EP2544922A1

EP2544922A1 - Automotive trim part for sound insulation and absorption

Info

Publication number: EP2544922A1
Application number: EP11708025A
Authority: EP
Inventors: Claudio Bertolini; Claudio Castagnetti
Original assignee: Autoneum Management AG
Current assignee: Autoneum Management AG
Priority date: 2010-03-09
Filing date: 2011-03-09
Publication date: 2013-01-16
Also published as: BR112012018864B1; EP2364881A1; EP2364881B1; CN102892631A; AR080393A1; CA2788430C; WO2011110587A1; BR112012018864A2; ZA201205740B; PL2364881T3; RU2012142925A; JP5769737B2; MY166308A; KR101550234B1; KR20130038196A; RU2549581C2; CA2788430A1; ES2542886T3; CN102892631B; JP2013521190A

Abstract

A sound insulating trim part with at least one area with predominantly sound absorbing characteristics (absorbing area), at least one other area with acoustic mass-spring characteristics (insulating area), characterised in that the mass layer of the insulating area consists of the same porous fibrous layer as the absorbing area, in the insulating area adjusted to a dynamic Young's modulus (Pa) of at least (96-AW-t) with AW area weight (g/m²), and t thickness (mm) of the porous fibrous layer, and at least an impervious thin barrier layer between the porous fibrous layer and the decoupling layer.

Description

Automotive trim part for sound insulation and absorption

Technical Field

The invention relates to an automotive trim part for noise attenuation in a vehicle.

Background Art

The sources of noise in a vehicle are many and include, among others, power train, driveline, tire contact patch (excited by the road surface), brakes, and wind. The noise generated by all these sources inside the vehicle’s cabin covers a rather large frequency range that, for normal diesel and petrol vehicles, can go up to 6.3kHz (above this frequency, the acoustical power radiated by the noise sources in a vehicle is generally negligible). Vehicle noise is generally divided into low, middle and high frequency noise. Typically, low frequency noise can be considered to cover the frequency range between 50Hz and 500Hz and is dominated by “structure-borne” noise: vibration is transmitted to the panels surrounding the passengers’ cabin via a variety of structural paths and such panels then radiate noise into the cabin itself. On the other hand, typically high-frequency noise can be considered to cover the frequency range above 2kHz. High-frequency noise is typically dominated by “airborne” noise: in this case the transmission of vibration to the panels surrounding the passengers’ cabin takes place through airborne paths. It is recognised that a grey area exists, where the two effects are combined and neither of the two dominates. However, for passenger comfort, it is important that the noise is attenuated in the middle frequency range as well as in the low and high frequency ranges.
For noise attenuation in vehicles like cars and trucks the use of insulators, dampers and absorbers to reflect and dissipate sound and thus reduce the overall interior sound level is well known.
Insulation is traditionally obtained by means of a mass-spring barrier system, whereby the mass element is formed by a layer of high density impervious material normally designated as heavy layer and the spring element is formed by a layer of low density material like a non compressed felt or foam.
The name “mass-spring” is commonly used to define a barrier system that provides sound insulation through the combination of two elements “mass” and “spring”. A part or a device is said to work as a “mass-spring” if its physical behaviour can be represented by the combination of a mass element and a spring element. An ideal mass-spring system acts as a sound insulator thanks mainly to the mechanical characteristics of its elements, which are bonded together.
A mass-spring system is normally put in a car on top of a steel layer with the spring element in contact with the steel. If considered as a whole, the complete system (mass spring plus steel layer) has the characteristics of a double partition. The insertion loss is a quantity that describes how effective is the action of the mass-spring system when put on top of the steel layer, independently from the insulation provided by steel layer itself. The insertion loss therefore shows the insulation performance of the mass-spring system.
The theoretical insertion loss curve (IL, measured in dB) that characterizes a mass-spring system has in particular following features. On most of the frequency range, the curve increases with the frequency in an approximately linear way, and the rate of growth is about 12dB/octave; such linear trend is considered very effective to guarantee a good insulation against the incoming sound waves and, for this reason, mass-spring systems have been widely used in the automotive industry. This trend is achieved only above a certain frequency value, called “resonance frequency of the mass-spring system”, at which the system is not effective as a sound insulator. The resonance frequency depends mainly on the weight of the mass element (the higher the weight, the lower the resonance frequency) and on the stiffness of the spring (the higher the stiffness, the higher the resonance frequency). At this frequency, the vibration of the mass element is even higher than that of the underlying structure, and thus the noise radiated by the mass element is even higher that the one that would be radiated by the underlying structure without mass-spring system. As a consequence, around the resonance frequency of the mass-spring system the IL curve has a minimum.
Both absorbing and insulating systems on their own have only a small bandwidth of frequencies where they work optimally. The absorber generally works better in the high frequencies, while the insulator generally works better in the low frequencies. Furthermore, both systems are sub optimal for use in a modern vehicle. The effectiveness of an insulator is strongly dependent on its weight, the higher the weight the more effective the insulator. The effectiveness of an absorber on the other hand is strongly dependent on the thickness of the material, the thicker the better. Both thickness and weight are becoming increasingly restricted, however. For example, the weight impacts the vehicle’s fuel economy and the thickness of the material impacts the vehicle’s spaciousness.
Recently a trend towards lower weights for the mass layer or heavy layer for classical mass-spring systems has decreased the average weight from about 3 (kg/m²) to around 2 (kg/m²). This drop in area weight also means less material used with common technology and therefore less cost. Even lower weights down to 1 (kg/m²) are possible and present on the market, but the technology to achieve this is expensive and has drawbacks in particular for low volume mass production. Typical classical mass layers are made of highly filled dense materials, such as EPDM, EVA, PU, PP etc. Since these materials have a high density, normally above 1000 (kg/m³), it is necessary to make a very thin layer to obtain the low area weight. This can increase production costs and cause production problems such as the material tearing easily during moulding.
The insulation performance of an acoustical barrier is assessed by sound transmission loss (TL). The ability of an acoustical barrier to reduce the intensity of the noise being transmitted depends on the nature of the materials forming the barrier. An important physical property controlling sound TL of an acoustical barrier is the mass per unit area of its component layers. For best insulating performance, the heavy layer of a mass spring will often have a smooth high-density surface to maximise reflection of noise waves, a non-porous structure and a certain material stiffness to minimize vibration. From this viewpoint, it is known that many textile fabrics, either thin and/or porous in structure, are not ideal for noise insulation.
JP 2001310672 discloses a multi-layer structure consisting of two absorbing layers with a sound reflecting film layer in between. The film layer reflects the sound penetrating the top absorbing layer back to the same layer, thereby increasing the absorbing effect of the multi-layer structure. The system can be tuned by optimising the film’s thickness and density.
JP 2001347899 discloses a classic mass-spring system with an additional absorbing layer on top of the mass layer. Thanks to the increase in noise attenuation guaranteed by the additional absorbing layer, the thickness and/or the density of the mass layer can be reduced.
EP 1428656 discloses a multi layer structure consisting of a foam layer and a fibrous layer with a film in between both layers. The fibrous layer, made with compressed felt, functions as an absorbing layer with an air flow resistance (AFR) of between 500 and 2500 (Nsm^-3)and an area mass of between 200 and 1600 g/m². The disclosed foam layer has a low compression force deflection with stiffness between 100 and 100000 (Pa), comparable with the stiffness of a felt layer normally used as a decoupler. The film used is preferably perforated or so thin that it does not have an impact on the absorption of both absorbing layers together. The film is called acoustically transparent to indicate that the sound waves can pass the film. For this purpose the film’s thickness disclosed is in the range of 0.01 (mm) or less.
Normally, to reduce the sound pressure level in the passengers’ compartment, a vehicle requires a good balance of the insulation and absorption provided by the acoustical trim parts. The different parts can have different functions (for example, insulation may be provided on the dash inner, absorption may be provided on the carpet). There is a current trend, however to achieve a more refined subdivision of the acoustical functions on the single areas, to optimise the global acoustical performance. As an example, a dash inner can be split in two parts, one providing high absorption and another providing high insulation. Generally, the lower part of the dash is more suitable for insulation, because the noise coming from the engine and the front wheels through this lower area is more relevant, while the upper part of the dash is more suitable for absorption, because some insulation is already provided by other elements of the car, for instance the instrumentation panel. In addition, the backside of the instrumentation panel will reflect sound waves coming through the part of the upper dash hidden behind the instrumentation panel itself. These reflected sound waves could be effectively eliminated using absorbing material. Similar considerations can be applied to other acoustical parts of the car. For instance the flooring: insulation is predominantly of use in the foot well areas and in the tunnel area, while absorption is predominantly of use underneath the front seat and in the rear floor panels.
For the above reasons, vehicle manufacturers typically make use of patches i.e. locally applied additional material ( US 2004150128 ). For instance US 5922265 discloses a method of applying additional heavy layer material in specified areas of a trim part, while the areas without the heavy layer material will act as absorber. These hybrid type of products have the disadvantage, that they still increase the area weight to obtain a combined noise absorbing and insulating solution. They can also be labour and cost intensive. In addition, a material used as a decoupler for an acoustic mass-spring system is generally not optimal for use as an absorber. Furthermore the use of different types of materials makes recycling of the parts and of the discarded material more difficult.

