GB2217951A - Active control of sound in enclosures - Google Patents

Abstract

The invention relates to a method and apparatus for reducing sound transmission to a body contained within an enclosure. The shell defining the enclosure is provided with a discontinuity which fixes, relative to the shell, the displacement minima of vibrations (50) induced therein by an external sound field. The invention provides for monitoring of the enclosed medium vibrations and introducing an adaptively adjusted sound field into the enclosure such that the pressure maxima (31, 41) and minima of the medium are maintained in alignment with the displacement minima and maxima respectively of the structural vibration of the enclosure (as Fig. 5b). This action is applied in a manner which is most suitable for each vibration mode of the shell and the medium of primary concern within the relevant octave bands. By this method, the sound transmission to an enclosed body is greatly reduced. <IMAGE>

Description

ACTIVE CONTROL OF SOUND IN ENCLOSURES The present invention relates to a method and apparatus for the control and reduction of sound pressure levels within enclosures and particularly to the use of active noise control techniques in an enclosure containing a gaseous atmosphere.

Enclosures are often provided around delicate or sensitive equipment for the purpose of protecting it from the injurious effect of vibration caused by intense sound waves. In most practical cases, the sound field to which an enclosure is to be subjected has no fixed periodicity, being randomly distributed in terms of the waveform against time, although this external field may be resolved into orthogonal components according to conventional sound scattering principles around a solid body of resolution. The dynamic pressure loadings from these scattering harmonics act upon the external surface of the enclosure shell causing it to vibrate.

In general, a shell of substantial uniformity of construction can only vibrate in certain modes of dynamic displacement relative to the normal to the undisplaced surfaces. In the absence of damping, the frequencies at which these modes occur are known as natural or eigen frequencies. Although external pressure components can force excite the structure at any frequency, it is at these natural frequencies that the largest amplitude vibrations occur.

Considering an enclosure which is symmetrical about a central axis and substantially uniform in cross-section, for example an enclosure having cylindrical form; the mode displacement shapes around the circumference at these natural frequencies are of sinusoidal form. Along the axial direction, the modes may take either exponential or sinusoidal forms depending on how the enclosure is constrained with respect to its surroundings, and on whether cantilever-like displacement occurs or if, as is more usual, the displacement consists of local symmetrical bending or shear deflection, It has been found that, in most practical cases, both the axial and circumferential displacement modes take sinusoidal forms. Other displacement mode forms may occur but can generally be decomposed into a series of sinusoidal forms using Fourier analysis or other methods.Although there are an infinite number of modes (and hence natural frequencies), only a finite number will be found within the low frequency octave bands arising in external sound fields.

Similarly, the gaseous medium within an enclosure will also vibrate in a way that may be decomposed into certain normal modes: as the modes are of gas particle displacement they may be expressed as modes of sound perturbation pressure. Since there can be no substantial component of particle displacement normal to the virtually impervious interior surface of the enclosure, the modal pressure patterns arise through reflections of the contained sound waves. These pressure patterns may, in general, be resolved into three orthogonal components: when expressed in cylindrical coordinates they are the axial, circumferential and radial components.

As before, the axial and circumferential components take sinusoidal forms while the radial component takes a Neumann function form. When there is no body, that is, no equipment to be protected within the enclosure, the Neumann function becomes a Bessel function of the first kind. When there is a body having a substantially cylindrical form within the enclosure, the Neumann function form is appropriate to describe the pressure pattern and when, as would often be the case, the majority of the space within the enclosure is taken up by the body it contains, the radial component form becomes closely approximated to a sinusoidal form.

Hence, in practice, all three components are approximately sinusoidal in form.

Since there is relatively little to dampen these sound pressure modes within the enclosure (between the impervious inner wall of the enclosure and outer surface of the enclosed body) they tend to occur at the natural or eigen frequencies corresponding to each combination of possible mode orders. As before, there are an infinite number of such modes and hence natural frequencies but again only a finite number of these fall within the low frequency octave bands of interest.

There may be a large number of vibration modes of the enclosure structure and even more of the internal gaseous medium sound pressure perturbations. The number of modes (natural frequencies) per unit bandwidth is called the modal density (modes/Hz) and, for a given octave frequency band, the number of modes will generally increase with the band centre frequency.

