CA1112674A - Fixed position variable frequency pendular-type vibration absorber - Google Patents

Fixed position variable frequency pendular-type vibration absorber

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Publication number
CA1112674A
CA1112674A CA332,863A CA332863A CA1112674A CA 1112674 A CA1112674 A CA 1112674A CA 332863 A CA332863 A CA 332863A CA 1112674 A CA1112674 A CA 1112674A
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Prior art keywords
mass
vibration
spring
vibration absorber
pendular
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Application number
CA332,863A
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French (fr)
Inventor
Kenneth C. Mard
Sylvester J. Washburn
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Raytheon Technologies Corp
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United Technologies Corp
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Abstract

FIXED POSITION VARIABLE FREQUENCY
PENDULAR-TYPE VIBRATION ABSORBER

ABSTRACT OF THE DISCLOSURE
A variable frequency vibration absorber adapted to be fixedly mounted in a fixed vibration prone system to coact with the system principal vibration excitation source to control system vibration. The vibration absorber is bifilar in construction and the natural frequency thereof is varied in proportion to the frequency of the vibration being generated by the principal system vibration.

Description

7~ ~:

-. BACKGROUND OF THE INVENTION

Field of Invention - This invention relates to ,;
vibration absorbers and more particularly to fixed , vibrat.ion absorbers ~hich utilize bifilar construction and in which the natural frequency of the vibration absorber :-~:
is vsried as a functiDn o~ the vibration bei~ng generated by the principal excitation source to thereby maintain the : , :
proper relationship between the absorber natural requency :~ `
and the:principal excitation source generated frequency to control system vibration. ;~
-.-~ Vescription of the Prior Art - In the fixed vibration ~ :~
absorber prior art the absorbers are basically fixed ~requency absorberswhich are capable of absorbing vibration ;~
over a relatively small range o frequency of the principal ~ .
excitation source, Typical of these absorbers are the swastika type absorber shown in U. S. Patent No, 3,005,520 ;~
to Mard and the battery absorber presently used in helicopters, which i.s basically a spring mounted weight and generally o~

~2-,, , " , , ., , . ~ .:, .. . .

~Z1~74 the type disclosed in Canadian Patent Application Ser. No.
329,227 entitled Tuned Spring-Mass Vibration Absorber by John Marshall II and filed on ~une 6, 1979. These prior art absorbers are fixed frequency abs~rbers which are capable of absorbing vibrations over a relatively 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 convention- ~ -ally been used solely on rotating mechanisms, such as crank-shafts of automobiles and aircraft engines and on helicoptar rotors as shown in Paul and Mard U. S. Patent No. 3,540,809. r In these instal1atlons, the centrlfugal force generated by rotation of the mechanism lnvolved is necessary for the operation of the bifilar-type vibration absorber. In a fixed position vibration absorber of the type sought in this application, centrifugal force is not present. In our absorber the force i8 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 friction and hence a high damping absorber and there fore a low amplification absorber so that it does not have `~
the advantages of our bifilar vibration absorber.

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67~ ~

S~MMARY OF ~IE INVENTION
____ A primary object of the present ;n~ention is to provide a vibration absorber o-f biilar construction and whose frequency varies as a function o~ the vibration frequency being generated by the system principal vibration excitation source so as to coact therewith in controlling ., system vibrations. ~ : ~
. .
In accordance with the present învention, the bifilar : vibration absorber:frequency is established by a biasing ~.
,:~ :
: IO spring force o selected spring rate acting against the ~ ~
;, ~ bifilar selected mass member or members to exert an . `
~: , internal orce thereon.. The spring-rate of the spring is selected to compensate for spring rate reduction normally caused by pendular e~cursions of the mass member ;:.
so:that the Sprlng rate ~and:hence the internal force~are ~-:substantially constant throughout pendular excursions of ~ ~;
:at least ~ 45 . In addition, mechanism is provided to control the spring deflection.as a unction of the vibra~
tion generated by~the pri.ncipal excitation source to ~ `; . .
:
thereby cause the bi~ilar absorber to become linear in :-:~
, nature and be ef~ectively operative over a substantially . .~.

