Proceedings of ASME Turbo Expo 2016: Turbomachinery Technical Conference and Exposition

GT2016
June 13 – 17, 2016, Seoul, South Korea

GT2016-57800

ACOUSTICALLY INDUCED VIBRATION MITIGATIONS IN COMPRESSOR PIPING

SYSTEMS

Timothy C. Allison, Ph.D. Jeffrey Bennett

Southwest Research Institute Southwest Research Institute
San Antonio, TX, USA San Antonio, TX, USA

ABSTRACT
Acoustically induced vibration (AIV) is a high-frequency NOMENCLATURE
vibration phenomenon that can occur downstream of pressure-
reducing devices such as control valves, restriction orifices, and Symbols
pressure relief or safety valves in compressor piping systems.
These vibrations can lead to high cycle fatigue failures of a Internal pipe radius [m]
downstream piping at side branches or welded supports. c Sound speed [m/s]
Existing methods for screening and analyzing acoustically
Di Internal pipe diameter [m]
induced vibration are not well-grounded in the underlying
physics and thus do not provide a methodology for evaluating a FL Liquid pressure recovery factor (from IEC) [-]
variety of mitigation strategies. Modeling of acoustically LP Sound pressure level [dB ref. 20µPa]
induced vibration is computationally challenging, as it requires LW Sound power level [dB ref. 1pW]
the interaction between tens or hundreds of higher-order
acoustic modes with a similar number of piping shell modes. P1 Pressure upstream of CV [MPa]
In order to obtain better insight into the underlying physics P2 Pressure between CV/PSV [MPa]
of AIV and to characterize the effectiveness of several P3 Pressure downstream of PSV [MPa]
mitigation methods, full-scale blowdown testing was performed
at Southwest Research Institute. Tests were performed using W Sound power [W]
20 MPa nitrogen gas vented at 28 kg/s through a 3x4” pressure ρ Density [kg/m3]
safety valve and multiple header pipe sizes ranging from 12” to
36”. Test configurations included baseline piping geometry at
each size and several AIV mitigations including stiffening INTRODUCTION
rings, viscous damping wrap, and internal acoustic mode Acoustically-induced vibration (AIV) refers to
disruptors. Test results from strain gauges, accelerometers, and high-frequency pipe wall vibration that is excited by an
dynamic pressure transducers show a broadband multimodal acoustic source. Typically, the source of this acoustic excitation
response with dynamic stresses up to 3 kHz near the safety is an upstream flow restriction such as a control valve, pressure
valve tailpipe connection to the test header, and various safety valve (PSV), or restriction orifice that passes a high mass
mitigations reduced dynamic stresses by 8-52% depending on flow of process fluid with a large pressure drop. In some cases,
the piping and type of mitigation. particularly with thin walled, large diameter pipe, the resulting
Acoustic and structural finite element models were vibration levels are severe enough to cause high-cycle fatigue
analyzed in order to determine the coincident modes that match damage and failure, usually at welded supports or branch
in both axial/circumferential shape and natural frequency and connections. Due to the high response frequencies (typically
compare coincident frequencies with measure stresses. The 100 - 3,000+ Hz), fatigue failures in some severe cases have
results show that observed peak stress frequencies do not been reported to occur within as little as several minutes of
generally correlate well with predicted coincident modes, and operation. The severity of AIV and likelihood of AIV-induced
that flow-induced turbulence excites frequencies below piping failures are complex physical phenomena that depend on
shell modes that can also result in significant stresses that multiple factors including an accurate characterization of the
combine with AIV. noise source(s), internal multimodal piping acoustics and