Not all of the components of the external sound field will spatially match at the natural modes of the enclosure structure and not all will therefore couple strongly with it. Strong coupling can only occur where the axial dimension of the enclosure corresponds to an integer number of wavelengths while the same or another integer number of wave cycles 'fit' around the circumference.

Similarly, not all structural vibration modes of the enclosure structure will strongly couple with the enclosed medium. Again there must be close correspondence between the natural frequencies of the structure and enclosed medium. Due to this, only a fraction of the components in the external sound field will he able to couple strongly to the enclosed medium.

Depending on the gaseous medium used within the enclosure, there may still be a large number of strongly coupled modes within each octave frequency band of interest.

Active noise control techniques use one or more secondary sound sources or emitters (such as loudspeakers) in unoccupied space within an enclosure to inject an additional well-controlled and adaptively adjusted sound field into the enclosure in order to minimise the sound intensity within certain low frequency bands over as much of the internal space as possible. U.K. Patent GB 2 149 614B describes a number of methods of providing active noise control within an enclosure. Successful application of these methods depends on being able to measure a fundamental frequency f 0 and harmonics fl-fn for an incident sound field in order to reproduce them in antiphase and inject them into the enclosure.

Although these methods are capable of active noise control in large enclosures such as aircraft cabins, they rely on the incident sound field having a relatively constant periodicity. Such techniques could not hitherto be considered for such semi-diffuse random timedomain noise within a large fairing or enclosure: the accuracy of sound field cancellation could not practically be maintained over so many possible acoustic modes.

Only when the internal cavity becomes small enough or narrow enough to limit the number of possible modes to one or two and/or the frequencies of interest are low enough to again reduce the number of possible modes, can active noise control of random time domain noise be considered. Implementation of active noise control techniques may also be by means of suitable wide frequency-band response force actuators fitted at one or more optimum points around the enclosure structure. By vibrating the shell in a suitably controlled manner, the required cancelling sound field can be generated in the structure using the dynamic mechanisms which ordinarily transmit the sound through the enclosure passively.

One method for reducing the number of modes available for energy transfer within the acoustic medium in an enclosure is to have a more suitable gaseous medium than air within the enclosure.

The enclosed atmosphere need not be totally replaced by the advantageous gas for there to be a worthwhile reduction in available modes. Helium is one known and used gas which, being inert, will have no adverse chemical effect on any equipment placed in the enclosure, whilst reducing the number of available acoustic modes within the enclosure by a factor of approximately 1/27; this being the cube of the ratio of the speed of sound in air to that in the gas: We have appreciated that this reduction In the number of available modes makes the use of active noise control method feasible.

In accordance with thc invention, thcrc is provided a method for reducing sound transmission t-c a body contained within an enclosure, in which means are provided on a shell defining the enclosure to fix relative to the shell displacement minima of vibrations induced therein by an external sound ficld and in vibrations are generated in the medium contained within the shell, there being further provided means within the enclosure to fix pressure maxima of the vibrating medium relative to the shell, such that the potential pressure maxima of the vibrating internal medium are maintained in alignment with the displacement minime of the structural vibration of the enclosure.

Thus, transfer of acoustic energy from the vibration of the shell structure into the anclosed medium is minimised by optimising the modal locking. That is to say the maxima of the shell vibrations are "locked" into a pro-determined position relative to the maxima of the vibrations in the enclosed medium. The position of the structural vibrational modes of the shell of the enclosure is dictated by the shape and construction of the shell and the vibrations of the enclosed modium are monitored and controlled.

For a structure having a porfectly circular cross-section there is no fixed orientation of the circumforential modes in the structure's vibration displacements. The modal displacement maxima could be located at any orientation and may possibly b nonstationary with time. In a normal structure any discontinuity such as a corner or joint between sections will lock the structural modes in place so that the discontinuity corresponds to a node in the standing vibration. This discontinuity will not normally affect the orientation of the spatial modes within the enclosed medium. These can be 'locked' in position by means of fixed hardware baffles spanning most of the gap between the inner wall of the enclosure and any body contained therein to prevent or minimise movement of the modes.Such hardware baffles could be made from firm but flexible materials such as reinforced rubber sheet. The major problem with the use of hardware baffles is lack of adaptability; each time a new payload is carried in the enclosure, partial or total redesign of the baffle system is necessary.