large vibration span of the principal excitation source .:, to coact therewith in controlling s~stem vibration over that span.
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 bifilar principal to take ~ ~Z~7~

advantage of low inherent damping, low friction, low maintenance~ and high reliability characteristics, and ,~
which utilizes a spring o selected spring rate to compensate for the non-linear pendulum effect o the bifilar at high amplitudes.
It is a further object of this invention to teach a ~ ;~
vibration absorber which minimizes riction, and hence is a minimal damping absorber. This minimal friction and load 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 control 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 to provide such a vibration absor~er whose natural requency is varied as a function of the system principal vibration excitation source so that the vibration absorber is always operating at maximum effectiveness and so that the -frequency range over which the vibration absorber is effective to control system vibrat;on is increased. -~ t is still a further object of this invention to provide an improved vibration absorber utilizing pendular, preferably bifilar or trifilar principles, to obtain their low inherent damping, light weight and small envelope advantages and to utilize a spring to compensate for the ~ 67 ~

non-lirleclr pendulum effect when the absorber is used at high angular amplitudes, which would otherwise change the frequency of the system to maké the system ineffective.
By utilizing the pendular or bifilar principle and coil spring arrangement, the construction taught herein produces a vibration ahsorber having low inherent damping~ thereby permitting the use of lower dynamic masses in the biEilar `;
absorber, thereby not only reducing the weight of the -. .
vibration absorbex but also of the principal system, such a5 the helicopter.

It is an important teaching o~ our invention to utilize ,, :
a compre~sed coil spring 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 or a given rotor 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 teac-h 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 maximurn compression force and displacement of the ~ .

67~L ~

spring members at these maximum angular positions to thereby effect linear.ization of the vibration absorber so that its natural frequency is non-variant throughout its full range of pendular motion up to ~45.
In accordance with an embodiment of the invention, a variable freq~ency vibration absorber adapted to be fixedly attached to a vibration.prone system to cooperate with the ;
principal vibration excitation source which primarily gener- ~.
ates 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 tosaid base means for support and pendular motion therefrom, first means operatively con~
nected to said mass means in preloaded condition to exert a force on said mass means to thereby establish the natural frequency thereof and of the vibrat:ion absorber, and to also cause said mass means to move in pendular motion so 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, control means responsive to the .~ .
frequency of the vibrations generated by the principal vibration excitation source and operatively connected to ;
said first means to vary the force exerted thereby and thereby the natural frequency of said mass means as a function of the vibration frequency generated by said excitation source to thereby maintain the proper relation- ~:
ship between the mass means natural frequency and the principal source generated frequency to control system :.
vibration.

.

67~

In accordance with a further embodiment of the invention, a helicopter has: a fuselage, a lift rotor pxojecting from and supported from said fuselage for rotation ~::
and constituting the principal fuselage vibration ~xcitation .
source which primarily generates vibra-tions in a given direction, a variable frequency vibration absorber fixedly attached to said fuselage and operative to control fuselage 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 f-or ~; -.
support and pendular motion therefrom, spring means oper~
atively connected. bétween said mass means in preloaded condi- ;~
tion to impose a preload on each of said mass means to thereby establish the natural frequency of the vibration ~.
absorber, and to also cause said mass means to move in pendular motion so 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 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 essen- ~
tially linear spring rate reacted on the mass means for -angles of pendular motion of at least ~45~, means to vary the preload exerted by t:he spring means between the mass means and each of the base means and thereby establish the :
initial desired natural frequency of the vibration absorber system, and control means responsive to rotor RPM and oper-atively connected to the preload means to vary the spring force exerted on each of said mass means and thereby the natural frequency of said vibration absorber as a function ~-- 7a -~;2674 of rotor RPM, to thereby maintain the proper relationship between the vibration absorber natural frequency and the rotor RPM generated -frequency to control fuselage vibrations.
In accordance with a further embodiment of the invention, a variable frequency vibration absorber adapted to be fixedly attached to a vibration prone system which primarily generates vibrations in a given direciion so as :~
to cooperate with the principal vibration excitation source .
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 and pendulaI motion therefrom, :
controllable spring means operatively connected between said -~
mass means in preloaded condition to exert an lnitial force on each of said mass mPans to estab:Lish the initial desired natural frequency thereof and, to also cause said mass means :
to move in pendular motion~so that motion of said mass means : produces additive forces in:said gLven direction to absorb :~.
the vlbratlon force established by said principal source, ~ :~
and so that all other forces so produced mutually cancel, . .
said spring means being of selected spring rate to compen-sate 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, and control ~:
means responsive to the frequency of the vibrations generated ; by the principal vibration excitation source and operatively connected to said spring means to vary the force exerted thereby and hence the natural frequency of said mass means 7b -i' 67~
' '' as a function of the vibration frequency generated by said excitation source to thereby maintain the proper relationship - ~
bet~een the mass means natural frequency and the principal ~ ~:
source generated frequency to control system vibration. ~
In accordance with a further embodiment of the .. -invention, there is provided, in combination: a vibration prone system, a second system associated with the vibration prone system and operative to provide the principal vibration excitation force to said vibration prone system primarily in a given direction, a variable frequency vibration abosrber ~.
fixedly attached to said vibration prone system and oper-ative to coact with the principal vibration excitation force to control system vibrations and comprising: base means, :~
two mass means of selected equal mass, pendular connecting ~
means connecting said two mass means in opposed positions ~ -to said base means for support and pendular motion therefrom, at least one spring member operatively connected between said mass means in preloaded condition to impose a preload ~on said two mass means to establish the natural frequency ~;
of the vibration absorber, and to also cause said mass means . ~ ~
to move in pendular motion so that motion of said mass means produces additive forces in said given direction to absorb ~ :`
the vibration force established by said principal force, and so that all other forces so produced mutually, cancel, said ~ -at least one spring member being of selecte~ spring rate to compensate for the spring rate reduction caused by the pendular motion of the pendular connecting means and khereby provide an essentially linear spring rate reacted on the mass means for angles of pendular motion of at least ~5, means to vary the preload exerted by said at least one spring member on said two mass means and thereby establish - 7c -7~ -initial desired natural frequency of the vibration absorber system, and control means responsive to the frequency of the vibrations generated by the principal vibration excitation force and operatively connected to the preload means to vary the deflection of said at least one spring nlember and hence the spring force exerted on said two mass means and thereby the natural frequency of said vibration absorber as a function of the frequency generated by the principal vibration force, to thereby maintain tha proper relationship between the vibration absorber natural frequency and the principal vibration excitatlon force generated ~ -frequency to control vibration prone system vibrations.
Other objects and advantages of the present inven- ` "
tion may be seen by referring to the following description and claims, read in conjunctiQn with the accompanying drawings.
BRIEF_ DESCRIPTIO~I OF THE DRAWINGS
Fig. l is a graph showing helicopter fuselage ;~
vibration 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 bifilar connection between the bifilar base and the bifilar mass member to produce the desired low frictiQn, 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 mechanism illustrated as attached thereto, of the pre-ferred embodiment of our vibration absorber shown in Fig. 4.

