PROCEEDINGS of the
23rd International Congress on Acoustics
9 to 13 September 2019 in Aachen, Germany

Acoustic validation of calculation software for ducts, panels androom acoustics
Ir. Christiaan Cornelis van Dijk1
1

Alara-Lukagro, Netherlands

ABSTRACT
Alara-Lukagro is a project based manufacturer of acoustic enclosures, components and attenuators. These
are often placed in projects where they are part of a number of components that have a strong effect on the
acoustics. In order to predict such a variety of situations software was developed in line with international
noise mitigation standards. The attenuation values are based on laboratory measurements and measurements
in situ. Low frequency attenuation rarely results in shortcomings, a larger spread in results is found in thehigh frequency range.. This paper shows how the attenuation is determined on a geometric chain model, andhow it can be engineered in order to design a good attenuator in complex realistic acoustic environments.
Keywords: Attenuators, Directivity dependence, Noise attenuations

1. INTRODUCTION
As a manufacturer of acoustic solutions, the received information is limited. Mostly however thisinformation satisfies to calculate simple and complex projects based on calculation rules and test
results from laboratories. Noise requirements have become more stringent and there is a lot ofcompetition. Due to variable rotation velocities and the way noise source data is often only available
in octave band values, splitter silencers are often a more suitable option than resonator silencers.
Fitting a broadband solution to a set of data and geometries leads to a design meant to beat thecompetition. How certain is this outcome, and does the wat we calculate lead to systematic errors thatcan or should be avoided in order to make better predictions on the end result of a noise control
design? The evaluation considers a number of aspects:
x How accurate can absorption be predicted?
x How well does the division into near field and far field Eyring(1) work?
x How accurate is data provided by the suppliers?
x How well can transmission loss from laboratories be compared with transmission loss inthe field?

2. Model2.1 General design
The main philosophy is to use a standard calculation program for critical paths and for bulks. The
calculation program for critical paths is most interesting. It creates a chain of steps, that can be mere
algebraic, or unfold from 1D into 3D (room) acoustic models, always respecting calculation rules,
room acoustics and the energy balance. Some boundary assumptions for reversed sound apply.

Figure 1 – Chain of elements 1 D information potentially unfolded into 3D blocks
1ch.van.dijk@alara-lukagro.com

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2.2 Boundary conditions
In the current model, sound that is reflected back into the source channel opening is assumed todisappear completely. Multiple bent reflections therefore do not lead to a theoretical increase at thereceiver side. Boundary conditions and algebraic components were evaluated, but are not included inthis report.
2.3 Room acoustics
The direct noise, the ground reflections and the corner or edge reflections are summed up as the
direct sound. The reverberant noise is reduced by the direct source contribution, when direct soundcan be calculated. The sound acting upon a surface is the sum of the reverberant and the direct sound.

Figure 2 – Example of a model with 6 elements, optimization and free field problem
2.4 Flow generated noise
For each step (change from room to room, or channel to room, or room to channel, or room/channel
into free field) the flow generated noise power per area is calculated in accordance with VDI 2081(2)'worst case' this noise is added to the sound power. The levels are always based upon the highest
average velocity and smallest area. If there is any relevant throttling, that will generate flow generated
noise, the program will automatically account for it at the end of each block.
2.5 Geometric expansion
As a manufacturer, our solutions are generally evaluated at close proximity. The noise is still quite
directional and is likely less under a 90 o angle from the source plane than on a small angle from theradiating surface. This led to the implementation of a safety margin, This safety margin is in line withthe expansion of a theoretical cube (see figure 3). Further away the curvature of noise leads to a wider
expansion. Because as a manufacturer we are rarely informed about the specifics of the environment(reflections) currently there is no transition into normal halve bulb expansion (yet).

Figure 3 – Expansion in close proximity of the source
Besides expansion near field correction may apply. Because the calculation is done automaticallynear field correction always applies. Since there are so many steps in any given calculation, it was
decided to use a more accurate continuous equation for the near field correction, which is based on abest fit with the numbers in the ILHR13.01(3).
2.6 Available elements
Numerous restraints, requirements and conditions can be set. More interesting are the currently
available elements. The elements that are available in the calculation t ool are:

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A
B
C
D
E
F
G
H
I
J

Table 1 – Elements (phenomena)
Rounded bents
Sharp bents
Expansions
Room acoustics
Sound insulation elements
Splitter attenuators (optimization )LouvresAcoustic louvres
Sound sources (Lw, Lp )Algebraic user defined equations

This means any situation that was evaluated was built up in the calculation tool only with the above
mentioned elements.
2.7 Evaluated projects
A large number of projects ware evaluated. Projects that were measured after the introduction of
the calculation tool were measured more extensively with the aim to carry out this evaluation. The
project that were included to evaluate the evaluation tool are in table 2 (anonymized).
No.
1
2
3
4
5
6
7
8
9
10
11
12
13
14

