CA1113515A - Fixed position, fixed frequency pendular type vibration absorber with frequency linearization - Google Patents

Abstract

FIXED POSITION, FIXED
FREQUENCY PENDULAR-TYPE VIBRATION
ABSORBER WITH FREQUENCY LINARIZATION

ABSTRACT OF THE DISCLOSURE
A fixed frequency vibration absorber adapted to be fixedly mounted in a fixed vibration prone system. The vibration absorber is of the pendular-type with two dynamic masses suspended in pendular fashion from a base member, and with at least one coil spring acting upon the masses to establish and linearize the vibration absorber natural frequency.

Description

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BACKGROUND OF THE INVENTION
Field of Invention - This invention relates to vibration absorbers and more particularly to fixed vibration absorbers which utilize pendular construction and in which the natural frequency of the vibration absorber remains constant and thereby efficient through-out full pendular excursions of + 45 of the dynamic mass members by the linearizing effect of the spring loading the dynamic mass members.
Description of the Prior Art - In the fixed vibration absorber prior art the absorbers are basically fixed frequency absorbers which are capable of absorbing vibration over a relativ~ly small range of frequency of the principal excitation source. Typical o these ~ ~:
absorbers are the swastika-type absorber shown in U.S.
Patent ~o, 3,005,520 to Mard and the battery absorber ~ :
presently used in helicopters, which is basically a -spring mounted weight and generally of the type dis-. closed in Canadian Patent Application Ser, ~o. 329,227, entitled Tuned Spring-Mass vibration Absorber by John Marshall II and filed on June 6, 1979.

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These prior art absorbers are fixed frequency absorbers which are capable o~ absorbing vibrations over a rela-tively small range of rotor RPM. In addition, they are generally.heavy, create substantial friction, and have bearings which are susceptible to wear.
Bifilar-type vibration absorbers have conventionally been used solely on rotating mechanisms, such as crank-shafts of automobiles and aircraft engines and on helicopter rotors as shown in Paul and Mard U. S. Patent No. 3,540,809. In these installations, the centrifugal force generated by rotation of the mechanism involved ~ is necessary for the operation of the bifilar-type -- vibration absorber. In a fixed position vibration absorber of the type sought in this application, centri-fugal force is not present. In this improved absorber, -the force is generated by a spring connected within the absorber either between the masses or between one mass and the base.
Another prior art absorber is shown in Desjardins et al U. S. Patent No. 3,536,165 but it should be noted that this is not a bifilar vibration absorber, that it is a high frictîon and hence a high damping absorber and therefore a low amplification absorber so that it does not have the advantages of our bifilar vibration absorber.

SU~MARY OF TXE INVENTION
A primary object of the present invention is to provide a vibration absorber of pendular construction and which operates at a fixed natural frequency through-out + 45 of absorber pendular motion.

-3-pendular vibration absorber Erequency is established by a biasing spring force of selected spring rate acting against the pendular selected mass member or members to exert an internal force thereon. The spring rate of the spring is selected to compensate for spring rate reduction normally caused by pendular excursions of the mass member so that the spring rate and hence the internal force are substantially constant throughout pendular excursions of at least + 45.
It is a further object of this invention to teach such a vibration absorber which is low in weight, small in envelope which utilizes the pendular principle to take advantage of low inherent damping, low friction, 1GW
maintenance, and high reliability characteristics, and which utilizes a spring of selected spring rate to compensate for the non-linear pendulum effect of the bifilar at high amplitudes.
It is a further object of this invention to teach a vibration absorber which minimizes friction, and hence is a minimal damping absorber. This minimal friction and low damping characteristic of our absorber results -in higher absorber amplification, that is a higher quotient of mass motion divided by aircraft motion, so that higher mass reaction loads can be realized to con-trol fuselage vibrations. Thus, lower damping in our vibration absorber results in lower weight required to achieve the desired vibration suppression.
; It is still a further object of this invention to provide an improved vibration absorber utilizing pendu-lar, preferably bifilar or trifilar principles, to obtain their low inherent damping, light weight and small .

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envelope advantages and to utiliz~ a spring to compensate for the non-linear pendulum effect when the damper is used at hi~h angular amplitudes, which would otherwise change the frequency of the system to make the system ineffective.
By utilizing the pendular principle and coil spring arrange-ment, the construction taught herein produces a vibration absorber having low inherent damping, thereby permitting the use of lower dynamic masses in the pendular absorber, thereby not only reducing the weight of the vibration absorber but also of the principal system, such as the helicopter.
It is an important teaching of our invention to utilize a compressed coil sprin~ member acting on the ;~ movable mass means in our pendular-type vibration ; absorber so that a particular selected compression in the spring height establishes the tuning frequency of the absorber for a given rctor RPM, or other principal excitation source. The spring rate is selected to linearize the system so that the absorber natural frequency is substantially invariant. This invariant feature is important in maintaining high amplification and high dynamic mass motions in the vibration absorber so as to permit the reduction of absorber weight.
It is a further object of this invention to teach such a vibration absorber in which the spring members impose maximum internal loads on the mass members when the mass members are at their end travel, maximum angular positions in their arcuate, pendular excursions, since the mass members impose maximum compression force and displacement of the spring members at these maximum angular positions to thereby effect linearization of the _5_ 3~

vibration absorber so that its natural frequency is non-variant throughout its full range of pendular motion up to + 45~.
In accordance with a further embodiment of the invention, a fixed frequency vibration absorber adapted to be fixedly attached to a vibration-prone system to cooperate with the principal vibration excitation source which primarily generates vibrations in a given direction so as to control system vibrations comprises base means, two mass means of selected equal mass, pendular connecting means connectins said mass means in opposed positions to said base means for support and pendular motion therefrom, and spring means operatively connected between said mass means in preloaded condition to exert a fixed force on said mass means to thereby establish the fixed natural frequency thereof and of the vibration absorber, and to also cause said mass means to move .
in pendular motion such that motion of said mass means produces additive forces in said given direction to absorb the vibration force established by said principal source, and : 20 so that all other forces so produced mutually cancel.
: In accordance with a further embodiment of the invention, a fixed frequency vibration absorber adapted to be fixedly attached to a vibration prone system to cooperate with the principal vibration excitation source which primarily : .
generates vibrations in a given direction so as to control system vibrations comprises: base means, two mass means of selected equal mass, pendular connecting means connecting :
said mass means in opposed positions to said base means for support from and pendular motion in said given direction, controllable spring means operatively connected between said :~
: mass means in selectively preloaded condition to exert a -, ~ ' .

force on said mass means to establish the desired natural frequency thereof, and to also cause said mass means to move in pendular motion such that motion of said mass means produces additive forces in said given direction to absorb the vibration force established by said principal source, and so that all other forces so produced mutually cancel, said spring means also being-of selected spring rate to compensate for the spring rate reduction caused by pendular motion of the pendular connecting means and thereby provide an essen-tially linear spring rate acting on said mass means for anglesof pendular motion of at least +45, so that the natural frequency of the vibration absorber is substantially constant throughout this range of operation.
In accordance with a still further embodiment of the invention, a fixed frequency vibration absorber adapted to be fixedly attached to a vibration prone system to co-operate with the principal vibration excitation source which primarily generates vibrations in a given direction so as to control system vibrations comprises: base means, two mass means of selected equal mass, pendular connecting means connecting said mass means in opposed positions to said base means for support from and pendular motion in said given direction, at least one spring member operatively connected between said mass means in selected preloaded condition to impose an internal force on said mass means to establish ; :
the natural frequency of the vibration absorber and to also cause said mass means to move in pendular motion such that motion of said mass means produces additive forces in said given direction to absorb the vibration force established by said principal source, and so that all other forces so produced mutually cancel, said spring member also being of - 6a -~ " ~

