CA2326193A1 - Acoustic device - Google Patents
Acoustic device Download PDFInfo
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- CA2326193A1 CA2326193A1 CA002326193A CA2326193A CA2326193A1 CA 2326193 A1 CA2326193 A1 CA 2326193A1 CA 002326193 A CA002326193 A CA 002326193A CA 2326193 A CA2326193 A CA 2326193A CA 2326193 A1 CA2326193 A1 CA 2326193A1
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- panel
- cavity
- acoustic
- enclosure
- resonant
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- XAGFODPZIPBFFR-UHFFFAOYSA-N aluminium Chemical compound [Al] XAGFODPZIPBFFR-UHFFFAOYSA-N 0.000 description 1
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Classifications
-
- H—ELECTRICITY
- H04—ELECTRIC COMMUNICATION TECHNIQUE
- H04R—LOUDSPEAKERS, MICROPHONES, GRAMOPHONE PICK-UPS OR LIKE ACOUSTIC ELECTROMECHANICAL TRANSDUCERS; DEAF-AID SETS; PUBLIC ADDRESS SYSTEMS
- H04R7/00—Diaphragms for electromechanical transducers; Cones
- H04R7/02—Diaphragms for electromechanical transducers; Cones characterised by the construction
- H04R7/04—Plane diaphragms
- H04R7/045—Plane diaphragms using the distributed mode principle, i.e. whereby the acoustic radiation is emanated from uniformly distributed free bending wave vibration induced in a stiff panel and not from pistonic motion
-
- H—ELECTRICITY
- H04—ELECTRIC COMMUNICATION TECHNIQUE
- H04R—LOUDSPEAKERS, MICROPHONES, GRAMOPHONE PICK-UPS OR LIKE ACOUSTIC ELECTROMECHANICAL TRANSDUCERS; DEAF-AID SETS; PUBLIC ADDRESS SYSTEMS
- H04R1/00—Details of transducers, loudspeakers or microphones
- H04R1/20—Arrangements for obtaining desired frequency or directional characteristics
- H04R1/22—Arrangements for obtaining desired frequency or directional characteristics for obtaining desired frequency characteristic only
- H04R1/24—Structural combinations of separate transducers or of two parts of the same transducer and responsive respectively to two or more frequency ranges
Landscapes
- Engineering & Computer Science (AREA)
- Signal Processing (AREA)
- Physics & Mathematics (AREA)
- Acoustics & Sound (AREA)
- Otolaryngology (AREA)
- Health & Medical Sciences (AREA)
- Multimedia (AREA)
- Details Of Audible-Bandwidth Transducers (AREA)
- Diaphragms For Electromechanical Transducers (AREA)
- Obtaining Desirable Characteristics In Audible-Bandwidth Transducers (AREA)
- Measurement Of Mechanical Vibrations Or Ultrasonic Waves (AREA)
- Soundproofing, Sound Blocking, And Sound Damping (AREA)
- Piezo-Electric Transducers For Audible Bands (AREA)
- Fittings On The Vehicle Exterior For Carrying Loads, And Devices For Holding Or Mounting Articles (AREA)
Abstract
From one aspect the invention is an acoustic device, e.g. a loudspeaker, comprising a resonant multi-mode acoustic radiator panel having opposed faces, a vibration exciter arranged to apply bending wave vibration to the resonant panel to produce an acoustic output, means defining a cavity enclosing at least a portion of one panel face and arranged to contain acoustic radiation from the said portion of the panel face, wherein the cavity is such as to modify the modal behaviour of the panel. From another aspect the invention is a method of modifying the modal behaviour of a resonant panel acoustic device, comprising bringing the resonant panel into close proximity with a boundary surface to define a resonant cavity therebetween.
Description
2 PC'T/GB99/01048 TITLE: ACOUSTIC DEVICE
DESCRIPTION
TECHNICAL FIELD
The invention relates to acoustic devices and more particularly, but not exclusively, to loudspeakers incorporating resonant mufti-mode panel acoustic radiators, e.g. of the kind described in our International application W097/09842. Loudspeakers as described in W097/09842 have become known as distributed mode (DM) loudspeakers.
Distributed mode loudspeakers (DML) are generally associated with thin, light and flat panels that radiate acoustic energy equally from both sides and in a complex diffuse fashion. While this is a useful attribute of a DML
there are various real-world situations in which by virtue of the applications and their boundary requirements a monopolar form of a DML would be preferred.
In such applications the product may with advantage be light, thin and unobtrusive.
BACKGROUND ART
It is known from International patent application W097/09842 to mount a multi-mode resonant acoustic radiator in a relatively shallow sealed box whereby acoustic radiation from one face of the radiator is contained. In this connection it should be noted that the term 'shallow' in this context is relative to the typical proportions of a pistonic cone type loudspeaker drive unit in a volume efficient enclosure. A typical volume to pistonic diaphragm area ratio may be 80:1, expressed in ml to cm2. A shallow enclosure for a resonant panel loudspeaker where pistonic drive of a lumped air volume is of little relevance, may have a ratio of 20:1.
DISCLOSURE OF INVENTION
According to the invention an acoustic device comprises a resonant multi-mode acoustic resonator or radiator panel having opposed faces, means defining a cavity enclosing at least a portion of one panel face and arranged to contain acoustic radiation from the said portion of the panel face, wherein the cavity is such as to modify the modal behaviour of the panel. The cavity may be sealed. A vibration exciter may be arranged to apply bending wave vibration to the resonant panel to produce an acoustic output, so that the device functions as a loudspeaker.
The cavity size may be such as to modify the modal behaviour of the panel.
The cavity may be shallow. The cavity may be sufficiently shallow that the distance between the internal cavity face adjacent to the said one panel face and the one panel face is sufficiently small as to cause fluid coupling to the panel. The resonant modes in the cavity can comprise cross modes parallel to the panel, i.e. which modulate along the panel, and perpendicular modes at right angles to the panel. Preferably the cavity is sufficiently shallow that the cross modes (X,Y) are more significant in modifying the modal behaviour of the panel than the perpendicular modes (Z). In embodiments, the frequencies of the perpendicular modes can lie outside the frequency range of interest.
The ratio of the cavity volume to panel area (ml:cm2) may be less than 10:1, say in the range about 10:1 to 0.2:1.
The panel may be terminated at its edges by a generally conventional resilient surround. The surround may resemble the roll surround of a conventional pistonic drive unit and may comprise one or more corrugations. The resilient surround may comprise foam rubber strips.
Alternatively the edges of the panel may be clamped in the enclosure, e.g. as described in our co-pending PCT
patent application PCT/GB99/00848 dated 30 March 1999.
Such an enclosure may be considered as a shallow tray containing a fluid whose surface may be considered to have wave-like behaviour and whose specific properties depend on both the fluid (air) and the dimensional or volume box 4 ' geometry. The panel is placed in coupled contact with this active wave surface and the surface wave excitation of the panel excites the fluid. Conversely the natural wave properties of the fluid interact with the panel, so modifying its behaviour. This is a complex coupled system with new acoustic properties in the field.
Subtle variations in the modal behaviour of the panel may be achieved by providing baffling, e.g. a simple baffle, in the enclosure and/or by providing frequency selective absorption in the enclosure.
Fr~n another aspect the invention is a method of modifying the modal behaviour of a resonant panel loudspeaker or resonator, comprising bringing the resonant panel into close proximity with a boundary surface to define a resonant cavity therebetween.
BRIEF DESCRIPTION OF DRAWINGS
Figure 1 is a cross section of a first embodiment of sealed box 1 5 resonant panel loudspeaker;
Figure 2 is a cross-sectional detail, to an enlarges scale, of the e~r~bodiment of Figure 1;
Figure 3 is a cross section of a second embodiment of sealed box resonant panel loudspeaker;
2 0 Figure 4 shows the polar response of a DNIL free-radiating on both sides;
Figure 5 shows a ccc~arison between the sound pressure level in Free Space (solid line) and with the ~ arranged 35mm frarn the wall (dotted line) 2 5 Figure 6 shows a comparison between the acoustic power of a DNIL
in free space (dotted line) and with a baffle around the panel between the front and rear;
Figure 7 shows a loudspeaker according to the invention;
Figure 8 shows a DML panel system;
Figure 9 illustrates the coupling of components;
Figure 10 illustrates a single plate eigen-function;
Figure 11 shows the magnitudes of the frequency 5 response of the first ten in-vacuum panel modes;
Figure 12 shows the magnitudes of the frequency response of the same modes in a loudspeaker according to the embodiment of the invention;
Figure 13 shows the effect of the enclosure on the panel velocity spectrum;
Figure 14 illustrates two mode shapes;
Figure 15 shows the frequency response of the reactance;
Figure 16 illustrates panel velocity measurement;
Figure 17 illustrates the microphone set up for the measurements;
Figure 18 shows the mechanical impedance for various panels;
Figure 19 shows the power response of various panels;
Figure 20 shows the polar response of various panels;
Figure 21 shows a microphone set up for measuring the internal pressure in the enclosure;
Figure 22 shows the internal pressure contour;
Figure 23 shows the internal pressure measured using the array of Figure 21;
Figure 24 shows the velocity and displacement of various panels;
Figure 25 shows the velocity spectrum of an A5 panel in free space and enclosed;
Figure 26 shows the velocity spectrum of another A5 panel in free space and enclosed;
Figure 27 shows the power response of an A2 panel in an enclosure of two depths, and Figure 28 illustrates equalisation using filters.
In the drawings and referring more particularly to Figures 1 and 2, a sealed box loudspeaker 1 comprises a box-like enclosure 2 closed at its front by a resonant panel-form acoustic radiator 5 of the kind described in W097/09842 to define a cavity 13. The radiator 5 is energised by.a vibration exciter 4 and is sealed to the enclosure round its periphery by a resilient suspension 6.
The suspension 6 comprises opposed resilient strips 7, e.g.
of foam rubber mounted in respective L-section frame members 9,10 which are held together by fasteners 11 to form a frame 8. The interior face 14 of the back wall 3 of the enclosure 2 is formed with stiffening ribs 12 to minimise vibration of the back wall. The enclosure may be a plastics moulding or a casting incorporating the stiffening ribs.