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 acoustic loading of the pipe wall, multimodal shell mode methodology, selected results, and initial attempts at comparing
response of piping, and fatigue behavior including weld effects physics-based models with experimental data.
and static loading of the piping2,3.
Due to the complexity and uncertainty in physics-based METHODOLOGY
modeling approaches, typical practice for AIV analysis and Three 20’ long test headers were constructed using a
mitigation generally involves a simplified screening procedure common tailpipe configuration fed by a control valve and PSV
such as the methodology presented by the Energy Institute (EI) in series. Nitrogen is pumped, vaporized, and stored in a
4
, Southwest Research Institute® (SwRI®)5 and in collaboration 3,000 psi (20 MPa) reservoir, which is then vented to
with Fluor6, and CSTI7. These screenings are based on the atmosphere at a rate of 70 MMSCFD (28 kg/s) for a period of
seminal papers by Carucci and Mueller8 and Eisinger9, which at least 10 seconds during each of a series of 12 tests. A
present the majority of AIV failure data available in the diagram of the test configuration is shown in Figure 2; the three
literature today. All of these screening methods involve a carbon steel test headers are shown in Figure 3 (top, 12” sch.
simplified characterization of source noise, attenuation of noise 80s; middle, 20” sch. 40s with damping wrap; bottom, 36” sch.
levels to downstream connections, and a geometry parameter 40s with pipe clamps); these sizes and schedules were overlap
used to characterize each connection (e.g. pipe diameter or with and extend the low-end range of diameter:thickness ratios
diameter:thickness ratio). Data can be plotted and compared for which data exist. Internal pipe dynamic pressure
with failure/non-failure data (e.g. see Figure 1) and new designs measurement locations are shown in Figure 4, where three rings
are required to fit below a failure criterion line that separates of four taps (rings A, B, and C) were located on the tailpipe,
failures from non-failures. and four taps were located in each header. Six strain gauge
rosettes were applied to the welded tailpipe connection on each
180 Historical Failure/Non-failure Data header in two orthogonal rows of three rosettes each (Figure 5).
C-M Failure Accelerometers were located at positions A1 – A4 indicated in
C-M Non-failure
175 Figure 3 (center panel).
Four test configurations were run on each header:
170 (1) untreated to establish a baseline case; (2) a pair of pipe
clamps to act as stiffening rings added on either side of the
PWL (dB)

165 welded branch connection; (3) adhesive viscous damping wrap,

a length of two diameters applied on either side of the welded
160
branch connection; and (4) two tube bundles inserted into the
header, one bundle on each side of the welded branch
155
connection, each bundle two header diameters long. The tube
150
bundle test case data for the 12” header were contaminated by
excessive vibration of the tube bundle and are omitted from the
145 analysis presented here.
20 30 40 50 60 70 80 90 100 110 120
D/T

Figure 1. Example Screening with AIV Historical Data5

Few existing approaches are based on modeling the
physics of AIV in significant detail, including consideration of
the internal fluid acoustic modes and pipe shell modes.
Proposed approaches that are physics-based, such as those
presented by Skailes1, Izuchi10, and Ghosh11 do not present test
data for validation. Although experiential methods may be
sufficient for design screenings for new installations, detailed
modeling is required for implementation of AIV mitigations on Figure 2. Experimental Setup
existing systems where piping replacement may not be an
Pipe clamps placed on either side of the welded branch
option. A better understanding of AIV mitigations may also be
connection act as stiffening rings, restricting motion of the pipe
applied to reduce the cost and weight of new installations, if
wall and increasing the resonant frequencies of the header pipe.
AIV considerations are driving the design.
A finite element (FE) model utilizing shell elements was used
Recent testing at SwRI seeks to expand the understanding
to determine the optimal width and spacing of the pipe clamps,
of the science of AIV, including noise generation and
which were equal to the header pipe thickness in each case. The
propagation, the dynamic strain response at a stress
maximum stress at the header-branch connection based on an
concentration, and the effectiveness of several AIV mitigation
analysis of the model’s natural frequencies from 200 - 2,000 Hz
strategies. This paper presents the test hardware and
was compared to a baseline case for several stiffening ring
configurations on each header geometry. The model was driven