selected spring rate to compensate for the spring rate reduction caused by the pendular motion of the pendular con-necting means and thereby provide an essentially linear spring rate reacted on the mass means for angles of pendular motion of at least +45, so that the natural frequency ~f the vibration absorber is substantially constant throughout this range of operation.
In accordance with a still further embodiment of the invention, a fixed frequency vibration absorber adapted to be fixedly attached to a vibration prone system to co- :
operate with the principal vibration excitation source which primarily generates vibrations in a given direction so as to control systems vibrations comprises: base means, two mass means of selected equal mass, pendular connecting means connecting said mass means in opposed positions to said base means for support from and pendular motion in said given direction and comprising: at least one set of overlapping apertures of selected diameter in said base means and said mass means and having parallel axes perpendicular to said given direction, and a pin member of selected diameter and having an axis parallel to the aperture axes and extending through the overlapping apertures of each set thereby joining .the mass means to the base means for pendular motion with respect thereto, at least one spring member operatively con-nected between said base means in selected preloaded condi-tion to impose an internal force on said mass means to establish the natural frequency of the vibration absorber and to also cause said mass means to move in pendular motion such that motion of said mass means produces additive forces :~
in said given direction to absorb the vibration force esta-blished by said principal source, and so that all other .

- 6b -forces so produced mutually cancel, said spring member being of selected spring rate to compensate for the spring rate reduction caused by the pendular motion of the pendular con-necting means and thereby provide an essentially linear spring rate reacted on the mass means for angles of pendular motion of up to at least +45, so that the natural frequency of the vibration absorber is substantially constant through- -out this range of operation, and so that said apertures and pin are of selected diameters to produce a minimal envelope vibration absorber consistent with load carrying requirements and so as to coact with said at least one spring mem~er in establishing desired vibration absorber natural frequency.
: In accordance with a still further embodiment of the invention, a helicopter has: a fuselage, a lift rotor projecting from and supported from said fuselage for rotation and constituting the principal fuselage vibration excitation source which primarily generates vibrations in a vertical direction, a fixed frequency vibration absorber fixedly ;attached to said fuselage at a selected station therein to control fuselage vibration and comprising: base means, two mass means of selected equal mass, pendular connecting means connecting said mass means in opposed positions to said base means for support from and pendular motion in said vertical direction, and spring means operatively connected between said mass means in s01ected preloaded condition to exert a fixed force on said mass means to thereby establish the fixed natural frequency thereof and of the vibration absorber and to also cause said mass means to move in pendular motion such that motion of said mass means produces additive forces in said vertical direction to coact with the vertical vibration force established by said rotor to reduce fuseIage vibration :~ - 6c -, ~ ~, and so that all other forces so produced by mass means motion mutually cancel and are there~ore not imparted to the fuse-lage.
In accordance with a still further embodiment of the invention, a helicopter has: a fuselage, a lift rotor projecting from and supported from said fuselage for rotation and constituting the principal fuselage vibration excitation source which primarily generates vibrations in a vertical direction, a fixed frequency vibration absorber fixedly : 10 attached at a selected station to said fuselage to control fuselage vibration and comprising: base means, two mass means of selected equal mass, pendular connecting means connecting said mass means in opposed ~ositions to said base means for support from and pendular motion in said - vertical direction, controllable spring means operatively :~ connected between said mass means in selected preloaded condition to exert a force on said mass means to establish the desired natural frequency thereof and to also cause said mass means to move in pendular motion such that motion of said mass means produces additive forces in said vertical direction to coact with the vertical vibration force esta-blished by said rotor to reduce fuselage vibration and so that all other forces so produced by mass means motion mutually cancel and are therefore not imparted to the fuse-lage, said spring means being of selected spring rate to compensate for the spring rate reduction caused by pendular motion of the pendular connecting means and thereby provide an essentially linear spring rate acting on said mass means for angles of pendular motion of at least +45, so that the natural frequency of the vibration absorber is substantially constant throughout this range of operation.

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Other objects and advantages of the present inven-tion may be seen by referring to the following description and claims, read in conjunction with the accompanying draw-ngs .
BRIEF DESCRIPTIO~ OF THE D~AWIl~GS
Fig. 1 is a graph showing helicopter fuselagevibration plotted against rotor RPM to show the operation of the prior art fixed vibration absorbers.
Fig. 2 is a schematic representation of one embodi-ment of our vibration absorber.
Fig. 3 is a schematic representation of a portion of the connection between the base and the mass member to produce the desired low fr~ction, low inherent damping, pendular result.
Fig, 4 is a schematic representation of the pre-ferred embodiment of our vibration absorber.
Fig. 5 is a top view, partially broken away and with control mechanlsm illustrated as attached thereto, of the preferred embodiment of our vibration absorber shown in Fig. 4.
Fig. 6 is a side view, partially broken away, of the vibration absorber of Fig. 4.
Fig. 7 is a view, partially broken away, taken along line 7-7 of Fig. 6.
Fig. 8 is a view, partially broken away, taken along line 8-8 of Fig, 5.
Fig. 9 is a cross-sectional showing of an actuator which could be used with our vibration absorber.

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Fig. 10 is a showing of the nat~lral frequency control mechanism utilized with our absorber in the helicopter environment.
Fig. 11 is a graph of the fluid pressure acting on or the internal force gene~ated in the pendular mass means plotted against rotor RPM.
Fig. 12 is a partial showing of the vibration absorber used as a fixed frequency fixed position absorber.
Figs. 13 and 14 are cross-sectional illustrations of spacer means used in combination with the spring or springs in the Fig. 12 embodiment.
Fig. 15 is a graph of the vibration absorber natural frequency ratio plotted against pendular angular motion to illustrate the difference in operation between this -~
linear vibration absorber and a conventional pendular vibration absorber.
Fig. 16 is an illustration of the pendular motion of the dynamic mass members of our vibration absorber to illustrate the amplitude of motion, angular motion amplitude, pendular arm length, and coil spring com-pression motion of the dynamic mass member.
Fig. 17 is a perspective showing of the base member of the vibration damper with the other parts of the vibration absorber removed therefrom.