- 7d s, Z~74 Fig. 6 is a side view, partially broken away, of. .
the vibration absorber of Fig. 4.
Fig. 7 is a view, partially broXen away, taken :
along line 7-7 of Fig. 6.

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1, ~ : ,; ' , ,' .
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Fig. 8 is a v;ew, partially broken away, talcen along line 8-8 of Fig. 5. ;~
Fig. 9 is a cross sectional showing o~ an actuator which could be used with our vibration absorber.
Fig. 10 is a showing of the natural frecluency control mechanism utili~ed with our absorber in the helicopter environment.
Fig. 11 is a graph of the fluid pressure acting on or the internal orce generated in the biEilar mass means -~

plotted against rotor RPM.~

Fig. 12 is a partial showing of the vibration absorber used as a fixed ~requency fixed position absorber, Figs. 13 and 14 are cross-sectional illustrations o-E
spacer means used in combination with the spring or springs ~;~
in the Fig. 12 embodirnent, ; ~;
Fig, 15 is a graph o the vibration absDrber natural ~ ;
frequency ratio plotted against pendular angular motion to illustrate the dif~erence in operation be~ween this linear vibration absorber and a conventional bifilar vibration absorber.

Fig. 16 is an illustration of the pendular motion of the dynamic mass members of our vibration absorber to illustrate the amplitude oE motion, angular motion amplitude, pendular arm length, and coil spring compression motion of the dynamic mass member.
Fig. 17 is a perspective showing of the base member o the vibration damper with the other parts of the vibration ab~orber rernoved therefrom, 7~ :~

DESCR:I.PTIO~I OF THE PF~EFERRED E2/lBODIMENT
The vibration absorber taught in this application will be described in the envi.ronment of a helicopter in which the vibration absorber is fi~edly mounted in a helicopter fuselage to coact with the principal helicopter vibration excitation source, n~mely, the rotor or rotDrs, to reduce the vibrati.on imparted to the fuselage thereby.
To appreciate the operation and advantages of this :
variable frequer..cy vibration absorber~ the short-comings o the fi~ed frequency, fixedly positioned prior art ~.
vibrat:ion absorbers will be discussed. Referring to Fig. 1 we see a graph A of helicopter fuselage vibrations plotted against helicopter rotor RPM. It is well known ::-that the vibration generated by a rotor and the response of the helicopter structure thereto varie~ and is a function o rotor RPM, as shown typically by graph ~. Line B
indicates the vibration line below which acceptable ~uselage .
vibrativn occurs The prior art ixed frequency~ fixed position vibration absorbers would operate generally along ;~.
graph C~ and it will be noted that such a vibration absorber is effective over range D between lines E and F. It will be noted that range D covers a 5mall variation or span in rotor R~M over whieh the prior art vibration absorbers are effective. Range D is determined by the mass ratio of the absorber, that îs the ratio of the weight of the absorber :~
to the effective weight of the substructure in which the absorber is fi~edly mounted~ and in part by the inherent ~ ~ ~ 2 6~ ~ ~

