Abstract:
An electromechanical force transducer comprising a plurality of resonant elements, a low stiffness member coupled between the adjacent faces of at least two adjacent resonant elements, and a stub member on which the resonant elements are supported and for coupling the transducer to a site to which force is to be applied. An electromechanical force transducer comprising a plate-like resonant element having a frequency distribution of modes in the operative frequency range of the transducer, and a stub member for coupling the transducer to a site to which force is to be applied and on which the resonant element is supported and arranged such that whole body non-bending modes are introduced into the resonant element. An electromechanical force transducer comprising a resonant element; a coupling member on the resonant element for mounting the transducer to a site to which vibration force is to be applied or taken; the transducer further comprising a member for increasing the rotational impedance presented to the coupling member adjacent said site.

Description:
[0001]     This application is a continuation-in-part of International patent application No. PCT/GB2004/003843, filed Sep. 9, 2004, which is incorporated herein by reference in its entirety. 
     
    
     TECHNICAL FIELD  
       [0002]     The invention relates to electromechanical force transducers, actuators, exciters and the like devices and more particularly but not exclusively, to such devices for use in acoustic apparatus, e.g. loudspeakers and microphones.  
       BACKGROUND  
       [0003]     The invention relates particularly, but not exclusively, to electromechanical force transducers of the kind described in International patent application WO01/54450 to the present applicant (incorporated herein by reference), and comprising one or more resonant elements or beams having a frequency distribution of modes in the operative frequency range of the transducer. Such transducers are known as “distributed mode actuators” or DMA for short.  
         [0004]     It is an object of the invention to provide a transducer in which damping is provided to result in a reduction of Q of the modes and a reduction in the severity of cancellation between modes to give an increased smoothness of acoustic pressure.  
         [0005]     It is also an object of the invention to improve the robustness of the transducer e.g. to give a reduction of chance of failure during drop or impact tests.  
         [0006]     Another object of the invention is to reduce the first resonant mode frequency of an actuator or transducer, e.g. a DMA transducer.  
         [0007]     Yet another object of the invention is to reduce the sensitivity of transducer performance to the properties of the panel to which it is attached.  
       SUMMARY DISCLOSURE OF THE INVENTION  
       [0008]     From one aspect, the invention is a transducer of the kind described wherein a low stiffness layer is inserted between, and bonded to the adjacent faces of a plurality of resonant elements. We have found that simply adding a damping layer to one face of a resonant element or beam gives poor damping performance as the layer stretches with the element as the element face changes dimensions. However, using a flexible reference layer with a high resistance to dimensional change, such as a foil, on the other side of the damping layer results in an improvement in damping as the damping layer now shears between the changing element face dimension and the non-stretching foil. If the reference layer can be made to change dimension in opposition to the damped face, the damping effect will be doubled. This is the effect gained by adhering the damping layer between adjacent element faces.  
         [0009]     From another aspect the invention is a DMA transducer wherein out of plane DMA modes are introduced into the audio band.  
         [0010]     From yet another aspect, the invention provides a bending inertial vibration transducer comprising a resonant element; a coupling member on the resonant element for mounting the transducer to a site to which vibration force is to be applied or taken; the transducer further comprising a member for increasing the rotational impedance presented to the coupling member adjacent said site.  
         [0011]     The member for increasing the rotational impedance may be configured as a foot extending laterally from the coupling member, i.e. transversely to the first direction in which the vibration force is applied or taken.  
         [0012]     The foot may be integral with the coupling member. The foot may extend further in a lateral direction than does the coupling member.  
         [0013]     The resonant element may be attached to the coupling member as a cantilever. The foot may extend from the coupling member in the same direction as the cantilever. The foot may extend from the coupling member solely in the same direction as the cantilever. 