Disclosure of Invention

It is therefore object of the invention to obtain an sound insulating trim part, which works over the range of frequencies important for noise reduction in a vehicle, without the drawbacks of the state of the art, in particular obtaining an alternative solution to the classical mass layers made of highly filled dense materials, like EPDM, EVA, PU, PP used today in acoustic mass-spring systems.
This object is achieved by the trim part according to claim 1. By a sound insulating trim part divided in areas with at least one area with predominantly sound absorbing characteristics (absorbing area), whereby the absorbing area comprises at least one porous fibrous layer, and at least one other area with acoustic mass-spring characteristics (insulating area), whereby the insulating area consists of at least a mass layer and a decoupling layer, the different local requirements can be covered. However, by using in the insulation area a mass layer consisting of a thin impervious barrier layer and of the same porous fibrous layer as the absorbing area, but adjusted to a dynamic Young’s modulus (Pa) of at least (96·AW·t) with AW being an area weight (g/m²), and t being a thickness (mm) of the fibrous layer, and whereby the barrier layer is between the porous fibrous layer and the decoupling layer and all layers are laminated together, the parts become less complex. The same porous fibrous layer is used for both areas, whereby the thickness of the porous fibrous layer in the absorbing area is larger than the thickness of the same layer in the insulating area.
High transmission loss is expected for a mass-spring system, where the mass layer consists of a traditional heavy layer, which is impervious. With impervious is meant air impermeable. Unexpectedly, it was found that it is possible to create a mass layer for a mass-spring system by means of a porous fibrous material on top of a thin air impervious barrier layer using the same fibrous layer as normally used for sound absorption. This enables the use of one porous fibrous layer for both insulation and absorption by adjusting the layer locally in those areas where predominantly insulation characteristics are of advantage. However to obtain a satisfying insulation, it is necessary that the dynamic Young’s modulus of the porous fibrous material is at least: 96·AW·t (Pa) to obtain a radiation frequency of such porous fibrous layer of at least 4900 (Hz), thus obtaining a good insulation performance over all the frequency range of interest, without a disturbing frequency dip in the sound TL spectrum.
The resonance frequency of the mass-spring system as described in the introduction and the radiation frequency of the fibrous top layer as described in the invention result in different and independent effects on the IL curve. Both appear in the IL curve of a multilayer according to the invention and produce a negative effect on the insulation performance, both causing the presence of a dip in the IL curve. But two dips are normally observed in two separate sections of the IL curve. For the considered types of multilayers, the mass spring resonance frequency is normally observed in the range of 200 to 500 Hz, while the porous fibrous layer’s radiation frequency is in the range above 1000Hz. For clarity it is chosen to use two different terms to distinguish between the two different frequencies.
The trim part according to the invention is based on the idea that both insulating and absorbing areas are needed to fine-tune the sound attenuation in a car. By using the same porous fibrous layer throughout the whole area of the trim part for both the insulating area and the absorbing area, it is possible to integrate both functions in a trim part, preferably in separate areas. The skilled person knows from experience which areas need what type of insulation, he is now able to supply parts using this knowledge and at the same time using less types of material and he is able to design the part according to the needs. A trim part according to the invention has at least one absorbing area and one insulating area, however the actual number of areas per each acoustic function (insulation or absorption) and/or the sizes of the areas can differ depending on the part and the location the part is used and last but not least dependent on the actual requirements.
An absorbing area is defined as an area of the trim part that behaves predominantly as absorber and shows bad insulation performance.
An insulating area is defined as an area on the trim part that behaves at least as a good insulator.

