CN1143593C - Acoustic devices relying on bending wave action - Google Patents

Abstract

Acoustic device comprising a member relying on bending wave action with beneficial distribution of resonant modes thereof, wherein the member has its acoustically active area at least partly bounded by means having a substantially restraining nature in relation to bending wave vibration. Operation can be below coincidence, or above if desired for active acoustic device further having beneficial location of bending wave transducer means determined with reference to and taking account of such bounding means.

Description

Acoustic devices relying on bending wave action

The present invention relates to an acoustic device comprising a sound emitting member which generates sound output by means of bending wave action and resultant surface vibrations.

It is known from US Patent No. 3,247,925 to WARNAKA to propose a woofer consisting of an extremely rigid resonant panel, the peripheral edge of which is bolted or glued to a rigid frame which supports conventional voice coil transduction A transducer that transmits the energy of the bending wave to the center of the panel. This woofer unit is said to operate entirely above wave coherent frequencies.

It is also known from US Patent No. 3,596,733 to BERTAGNI to propose a loudspeaker with a diaphragm formed from an extended polystyrene plate-like member having pre-tensioned front and rear surfaces, wherein The surface is or includes irregular shapes.

Instructive teachings of bending wave acting acoustic devices, considered for convenience to be of the general resonant panel type, are presented in International Patent Application WO97/09842, involving the enhancement or optimization of sound performance in terms of panel parameters, including geometric and bending stiffness, especially including usable operation at or below the coherent frequency. Geometric parameters of interest include the proportions or aspect ratios of such panels, including for use as passive acoustic devices. Bending stiffness parameters can effectively interact with geometric parameters, including their anisotropy, that is, along the axis of the geometry, it has different bending stiffness or along the axis of the geometry, the bending stiffness can be decomposed into In substantially consistent directions, these geometric shapes are feasible variations of the proportions of this shape. The preferred in-board positions of the transducers for the active acoustic device effectively have proportional defined coordinates. Other areal distributions of bending stiffness may be useful to provide other useful locations for the transducer, such as substantially at the geometric center or at the center of mass, see International Patent Application WO 98/00621 which includes the The aforementioned bending wave action is combined with an acoustically further related piston action. Acoustic operation is described and claimed at least in WO97/09842 for the entire panel and only a part thereof that is acoustically active.

Based on intuition, the specific analysis and design approach we implement for this resonant mode bending wave action acoustic device focuses on the entire panel where the edges are fully Or substantially free to vibrate, these edges include portions subject to light edge damping. The present invention arose from the surprising and counterintuitive result of further consideration, research and experimentation.

Certain underlying requirements continue to apply and are of profound technical/creative importance, especially for an acoustic device component extending transversely through its thickness and capable of sustaining bending waves through its acoustically active region, which is important for the A fundamental requirement for a component or panel called resonant sound; this component is also used for parameters such as geometrical parameters and for bending stiffness to have a value consistent with the resultant distribution of the natural bending wave vibrations of the component that is of interest The frequency range above which it is effective or beneficial to obtain the desired or acceptable acoustical operation of the device is a further requirement for resonant acoustic components or panels here. Certain embodiments of the present invention also provide a device that generally provides significant confinement of bending wave vibrations at the edge, perimeter, or other boundary of such components or panels or their acoustically active zone, and that is also generally capable of at least coherent Operate on at least a portion of the frequencies below the frequency. The word "substantial restriction" as used herein means a greater restriction of at least a portion of the edge of the part than is specifically disclosed in WO97/09842, preferably a restriction of the edge extent and payload, clamping or effective grounding effect .

It is useful to consider such firm edges/regional borders/surrounding constraints from two perspectives, effect or creative.

One is that the limitation/reduction of the available bending wave vibration edge/around/boundary movement of the component (compared to the specific disclosure in WO97/09842) can generate vibrational bending wave energy from the middle of the acoustically active zone Useful composite for obtaining sound output. Another is that the acoustically relevant and effective natural modes of resonant bending wave action are different (compared to the specific disclosure in WO97/09842) due to the confinement/suppression of bending waves at the edge/surrounding/boundary of the component The vibration moves, thereby effectively reducing and canceling the effect from the lowest resonant mode, whereas if the edge/surrounding/boundary of the acoustically active area of the part is free, with the same bending wave distribution as the specific disclosure in WO97/09842, then the The lowest resonant mode is effective; also due to the reduction/significant suppression of resonant modes involving twisting.