     
    
     BRIEF DESCRIPTION OF THE DRAWING  
       [0014]     Embodiments that incorporate the best mode for carrying out the invention are described in detail below, purely by way of example, with reference to the accompanying drawing, in which:  
         [0015]      FIG. 1  is a side view of a first embodiment of electromechanical force transducer of the present invention;  
         [0016]      FIG. 2   a  is a side view of part of an electromechanical force transducer;  
         [0017]      FIG. 2   b  is a side view of a first embodiment of electromechanical force transducer of the present invention;  
         [0018]      FIG. 3  is a graph comparing blocked force of a single beam transducer and the transducer of  FIG. 1 ;  
         [0019]      FIG. 4  is a graph comparing acoustic pressure between an undamped double beam DMA, a ½ damped DMA (that is with damping material bonded between the resonant elements over half the lengths of the resonant elements) and a fully damped double beam DMA transducer;  
         [0020]      FIG. 5  is a side view of a single beam actuator;  
         [0021]      FIG. 6  is a side view of a second embodiment of electromechanical force transducer of the present invention;  
         [0022]      FIG. 7  is a graph comparing blocked force under different conditions;  
         [0023]      FIG. 8   a  is a graph comparing acoustic pressure under different conditions;  
         [0024]      FIG. 8   b  is a perspective view of a transducer of the kind shown in  FIG. 6  mounted at a panel edge, and  
         [0025]      FIG. 9  is a graph comparing blocked force with different compliant stubs;  
         [0026]      FIGS. 10A  and B show the variation with frequency F in sound pressure level generated by a transducer applied to a thinner panel and a thicker panel respectively;  
         [0027]      FIG. 11  is a perspective framework view of an embodiment of a third aspect of the present invention;  
         [0028]      FIG. 12  is a rear elevation of the embodiment of  FIG. 11 ;  
         [0029]      FIG. 13  is a similar view to that of  FIG. 12  and showing a further embodiment of a third aspect of the present invention. 
     
    
     DETAILED DESCRIPTION  
       [0030]      FIG. 1  shows a double beam transducer of the kind generally described in WO01/54450, the text of which is incorporated in the present application. The transducer ( 1 ) comprises a first piezoelectric beam ( 2 ) on the back of which is mounted a second piezoelectric beam ( 3 ) by connecting means in the form of a rigid stub ( 4 ) located near to the centre of both beams. Each beam is a bimorph.  
         [0031]     The transducer ( 1 ) is mounted on a structure ( 5 ), e.g. a bending-wave loudspeaker panel, e.g. a distributed mode loudspeaker (DML), by coupling means in the form of a rigid stub ( 6 ) located near to the centre of the first beam.  
         [0032]     In the present invention a low stiffness layer ( 7 ) of foamed plastics is bonded between adjacent faces of the two beams ( 2 , 3 ). The bonded layer may cover substantially the whole of the adjacent faces or may be discontinuous, e.g. to damp certain modes.  
         [0033]     The following sets out some parameters for one suitable foam damping material.  
         [0034]     “Poron” slow rebound foam polyurethane plastics material.  
         [0035]     Type: 4790-92-25041-04S.  
         [0036]     Thickness: 1.05 mm (we have also tried 1.0 mm with success).  
         [0037]     Density: 400 kg/m3.  
         [0038]     Compressional E (Young&#39;s Modulus with the foam in  
         [0039]     compression)=2 MPa at 1 kHz.  
         [0040]     The measured resistance, R, is approx 8×10 5  Ns/m3. These figures are the measured ‘real’ part of the mechanical resistance when in compression, not shear. Shear figures are not available.  
         [0041]     Use of a thinner foam (0.6 mm) also gave good results. A thicker foam, say up to 1.5 mm would be expected to give good results with this material. We suggest thickness limits between 0.3 and 2.0 mm.  
         [0042]     The density (in isolation from E and R) is expected to be irrelevant, and could vary by a factor of 100 and have little effect. E is important but the shearing that is occurring makes the importance of E difficult to identify. We suggest a factor of 4 increase in E would start to stiffen the beam, so is to be avoided. A reduction of E would have little effect as it appears the system stiffness is not being affected too much by the addition of the foam. The R figure is important. Reducing R is expected to effect damping in a linear fashion. We suggest that it is not reduced by more than a factor of say 4. Increasing R is good but cannot be achieved without affecting the other parameters.  