Porous fibrous layer

The use of porous fibrous materials, like felt or nonwoven, for the construction of acoustic absorbing parts is known. The thicker the fibrous layer is the better the acoustic absorption. The use of this type of material in a mass spring system to obtain a mass layer is not known in the art.
It was found that the dynamic Young’s modulus is related to the radiation frequency of the porous fibrous layer with E being dynamic Young’s modulus (Pa), ν being radiation frequency (Hz), AW being area weight (kg/m²), and t being thickness (m). According to this relation a proper value of the dynamic Young’s modulus enables the design of a trim part with the radiation frequency outside the frequency range of interest and therefore an undisturbed insertion loss in the frequency range of interest. In particular, if the dynamic Young’s modulus is higher than the minimum value defined as E_min=AW·4·t·ν₀ ², with ν₀=4900Hz, then the radiation frequency of the porous fibrous layer will appear above the frequency range of application of the trim parts. Therefore the dynamic Young’s modulus should be at least 96·AW·t (Pa) with AW (g/m²) and t (mm). This gives a high dynamic Young’s modulus at which the material cannot be compressed easily anymore. The trim part area containing the porous fibrous layer with at least a dynamic Young’s modulus of 96·AW·t (Pa), a decoupling layer and a thin impervious barrier layer, for instance an impervious film layer, between the porous fibrous layer and the decoupling layer, all layers laminated together, will function as an acoustic mass-spring system, ergo as an insulating area. The porous fibrous layer together with the impervious barrier layer is an alternative mass layer and can replace the heavy layer material normally used. The material is cheaper and the overall part is easier to recycle in comparison to mass-spring systems using the classical filled heavy layer materials.
Normally a fibrous material is produced in blanks, i.e. a semi-finished good in which the fibres are assembled together. A blank is at a reasonable approximation homogeneous. A blank is composed by a sheet of material having an initial thickness and is characterized by its area weight, because the fibres are evenly distributed on the area. When a blank is formed, for example by compression, it assumes a final shape. Finally, a layer with a certain thickness is obtained. The area weight, i.e. the weight of the material in the unit area, is maintained after the forming process. From the same blank, several final thicknesses can be obtained, depending on the level of compression.
The Young’s modulus of a fibrous material depends on several parameters. Firstly the characteristics of the material itself, i.e., the material composition, type and amount of fibres, type and amount of binders, etc. In addition for the same fibre recipe, it depends on the density of the material, which is linked to the thickness of the layer. Therefore, for a certain composition of felt, the Young’s modulus can be measured at the different thicknesses and will consequently assume different values, normally increasing when the thickness is decreased (for the same initial blank).
A given porous fibrous layer will be according to the invention if its measured Young’s modulus is higher than the minimum necessary to make it act as a rigid mass on the frequency range that is important for noise attenuation in vehicles, given by the formula 96·AW·t. When this condition is fulfilled, the layer will act, when put on top of a thin impervious barrier layer, as a rigid mass and will have the optimal insulation performance, according to the present invention.
The design of a fibrous porous layer acting as a rigid mass according to the present invention involves therefore the following steps.

A felt composition and an area weight are chosen.
The material is then formed at a certain thickness.
The area weight (AW, g/m²) and the thickness (t, mm) of the formed material are measured.
The Young’s modulus is measured through Elwis-S, for a formed sample at the thickness t (measured Young’s modulus: E_meas).
The minimum necessary Young’s modulus (E_min) is calculated by the formula 96·AW·t, where AW is the area weight (g/m²) and t the thickness (mm), both just measured.
It has to be verified that the condition E_meas > E_min is fulfilled.

If the condition is fulfilled, the choice of the material is satisfactory according to the present invention and the fibrous material can be used at the determined thickness, acting as a rigid insulating mass. Otherwise, the choice has to be changed and re-iterated, restarting from one of the points 1 to 4, where the parameters (felt composition and/or area weight and/or thickness) must be changed.
The porous fibrous layer can be any type of felt. It can be made from any thermo formable fibrous materials, including those derived from natural and/or synthetic fibres. Preferably the felt is made from recycled material like shoddy cotton, or other recycled fibres, like polyester.
The fibrous felt material comprises preferably a binding material, either in binding fibres or in resinous material, for instance thermoplastic polymers. At least 30% Epoxy resin or at least 25% bi-component binder fibres is preferred. Other binding fibres or materials achieving the porous fibrous layer according to the invention are possible and not excluded.
Preferably the area weight is between 500 and 2000 (g/m²), more preferably between 800 and 1600 (g/m²).
An additional restriction is normally also the available space in the car where the acoustical trim part can be put. This restriction is normally given by the carmaker and is in the range of maximum 20-25 mm. All layers of the trim part must share this space. Therefore the thickness of the porous fibrous layer in the insulating area is preferably between 1 and 10 (mm), leaving enough space for the decoupling layer. The thickness of the porous fibrous layer in the absorbing area is basically only restricted by the space available. The thickness can vary throughout the areas and between areas. However the thickness of the porous fibrous layer in the absorbing area is larger than in the insulating area.
The airflow resistance (AFR) of the porous fibrous layer in the absorbing area is preferably between 300 and 3000 (Nsm^-3), preferably between 400 and 1500 (Nsm^-3). A higher AFR is better for absorption. However it decreases with increasing thickness therefore the AFR is preferably between 400 and 1500 (Nsm^-3) for a thickness of between 8 and 12 (mm). Adding additional absorbing layers can further enhance the absorption; either locally on the absorbing areas or as an additional layer on basically the whole trim part. In the insulating area this will effectively form a combined absorption and insulating area. The additional layers can be in the form of felt material similar or the same as used for the porous fibrous layer and/or additional scrim layers.
Next to the absorbing areas and the insulating areas also intermediate areas will exist, that form the areas between an insulating area and an absorbing area or around the rim of the part. These areas are less easy to identify as absorbing area or insulating area mainly due to process conditions creating a type of intermediate zones with changing thickness, increasing in the direction of the absorbing zone and therefore behaving between a good absorber and a not so bad insulator.
Another type of intermediate areas can exist locally to follow the 3-dimensional shape of the part that has to match with the space available in the car. In the state of the art, highly compressed areas exist around the holes in the trim part that are needed for throughput of cables or mounting fixtures. These latter areas are normally not dedicated to acoustic insulation as the acoustic weakness of the holes compromises any insulating characteristic in their close vicinity.

