PROCEEDINGS of the23rd International Congress on Acoustics
9 to 13 September 2019 in Aachen, Germany

Sound transmission through double-glazed window:
numerical and experimental analyses
Chaima SOUSSI 1; Mathieu AUCEJO1; Walid LARBI1; Jean-François DEÜ1
1

Conservatoire national des arts et métiers, LMSSC, Paris, FranceABSTRACT
Double-glazing is widely used in constructions due of their sound insulation performance. According to thestandards, the prediction of their acoustic performances is carried out in laboratory using two reverberantrooms. It’s well known that, in the low frequency range below the Schroeder frequency, the presence of theserooms can affect the measurements due to their modal behavior. Thus, the prediction of their acoustic andvibratory behavior is very complicated and depends on several factors such as the used materials, theboundary conditions and the diffuseness of the acoustic field.
In this work, experimental and numerical studies on the vibroacoustic behavior in low frequency range of
double-glazing are presented. First, a finite element model for the proposed fluid-structure coupled problem
is proposed. Experimental modal analysis of the structure is then detailed. The results of this analysis areused to validate a numerical model of double-glazing. Finally, various models for numerical sound reduction
prediction of double glazing are presented and compared.
Keywords: Double-glazing, Experimental modal analysis, Finite element.

1. INTRODUCTION
Double glazing is often used in noise reduction since a relatively high transmission loss is achievedby introducing of an acoustic gap between the panels. Due to various effects of fluid-structureinteraction, the prediction of the acoustic performance of the double wall remains very complicatedin the low frequencies whatever the method used. We propose in this paper a reliable method base don experimental measurements and numerical analyzes able to predict the vibroacoustic behavior of
double-glazing. In the literature, the article (1), (2) and (3) studied experimentally the transmission
loss through single, double and triple glazing. In (4), the effect of boundary conditions on this acoustic
indicator of double panels was investigated. A comparison of different type of glass models can befound in (5). For the numerical prediction approaches, several methods are available. The choice ofthe numerical method depends on the computational cost and the treated frequency band. The Finite
Element Method (FEM) is suitable to treat problems in low frequencies. In (6), this method is applied
to develop a model of a sandwich plate with viscoelastic core. The FEM is applied in (7) by the authors
for the different layers of the sound barrier coupled to a variational Boundary Element Method (BEM)
to account for fluid loading.
The present paper is organized as follows: In the first part, a finite element formulation of thevibro-acoustic problem is presented. The second section is devoted to the modal experimental analysisof the double-glazing. The impact hammer test is used to evaluate the modal parameters (naturalfrequencies, mode shape and damping) which were used to recalibrate a numerical model of thestudied system. In the last part, numerical evaluation of the acoustic performances of double -glazing
is presented. The difficulties of the numerical models are related to the representation of (i) theacoustic excitation and (ii) the effects of the emitting and receiving rooms. Thus, four configurations
are proposed from which two are selected. We study the difference between the case where an acoustic
room is modeled and where the Rayleigh integral is used. The effect of the dynamic damping is also
analyzed.
1

Chaima.soussi@lecnam.net

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2. Finite element formulation for the vibro-acoustic problem
Let us consider the linear vibrations of an elastic structure filled with a homogeneous, inviscid andcompressible fluid, neglecting gravity effects. This configuration is chosen to model the doubleglazing system. We establish in this section the local equations of the coupled problem and thecorresponding finite element matrix equations.
2.1 Local equations of the problem
We consider an elastic structure occupying the domain ȳௌ ೔ (݅=1, 2) at the equilibrium. The structure
is subjected to a prescribed displacement ‫ ܝ‬on a part Ȟ௨ and to surface force density ۴ ௗ on thecomplementary part of its external boundary. The interior fluid domain is denoted by ȳி and the fluid–structure interface by Σ (see Figure 1).
We note ࣌, the stress tensor and ߩ௦ , the density of the structure. In addition, since the compressible
fluid is assumed to be inviscid, instead of describing its motion by a fluid displacement vector field,
which requires an appropriate discretization of the fluid irrotationality constraint , we use the pressure
scalar field ‫݌‬. Let us note ݇ the wave number in the fluid (݇ ൌ  ߱Τܿ‫ ܨ‬ሻwhere ܿி is the speed of wave
propagation in the fluid, ߩி the mass density of the fluid, and ‫ܖ‬௦ the normal of the structure. The local
equations of the fluid-structure coupled problem described previously are:ȳௌ ೔(1)࣌ሺ‫ܝ‬ሻ ൅ ߩ௦ ߱ଶ ‫ ܝ‬ൌ Ͳ࣌ሺ‫ܝ‬ሻ‫ܖ‬௦ ൌ ۴ ௗ Ȟఙ(2)‫ ܝ‬ൌ  ‫ܝ‬ௗ Ȟ௨(3)ி࣌ሺ‫ܝ‬ሻ‫ ܖ‬ൌ ‫ ܖ݌‬ȭ(4)ο‫ ݌‬൅ ݇ ଶ ‫ ݌‬ൌ Ͳȳி