The resulting nominally sparser or less effective/correlated bending wave components of the sound can then be instantiated for simplified simulations or analyzes based on the consideration of the interaction of the relevant resonance plate modes The equivalent simple bundle of these resonant plate modes is associated with each down and starts at the resonant modal frequency f1 instead of f0, and "removes" the combined mode involving the f0 frequency, with interesting and useful effects , this effect can be derived from the flatness of the spacing with respect to the directly and combinatorially related natural resonant modes involving the f1 frequency.

Ramifications are extensive and advantageous, including the availability of improved acoustic efficiency of energy conversion; and/or often very useful augmentation of candidate sub-regions to provide availability of feasible/optimal transducer locations properties, at least as identified by mechanical impedance analysis as taught in co-pending International Patent Application PCT/GB99/00404; and/or, compared to the illustration for isotropic bending stiffness, The available range of areal shapes/proportions of the part is generally much greater, even with shape ratios of about 1:1 up to about 1:3 or more; and/or, inherently low bending stiffness and at least due to Availability of the sound performance of the panel member material which is effectively hardened overall by the effect of /surrounding/boundary limitation; and/or, the performance associated with the high power input transducer arrangement for the loudspeaker embodiment, all including such Constraints can provide significant loads, either based on inertial foundations, or made more practical by actually fixing in stiffer/heavier load bearings or other heavy load means.

An important advantage of the present invention is to provide novel and useful resonant panel acoustic devices, including active acoustic devices as loudspeakers, which can be manufactured very conveniently, as robust and easy-to-install panel-type devices, especially for International Patent Application WO97 /09842 specifically illustrated and described in the acoustic device improved.

According to an aspect of the present invention there is provided an acoustic device which relies on bending wave action and is capable of operation below coherent frequencies, comprising a component which provides said acoustic operation due to the advantageous distribution of the resonant modes of the bending wave action therein , wherein the acoustically active region of the component is bounded at least in part by means having substantially limiting bending wave vibration properties.

According to another aspect of the present invention, there is provided an active acoustic device comprising a component having a beneficial distribution of resonant modes, a beneficial location of a bending wave transducer device and relying on bending wave action, wherein the component is acoustically effective The zone is delimited at least in part by means having substantially limiting bending wave vibration properties, and the position of its transducer means is determined with reference to and taking into account such delimiting means.

The entire circumference of the sound part may be constrained or clamped substantially; or the part, for example a rectangular plate, is constrained or clamped at one or more up to all of its side edges at one or more, but not all, of its circumference. This can be used as a flag mount providing said significant restriction on one side, allowing the sound active area to protrude from there, or as a mount on parallel The active area lies between these mounting restriction sides. And this can facilitate the production of loudspeakers, such as mid/high frequency devices, with fully sealed or only highly selectively open diaphragms. Encapsulating the diaphragm entirely or nearly entirely enables the manufacture of so-called infinite baffle loudspeakers, containing/controlling the rear sound emission which would otherwise be detrimental at mid to low frequencies.

The overall significant restraint and clamping frame also makes the design of speaker assemblies more mechanically predictable, while facilitating the manufacture of speaker assemblies that are structurally strong (with the panel edges being largely free or only compared to a resonant panel loudspeaker suspended in a slightly damped elastic manner).

Significant confinement or clamping of the surrounding parts or edges of the sound part may be achieved in any desired manner, for example by adhesively fastening the edge tightly to a solid frame or the like, or by e.g. clamping the edge to a frame part between mechanical methods. The edge confinement/clamping required here can also be formed by molding techniques (such as injection molding of plastic materials) by forming the edge of the part with a monolithic or integrally thickened surrounding portion of sufficient rigidity to restrain movement of the edge of the sound part to fulfill. Co-molding of sound components and thickened edge settings is suitable. This molding technique is particularly suitable for use where the acoustic component is formed in a single piece and is readily available in an economical manner.