Table 2 - Projects
description
Energy plant in residential area with ten 2.5 MW el. Gas engines (United Kingdom)
Test facility with three test cells for diesel and gas engines for large vessels (Germany)
Energy plant in residential area with one 2.5 MW el. Gas engine (France)
Heat and power mini plant 1.6 MW el.in a greenhouse (Belgium)
Silencer on a chimney (Netherlands)
Power pack test 500 kW el. (Netherlands)
Mobile Pump installation test (Netherlands)
Heat and power mini plant 1.6 MW el. In a greenhouse (Netherlands)
Emergency power with 9 diesel engines 2 MW el.(Germany)
Existing silencer on a production facility (Netherlands)
Enclosure around an electrical pump installation (Netherlands)
New silencer in an underground concrete structure (Netherlands)
Gas power plant with more than three 10 MW el. gas engines (Germany)
Power containers for emergency power with 3.5 MW el. diesel engines (Netherlands)

For privacy reasons of our customers no more details than mentioned above will be made available.
This list is to illustrate the nature and the magnitude of this evaluation.

3.

Evaluation

3.1 Different reasons for measurements
The projects that were evaluated were not chosen randomly. In general projects are not evaluated,
unless there is a suspicion, or an agreed request before sales to do so. When the data of the evaluation
is processed, this intention was to keep this distinction. In practice other factors can disturb themeasurements. In order to maintain a significant amount of measurements all projects are included.
3.2 Absorption values
As a noise control solution firm a lot of projects involve big gas or diesel internal combustion
engines, which typically require attenuation in the lower frequencies (31 -250 Hz octave bands). The
rate between direct noise radiated on acoustic elements and reverberant noise acting upon the element
strongly favors the reverberant noise. So the absorption coefficient is a very important paramete r. The
absorption coefficient in an impedance tube is typically lower than in a reverberant room. The
situation in practice is most often more like a reverberant room than an impedance tube. However the
ISO 354 (4) allows to use sample sizes that increase the area of absorption to above the limit where
Sabine applies (1).

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Therefore it is important to determine the right absorption coefficient. In some projects the soundsource couldn't be switched off, or the room was too big, or too noisy to determine reverberation times.
Therefore not all noise measurements include reverberation measurements.
Initially the absorption coefficient used in the calculation tool was a safe estimate based on the
Sabine absorption coefficient according to ISO 354 (4). Improvement by using measured Eyring
absorption values already resulted in a 0.8 dB better prediction of the acoustic pressure in themachinery room. Taking into account that measurements were done manually and an acoustician w as
therefore present in the room another 0.05 dB of the total dB(A) level can be accounted for. The usedaverage absorption with a different material property list is given in see table 3, the influence ofmanned measurements in table 6.
Table 3 – Eyring average shell Absorption values [-]See Table 2
1 old
1 new
3 old
3 new
4 old
4 new
8 old
8 new
9 old
9 new
11 old11 new
13 old13 new
14 old14 new

31
Hz
0.10
0.09
0.03
0.03
0.11
0.12
0.11
0.12
0.05
0.05
0.09
0.08
0.04
0.05
0.10
0.11

63
Hz
0.13
0.16
0.03
0.04
0.15
0.22
0.15
0.22
0.05
0.05
0.13
0.16
0.04
0.05
0.15
0.22

125
Hz
0.31
0.26
0.11
0.10
0.32
0.31
0.32
0.31
0.04
0.05
0.32
0.27
0.04
0.05
0.32
0.32

250
Hz
0.59
0.41
0.13
0.12
0.61
0.46
0.61
0.46
0.05
0.08
0.64
0.44
0.04
0.08
0.63
0.47

500
Hz
0.66
0.44
0.14
0.14
0.69
0.48
0.69
0.48
0.06
0.09
0.71
0.46
0.06
0.09
0.71
0.49

1
kHz
0.73
0.55
0.18
0.17
0.76
0.58
0.76
0.58
0.08
0.10
0.79
0.59
0.08
0.10
0.79
0.59

2
kHz
0.73
0.61
0.17
0.17
0.76
0.64
0.76
0.64
0.07
0.08
0.79
0.66
0.07
0.08
0.78
0.65

4
kHz
0.69
0.64
0.15
0.18
0.73
0.67
0.73
0.67
0.06
0.07
0.75
0.70
0.06
0.06
0.74
0.68

8
kHz
0.69
0.59
0.14
0.18
0.73
0.62
0.73
0.62
0.07
0.07
0.75
0.65
0.07
0.06
0.74
0.62