When the modal displacement maxima coincide (that is to say, when the structural wave field in the enclosure and the acoustic wave field in the enclosed medium have the same spatial orientation of their respective circumferential orders) maximum energy transfer to the acoustic modes in the enclosed medium may occur. Conversely, if the acoustic wave field (the spatial modes within the enclosed medium) are displaced such that the two are mismatched with the displacement minima of the structural modes aligned with the perturbation pressure maxima of the internal medium vibration modes (as in the method of the invention), in theory, no energy transfer should take place. In practice, total cancellation of vibration in the enclosed medium is difficult to achieve although a useful degree of attenuation of the transmitted energy is still obtained.

The method of the invention could be used effectively in the space industry, to reduce noise pressure levels in the payload enclosures of disposable launch vehicles. Reduction by 10 decibels or more for the sound energy spectrum over the lower octave frequency bands of centre frequencies 31.5, 63 and 125Hz for example, would be of use to prevent damage to the payloads carried from the noise levels generated by the rocket exhaust. For the initial rocket firing sequence the method would be of particular use since the external sound levels are most intense over the first 15 seconds or so of this phase of the launch.

For an enclosure which humans or animals will occupy, the use of the method including the enclosed gaseous medium is not particularly suitable. This may be overcome however, by constructing the enclosure with inner and outer walls and applying the active noise control techniques in the cavity therebetween.

The method described is also generally applicable to cases where the enclosure structure excitation forces arise, not from an external sound field, but from mechanical dynamic loadings to some part or parts of the structure. The structure will still respond and its dominant energy transfer modes to the enclosed acoustic medium may still be controlled by the same modal locking techniques as described above, the action being applied in a manner most adaptively suitable for each vibration mode of the shell and medium of primary concern within the relevant octave bands.

Preferably the enclosure contains a gaseous medium such that the number of possible modes of vibration of the medium is less than in air. The method is preferably implemented by having detection sensors (such as microphones or accelerometers) mounted outside the enclosure to monitor the incident sound field or vibration.

Further detection sensors are mounted inside the enclosure to provide an error signal. A state of the art fast microprocessor operating in real time processes the signals from the detection and error sensors to generate a continuously adapted noise control signal. Following suitable filtering and power amplification, the signal is output from secondary sources (such as loudspeakers or force actuators) within the enclosure. This controlled noise and/or vibration in combination with a suitable gaseous medium controls the modal locking for the low frequency bands of interest, to reduce the sound wave intensity in the enclosure.

One way of ensuring the most advantageously uniform reduction of the sound field within the free space of the helium gas (or other enclosed medium) is to control the secondary sound sources or shell vibration by means of the least mean square error of the actual to ideal sound fields for the required enclosed volume.

It should be noted that, although only continuous random noise signals with suitable processors and predictive algorithms have been discussed, the same principles are also suitable for attenuation of noise generated by transient sound and force loadings including impulses. Separation of the component stages of a disposable space launch vehicle is one context where such acoustic shocks are known to cause problems.

One preferred application of the method of the invention in the area of orbital launch vehicles will now be described in detail with reference to the accompanying drawings in which: Fig. 1 shows a launch vehicle at the point of lift-off with a payload enclosure mounted atop the main body, with incident sound waves reflected from the ground.

Fig. 2 is a schematic example of structural vibration modes in a hollow cylinder.

Fig. 3 is an example of a response spectrum showing the changes in amplitude of vibration (V) with increasing frequency (F).

Fig. 4 shows the axial, radial and circumferential acoustic mode patterns for an enclosed cylindrical volume.

Fig. 5 is a diagrammatic section through a cylindrical enclosure showing the deflection of the shell structural modes and internal medium pressure modes in aligned and offset positions.

Fig. 6 is a schematic layout of a multiple secondary source feed-forward active noise control system.

Fig. 7 is a block diagram of the feed-forward active noise control system.

Fig. 8 is a block diagram representation of the generation of the signal at the I'th error sensor.

Referring now to Fig. 1, a disposable or recoverable launch vehicle 10 with additional booster rockets 12 is shown at the point of lift-off from a launch pad 20. The launch vehicle 10 has a payload fairing 14 at its upper end, which would normally contain a payload 16 such as a satellite or systems for observation and scientific exploration to be placed in orbit. The fairing 14 is usually constructed as a lightweight shell and may be formed from a composite of strong, light carbon fibre, glass fibre and epoxy resin bonded skins forming a sandwich with a lightweight aluminium honeycomb core.

