with an addi-tional damping layer on the inside. As acous-tic insulations provide a high insertion loss at higher frequencies they are an effective noise control measure for valves and the connected piping which, typically, are characterised by high frequency noise emissions.
Acoustic insulation of pipes and valves can only be effective if applied properly and if the involved interfaces are properly taken into account. As such, they require a good co- operation and co-ordination between the par-ties involved, e.g. plant manufacturer, valve supplier, pipe manufacturer, insulation con-tractor etc.
Today, ISO 15665 of 2003 “Acoustic insula-tion for pipes, valves and flanges” [5] is the most important standard with respect to acous-tic insulations of pipes. It defines and de-scribes three classes of acoustic insulation sys-tems (classes A, B and C) with different fre-quency-dependent minimum insertion losses and gives basic details on their design. As an example, Ta b l e 2 lists the minimum inser-tion losses as specified in [5] for the most ef-fective insulation class C. It is seen that a dis-tinction is made between three different ranges of nominal pipe diameters because the achiev-
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Table 2. Minimum insertion losses for the acoustic insulation systems class C according to [5].
Class Range of nominal pipe
diameter D [mm]
Minimum insertion loss [dB]
Octave band centre frequency [Hz]
125 250 500 1000 2000 4000 8000
C1 D < 300 –5 –1 11 23 34 38 42
C2 300 ≤ D < 650 –7 4 14 24 34 38 42
C3 650 ≤ D < 1000 1 9 17 26 34 38 42
able insertion loss of a specific insulation is also dependant upon the diameter of the pipe to be insulated.
The insertion losses given in [5] can only be achieved if the insulation is installed accord-ing to the specifications in the standard and when all relevant details are taken into ac-count. Examples for such details are given in the following.
In many plants, pipe supports can be found which are rigidly connected with the pipe and which penetrate the insulation ( F i g u r e 5 , right hand side picture). In doing so, a sec-ondary sound transmission path or “sound bridge” is created: The supports transmit structure-borne sound from the pipe to the outside of the insulation cladding, where it is radiated from the supports themselves, the supporting steelwork or the cladding (if in contact with the supports) in the form of sec-ondary airborne sound. As a result, a large part of the sound energy in the pipe wall ef-fectively “by-passes” the insulation. The effect is especially pronounced for high frequency noise, as typically generated in a valve, which is radiated well from nearly any structural part. Similar effects can occur with connected safety valves and the yoke of the shut-off valves in steam pipes, for example (Figure 5, left hand side picture).
Insufficient sealing of the outer cladding at the penetrations of supports, sensors, tap points etc. can create “noise leaks” through which airborne sound from the pipe is emitted into the open. In the case of stringent noise requirements, such leaks can weaken the ef-fectiveness of an acoustic insulation to more than the permissible extent.
However, the secondary sound transmission path that is underrated most frequently is the yoke, stem and actuator assembly of the steam control valve itself which often remain unin-sulated ( F i g u r e 6 ). In this case the struc-ture-borne sound is transmitted from the valve directly and then radiated into the open as air-borne noise. The detrimental effect of this phenomenon can be seen in the right column of Table 1. It shows the contributions of a steam conditioning valve’s sound transmis-sion paths (body, pipes, yoke etc.) to the total A-weighted external sound pressure level at 1 m distance from the valve for a high-quality acoustic insulation of class C according to [5] (“with insulation”). The noise from the pipe at the valve inlet and outlet and from the valve body is sufficiently reduced by the acoustic insulation. As a result, only the yoke, stem and actuator assembly protruding the insulation determines the total level. An ap-propriate noise control measure is to acousti-cally enclose yoke, stem and actuator assem-bly - which may involve adequate ventilation measures for cooling if sensitive electrical or hydraulic components are inside the enclo-sure.
In summary, acoustic insulations are only fully effective if all acoustically weak points and secondary sound transmission paths are avoided. Because of the numerous restric-tions, boundary conditions and interfaces as-sociated with typical piping, this requires that the insulations are planned together with the piping designer already at a very early stage.