The panel in this embodiment may be of A2 size and the depth of the cavity 13 may be 90mm.
The loudspeaker embodiment of Figure 3 is generally similar to that of Figures 1 and 2, but here the radiator panel 5 is mounted on a single resilient strip suspension 6, e.g. of foam rubber, interposed between the edge of the radiator 5 and the enclosure to seal the cavity. The 7 _ radiator panel size may be A5 and the cavity depth around 3 or 4 mm.
It will be appreciated that although the embodiments of Figures 1 to 3 relate to loudspeakers, it would equally be possible to produce an acoustic resonator for modifying the acoustic behaviour of a space, e.g. a meeting room or auditorium, using devices of the general kind of Figures 1 to 3, but which omit the vibration exciter 4.
It is shown that a panel in this form of deployment can provide a very useful bandwidth with quite a small enclosure volume with respect to the diaphragm size, as compared with piston speakers. The mechanisms responsible for the minimal interaction of this boundary with the distributed mode action are examined and it is further shown that in general a simple passive equalisation network may be all that is required to produce a flat power response. It is also demonstrated that in such a manifestation, a DML can produce a near-ideal hemispherical directivity pattern over its working frequency range into a 2Pi space.
A closed form solution is presented which is the result of solving the bending wave equations for the coupled system of the panel and enclosure combination. The system acoustic impedance function is derived and is in turn used to calculate the effect of the coupled enclosure on the eigen-frequencies, and predicting the relevant shifts and additions to the plate modes.
Finally, experimental measurement data of a number WO 99/52322 PC'T/GB99/01048 samples of varying lump parameters and sizes are investigated and the measurements compared with the results from the analytical model.
Figure 4 illustrates a typical polar response of a free DML. Note that the reduction of pressure in the plane of the panel is due to the cancellation effect of acoustic radiation at or near the edges. When a free DML is brought near a boundary, in particular parallel with the boundary surface, acoustic interference starts to take place as the distance to the surface is reduced below about l5cm, for a panel of approximately 500 cm2 surface area. The effect varies in its severity and nature with the distance to the boundary as well as the panel size. The result, nonetheless is invariably a reduction of low frequency extension, peaking of response in the lower midrange region, and some aberration in the midrange and lower treble registers as shown in the example of Figure 5. Because of this, and despite the fact that the peak can easily be compensated for, application of a 'free' DML near a boundary becomes rather restrictive. ' When a DML is placed in a closed box or so-called "infinite baffle" of sufficiently large volume, radiation due to the rear of the panel is contained and that of the front is generally augmented in its mid and low frequency response, benefiting from two aspects. First is due to the absence of interference effect, caused by the front and rear radiation, at frequencies whose acoustic wavelengths in air are comparable to the free panel dimensions; and second, from the mid to low frequency boundary reinforcement due to baffling and radiation into 2Pi space, see Figure 6. Here we can see that almost 20 dB
augmentation at 100Hz is achieved from a panel of 0.25 mz surface area.
Whilst this is a definite advantage in maximising bandwidth, it may not be possible to incorporate in practice unless the application would lend itself to such a solution. Suitable applications include ceiling tile loudspeakers and custom in-wall installation.
In various other applications there may be a definite advantage to utilise the benefits of the "infinite baffle"
configuration, without having the luxury of a large closed volume of air behind the panel. Such applications may also benefit from an overall thinness and lightness of the loudspeaker. It is an object of the present invention to bring understanding to this form of deployment and offer analytical solutions.
A substantial volume of work supports conventional piston loudspeakers in various modes of operation, especially in predicting their low frequency behaviour when used in an enclosure. It is noteworthy that distributed mode loudspeakers are of very recent development and as such there is virtually no prior knowledge of the issues involved to assist with the derivation of solutions for similar analysis. In what follows, an approach is adopted which provides a useful set of solutions for a DML deployed in various mechanoacoustic interface conditions including ". .....,.._~_ ...~,CV.,,,.~:v ,m . 1_ 6_CA 02326193 2000-09-271,$0 464~4~U5-.
+49 f39 WO 99/5232x PC'T'/GB99101~48 loading with a small enclosure.
The system under analysis is shown schematically in F~,gux~e 7 . In this exampl a the front side of the panel rad_ates into free space, whilst. the other side ~.s loaded 3 wit'.~. an enclc;:ure. This coupled system ray be treated as a nettuork of velocities and pxessuras as shown in the bloeF
diagram of Figure 8. The components are, from lef~ to right; the electromechanical driving section, the moda?
system of the panel, and the acoustical systems.
10. The normal ve~.ecity of the bending-wave field across a v:.bxating panel is responsible for its acoustic radiation.
This radiation in tuxn leads to a reacting force Which modifies the panel ~=ibration. In the case of a DML
radiating equally from both sides, the radiation impedance, which is the reacting element, is normally insignificant as compared with the mechanical impedance of the panel.
However, whey. the panel radiates into a small enclosure, the effect of acousCic impedance due to its rear radiation is na longer small, and in fact it Will modify and add to 2Q the scale of the modality of the panel.
This coupling, as shown in Figure 9, is equivalent to a mechanoacoustica~. closed loop system in which the reacting sound pressure is due to the velocity of the panel.
itself. This pressure modifies the modal distribution of the bending wave field which i.n turn has an effect on the sound pressure response and directivity of the panel.
In order to calculate directivity and to inspect forces and flaws within the system, it is necessary to AMENDED SHEET
solve for the plate velocity. This far-field sound pressure response can then be obtained with the help of Fourier transformation of this velocity as described in an article by PANZER, J; HARRIS, N; entitled "Distributed Mode Loudspeaker Radiation Simulation" presented at the 105th AES
Convention, San Francisco 1998 # 9783. The forces and flows can then be found with the help of network analysis.
This problem can be approached by developing the velocities and pressures of the total system in terms of the in-vacuum panel eigen-functions (3,4) as explained in CREMER,L; HECKL,M; UNGAR,E; "Structure-Borne Sound"
SPRINGER 1973 and BLEVINS, R.D. "Formulas for Natural frequency and Mode Shape", KRIEGER Publ., Malabar 1984.
For example, the velocity at any point on the panel can be calculated from equation (1).
~(x,Y) - ~ YPi (J~) ~ Foi (J~) ~ ~Pi (xo~Yo) ' ~Pi (x.Y) i=0 (1) This series represents a solution to the differential equation describing the plate bending waves, equation (2), when coupled to the electromechanical lumped element network as well as its immediate acoustic boundaries.
~g ~ ~(x,Y) ~- ~' w ' ~(x.Y) - )w' Pm(x.Y) -)~' pa(x,Y) (2) LB is the bending rigidity differential operator of fourth order in x and y, v is the normal component of the bending wave velocity. a is the mass per unit area and c~
is the driving frequency. The panel is disturbed by the mechanical driving pressure, pm. and the acoustic reacting sound pressure field, pa, Figure 7.
Each term of the series in equation (1) is called a modal velocity, or, a "mode" in short. The model decomposition is a generalised Fourier transform whose eigen-functions ~Pi share the orthogonality property with the sine and cosine functions associated with Fourier transformation. The orthogonality property of ~Pi is a necessary condition to allow appropriate solutions to the differential equation (2). The set of eigen-functions and their parameters are found from the homogenous version of equation (2) i.e. after switching off the driving forces.
In this case the panel can only vibrate at its natural frequencies or the so-called eigen-frequencies, ~i, in order to satisfy the boundary conditions.
In equation ( 2 ) , dpi ~x, y~ is the value of the ith plate eigen-function at the position where the velocity is observed. ~pi (xo,yo~ is the eigen-function at the position where the driving force Fpi~~w~ is applied to the panel. The driving force includes the transfer functions of the electromechanical components associated with the driving actuator at (xo,yo), as for example exciters, suspensions, etc. Since the driving force depends on the panel velocity at the driving point, a similar feedback situation as with the mechanoacoustical coupling exists at the drive point(s), albeit the effect is quite small in practice.
Figure 10 gives an example of the velocity magnitude distribution of a single eigen-function across a DML panel.
The black lines are the nodal lines where the velocity is zero. With increasing mode index the velocity pattern becomes increasingly more complex. For a medium sized panel approximately 200 modes must be summed in order to cover the audio range.
The modal admittance, Ypi~~~,~, is the weighting function of the modes and determines with which amplitude and in which phase the itn mode takes part in the sum of equation (1). Ypi, as described in equation (3), depends on the driving frequency, the plate eigen-value and, most important in the context of this paper, on the acoustic impedance of the enclosure together with the impedance due to the free field radiation.
Yp i (S, = 1 . Sp ' dpi Rpi Sp + Sp ~ dpi + Ypi v (3) sp - s/wp is the Laplace frequency variable normalised to the fundamental panel frequency, mp, which in turn depends on the bending stiffness Kp and mass Mp of the panel, namely (~p2= Kp/Mp. Rpi is the modal resistance due to material losses and describes the value of Ypi~~~,~ at resonance when sp - ~,pi. ~,Pi is a scaling factor and is a function of the itn plate eigen-value ~,pi and the total radiation impedance Z",ai as described in equation (4).
yp; (S~ = J~p~ + SP ~ Zma i (!w~ .
Kp Mp (4) In the vacuum case ( Zmai°O) the second term in equation (3) becomes a band-pass transfer function of second order with damping factor dpi. Figure 11 shows the magnitudes of the frequency response of the in-vacuum Ypi,~W~ for the first ten modes of a panel, when clamped at the edges. The panel eigen-frequencies coincide with the peaks of these curves.
If the same panel is now mounted onto an enclosure, the modes will not only be shifted in frequency but also modified, as seen in Figure 12. This happens as a result of the interaction between the two modal systems of the panel and the enclosure, where the modal admittance of the total system is no longer a second order function as in the in-vacuum case. In fact, the denominator of equation (3) could be expanded in a polynomial of high order, which will reflect the resulting extended characteristic function.
The frequency response graphs of Figure 13 shows the effect of the enclosure on the panel velocity spectrum.
The two frequency response curves are calculated under identical drive condition, however, the left-hand graph displays the in-vacuum case, whilst the right hand graph shows the velocity when both sides of the panel are loaded with an enclosure. A double enclosure was used in this example in order to exclude the radiation impedance of air.