Thin barrier layer

At least the insulating area(s) must contain a thin barrier layer. This thin barrier layer, situated between the porous fibrous layer and the decoupling layer, must be air impervious, however it does not have in itself the function of the mass element of the mass-spring system, like the heavy layer barriers normally found in classic mass-spring systems. Such function is accomplished only by the combination of the porous fibrous layer and of the thin barrier layer. Only if the thin barrier layer is air impervious, the porous fibrous layer according to the invention together with the thin barrier layer, will function according to the invention as a mass layer for a classic mass-spring system. Although a film is given in the examples alternative air impervious thin materials can be used.
If a film is used as a thin barrier layer, it preferably has a thickness of at least 40 (μm), preferably around 60 to 80 (μm). Although thicker films will work, they will not really add to the function and only to the price of the part. Furthermore thicker films might interfere with the forming of the felt.
The thin barrier layer, in particular a film, can be made from thermoplastic material like PVOH, PET, EVA, PE, or PP or dual layer materials like a PE/PA foil laminate. The choice of the barrier material is dependent on the porous fibrous layer and on the decoupling layer and should be able to form a laminate binding all layers together. Also materials that are used as an adhesive either as film or powder can be used. However after the binding and/or forming of the trim part, the formed barrier layer should be impervious to air in the final product.
The thin barrier layer should not necessarily be present also in the absorbing areas and/or intermediate areas, however for ease of production it is recommended.

Decoupling layer

As a decoupling layer, the standard material used for the spring layer in a classic acoustic mass-spring system can be used in at least the insulating area of the trim part according to the invention following the same principles. The layer may be formed from any type of thermoplastic and thermosetting foam, closed or open, e.g. polyurethane foam. It can also be made from fibrous materials, e.g. thermo formable fibrous materials, including those derived from natural and/or synthetic fibres. The decoupling layer has preferably a very low compression stiffness of less than 100 (kPa). Preferably the decoupling layer is also porous or open pored to enhance the spring effect. In principle the decoupling layer should be attached to the film layer over the entire surface of the part in the insulating areas to have the most optimised effect, however due to the production technique very locally this might not be the case. As the insulating area of the part should function overall as an acoustical mass-spring system, small local areas were the layers are not coupled will not impair the overall attenuation effect.
The thickness of the decoupling layer can be optimised, however it is mostly depending on space restrictions in the car. Preferably the thickness can be varied over the area of the part to follow the available space in the car. Normally the thickness is between 1 and 100 (mm), in most areas between 5 and 20 (mm).

Additional layers

The trim part according to the invention comprises at least 3 layers in the insulating area and at least one layer – the porous fibrous layer – in the absorbing area, whereby the at least one layer of the absorbing area is a shared layer. To function optimal as a spring mass system the at least 3 layers of the insulating area are laminated together. However it is possible to optimise the part further by adding on the porous fibrous layer an additional layer with absorbing qualities, either locally on certain type of areas, partially or wholly on the trim part. The area weight of the additional layer is preferably between 500 and 2000 (g/m²).
The absorbing layer may be formed from any type of thermoplastic and thermosetting foam, e.g. polyurethane foam. However for the purpose of absorbing noise the foam must be open pored and/or porous to enable the entrance of sound waves according to the principles of sound absorption, as known in the art. The absorbing layer can also be made from fibrous materials, e.g. thermo formable fibrous materials, including those derived from natural and/or synthetic fibres. It can be made of the same type of material as the fibrous porous mass layer but loftier. The airflow resistance (AFR) of the absorbing layer is preferably at least 500 (Nsm^-3), more preferable between 500 and 2500 (Nsm^-3). Preferably the AFR differs from the AFR of the porous fibrous layer.
Also an additional scrim can be put on top of either the absorbing material or on the porous fibrous layer directly to enhance even further the acoustic absorption and/or to protect the underlying layers, for instance against water etc. A scrim is a thin nonwoven with a thickness between 0.1 and around 1 (mm), preferably between 0.25 and 0.5 (mm) and an increased airflow resistance. It has preferably an airflow resistance (AFR) of between 500 and 3000 (Nsm^-3), more preferably of between 1000 and 1500 (Nsm^-3). Whereby the scrim and the underlying absorbing layer preferably differ in AFR, to obtain an increased absorption.
The area weight of the scrim layer can be between 50 and 250 (g/m²), preferably between 80 and 150 (g/m²).
The scrims can be made from continuous or staple fibres or fibre mixtures. The fibres can be made by meltblown or spunbond technologies. They can also be mixed with natural fibres. The scrims are for example made of polyester, or polyolefin fibres or a combination of fibres for instance of polyester and cellulose, or polyamide and polyethylene, or polypropylene and polyethylene.
These and other characteristics of the invention will be clear from the following description of preferential forms, given as non-restrictive examples with reference to the attached drawings.

Production method

The trim part according to the invention can be produced with cold and/or hot moulding methods commonly known in the art. For instance the porous fibrous layer with or without the thin barrier layer can be formed to obtain the wanted dynamic Young’s modulus in the insulating area(s) and at the same time to form the part in the 3-dimensional shape needed, and in a second step the decoupling layer can be either injection moulded or a foam or fibre layer can be added to the backside of the thin barrier layer at least in the insulating areas.

Definition of mechanical and compression stiffness and measurement

Mechanical stiffness is linked to the reaction that a material (a layer of material) offers to an external stress excitation. Compression stiffness is related to a compression excitation and bending stiffness is related to a bending excitation. The bending stiffness relates the applied bending moment to the resulting deflection. On the other hand, the compression or normal stiffness relates the applied normal force to the resulting strain. For a homogeneous plate made with an isotropic material, it is the product of the elastic modulus E of the material and the surface A of the plate.
For a plate made with an isotropic material both compression and bending stiffness relate directly to the material’s Young’s modulus and it is possible to calculate one from the other. However, if the material is not isotropic, as it is the case for most felts, the relationships just explained no longer apply, because bending stiffness is linked mainly to the in-plane material’s Young’s modulus, while compression stiffness is linked mainly to the out-of-plane Young’s modulus. Therefore, it is not possible any more to calculate one from the other. In addition, both compression stiffness and bending stiffness can be measured in static or dynamic conditions and are in principle different in static and dynamic conditions.
The radiation of a layer of material is originated from the vibrations of the layer orthogonal to its plane and is mainly linked to the dynamic compression stiffness of the material. The dynamic Young’s modulus of a porous material was measured with the commercially available “Elwis-S” device (Rieter Automotive AG), in which the sample is excited by a compression stress. The measurement using Elwis-S is described in for instance BERTOLINI, et al. Transfer function based method to identify frequency dependent Young's modulus, Poisson's ratio and damping loss factor of poroelastic materials. Symposium on acoustics of poro-elastic materials (SAPEM), Bradford, Dec. 2008.
As these types of measurements are not generally used yet for porous materials, there exist no official NEN or ISO norm. However other similar measurement systems are known and used, based on similar physical principles, as described in detail in: LANGLOIS, et al. Polynomial relations for quasi-static mechanical characterization of isotropic poroelastic materials. J. Acoustical Soc. Am. 2001, vol.10, no.6, p.3032-3040.
A direct correlation of a Young’s modulus measured with a static method and a Young’s modulus measured with a dynamic method, is not straightforward and in most of the cases meaningless, because the dynamic Young’s modulus is measured in the frequency domain over a predefined frequency range (for example 300-600 Hz) and the static value of the Young’s modulus corresponds to the limit-case of 0 (Hz), which is not directly obtainable from dynamic measurements.
For the current invention the compression stiffness is important and not the mechanical stiffness normally used in the state of the art.