damping of the vibratîon absorber and the substructure ;
With the prior art absorbers, to obtain a relatively wide range C oE absorbing, a very heavy vibration absorber would ;
be required. Dimension G~ the minimum achievable vibration level, would be determined by the amounk of inherent damping ~;
in the absorber, and in part by the inherent substructure damping, and in part by the aforementioned mass ratio. If, theoretically, a vibration absorber could be utilized which has zero inherent damping, maximum vibration absorptioh would occur so as to achieve minimum fuselage vibration~
i.e.; d-imension G would be reduced. Such a system cannot ,:~
be realized in practice.
The objective of this vibration abso-rber is to be able to get maximum vibration absorption, indicat~ed by any point along line H over a greater range J o~ rotor RPM~ lines K
and L representing minimum and maxim~n necessary operating or exci~ation ~Ms of the helicopter or o~r Frinc~l s~tn~e~
To understand the purpose and operation o~ this vibration absorber, it is first necessary to ~mderstand~;
the diference between a vibration absorber and a vibration damper. A vibration damper serves to dissipate the energy of the vibrations imparted to the ~uselage by the rotor.
Vibration dampers can use friction principles or any type ~;
of energy damping principle. A vibration absorber, on ~;
the other hand, does not dissipate already established vibration energy but establishes a second vibratory mode in the system so as to coact with the principal system mode, the substructure mode, to produce a resultant mode which has minimum vibration. Stated another way, a vibration damp~r 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 oE one form of this vibra- ~;
tion absorber lO is shown in Fig. 2 In Fig. 2 masses 11 and 12~ of selected mass, are supported rom base members by suspension arm members a, which can be considered to be -~
pend~lous members as ilLustrated by the phantom line m~tion for mass 11. In practice, pendular arm a is actually the ~pin and bushing connection shown representati~ely in Fig. 3 in which pin member 1~ of diameter d is positioned in hole 16 of one of the mass members ll Gr 12 and overlapping hole -~
-:
17 in the~base member so as to produce an equivalent pendulum motion of pendulum arm a7 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 ll and 12 and serves to draw them together and thereby preloads the selected `~
masses so suspended to establish an internal force thereîn and thereby establish the natural frequency of masses 11 and 12, and therefore the natural fre~uency of absorber lO.
The natural frequency of masses 11 and 12, and hence absorber ~
10, is determined by the preload of sprin~ 18 and the mass ;
of mass members 11 and 12, which are preferably of equal mass Spring 18 performs '11 - .

267~ ~

another important function, in particular, it makes linear the non-linear characteristi~s of the pendular constructio~.
To explain this linear/non-linear concept, reerence will be made to Fig 2 It will be noted by viewing Fig 2 that as arms a pivot to move mass 11 from its solid line to its phantom line position, the spring rate of the conventional bifilar system~ eonsidering only the preload from spring 18 and not the spring rate, is reduced and thereore the natural frequency o the bifilar system is reduced to thereby reduce its effectiveness. This reduction in natural .
frequency of the mass member with amplitude causes the system to be non-linear, and limits its range of effective~
ness. This non-linear vibration characteristic of a pend~lar system occurs immediately upon any angular motion although a practical angle of excess would typically be 10. We could prevent the system from swinging beyond 10 by , increasing the length of the pendulum arms a but this would be undesirable because this would produce a heavier .
system requiring a larger space envelope.
With spring 18 present, however, as mass 11 swings from its solid line to its phantom line position, the changing force of spring 18 acting on mass 11 is increased, 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 its original value.
In this vibration absorber, we maintain the low weight and small space envelope advantage o a short pendulum ~12-6~