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 by a circumferential forced unit load excitation at one end of The tube bundles were intended to break up internal pipe
the pipe with a small gap to ensure that all modes would be acoustic modes to potentially reduce coincidences with pipe
excited, and to provide repeatability between different runs. vibrational shell modes. Tube bundles consisted of 19 tubes
each of diameter 2.25” (12” header), 3.75” (20” header), and 7”
(36” header). Each tube bundle was two header diameters long,
and the pair of bundles was centered around the PSV tailpipe
inlet, with a 6” gap between the inner end of each tube bundle,
and the edge of the tailpipe weldolet. One of the 36” tube
bundles is shown in Figure 6.
All results were analyzed from a 10-second record of
approximately uniform flow from each test. Representative
normalized flow and pressure profiles are shown in Figure 7,
where the analysis period ranged from 30 – 40 seconds.

Figure 4. Internal Noise Measurement Locations

Figure 3. Test Header Configurations

Figure 5. Strain Gauge Placement
Polydamp® Acoustical Barrier with a weight of 1.0 lb/ft2
(PAB-010) was used to provide damping. This adhesive wrap
was selected over several other damping materials after
comparative modal tests showed that it had the highest damping
ratio. Damping wrap was applied around the entire header pipe,
starting at the tailpipe weld and extending two header diameters
in either direction. This dimension was chosen based on
material requirements and the desire to maintain a length of
undamped pipe on either side of the treatment to provide a
realistic vibration response of the total system.

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 Figure 8. Strain Time Series (12” Header Baseline)
Figure 6. 36” Tube Bundle

Figure 9. Strain Spectrum (12” Header Baseline)

Figure 7. Normalized Pressure and Flow The root main square (RMS) strain from 100 – 3,000 Hz
from the center gauge of rosette 4, which showed the highest
levels for all tests, was used to compare mitigations. Results are
MEASUREMENT RESULTS I: STRAIN shown in Table 1 as a percent reduction from the baseline cases
Measurements of the dynamic strain show the highest (1a, 2a, and 3a).
levels at rosette 4, along the circumference of the pipe. All mitigations provided some reduction in measured
Representative normalized strain time series and spectra from strain, with pipe clamps and tube bundles producing the
the 12” header measurements are shown in Figure 8 and Figure greatest effect. Data from the 12” header test were
9, where the three traces correspond to the three gauges contaminated by excessive vibration of the tube bundle and are
constituting a rosette. The highest levels are observed in the omitted from the analysis presented here. While the pipe
center gauge (green); the drift in offset is due to the cooling of clamps appear to provide the greatest potential reduction in
the pipe during testing caused by the isentropic expansion of measured dynamic strain, it is important to note that stiffening
nitrogen gas. The response of the gauges is highly-frequency ring and acoustic mode disruptors are highly geometry-
dependent corresponding to the header modes of vibration, with dependent and need to be engineered for optimization to each
peak response occurring at 95 Hz, 112 Hz, and 715 Hz for the specific application.
36”, 20”, and 12” header baseline tests, respectively.

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 Table 1. Mitigation Results 1,000 Hz between rings B and C, which may be due to
turbulence generation in the tailpipe elbow.
Test Configuration % Reduction
1a 36” baseline -
1b 36” clamps 51.8%
1c 36” damping wrap 11.1%
1d 36” tube bundles 30.1%
2a 20” baseline -
2b 20” clamps 17.8%
2c 20” damping wrap 17.0%
2d 20” tube bundles 28.2%
3a 12” baseline -
3b 12” clamps 15.8% Figure 10. Dynamic Pressure Spectra at
3c 12” damping wrap 8.4% Rings A, B, and C