The observation point is at the drive point of the exciter.
Clearly visible is the effect of the panel eigen-frequency shift to higher frequencies in the right diagram, which was also seen in Figure 12. It is noteworthy that as a result of the enclosure influence, and the subsequent increase in 5 the number and density of modes, a more evenly distributed curve describing the velocity spectrum is obtained.
The mechanical radiation impedance is the ratio of the reacting force, due to radiation, and the panel velocity.
For a single mode, the radiation impedance can be regarded 10 as constant across the panel area and may be expressed in terms of the, acoustical radiated power Pai of a single mode.
Thus the modal radiation impedance of the ith mode may be described by equation (5).
Pai 15 Zmai =2' 2 < v; >
(5) <vi> is the mean velocity across the panel associated with the ith mode. Since this value is squared and therefore always positive and real, the properties of the radiation impedance Z",ai are directly related to the properties of the acoustical power, which is in general a complex value. The real part of Pai is equal to the radiated far-field power, which contributes to the resistive part of Z",ai, causing damping of the velocity field of the panel.
The imaginary part of Pai is caused by energy storing mechanisms of the coupled system, yielding to a positive or negative value for the reactance of Zmai.
A positive reactance is caused by the presence of an acoustical mass. This is typical, for example, of radiation into free space. A negative reactance of Z"~i, on the other hand, is indicative of the presence of a sealed enclosure with its equivalent stiffness. In physical terms, a 'mass' type radiation impedance is caused by a movement of air without compression, whereas a 'spring' type impedance exists when air is compressed without shifting it.
The principal effect of the imaginary part of the radiation impedance is a shift of the in-vacuum eigen frequencies of the panel. A positive reactance of Z",~i (mass) causes a down-shift of the plate eigen-frequencies, whereas a negative reactance (stiffness) shifts the eigen frequencies up. At a given frequency, the pane-mode itself dictates which effect will be dominating. This phenomenon is clarified by the diagram of Figure 14, which shows that symmetrical mode shapes cause compression of air, 'spring' behaviour, whereas asymmetrical mode shapes shift the air side to side, yielding an acoustical 'mass' behaviour. New modes, which are not present in either system when they are apart, are created by the interaction. of the panel and enclosure reactances.
Figure 15 shows the frequency response of the imaginary part of the enclosure radiation impedance. The left-hand graph displays a 'spring-type' reactance, typically produced by a symmetrical panel-mode. Up to the first enclosure eigen-frequency the reactance is mostly negative. In-vacuum eigen-frequencies of the panel, which are within this frequency region; are shifted up. In contrast the right diagram displays a 'mass-type' reactance behaviour, typically produced by an asymmetrical panel mode.
If the enclosure is sealed and has a rigid wall parallel to the panel surface, as in our case here, then the mechanical radiation impedance for the ith-plate mode is (5) ~'(i,x,l)2.
1 O Zmai ° -l' w' Pa ' Ad x,i kz(x,l) . tan (kz(x,l) ' adz ) (6) yrci,x,i~ is the coupling integral which takes into account the cross-sectional boundary conditions and involves the plate and enclosure eigen-functions. The index, i, in equation (6) is the plate mode-number; Ldz is the depth of the enclosure; and kZ is the modal wave-number component in the z-direction (normal to the panel). For a rigid rectangular enclosure kz is described by equation(7):
2 k.n 2 ~.n kz(x,l) = ka ( + -~~dx, ~dY
(7) The indices, k and 1, are the enclosure cross-mode numbers in x and y direction, where Ldx and Ldy are enclosure dimensions in this plane. Ao is the area of the panel and Ad is cross-sectional area of the enclosure in the x and y plane.
Equation (6) is a complicated function, which describes the interaction of the panel modes and the enclosure modes in detail. In order to understand the nature of this formula, let us simplify it by constraining the.system to the first mode of the panel and to the z-modes of the enclosure only (k=1=0). This will result in the following simplified relationship.
ZmaO =-]'Za ' Ao cot (kZ ~~dz) d (8) Equation (8) is the well known driving point impedance of a closed duct (6). If the product kZ.LdZ « 1 then a further simplification can be made as follows.
Zmao = A2 .
~'w'Cab (9) where Cab = Vb/(Pa.ca2) is the acoustical compliance of the enclosure of volume Vb. Equation (9) is the low frequency lumped element model of the enclosure. If the source is a rigid piston of mass Mms with a suspension having a compliance Cms then the fundamental 'mode' has the eigen-value ~,po = 1 and the scaling factor of the coupled system of equation (9) becomes the well known relationship as shown in equation (10),[1].
~ + Cms Cmb (10) .. . .... , . ...~ . ~ IIIE\Cf fr:N UO , j. - 6 - a : 16 = 54 : U 1480 4~64~405--~ +49 89 PCfIGB 99/0 I 04 $
i,;ith the equi=~a].ent mechanicel compliance of the enclosure air vvl~~rnc C,;~, = C~b/Ao2.
~'ariou~ tests were carried out to znvestiGate tha prfect of a sh311o~w back enclosure on DM loudspeakers. zn addition to bringing general insight ir_to the behaviour ef pM~, panels in an enclosure, the experi~!ents were designed to help verify the theoretical model and establish the extent to which such models are accurate in predicting the behaviour of the coupled modal system of a DML panel and lt,s Pnclosuxe.
Two DML panels of d?fferent size and bulk properties were selected as our test objects. xt was decsded that these would be of sufficiently different size on the one hand, and of a useful difference in their bulk groperti.es on the other, to cover a good range in scale. ~ The first set 'A' was selected as a small A5 size panel of 149mm x 210mm With three different bulk mechanical properties.
These were AS-1, polycarbonate skin on polycarbonate honeycomb; A5-2 carbon fibre on Rohacell; and A5-3, Rohacell withcut skin. Set 'g' was chosen to be eight times larger, appros~imately to A2 size or 424rcan x 592mm. A2-1 was constructed with glass fibre skin on polycarbonate honeycomb core, whilst A2-2 was carbon fibre skin on aluminium honeycomb. Table-1 lists the bulk properties of these objects. Actuation was achieved . by a sing~,e electrodynamic t~toving coil exciter at the optimum position.
Two exciter types were used, where they suited most the AMENDED SHEET
WO 99/52322 PC'T/GB99/01048 size of the panels under test. In the case of A2 panels a 25mm exciter was employed with B1 - 2.3 Tm, Re = 3.7 ~ and Le - 60 ~H, whilst a 13mm model was used in the case of the smaller A5 panels with B1 - 1.0 Tm, Re=7.3 ~ and Le=36 ~,H.
Zm Size Panel Type (Nm) (Kg/m')(Ns/m) (mm) A2-1 Glass on PC Core 10.4 0.89 29.3 5 x 592 x 420 A2-2 Carbon on AI Core57.6 1.00 60.0 7.2 x 592 x A5-1 PC on PC core 1.39 0.64 7.5 2 x 210 x 199 A5-2 Carbon on Rohacell3.33 0.65 11.8 2 x 210 x 199 A5-3 Rohacell core 0.33 0 32 2 7 I 3 x 210 x 149 Panels were mounted onto a back enclosure with adjustable depth using a soft polyurethane foam for suspension and acoustic seal. The enclosure depth was made adjustable on 16,28,40 and 53mm for set 'A' and on 20,50,95 10 and 130mm for set 'B' panels. Various measurements were carried out at different enclosure depths for every test case and result documented.
Panel velocity and displacement were measured using a Laser Vibrometer. The frequency range of interest was 15 covered with a linear frequency scale of 1600 points. The set-up shown in Figure 16 was used to measure the panel mechanical impedance by calculating the ratio of the applied force to the panel velocity at the drive point.
2 0 Zm = F
V
In this procedure, the applied force was calculated -u~rvcrtr-_w os . 1_ ~ CA 023261 ~~ 2000-09-2 ~1~8U 4G~405-~ +4s as _ Gg W0 99I~232 2 PCTIG F39 ~3!O t 04 8 from the lump parameter information of the e~cc~~er.
Althour~h panel velocity in itself feeds back into the electromechanical circuit, its couplyng is quite weak. It can be shown that for small values of exciter H1, ( 1 -3 Tm) , - 5 providing that the driving amplifier output impedance is low (constant voltage), the modal coupling back to the - electromechanical system is su~ficiQntly weak to make this assumption plausible. Small error ariszng from this approximation was therefore iarnored, Figures 18a to f show ti-:e mechanical imgedance of the A5-1 and A5~2 panels, derived from the measurement of panel velocity and tree applied force measured by the Laser Vibrorzeter. Note that the impedance minima for each enclosure depth occur at the system resonance mode.
Sound pressure level and polar response 4f the various panels Grere measured in a large space of 350 cubic metres and gated at 12 to l9ms for anechoic -response using MLSSA, depending on the measurement. Power measurements were carried out employing a 9-microphone array system, as shown in Figure 17b and in a set-up shown in Figure 17a. These are plotted in Figures 19a to d fox various enclosure depths. System resonance is highlighted by markers on the graphs.
Polar response of the A5~1 and A5-2 panels 'were measured for a 28mm deep enclosure and the result is~~-shown in Figures 20a and b. When compared with the Pour plot of the free DML in Figure 1, they demonstrate the~significance of the closed-back DMZ in its improved directivity.
AMENDED SHEET
m.. .....,....... ..,~~NL~~ 06 . 1_ g_CA 0232619 2000-09-21480 484405-. +49 89 wo 9~fsz3zz rcT~c~99~o~oa~
z2 To investigate fur~her the nature and the effect of enclosure on the panel beraviour, especially at the combined system resonance, 3 special dig wa$ made to allow the measurement of the internal pressure of the enclosure at nine predetermined points as shown in Figure ~1. fhe microphone was inserted in the holes provided within zhe back-plate of an .F15 enclosure j ig at a predetermined depth, while the other eight position holes were tightly blocked with hard rubber grommets. The microphone was mechanically 1f3 isolated from the enclosure by an appropriate rubber grommet during the measurement..