arm a, yet produce a linear system by c~ntrolling the natural frequency of the vibration absorber by manipulation of the force generated by spring 18 and imparted to the masses ll 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 force~applied thereto to establish system ~ -natural frequency by spring 18~ of selected preload and sprin~, rate, which serves to force masses ll and 12 to ". ,:
separate, For a more particular description o~ the preferred ,. .. . .
embodiment re~erence will now be made to Figs, 5-8 in which base mem~er 20, which is fixed to the uselage as shown in Fig. 8, supports selected mass members 11 and 12 therefrom in pendular fashion, Each mass member 11 and 12 is supported from the base member 20 by three pendular ;`~-connections similar to Fig. 3, thereby forming a trifilar connection, and each of the three connections including, as best shown in Fig. 8; an aperture 22 in masses 11 and ;
12 and an over1apping aperture 24 in base 20 and each having a p;n member 26 extending therethrough. As best shown in Fig. 5~ each mass means 11 and 12 is connected to base member 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, ~ 67 ~

the perldular conn~ction at s~ation S2 is at the bottom of each mass ~hile pendular connections Sl and S3 are at the top o ~ach mass. In vie~ of this three stat;on connection~ reminiscent oE the tnree-legged stool, the mass is given geometri.c stability as supported from bas~
20 in both the yaw direction s~o-.m in Figs. 5 and ~ and the pitch direction shown in Fig. 8~ It will therefore be seen that to this point our vibrat;on absorber includes two mass members 11 and 12 supported in selectively spaced 1~ connecting stations from base.member 20. The connections ;~
may be of the type more ully disclosed in ~. S. Patent No. 3,5~0,809 to W. F. Paul e~ al. In Fi$~ 7, one of two spr.ing members 18 is shown extending between masses 11 and ~.
12, utilizing spring re~ainers 28 and 30. Springs 18 are o selected spring rate so that ~7hen installed and preloaded, --~
the springs provide the necessary internal force to mass - members 11 and 12 to establish a selected natural frequency o masses 11 and 12 and thereore of vibration absorber 10. -~
With sprin~ 18 assembled as shown in Fig. 7 and preioaded it will be observed that the spring serves to impart a separating force to mass means 11 and I2~
The construction of base member 20, which is preferably of one~piece construction, is very important to this invention ~s best shown in Fig. 17, base member 20 comprises flat platorm 51 extending longitudin~lly of the base member as sho~Jn in Fig 17 and constituting a solid base for the : base member 20 so that platform 51 may be attached in any : conventional fashion, such as by nuts and bolts~ to the fi~ed vibration prone system which our vibration absorber -1~

67~

is intended to operate in Three parallel, laterally extending plate members 53, 55 and 57 extend perpendicularly from platform 51 and extend in the lateral direction, which is the direction or plane of desired mass memberm~on. ~p~e -~
members 53 and 57 are identical in shape and project a substantially grea~er height out of platform 51 than does central plate member 55. Plate members 53, 55 and 57 each have equall~ l~terally spaced apertures 59 and 61, 63 and ;~
65, 67 and 69 therein, respectively. ~pertures 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 helght above platform 51g 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 diameter ap~rtures having parallel axes and with each aperture positioned at the corner of a triangle. The first three aperture set consists of apertures 61, 65 and 69, while the second aperture set consists o apertures 59, 63 and 67 These two aperture sets are parallel to one another and, in view of the fact 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 79~ ~

pendular connection are offset in two perpendicular direc-~ions, ~hich 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 e~ual height apertures 59 and 67 This three point triangular-type connection between the mass members and the base member provide geometric stability so as to prevent both roll and yaw tumbllng of the mass members with respect to the base member.
With respect to the construction of plate members 53 55 and 57 and in particular their construction 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 p~ate member 57, apertures 67 . . .
and 69 have two parallel beam portions 71 and 73 extending la~erally across the plate member above and below the apertures and structural web section 75 extending between beam members 71 and 73 at a station between apertures 67 and 69 so as to form an I--shaped structure, formed by beam members 71 and 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 located. In act, this X~shaped structure is strengthened by the fact that its ends are closed at portions 77 and 79 to form a closed 6~

box construction consisting of sections 71~ 77, 73 and 79, with structural web section 75 e~tending through the center `~
thereo0 Mass member loads reacted by plate mernber 57 at apertures 67 and 69 are imparted to plate member 57 at this high strength structural section and therefrom into platfo~n member 51 for transmittal to the fixed vibration prone `~ ;~
system, such as the fuselage o the helicopter. The 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 oE the pLate members `
in the vicinity of the apertures7 plate members 53 and 55 ;
are similarly constructed. ~ ~-As best shown in Figs. 5 and 6~ the ùynamic mass members 11 and 12 ext~nd 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, 25, 27 29 and 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 spaed longitudinally with respect to base member plate member 53, the second set 25 and 27 being ~, ; positioned on opposite sides of base plate member SS and selectively spaced longitudinally with respect thereto, and third set 29 and 31 positioned on opposites o-f base plate member 57 with selected longitudinal spacing there-between.
As best shown in Figs. 5 and 7,each parallel cDmpression coil spring 1~ is received at its opposite ends in spring end retainers 28 and 30, which retainers are supported in m~ss 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 ~ -o 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 ~f the mass members and each hag a cylindrical, flanged bushing inserted therein as shown, which bushing is fabricated 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. 5 and 6.