         [0043]      FIG. 2  shows the effect of bonding to one face or to both faces of multibeam transducer.  FIG. 2   a  shows the case where the damping layer ( 7 ) is only bonded to one beam ( 2 ). When the other beam ( 3 ) moves in relation to ( 2 ), it slides over the upper surface of the damping layer, which therefore does not deform and adds little damping to the bending resonances. However, in  FIG. 2   b,  the damping layer is bonded to both beams, and so is forced into shear by the relative movement of beam ( 3 ) in relation to beam ( 2 ). It is this shearing which applies damping.  
         [0044]     The beam lengths need not be the same but maximum damping effect is expected if they are. The measured effect of adding a damping layer between two beams on the blocked force of a centrally mounted transducer is shown in  FIG. 3 . The Q of all modes is reduced and the natural frequencies have not changed implying extremely low stiffness of bond material ( 7 ). Adding the damping layer increases output when cancellation inside the transducer is occurring, such as between the resonances of dissimilar length beams.  
         [0045]      FIG. 4  shows the simulated effect on acoustic pressure of adding a damping between the faces of a 36 mm/34 mm beam length DMA transducer. Output at the transducer fundamental is slightly reduced, but a broad increase in output occurs in 3-4 kHz region. This is the region of internal cancellation in the transducer. The acoustic pressure response is also smoother.  
         [0046]     Drop test failure rates are expected to be reduced. At impact most of the energy will be present in the exciter at its fundamental resonance. Since the damping reduces the Q of this resonance, the instantaneous maximum displacement will be reduced, resulting in reduced stress in the beam. This stress reduction is expected to improve drop test reliability. In addition, the build height of the transducer can be reduced by the present invention.  
         [0047]     The stub used to couple a transducer of the kind described above to its load is stiff in all 3 Cartesian axes and rotational stiffness is usually ignored, and is assumed to be high. For the case of a beam with stub position halfway along its length, 0 rotation occurs at the stub for the beam fundamental resonance. If this 0 rotation boundary condition is replicated at the end of a half length beam the fundamental will occur at the same frequency as the full length beam, with half the force. This is the cantilever condition, see  FIG. 5 .  FIG. 5  is a diagram showing fundamental mode shape of a cantilever beam (that is an extreme offset stub). The displaced shape shows pure bending motion.  
         [0048]     However by reducing the stub rotational stiffness from this high value to a lower one, the f 0  of the beam drops and becomes less dependent on bending motion of the beam and more rigid body-like, see  FIG. 6 .  FIG. 6  is a diagram of a modeshape of a beam coupled to a panel with a soft stub allowing rotation of the beam, the modeshape showing some bending in the beam and some rotational translations. In the limiting case of a rotational stiffness of 0, the mode drops to 0 Hz and is a rigid body mode. Reference ( 9 ) represents a trapped air layer behind the panel ( 5 ), which in the simulation couples to the panel and affects the modal set of resonances in the panel, and reference ( 10 ) represents the body of a cell phone containing a loudspeaker formed by the panel ( 5 ) and transducer ( 1 ). It will be noted that the deflection of the beam ( 2 ) is greatly exaggerated so that it is visible.  
         [0049]     By choosing this rotational compliance the f 0  of the beam can be lower than the f 0  of a beam twice its length, mounted at its centre—FE analysis has been used to show this effect, see  FIG. 7 .  FIG. 7  is a graph of simulated blocked force generated by 3 conditions: a 36 mm beam centrally mounted, a half length beam with stiff stub at end and half length beam with compliant stub at the end. The hard stub case causes a stiffening of the beam, effectively reducing its length slightly.  
         [0050]     A solid stub will have the same stiffness in the 3 translational and rotational axes. By suitably profiling the cross-sectional shape of the stub, different stiffnesses in the 6 different axes can be generated. The result is that modes in the different axes occur at different frequencies. If the load impedance is asymmetric, modes involving movement in directions other than normal to the beam surface can couple into the panel, providing increased modal density, see  FIG. 8 .  FIG. 8   a  is a graph of simulated effect on acoustic pressure generated by changing stub stiffness.  FIG. 8   b  is a perspective view of a panel-form loudspeaker having a panel ( 5 ) with an attached transducer mounted on a soft stub ( 6 ) of I-beam section and showing the DMA moving in-plane. In the case of the in-plane mode illustrated in  FIG. 8 , this mode is not present if the rotational stiffness around the axis ( 8 ) normal to the plane of the panel is ignored. In this case the first mode is partly due to rotational stiffness around the axis along the short edge of the beam, the second mode is due to the stiffness around the axis normal to the beam. The last rotational axis, around the axis moving along the length of the beam will also generate a mode.  