Without such an enclosure 14 the payload 16 would not only become damaged by high speed aerodynamic flows whilst still within the Earth's atmosphere but would also risk failures, particularly of electronic equipment vital to the functioning of launcher 10 and payload 16 due to strong vibration responses to the acoustic energy transferred from the highly turbulent rocket exhaust 22 into the region occupied by the payload.

The launch pad 20 is shown having exhaust deflection chutes 23 around the base of the launch vehicle 10 to divert the exhaust thrust 22 away from the rocket motor exhaust nozzles 18. In addition to vibration transmitted through the body of the vehicle 10 to the payload enclosures 14, there is a component of acoustic noise which causes the enclosure 14 to vibrate and which is in the form of sound waves 24 generated from the turbulent exhaust 22 as modified by reflections from the ground. These incident sound waves 24 are most intense and thus, most damaging at and immediately after lift-off.

Due to local ground and launch pad reflection and scattering effects the sound waves 24 propagating from the rocket exhaust 22towards the payload fairing 14 are not simple spherical waves (which would be the output from a single point sound source). Instead they consist of a dominant, almost plane, series of wavefronts which arrive at the fairing 14 at a more or less well defined angle relative to the vertical axis 26 of the fairing 14. The mean angle of incidence is dependent on the frequency band since the rocket motor noise dominating each band originates from a region of turbulent mixing which exists at a different distance from the rocket motor exhaust nozzles 18. In general, the lowest frequency bands at any given time originate from regions well separated along the ground from the exhaust nozzles 18 for the case where the launch pad 20 has exhaust deflector chutes 23.This means that the lower the frequency, the greater the mean angle of incidence of the wave fronts 24 on the fairing 14. Generally, the angles of incidence range from 40 down to around 100 for higher frequencies or when the launch vehicle 10 is clear of the launch pad 20.

It is this incident noise 24 which may cause damage, particularly to sensitive electronic control equipment for the launch vehicle 10 and payload 16. Because there is a small amount of variation in the angle of incidence, the external sound field is random with respect to time and spatially not totally diffuse. The previously decribed method of modal locking controlled by the microprocessor circuitry of an active noise control system is therefore a viable alternative to additional layers of acoustic insulation and hardware baffles within the enclosure. For this application, helium is a suitable gas for use as the fairing enclosed medium as it will have no adverse chemical effect on the payload and electrical systems carried.

In Fig. 2 a cylindrical enclosure such as part of the payload fairing 14 is shown, illustrating how the circumferential 60 and axial 62 modes take substantially sinusoidal forms. These modes can be locked in place by discontinuities in the fairing structure (which may for example have to split or separate at some stage of the launch). The combined displacement 63 of the shell normal to its surface relative to the undisplaced surface for both outwards 64 and inwards 65 deflection can be seen.

The circumferential 70,-axial 72 and radial 74 components of the enclosed acoustic modes can be seen in Fig. 4 with the circumferential 70 and axial 72 components taking sinusoidal form whilst the radial component 74 takes a Bessel functional form with the outer wall maxima 'fitting' the circumferential form 70 and reducing to zero at the axis 26 of the enclosure. When a substantially cylindrical payload is enclosed, then the radial component also approximates to a sinusoidal form.

Fig. 3 is an example of a response spectrum showing variations in the amplitude of vibration (V) against frequency (F) for a large launcher fairing showing individual mode responses 66 for the three most significant octave frequency bands 68. A calculation for such a fairing of typical size and having air as the enclosed medium shows: Table 1: Octave Band Centre Frequency (Hz) 31.5 63 125 Number of Structure Modes 4 14 40 Number of Modes in Air Medium (empty) 2 7 36 Number of Modes in Air Medium (payload) 1 4 18 From Table 1 it can be seen that, for a fairing having air as the enclosed medium, there are too many modes (even with payload) for active noise control methods using currently available microprocessor circuitry to cope with.When, however, the enclosed medium is helium, the same calculation for the same fairing gives: Table 2: Octave Band Centre Frequency (Hz) 31.5 63 125 Number of Structure Modes 4 14 40 Number of Modes in He Medium (empty) O 0 2 Number of Modes in He Medium (payload) O 0 2 Thus the use of helium not only removes any chance of coupling between the internal medium and structural modes for the two lowest octave frequency bands (31.5 and 63Hz), it also makes the use of active noise control techniques feasible in the next frequency band (125Hz) by reducing the number of possible modes. Although the physical size of the enclosure may have an effect on the results, typical calculations have shown that the dominant energy transfer occurs for the two resonant modes shown on Table 2, and not for any form of non-resonant or forced vibration between modes in adjacent octave frequency bands.