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These pin members which are visable are designated as 71, 73 and 75 but it shouLd be noted that each mass member 11. `:`
and 12 is connected to and 5upported 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 member5 71-75 are fabricated of an , :
anti-friction material such as a car~onized steel, As can - be best seen in Fig. 67 these ~endular connecting stations Sl, S2 and S3 are longitudinally of~set from each other to `~
provide geometric stability between the mass members and the base members 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 base member, Due t~ 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 transl2ti~nal motion with respec~ to the base member so as to be parallel : :
20 thereto at all times, To minimize friction and hence damping of the system, each pin member includes a tapered c;rcumferential flange illustrated in Figs. 5 and 6 in connection with pin 73 only and indicated at 83 and 85g however all pin members have such tapered flanges, Flanges 83 and 85 are positioned in the longitudinal spacing 87 `~
and 89 between the bushing apertures th~ough which pin member 73 extends and are tapered in a radially outward -19~ :~

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 full mode of pendular operation therebetween~
It will therefore be seen that this vibration absor.ber produces min-imal friction, solely the minimal .
flex;ng -friction o-f the coil spring members 18 and the rolling friction of roller members:71-75. This vibration ., ~ .
absorber is thereore low in damping~ high in amplificat-lon, 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 de1ection, free length and mean diameter :be selected so that the coil spring îs statically stable when its- ground ends are positioned between spring retainers
2~ and 30. I~e importance of this spring static stability ;~
is that it does not require additional spring support :;
mechanisms7 such as a center spring guide, since such `.~`
would.add weight, friction and damping to the sys~em to thereb~ reduce the efectiveness o the vibration absorber. .
It should be noted that maxim~n spring deflection 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 ~20-7~

could cause fretting or friction, both of which are detrimental to absorber life or performance. Xn addition, both the transverse and axial natural requencies of the spring are selected to be detuned from the system excitation frequencies so as to avoid exc~ssive spring motions, since any touching ~-of parts caused thereby could produce fretting or friction, both of which are detrimental to absorber life or performance.
~ ctuator 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 ~ -requency of vibration absorber 10. Actuator 32 may thereafter b& 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 spring 1~ 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 from 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 o-f rotor RPM to coact with vibration excitation forces imposed on the fuselage by the rotor to thereby reduce fuselage vibration~
The construction of actuator 32 may best be understood by viewing Fi~. 9. The actuator consists of telescoping -21~

2~7~ :~

sleeve members 34 and 36, the ormer 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 1uid 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 orce causing sleeve mem~er ~ :
~4 to move letwardly with respect to ~ixed member 36 to : :~
thereby compress spring 18 as it so moves. This compression of spring 18 adds to the internal force applied to mass means 11 agalnst which it directly bears through retainer 28 Similarly, due to the fluid pressure so exerted 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 wh;ch is o conventional design and operates .: ~
, ~
in typical rheostat fashion to send a position feedback ~:
signalS representative of the position o~ movable member 34 as determined by the pressure in chamber 44, to the actuator control system 47. rnere are other prior art actuators which could be used in this vibration absorber, for example, the positioning actuator sold under part number A-24553~2 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 Division of Simmonds Precision, Cedar Knolls~ New Jersey .

~22-,.':;
Attention is now directed to Fig. 10 for an e~planation ~ -of the control system 47 used to vary the natural frequency of absorber 10 as a function of rotor RPM. As shown in Fig. 10, helicopter rotor 42, possibly through a tachometer, imparts a rotor speed (RPM) signal to controller 44. The ~`
,;, ~. ~,.
control~er 44 operates to provide a signal on a line 74 to the absorber 10 that is proper to control the valves in the a~sorber 10 to provide displacement as the square o rotor ``~
....
speed within an operating range of rotor speeds as is ~
''. '.
described with respect to Fig. 11 hereinater. Assuming the rotor 42 provides a tachometer signal on a line 76 , which varies in frequency as a f~mction of rotor speed, conversion to a DC voltage proportional to ro~ r speed may ~ - --be made by any conventional frequency-to-voltage converter 78a which mayg for instance, comprise a simple integraLor, ~;
or a more complex converter employing a Teledyne Philbrick ~, 4708 frequency-to-vol~age 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 both inputs of an analog multiplier circuit 82, of any well known type, so as to provide a signal on a line 84 which is a unction of the square of rotor speed. A potentiometer ~
86 is provided to allow a gain adjustment, whereby the ~`
overall efect of the control can be adjusted to suit each particular aîrcraft. This provides a suitable signal on a line 88, which is some constant t-imes the square of ~ 2679L