Significant confinement or clamping can also be used to confine one sound part within another larger sound part. Thus a large sound panel intended for mid/low frequency operation can be molded to include a smaller sound panel intended for high frequency operation and defined by a central stiff rib.

Can be designed for significant limiting or clamping action, providing a mechanical termination impedance designed to control reverberation time within the sound component, helping to control the frequency response to the component, perhaps especially at low frequencies.

The proportions of suitable resonant panel components may be as or substantially different from the specific teachings in WO97/09842 in terms of specific shape variations. For example, a substantially rectangular resonant panel member having a substantially isotropic bending stiffness has an aspect ratio below 1:1.5, thus encompassing generally the teachings of the prior art for substantially free-edged panel members, But it is not limited thereto, which will be described in detail later, or the shape ratio is greater than 1:1.5, which will be described in detail later. Anisotropic/complex distribution variations of the bending stiffness, as above, are also conceivable.

The delimiting means may be at least partially around said sound active area and confine, and/or, around the panel form part so as to be sound active as a whole, typically up to 25% of the entire area boundary/surrounding edge extent or a larger extent, usually its entire extent.

Resonant panel components are generally self-supporting and do not require pre-tensioning for mechanical stability, especially of the type used in free-edge forms or simple edge-supported applications.

For clamped panel components, there is a 10-fold or so increase in the first bending frequency due to the natural hardening of the panel components when clamped. It is reasonable and sensitive to reduce the first mode frequency by significantly reducing the bending stiffness performance before the lower frequency range. It is conceivable that in this case the stiffness of the panel member could be as low as 0.001 Nm, with an areal density as low as 25 g/m ² .

From one point of view, these end points of the range of values describe a panel component that can benefit from the application of these tensions for mechanical stability as well as for the functions supported by the drive means. These tensions can be applied uniformly with respect to the effective geometry of the part, or they can be applied differently, ie different directions or different magnitudes of tension.

At the extreme, the tensioned panels exhibit a highly balanced and predominantly second order or non-discrete wave action (velocity is constant with frequency) of the tensioned membrane properties supporting bending waves. For such "panel" components, the resonance distribution can be optimized for the desired acoustic behavior by controlling the tension and boundary.geometry broadly consistent with the distributed mode teaching, see WO97/09842. Similarly, a preferred mode distribution can be further enhanced to act as a preferred/optimized transducer instead of an actuator/sensor.

Depending on the magnitude of the tension and with increasing density and especially the bending stiffness, there is a range in which second-order bending wave actions are superimposed and enhanced by fourth-order, divergent bending actions due to stiffness. Optimization of the two waves can be achieved by calculation and/or experimentation to provide the best results for a given application.

The design field of small bandwidth sound panels with clamped edges is also conceivable.

A practical implementation of the invention is shown diagrammatically by way of illustration, in the accompanying drawings:

1 and 1A are exploded perspective and partial cross-sectional views of a generally rectangular resonant panel acoustic component 10 clamped at its edges to opposing rectangular surrounding frame components 11A, 11A, 12B by nuts, bolts 12A and 12B. Between 11B, the surrounding frame parts 11A, 11B may further be used for mounting to a base or other parent structure.

2 is a partial sectional view showing another edge of clamping/fixing of the resonant panel acoustic component 20 fixed to the frame 21 by an adhesive 22;

Figures 3 and 3A are partial perspective and partial cross-sectional views of a plastic injection molded part 35 formed as a wall member having stiffening ribs 36 intersecting a rectangular border 31 serving as an acoustically active panel area 30, the boundary area 31 is also formed by raised ribs that reinforce the edge of the operator panel area 30;

Figure 4 is a partial perspective view of a resonant panel acoustic component 40 stretched on frame 41 and clamped at its edges by surrounding clamping frame components 42;

Figure 4A is a partial cross-sectional view of the embodiment of Figure 4;

Figure 4B is a partial cross-sectional view similar to Figure 4A of another embodiment of a resonantly extended panel acoustic component;

Figures 5A, 5B are graphs showing the frequency response of various resonant panel components of A4 and A5 sizes, respectively, where the thick line traces represent clamped edge panels and the thin line traces represent free or elastic edge suspended panels.