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nESCRX:PTION OF TEIE P~EFERRE:D EMIIODIMENT
The vibration absorber taught in this application will be described in the environment of a helicopter in which the vibration absorber is fixedly mounted in a helicopter uselage to coact with the principal helicopter ~ibration excitation source~ namely, t~e rotor or ro~ors, to reduce the vibration ~mparted to the ~uselage thereby.
To appreciate the operation and advantages o~ this variable frequency vibxation absorber, ~he short-comings of the ixed frequency, ~ixedly positioned prior art vi~ratio~ absorbers will be di5cussed. Referring to ~ig. 1 we see a graph ~ o helicopter ~uselage ~ibrations p~otted against helicopter r~tor RPM~ It is well known that the vibration generated by a rotor and the response o~ the helicopter structure thereto varies and ls a func~
of rotor RPM, as shown typically by graph A~ Line B
indicates the vibration line below which accep~able fuselage : vibration occurs The prior art fixed frequency, fixed position vibration absorbers would operate ~enerally along graph C, and it will be noted that such a vibr~tion absor~er is effective over range D between lines E and :F. It will : be noted that ran~e D covers a small variatio~ or s~an in rotor RP~ over which the prior art ~ibration absorbers are efective. Range D i9 determined by the mass ratio o the absorber, that is the ratio of ~he weight of the absorber to the e:fective weight of the substructure in which the absor~er is fixedly mounted, and in part by the inherent . - ~ -_7~--damping o the vibration absorber and the substructure.With t~e prior art absorb~rs~ to obtain a reLatively wide range C o absorbing~ a very heavy vibration absorber would be required. Dîmension G, the minimum achievable vibration ; level, would ~è determined by the amount of inherent damping in the absorber, and in part by the inherent substructure dampin~, and in part by the afore~entioned mass ratio. If, t~eoretically, a vibration absorber could be utilized which has zero inherent damping, maximum vibration absorptio~
would occur so as to achieve minimum fuselage ~ibration, l.e., d~mension G would be reduced. Such a system cannot - be realized in practice.
The obj ec t;ve of this vibration absor~er is to be able to get maximum vib~ation absorption, îndicated by any point along line H over a greater range J o~ rotor RP~, li~es K
and L representing minimum and maximum necessary operating-~r excitation ~PMs of the helicopter or o~r p~ sb~n~.
To u~derstand the purpose and operation of this vi~ratio~ absor~er, it is first necessary to understana the difference between a vibration absorber a~d a vib~ation damper. A vibration damper serves to dissipate the energy o the vib~ation5 imparted to the fuselage by the roeor.
~ibration dampers ca~ use friction principles or a~y type of energy damping principle. A vibration absorber, on the other hand, doe~ not dissipate alreaay established v~brativn energy but e~tabllshes a second ~i~ra~ory mode in the system so as to coact w~th the principal ~yseem :~ .
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mode, the substructure mode, to produce a resultant mode which has minimum vibration. Stated another way, a vibration damper damps already created principal system vibrations, while a vibration absorber coacts with the system principal vibration excitation source to change its characteristics to a low vibration system.
A schematic representation of one form of this vibra-tion absorber 10 is shown in Fig. 2. In Fig. 2 masses 11 and 12, of selected mass, are supported from base,members by suspension arm members a, which can be considered ~o be pendulous members as illustrated by the phantom ~ine motion for mass 11. In practice, pendular arm a is ac~ually the pin and bushing connection shown representatively in Fig. 3 in which pin member 14-of diameter d is positioned in hole 16 of one of thP mass members 11 or 12 and overlapping hole 17 in the base member so as to produce an equivalent pendulum motion of pendulum arm a, in which arm a equals the difference between hole diameter D and pin diameter d, i.e.~ a - D - d.
Spring 18 is positioned between masses 11 and 12 and serves to draw them together and thereby pre~oads the selected masses so suspended to establish an in~ernal force therein and thereby establish the natural frequency of masses 11 and ,12, and therefore the natural frequency of absorber 10.
The natural frequency of masses 11 and 12, and hence absorber 10, is determined by the preload of spring 18 and the mass ; of mass members 11 and 12, which are preferably of ' equal mass. Spring 18 performs :, - . . . . .
.. . . ~ -another important functi~n, in particular, it makes linear the non-linear characteristics o~ the pendular construction~
To explain this linear/non-linear concept, reference will be made to Fig. 2. It will be noted by viewing Fig. 2 that as arms a pivo~ to move mass ll from its solid line to its phantom line position, the spring rate of the conventionaL
bifilar system, considaring only the preload from spring 18 and not the spring rate~ is reduced and therefore the natural frequency o~ the bifil~r system is reduced to thereby reduce its effectiveness. This reduction in natural frequency of the mas~ member with amplitude causes the system to be non-linear, and limits its range o effective-ness. ThiS non-linear vibration characteristic of a pend~lar system occurs immediately upon any angular motion although a practical angle o~ excess would typically be 10. We cou~d prevent the system from swinging beyond lO by increasing the length of the pendulum arms a but this would be undesirable because this would produce a heavier system requi~ing a larger space e~velope.
With spring 18 present, however, as mas~ ll swing~
from Lt solid line to its phantom li~e positno~, the chang~ng force of spring 18 acting on mass ll is increase~
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thereby tending to keep the system linear by keeping the equivalent absorber spring rate and natural frequency of .~ , the bifilar system shown in Fig. 2 at îts original value.
In this vibration absorber~ we maintain the low weight and small space envelope advantage of a shor~ pendu~um g _ , arm a, yet produce a linear system by controlling the natural frequency of the vibration absorber by manipulation of the force generated by spring 18 and imparted to the masses 11 and 12.
The preferred embodiment of vibration absorber 10 is shown schematically in Fig~ 4 in which masses 11 and 12, of selected mass, are supported from central base member or ground 20 by pendular-type connections represented by arms a and have internal orce applied thereto to establish system natural frequency by spring 18, of selected preload and spring rate, which serves to force masses 11 and 1~ to separate.
For a more particular description of the preferred embodiment reference will now be made to Figs. 5-8 in which base member 20, which is fixed to the fuselage as shown in Fig. 8, supports selected mass members 11 and 12 therefrom in pendular fashion Each mass m~mber 11 and 12 is supported from the base member 20 by three pendular -connection~ similar to Fig. 3, thereby forming a tri~ilar connection, and each o~ the three connections including, as best shown in Fig. 8, an aperture 22 in mas~es 11 and 12 and an overlapping aperture 24 in base 20 and each having a pin member 26 extending therethroug~. As best shown in Fig. 5, each mass means 11 and 12 i5 connected to base ~ember 20 at three such pendular connecting stations along the mass length, which stations are designated as Sl, S2 and S3. As best shown in Fig 6, .

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the pendular connection at station S2 is at the bottom of each mass while pendular connections Sl and S3 are at the top of each mass. In view of this three station connection, reminiscent of the three-legged stool, the mass is given geometric stability as supported ~rom base 20 in both the yaw direction shown in Figs. 5 and 8 and the pitch direction shown in Fig. 8. It will therefore be seen that to this point our vibration absorber inclu~e~
two mass members 11 and 12 supported in selectively spaced connecting stations from base member 20. The connections may be of the type more fully disclosed in U. S. Patent No. 3,540,809`to W. F. Paul et al. ~n Fig. 7, one of two spring members 18 is shown extending between masses ll and 12, utilizing spring retainers 28 and 30. Springs 1~ are of selected spring rate so that when ins~aLled and preloaded, the springs provide the necessary internal force to mass members 11 and 12 to establish a selected natural frequency of masses 11 and 12 and therefore o vibration absorber 10.
With spring 18 assembled as shown in Fig. 7 and pre7Oaded, it will be observed that the spring serves to impart a separating force to mass means 11 and 12.
The construction of base member 207 which is preferably of one-piece const~uction, is very important to this invention. As best shown in Fig~ 17, base member 20 comprises flat platform 51 extending longitudinally of the base member as shown in Fig. 17 and constituting a solid base for the base member 20 so that platform 51 may be attached in any conventional ~ashion, such as by nuts and botts, to the fixed vibration prone system which our vibration absorber is in~ended to operate in. Three parallel, laterally extending plate members 53, 55 and 57 extend perpendicularly from platform 51 and e~tend in the lateral direction, which is the direction or plane of desired mass membermdi~n. ~ p~e members 53 and 57 are identical in shape and project a substantially greater height out of plat~orm 51 than does central plate member 55. Plate members 53, 55 and 57 each have equally laterally spaced apertures 59 and 61, 63 and 65, 67 and 69 therein, respectively, Apertures 59-69 are of equal diameter and their axes extend perpendicular to plate members 53, 55 and 57, and therefore perpendicular to the direction of desired dynamic mass motion for the - vibration absorber Apertures 59 and 61, and 67 and 69 are the same height above platform 51, while apertures 63 and 65 are substantially closer thereto. By viewing Fig. 17 it will be observed that apertures 59-69 form two sets of three equal di~meter apertures having parallel axes and with each aperture positioned at the corner of a triangle. The first three aperture set consists of ~ -apertures 61, ~5 and 69, while the second aperture set consists of apertures 59, 63 and 67. These two aperture sets are parallel to one another and, in view of the ~act --that the apertures in each set are positioned at the corner of a triangle, they form the basis, when joined to mass members 11 and 12 as more fully disclosed in Figs. 5 and 6, for three point pendular of bifilar type connection between the mass members and the base member, which three points of pendular connection are ofset in two perpendicular direc-tions, which are coplanar. To be more specific, for example, aperture set 59, 63 and 67 includes three longitudinally offset apertures 59, 63 and 67, and also includes aperture 63 which is vertically offset from equal height apertures 59 and 67. This three point triangular-type connection be~een the mass members and the base member provide geometric stability so as to prevent both ~ -roll and yaw tumbling of the mass members with respect to the base member.
~ ith respect to the construction of pla~e me~bers 53, 55 and 57 and in particular their c~nstruction in the areas where the apertures pass therethrough, it is important to note that these plate members provide substan- -tial structural support to the mass members which will be supported therefrom in that, as best shown in Fig. 17 and illustrated with respect to plate member 57, apertures 67 and 69 have two parallel beam portions 71 and 73 extending laterally across the pLate member ab~ve ana below ~he apertures and structural web section 75 ex~ending between beam members 71 and 73 at a station betwee~ apertures 67 and 69 SG as to form an I-shaped structure, ~Drmed by beam members 71 ~nd 73 and support web 75, at the load carrying station of plate member 57 in which dynamic mass member supporting apertures 67 and 69 are ~ocated. In fact~ this I-shaped structure is strengthened by the fact that its ends are closed at portions 77 and 79 to form a closed : ..