rotor speed, to a surnming amplifier 90, the other input of which is a feedback error signal on a li~e 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 o~ the summing ampliier:: ' 90 provides a signal on the line 7~ to direct the actuator : .
to a posltion determined as some constant times the square o~ rotor speed, which position is maintained in closed. ~;

loop fashion by the`~eedback signal~on:a line 94, as modified by the bias provided by the source 96, The ,`~
blas resulting from the source 96 will cause the 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 ll and:, 12 and therefore absorber lO. This actuator preloading ;
is done so that actuator 32 can:reciprocate either leftward~
ly or rightwardly and thereby vary the internal force being imposed upon mass means ll and 12 in response to bot~
rotor RPM increases and rotor RPM decreases. It will be :~
realized that if actuator 32 were installed in its end travel position, it could respond to rotor RPM changes ~
in one direction only, ;~i The controller 4~ is thus prograrr~ed to send a hydraulic pressure signal proportional to rotor RP~ ~:
to absorber lO and absorber lO provides an actuator -2~-.. . ....

:' ~ 7 position feedbclck signal to the controller 4~. It will be noted that this absorber is fixedly mounted frorn the `~
fuselage.
Attention ;s now directed to Fig. 5 for a further explanation of this control system ~7. The pressure signal ~rom controller 44 goes to hydraulic valve 46, which receives aircr~ft supply pressure through line 48 ;;
and has hydraulic return line 50O Selectively pressurized ~ -~
hydraulic 1uid passes through 1exible pressure line 52 -~
into common pressure line S4 from which it enters the two actuators 32a and 32b to selectively change the force being e~erted by springs 18 on mass means 11 and -12 and hence the natural frequency~thereof and of the vibration absorbPr 10. siInilarlyg position feedback : - . , signals ~rom each actuator 32a and 32b are brDught through position feedback line 56 to controller 44.
This control system 47 is an open loop position ~eedback system because it is preprogrammed, that is, it has been calibrated in the laboratory to return to a given position. It will be evident to those skilLed in the art that this control system also has the capability of acting as a closed loop position feedback system.
The operation o our absorber is illustrated in the graph shown in ~ig. 11 in which the pressure in the flexible pressure line 52 or the internal force imparted to the mass means 11 and 12 by spring 18 is plotted against rotor RPM (NR). Biasing put into the system -~5- ,:

~ 7 causes the pressure to be flat in the low ~PM range, which is below the operating range, and then follows the curved - ~

graph portion representative of the formula ~ NR
where~ is the pressure and NR is the rotor speed (RPM~
It will therefore be seen that over ~he region designated as l'Operating Range" the force acting upon the vibrations absorber 10 to vary its natural frequency varies as a ~- -function o rotor speed, in particular, rotor speed squared.
This IlOperating 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 li~it the useful motion~o~ 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 absor~er is an improved vibration absorber -utilizing bifilar principles to take advantage of the ligh~eight~ small dimensional envelope, the low inherent -~
damping thereof~ the high reliability thereof, the low ~-riction generated thereby, and the minimum maintenance ~, .
required therefore Ihis vibration absorber also utilizes a spring to compensate for the non-linear pendulum e~fect of the pendular-type vibration absorber at high amplitudes, thereby making the absorber linear. It will further be realizecl that this ~ibration absorber changes its natural ~ L26~
frequency as a function of rotor RPM so that the absorber will always be operating at its maximum level of effective-ness to reduce fuselage vibration due to rotor eæcitation.
The absorber spring 18 is a selected spring rate which is controLled to initially preload the selected bifilar mass mem~ers to establish the initial natural frequency of the mass mem~ers and the absorber. The ~ibration absorber is thereafter controlled to vary the amount of loading by , . ~.
the spring on the absorber mass members as a unction of 10 rotor RPM to permit effective vibration absorption over a large span o 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 skilled 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 vibratio~ excitation source, as a function 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 vibration absorber o the pendular~type with provisions for absorber natural requency 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 capability, ,, ~"

67~
: .