Figures 6A, 6B and 7A, 7B and 8A and 8B are graphical representations of mechanical impedance versus frequency for selected aspect ratios of panel members of clamped edges;

9A, 9B, 9C are graphical representations of the relative smoothed inverse mean square error for the position of the transducer device;

Figure 10 is a calculated curve of 1/4 panel mechanical impedance of a panel assembly with clamped edges;

Figure 11 is a graphical representation for different aspect ratios of panels with clamped edges;

12A-H are 1/4 panel mechanical impedance curves measured for different aspect ratios;

Figures 13A-H are associated sound output curves for a reference value;

Figure 14 shows the maximum value of the mean square error of the inverse power for different shape ratios;

15A-J are combined polar plots of sound output for low resonance modes with clamped edge panel members having a 1:3 aspect ratio;

16A-E are graphs comparing sound output power for particular panel configurations that differ in size and/or stiffness.

Referring to Figures 1 and 2, sound components may be and are represented as generally rectangular, and may have aspect ratios that are considered to be preferred in WO97/09842, although a wider range of aspect ratios appears to have Useful potential to obtain high modal density and uniform modal distribution within the component.

Figures 4 and 4A show an embodiment of a resonant acoustic component 40 stretched over a rectangular surrounding frame 41 and clamped thereto by a clamping frame 42 to hold the acoustic component in place. Tension is applied to component 40 in the direction of arrow F. As shown in FIG. Alternatively, as shown in Figure 4B, the clamping frame 42 can be replaced by a tensioning device 43, such as a tensioning device 43 comprising a tensioning spring 44 on a frame 45, which is fixed to the edge of the sound component to stretch the part over the rectangular surrounding frame.

A vibration exciter, such as that described in WO97/09842, can be placed on the sound part in the embodiment of Figures 4, 4A and 4B to excite resonance in the sound part to produce sound output, thereby The sound part can be used as a speaker or a speaker driver unit. For clarity, these vibration exciters are not shown in Figures 4, 4A and 4B.

The strong confinement or clamping of the panel edges enables the use of relatively low stiffness materials (compared to typical applications of panels with substantially free edges), which can help by lowering the panel's fundamental bending mode frequencies, This includes applications even below the level practical for generally stiffer edge substantially free panels (despite the fact that the lowest frequency free edge modes are lost throughout the clamped panel). For example, practical examples of free edge panels of the kind described in WO97/09842 may have stiffnesses in the order of 0.1 to 50 Nm, clamped edge panels of the same general kind may be at least an order of magnitude lower, even lower to 0.001Nm. While the surface density of the described practical examples of free edge panels can range from 100 to 1000 g/m ² , the surface density of clamped edge panels is only a fraction of that, even as low as 25 g/m ² . But it should be understood that very stiff and/or denser materials could be used for the sound panels here, at least where the lowest frequency performance is not required, by firm edge confinement or clamping. Such applications include tweeters, sirens, ultrasonic regenerators.

Using a lower stiffness panel material results in higher coherent frequencies, for example above the normal acoustic frequency band, which improves the uniformity of sound directionality from the resonant speaker panel. And the lower stiffness panel can effectively enhance the modal density in the low range, thus improving the sound quality.

Useful changes are made to the confinement/clamping of the entire surrounding edge/boundary as shown, including any effective lesser degree of significant confinement/clamping, which for a roughly rectangular panel part/active area would be The constraints shown on the other three sides are omitted on one side; or the constraints shown on the other two sides are omitted on two sides that are generally parallel.

The sound emitting means herein may be activated in any of the ways suggested in WO97/09842, for example by at least one inertial electromechanical actuator means. The or each exciter means may be arranged to excite the emitting part at any suitable geometrical position in the area of the sound part, either according to the principles in WO97/09842 or according to the mechanical impedance analysis in PCT/GB99/00404 Or experiment to be sure. Such a vibration exciter has been omitted from FIG. 1 for clarity.