Frp;Tc this data, a contour plot was created to shoal the pressure distribution at system resonance and at eithex side of this frequency as shown in Figures 22a to c_ The pressure frequency response was also plotted for the nine positions as shot~n in Figure 23. This graph eXhlblts good definition in the region of resonance for all curves associated with the measurement points within the enclosure. However, the pressure tends to vary across the 2D enclosure cross-sectional area as the frequency is increased.
The normal component of velocity and displacement across the panels was measured with a Scanning Laser Vibrometer. The velocity and displacement distribution across the panels were plotted to investigate the behaviour of the panel around the coupled system resonance. The results were dccumented and a number of the cases are shown in Figures 2~a to d. These results suggest a timpanic AMENDED SHEET
modal behaviour of the panel at resonance, with the whole of the panel moving, albeit at a lesser velocity and displacement as one moves towards the panel edges.
In practice this behaviour is consistent for all boundary conditions of the panel, although the mode shape will vary from case to case depending on a complex set of parameters, including panel stiffness, mass, size and boundary conditions. In the limit and for an infinitely rigid panel, this system resonance will be seen as the fundamental rigid body mode of the piston acting on the stiffness of the enclosure air volume. It was found to be convenient to call the DML system resonance, the 'Whole Body Mode' or WBM.
The full theoretical derivations of the coupled system has been implemented in a suite of software by New Transducers Limited. A version of this package was used to simulate the mechanoacoustical behaviour of our test objects in this paper. This package is able to take into account all the electrical, mechanical and acoustical variables associated with a panel, exciter(s) and mechanoacoustical interfaces with a frame or an enclosure and predict, amongst other parameters, the far-field acoustic pressure, power and directivity of the total system.
Figure 25a shows the log-velocity spectrum of a free radiating, A5-1 panel clamped in a frame, radiating in free space equally from both sides. The solid line represents the simulation curve and the dashed line is the measure velocity spectrum. At low frequencies the panel goes in resonance with the exciter. The discrepancy in the frequency range above 1000 Hz is due to the absence of the free field radiation impedance in the simulation model.
Figure 25b shows the same panel as in Figure 25a but this time loaded with two identical enclosures, one on each side of the panel, with the same cross-section as the panel and a depth of 24mm. A double enclosure was designed and used in order to exclude the radiation impedance of free field on one side of the panel and make the experiment independent of the free field radiation impedance. It is important to note that this laboratory set-up was used for theory verification only.
In order to enable velocity measurement of the panel, the back walls of the two enclosures were made from a transparent material to allow access by the laser beam to the panel surface. This test was repeated using panel A5-3 Rohacell without skin, with different bulk properties and the result is shown in Figures 26a and b. In both cases simulation was performed using 200 point logarithmic range, whilst the laser measurement used 1600 point linear range.
From the foregoing theory and work, it is clear that a small enclosure fitted to a DML will bring with it, amongst a number of benefits, a singular drawback. This manifests itself in an excess of power due to WBM at the system resonance as shown in Figures 27a and b. It is noteworthy that apart from this peak, in all other aspect the enclosed DML can offer a substantially improved performance including increased power bandwidth.
It has been found that in most cases a simple second order band-stop equalisation network of appropriate Q
matching that of the power response peak, may be designed to equalise the response peak. Furthermore in some cases a single pole high-pass filter would often adjust for this by tilting the LF region, to provide a broadly flat power response. Due to the unique nature of DML panels and their resistive electrical impedance response, whether the filter is active or passive, its design will remain very simple. Figure 28a shows where a band-stop passive filter has been incorporated for equalisation. Further examples may be seen in Figures 28b and c that show simple pole EQ
with a capacitor used in series with the loudspeakers.
When a free DML is used near and parallel to a wall, special care must be taken to ensure minimal interaction with the latter, due to its unique complex dipolar characteristics. This interaction is a function of the distance to the boundary, and therefore, cannot be universally fixed. Full baffling of the panel has definite advantages in extending the low frequency response of the system, but this may not be a practical proposition in a large number of applications.
A very small enclosure used with a DML will render it independent of its immediate environment and make the system predictable in its acoustical performance. The mathematical model developed demonstrates the level of complexity for a DML in the coupled system. This throws a sharp contrast between the prediction and design of a DML
and that of the conventional piston radiator. Whilst the mechanoacoustical properties of a cone-in-box may be found by relatively simply calculations (even by a hand calculator) those associated with a DML and its enclosure are subject to complex interactive relationships which render this system impossible to predict without the proper tools.
The change in system performance with varying enclosure volume is quite marked in the case where the depth is small compared with the panel dimensions. However, it is also seen that beyond a certain depth the increase in LF response become marginal. This of course is consistent with behaviour of a rigid piston in an enclosure. As an example, an A2 size panel with 50mm enclosure depth can be designed to have a bandwidth extending down to about 120Hz, Figure 24.
Another feature of a DML with a small enclosure is seen to be a significant improvement in the mid and high frequency response of the system. This is in many of the measured and simulated graphs in this paper and of course anticipated by the theory. It is clear that the increase in the panel system modality is mostly responsible for this improvement, however, enclosures losses might also influence this by increasing the overall damping of the system.
As a natural consequence of containing the rear radiation of the panel, the directivity of the enclosed system changes substantially from a dipolar shape to a near cardioid behaviour as shown in Figure 17. It is envisaged that the directivity associated with a closed-back DML may find use in certain applications where stronger lateral coverage is desirable.
Power response measurements were found to be most useful when working with the enclosed DM system, in order to observe the excessive energy region that may need compensation. This is in line with other work done on DM
loudspeakers, in which it has been found that the power response is the most representative acoustic measurement correlating well to the subjective performance of a DML.
Using the power response, it was found that in practice a simple band-pass or a single pole high-pass filter is all that is needed to equalise the power response in this region.
DESCRIPTION
TECHNICAL FIELD
The invention relates to acoustic devices and more particularly, but not exclusively, to loudspeakers incorporating resonant mufti-mode panel acoustic radiators, e.g. of the kind described in our International application W097/09842. Loudspeakers as described in W097/09842 have become known as distributed mode (DM) loudspeakers.
Distributed mode loudspeakers (DML) are generally associated with thin, light and flat panels that radiate acoustic energy equally from both sides and in a complex diffuse fashion. While this is a useful attribute of a DML
there are various real-world situations in which by virtue of the applications and their boundary requirements a monopolar form of a DML would be preferred.
In such applications the product may with advantage be light, thin and unobtrusive.
BACKGROUND ART
It is known from International patent application W097/09842 to mount a multi-mode resonant acoustic radiator in a relatively shallow sealed box whereby acoustic radiation from one face of the radiator is contained. In this connection it should be noted that the term 'shallow' in this context is relative to the typical proportions of a pistonic cone type loudspeaker drive unit in a volume efficient enclosure. A typical volume to pistonic diaphragm area ratio may be 80:1, expressed in ml to cm2. A shallow enclosure for a resonant panel loudspeaker where pistonic drive of a lumped air volume is of little relevance, may have a ratio of 20:1.
DISCLOSURE OF INVENTION
According to the invention an acoustic device comprises a resonant multi-mode acoustic resonator or radiator panel having opposed faces, means defining a cavity enclosing at least a portion of one panel face and arranged to contain acoustic radiation from the said portion of the panel face, wherein the cavity is such as to modify the modal behaviour of the panel. The cavity may be sealed. A vibration exciter may be arranged to apply bending wave vibration to the resonant panel to produce an acoustic output, so that the device functions as a loudspeaker.
The cavity size may be such as to modify the modal behaviour of the panel.
The cavity may be shallow. The cavity may be sufficiently shallow that the distance between the internal cavity face adjacent to the said one panel face and the one panel face is sufficiently small as to cause fluid coupling to the panel. The resonant modes in the cavity can comprise cross modes parallel to the panel, i.e. which modulate along the panel, and perpendicular modes at right angles to the panel. Preferably the cavity is sufficiently shallow that the cross modes (X,Y) are more significant in modifying the modal behaviour of the panel than the perpendicular modes (Z). In embodiments, the frequencies of the perpendicular modes can lie outside the frequency range of interest.
The ratio of the cavity volume to panel area (ml:cm2) may be less than 10:1, say in the range about 10:1 to 0.2:1.
The panel may be terminated at its edges by a generally conventional resilient surround. The surround may resemble the roll surround of a conventional pistonic drive unit and may comprise one or more corrugations. The resilient surround may comprise foam rubber strips.
Alternatively the edges of the panel may be clamped in the enclosure, e.g. as described in our co-pending PCT
patent application PCT/GB99/00848 dated 30 March 1999.
Such an enclosure may be considered as a shallow tray containing a fluid whose surface may be considered to have wave-like behaviour and whose specific properties depend on both the fluid (air) and the dimensional or volume box 4 ' geometry. The panel is placed in coupled contact with this active wave surface and the surface wave excitation of the panel excites the fluid. Conversely the natural wave properties of the fluid interact with the panel, so modifying its behaviour. This is a complex coupled system with new acoustic properties in the field.
Subtle variations in the modal behaviour of the panel may be achieved by providing baffling, e.g. a simple baffle, in the enclosure and/or by providing frequency selective absorption in the enclosure.
Fr~n another aspect the invention is a method of modifying the modal behaviour of a resonant panel loudspeaker or resonator, comprising bringing the resonant panel into close proximity with a boundary surface to define a resonant cavity therebetween.