A typically cluttered room has a higher absorption value than an uncluttered room. Taking into
consideration the improved absorption from typical clutter in three test cells, the indirect noise can befurther reduced. In table 4 the increase in average Eyring absorption in three test facilities with a shellarea of 284.4 m 2 as a result of typical clutter is shown.
Table 4 – Eyring average shell Absorption values empty and cluttered measured 9[-]See Table 2

31
Hz

63
Hz

125
Hz

250
Hz

500
Hz

1
kHz

2
kHz

4
kHz

8
kHz

2 empty

0.03

0.05

0.08

0.18

0.25

0.35

0.40

0.42

0.38

2 cluttered

0.07

0.11

0.15

0.23

0.30

0.34

0.32

0.32

0.37

The shell acoustic model representation is chosen, because adding absorption has its limitations.
In a room it is not impossible to increase the product of the absorption times area in some octave
bands to a value more than three times the initial shell area, using the ceiling only, see figure 4.
However this will change the direction of the sound field and the average absorption will bot rise far
above 0.2 when the ceiling area is 0.2 times the total shell area and the rest of the shell is still fullyreflective.

Figure 4 side view of a hypothetical room with three times the shell area as absorption on the ceiling

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3.3 Source data
Based on the room acoustics model and the validation measurements on panels as used in thesemeasured situations the sound power level that was provided by the customer or machine manufacturer
was compared with the measurements taken around the engine or sound source. Before such a
comparison can be made the consistency of 16 equal engine rooms with the same measurement device
and procedure was done The standard deviation of this exercise as shown in table 5 g ives an upper
limit to the inaccuracy of the measurements of the calibrated device.
Table 5 – Standard deviation of average noise levels in equal engine rooms under equal load in dB( -)
16
Hz

31
Hz

63
Hz

125
Hz

250
Hz

500
Hz

1
kHz

2
kHz

4
kHz

8
kHz

dB(A)

0.9

1.2

1.8

0.9

0.8

0.7

0.9

0.8

1.2

0.8

1.0

Initially the same 8 projects as in table 3 were compared. There is a striking inaccuracy found in asingle manufacturer, another engine was not operating on the designed operating pressure. After theseprojects were removed 3 projects and 5 engine rooms remained representative. This improved theaccuracy of the 125 Hz values, but had limited effect on the 63 Hz values. Furthermore no dB(A)value in any machinery room was at least 6 dB(-) less than predicted, see figure 5.

Figure 5 overestimation of sound pressure in a machinery room
When the additional absorption from clutter as shown in table 4 is assumed to be half related to
the floor and half to the walls a hypothetical change in absorption as a result of clutter can be derived
as shown in table 6. The room acoustics can be re-evaluated as shown in figure 6. The dotted red linesthat are not mentioned in the legend show the standard deviation for identical systems from table 5.
When poor predictions are left out and the average clutter is included the calculated and measured
sound pressure values differ roughly as much as the standard deviation shown in table 5. Around 2 dB
more sound power in the 63 Hz octave band is needed in order to create a perfect fit.
Table 6 Influence of clutter on average shell absorption values
See table 2
1 manned
1 cluttered
3 manned
3 cluttered
4 manned
4 cluttered
8 manned
8 cluttered
9 manned
9 cluttered

31
Hz
0,09
0,10
0,03
0,05
0,12
0,12
0,12
0,12
0,06
0,08

63
Hz
0,16
0,32
0,04
0,07
0,22
0,32
0,22
0,32
0,05
0,15

125
Hz
0,26
0,41
0,11
0,14
0,32
0,41
0,32
0,41
0,06
0,14

250
Hz
0,42
0,53
0,13
0,15
0,47
0,53
0,47
0,53
0,08
0,15

500
Hz
0,45
0,57
0,14
0,17
0,48
0,56
0,48
0,56
0,09
0,17

1
kHz
0,56
0,56
0,18
0,18
0,58
0,58
0,58
0,58
0,11
0,11

2
kHz
0,62
0,62
0,18
0,18
0,64
0,64
0,64
0,64
0,09
0,09

4
Hz
0,65
0,65
0,18
0,18
0,68
0,68
0,68
0,68
0,08
0,08

8
kHz
0,60
0,60
0,18
0,18
0,62
0,62
0,62
0,62
0,08
0,08

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11 manned
11 cluttered
13 manned
13 cluttered
14 manned
14 cluttered

0,08
0,09
0,05
0,07
0,11
0,11

0,16
0,37
0,05
0,07
0,22
0,33

0,27
0,47
0,05
0,07
0,32
0,43

0,45
0,59
0,08
0,09
0,48
0,56

0,47
0,64
0,09
0,10
0,50
0,59

0,60
0,60
0,10
0,10
0,60
0,60

0,68
0,68
0,08
0,08
0,66
0,66

0,72
0,72
0,06
0,06
0,69
0,69

0,66
0,66
0,06
0,06
0,63
0,63

Figure 5 For clutter corrected predicted sound pressure levels in machinery rooms
3.4 Attenuator values in practice
In line with the last paragraph the attenuation was evaluated, see figure 6. Therefore the adjusted
sound power spectrum was used based on a uncluttered shell in line with individual source
measurements and the average shell absorption value mentioned as 'new' in table 3.