Fig. 5 shows a cross-section through the enclosure 14, having structural vibrations 50 of a second order mode showing the outward 51 and inward 52 deflection of the enclosure shell. The instantaneous modal perturbation pressures (the pressure modes of the enclosed medium) are shown for both positive 30 and negative 40 pressures in the radial Bessel 32,42 and circumferential 31,31 components.

In Fig. 5a, the maxima of the structural modes represented by the deflection of the shell 51,52 and the maxima of the pressure modes 31,41 are matched: with the modes so aligned, the strongest possible coupling occurs such as to allow maximum transfer of the acoustic noise energy from the shell of the enclosure to the payload space within.

By use of the methods described in detail below, when the pressure modes have been reoriented, as in Fig. 5b, the minimum transfer of acoustic noise occurs, between the incident sound field and the medium enclosed within the fairing.

Fig. 6 shows a schematic arrangement of an active control system for achieving the desired modal locking. A number 'K' of detection sensors 86 (microphones) are placed such that the correlation between their signals and the error sensor signals is maximised, but with the maximum possible "propagation delay" between the arrival of signals at the detection sensors 86 and their subsequent arrival at the error sensors 84. The choice of detection sensor placement will therefore be a compromise between the requirements of correlation and sufficient delay to allow the processing of the detected signals.One possible choice is provided by placing the detection sensors (accelerometers in this case) on the vibrating surface of the enclosure, in which case little delay for processing the signals is provided, and the filters may have to provide an "optimal prediction" of future values of the error sensor signals. (This optimum prediction is automatically ensured by the algorithm described below). In this case it is likely that the number of detection sensors 86 will have to be at least equal to or greater than the number of dominantly contributing modes of vibration of the structural shell which excites the sound field. Furthermore they should be placed so as to efficiently detect these structural modes (i.e. accelerometers placed at antinodes of the dominant modes).In general the number "K" of detection sensors required will have to be at least equal to or greater than the effective number of independent (uncorrelated) primary sources contributing to the sound field. The number 'M' of secondary sources 80 required is in general at least equal to or greater than the number of dominantly contributing acoustic modes of the enclosure. These should be placed so as to most effectively couple with the acoustic modes of the enclosure. The number L of error signals should be at least equal to or greater than the number M of secondary sources and they should be placed so as to most effectively detect the acoustic modes of the enclosure (i.e. at pressure maxima). It is assumed the detection sensor signals are input to the matrix of filters H after appropriate conditioning, amplification and anti-alias filtering.The matrix is controlled by a fast microprocessor 92 which is also connected to the error sensors 84 in the enclosure 14. The outputs from H are passed through appropriate reconstruction filters and power amplification before being fed to the secondary sources 80 in the enclosure 14.

The K x M matrix H of adaptive digital filters 90 required can be determined by solving the optimisation problem of reducing the time average (expectation) of the sum of the squared error signals to a minimum. This problem is most easily addressed using the discrete time formulation presented by Elliott et al. (IEEE Transactions on Acoustics, Speech and Signal Processing, Vol ASSP35, No. 10, 1423-1434) which for this case, is generalised to include multiple detection sensor signals x(t) and assumes there is no feedback of secondary source outputs to the detection sensor inputs (a reasonable assumption for the detection sensor locations described above. The general problem can be represented in block diagram form as illustrated in Figure 7.In the diagram, A represents the matrix of transfer functions relating the error sensor signals d(t) to the primary noise source P; B represents the matrix of transfer functions relating the detection sensor signals x(t) to the primary noise source P; C represents the matrix of transfer functions relating the outputs of the filter matrix H (denoted by y(t)) to the error sensor signals due to secondary sources dl(t). The output of this feed forward noise control system, e(t), represents the L error sensor signals.