Other measurements

Airflow resistance was measured according to ISO9053.
The area weight and thickness were measured using standard methods known in the art.
The transmission loss ( TL ) of a structure is a measure of its sound insulation. It is defined as the ratio, expressed in decibels, of the acoustic power incident on the structure and the acoustic power transmitted by the structure to the receiving side. In the case of an automotive structure equipped with an acoustical part, transmission loss is not only due to the presence of the part, but also to the steel structure on which the part is mounted. Since it is important to evaluate the sound insulation capabilities of an automotive acoustical part independently from the steel structure on which it is mounted, the insertion loss is introduced. The insertion loss ( IL ) of an acoustical part mounted on a structure is defined as the difference between the transmission loss of the structure equipped with the acoustical part and the transmission loss of the structure alone:
The insertion loss and the absorption coefficient were simulated using SISAB, a numerical simulation software for the calculation of the acoustical performance of acoustical parts, based on the transfer matrix method. The transfer matrix method is a method for simulating sound propagation in layered media and is described in for instance BROUARD B., et al. A general method for modelling sound propagation in layered media. Journal of Sound and Vibration. 1995, vol.193, no.1, p.129-142.

Brief Description of Drawings

Figure 1 Insertion loss of samples A-C
Figure 2 Example of an inner dash trim part with regions of sound insulation and regions of sound absorption.
Figure 3 Schematic of the multilayer for either the insulating area or the absorbing area according to the invention
Figure 4 Schematic of an alternative multilayer for either the insulating area or the absorbing area according to the invention
Figure 5 Insertion loss of multilayer according to figure 3 or 4
Figure 6 Absorption of multilayer according to figure 3 or 4
Figure 7 Schematic of an alternative multilayer for either the insulating area or the absorbing area according to the invention
Figure 8 Insertion loss of multilayer according to figure 7
Figure 9 Absorption of multilayer according to figure 7
Figure 10 Graph of the dynamic E modulus in relation to the area weight and the thickness of the porous fibrous layer.
Figure 11 Graph comparison of the insertion loss of different samples.