(5)

ȭ

(6)

௦

ி

Figure 1 – Double-glazing systemଶ‫݌׏‬Ǥ ‫ ܖ‬ൌ ߩி ߱ ‫ܝ‬Ǥ ‫ܖ‬Eq. (1) corresponds to the elasto-dynamic equation in absence of body force; Eqs. (2), (3) are theprescribed mechanical boundary conditions represent the surface forces applied on Ȟఙ and thedisplacement imposed on Ȟ௨ ,respectively; Eq. (4) results from the action of pressure forces exerted
by the fluid on the structure; Eq. (5) is the Helmholtz equation; and Eq. (6) is the kinematic interface
fluid–structure condition on Σ.
2.2 Finite element discretization
In the finite element context, the previous vibro-acoustic problem is described in terms of structure
displacement ‫ ݑ‬and acoustic pressure ‫݌‬. The discretization of the variational formulation associatedto the previous equations (1) to (6) leads to the following matrix system (see (6) for details):‫ۻ‬௨ Ͳ۹ ௨ െ۱۲௨ Ͳ‫܃‬۴൬൤ Ͳ ۹ ൨ ൅ ݅߱ ൤ Ͳ ۲ ൨ െ ߱ଶ ൤ ୘(7)൨൰ ቂ ቃ ൌ ቂ ቃ۱‫ۻ‬௣‫۾‬Ͳ୮௣with ۹ ௨ , ‫ۻ‬௨ and ۲௨ are the structural stiffness, mass and damping matrices, ۹ ୮ , ‫ۻ‬௣ and ۲௣ arethe associated acoustic matrices, ۱ is the fluid-structure coupling matrix and ۴ is the structural load.

3. Experimental analysis of the double-glazing
The modal analysis of the structure is essential for understanding its dynamic performance. In thissection, we are interested in experimental modal characterization (resonant frequencies and dampingcoefficients) of double glazing. The experimental results will be used for the validation of a developed
numerical model.
3.1 Description of the structure
The complete structure presented in Figure 2.a is a domestic wooden window with two symmetrical
opening mounted on a frame. The model is 1.45 m wide by 1.5 m high. It weights approximately 76 kg.
Each opening is composed of double-glazed surrounded by a rubber seal and incorporated in a sash
which is composed of two stiles and two rails (Figure 2.b). The complete, the window is composed of92 pieces of different sizes and different materials such are wood, glass, rubber, etc.

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Figure 2 – (a) Wooden window, (b) Components of one opening
The double-glazing, presented in Figure 3, combines two glass panes separated by a 18 mm spacer
and an argon gap. It is 0.581 m wide by 1.309 m high. The polyamide spacer is filled at least half by
the desiccant, which is a silicon material used to absorb the moisture within the cavity through theperforations of the spacer. The gluing of the spacer on the window is done with a butyl type sealantor resistant silicone. All the parts are connected and sealed with a visible silicone located around theoutside edge.

Figure 3 – Details of the double-glazing
3.2 Experimental Modal Analysis
The Experimental Modal Analysis (EMA) is a tool to characterize the modal properties of acomplex structure through the identification of its natural frequencies, mode shapes and modal
damping ratios. Basically, the implementation of an EMA requires the measurement of the Frequency
Response Function (FRF) of the considered structure at different points of the latter (Figure 4). In thiswork, the “roving hammer” test is used with an impact hammer and three reference accelerometers.
This method consists in fixing accelerometers and then impact the structure at several points.
3.3 Impact testing considerations
The basic principle of the roving hammer test is to hit the structure with an impulse and thenmeasure its response. This needs to be done at several points on the structure to obtain a fair
representation of the mode shapes. The impact hammer is equipped with a force sensor and a nylontip.

Accelerometer
Impact hammerFigure 4 – The device of modal analysis
The suspended double-glazing (see Figure 5) is meshed into 153 nodes. Excitation is done in all
nodes with an average of three impacts and measurements are performed up to 400 Hz, with 0.5 Hz
frequency resolution.