box construction consisting of sections 71, 77, 73 and 79, with structural web section 75 extending through the center thereof. Mass member loads reacted by plate member 57 at apertures 67 and 69 are imparted to plate member 57 at this high strength structural section and therefrom into platform member 51 for transmittal to the fixed vibration prone system, such as the fuselage of the helicopter. m e load carrying demands on plate member 57 might be such that the plate may include lightening and maintenance access holes 81 and 83. It will be noted that while plate member 57 has been used to describe the structure o~ the plate members in the vicinity of the apertures, plate members 53 and 55 are similarly constructed.
As best shown in Figs. S and 6, the dynamic mass members 11 and 12 extend longitudinally along opposite lateral sides of base member 20 and each is preferably of one-piece construc-tion and fabricated to include plate members 21, 23~ Z5~ 27, 29 a~d 31 which extend parallel to plate members 53, 55 and 57 of base member 20 and extend in the direction of mass member motion or in the plane of mass member motion. The mass member plate members constitute three sets, with the first set 21 and 23 being positioned on opposite sides of and selectively spaced longitudinally with respect to base member plate member 53~ the second set 25 and 27 being positioned on opposite side of base plate member 55 and selectively spaced longitudinally with respect thereto~
and third set ~9 and 31 positioned on opposites of base plate member 57 with selected longitudinal spacing there-between.
As best shown in Figs. 5 and 7~each parallel compression coil spring 18 is received at its opposite ends in spring end retainers 28 and 30, which retainers are supported in mass members 11 and 12 as shown. In addition, the opposite ends of coil spring 18 are ground to properly fit into retainers 28 and 30 and thereby aid the spring static stability so that it needs no support between its ends.
Each mass member plate member has an aperture therein of equal diameter with the apertures in all other mass member plate members and of equal diameter with the apertures in the plate members of the base member 20.
Each plate member aperture is concentric about axes which are not shown but which are perpendicular to the plate member and parallel to each other. As best shown in Figs. 5 and 6, these mass members apertures include apertures 33, 35, 37, 39, 41 and 43 in plate members 21-31, respectively. As will be seen in Figs. 5 and 6, the apertures in the plate members of the base member overlap with the apertures in the plate members of the mass members and each has a cylindrical, flanged bushing inserted therein as shown, which bushing is ~abricated of an anti-friction material, such as hardened stainless steel.
A solid, substantially cylindrical pin extends through each set of aligned apertures as shown in Figs. S and 6. ;~

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These pin members which are visable are designated as 71, 73 and 75 but it sh~uLd be noted that each mass member 11 and 12 is connected to and supported from base member 20 at three pendular or trifilar type connecting stations Sl, S2 and S3, which stations are defined by the overlapping apertures of the base member and the mass members and the pin members. The pin members 71-75 are fabricated of an anti-friction material such as a carbonized steel. As can be best seen in Fig. 6, these ~ ndular connecting station~ -Sl, S2 and S3 are longitudinally offset from each other to provide geometric stability between the mass members and the base m~mbers to prevent roll moments therebetween, and are also vertically offset to-provide the necessary geometric stability to prevent yaw moments between the mass members and the ba9e member~ Due to this three position pendular, trifilar-type connection between each mass member 11 and 12 and the base member 20, each mass member moves in pendular, arcuate translational motion with respect to the base member so as to be parallel thereto at all times. To minimize ~riction and hence damping of the system~ each pin member includes a tapared circumferential ~lange illu9trated in Figs. 5 and 6 in c~nnection with pin 73 only and indicated at 83 and 85, however all pin members have such tapere~ ~langes Flanges 83 and 85 are positioned in the longitudinal spacing 87 and 89 between the bushing apertures through which pin member 73 extends and are tapered in a radially outward --. .

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direction so as to be of minimal thickness at their outer periphery and hence serve to produce minimum friction contact between the reLatively movable mass members and base member during the ~ull mode o pendular operation therebetween.
It will therefore be seen that this vibration absorber produces minimal friction, soLely the minimal flexing friction of the coil spring members 18 and the rolling friction of roller members 71-75. This vibration absorber is therefore low in d~mping, high in amplification, with lower weight supported masses 11 and 12, thereby . reducing the weight of the absorber and the overall aircraft.
With respect to spring members 18, it is important that the spring deflection, free length and mean di~meter be selected so that the coil spring i5 staticalty stabLe when its ground ends are positioned between spring retainers 28 and 30. The importance of this spring static stabilLty is tha t it does not require add;tional spring support mechanisms, such as a center spring guide, since such would add weight, :Eriction and damping to the sy~tem to there~y reduce the effectiveness of the vibration absorber.
It should be noted ~hae maximum spring defIection is achieved when first, the absorber is tuned to its highest operating frequency and second, the absorber is operating at its maximum pendular amplitude so as to avoid excessive transverse spring deflections, since any touching of parts could cause fretting or friction, bot~ of which are detrimental to absorber life or performance. In addition, both the transverse and axial natural frequencies of the spring are selected to be detuned from the system excitation ~requencies so as to avoid excessive spring motions, since any touching of parts caused thereby could produce fretting or friction, both of which are detrimental to absorber life or performance.
Actuator 32, shown in Fig. 7, is positioned in series with spring 18 between masses 11 and 12. Actuator 32 may be actuated initially to impose a force to selectively preload spring 18 and establish the initial natural frequency of vibration absorber 10. Actuator 32 may - - thereafter be actuated to either increase or~decrease the natural frequency of vibration absorber 10. When-actuator - 32 is controlled as a function of helicopter rotor RPM, the actuator is then varying the deflection of spri~g 18 to thereby vary the internal forces on mass means 11 and 12, and hence to vary the natural frequency of absorber 10 as a function of rotor RPM fr~m its initial natural requency caused by initial preloading or from its last actuator established natural frequency. In this fashion, the natural frequency of vibration absorber 10 is controlled as a function of rotor RPM to coact with vibration excitation forces impo~ed on the fuselage by the rotor to thereby reduce fuselage vibration.
The construction of actuator 32 may best bP understood - by viewing Fig. 9. The actuator consists of telescoping sleeve members 34 and 36, the former being translatable with respect to the latter? and the latter being fixedly connected to the mass means 12 by conventional connecting means 38~ Selectively pressurized fluid from a control system to be described hereinafter enters adapter 40 and flows therethrough and through passage 42 into hydraulic chamber 44 where it exerts a force causing sleeve member 34 to move leftwardly with respect to fixed member 36 to thereby compress spring 18 as it so moves. This compression of sprîng 18 adds to the internal force applied to mass . means 11 against which it directly bears through retainer ; 28. Similarly, due to the fluid pressure so e~erted on fixed sleeve 36, which is attached to mass means 12, actuator 32 similarly creates greater internal force in :
mass means 12. Actuator 32 also includes a position transducer 45 which is o conventional design and operates in typical rheostat fashion to send a position feedback signal, representative of the position of mova~le member 34 as determined by the pressure in chamber 44, to the actuator control system 47. There are other prior art actuators which could be used in thi5 vibration absorber, for example, the positioning actuator sold under part -number A-24553~ by Moog, Inc., Aerospace Division of Proner Airport, East Aurora, New York 14052~ Another prior art actuator is an electric screw-type actuator with feedback of the type manufactured by Motion Controls Divisîon of Simmonds Precision, Cedar Knolls, New Jersey .