Figures

Figure 1 shows the insertion loss curves of the comparative samples A-B and sample C. The simulated insertion loss shown is the transmission loss of the system constituted by the multilayer and the steel plate on which it is applied minus the transmission loss of the steel plate itself.
The insertion loss and the sound absorption of different noise attenuation multilayer constructions of the state of the art were simulated using measured material parameters and compared with the insertion loss and sound absorption of a noise attenuation multilayer according to the invention. All samples have the same total thickness of 25 (mm).
Comparative sample A is a classic mass-spring systems with the mass layer formed by an EPDM heavy layer material of 1 (kg/m²) and injected foam as the decoupling layer. The total area weight of sample A was 2370 (g/m²).
Comparative sample B is made according to the principles of EP 1428656 which discloses a multilayer structure consisting of a foam decoupling layer and a top fibrous layer with a film in between both layers. The top fibrous layer is an air-laid soft felt layer with an area weight of 1000 (g/m²), a thickness of 6 (mm) and an AFR of 1000 (Nsm^-3). The total area weight of this multilayer is 2150 (g/m²). The dynamic Young’s modulus of the fibrous layer was measured and is around 70000 (Pa). According to the equation given, this fibrous layer will have a radiation frequency in the area of around 1700 (Hz). The film used is 0.06 (mm) and impervious. The decoupling layer is injected foam with an area weight of 1100 (g/m²).
Sample C was made according to the invention, and contains the same decoupling layer and film layer as comparative sample B. The porous fibrous layer on top of the film layer was made of a compressed rigid felt layer with an area weight of 900 (g/m²), a thickness of 3 (mm) and a dynamic Young’s modulus of 550000 (Pa). According to the equation given, this porous fibrous layer will have a radiation frequency in the area of around 7100 (Hz).
Sample A is a classical mass-spring system with an area weight for the heavy layer of 1 (kg/m²). The insulating performance is high over a large range of frequencies and therefore this sample represents the preferred system to use for noise attenuation in a car, however the system is much too heavy. Furthermore, the material normally used for heavy layer, in this case EPDM, is difficult to recycle. In terms of overall noise attenuation, the classical mass-spring system – A – is still superior, since in comparative sample B the top felt layer has a radiation frequency of about 1700 (Hz) that compromises the insulation characteristics of the multilayer. This is made visible in figure 1 in the IL curve of comparative sample B by a dip in the 1/3 octave frequency band centred at 1600 (Hz), which is the frequency band including the radiation frequency of the top felt layer used for this sample.
It was now found that by increasing the dynamic stiffness of the material constituting the porous fibrous layer, in particular by increasing its compression stiffness in the layer’s out-of-plane direction, the radiation frequency of the layer can be shifted to a higher frequency.
By choosing the dynamic Young’s modulus of the fibrous material constituting the porous fibrous layer in the insulating area in such a way that the radiation frequency of this area lies outside of the frequency range where noise needs to be attenuated, the layer will behave, when put on top of a thin barrier layer, as a mass-spring system over the wanted frequency range.
Sample C for instance has a porous fibrous layer on top of the film layer made with a compressed rigid felt layer with an area weight of 900 (g/m²), a thickness of 3 (mm) and a dynamic Young’s modulus of 550000 (Pa). It shows an insertion loss comparable and even better than that of comparative sample A, the classic mass-spring system with (1 kg/m²) heavy layer. And the radiation frequency only appears as a dip in the insertion loss curve in the 1/3 octave frequency band centred at 6300 (Hz). This is far above the frequency range normally considered interesting for noise attenuation in a vehicle.
The effect, that a thin barrier layer coupled with a porous fibrous layer having a dynamic Young’s modulus of at least 96·AW·t (Pa) can form the mass layer of a classical acoustic mass-spring system is not only dependent on the compression of the porous fibrous layer. It can also be dependent on the type of material used for such porous fibrous layer and on the amount of binding between the material components, for instance between the fibres or the resin and fibres. The equation gives therefore only guidance to how to design a trim part according to the invention. The actual frequency where the porous fibrous layer’s radiation frequency in reality occurs can deviate from the calculated one, however as long as it appears above at least 4900 (Hz) it will not interfere any more with the noise attenuation necessary and mostly wanted in vehicles. For other applications the minimal dynamic Young’s modulus needed might differ, however a skilled person will be able to adjust the equation following the invention guidance.
All optimisations of sound attenuation of trim parts, as given in the state of the art, are directed to defining the airflow resistance of at least the absorbing layers. It was found that for the trim part according to the invention, the radiation in general and the radiation frequency in particularly of the upper porous fibrous layer does not depend strongly on its airflow resistance. The airflow resistance was found to have mainly a damping influence on the slope of transmission loss over the whole frequency measured. The damping effect is larger with increased airflow resistance.
Figure 2 shows an example of an inner dash part with two separate areas having different acoustic functions, with the aim of obtaining an optimised compromise of insulation and absorption. Generally, the lower part of an inner dash part is more suitable for insulation (I), because the noise paths coming from the engine and the front wheels through this lower area are more relevant, while the upper part of the dash (II) is more suitable for absorption, because some insulation is already provided by other elements of the car, for example the instrumentation panel. Between these areas, in areas where the packaging space is minimal or in heavily 3D-shaped areas, it is normally not possible to identify the actual acoustical characteristics for instance due to either impairing of the decoupling layer or compression of a lofty layer that should function as an absorbing layer. Inner dashes where well-defined areas with different sound attenuation characteristics (insulation or absorption) are needed, as shown here, are normally made as two parts instead of one. Another option is to put additional heavy-layer pads on top of an absorbing material, generally foam, to obtain a local mass-spring system.
To achieve an overall better sound attenuation for an inner dash trim part the whole part can be built with different distinctive areas: the insulating area (I) can be formed by combining a thin impervious barrier layer and a porous fibrous layer with adjusted dynamic Young’s modulus to form together the alternative mass layer according to the invention, and the absorbing area (II) can be formed by the same porous fibrous layer not adjusted for insulation, ergo not rigidified. Thus area I of the trim part in the inner dash shown would contain the alternative mass-spring system according to the invention. Area II would contain the non-rigidified porous fibrous layer functioning as a standard absorber known in the art.
Figure 3 shows a schematic cross section of a multilayer according to the invention. The multilayer according to the invention contains at least an area with sound insulating characteristics (I), afterwards called insulating area, and an area with sound absorbing characteristics (II), afterwards called absorbing area. The location of the areas on the part is depending on the area of the vehicle where the part is used and on the expected noise levels and frequency characteristics in that specific area. (See as an example the inner dash previously described.)
The insulating area (I) and the absorbing area (II) have at least the same porous fibrous layer (1), whereby this layer in the insulating area is compressed to form a rigid layer (1), such that the dynamic Young’s modulus of this porous fibrous layer is at least 96·AW·t (Pa), with AW area weight (g/m²), and t thickness (mm) of the porous fibrous layer.
The insulation characteristic is formed with a mass layer A consisting of the thin barrier layer 2 and the porous fibrous layer 1, according to the invention, and with a spring layer B consisting of a decoupling layer (3), together forming an acoustic mass-spring system. In Area I predominantly a sound-insulating characteristic can be expected accordingly.
In area II the porous fibrous layer 1 is not compressed but kept lofty enabling sound absorbing characteristics in this area. Preferably an additional scrim layer 4 can be put on top of the absorbing layer 4 to enhance the sound absorbing effect even further.
Figure 4 shows an alternative multilayer according to the invention, based on the same principles as in Figure 3 (see there for reference.) The difference is that the area underneath the compaction is used for the addition of the thin barrier layer and the decoupling layer, producing a more even part. In praxis the part will be more a crossover between figure 3 and 4, in particularly the shape of automotive trim parts is normally a 3D form and this will influence the final layout of the layering as well. Also between the insulating area and the absorbing area there will not be clear-cut boundaries, rather intermediate areas.
The insertion loss and absorption curves were simulated for multilayer construction according to figure 3 or 4, without the scrim layer, and they are shown, in figure 5 and figure 6 with following features for the different layers.
The absorbing area (II) consists of a porous fibrous layer in the form of a cotton based felt with 30% Epoxy binder, with a thickness of 20 mm and an area weight of 1100 (g/m²). The absorption and insulation simulated is indicated with ABS for the absorbing area.
The insulation area contains the same porous fibrous layer compacted to 2.7 (mm), to comply with the dynamic Young’s modulus requirements according to the invention, a film layer and a foam decoupling layer. The total thickness of the insulation layer is 20 mm. The simulated absorption and insulation is indicated with INS for the insulation area.
From the insertion loss (figure 5), it is clear that the resonance dip only appears around 6300 (Hz) and therefore falls in the scope of the claim. Also an increase in overall insertion loss can be observed showing the improvement in insulation characteristics for a part with such an insulation area. For the absorbing area there is no material to act as a spring and in this case the insertion loss curve shows values that are close to zero over all the frequency range.
From the absorption curve (figure 6), it is clear that the absorption of the insulation area is poor, and the absorption of the absorbing area is much better as expected. From an absorbing point of view a pure absorber like simulated here for the absorbing area will work best.
Having a trim part sharing both insulating areas and absorbing areas would however improve the overall sound attenuating performance of the part, in particularly because the dedication of an area to a certain type of sound attenuation (insulation or absorption) on one and the same trim part can be defined in advance and can be formed in a part locally without the need of additional patches or other materials. This means also that large trim parts like carpets or inner dash insulators now can be made in one piece even if different sound characteristics are required on different part areas.