Reference may be made to WO97/09842 for the types of actuators applicable, and the positioning of these actuators may be determined according to the same principles as taught in WO97/09842 and/or PCT/GB99/00404, compared to WO97/09842 , the difference being the actual location that can be used.

Some useful research on resonant panel components clamped as or located across the edges of active acoustic devices, particularly loudspeakers, was first published in co-pending PCT application PCT/GB99/00404 and related to its Figures 11 to 16 , which was filed on February 9, 1999; these figures are reproduced here as Figures 6 to 11, respectively. These studies are of course based on the analysis of parameters involved in power transfer, especially the smoothness of the input power in relation to the mechanical impedance; and especially on the available-optimized transducer position and the shape of the panel parts, especially for at least is the aspect ratio of the roughly rectangular panel member and the transducer position on a scaled coordinate basis. Thus, Figures 6A, 6B and 7A, 7B and 8A, 8B are graphical representations of mechanical impedance versus frequency for panel components having selected aspect ratios and bending stiffness isotropic, accompanying the graphs of Figures 9A, 9B and 9C. The curves represent the smoothed mechanical impedance measured by the standard inverse mean square error for the localization of particularly promising transducer positions. Precisely calculated preferred aspect ratios of 1.160, 1.341 and 1.643 are shown together with equally precisely calculated preferred transducer position coordinates (0.437, 0.414), (0.385, 0.387), (0.409, 0.439), respectively. Figure 10 is the calculated 1/4 panel mechanical impedance curve for an aspect ratio of 1.16, and shows a substantial range of areas that are promising for transducer location, even two such separate areas (hatched area ). Figure 11 gives a comparison of this preferred clamping edge aspect ratio and transducer location, also including for an aspect ratio of 1.138.

Further studies done here are based on actual measurements of mechanical input power involving substantially rectangular resonant panel members with increased aspect ratios; and in each case for a reference value of ten times the lowest effective resonant modal frequency or A nearly straight line is used to fit the frequency response. Quarter-panel contours of the inverse mean square error of this fit are given in Figs. 12A-H, including for cases equal to or close to the aforementioned shape ratios (Figs. 12A, 12B, and 12D), corresponding to Fig. 13A -H is used to fit straight line frequencies respectively, the lightest color/shading represents the most feasible transducer position and enters the discontinuity region of feasibility indicated by the higher shape ratio.

It would be worthwhile to extend these further studies to aspect ratios as high as 1:4, perhaps especially in establishing the feasibility of starting with a square or near-square shape. This was unexpected, to say the least, in the context of our previous enlightening work and teachings on resonant panel components with edges substantially free of bending wave vibrations. Likewise, the unexpected increase in operating power for aspect ratio 1.41 identified here from Figures 5A, 5B was further determined to be consistent with other aspect ratios currently being investigated. Even more unexpectedly, a significant reduction of the shape ratio threshold gives an average spacing of acoustically favorable resonant modal frequencies, the reason for which has been given for further careful consideration and analysis. The following results are given in terms of simplified beam theory, the resonant panel part is roughly rectangular, and the bending stiffness of the resonant panel part is substantially isotropic.

In general, it is confirmed from previous work/teachings, that is, for panel parts with substantially free edges, that the lowest resonant modal frequency determined by the long side dimension is optimal and corresponds to the next higher Together, the short side dimensions of the resonant modal frequencies give a respective sequence of interrelated higher resonant modal frequencies. These sequences are substantially interleaved with each other numerically. In fact, for panels with essentially free edges, a large aspect ratio results in a second order (or higher) of the resonant modal frequencies of the panel components based directly on the dimensions of the long side, while lower than the first order based on the short side size. The band gap is thus too large to actually achieve satisfactory acoustic performance relying on bending wave action at the low frequencies involved.