BRIEF DESCRIPTION OF DRAWINGS
Figure 1 is a cross section of a first embodiment of sealed box 1 5 resonant panel loudspeaker;
Figure 2 is a cross-sectional detail, to an enlarges scale, of the e~r~bodiment of Figure 1;
Figure 3 is a cross section of a second embodiment of sealed box resonant panel loudspeaker;
2 0 Figure 4 shows the polar response of a DNIL free-radiating on both sides;
Figure 5 shows a ccc~arison between the sound pressure level in Free Space (solid line) and with the ~ arranged 35mm frarn the wall (dotted line) 2 5 Figure 6 shows a comparison between the acoustic power of a DNIL
in free space (dotted line) and with a baffle around the panel between the front and rear;
Figure 7 shows a loudspeaker according to the invention;
Figure 8 shows a DML panel system;
Figure 9 illustrates the coupling of components;
Figure 10 illustrates a single plate eigen-function;
Figure 11 shows the magnitudes of the frequency 5 response of the first ten in-vacuum panel modes;
Figure 12 shows the magnitudes of the frequency response of the same modes in a loudspeaker according to the embodiment of the invention;
Figure 13 shows the effect of the enclosure on the panel velocity spectrum;
Figure 14 illustrates two mode shapes;
Figure 15 shows the frequency response of the reactance;
Figure 16 illustrates panel velocity measurement;
Figure 17 illustrates the microphone set up for the measurements;
Figure 18 shows the mechanical impedance for various panels;
Figure 19 shows the power response of various panels;
Figure 20 shows the polar response of various panels;
Figure 21 shows a microphone set up for measuring the internal pressure in the enclosure;
Figure 22 shows the internal pressure contour;
Figure 23 shows the internal pressure measured using the array of Figure 21;
Figure 24 shows the velocity and displacement of various panels;
Figure 25 shows the velocity spectrum of an A5 panel in free space and enclosed;
Figure 26 shows the velocity spectrum of another A5 panel in free space and enclosed;
Figure 27 shows the power response of an A2 panel in an enclosure of two depths, and Figure 28 illustrates equalisation using filters.
In the drawings and referring more particularly to Figures 1 and 2, a sealed box loudspeaker 1 comprises a box-like enclosure 2 closed at its front by a resonant panel-form acoustic radiator 5 of the kind described in W097/09842 to define a cavity 13. The radiator 5 is energised by.a vibration exciter 4 and is sealed to the enclosure round its periphery by a resilient suspension 6.
The suspension 6 comprises opposed resilient strips 7, e.g.
of foam rubber mounted in respective L-section frame members 9,10 which are held together by fasteners 11 to form a frame 8. The interior face 14 of the back wall 3 of the enclosure 2 is formed with stiffening ribs 12 to minimise vibration of the back wall. The enclosure may be a plastics moulding or a casting incorporating the stiffening ribs.
The panel in this embodiment may be of A2 size and the depth of the cavity 13 may be 90mm.
The loudspeaker embodiment of Figure 3 is generally similar to that of Figures 1 and 2, but here the radiator panel 5 is mounted on a single resilient strip suspension 6, e.g. of foam rubber, interposed between the edge of the radiator 5 and the enclosure to seal the cavity. The 7 _ radiator panel size may be A5 and the cavity depth around 3 or 4 mm.
It will be appreciated that although the embodiments of Figures 1 to 3 relate to loudspeakers, it would equally be possible to produce an acoustic resonator for modifying the acoustic behaviour of a space, e.g. a meeting room or auditorium, using devices of the general kind of Figures 1 to 3, but which omit the vibration exciter 4.
It is shown that a panel in this form of deployment can provide a very useful bandwidth with quite a small enclosure volume with respect to the diaphragm size, as compared with piston speakers. The mechanisms responsible for the minimal interaction of this boundary with the distributed mode action are examined and it is further shown that in general a simple passive equalisation network may be all that is required to produce a flat power response. It is also demonstrated that in such a manifestation, a DML can produce a near-ideal hemispherical directivity pattern over its working frequency range into a 2Pi space.
A closed form solution is presented which is the result of solving the bending wave equations for the coupled system of the panel and enclosure combination. The system acoustic impedance function is derived and is in turn used to calculate the effect of the coupled enclosure on the eigen-frequencies, and predicting the relevant shifts and additions to the plate modes.
Finally, experimental measurement data of a number WO 99/52322 PC'T/GB99/01048 samples of varying lump parameters and sizes are investigated and the measurements compared with the results from the analytical model.
Figure 4 illustrates a typical polar response of a free DML. Note that the reduction of pressure in the plane of the panel is due to the cancellation effect of acoustic radiation at or near the edges. When a free DML is brought near a boundary, in particular parallel with the boundary surface, acoustic interference starts to take place as the distance to the surface is reduced below about l5cm, for a panel of approximately 500 cm2 surface area. The effect varies in its severity and nature with the distance to the boundary as well as the panel size. The result, nonetheless is invariably a reduction of low frequency extension, peaking of response in the lower midrange region, and some aberration in the midrange and lower treble registers as shown in the example of Figure 5. Because of this, and despite the fact that the peak can easily be compensated for, application of a 'free' DML near a boundary becomes rather restrictive. ' When a DML is placed in a closed box or so-called "infinite baffle" of sufficiently large volume, radiation due to the rear of the panel is contained and that of the front is generally augmented in its mid and low frequency response, benefiting from two aspects. First is due to the absence of interference effect, caused by the front and rear radiation, at frequencies whose acoustic wavelengths in air are comparable to the free panel dimensions; and second, from the mid to low frequency boundary reinforcement due to baffling and radiation into 2Pi space, see Figure 6. Here we can see that almost 20 dB
augmentation at 100Hz is achieved from a panel of 0.25 mz surface area.
Whilst this is a definite advantage in maximising bandwidth, it may not be possible to incorporate in practice unless the application would lend itself to such a solution. Suitable applications include ceiling tile loudspeakers and custom in-wall installation.
In various other applications there may be a definite advantage to utilise the benefits of the "infinite baffle"
configuration, without having the luxury of a large closed volume of air behind the panel. Such applications may also benefit from an overall thinness and lightness of the loudspeaker. It is an object of the present invention to bring understanding to this form of deployment and offer analytical solutions.
A substantial volume of work supports conventional piston loudspeakers in various modes of operation, especially in predicting their low frequency behaviour when used in an enclosure. It is noteworthy that distributed mode loudspeakers are of very recent development and as such there is virtually no prior knowledge of the issues involved to assist with the derivation of solutions for similar analysis. In what follows, an approach is adopted which provides a useful set of solutions for a DML deployed in various mechanoacoustic interface conditions including ". .....,.._~_ ...~,CV.,,,.~:v ,m . 1_ 6_CA 02326193 2000-09-271,$0 464~4~U5-.
+49 f39 WO 99/5232x PC'T'/GB99101~48 loading with a small enclosure.
The system under analysis is shown schematically in F~,gux~e 7 . In this exampl a the front side of the panel rad_ates into free space, whilst. the other side ~.s loaded 3 wit'.~. an enclc;:ure. This coupled system ray be treated as a nettuork of velocities and pxessuras as shown in the bloeF
diagram of Figure 8. The components are, from lef~ to right; the electromechanical driving section, the moda?
system of the panel, and the acoustical systems.
10. The normal ve~.ecity of the bending-wave field across a v:.bxating panel is responsible for its acoustic radiation.
This radiation in tuxn leads to a reacting force Which modifies the panel ~=ibration. In the case of a DML
radiating equally from both sides, the radiation impedance, which is the reacting element, is normally insignificant as compared with the mechanical impedance of the panel.
However, whey. the panel radiates into a small enclosure, the effect of acousCic impedance due to its rear radiation is na longer small, and in fact it Will modify and add to 2Q the scale of the modality of the panel.
This coupling, as shown in Figure 9, is equivalent to a mechanoacoustica~. closed loop system in which the reacting sound pressure is due to the velocity of the panel.
itself. This pressure modifies the modal distribution of the bending wave field which i.n turn has an effect on the sound pressure response and directivity of the panel.
In order to calculate directivity and to inspect forces and flaws within the system, it is necessary to AMENDED SHEET
solve for the plate velocity. This far-field sound pressure response can then be obtained with the help of Fourier transformation of this velocity as described in an article by PANZER, J; HARRIS, N; entitled "Distributed Mode Loudspeaker Radiation Simulation" presented at the 105th AES
Convention, San Francisco 1998 # 9783. The forces and flows can then be found with the help of network analysis.
This problem can be approached by developing the velocities and pressures of the total system in terms of the in-vacuum panel eigen-functions (3,4) as explained in CREMER,L; HECKL,M; UNGAR,E; "Structure-Borne Sound"
SPRINGER 1973 and BLEVINS, R.D. "Formulas for Natural frequency and Mode Shape", KRIEGER Publ., Malabar 1984.
For example, the velocity at any point on the panel can be calculated from equation (1).
~(x,Y) - ~ YPi (J~) ~ Foi (J~) ~ ~Pi (xo~Yo) ' ~Pi (x.Y) i=0 (1) This series represents a solution to the differential equation describing the plate bending waves, equation (2), when coupled to the electromechanical lumped element network as well as its immediate acoustic boundaries.
~g ~ ~(x,Y) ~- ~' w ' ~(x.Y) - )w' Pm(x.Y) -)~' pa(x,Y) (2) LB is the bending rigidity differential operator of fourth order in x and y, v is the normal component of the bending wave velocity. a is the mass per unit area and c~
is the driving frequency. The panel is disturbed by the mechanical driving pressure, pm. and the acoustic reacting sound pressure field, pa, Figure 7.
Each term of the series in equation (1) is called a modal velocity, or, a "mode" in short. The model decomposition is a generalised Fourier transform whose eigen-functions ~Pi share the orthogonality property with the sine and cosine functions associated with Fourier transformation. The orthogonality property of ~Pi is a necessary condition to allow appropriate solutions to the differential equation (2). The set of eigen-functions and their parameters are found from the homogenous version of equation (2) i.e. after switching off the driving forces.
In this case the panel can only vibrate at its natural frequencies or the so-called eigen-frequencies, ~i, in order to satisfy the boundary conditions.
In equation ( 2 ) , dpi ~x, y~ is the value of the ith plate eigen-function at the position where the velocity is observed. ~pi (xo,yo~ is the eigen-function at the position where the driving force Fpi~~w~ is applied to the panel. The driving force includes the transfer functions of the electromechanical components associated with the driving actuator at (xo,yo), as for example exciters, suspensions, etc. Since the driving force depends on the panel velocity at the driving point, a similar feedback situation as with the mechanoacoustical coupling exists at the drive point(s), albeit the effect is quite small in practice.
Figure 10 gives an example of the velocity magnitude distribution of a single eigen-function across a DML panel.
The black lines are the nodal lines where the velocity is zero. With increasing mode index the velocity pattern becomes increasingly more complex. For a medium sized panel approximately 200 modes must be summed in order to cover the audio range.