Attention is now directed to Fig~ 10 or an explanation of the control system 47 used to vary the natural frequency of absorber 10 as a function of rotor RPM. As sho~n in Fig. 10, helicopter rotor 42, possibly t~rDugh a tachometer, imparts a rotor speed (RPM) signal to controller 44. The controller 44 operates to provide a signal on a line 74 to the absorber 10 that is proper to control the valve~ in the absorber 10 to provide displacement as the square of rotor speed within an operating range`of rotor speeds as is described with respect to Fig. 11 hereinafter. Assuming the rotor 42 provides a tachometer signal on a line 76 -~
which varies in frequency as a function of rotor speed, conversion to a DC voltage proportional t~ ro~ r speed may be made by any conventional frequency-to-voltage ~onverter 78, which may, for instance, comprise a simple integrator, ..
or a more complex converter employing a Teledyne Philbrick 470~ frequency-to-voltage conversion circuit, or~ the like.
In any event, a DC signal on a line 80 as a function of rotary speed of the rotor 42 is provided to bot~ inputs of an analog multiplier circuit 82, of any weIl known -~
t~pe, so as to provide a signal on a line 84 which is a .
function of t~e square of rotor speed. A potentiomster .
: 86 is provided to allow a gain adjustment~ whereby the overall effect of the control can be adjusted to suit each particular aircraft. This provides a suitable signal on ,:
a line 88, wh~ch is some constant times the square of , . . , -i' P ~
rotor speed, to a summing ampliier 90, the other input of which is a feedback error signal on a line 92 which combines the actual position of the actuator 32 in response to the position sensing potentiometer 45 (Fig. 9), on a line 94, and a bias reference provided by a source 96 on a line 98. Thus the output of the summing amplifier 90 provides a signal on the line 74 to direct the actuator to a position determined as some constant times the square of rotor speed, which position is maintained in closed loop ~ashion by the feedback signal on a line 94,-as modified by the bias provided by the source 96. The bias resulting from the source 96 will cause t~e pressure ~' :
signal on a line 74 to bring the actuator 32 to a selected initial position, thereby compressing spring 18 as shown in Fig. 7 to an initial position which will produce.the desired initial natural frequency in mass means 1~ and 12 and therefore absorber 10. This actuator preloading is done so that actuator 32 can reciprocate either leftward-ly or rightwardly and thereby vary the internal force being 20 imposed upon mass means 11 and 12 in response to bo~
rotor RPM increases and rotor RPM decreases. It will be realiæed that if actuator,32 were installed in its end travel position, it could respond to rotor RPM changes in one direction only.
The c~ntroller 44 is thus progræmmed to send a hydra~lic pressure signal proportional to rotor RPM
to absorber 10 and absorber 10 provides an actuator ~ . . .

position feedback signal to the controller 44. It will be noted that this absorber is fixedly mounted from the fuselage.
Attention is now directed to Fig. 5 for a further explanation of this control system 47. The pressure signal from controller 44 goes to hydraulic valve 46, which receives aircraft supply pressure through line 48 and has hydrad ic return line 50. Selectively pressurized hydraulic fluid passes throug~ fle~ible pressure line 52 into common pressure line 54 from which it enters the two actuators 32a and 32b to selectively change the foroe being exerted by springs 18 on mass means 11 and 12 and hence the natural frequency thereo~ and of the vibration absorber 10. Similarly, position feedback signals from each actuator 32a and 32b are brought through position feedback line 56 to controller 44.
: This control system 47 is an open loop position feedback system because it is preprogrammed, that is, , . - . .
it has been calibrated in the laboratory to return to ~0 a given position. It will be evident to those sk~lled in the art that this control system also has t~e capability of acting as a closed loop position feedback system.
The operation of our absorber is illustrated in the . ~- .
graph ~hown in Fig. 11 in which the pressure in the flexible pressure line sa or the internàl force ~mpar~ed to:the mass means 11 and 12 by spring 18 is plotted against rotor RPM (NR). ~iasing put into the system .

,, ' ', , causes the pressure to be flat in the low RPM range, which is below the operating range, and then follows the curved graph portion representative of the formula f7 ~NR
where~ is the pressure and NR is the rotor speed (RPM).
It will therefore be seen that over ~he region designated as "Operating Range" the force acting upon the vibrations absorber 10 to vary its natural frequency varies as a function of rotor speed3 in particular, rotor speed squared.
This "Operating Range" is approximately 90 percent - 120 percent.
Positive stop 99, which may be made of rubber, are attached to mass means 11 and 12 as best shown in Fig. 6 ; and serve to limit the useful motion of the mass members relative to the base member, to prevent metal-to-metal contact between the mass members and the associated vibration absorber parts.
It will therefore be seen that our variable frequency vibration absorber i9 an improved vibration absorber utilizing bifilar principles to take advantage of ~he lightweight, small dimensional envelope, the low inherent damping thereof, the high reliability thereof, the low friction generated thereby, and the minimum maintenance required therefore~ This vibration absorber a7so u~ili2es a spring to compensate for the non-linear pendulum effect of the pendular-type vibration absorber at high amplitude~, ; thereby making the absorber linear. It will ~urther be realized that this vibration absorber changes its natural -~3-, frequency as a unction of rotor RPM so thQt the absorber will always be operating at its maximum level of effective-ness to reduce fuselage vibration due to rotor excitation.
The absorber spring 18 is a selected spring rate which is controlled to initially preload the selected bifilar mass m~mbers to establish the initial natural frequency of the ...
mass members and the absorber. The vibration absorber is thereafter controlled to vary the amount of loading by the spring on the absorber mass members as a fun tion of .rotor RPM to permit effective vi~ration absorption over a large span of rotor operating frequencies.
While this vibration absorber has been described in the helicopter environment to control the vibrations -~;.
generated by the helicopter rotor and imparted thereby to the helicopter fuselage, it will be evident to those s~illed in the art that it can be utilized in any fixed vibration prone system as a fixed vibration absorber operative to coact with the system principal vibration excitation source, as a unction of the vibrations generated by the principal source, to reduce system vibration.
Further, while the preferred embodiment of the inven-tion is directed to a fixed vibrat;on absorber of the pendular-type with provisions for absorber natural frequency variation, it should be noted that the fixed bifilar vibration absorbers provides substantial advantages over prior art fixed vibration absorbers, even when used without the natural frequency variation capabi~ity, because the vibration absorber so used as a fixed natural frequency absorber will still have the inherent advantages of a bifilar-type system, namely its low inherent damping, lightweight, minimum space envelope, high reliability and minimum maintenance.
- Viewing Fig 12, we see vibration absorber 10 as a fixed frequency vibration absorber. When used as a fixed frequency vibration absorber as shown in Fig. 12, the absorber construction w;ll be precisely as shown in Figs.
5-8 in the preferred embodiment except that actuator 32 will be removed and preferably replaced by a spring retainer 60, which is preferably identical with retainer 28 but positioned at the opposite end of spring 18 there-from and acting against mass member 12. Nith the removal of actuator 32, the actuator control mechanism 47 shown in Figs. 5 and 10 is also eliminated. By viewing Fig. 12 it will be noted that the fixed frequency vibration absorber 10 includes mass members 11 and 12 supported by the same i` pendular-type connections shown in Figs. 5-8 from base member 20 and wi~h spring or springs 18 applying a force thereto tending to separate the mass means 10 and 12.
The natural frequency of the Fig. 12 fixed ~requency vibration absorber is determined by the mass o~ mass means ll and 12 and the spring preload and spring rate of spr~ng or springs 18. As in the Figs. 5-8 variable frequency absorber, the Fig. 12 fixed frequency absorber is also linear in the same fashion.