In contrast, for fully edge-clamped resonant panel components, the first effective resonant modal frequency actually requires contributions from the first resonant mode based on the length of the short side, that is, the two axes parallel to the side x, y The first combined mode of the plane vibration action of a sequence (fx1, fx2,:...fxn) and (fy1, fy2:...fym), as shown in the resonance mode spectrum equation:

${\mathrm{<span\; class="notranslate">\; fxy</span>}}_{\mathrm{<span\; class="notranslate">\; nm</span>}}<span\; class="notranslate">\; =</span>\sqrt{<span\; class="notranslate">\; (</span>{\mathrm{<span\; class="notranslate">\; fx</span>}}_{\mathrm{<span\; class="notranslate">\; no</span>}}{<span\; class="notranslate">\; )</span>}^{\mathrm{<span\; class="notranslate">\; 2</span>}}<span\; class="notranslate">\; +</span>{<span\; class="notranslate">\; (</span>{\mathrm{<span\; class="notranslate">\; fx</span>}}_{\mathrm{<span\; class="notranslate">\; m</span>}}<span\; class="notranslate">\; )</span>}^{\mathrm{<span\; class="notranslate">\; 2</span>}}}<span\; class="notranslate">\; -</span><span\; class="notranslate">\; -</span><span\; class="notranslate">\; -</span>\mathrm{<span\; class="notranslate">\; no</span>}<span\; class="notranslate">\; \&Greater\; Equal;</span>\mathrm{<span\; class="notranslate">\; 1</span>}<span\; class="notranslate">\; -</span><span\; class="notranslate">\; -</span>\mathrm{<span\; class="notranslate">\; m</span>}<span\; class="notranslate">\; \&Greater\; Equal;</span>\mathrm{<span\; class="notranslate">\; 1</span>}$

The effect of this orthogonal relationship is that large aspect ratios can produce a series of resonant modal frequencies in close proximity based on the next-higher-order contributions in those sequences associated with the long sides, followed by those in the sequences associated with the short sides Sub-higher-order contributions. Figure 14 plots the inverse mean square error of maximum power versus aspect ratio and shows that as aspect ratio increases, the power flatness (greater than the lowest effective resonant mode) increases, peaking at about 1:3. Effectively, large shape ratios have closer resonant modal frequencies for edge-constrained components, whereas the reverse holds true for relatively free-edge panels in document WO 97/09842.

Of course, this result does not in any way detract from the beneficial and possessive effect of acoustic devices using smaller aspect ratio, whole edge clamped resonant panel components, which can also be fully applied to the above analysis from e.g. PCT/GB99/00404 The desired acoustic device operation begins at the interleaved resonant modal frequencies predicted.

However, there are clearly greater design possibilities here. In any particular case and desired application of the acoustic device here, for a given bending stiffness or shape ratio, the specific spectrum of resonant modal frequencies will vary significantly with shape ratio, and for a desired or acceptable Acoustic device performance is often selected based on calculated, measurable or perceived results.

Another relevant factor has been identified and studied, namely axial and/or azimuth-related sound action and performance, for which differences are apparent and are important when designing a specific acoustic device for a specific application. Useful/effective, especially if such a gap is indeed necessary or unnecessary, or some particular combination of these factors that is preferred or acceptable is useful/effective. Figures 15A-J are combined polar plots, respectively, for the lower resonant modal frequencies of a resonant panel member with an aspect ratio of 1:3, in each case representing the transverse (solid) and longitudinal (dashed) planes, That is, expressed by the horizontal or vertical longer dimension. In general, as expected, the emission pattern is distinct, generally smoother in smaller length faces and more diffuse in longer length faces. Design options include the acceptability of higher frequencies for the lowest resonant mode, which is directly dependent on the shape ratio for any particular panel member configuration; including the acceptability of significantly different orientations of vibration of the panel member in different axes; The relative selection of the orientation or spatial azimuth of the panel components is thus used differently in the power smoothness at the respective emission planes; The similarity of responses in elevation planes or other effective trade-offs.