Valve Noise and Structure-borne Sound
The examples on secondary sound transmis-sion paths or “sound bridges” in the previous section show that structure-borne sound plays a decisive role with respect to valve noise. In general, with the exception of steam discharg-ing into the atmosphere, the noise generated in a valve is, ultimately, emitted by the valve components and the connected pipes as so-called secondary airborne noise caused by structure-borne sound-induced surface vibra-tions. The sound power level of secondary air-borne noise LWa radiated from a surface can be determined from the vibration velocity level of the surface Lv (usually related to a reference value of 5 . 10–8 m/s), the area of the radiating surface S and the so-called radiation efficiency factor σ:
LWa = Lv + 10 lg (S) + 10 lg (σ) dB (1)
The frequency-dependent radiation efficiency factor describes the portion of structural vi-bration energy, which is converted into air-borne sound (0 < σ ≤ 1). Equations for the calculation of radiation efficiency factors for pipes can be found in literature (e.g. [6]). Generally, σ increases with frequency.
Structure-borne sound is generated by the acoustic field inside the pipe (acoustically in-duced vibration) and by alternating forces, which are created during pressure reduction at the individual stages in the valve. In some applications it is possible to reduce secondary noise emissions from structure-borne sound by isolating the equipment in which the alter-nating forces are generated from the remain-ing parts of the system. An example are high-performance blow-off silencers as they are
Figure 5. Secondary sound transmission paths. Left hand figure: Safety valve (left) and yoke of shut-off valve (right). Right hand figure: pipe support.
Figure 6. Secondary sound transmission path “yoke, stem and actuator assembly”.
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used in special situations to reduce noise from venting large amounts of pressurised steam directly into the atmosphere. To avoid the transfer of structure-borne sound from the si-lencer’s pressure relief stages to the silencer casing and the attached piping, the stages are acoustically decoupled from the rest of the system ( F i g u r e 7 ).
The above design solution of decoupling the pressure-reducing stages cannot be applied to valves directly but comparable principles of reducing the transmission of structure-borne
To atmosphere
Absorptionsilencersystem
Casing
Pressurereliefstages
Decoupling
Inlet pipe
Figure 7. Blow-off silencer with structure-borne sound decoupling (with per-mission of BBM Akustik Technologie GmbH).
sound have successfully been applied in a number of supplier-specif ic solutions of acoustically optimised valves.
Measuring Structure-borne Sound
An important practical tool for assessing structure-borne sound from valves and pipes on site are measurements of surface vibrations with accelerometers. The measured vibration acceleration is converted into velocity by inte-gration, from the velocity levels the second-ary airborne noise LWa radiated from a surface can be determined according to (1). In addi-tion, an indication about the internal sound power level LWi can be obtained. The results of such measurements can be important if compliance with guaranteed values needs to be demonstrated and as a basis for remedial actions with respect to airborne sound and vi-brations (proper design and dimensioning of
insulation etc.). The method is also applicable in cases where direct airborne noise measure-ments are not possible because of high back-ground noise.
For vibration measurements on site, acceler-ometers may not always be applicable as they are usually mounted via adapter plates that are glued on to the vibrating surface. This presents a problem with steam pipes because of the high surface temperatures. In this case laser vibrometers – which measure the vibra-tion velocity without contact to the pipe – are an interesting alternative. Such a measure-ment is shown in the photo in F i g u r e 8 .
References
[1] VDMA 24422: Armaturen. Richtlinien für die Geräuschberechnung Regel- und Absperrarma-turen. January 1989.
[2] IEC 60534-8-3: Industrial process control valves. Part 8-3: Noise considerations control valve, aerodynamic noise prediction method. July 2000.
[3] ANSI/ISA-S75.17: Control valve aerodynamic noise prediction. 1989.
[4] Borden G., and Friedmann P. (Editors): Con-trol valves, practical guides for measurement and control, ISA, 1998
[5] ISO 15665: Acoustics – Acoustic insulation for pipes, valves and flanges. First edition 2003-08-15.
[6] Müller G., and Möser M. (Editors): “Taschen-buch der Technischen Akustik”, 3. Auflage, Springer Verlag, Berlin, (2004). ∙
Figure 8. Vibration measurement on a steam pipe with a laser vibrometer (on left side). On the right side the red laser dot on the pipe with the insulation removed can be seen.