To aid the formulation of the optimisation problem, this block diagram can be rearranged to effectively invert the order of operation of the matrices H and C. Since these matrices will not in general commute, a different method of re-arrangement is necessary. This method may be demonstrated by the following argument. It must first be assumed that H and C are matrices of linear time invariant systems corresponding to the elements Hmk and CIm which are the frequency domain representations of these matrix elements. Working in the frequency domain, the vector y of filter outputs to the M secondary sources and the vector # of the corresponding signals at the L error sensors are given by y = H x and # = C y which can be written as y = [h1, h2, . .. hk] x,

where the hk are the columns of H and the CT are the rows of C.

C1@ This enables the signal at the l'th error sensor to be expressed as #1 = C1T (h1x1 + h2x2 + . . . hkxk) (2) which, transposing each term, can be alternatively written as #1 = h1T (c1x1) + h2T(c1x2) ... hkT(c1xk) (3) The result of passing the detection signal xk through the m'th element of the row c1 defines the "filtered reference signal" r1mk.

The equivalent block diagram representation of the generation of which follows from equation (3) is now shown in Figure 8. The vectors hk are defined in terms of the elements Hmk, since the production of the contribution to the I'th error sensor signal is due to the M secondary sources when the elements of the matrices H and C operate on the K detection sensor signals in reverse order.

Having effectively reversed the order of operation of the transfer functions, it is now possible to work in discrete time.

Thus, the signal at the n'th time increment at the l'th error sensor can be expressed as el(n) = dl(n) + r1T w (4) where the signal #1(n) = rT w has been expressed as a convolution of the filtered reference signals r lmk and the corresponding finite impulse lmk response representations of the filters Hmk. Thus r1T = [r111T r121T ... rlm1T r112T r122T..rlm2T ... r11kT r12kT ..rlmkT ] (5) wT = [w11T w21T ... wm1T w12T w22T ...wm2T ... w1kT w2kT ...wmkT] (6) where each of the component vectors r1mk constitutes a discrete time sequence convolved with a vector of filter tap weights w where r1mkT = [r1mk(n) r1mk(n-1) ... r1mk(n-I+1)] (7) wmk = [wmko wmk1 . .. wmk(I-1)] (8) and the relevant convolutions are given by

Thus the transfer functions Hmk have been represented by I'th order FIR filters.

The analysis now follows exactly that presented by Elliott et al (IEEE Trans ASSP-35, 1423-1434). Equation (4) above corresponds exactly to equation (8) of Elliott's paper, but the composite tap-weight vector w now includes the additional filters necessary to deal with multiple detection sensor signals, and r includes the additional necessary filtered reference signals.

Now defining the vector of L discrete time error signals as eT = [e1(n) e2(n) . . eL(n)] (10) and the vector of discrete time signals produced by the error sensors due to the primary sources only as dT = [d1(n) d2(n) dL (n)] (11) it can therefore be written e = d + R w (12) where the matrix R is defined by RT = Er1 r2 .. ... rL] (13j The quadratic cost function which it is now sought to minimise is defined by J = E[eTe] = E(dT d] + 2J E [RTd] + wTE [RTR]w (14) where E[ ] is the expectation operator.This function is a quadratic form which has the unique minimum defined as the optimal filter coefficient vector wopt = -[E [RTR] ]-1 E[RTd] (15) which corresponds to a minimum mean square error of T - E@ dTd] - E@dT@@ @E @RTR] ]-1 E[RTd] (16) min Elliott's stochastic gradient algorithm can now be used to find the optimum. Thus, based on the method of steepest descent, the filter coefficients are updated by an amount proportional to the negative of the gradient of the quadratic surface. Using an instantaneous estimate of the gradient however the coefficients are updated in accordance with the algorithm w(n+1) = w(n) - &alpha;R(n)e(n) where a is a convergence coefficient which must be chosen to provide the fastest adaptation of the digital filters without producing an unstable response.

By taking account of multiple error signals and reducing the number of modes in the enclosed medium, an active noise control method is provided which is capable of handling transient and random time-domain noise and vibration.

Although the system has been described in relation to orbital launch vehicles having cylindrical enclosures, it is generally applicable to any enclosure structure through which acoustic noise or mechanical vibration is transmitted by resonant mode coupling.

Similar principles would apply to non-cylindrical forms of enclosure, although a different set of structural vibration mode forms would have to be prevented by the active control system from efficiently transferring noise to, or generating noise within the enclosure. For a non-cylindrical enclosure, the modal locking and use of active control methods in the presence of a favourable acoustic medium such as helium would still be applicable.