         [0051]     An example of a stub shape giving different stiffnesses in different axes is an I-section, see  FIG. 9 .  FIG. 9  is a graph of simulated effect on blocked force of polycarbonate I-section stub with varying vertical bar lengths. The stub is 3 mm wide in total with inner bar of 1 mm width, bar length being specified on the plot.  
         [0052]     By changing the fundamental resonance from a purely bending motion in the beam to a partly translatory motion, the stress in the beam is reduced at the fundamental. Since the fundamental resonance will receive the most energy during impact, the beam is more likely to survive without damage as most of the deformation will occur in the stub.  
         [0053]     Although a stub of I-beam section has been described, many other stub cross-sections could be used, for example, trapezoidal, cylindrical and so forth.  
         [0054]     When a transducer of the kind mentioned above is mounted as shown in  FIG. 1 , namely on a bending-wave acoustic panel of the kind known from WO97/09842 (incorporated herein by reference), it may show bandwidth sensitivity to panel impedance as a result of the fundamental frequency, f 0 , of the transducer depending on the rotational impedance presented to its stub or coupling member. This will be evident from a comparison of  FIGS. 10A  and B, which show the variation with frequency F in sound pressure level (in dB, measured at a distance of 10 cm, 1V RMS) generated by a transducer applied to a thinner panel and a thicker panel respectively.  
         [0055]     The dashed line in each figure shows the response for the loudspeaker arrangement shown in  FIGS. 11 and 12  in which a transducer  1  comprises first and second piezoelectric beams  2 , 3  attached in cantilever fashion to a stub  6 . The stub has an integral foot  12  extending transversely in the same direction as the cantilever piezoelectric beams. Stub and foot are in turn connected to panel  5  to apply a force substantially in direction  15 .  
         [0056]     Foot  12  extends substantially transversely to direction  15  and solely in the same direction as the cantilever. It has a length L of 2 mm, a thickness of 0.7 mm and is made of plastic, for example polycarbonate or an engineering plastic such as Grilamid™ or Grivory™ sold by EMS-Grivory. The latter material has a Young&#39;s Modulus of 68 GPa, many times larger than polycarbonate, and a density of 2700 kg/m3.  
         [0057]     Panel  5  is made of polycarbonate of typically 1 mm thickness and has at its rear surface a relatively thick air pocket  13  that extends across the panel at that end of the panel at which the transducer  1  is mounted and a thinner air pocket  14  extending over the remainder of the panel.  
         [0058]     The dashed line of  FIG. 10A  shows results for the thinner panel having 0.6 mm thickness whilst the dashed line of  FIG. 10B  shows results for the thicker panel having 1.5 mm thickness. It will be seen that in moving from the thinner panel to the thicker panel, the lowest (f 0 ) resonant frequency of the transducer increases by 130 Hz from 403 Hz to 533 Hz, effectively reducing the bandwidth of the loudspeaker, in particular at its lower end.  
         [0059]     The solid lines of  FIGS. 10A and 10B  illustrate how this increase can be reduced—from a jump of 130 Hz to a jump of 95 Hz (from an f 0  of 464 Hz to an f 0  of 559 Hz)—by the use of a foot having greater length and thickness of 4 mm and 1 mm respectively.  FIG. 13  illustrates the arrangement of  FIG. 12  incorporating a longer foot that extends further in the direction transverse to axis  15  than does the stub  6 . Such increased foot length and thickness result combine to present an increased rotational impedance to the transducer stub  6 , particularly adjacent the point at which the stub is attached to the panel. As evidenced by  FIGS. 10A  and B, increasing the rotational impedance presented to the stub reduces the sensitivity of transducer performance to the properties of the panel to which it is attached. Such a reduction in sensitivity may be particularly desirable where the transducer and/or bending wave member are mass-produced items and modifications to suit particular configurations would involve significant costs.