The modal admittance, Ypi~~~,~, is the weighting function of the modes and determines with which amplitude and in which phase the itn mode takes part in the sum of equation (1). Ypi, as described in equation (3), depends on the driving frequency, the plate eigen-value and, most important in the context of this paper, on the acoustic impedance of the enclosure together with the impedance due to the free field radiation.
Yp i (S, = 1 . Sp ' dpi Rpi Sp + Sp ~ dpi + Ypi v (3) sp - s/wp is the Laplace frequency variable normalised to the fundamental panel frequency, mp, which in turn depends on the bending stiffness Kp and mass Mp of the panel, namely (~p2= Kp/Mp. Rpi is the modal resistance due to material losses and describes the value of Ypi~~~,~ at resonance when sp - ~,pi. ~,Pi is a scaling factor and is a function of the itn plate eigen-value ~,pi and the total radiation impedance Z",ai as described in equation (4).
yp; (S~ = J~p~ + SP ~ Zma i (!w~ .
Kp Mp (4) In the vacuum case ( Zmai°O) the second term in equation (3) becomes a band-pass transfer function of second order with damping factor dpi. Figure 11 shows the magnitudes of the frequency response of the in-vacuum Ypi,~W~ for the first ten modes of a panel, when clamped at the edges. The panel eigen-frequencies coincide with the peaks of these curves.
If the same panel is now mounted onto an enclosure, the modes will not only be shifted in frequency but also modified, as seen in Figure 12. This happens as a result of the interaction between the two modal systems of the panel and the enclosure, where the modal admittance of the total system is no longer a second order function as in the in-vacuum case. In fact, the denominator of equation (3) could be expanded in a polynomial of high order, which will reflect the resulting extended characteristic function.
The frequency response graphs of Figure 13 shows the effect of the enclosure on the panel velocity spectrum.
The two frequency response curves are calculated under identical drive condition, however, the left-hand graph displays the in-vacuum case, whilst the right hand graph shows the velocity when both sides of the panel are loaded with an enclosure. A double enclosure was used in this example in order to exclude the radiation impedance of air.
The observation point is at the drive point of the exciter.
Clearly visible is the effect of the panel eigen-frequency shift to higher frequencies in the right diagram, which was also seen in Figure 12. It is noteworthy that as a result of the enclosure influence, and the subsequent increase in 5 the number and density of modes, a more evenly distributed curve describing the velocity spectrum is obtained.
The mechanical radiation impedance is the ratio of the reacting force, due to radiation, and the panel velocity.
For a single mode, the radiation impedance can be regarded 10 as constant across the panel area and may be expressed in terms of the, acoustical radiated power Pai of a single mode.
Thus the modal radiation impedance of the ith mode may be described by equation (5).
Pai 15 Zmai =2' 2 < v; >
(5) <vi> is the mean velocity across the panel associated with the ith mode. Since this value is squared and therefore always positive and real, the properties of the radiation impedance Z",ai are directly related to the properties of the acoustical power, which is in general a complex value. The real part of Pai is equal to the radiated far-field power, which contributes to the resistive part of Z",ai, causing damping of the velocity field of the panel.
The imaginary part of Pai is caused by energy storing mechanisms of the coupled system, yielding to a positive or negative value for the reactance of Zmai.
A positive reactance is caused by the presence of an acoustical mass. This is typical, for example, of radiation into free space. A negative reactance of Z"~i, on the other hand, is indicative of the presence of a sealed enclosure with its equivalent stiffness. In physical terms, a 'mass' type radiation impedance is caused by a movement of air without compression, whereas a 'spring' type impedance exists when air is compressed without shifting it.
The principal effect of the imaginary part of the radiation impedance is a shift of the in-vacuum eigen frequencies of the panel. A positive reactance of Z",~i (mass) causes a down-shift of the plate eigen-frequencies, whereas a negative reactance (stiffness) shifts the eigen frequencies up. At a given frequency, the pane-mode itself dictates which effect will be dominating. This phenomenon is clarified by the diagram of Figure 14, which shows that symmetrical mode shapes cause compression of air, 'spring' behaviour, whereas asymmetrical mode shapes shift the air side to side, yielding an acoustical 'mass' behaviour. New modes, which are not present in either system when they are apart, are created by the interaction. of the panel and enclosure reactances.
Figure 15 shows the frequency response of the imaginary part of the enclosure radiation impedance. The left-hand graph displays a 'spring-type' reactance, typically produced by a symmetrical panel-mode. Up to the first enclosure eigen-frequency the reactance is mostly negative. In-vacuum eigen-frequencies of the panel, which are within this frequency region; are shifted up. In contrast the right diagram displays a 'mass-type' reactance behaviour, typically produced by an asymmetrical panel mode.
If the enclosure is sealed and has a rigid wall parallel to the panel surface, as in our case here, then the mechanical radiation impedance for the ith-plate mode is (5) ~'(i,x,l)2.
1 O Zmai ° -l' w' Pa ' Ad x,i kz(x,l) . tan (kz(x,l) ' adz ) (6) yrci,x,i~ is the coupling integral which takes into account the cross-sectional boundary conditions and involves the plate and enclosure eigen-functions. The index, i, in equation (6) is the plate mode-number; Ldz is the depth of the enclosure; and kZ is the modal wave-number component in the z-direction (normal to the panel). For a rigid rectangular enclosure kz is described by equation(7):
2 k.n 2 ~.n kz(x,l) = ka ( + -~~dx, ~dY
(7) The indices, k and 1, are the enclosure cross-mode numbers in x and y direction, where Ldx and Ldy are enclosure dimensions in this plane. Ao is the area of the panel and Ad is cross-sectional area of the enclosure in the x and y plane.
Equation (6) is a complicated function, which describes the interaction of the panel modes and the enclosure modes in detail. In order to understand the nature of this formula, let us simplify it by constraining the.system to the first mode of the panel and to the z-modes of the enclosure only (k=1=0). This will result in the following simplified relationship.
ZmaO =-]'Za ' Ao cot (kZ ~~dz) d (8) Equation (8) is the well known driving point impedance of a closed duct (6). If the product kZ.LdZ « 1 then a further simplification can be made as follows.
Zmao = A2 .
~'w'Cab (9) where Cab = Vb/(Pa.ca2) is the acoustical compliance of the enclosure of volume Vb. Equation (9) is the low frequency lumped element model of the enclosure. If the source is a rigid piston of mass Mms with a suspension having a compliance Cms then the fundamental 'mode' has the eigen-value ~,po = 1 and the scaling factor of the coupled system of equation (9) becomes the well known relationship as shown in equation (10),[1].
~ + Cms Cmb (10) .. . .... , . ...~ . ~ IIIE\Cf fr:N UO , j. - 6 - a : 16 = 54 : U 1480 4~64~405--~ +49 89 PCfIGB 99/0 I 04 $
i,;ith the equi=~a].ent mechanicel compliance of the enclosure air vvl~~rnc C,;~, = C~b/Ao2.
~'ariou~ tests were carried out to znvestiGate tha prfect of a sh311o~w back enclosure on DM loudspeakers. zn addition to bringing general insight ir_to the behaviour ef pM~, panels in an enclosure, the experi~!ents were designed to help verify the theoretical model and establish the extent to which such models are accurate in predicting the behaviour of the coupled modal system of a DML panel and lt,s Pnclosuxe.
Two DML panels of d?fferent size and bulk properties were selected as our test objects. xt was decsded that these would be of sufficiently different size on the one hand, and of a useful difference in their bulk groperti.es on the other, to cover a good range in scale. ~ The first set 'A' was selected as a small A5 size panel of 149mm x 210mm With three different bulk mechanical properties.
These were AS-1, polycarbonate skin on polycarbonate honeycomb; A5-2 carbon fibre on Rohacell; and A5-3, Rohacell withcut skin. Set 'g' was chosen to be eight times larger, appros~imately to A2 size or 424rcan x 592mm. A2-1 was constructed with glass fibre skin on polycarbonate honeycomb core, whilst A2-2 was carbon fibre skin on aluminium honeycomb. Table-1 lists the bulk properties of these objects. Actuation was achieved . by a sing~,e electrodynamic t~toving coil exciter at the optimum position.
Two exciter types were used, where they suited most the AMENDED SHEET
WO 99/52322 PC'T/GB99/01048 size of the panels under test. In the case of A2 panels a 25mm exciter was employed with B1 - 2.3 Tm, Re = 3.7 ~ and Le - 60 ~H, whilst a 13mm model was used in the case of the smaller A5 panels with B1 - 1.0 Tm, Re=7.3 ~ and Le=36 ~,H.
Zm Size Panel Type (Nm) (Kg/m')(Ns/m) (mm) A2-1 Glass on PC Core 10.4 0.89 29.3 5 x 592 x 420 A2-2 Carbon on AI Core57.6 1.00 60.0 7.2 x 592 x A5-1 PC on PC core 1.39 0.64 7.5 2 x 210 x 199 A5-2 Carbon on Rohacell3.33 0.65 11.8 2 x 210 x 199 A5-3 Rohacell core 0.33 0 32 2 7 I 3 x 210 x 149 Panels were mounted onto a back enclosure with adjustable depth using a soft polyurethane foam for suspension and acoustic seal. The enclosure depth was made adjustable on 16,28,40 and 53mm for set 'A' and on 20,50,95 10 and 130mm for set 'B' panels. Various measurements were carried out at different enclosure depths for every test case and result documented.
Panel velocity and displacement were measured using a Laser Vibrometer. The frequency range of interest was 15 covered with a linear frequency scale of 1600 points. The set-up shown in Figure 16 was used to measure the panel mechanical impedance by calculating the ratio of the applied force to the panel velocity at the drive point.
2 0 Zm = F
V
In this procedure, the applied force was calculated -u~rvcrtr-_w os . 1_ ~ CA 023261 ~~ 2000-09-2 ~1~8U 4G~405-~ +4s as _ Gg W0 99I~232 2 PCTIG F39 ~3!O t 04 8 from the lump parameter information of the e~cc~~er.