, s It may be desired to modify the Fig. 12 fixed frequency modification as shown in Figs. 13 or 14 to permit a degree of adjusbment in establishing the preload force exerted by spring 18 and hence the natural frequency o~ absorber 10 prior to or after its installation either as a subassembly or after installation in the s~bstructure requiring vibration - suppression but not during operation. Viewing Fig. 13 we see a cross-sectional showing of spacer me~ber 62 comprising inner and outer continuous and threaded ring members 64 and 66 in threaded engagement with one another so that the ring m~mbers 64 and 66 may be rotated manually relative to one another throu~h the space shown in Fig. 12 between mass members 11 and 12 thereby varyîng the width or spacing dimension of variable spacer 62 to vary the ~orce exerted by spring 18 on members 11 and 12. Spring 18 may be of one or two-piece construction. Viewing Fig. 14 we see spacer ring 68, shown in partial cross-section, between one or two-piece spring 18 to serve as a spacer ring therebetween to vary the force exerted by ~he combination , of spacer 68 and spring or springs 18 on masses 11 and 12.
Spacer 68 i8 ~referably of two or more piece, segmented construction so as to be manually position~ble through the area shown in Fig. 12 between members 11 and 12 and joined by conventional connecting means to form a continuous spacer ring 68 as illustrated. Of course, for fixed frequency operation actuator 32 could be used but ad~usted to a fixed position to preload springs 18 to establish a .

~ . .

fixed natural frequency for absorber 10. To provide a better understanding o~ the operation of the vibration absorber, the design steps and considerations taken into account in optimizing the design will now be discussed.
We first determined the useful motion which would be required and ~hich is available in our pendular-type vibration absorber by estimating the impedance of the sturcture to be suppressed~ such as helicopter fuselage, and considering both the location of the vibration absorber in the helicopter and the locations in the fuselage where vibrations are to be controlled, such as the cockpit or various cabin locations, one can determine the absorber dynamic mass required to reduce the fuselage to the desired vibration level. Knowing this and the frequency of operation o the vibration absorber dynamic masses, which, for example, happens to be four (4) times rotor RPM for a four bladed rotor, the required absorber dynamic masses displacement ~perating travel, which is the absorber useful amplitude~
can be established.
Having determined this useful motio~ or useful amplitude o~ our pendu~ar-type vibration absorber, one can then : determine the pendular length necessary to achieve max~mum .

; mass member desired angular displacement which we chose to be + 45. This was done by utilizing the equation:
1 X . .

Where: ~ = angular displacement of the ma5s member relative to the base memberO
X = useful amplitude or motion, and a = the pendular lengths and is equal to D - d, where D is the bushing diameter of the base member and mass members apertures, and d is the pin diameter.
The signific~nce of what has been done to this point can best be realized by viewing Fig. 16 which show the pendular arc through which each part of each mass member moves relative to the base member. In Fig. 16, the mass c.g. is illustrated as having an angular displacement of through - ~on opposite sides of its illustrated neutral position, and with pendular length being "a", where a = D - d~
This arcuate, translational pendular motion illustrated in Fig. 16 shows mass member amplitude, which is ~ X and - X, i.e. 2 X total amplitude, and also shows mass member motion "Y", which detexmines the amount of compression of the spring members. It is important to note that the springs are defected or cycled twice for each full cycle o~ "X"
motion of the mass members. In this connection~ it wilt be noted that when the mass member starts its downward motion from its ~ X position, which is also its ful~ angular motion ~ ~ position, the spring is maximally compressed the ; full distance Y and that the spring is al~o maximally compres-sed the full distance Y when the mass concludes its downward motion at position - X, which is also its full angular mot~on - ~po~ition. This is the characteristic of our pendular-type v~bration absorber which produces the internal force being imposed by the spring members on the mass members as the mass moves through its arcuate motion, and ~ence the ;. ,, non-variant natural frequency of the vibration absorber.
While it is an in~erent disadvantage in a pendular construc-tion that it becomes more non-linear as the angular dis-placement of the absorber mass members increase, this îs overcome in our construction in that the spring is compres-sed its greatest at the points of maximum angular dis-- placement to thereby maximize the internal force exerted by the springs on the mass members at that point, and thereby retain a first order linearity so that the natural frequency of the absor~er is non-variant with angular displacement. Maintaining linearity is important to maintaining high absorber amplification so ~hat small dynamic masses can be operated at large useful amplitudes to obtain the necessary inertial reaction forces to suppress aircraft vibration.
A vibration system, such as this vibration absorber, can be described in terms of its effective mass and its effective spring rate (~). Since the effective mass has ~ -already been established, we determined the effective spring rate, or progra~med rate in the case of a variable tuned absorber, necessary to achieve the desired absorber natural frequency or frequencies. This proceaure is fully outlined in Den Hartog's work on "Mechanical Vibrations'!.
; The internal steady load requirement for the absorber can be arrived at by the formula:
Fnr = (Kx) (a) Where Fnr = the internal steady load between the absorber masses for the various rotor speeds.
Kx = the effective spring rate, and a is the length of the pendular arm, i.e., D - d.
Now this steady load, or loads, Fnr is achieved by placing a spring between the mass member spring retainers -having compressed the spring into position so that its internal loads will satisfy the requirement to establish the systems natural frequency (ies) in proper relation to the aircraft's impedance and excitation frequencies. This lo force Fnr is comparable to the centrifugal force for an absorber installed in a rotating system.
- The derived equations of motion will show that there is a preferred spring rate to maintain the absorber's linearity, and that this spring rate is dependent upon the internal load, Fnr, the pendular length, a, and the angular displacement~. The following equation expresses this relationship: -~ _ .
s nr / ~ 1 ~ t~ ~ ) ( 2 sin ~ ) ¦
.~ . .
,., _ ( 1 - cos ~ ) Where: K9 = the preferred spring rate of the physical spring.
~y considering normal operating conditions, typical values of Fnr and ~ can be chosen to select the desired spring rate Ks. This linearization is comparable to incorporating the -cyclodial bushing taught in Canadian Patent Application -Ser. No. 331,688 by John Madden, filed on July 12, 1979 and entitled '`Constant Frequency Bifilar Vibration Absorber'`, .