The width, thickness, and spacing of the stiffening rings all The IEC predicted noise for the series combination of the
affect the vibrational response of the header pipe and the control valve and PSV is 192 dB assuming a conservative
resulting dynamic strain at the tailpipe weld. Similarly, the liquid pressure recovery factor FL of 1 for both valves. A more
geometry of the tube bundles, while reducing low-frequency realistic FL estimate of 0.6 for the PSV produces a predicted
acoustic modes in the header, will set up additional high level of 187 dB.
frequency modes that may have not been present in the baseline The measured and predicted noise levels for several tests
case. Introducing tubes into the mean flow of the header will are shown in Table 2, where IEC (a) is computed with a PSV FL
also result in additional flow restriction and increased of 0.6, and IEC (b) is computed with a PSV FL of 1.0. The EI
turbulence generation. calculation is shown both in terms of power level, as the
standard produces, and converted to pressure by the method
MEASUREMENT RESULTS II: DYNAMIC PRESSURE described by Evans12, shown in Equation 2. The 6 dB sonic
Physical testing was performed at the valve manufacturer’s flow factor was included for the PSV for all EI calculations.
blowdown test rig in an attempt to replicate the scenario that The noise is observed to track both the flow and pressure ratio
led to valve d. An overview of several valve noise predictive as shown in Figure 7.
methods and measurement results from a test configuration
1
similar to this study have been presented by Evans 12. The 𝐿𝑊,𝑝𝑖𝑝𝑒 = 𝐿𝑃 − 10 log10 ( ) (2)
𝜋𝑎2
predicted valve noise is best described by International
Electrotechnical Commission (IEC) standard13, based on Comparing noise measurements from different locations
several parameters which define a noise generation regime and provides an assessment of the degree of decay with distance.
associated acoustic efficiency. The standard provides a sound The distance between ring A and B was 100” for all tests; the
power W which can be used to produce the predicted sound total straight line distance between ring B and C varied
pressure level as shown in Equation 1, where 𝜌 and 𝑐 are depending on the header pipe diameter, where the horizontal
evaluated downstream of the valve: distance from ring B to the center line of the elbow was 31” for
all tests. The distance from the center line of the elbow to
3.2𝑥109 𝑊𝜌𝑐
𝐿𝑃 = 10 log10 ( ) (1) ring C was 16”, 32”, and 41.5” for the 36”, 20”, and 12” header
𝐷𝑖2
tests, respectively, so measurements between rings B and C
compare results from non-uniform spacing.
The average RMS pressure level measured at the ring A
locations (12.75” from PSV throat) was 193.9 dB ref. 20µPa, which
is consistent with past measurements at similar conditions12.
The measured spectra at positions 1 (ring A), 5 (ring B), and 12
(ring C) for a representative case are shown in Figure 10, where
the response is broadband up to approximately 1 kHz, after
which the modal response in the tailpipe begins to contribute.
These spectra indicate a slight increase in noise from 300 –