Enclosures in which this system may suitably be used include rail freight and passenger compartments, motor vehicle bodies, airship gondolas, aircraft cabins including helicopters, and undersea vessels such as diving bells (where, since divers are often required to breath an oxygen-helium mixture, a jacket layer may not be necessary). The system may also be applied to industrial buildings and installations subject to high noise levels such as engine or rocket motor test cells, petrochemical process plants and nuclear power systems.

Claims (23)

1. A method for reducing sound transmission to a body contained within an enclosure, in which means are provided on a shell defining the enclosure to fix relative to the shell displacement minima of vibrations induced therein by an external sound field and in which vibrations are generated in the medium contained within the shell, there being further provided means within the enclosure to fix pressure maxima of the vibrating medium relative to the shell, such that the potential pressure maxima of the vibrating internal medium are maintained in alignment with the displacement minima of the structural vibration of the enclosure.

2. A method according to claim 1 in which the enclosure contains a gaseous medium such that the number of possible modes of vibration of the medium is less than in air.

3. A method according to claims 1 or 2 in which the gas used is chemically inert.

4. A method according to any of claims 1 to 3 in which vibrations are generated in the medium contained within the shell by means of one or more sound emitters.

5. A method according to any of claims 1 to 3 in which vibrations are generated in the medium contained within the shell by means of force actuators acting on the shell structure.

6. A method according to any preceding claim in which the intensity of vibrations within the medium in the enclosure is monitored and the vibrations generated varied in response to the monitored intensity.

7. A method according to any of claims 1 to 5 in which the intensity of vibrations within the shell structure is monitored and the vibration generated varied in response to the monitored intensity.

8. A method according to any of claims 1 to 5 in which the intensity of vibrations in the enclosed medium is monitored by one or more error sensors, the vibrations being generated in response to the monitored intensity of vibrations in the shell structure and adaptively varied in response to the error sensor signals.

9. A method according to claim 8 in which the generated vibrations are adaptively varied such that the mean value of the squared error sensor signals is minimised.

10. A method according to any preceding claim in which the means for fixing the displacement minima of the vibration induced in the shell is a discontinuity in the shell structure.

11. A method for reducing sound transmission to a body contained within an enclosure, the method being substantially as hereinbefore described with reference to the drawings.

12. Apparatus for reducing sound transmission to a body contained within an enclosure comprising means provided on a shell defining the enclosure to fix, relative to the shell, displacement minima of vibrations induced therein by an external sound field, and means within the enclosure to fix pressure maxima of the vibrating medium relative to the shell such that the potential pressure maxima of the vibrating medium are maintained in alignment with the displacement minima of the structural vibration of the enclosure.

13. Apparatus according to claim 12 in which the enclosure contains a gaseous medium such that the number of possible modes of vibration of the medium is less than in air.

14. Apparatus according to claims 12 or 13 in which the gas used is chemically inert.

15. Apparatus according to any of claims 12 to 14 in which vibrations are generated in the medium contained within the shell by means of one or more sound emitters.

16. Apparatus according to any of claims 12 to 14 in which vibrations are generated in the medium contained within the shell by means of force actuators acting on the shell structure.

17. Apparatus according to any of claims 12 to 16 including means to monitor the intensity of vibrations within the medium in the enclosure and to vary the generation of vibrations in response to the monitored intensity.

18. Apparatus according to any of claims 12 to 16 including means to monitor the intensity of vibrations in the shell structure and to vary the generation of vibration in response to the monitored intensity.

19. Apparatus according to any of claims 12 to 16 including error sensors in the enclosed medium monitoring the vibrations therein, means for monitoring the intensity of vibrations in the shell structure, means for generating vibration in the enclosed medium and means for adaptively varying the generated vibrations in response to the error sensor signals.

20. Apparatus according to claim 19 including means to adaptively vary the generated vibrations such that the mean value of the squared error sensor signals is minimised.

21. Apparatus according to any of claims 12 to 20 in which the means provided to fix the displacement minima of vibrations induced in the shell, relative to the shell, is a discontinuity in the structure of the shell.

22. Apparatus according to any of claims 12 to 21 in which the shell is formed from inner and outer walls, the enclosure being defined as the cavity defined therebetween.

23. Apparatus for reducing sound transmission to a body contained within an enclosure, the apparatus being substantially as hereinbefore described with reference to the drawings.