Using conventional methods, the steady and vibratory loads o~ the spring can be determined from previous data selected or es~ablished.
~ hen, using spring stress allowables, both steady and vibratory, the various spring designs available can be calculated using conventional approaches. Of the springs so selected, each must be chec~ed with respect to stat;c stability of the pnysical spring when placed ~etween the spring retainers of the mass means under the load conditi~ns imposed. ~gain, conventional approaches can be used ~o establish the permissibl~ relationships ~or the compression coil spring which was chosen, for example, between the spring ~-free length~ compressed length, and mean diameter of the particular type of spring end constraints chosen. It is important to achieve the spring design with static stability ; without the ~eed of guides, since such guides are likely - to result in points of contact and introduce sliding friction which will increase the absorber's damping ~nd reduce its per~ormance. This basic spring technology is well known and fully explained in h. M. Wahlts book entitled "Mechanical Springs".
The transverse and axial installed spring natural frequencies or the springs under consideration must be checked out to determine that neither ~s close to the excitation frequencie5 of other absorber elements so a~
to avoid resonance therebetween, which could bring about metal-to-metal contact and cause fretting or introduce friction damping. The final relative mo~ions determlned for the selected spring then determine the clearances between vibration absorber components, for example, the radial clearance between the springs and the dynamic masse~
Since the connecting pins of the pendular-type connection will on~y contact the aperture bushings when the pins are subjected to compressive loading, it is necessary to determine ^ - all of their instantaneous applied loads from the spring and the inertia loads of all the moving masses, and then it is necessary to place the pins and bushings in such locations that their reaction forces maintain compressive loads on the pins at all times. This occurs when the c~mbined applied force of the spring and the mass members inertia loads have a resu7tant vector with a line of action which at all times extends between two sets of overlapping apertures and pins, to thereby assure both pitch and yaw stability, particularly pitch, see Fig. 8~ of the mass members relative to the base member. It will further be - seen that spreading the sets of pins/aperture ~ushing~
results in positive stability. Also, locating the dynamic mass c.g. close to the pinslaperture bushings results in positive stability by minimizing vertical pitch coupling.
Pin inertia must be kept sufficiently low so that the pins do not skid under rotational accelerated loading which is characteristic of vibratory motion~ Positive reaction capability is determined by determining the pi~
instantaneous loading and it8 coefficient of friction with respect to the bushin~. This absorber was determined to have no problems in this regard and therefor one-piece, , solid pin members were used.
Knowing the maximum pin/aperture bushing applied loads, from above, the pin and bushing diameters, and using applicable stress allowables and modulus of selected materials, the widths of the pins and aperture bushings can be established by conventional means.
It will therefore be seen, as described and shown in greater particularity supra, that the fixed fre~uency vibration absorber taught herein is adapted to be fixedly attached to a vibration-prone system to cooperate with the principal vibration excitation source which primarily generates vibrations in a given direction, such as the vertical direction for a helicopter rotor, so as to control system vibrations. This fixed frequency vibration absorber comprises a base member having two mass members of selected equal mass supported from the base member in opposed positions preferably on opposite sides thereof, through pendular connecting means which support the mass means for allochiral pendular motion in the direction of the primary 20 source vibrations. As used herein, allochiral means mirror-image. Spring members extend between the mass means in pre-loaded condition to perform the dual function of exerting a fixed force on the mass means to thereby establish the fixed frequency thereof, and of the vibration absorber, and also to cause the mass means to move in coincident, allochiral pendular motion so that the motion of the mass members pro-duces additive forces in the direction of the principal source vibrations to absorb or coact with the vibration force esta-blished by the principal source so that minimal vibration is imparted from the principal source to the area where the vibration absorber is mounted, such as a helicopter fuselage, , ,~: .' and so that all other forces ~roduced by the mass means pendular motion are mutually cancelled.
This will be best understood by viewing Fig. 16 which shows the centers of gravity of mass members 11 and 12 mounted on opposite sides of frame 20 through pendular, bifilar connections thereto, so that due to the force being exerted against masses 11 and 12 by preloaded spring 18 as shown supra, the mass members ll and 12 are caused to move in allochiral, coincident pendular motion so that the mass members 11 and 12 coact to impart additive loads in the ~
direction X and - X to absorb or coact with the vibrations traveling in that direction from the principal excitation foxce, such as a helicopter rotor. It will also be noted that all other forces generated by the pendular motion of the opposed mass means 11 and 12 will be mutually cancelling in that the forces generated by each mass members in direction Y will be cancelled by an equal force in the opposite direc-tion generated by the oppositely mounted mass means. In ~`
addition, the spring rate of the spring members are selected, so that, as best described in connection with the earlier description of Fig. 16, the force imparted by the spring members to the mass members increases with mass members angular motion amplitude, thereby causing the fixed frequency of the vibration absorber to remain substantially constant to thereby produce a substantially linear vibration absorber.
We wish it to be understood that we do not desire to be limited to the exact details of construction shown and described, for obvious modifications will occur to a person skilled in the art. -~
~ '.

, ~

Claims (33)

The embodiments of the invention in which an exclusive property or privilege is claimed are defined as follows:-

1. A fixed frequency vibration absorber adapted to be fixedly attached to a vibration-prone system to cooperate with the principal vibration excitation source which primarily generates vibrations in a given direction so as to control system vibrations and comprising:
base means, two mass means of selected equal mass, pendular connecting means connecting said mass means in opposed positions to said base means for support and pendular motion therefrom, and spring means operatively connected between said mass means in preloaded condition to exert a fixed force on said mass means to thereby establish the fixed natural frequency thereof and of the vibration absorber, and to also cause said mass means to move in pendular motion such that motion of said mass means produces additive forces in said given direction to absorb the vibration force established by said principal source, and so that all other forces so produced mutually cancel.

2. A vibration absorber according to Claim 1 wherein said spring member exerts a force to separate said mass means or to move said mass means closer together.

3. A vibration absorber according to Claim 2 wherein said connecting means comprises three pendular connections between said base means and each of said mass means, which connections are spaced in two substantially perpendicular directions to provide stability in the support of the mass means from the base means.

4. A vibration absorber according to Claim 3 and wherein said three pendular connections are located at the apex of a triangle to provide two directional geometric stability to each of the mass means from the base means.

5. A vibration absorber according to Claim 4 wherein each of said pendular connections comprises overlapping apertures of selected diameters and having parallel axes perpendicular to said given direction in said base means and said mass means, and a pin member of selected diameter and having an axis parallel to the aperture axes and extend-ing through the overlapping apertures thereby joining the mass means to the base means for pendular motion with respect thereto.

6. A fixed frequency vibration absorber adapted to be fixedly attached to a vibration prone system to cooperate with the principal vibration excitation source which primarily generates vibrations in a given direction so as to control system vibrations and comprising:
base means, two mass means of selected equal mass, pendular connecting means connecting said mass means in opposed positions to said base means for support from and pendular motion in said given direction, controllable spring means operatively connected between said mass means in selectively preloaded condition to exert a force on said mass means to establish the desired natural frequency thereof, and to also cause said mass means to move in pendular motion such that motion of said mass means produces additive forces in said given direction to absorb the vibration force established by said principal source, and so that all other forces so produced mutually cancel, said spring means also being of selected spring rate to compensate for the spring rate reduction caused by pendular motion of the pendular connecting means and thereby provide an essentially linear spring rate acting on said mass means for angles of pendular motion of at least ?+45°, so that the natural frequency of the vibration absorber is substantially constant throughout this range of operation.

7. A vibration absorber according to Claim 6 wherein said spring member exerts a force causing said mass means to move relative to one another.

8. A vibration absorber according to claim 7 wherein said connecting means comprises three pendular connections between said base means and each of said mass members, which connections are spaced in two substantially perpendicular directions to provide stability in the support of the mass means from the base means.

9. A vibration absorber according to Claim 8 and wherein said three pendular connections are located at the apex of a triangle to provide two directional geometric stability to each of the mass means from the base means.

10. A vibration absorber according to Claim 9 wherein each of said pendular connections comprises overlapping apertures of selected diameters and having parallel axes perpendicular to said given direction in said base means and said mass means, and a pin member of selected diameter and having an axis parallel to the aperture axes and extend-ing through the overlapping apertures thereby joining the mass means to the base means for pendular motion with respect thereto.