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 Table 2. Measured and Predicted Noise Table 4. Header Pipe Internal Noise Measurements –
Baseline
Test Measured IEC (a) IEC (b) EI (LP) EI (LW)
(36”) (20”) (12”)
36” bare 189.4 186.9 191.4 191.0 170.1
36” damp 191.9 186.9 191.4 190.9 170.0 Ring C 186.9 187.6 188.4
13 176.2 180.2 186.3
36” tubes 191.9 187.0 191.6 191.1 170.2
14 170.7 - 179.5
20” bare 195.1 187.6 192.1 191.4 170.6
15 170.4 174.3 176.4
20” damp 194.7 187.4 191.8 191.1 170.3
16 167.8 - 173.9
20” tubes 194.6 187.4 191.8 191.2 170.3
12” bare 194.2 187.4 191.7 191.0 170.1
Table 5. Header Pipe Internal Noise Measurements –
12” damp 194.7 187.1 191.6 191.1 170.3
Tube Bundles
12” tubes 194.6 187.3 191.7 191.0 170.2
(36”) (20”) (12”)
Average 193.4 187.2 191.7 191.1 170.2
Ring C 188.1 187.7 *
Assessment of noise propagation from measurement 13 177.1 180.8 *
rings A to B indicate a decay of 0.2 dB/diameter through the 4” 14 - - *
sch. 40 tailpipe, as shown in Table 3. This is significantly
higher than the EI standard which computes 0.06 dB/diameter, 15 - - *
and suggests either that the EI method is conservative or that at 16 171.3 - *
very high levels (close to the source), the radiated noise and
*Data contaminated by tube bundle vibration.
associated decay is higher than at low levels. Comparing
measurements from rings B and C (Table 3) indicates that noise The presence of the tube bundles led to a minor increase in
may be generated by the elbow as additional turbulence is the overall measured noise inside the header pipe. A
introduced, since the effective attenuation was observed to be comparison of the spectra from measurements at position 13
lower. The noise was observed to increase during the showing the effect of the tubes in the 20” and 36” headers is
36” header test, where rings B and C were closest together. shown in Figure 11 and Figure 12. A dramatic decrease in the
Table 3. Noise Decay in PSV Tailpipe with Distance response of the mode at 255 Hz is observed on the 20” header,
and a slight decrease is observed at 1,160 Hz. A shift is
Location (test) dB/D Total Distance (in) observed in the 36” header data, where primary responses at
A-B (36” header) 0.21 100 94 and 216 Hz are replaced by 275 and 791 Hz; an overall
increase in noise was observed which may be due to additional
A-B (20” header) 0.22 100
turbulence generation.
A-B (12” header) 0.20 100
B-C (36” header) -0.13 47
B-C (20” header) 0.12 63
B-C (12” header) 0.06 72.5

The average level measured at ring C from all tests was

188.0 dB; the measured levels at locations 13 – 16 are shown in
Tables 4 and 5 for tests with and without tube bundles present.
These results suggest that the decay factor when noise
propagates from a tailpipe into a larger header depends upon
the ratio of pipe diameters, with greater attenuation occurring in
larger headers. The observed attenuation was approximately
11 dB into the 36” header, 7 dB into the 20” header, and 2 dB
into the 12” header. It is important to note when observing Figure 11. Dynamic Pressure Spectra Inside 20”
measurements at locations 14 – 16 that flow was allowed to Header
propagate in both directions through the header, and it is
possible that restricting flow to one direction could affect the
noise propagation.

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 Figure 12. Dynamic Pressure Spectra Inside 36” (a)
Header

PHYSICS-BASED MODELING AND COMPARISON

WITH TEST DATA
In an effort to improve physics-based computational
techniques for AIV analysis, three-dimensional structural and
acoustic FE models of each baseline test configuration were
created and analyzed in order to compare with the test data
described previously. The structural and acoustic model results
are compared with each other and to the test data using
coincidence plots that show frequency and spatial similarity
between acoustic and predicted modes. Examples of the
analysis approach and comparison results for these three
headers are presented in this section.
Finite element model geometry for the three header (b)
configurations is shown in Figure 13. Using these models, a
modal analysis for each model was performed using ANSYS
R16.0 finite element software. The header and branch pipe of
each model was meshed using a sweep method with all quad
elements. The rest of the connecting parts were automatically
meshed, with confirmation of sufficiently fine element sizing
between bodies. Two 1” wide strips around the circumference
of the header, each positioned a specified distance from the
header ends, were constrained as a fixed support in order to
match tie-down locations on the test header.
In addition to the solid-element models, simplified shell-
element models were created that incorporated 9 diameters of
header pipe and 4.5 diameters of branch piping with symmetry
boundary conditions in order to compare results with the solid
models and investigate a more efficient modeling approach for (c)
AIV piping.
Figure 13. (a) 12” Header Model; (b) 20” Header
Model; (c) 36” Header Model
A modal analysis was performed up to 3,000 Hz (12” and
20” headers) and 2,000 Hz (36” header) based on enveloping
the observed range of peak frequencies in test data. The modal
analysis results for the 12”, 20”, and 36” headers show 273,
693, and 1,154 modes for each model, respectively.
A ring of nodes on the outside circumference of the header
pipes and roughly 4.8” from the center of the branch pipe was