The panel part of Figure 16A includes a 0.05 mm thick black glass skin on a 4 mm thick aluminum honeycomb structure, resulting in a substantially isotropic bending stiffness of 12.26 Newton meters, a mass density of 0.76 kg/square meter and a coherence frequency of 4.6 kHz . The panel part of Figure 16B comprises a 0.102 mm thick black glass skin on a 1.8 mm thick Rohacell core, resulting in a substantially isotropic bending stiffness of 2.47 Newton meters, a mass density of 0.60 kg/m2 and a coherence frequency of 9.1 kHz . The panel part of Figure 16C comprised a 0.05 mm thick Melinext ^™ skin on a 1.5 mm thick Rohacell core, resulting in a substantially isotropic bending stiffness of 0.32 Newton meters, a mass density of 0.35 kg/m2 and a coherence frequency of 19.2 kHz . These panels all had similar aspect ratios between 1.13 and 1.14 and were driven with similar drivers with 13 mm effective diameter and 8 ohm input impedance. The sound power of each panel component is measured with all panel edges free to vibrate against the panel's resonant bending wave action, and with all edges clamped against this vibration. Figures 16A-C show the improvement in sound output power achieved by clamping below the coherent frequency. Although not shown above the coherent frequency, the higher the coherent frequency, the greater the beneficial effect of clamping. Therefore, only the Lower bending stiffness of panel components.

The panel components used in Figures 16D-E have the same stiffest structure as Figure 16A, but have larger dimensions than the 260mm x 230mm dimensions of Figures 16A-C, namely 360mm x 315mm and 545mm x 480mm respectively . Here it is demonstrated that complete clamping is achieved from the coherence frequency down to the lowest resonant modal frequency of the particular involved panel part to about 400 Hz for the smallest panel part (Fig. 16A-C) and lower for the larger or largest panel part. Tightness produces improved sound output power. It is also worth noting that the larger the panel part, the closer the shape of the mode for a given frequency is to a sine wave.

Examination measures consisted of the mechanical input power to drive all of these panel parts when the edges were free to vibrate and when the edges were fully clamped, and showed that all panel parts took up almost the same power.

Of interest is the consideration of the hypothesis before the opposite teaching of WO97/09842: based on the theory that in an infinite panel expecting a perfect sine wave, no useful sound emission below the coherent frequency is obtained; following the teaching of WO97/09842 It is clearly established that useful sound emissions below the coherent frequency are available on finite panels, i.e. such emissions come from components of the finite panel that vibrate differently from the full sinusoidal distribution, manifested primarily by For the lowest frequency mode and close to the exciting transducer and at the edge of free vibration, of course the latter is emphasized here. However, from its teaching it is clear that the ability to specifically confine the edges to confine bending wave vibrations has beneficial effects on the coupling of sound to air, especially below coherent frequencies where efficiency is improved. Of course this happens in a self-evident context of sound output power which is necessarily associated with losses in the resonant panel part and in the near field of sound, at least the latter being significantly reduced by its edge limits, Acoustic shorting around such edges subject to this limitation is effectively eliminated.

It seems reasonable to attribute the increase in the coupling of sound to air below the coherent frequency to the reflection of energy that would otherwise be lost in the near-field of sound, if only on this basis that this energy is in a state of interest. In the acoustic range of bending wave vibration at the resonant modal frequency of the panel member, the panel member must be retained as acoustic energy, either at the confined edges or in the middle of the panel member as the coupling to the air is improved. The case above the coherent frequency is of course unaffected. Of course this is all within a step closer to the available resonant modes of the panel part with edge confinement which is necessary not to generate torsional modes of vibration which are captured by Edge confinement and especially clamping are effectively reduced or eliminated.

Locations beneficial for the second transducer were further investigated using a relocatable/unfixed second transducer based on the measurements; edges of the discontinuity were investigated using inertial mass at the positioned location Constrain/clamp. The results regarding the location of the second transducer mainly emphasize the extent and complexity of the interaction between the effects in the resonant panel components of the two transducers. In fact, the best indicated position for a transducer next to the advantageously placed first transducer for a resonant panel member of substantially rectangular shape and substantially isotropic bending stiffness is actually at or near Near the center or close to a position along the 3/4 length of the axis bounding the panel square where the first transducer is placed, the quality of the sound output tends to be negatively affected (although this is certainly practicable for some applications). Results for discontinuous confinement/clamping are of particular interest in indicating potentially useful transitions from near equivalence of continuous confinement/clamping to bending in the considered panel components The wavelength dependence of the wave and the larger spacing dependence of the bandpass filter effect on the acoustic frequency.