11. A vibration absorber according to Claim 10 wherein said diameters of said overlapping apertures are equal.

12. A fixed frequency vibration absorber adapted to be fixedly attached to a vibration prone system to cooperate with the principal vibration excitation source which primarily generates vibrations in a given direction so as to control system vibrations and comprising:
base means, two mass means of selected equal mass, pendular connecting means connecting said mass means in opposed positions to said base means for support from and pendular motion in said given direction, at least one spring member operatively connected between said mass means in selected preloaded condition to impose an internal force on said mass means to establish the natural frequency of the vibration absorber and to also cause said mass means to move in pendular motion such that motion of said mass means produces additive forces in said given direction to absorb the vibration force established by said principal source, and so that all other forces so produced mutually cancel, said spring member also being of selected spring rate to compensate for the spring rate reduction caused by the pendular motion of the pendular con-necting means and thereby provide an essentially linear spring rate reacted on the mass means for angles of pendular motion of at least ?45°, so that the natural frequency of the vibration absorber is substantially constant throughout this range of operation.

13. A vibration absorber according to Claim 12 wherein said spring member exerts a force to separate said mass means or to move said mass means closer together.

14. A vibration absorber according to Claim 13 wherein said connecting means comprises three pendular connections between said base means and each of said mass members which connections are spaced in two substantially perpendicular directions to provide stability in the support of the mass means from the base means.

15. A vibration absorber according to Claim 14 and wherein said three pendular connections are located at the apex of a triangle to provide two directional geometric stability to each of the mass means from the base means.

16. A vibration absorber according to Claim 15 wherein each of said pendular connection comprises overlapping apertures of selected diameters and having parallel axes perpendicular to said given direction in said base means and said mass means, and a pin member of selected diameter and having an axis parallel to the aperture axes and extending through the overlapping apertures thereby joining the mass means to the base means for pendular motion with respect thereto.

17. A vibration absorber according to Claim 12 wherein said spring member exerts a preload on said mass means to establish the initial natural frequency of the vibration absorber, and including means to change the preload exerted by the spring means between the mass means and thereby establish the desired fixed natural frequency of the vibration absorber.

18. A fixed frequency vibration absorber adapted to be fixedly attached to a vibration prone system to cooperate with the principal vibration excitation source which primarily generates vibrations in a given direction so as to control systems vibrations and comprising:
base means, two mass means of selected equal mass, pendular connecting means connecting said mass means in opposed positions to said base means for support from and pendular motion in said given direction and comprising:
at least one set of overlapping apertures of selected diameter in said base means and said mass means and having parallel axes perpendicular to said given direction, and a pin member of selected diameter and having an axis parallel to the aperture axes and extending through the overlapping apertures of each set thereby joining the mass means to the base means for pendular motion with respect thereto, at least one spring member operatively connected between said base means in selected preloaded condition to impose an internal force on said mass means to establish the natural frequency of the vibration absorber and to also cause said mass means to move in pendular motion such that motion of said mass means produces additive forces in said given direction to absorb the vibration force established by said principal source, and so that all other forces so produced mutually cancel, said spring member being of selected spring rate to compensate for the spring rate reduction caused by the pendular motion of the pendular connecting means and thereby provide an essentially linear spring rate reacted on the mass means for angles of pendular motion of up to at least ?45°, so that the natural frequency of the vibration absorber is substantially constant through-out this range of operation, and so that said apertures and pin are of selected diameters to produce a minimal envelope vibration absorber consistent with load carrying require-ments and so as to coact with said at least one spring member in establishing desired vibration absorber natural frequency.

19. A vibration absorber according to Claim 18 and including:
means to adjust the fixed preload of said spring member to adjust the fixed natural frequency of the vibration absorber.

20. A vibration absorber according to Claim 19 wherein said spring member exerts a force to separate said mass means or to move said mass means closer together.

21. A vibration absorber according to Claim 20 wherein said connecting means comprises three pendular connections between said base means and each of said mass members, which connections are spaced in two substantially perpendicular directions to provide stability in the support of the mass means from the base means, and wherein said three pendular connections are located at the apex of a triangle to provide two directional geometric stability to each of the mass means from the base means.

22. A helicopter having:
a fuselage, a lift rotor projecting from and supported from said fuselage for rotation and constituting the principal fuselage vibration excitation source which primarily generates vibrations in a vertical direction, a fixed frequency vibration absorber fixedly attached to said fuselage at a selected station therein to control fuselage vibration and comprising:
base means, two mass means of selected equal mass, pendular connecting means connecting said mass means in opposed positions to said base means for support from and pendular motion in said vertical direction, and spring means operatively connected between said mass means in selected preloaded condition to exert a fixed force on said mass means to thereby establish the fixed natural frequency thereof and of the vibration absorber and to also cause said mass means to move in pendular motion such that motion of said mass means produces additive forces in said vertical direction to coact with the vertical vibration force established by said rotor to reduce fuselage vibration and so that all other forces so produced by mass means motion mutually cancel and are therefore not imparted to the fuselage.

23. A vibration absorber according to Claim 22 wherein said spring member exerts a force to separate said mass means or to move said mass means closer together.

24. A vibration absorber according to Claim 23 wherein said connecting means comprises three pendular connections between said base means and each of said mass members, which connections are spaced in two substantially perpendicular directions to provide stability in the support of the mass means from the base means.

25. A vibration absorber according to Claim 24 and wherein said three pendular connections are located at the apex of a triangle to provide two directional geometric stability to each of the mass means from the base means.

26. A vibration absorber according to Claim 25 wherein each of said pendular connections comprises overlapping apertures of selected diameters in said base means and said mass means and having parallel axes perpendicular to said vertical direction, and a pin member of selected diameter and having an axis parallel to the aperture axes and extend-ing through the overlapping apertures thereby joining the mass means to the base means for pendular motion with respect thereto.

27. A helicopter having:
a fuselage, a lift rotor projecting from and supported from said fuselage for rotation and constituting the principal fuselage vibration excitation source which primarily gener-ates vibrations in a vertical direction, a fixed frequency vibration absorber fixedly attached at a selected station to said fuselage to control fuselage vibration and comprising:
base means, two mass means of selected equal mass, pendular connecting means connecting said mass means in opposed positions to said base means for support from and pendular motion in said vertical direction, controllable spring means operatively connected between said mass means in selected preloaded condition to exert a force on said mass means to establish the desired natural frequency thereof and to also cause said mass means to move in pendular motion such that motion of said mass means produces additive forces in said vertical direc-tion to coact with the vertical vibration force established by said rotor to reduce fuselage vibration and so that all other forces so produced by mass means motion mutually cancel and are therefore not imparted to the fuselage, said spring means being of selected spring rate to compensate for the spring rate reduction caused by pendular motion of the pendular connecting means and thereby provide an essen-tially linear spring rate acting on said mass means for angles of pendular motion of at least ?45°, so that the natural frequency of the vibration absorber is substantially constant throughout this range of operation.

28. A vibration absorber according to Claim 27 wherein said spring member exerts a force to separate said mass means or to move said mass means closer together.

29. A vibration absorber according to Claim 28 wherein said connecting means comprises three pendular connections between said base means and each of said mass members, which connections are spaced in two substantially perpendicular directions to provide stability in the support of the mass means from the base means.

30. A vibration absorber according to Claim 29 wherein said three pendular connections are located at the apex of a triangle to provide two directional geometric stability to each of the mass means from the base means.

31. A vibration absorber according to Claim 30 wherein each of said pendular connections comprises overlapping apertures of selected diameters in said base means and said mass means and having parallel axes perpendicular to said vertical direction, and a pin member of selected diameter and having an axis parallel to the aperture axes and extend-ing through the overlapping apertures thereby joining the mass means to the base means for pendular motion with respect thereto.

32. A vibration absorber according to Claim 2 wherein said spring means is at least one spring of selected spring rate.

33. A vibration absorber according to Claim 2 wherein said spring means is at least one spring of selected spring rate and at least one spacer member positioned in series with said spring between said mass means cooperating with said spring to exert a combined force on said mass means.