Althour~h panel velocity in itself feeds back into the electromechanical circuit, its couplyng is quite weak. It can be shown that for small values of exciter H1, ( 1 -3 Tm) , - 5 providing that the driving amplifier output impedance is low (constant voltage), the modal coupling back to the - electromechanical system is su~ficiQntly weak to make this assumption plausible. Small error ariszng from this approximation was therefore iarnored, Figures 18a to f show ti-:e mechanical imgedance of the A5-1 and A5~2 panels, derived from the measurement of panel velocity and tree applied force measured by the Laser Vibrorzeter. Note that the impedance minima for each enclosure depth occur at the system resonance mode.
Sound pressure level and polar response 4f the various panels Grere measured in a large space of 350 cubic metres and gated at 12 to l9ms for anechoic -response using MLSSA, depending on the measurement. Power measurements were carried out employing a 9-microphone array system, as shown in Figure 17b and in a set-up shown in Figure 17a. These are plotted in Figures 19a to d fox various enclosure depths. System resonance is highlighted by markers on the graphs.
Polar response of the A5~1 and A5-2 panels 'were measured for a 28mm deep enclosure and the result is~~-shown in Figures 20a and b. When compared with the Pour plot of the free DML in Figure 1, they demonstrate the~significance of the closed-back DMZ in its improved directivity.
AMENDED SHEET
m.. .....,....... ..,~~NL~~ 06 . 1_ g_CA 0232619 2000-09-21480 484405-. +49 89 wo 9~fsz3zz rcT~c~99~o~oa~
z2 To investigate fur~her the nature and the effect of enclosure on the panel beraviour, especially at the combined system resonance, 3 special dig wa$ made to allow the measurement of the internal pressure of the enclosure at nine predetermined points as shown in Figure ~1. fhe microphone was inserted in the holes provided within zhe back-plate of an .F15 enclosure j ig at a predetermined depth, while the other eight position holes were tightly blocked with hard rubber grommets. The microphone was mechanically 1f3 isolated from the enclosure by an appropriate rubber grommet during the measurement..
Frp;Tc this data, a contour plot was created to shoal the pressure distribution at system resonance and at eithex side of this frequency as shown in Figures 22a to c_ The pressure frequency response was also plotted for the nine positions as shot~n in Figure 23. This graph eXhlblts good definition in the region of resonance for all curves associated with the measurement points within the enclosure. However, the pressure tends to vary across the 2D enclosure cross-sectional area as the frequency is increased.
The normal component of velocity and displacement across the panels was measured with a Scanning Laser Vibrometer. The velocity and displacement distribution across the panels were plotted to investigate the behaviour of the panel around the coupled system resonance. The results were dccumented and a number of the cases are shown in Figures 2~a to d. These results suggest a timpanic AMENDED SHEET
modal behaviour of the panel at resonance, with the whole of the panel moving, albeit at a lesser velocity and displacement as one moves towards the panel edges.
In practice this behaviour is consistent for all boundary conditions of the panel, although the mode shape will vary from case to case depending on a complex set of parameters, including panel stiffness, mass, size and boundary conditions. In the limit and for an infinitely rigid panel, this system resonance will be seen as the fundamental rigid body mode of the piston acting on the stiffness of the enclosure air volume. It was found to be convenient to call the DML system resonance, the 'Whole Body Mode' or WBM.
The full theoretical derivations of the coupled system has been implemented in a suite of software by New Transducers Limited. A version of this package was used to simulate the mechanoacoustical behaviour of our test objects in this paper. This package is able to take into account all the electrical, mechanical and acoustical variables associated with a panel, exciter(s) and mechanoacoustical interfaces with a frame or an enclosure and predict, amongst other parameters, the far-field acoustic pressure, power and directivity of the total system.
Figure 25a shows the log-velocity spectrum of a free radiating, A5-1 panel clamped in a frame, radiating in free space equally from both sides. The solid line represents the simulation curve and the dashed line is the measure velocity spectrum. At low frequencies the panel goes in resonance with the exciter. The discrepancy in the frequency range above 1000 Hz is due to the absence of the free field radiation impedance in the simulation model.
Figure 25b shows the same panel as in Figure 25a but this time loaded with two identical enclosures, one on each side of the panel, with the same cross-section as the panel and a depth of 24mm. A double enclosure was designed and used in order to exclude the radiation impedance of free field on one side of the panel and make the experiment independent of the free field radiation impedance. It is important to note that this laboratory set-up was used for theory verification only.
In order to enable velocity measurement of the panel, the back walls of the two enclosures were made from a transparent material to allow access by the laser beam to the panel surface. This test was repeated using panel A5-3 Rohacell without skin, with different bulk properties and the result is shown in Figures 26a and b. In both cases simulation was performed using 200 point logarithmic range, whilst the laser measurement used 1600 point linear range.
From the foregoing theory and work, it is clear that a small enclosure fitted to a DML will bring with it, amongst a number of benefits, a singular drawback. This manifests itself in an excess of power due to WBM at the system resonance as shown in Figures 27a and b. It is noteworthy that apart from this peak, in all other aspect the enclosed DML can offer a substantially improved performance including increased power bandwidth.
It has been found that in most cases a simple second order band-stop equalisation network of appropriate Q
matching that of the power response peak, may be designed to equalise the response peak. Furthermore in some cases a single pole high-pass filter would often adjust for this by tilting the LF region, to provide a broadly flat power response. Due to the unique nature of DML panels and their resistive electrical impedance response, whether the filter is active or passive, its design will remain very simple. Figure 28a shows where a band-stop passive filter has been incorporated for equalisation. Further examples may be seen in Figures 28b and c that show simple pole EQ
with a capacitor used in series with the loudspeakers.
When a free DML is used near and parallel to a wall, special care must be taken to ensure minimal interaction with the latter, due to its unique complex dipolar characteristics. This interaction is a function of the distance to the boundary, and therefore, cannot be universally fixed. Full baffling of the panel has definite advantages in extending the low frequency response of the system, but this may not be a practical proposition in a large number of applications.
A very small enclosure used with a DML will render it independent of its immediate environment and make the system predictable in its acoustical performance. The mathematical model developed demonstrates the level of complexity for a DML in the coupled system. This throws a sharp contrast between the prediction and design of a DML
and that of the conventional piston radiator. Whilst the mechanoacoustical properties of a cone-in-box may be found by relatively simply calculations (even by a hand calculator) those associated with a DML and its enclosure are subject to complex interactive relationships which render this system impossible to predict without the proper tools.
The change in system performance with varying enclosure volume is quite marked in the case where the depth is small compared with the panel dimensions. However, it is also seen that beyond a certain depth the increase in LF response become marginal. This of course is consistent with behaviour of a rigid piston in an enclosure. As an example, an A2 size panel with 50mm enclosure depth can be designed to have a bandwidth extending down to about 120Hz, Figure 24.
Another feature of a DML with a small enclosure is seen to be a significant improvement in the mid and high frequency response of the system. This is in many of the measured and simulated graphs in this paper and of course anticipated by the theory. It is clear that the increase in the panel system modality is mostly responsible for this improvement, however, enclosures losses might also influence this by increasing the overall damping of the system.
As a natural consequence of containing the rear radiation of the panel, the directivity of the enclosed system changes substantially from a dipolar shape to a near cardioid behaviour as shown in Figure 17. It is envisaged that the directivity associated with a closed-back DML may find use in certain applications where stronger lateral coverage is desirable.
Power response measurements were found to be most useful when working with the enclosed DM system, in order to observe the excessive energy region that may need compensation. This is in line with other work done on DM
loudspeakers, in which it has been found that the power response is the most representative acoustic measurement correlating well to the subjective performance of a DML.
Using the power response, it was found that in practice a simple band-pass or a single pole high-pass filter is all that is needed to equalise the power response in this region.
Claims (7)
1. An acoustic device comprising a resonant bending wave acoustic panel having opposed faces, means defining a cavity enclosing at least a portion of one panel face and arranged to contain acoustic radiation from the said portion of the panel face, characterised in that the cavity is sufficiently shallow that the rear face of the cavity facing the said one panel face causes fluid coupling to the panel such that X and Y cross modes in the fluid are generally dominant, and such as to modify the modal behaviour of the panel.
2. An acoustic device according to claim 1, wherein the cavity is sealed.
3. An acoustic device according to claim 1 or claim 2, wherein the ratio of the cavity volume to enclosed panel area (ml:cm2) is in the range 10:1 to 0.2:1.
4. An acoustic device according to any preceding claim, wherein the panel is mounted in and sealed to the cavity defining means by a peripheral surround.
5. An acoustic device according to claim 4, wherein the surround is resilient.
6. A loudspeaker comprising an acoustic device as claimed in any preceding claim, and having a vibration exciter arranged to apply bending wave vibration to the resonant panel to produce an acoustic output.
7. A method of multiplying the modal behaviour of a resonant bending wave panel acoustic device, comprising bringing the resonant panel into close proximity with a boundary surface to define a resonant cavity therebetween in which X anal Y cross modes in the cavity are dominant.