Figure 6 Overestimations on the transmission loss for splitter attenuators
Note that how poor the prediction of absorption values the sound pressure is fitted to themeasurements, so the sound pressure in the room is always correct. Only the direc t noise is potentially
underestimated. The reverberant noise is much more than the direct noise contribution, so the potential
underestimation of the noise acting upon the silencers is negligible.
Two silencers were excluded from figure 6. One of them had such high attenuation in the highfrequencies that the noise didn't exceed the background noise, even not at 0,1 m. from the silencer.
The other one was around 30 dB worse for frequencies of 1 kHz and above, but was not our own makeand had unknown infill material. So the numbers cannot be properly compared. Because of the2564interesting situation a schematic of this underperforming silencer is given in figure 7. The noise in thepipe (on the right of figure 7) can be assumed omnidirectional, so the square diffusor, between thepipe bend and the attenuator directs the noise in a straight line through the splitters.

Figure 7. Silencer with serious high frequency shortcomings.
Directionality has a strong influence on the high frequency transmission loss, see figure 8.

Figure 8 The effect of angle of incidence on high frequency attenuation (3, from table 2)

Figure 9 Visual representation of the dispersion of acoustic pressure for different frequencies
As can be seen in figure 8, the directivity of sound can have a strong effect in higher frequencies
and has no effect on the lower frequency attenuation. The effect in high frequencies can easily exceed
10-20 dB above 2 kHz. In the low frequencies the position of the attenuator can have a strong effectin relation to standing waves. This effect only occurs in lower frequencies.
In figure 6 the spread in attenuation is high in the 63 Hz, than reduces and keeps increasing above
250 Hz. The average shortcoming in 63 Hz in figure 6 is again 2 dB, the same value as in figure 5.
The measurements are done by a pressure measurement device only, the procedure (5) prescribes,
measurements are not done closer than 0.5 m from a hard surface. Therefore measurements are

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potentially carried out in a standing wave were sound propagates as velocity rather that pressure. A
2 dB underestimation of the sound pressure is consistent with physics. Thus when deriving a soundpower number in 63 Hz from measurements in an echoic room, a 2 dB safety margin should be applied.
The numbers are yet to be fit to a new theoretical model, but as a first impression from typicalmachinery rooms, the performance of the splitter attenuator can differ in accordance to figure 10.
Differences work both ways, positive and negative. Which is a good motivator in order to deduct
simplified algebraic equations that predict these design features.

Figure 10. Typical design effects for Transmission loss values
3.5 Risk from vibrations
There is limited space I n this paper to go into detail. In practice a 1.7 dB margin for constructions
with a good vibration insulation is sufficient. When vibration is an issue no single safety margin canrealistically account for the potential defects.

4. Conclusions
When using only the shell and the average absorption value for room acous tics, it is important to
use the Eyring absorption values and to account for the clutter. If this is not done the sound pressurelevel in a room as a result if the engine power is overestimated by respectively around 0.8 and 3.5 dB.
Using a sound pressure meter and a room acoustic model to evaluate the sound power emitted in amachinery room will likely result in a 2 dB underestimation of the sound power in the 63 Hz octave
band.
Splitter attenuators are highly unpredictable but behave on average the same as under laboratory
circumstances. Location affects the low frequencies and angle of incidence the high frequencies. It is
worthwhile to deduct equations and differentiate direct and reverberant noise to predict these effects.

ACKNOWLEDGEMENTS
I'd like to acknowledge Theo Campmans of LBP sight for addressing me to the magnitude of thedifference in high frequency attenuation for different angles of incidence. Furthermore I'd like toacknowledge André Hameete and Alara-Lukagro for sponsoring this extensive evaluation.

REFERENCES
1. van Dijk, CC, De analyse van isolatie- en absorptiewaarden in absorberende ruimten, Blad geluid no. 4
demeber2015; Wolters Kluwer; p. 24-27.
2. VDI 2081 Noise generation and noise reduction in air-conditioning systems, Verein Deutscher Ingenieure,
Düsseldorf 2001. ICS 17.140.20; 91.140.30; 2003-08.
3. ILHR13.01, Handleiding meten en rekenen industrielawaai, ir W.C. Reij; Interdepartementale commissie
geluidhinder; 1981
4. ISO 354. Acoustics — Measurement of sound absorption in a reverberation room. INTERNATIONAL
STANDARD; 2003(E).
5. ISO 11957, Acoustics - Determination of sound insulation performance of cabins - Laboratory and in situ
measurements. INTERNATIONAL STANDARD; 2009(E).

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