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Figure 5 – Suspended double-glazing
3.4 Numerical ModelIn this section, the double glazing experimentally tested is modeled numerically with the finite
element method. The objective is to find a numerical representation of the structure able ofreproducing its dynamic modal behavior. Thus, the model found will be used for acoustic studies.
The numerical model consists of two 2D shells representing the two glass panels separated by athree-dimensional cavity filled with argon (sound speed 317 m/s and density 1.6 kg/m 3 ). The structure
is considered free in the space. The properties of the glass defined for the two planes are ‫ ܧ‬ൌ ͸Ͳ ;ߩ ൌ ʹͶͷͲ (kg/m3 ) and Q =0.23. An equivalent material, whose properties are ‫ ܧ‬ൌ ͲǤͳ  ;ߩ ൌ ͳͲͲͲ(kg/m3 ) and Q=0.49, is defined for the group «spacer + butyl + desiccant». Quadrilateralshell elements for the solid and hexahedral elements for the fluid are used for the finite element
discretization. The mesh sizes for the structure and the acoustic domain are controlled by thewavelength ߣ (m) which depends on the frequency range of interest. For the acoustic domain, theacoustic wavelength is ߣ௔ ൌ  ܿΤ݂ where ܿ (m/s) is the speed of sound. For the structure, the flexuralwavelength is ߣ௙ ൌ ඥʹߨΤ݂ ሺ‫ܦ‬Τ‫ܯ‬ሻଵȀସ where ‫( ܦ‬Pa.m3 ) is the flexural rigidity and ‫( ܯ‬kg/m²) thesurface mass density. Since it is recommended to use 6 elemen ts per wavelength, we obtain a model
containing 35604 degrees of freedom.
3.5 Results and discussion
Firstly, in order to validate the experimental measurements, the principle of reciprocity must beverified. It means that the frequency response function measured between two points must be the same
regardless of which of them is input or output. Two cases are tested; the first one where the excitation
point P1 and the measurement P2 are in the same glass pane. The second one where the excitation
point is in the first glass pane and the measurement point is in the second one. The results presented
in Figure 6 show that the reciprocity is insured since the curves are superposed following thepermutation between the excitation and measurement points.

Figure 6 – Verification of the reciprocity
For analyzing the experimental data, a home-made EMA toolbox based on the pLSCF(polyreference Least Squares Complex Frequency domain) (8) has been implemented. Figure 7 shows
the comparison of the first ten natural frequencies of the structure obtained from the experimental andnumerical analysis. For a such complex structure, results are considered satisfying since the maximum
difference of natural frequency is 6.8 %, observed for Mode 8.

198

100Frequency (Hz)

120
100

Numerical

90

Experimental

80

Gap

70
60

80

506040Gap (%)

140

30

40

20

20

100012345678

910ModeFigure 7 – Comparison of the first ten experimental and numerical eigenfrequencies of the double-glazing
Table 1 presents a comparison of modes shapes 1, 2 3 and 6 experimentally measured and foundby finite element method. Once again, a good agreement between the two methods is observed.
Table 1–Comparison of experimental and numerical mode shapes
Mode1236Experimental
29.9 Hz37.7 Hz41.2 Hz75.5 Hz

Numerical
29.8 Hz36.6 Hz38.6 Hz73 Hz
The modal damping is often the parameter extracted with the greatest uncertainty (9). Results ofthe first ten modal damping presented in the Figure 8 show that this parameter variates from 0.9 % at
the 6th mode to reaches the maximum at the 2 nd mode with a value of 11.3 %. These data will be usedfor vibroacoustic numerical simulations of the problem.
12Damping (%)

108642012345678910ModeFigure 8 – Experimental modal damping of the double-glazing

4. Measurements of sound transmission
4.1 Experimental set-up
The airborne sound insulation of building elements such as walls, doors, windows can be evaluatedin laboratory according to standards norms (10). These standards allow the determination of theacoustic performances of the structures by the measurement of the sound reduction index.
Measurement takes into consideration two adjacent reverberation rooms, separated by a common wallcontaining an opening in which the test element is mounted (see Figure 9). During the laboratory

measurements, ܴ is defined as the difference between the average sound pressure levels in the source

199room and in the receiving room with a correction term considering the acoustic absorption ‫( ܣ‬m²)
and the area of the test specimen ܵ (m²):ܴ ൌ ‫ܮ‬௦ െ ‫ܮ‬௥ ൅ ͳͲ ଵ଴ ሺܵΤ‫ܣ‬ሻ(8)Source chamber