Figure 7 shows an alternative layering according to the invention where film and decoupler are available over the whole surface of the part, including the absorbing area(s). However this changes the behaviour of the part, in particular the insulation properties as can be seen from figure 8 and figure 9. In this example a multilayer with a total thickness of 25 mm, where the porous fibrous layer is made with a cotton felt with 30% Epoxy resin as a binder and has an area weight of 1100 (g/m²) is considered. The thickness of such porous fibrous layer is adjusted to be 2.7 (mm) in the insulation area and 17.3 (mm) in the absorbing area and underneath the porous fibrous layer a impervious film and a foam decoupler are considered (both in the absorbing and in the insulating area), being the thickness of the foam adjusted in such a way to obtain a total thickness of 25 (mm) for both areas. The absorbing area will now start working as a mass-spring system, however its insertion loss curve will show an insulation dip around 1000 and 1600 (Hz), disturbing the noise attenuation in the wanted range of frequencies. Whereas the dip for the insulating area is around 6300 (Hz), comparable with our previous sample. For the same samples however the absorption curve found for the absorbing area is slightly impaired due to the own radiation frequency. Although this solution might have some acoustical disadvantages in comparison with the previous layout it would still have advantage in comparison with the current state of the art. The use of the foam layer over the whole area of the trim part would help to form a smoother part in comparison with only locally applied foam, in particularly if the foam is injected in a mould.
Other alternative solutions might be to have the thin barrier layer only in the insulating areas, but the decoupler over the whole surface so that in the areas without thin barrier layer the decoupler would function together with the porous fibrous layer as a double layer absorber.
To enhance the overall function of either the insulating area and/or the absorbing area additional absorbing material can be put on top of the porous fibrous layer. Also the use of a nonwoven (4) will enhance the absorbing characteristics of the trim part.
The effect, that a thin barrier layer together with a porous fibres top layer with a dynamic Young’s modulus of at least 96·AW·t (Pa) can form a mass layer with characteristics comparable with a classical acoustic mass-spring system is not only dependent on the compression of the felt. It can also be dependent on the type of material used and the amount of binding between the material components, for instance between the fibres or the resin and fibres. The equation gives therefore only guidance to how to design a trim part according to the invention. The actual frequency where the radiation frequency in reality occurs can deviate from the calculated one, however as long as it appears above at least 4900 (Hz) it will not interfere any more with the noise attenuation necessary and mostly wanted in vehicles. For other applications the minimal dynamic Young’s modulus needed might differ, however a skilled person will be able to adjust the equation following the invention guidance.
All optimisations of sound attenuation of trim parts, as given in the state of the art, are directed to defining the airflow resistance of at least the upper layer or the absorbing layers. It was found that for the trim part according to the invention, the radiation frequency of the upper porous fibrous layer does not depend strongly on its airflow resistance. The airflow resistance was found to have mainly a damping influence on the slope of the insertion loss curve over the whole frequency measured. The damping effect is larger with increased airflow resistance.
In the following, an example of how a skilled person can use the equation to design a trim part according to the invention is given. Figure 10 shows a graph of dynamic Young’s modulus vs. thickness for the insulating mass layer according to the invention. In this case a felt layer made primarily of recycled cotton with 30% phenolic resin was taken. This material was used until not long ago as decoupler or absorbing layer, mainly in multilayer configurations. Nowadays the phenol binder is no longer applicable in interior parts for vehicles due to regulations on vapours in the car interior. The material however can still be used in car external parts, in the engine bay area or in trucks. It is here not chosen as a restrictive sample but more as an example to show how to design the material according to the invention.
In figure 10, line L1000gsm shows, as a function of the layer’s thickness, the minimum dynamic Young’s modulus that a porous fibrous layer with an area weight of 1000 (g/m²) must have to be according to the invention. This was calculated with the formula with v is 4900 Hz and it is shown then in Figure 5 as a straight line. Lines L1200gsm, L1400gsm and L1600gsm in the same figure show similar data for the area weights of 1200, 1400 and 1600 (g/m²). The dynamic Young’s modulus of a porous fibrous layer with a given thickness and one of these area weights should be above the line corresponding to its area weight, to make sure that the layer’s radiation frequency is shifted to at least 4900Hz and thus outside of the frequency range of primary interest for noise attenuation in vehicles.
In Figure 10, line A1000gsm shows, as a function of the layer’s thickness, the measured dynamic Young’s modulus of a layer of primarily cotton felt with 30% phenolic resin having an area weight of 1000 (g/m²). In the same figure lines A1200gsm, A1600gsm show similar data for the area weights of 1200 (g/m²) and 1600 (g/m²) respectively. For certain points the dynamic Young’s modulus was measured and the behaviour as depicted was extrapolated from these measurements. This material shows a quick increase in the dynamic Young’s modulus already showing a radiation frequency above 4900 (Hz) at an area weight of 1000 (g/m²) and a thickness of around 8 (mm). However due to space restrictions this thickness would not be preferred in the interior of a car for instance for an inner dash. Although in theory it would be possible to come to the right dynamic Young’s modulus with much lower densities, the weight of the porous fibrous layer trim part would no longer be enough to guarantee that the part can function as a good insulating part.
In figure 10, line B1200gsm shows, as a function of the layer’s thickness, the dynamic Young’s modulus of a layer of primarily cotton felt material with 30% epoxy resin and an area weight of 1200 (g/m²). Line B1600gsm shows similar data for the case of the area weight of 1600 (g/m²). For certain points the dynamic Young’s modulus was measured and the behaviour as depicted was extrapolated from these measurements. If one compares these data with those for phenolic resin felt discussed before, it is clearly visible that the binding material has an effect on the compression stiffness of the material and hence on the dynamic Young’s modulus at a certain area weight and thickness.
Line C1400gsm shows, as a function of the layer’s thickness the dynamic Young’s modulus of a layer of primarily cotton felt material bound with 15% bi-component binding fibres and having an area weight of 1400 (g/m²). For certain points the dynamic Young’s modulus was measured and the behaviour as depicted was extrapolated from these measurements.
In a second set of samples, the influence of binder material, in particular the type and amount of binder is looked at in more detail.
Sample EPOXY30% of cotton felt with 30% Epoxy resin with a measured area weight of 1090 (g/m²) and a thickness of 2.7 (mm) was found having a measured dynamic Young’s modulus of 5.55E5 (Pa), thus higher than the required Young’s modulus as calculated according to the invention.
Sample EPOXY20% of cotton felt with 20% Epoxy resin with a measured area weight of 1450 (g/m²) and a thickness of 4 (mm) was found having a measured dynamic Young’s modulus of 2.2E5 (Pa), thus much lower than the required Young’s modulus as calculated according to the invention.
Sample BICO25% of cotton felt with 25% bi-component binding fibres with a measured area weight of 1040 (g/m²) and a thickness of 2.1 (mm), was measured having a dynamic Young’s modulus of 5.08E5 (Pa), thus much higher than the required Young’s modulus as calculated according to the invention.
Sample BICO15% of cotton felt with 15% bi-component binding fibres with a measured area weight of 1280 (g/m²) and a thickness of 4 (mm) was found having a dynamic Young’s modulus of 9.91E4 (Pa) thus much lower than the required Young’s modulus as calculated according to the invention.
For these samples in addition the insertion loss was simulated. Figure 11 shows the simulated insertion loss of the samples comparing 25 (mm) thick samples with top layer as defined, a 70 (μm) film and the remaining thickness covered with foam as the decoupler.
The insulation curve of Sample A, that is the classical mass-spring system with an area weight for the heavy layer of 1 (kg/m²) earlier introduced is also here used as reference sample.
The measured and calculated radiation frequencies for the samples’ top porous fibrous layers are appearing as a dip D in the IL curves. For the samples EPOXY25% and BICO15% the radiation frequency found was at 3150 (Hz) (D2) and 1600 (Hz) (D1), both in the area of interest for the attenuation of sound in a car. While the radiation frequencies of EPOXY30% and BICO20% were found both at around 6300 (Hz) (D3 and D4), outside the area of interest for the automotive industry.
Surprisingly, an insulation effect is obtained, that is not strongly related to the AFR of the top layer. On the other hand, it was found that the driving factor to obtain a consistent insulation without any dip effect in the range of frequency of interest for instance for automotive applications, is the Young’s modulus of the top layer according to the invention.
When the thickness of the upper layer is changed, both the AFR and the Young’s modulus change and, in general, both the AFR and the Young’s modulus are increasing when the thickness of the layer is decreased. However, the value of each of those parameters is related to the characteristics of the material. The AFR and the Young’s modulus, as well as other acoustical and mechanical parameters of a porous material, are not only a function of the thickness.
As an example, the AFR of two comparable felt materials with the same thickness are compared. An “air laid” felt normally used for automotive application with an area weight of 1000g/m² shows an AFR of 3200 Nsm^-3at approximately 2.5mm. The same material at a thickness of 6 mm shows an AFR of 1050 Nsm^-3. In comparison a “needled” felt normally used for automotive applications, having approximately the same area weight of 1000g/m² shows an AFR 220 Nsm^-3 at approximately 6mm. At the same thickness, the two materials have different AFR. The two felts mainly differ in the way the fibres are processed to form a layer of material and this has an impact on the AFR.
The same consideration applies for the Young’s modulus: for every material, the Young’s modulus is increasing when the thickness is decreasing, however two different materials at the same thickness do not necessarily have the same value of the Young’s modulus and can be characterized by very different Young’s moduli, depending mainly on their composition and on the way they are produced.
Moreover, the AFR and the Young’s modulus are independent parameters, the first being linked to the acoustical characteristics of the material and the second one being linked to the mechanical characteristics of the material. As an example, two materials with the same AFR (linked, for example, to a similar distribution of the fibres in the materials) can have a different Young’s modulus (linked, for example, to a different amount of binders in the material) and therefore a different performance. (See for example figures 10 and 11).
As can also be seen from the materials depicted certain materials are not suitable to form the mass layer according to the invention, basically because they must be compressed to a thickness no longer possibly achievable or at a cost of extreme high pressure forces, making the process no longer cost effective. However by adjusting the ratio of binding material vs. fibrous material, the binding material used, and the area weight and/or thickness it is possible to design materials suitable to be used as a porous fibrous mass layer according to the invention.
The sound insulating trim part according to the invention, with some areas dedicated to absorption and other areas dedicated to insulation, whereby both areas share the same porous fibrous layer can be used in a car for instance as an inner dash as described previously. However it can also be used as a floor covering, possibly with a decorative layer or a carpet layer on top, whereby the carpet layer is preferably a porous system for instance a tufted carpet or a nonwoven carpet. It can also be used in outer or inner wheel liners. All applications can be in vehicles like a car or a truck.