Applications Claiming Priority (3)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| GB9807316.6 | 1998-04-07 | ||
| GBGB9807316.6A GB9807316D0 (en) | 1998-04-07 | 1998-04-07 | Loudspeaker |
| PCT/GB1999/001048 WO1999052322A1 (en) | 1998-04-07 | 1999-04-06 | Acoustic device |
Publications (1)
| Publication Number | Publication Date |
|---|---|
| CA2326193A1 true CA2326193A1 (en) | 1999-10-14 |
Family
ID=10829902
Family Applications (1)
| Application Number | Title | Priority Date | Filing Date |
|---|---|---|---|
| CA002326193A Abandoned CA2326193A1 (en) | 1998-04-07 | 1999-04-06 | Acoustic device |
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| Country | Link |
|---|---|
| EP (1) | EP1070437B1 (en) |
| JP (1) | JP2002511681A (en) |
| KR (1) | KR20010042491A (en) |
| CN (1) | CN100417304C (en) |
| AR (1) | AR019019A1 (en) |
| AT (1) | ATE334567T1 (en) |
| AU (1) | AU3340099A (en) |
| BG (1) | BG104811A (en) |
| BR (1) | BR9909496A (en) |
| CA (1) | CA2326193A1 (en) |
| DE (1) | DE69932507T2 (en) |
| EA (1) | EA200001038A1 (en) |
| GB (1) | GB9807316D0 (en) |
| HK (1) | HK1030327A1 (en) |
| HU (1) | HUP0103916A3 (en) |
| ID (1) | ID27518A (en) |
| IL (1) | IL138310A0 (en) |
| NO (1) | NO20005056L (en) |
| NZ (1) | NZ506732A (en) |
| PL (1) | PL343440A1 (en) |
| SK (1) | SK14932000A3 (en) |
| TR (1) | TR200002920T2 (en) |
| TW (1) | TW462201B (en) |
| WO (1) | WO1999052322A1 (en) |
| ZA (1) | ZA200004675B (en) |
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| GB0114501D0 (en) * | 2001-06-14 | 2001-08-08 | New Transducers Ltd | Mobile telephone |
| DE50115744D1 (en) * | 2001-07-13 | 2011-01-27 | Harman Becker Automotive Sys | PANEL SPEAKER |
| US8284955B2 (en) | 2006-02-07 | 2012-10-09 | Bongiovi Acoustics Llc | System and method for digital signal processing |
| US11431312B2 (en) | 2004-08-10 | 2022-08-30 | Bongiovi Acoustics Llc | System and method for digital signal processing |
| US10158337B2 (en) | 2004-08-10 | 2018-12-18 | Bongiovi Acoustics Llc | System and method for digital signal processing |
| US10848118B2 (en) | 2004-08-10 | 2020-11-24 | Bongiovi Acoustics Llc | System and method for digital signal processing |
| JP4072542B2 (en) * | 2005-03-14 | 2008-04-09 | Ｎｅｃアクセステクニカ株式会社 | Speaker integrated display |
| US10701505B2 (en) | 2006-02-07 | 2020-06-30 | Bongiovi Acoustics Llc. | System, method, and apparatus for generating and digitally processing a head related audio transfer function |
| US10848867B2 (en) | 2006-02-07 | 2020-11-24 | Bongiovi Acoustics Llc | System and method for digital signal processing |
| US9615189B2 (en) | 2014-08-08 | 2017-04-04 | Bongiovi Acoustics Llc | Artificial ear apparatus and associated methods for generating a head related audio transfer function |
| US11202161B2 (en) | 2006-02-07 | 2021-12-14 | Bongiovi Acoustics Llc | System, method, and apparatus for generating and digitally processing a head related audio transfer function |
| US10069471B2 (en) | 2006-02-07 | 2018-09-04 | Bongiovi Acoustics Llc | System and method for digital signal processing |
| JP2010206557A (en) * | 2009-03-04 | 2010-09-16 | Panasonic Corp | Building material |
| KR101496911B1 (en) * | 2009-03-23 | 2015-02-27 | 삼성전자 주식회사 | Sound test system and method for testing the same |
| CN102711029A (en) * | 2012-01-09 | 2012-10-03 | 瑞声声学科技(深圳)有限公司 | Testing method of intrinsic noise voltage of microphone and testing device of testing method |
| US9883318B2 (en) | 2013-06-12 | 2018-01-30 | Bongiovi Acoustics Llc | System and method for stereo field enhancement in two-channel audio systems |
| US9264004B2 (en) | 2013-06-12 | 2016-02-16 | Bongiovi Acoustics Llc | System and method for narrow bandwidth digital signal processing |
| US9906858B2 (en) | 2013-10-22 | 2018-02-27 | Bongiovi Acoustics Llc | System and method for digital signal processing |
| US10820883B2 (en) | 2014-04-16 | 2020-11-03 | Bongiovi Acoustics Llc | Noise reduction assembly for auscultation of a body |
| US10639000B2 (en) | 2014-04-16 | 2020-05-05 | Bongiovi Acoustics Llc | Device for wide-band auscultation |
| US9615813B2 (en) | 2014-04-16 | 2017-04-11 | Bongiovi Acoustics Llc. | Device for wide-band auscultation |
| US9564146B2 (en) | 2014-08-01 | 2017-02-07 | Bongiovi Acoustics Llc | System and method for digital signal processing in deep diving environment |
| US9638672B2 (en) | 2015-03-06 | 2017-05-02 | Bongiovi Acoustics Llc | System and method for acquiring acoustic information from a resonating body |
| US9621994B1 (en) | 2015-11-16 | 2017-04-11 | Bongiovi Acoustics Llc | Surface acoustic transducer |
| JP2018537910A (en) | 2015-11-16 | 2018-12-20 | ボンジョビ アコースティックス リミテッド ライアビリティー カンパニー | Surface acoustic transducer |
| KR101704517B1 (en) | 2016-03-28 | 2017-02-09 | 엘지디스플레이 주식회사 | Display device for generating sound by panel vibration type |
| KR102663406B1 (en) | 2016-04-04 | 2024-05-14 | 엘지디스플레이 주식회사 | Sound generation actuator of panel vibration type and double faced display device with the same |
| KR20170114471A (en) | 2016-04-05 | 2017-10-16 | 엘지디스플레이 주식회사 | Organic light emitting display device |
| US10219079B2 (en) | 2016-03-28 | 2019-02-26 | Lg Display Co., Ltd. | Display device for generating sound by vibrating panel |
| US11211043B2 (en) | 2018-04-11 | 2021-12-28 | Bongiovi Acoustics Llc | Audio enhanced hearing protection system |
| US10959035B2 (en) | 2018-08-02 | 2021-03-23 | Bongiovi Acoustics Llc | System, method, and apparatus for generating and digitally processing a head related audio transfer function |
| US20200388265A1 (en) * | 2019-06-10 | 2020-12-10 | Toyota Motor Engineering & Manufacturing North America, Inc. | Sound isolation device |
| KR102215716B1 (en) * | 2019-10-21 | 2021-02-18 | 삼원액트 주식회사 | Panel with sound reproduction |
| AT525365B1 (en) | 2022-05-25 | 2023-03-15 | Cale3D Prime Gmbh | Electroacoustic Transducer |
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| US4914750A (en) * | 1987-07-13 | 1990-04-03 | Avm Hess, Inc. | Sound transducer |
| ATE179045T1 (en) * | 1995-09-02 | 1999-04-15 | New Transducers Ltd | INERTIAL VIBRATION TRANSDUCER |
| UA51671C2 (en) * | 1995-09-02 | 2002-12-16 | Нью Транзд'Юсез Лімітед | Acoustic device |
| EA000836B1 (en) * | 1995-09-02 | 2000-04-24 | Нью Трэнсдьюсерз Лимитед | Inertial vibration transducer |
-
1998
- 1998-04-07 GB GBGB9807316.6A patent/GB9807316D0/en not_active Ceased
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1999
- 1999-04-06 NZ NZ506732A patent/NZ506732A/en unknown
- 1999-04-06 SK SK1493-2000A patent/SK14932000A3/en unknown
- 1999-04-06 CA CA002326193A patent/CA2326193A1/en not_active Abandoned
- 1999-04-06 HU HU0103916A patent/HUP0103916A3/en unknown
- 1999-04-06 BR BR9909496-7A patent/BR9909496A/en not_active Application Discontinuation
- 1999-04-06 AU AU33400/99A patent/AU3340099A/en not_active Abandoned
- 1999-04-06 WO PCT/GB1999/001048 patent/WO1999052322A1/en active IP Right Grant
- 1999-04-06 EA EA200001038A patent/EA200001038A1/en unknown
- 1999-04-06 DE DE69932507T patent/DE69932507T2/en not_active Expired - Lifetime
- 1999-04-06 JP JP2000542951A patent/JP2002511681A/en not_active Abandoned
- 1999-04-06 AR ARP990101546A patent/AR019019A1/en unknown
- 1999-04-06 EP EP99914672A patent/EP1070437B1/en not_active Expired - Lifetime
- 1999-04-06 ID IDW20001944A patent/ID27518A/en unknown
- 1999-04-06 CN CNB998047651A patent/CN100417304C/en not_active Expired - Fee Related
- 1999-04-06 AT AT99914672T patent/ATE334567T1/en not_active IP Right Cessation
- 1999-04-06 KR KR1020007011113A patent/KR20010042491A/en not_active Application Discontinuation
- 1999-04-06 TR TR2000/02920T patent/TR200002920T2/en unknown
- 1999-04-06 PL PL99343440A patent/PL343440A1/en unknown
- 1999-04-06 IL IL13831099A patent/IL138310A0/en unknown
- 1999-04-12 TW TW088105770A patent/TW462201B/en not_active IP Right Cessation
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2000
- 2000-09-06 ZA ZA200004675A patent/ZA200004675B/en unknown
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- 2000-10-06 NO NO20005056A patent/NO20005056L/en not_active Application Discontinuation
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Also Published As
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| BR9909496A (en) | 2000-12-12 |
| NO20005056L (en) | 2000-12-06 |
| JP2002511681A (en) | 2002-04-16 |
| AR019019A1 (en) | 2001-12-26 |
| WO1999052322A1 (en) | 1999-10-14 |
| ZA200004675B (en) | 2002-02-27 |
| IL138310A0 (en) | 2001-10-31 |
| BG104811A (en) | 2001-07-31 |
| CN1296719A (en) | 2001-05-23 |
| DE69932507T2 (en) | 2007-07-19 |
| EP1070437B1 (en) | 2006-07-26 |
| TR200002920T2 (en) | 2000-12-21 |
| NO20005056D0 (en) | 2000-10-06 |
| EP1070437A1 (en) | 2001-01-24 |
| HUP0103916A2 (en) | 2002-03-28 |
| GB9807316D0 (en) | 1998-06-03 |
| HK1030327A1 (en) | 2001-04-27 |
| NZ506732A (en) | 2002-11-26 |
| TW462201B (en) | 2001-11-01 |
| ID27518A (en) | 2001-04-12 |
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| EA200001038A1 (en) | 2001-04-23 |
| ATE334567T1 (en) | 2006-08-15 |
| CN100417304C (en) | 2008-09-03 |
| SK14932000A3 (en) | 2001-05-10 |
| AU3340099A (en) | 1999-10-25 |
| PL343440A1 (en) | 2001-08-13 |
| HUP0103916A3 (en) | 2002-12-28 |
| DE69932507D1 (en) | 2006-09-07 |
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| Date | Code | Title | Description |
|---|---|---|---|
| FZDE | Discontinued |