Rec eiv ing c hamber

Figure 9 – Experimental set-up of acoustic experimental test
4.2 Other configurations
Standards (10) assume that acoustic excitation is a diffuse field for the entire frequency band ofmeasurement [100 Hz - 5000 Hz]. This condition is respected only beyond the “Schroeder frequency”
of the room which marks the limit between the modal behavior and the diffuse field. This frequency
is generally high and inversely proportional to the dimensions of the chamber.
At low frequencies, to treat the problem numerically by respecting the condition of the diffuse
field, it is necessary to model a room whose dimensions are larger than the wavelength which ispenalizing for the cost of computation. To overcome this problem, the plane wave superposition
method is an effective solution. The technique consists in exciting the structure by a diffuse soundfield (Figure 10.b). This solution therefore makes it possible to avoid modeling the emitting chambe r.
On the reception side, it can be assumed that the window radiates in a semi -infinite field without
considering the receiving acoustic cavity (Figure 10.c). The combination of these two solutions,presented by the last configuration (Figure 10.d), makes it possible to reduce the size of the modeland to avoid the modal effects of the rooms.
For numerical methods, the FEM method is used to model the window. For the modeling of thetransmission and reception sides, the method used will depend on the chosen configuration. For
example, for cases where both chambers are modeled, the FEM will be used since it is a bounded
problem. For the cases of radiation in a free field corresponding to an unbounded problem, se veral
solutions are presented (Rayleigh, BEM or also FEM combined with a perfectly matched layer (PML)
for the far field.

Figure 10 – Different configurations considered for problem resolution
4.3 Numerical analysis
In this section, we focus our attention on the numerical analysis of configurations (b) and (d) tocalculate the sound transmission of the double-glazing model previously established. For the two cases
considered, the excitation of the structure is a diffuse sound field which is a superposition of random

200

plane waves in phase and direction making a uniform sound pressure at every point. The difference
between the two analyses is the modelling of the receiving side: In the first configuration, the structureradiates in a 3.25ൈ4ൈ5 m3 rigid rectangular cavity. The structure is clamped to a wall of the room. In
the second case, the Rayleigh integral is used in the receiving side. This condition represents theacoustic radiation of a plane structure mounted in a rigid baffle in a half-space. Since direct frequency
analyses are carried out, only materials damping is defined in the models.
4.3.2 Results and discussion
The carried numerical models allow to determinate the Transmission Loss of the structure (TL).
This parameter is calculated in decibels (dB) as the ratio of the sound power incident ‫ݓ‬௜ by the soundpower transmitted ‫ݓ‬௧ by the structure: ൌ ͳͲ ଵ଴ ሺ‫ݓ‬௜ Τ‫ݓ‬௧ ሻ(9)A comparison of the results in narrow band and 1/3 octave is established. The fr equency range fornarrow band figure goes from 1Hz to 600 Hz and from 31 Hz to 500 Hz for the 1/3 octave bandpresentation. A damping of 1% is used for all the material of the models. Firstly, a comparison of theTL calculated with the two configurations (b) and (d) is illustrated in the Figure 11. We can see in thenarrow frequency band, a fluctuation of the TL at the low frequencies with the presence of thereceiving room. This is explained by the modal behavior of the acoustic cavity. In fact, the coincidenceof the excitation with the acoustic natural frequencies, for example Mode 2 at 43 Hz, Mode 7 at 68 Hz
and Mode 28 at 132 Hz presented in Figure 11, causes an increase in the sound pressure level in theroom, which leads to a decrease of the measured TL. This is not the case when the structure radiatedin a semi-infinite domain. The sound reduction calculated in the two configurations becomes similarfrom 200 Hz, a value around the cutoff frequency of the acoustic cavity.

Figure 11 – Comparison of the TL between the two configurations in the narrow and 1/3 octave band
The second aim of this section is to study the influence of the damping in the TL of the structure.
For this, two analyses are made with the configuration (d). Results presented by the Figure 12
disclosed the fact that the damping smooths the curve of the TL and influences considerably the results
above 100 Hz. At 500 Hz, the model without damping present a results 15 dB lower than the case with1% damping. Below 100 Hz, the TL of all models present a similar behavior.

Figure 12 – Effect of the damping in the TL of the double-glazing (configuration d)201

5. CONCLUSIONS AND PERSPECTIVESIn this work, the vibroacoustic behavior of the double-glazing glass is studied. First, the modal
behavior of the structure is experimentally identified and validated with the finite element method.
The established numerical model was used to carry out different acoustic analyses. The numericalexamples show that the presence of the acoustic room has a significant influence on the TL in the very
low frequencies due to its modal behavior below the Schneider frequency. In addition, the dynamic
damping doesn’t have an influence below 100 Hz but improve the acoustic performances in the higher
frequencies.
Within this framework, many future investigations and developments can be considered among
which: the calculation of the TL with other configurations to study the influence of the acoustic
cavities and their geometries; the influence of the structural and the acoustic boundary conditions; therecalibration of a model of one opening and then the hole window.

ACKNOWLEDGEMENTS
The authors would like to express their thanks to CODIFAB (Comité professionnel dedéveloppement des industries françaises de l’ameublement et du bois) for its financial support.

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