Claims (26)

1. acoustic apparatus, depending on bending wave action also can work under coincidence frequency, comprise a sound components, sound by in this sound components, forming resonance bending wave mode, it is characterized in that also comprise the device that defines this sound components at least in part, this defines device and has limited characteristic for bending wave vibration, the described device that defines is positioned on the described parts edge on every side, and along occupying at least 25% of described edge on every side continuously.

2. acoustic apparatus, comprise a sound components, sound by in this sound components, forming resonance bending wave mode, it is characterized in that, also comprise a device that defines described sound components at least in part, this defines device bending wave vibration is had limited characteristic, transducer apparatus position on described sound components by with reference to and consider that this device that defines determines, the described device that defines is positioned on the described sound components edge on every side, and along occupying at least 25% of described edge on every side continuously.

3. according to the acoustic apparatus of claim 1 or 2, it is characterized in that define device and be used for and limit described acoustics active zone, this sound components is the part of a structure or this structure.

4. according to the acoustic apparatus of claim 1 or 2, it is characterized in that, define device and extend along the described zone of rectangle or a limit of panel component.

5. according to the acoustic apparatus of claim 4, it is characterized in that, define device and also extend along the relative parallel edges of described zone or panel component.

6. according to the acoustic apparatus of claim 1 or 2, it is characterized in that, define device and extend in the entire circumference at described zone or edge.

7. according to the acoustic apparatus of claim 1 or 2, it is characterized in that, define the integral rigidity that device obviously promotes these parts, this is effective in required acoustic operation.

8. according to the acoustic apparatus of claim 7, it is characterized in that the bending stiffness of these parts is below 5 Newton meters in described acoustics active zone.

9. acoustic apparatus according to Claim 8 is characterized in that, described bending stiffness is more than 0.001 Newton meter.

10. according to the acoustic apparatus of claim 9, it is characterized in that the superficial density of these parts is from 25g/m in described acoustics active zone ²Beginning.

11. the acoustic apparatus according to claim 1 or 2 is characterized in that, the described zone of described parts has isotropic bending stiffness.

12. the acoustic apparatus according to claim 1 or 2 is characterized in that, the described zone of described parts has two-way anisotropic bending stiffness.

13. the acoustic apparatus according to claim 1 or 2 is characterized in that, the described zone of described parts has the bending stiffness that is suitable for the needed vantage point of transducer apparatus and distributes.

14. the acoustic apparatus according to claim 1 or 2 is characterized in that, described zone is a rectangle, and its bending stiffness is isotropic, and has the shape ratio between 1: 1 to 1: 1.5.

15. the acoustic apparatus according to claim 1 or 2 is characterized in that, described zone is a rectangle, and its bending stiffness is isotropic, and has the shape ratio greater than 1: 1.5.

16. the acoustic apparatus according to claim 15 is characterized in that, described shape ratio is 1: 3 or bigger.

17. the acoustic apparatus according to claim 1 or 2 is characterized in that, defines device and is used for limiting at least the resonance mode of the bending wave action that relates to distortion.

18. the acoustic apparatus according to claim 1 or 2 is characterized in that, defines the near field that device is used for reducing at least voice output and offsets.

19. the acoustic apparatus according to claim 1 or 2 is characterized in that, defines device and is used for improving voice output from the energy of the resonance bending wave action in the described zone.

20. the acoustic apparatus according to claim 1 or 2 is characterized in that, these parts are tensioned.

21. the acoustic apparatus according to claim 20 is characterized in that, these parts are tensioned in a lateral direction at two.

22. the acoustic apparatus according to claim 21 is characterized in that, tensioning equates on both direction.

23. the acoustic apparatus according to claim 22 is characterized in that, the bending stiffness of these parts in the acoustics active zone is low.

24. the acoustic apparatus according to claim 20 is characterized in that, the superficial density of these parts in the acoustics active zone is low.

25. the acoustic apparatus according to claim 1 or 2 is characterized in that, defines device and this parts one.

26. the acoustic apparatus according to claim 25 is characterized in that, defines device and this parts one molding.

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