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 selected as a location to analyze the circumferential In addition to structural models, three-dimensional acoustic
deformation of the header at the resultant frequencies of this models were developed in order to determine the higher-order
simulation. Examples of this circumferential deformation are acoustic mode shapes that interact with the piping to produce
shown in Figure 14. AIV stresses. These models were simple cylinders
corresponding to the internal diameter and length of each test
header. The branch connection and instruments that were a part
of the solid model analysis were not incorporated into this
acoustic modal analysis since they were not expected to have a
strong effect on the modal analysis results.
The structural and acoustic natural frequencies and
circumferential mode shape patterns from each analysis are
combined and overlaid on measured strain data in Figures 16 –
18 for the three baseline header geometries. In these plots, only
the most dominant circumferential and axial mode shape
patterns have been considered. The modal results displayed
have been filtered to remove structural modes do not have
frequencies within 5% of an acoustic mode with the same
dominant patterns. For the 12” diameter header (Figure 16),
there was reasonable agreement between measured peak stress
frequencies and coincidence predictions above approximately
600 Hz. Many coincident modes still existed, but dense clusters
of coincident modes (particularly at the 3ND and 4ND patterns)
occurred near the same frequencies where many stress peaks
were measured. The 20” (Figure 17) and 36” (Figure 18)
arrangements showed less agreement, although some of the
Figure 14. Example Shell Mode Shape Results major stress peaks did align with predicted coincident
frequencies. A greater degree of matching could be acquired by
considering less-dominant spatial patterns in mixed mode
Most of the resulting mode shapes are composed of shapes. One consistent observation among all the test headers
multiple spatial patterns, such as the example mode shape was that major stress peaks occurred at frequencies below the
shown in Figure 15. Thus, it is necessary to quantify the cut-on frequency for 5ND modes, indicating that most AIV
relative contribution of each nodal diameter component and stresses likely occurred at 4ND patterns or lower. Stress peaks
perform analysis for various thresholds of components, e.g., we at frequencies below the first shell mode are caused by beam-
define a threshold number between 0 and 1 that is associated type bending modes of the piping that are excited by flow-
with the amplitude ratio of each nodal diameter component to induced turbulence. These results highlight the fact that AIV
the maximum nodal diameter component. High threshold does not generally occur in isolation from other flow- or
numbers near 1 denote the dominant nodal diameter mechanically-induced vibrations, and overall stress evaluations
components in the mode shape, and lower threshold numbers must include these phenomena as well.
refer to nodal diameter patterns that are less dominant.

Figure 16. 12” Header Coincidence Plot

Figure 15. Waveform Incorporating a 1 and 3 Nodal

Diameter Pattern

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 Initial comparisons with three-dimensional acoustic and
structural model results showed that a very large number of
coincidence conditions (where frequencies and spatial patterns
of acoustic and shell modes match) are predicted, even when
considering both axial and circumferential deformations. The
number of these conditions exceeds the number of peaks
observed in measured stress data. The data also indicate
significant stresses at frequencies below shell modes due to
flow-induced turbulence excitation of bending modes. Thus,
analysis approaches to predict overall piping stresses and
fatigue life should include both types of phenomena in order to
obtain an accurate prediction.

Figure 17. 20” Header Coincidence Plot REFERENCES

[1] B. Skailes, S. S. Ngiam, “A Finite Element Model Analysis
Approach to Assess Piping Failures due to Acoustically Induced
Vibration,” Institute of Noise Control Engineering, Proceedings
of Inter-Noise (2012), New York, NY.