Legend to the figures

I. Insulating area comprising
A Mass layer comprising at least fibrous porous layer and thin barrier layer
B Spring layer

II. Absorbing area
1. Fibrous porous layer
2. Thin barrier layer
3. Decoupling layer
4. Scrim layer

Claims

A sound insulating trim part comprising at least one area with predominantly sound absorbing characteristics (absorbing area), whereby the absorbing area comprises at least one porous fibrous layer, and at least one other area with acoustic mass-spring characteristics (insulating area), whereby the insulating area consists of at least a mass layer and a decoupling layer, characterised in that the mass layer consists of an impervious barrier layer and of the same porous fibrous layer as the absorbing area, in the insulating area adjusted to a dynamic Young’s modulus (Pa) of at least (96·AW·t) with AW being the area weight (g/m²), and t being the thickness (mm) of the porous fibrous layer, and whereby the impervious thin barrier layer is at least between the porous fibrous layer and the decoupling layer and all layers are laminated together, and whereby the thickness of the porous fibrous layer in the absorbing area is larger than the thickness of the same layer in the insulating area.
Sound insulating trim part according to claim 1, wherein the absorbing and/or the insulating areas are more than one area and the thickness of the same type of areas is different among the separate areas.
Sound insulating trim part according to claim 1, wherein the AW of the porous fibrous layer is between 500 and 2000 (g/m²).
Sound insulating trim part according to one of the preceding claims, wherein the thickness t of the porous fibrous layer in the insulating area is between 1 and 10 (mm),
Sound insulating trim part according to one of the preceding claims, wherein the thickness t of the porous fibrous layer in the absorbing area is at least 4, preferably between 4 and 25 (mm).
Sound insulating trim part according to one of the preceding claims, further comprising at least partially an additional absorbing layer on top of the porous fibrous layer.
Sound insulating trim part according to one of the preceding claims further comprising at least partially an additional scrim on top of the porous fibrous layer.
Sound insulating trim part according to one of the preceding claims further comprising at least partially a decorative layer or a carpet layer, preferably a tufted carpet or a nonwoven carpet.
Use of the sound insulating trim part containing absorbing areas or insulating areas or insulating areas with additional absorbing properties, according to one of the preceding claims for an automotive part, like an inner dash, a floor covering or a wheel liner in a vehicle like a car or a truck.