[2] T. Allison, N. Evans, and N. Poerner, “An Efficient Finite

Element Analysis Method for Acoustic Induced Vibration
Analysis,” Institute of Noise Control Engineering, Proceedings of
Inter-Noise (2012), New York, NY.

[3] T. Allison and N. Evans, “Acoustically Induced Vibration in Gas

Piping Systems,” Proceedings of Gas Machinery Conference,
Albuquerque, NM, 2013.

[4] Guidelines for the Avoidance of Vibration Induced Fatigue

Figure 18. 36” Header Coincidence Plot Failures in Process Pipework, Second Edition, Energy Institute,
London, (2008).
CONCLUSIONS AND FUTURE WORK
[5] T. C. Allison and L. J. Goland, “An FEA-Based Acoustic Fatigue
This paper has presented test results from full-scale AIV Analysis Methodology,” ANSYS Regional Conference:
testing, including data for baseline piping geometry and several Engineering the System. August 31 - September 1, 2011.
mitigation methods. Results of the dynamic strain Houston, TX.
measurements suggest that stiffening rings, acoustic mode
disruptors, and viscous damping wrap are all viable AIV [6] N. Evans, D. Arnett, T. Allison, and N. Poerner, “Practical
mitigations, reducing measured dynamic strain by up to and application of AIV analysis methods for screening, qualification,
slightly over 50% for some configurations. However, stiffening and redesign of complex piping systems,” Institute of Noise
rings and tube bundle solutions require engineering for specific Control Engineering, Proceedings of Inter-Noise (2012, New
geometries to obtain optimal results, and improper application York, NY.
of these mitigations could potentially increase AIV stresses.
Viscous damping wrap is effective at providing strain reduction [7] R. D. Bruce, A. S. Bommer, and T. E. Page, “Solving Acoustic-
Induced Vibration Problems in the Design Stage,” Sound and
without the requirement of a detailed design effort. Vibration, August 2013.
Results from the noise measurements suggest that when
computing PSV noise via the IEC method, a liquid pressure [8] V. A. Carucci and R. T. Mueller, “Acoustically induced piping
recovery factor of 1.0 may be appropriate. Additionally, the vibration in high capacity pressure reducing systems,” American
attenuation factor provided in the EI guideline may be Society of Mechanical Engineers, (1983).
conservative when applied to high amplitude noise sources and
when working close to the source. The attenuation applied to [9] F. L. Eisinger, “Designing piping systems against acoustically
propagation into a large header depends strongly upon the ratio induced structural fatigue”, Journal of Pressure Vessel
of tailpipe to header diameter, and can be significant for large Technology. 119, 379-383, (1997).
header diameters. Additional work would be valuable in
quantifying the noise generation effects of pipe fittings such as [10] H. Izuchi, M. Nishiuchi, and G.Y.H. Lee, “Fatigue Life
elbows. Estimation of Piping System for Evaluation of Acoustically
Induced Vibration (AIV),” Institute of Noise Control

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 Engineering, Proceedings of Inter-Noise (2014), Melbourne,
Austrailia.

[11] A. Ghosh, Y. Niu, and R. Arjunan, “Difficulties in Predicting

Cycles of Failure Using Finite Element Analysis of Acoustically
Induced Vibration (AIV) Problems in Piping Systems,”
Proceedings of the ASME 2014 Pressure Vessels & Piping
Division Conference. PVP2014-28325, July 20-24. Anaheim,
CA.

[12] N. D. Evans, “Measurement of High Amplitude Relief Valve

Noise for Acoustically Induced Vibration and Comparison to
Predictive Methods,” Proceedings of the ASME 2014 Pressure
Vessels & Piping Division Conference. PVP2014-28794, July 20-
24. Anaheim, CA.

[13] IEC 60534-8-3. Industrial-process control valves – Part 8-3:

Nose considerations – Control valve aerodynamic noise
prediction method, Edition 3.0, 2010-11, International
Electrotechnical Commission, Geneva Switzerland (2010).

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