because the vibration absorber so used as a fixed natural freq-lency absorber will still have the inhe~ent 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 absor~er 10 as a fixed frequency vibration absorber~ When used as a fixed frequency vibration absorber as shown in Fig. 12, the absorber construction will 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 2~ but positioned at the opposite end of spring 18 there-rom and acting aga;nst mass member 12. With 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 pendular-type connections shown in Figs. 5-8 from base member 20 and with spring or springs 18 applying a force thereto tending to separate the mass means 10 and 12 ~ ;~
The natur~l frequency of the Fig. 12 fixed frequency vibration absorber is determined by the mass of mass means 11 and 12 and the spring preload and spring rate of spring or springs 18. ~s in the Figs. 5-8 varia~7e frequency absorber~ the Fig 12 fixed frequency absorber is also linear in the same fashion.

-~8-~$1~67~

It may be desired to modlfy the Fig~ ~.2 Eixed frequency .
modificatian as shown in FLgs~ 13 or 74 t:o permit a degree of adjus~ment in establishing the preload .Force exerted by spring 18 and hence the natural frequency of a~sorber lO prior to or ater its instal.l.ation ~it:her as a subassembly or aiter i.nstallation in the substrtlcttlre requiring vibration suppression ~ut not during operation. Viewing Fig. 13 we see a cross-sectional showing of spacer member 62 comprlsing .
inner and outer continuous ancl t~.readed ri.ng members 64 and 66 in threaded engagement. with on~ anot~er so that the ring members 64:and 66 may be ~otated manually rela~îve , .
to one another through ~he~space æhown in Fîg~ 12 between :
mass members ll and 12 thereby -varying the ~idth or spac1ng . ;~
dimension o:~ variable spacer ~2 to vary lhe orce exerted by spring l8 on mel~ber~:l.l and 12~ 5pring 18 mar be ~
~ne or two-piece constructioLlO Viewi~.g ?Ig~ 14 we see spacer ring 683 sho~m in par~ial cross~secLion9 bet~ween : ;
one or two-piece spring 18 to serve as a .spacer ring therebetween to vary the force e-xerted l~y the comb;nation of spacer 68 and spr;ng or springs 18 ~l~ mas.ses ll and 12 Spacer 68 is preferably of two or more piece~ segmented ;:
construction so as to be manually posLtionabLe through . : ~;
the area shown in Figo 12 between mem~ers 11 a.nd 12 and joined by conventional connecting means to fonm a oontinuous spacer ring 68 as illustrated. 0~ cou~seg for fixed frequency operation actuator 32 could l~e used but adjusted to a ~ixed posi tion to preload springs 18 t:o establish a 67~ ~

fixed natural frequency for absorber 10. To provide a better understanding of 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 which is available in our pendular-type vibration absorber by est-imating 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 of the vibration absorber dynamic masses, which~ for example, happens to be four (4) times rotor RPM for a four bladed ~;
rotor, the required absorber dyna~ic masses displacement operating travel, which is the absorber useful amplitude, can be established.
~.
ZO Having determined thi.s useful motion or useful amplitude of our pendu~ar-t~pe vibration abæorber3 one can then determine the pendular length necessary to achieve maxim~n mass member desired angular displacement which we chose to be ~ 45~. Thiæ was done by utilizing the equation:
~ sin 1 X

Where: ~ = angular di.splacement of the mass rnember relative to the base member.
X = useful amplitude or motion, and a = the pendular lengths and is equal to D - d, where ~ is the bushing diameter of the base member and mass members apertures, and d is the pin diameter The significance 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 disp~acement of -~4~ 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 determines the amount of compression of the spring members. It i 5 important to note that the springs are deected or cycled twice for each full cycle of "X' motion of the mass members. In this connection, it will be noted that when the mass member starts its downward motion from its -~ X pOSitiOll, which is also its full angular ;
motion ~ ~ position, the spring is maximally compressed the full distance Y and that the spring is also maximaLly compres-sed the ~ull distance Y when the mass concludes its downward motion at position - ~, which is also its full angular motion ~
- ~position. This is the characteristic of our pendular- ~ -type vibration 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 hence the -:

67~

non-variant natural frequency of the vibration absorber.
While it is an inherent disadvantage in a pendular cons~ruc~
tion that it becomes more non-linear as the angular dis~ ; -placement of the absorber mass members increase~ this is overcome in our construction in that the spring is compres-sed its greatest at the points of ma~imum angular dis-placement to thereby maximize the inter~al force exerted -:
by the springs on the mass members at that point, and thereby retain a first order linearity so that the natural ~ ~~
~:.
frequency o~ the absorber is non-variant with angular displacement. Maintaining linearity is important to maintaining high absorber amplification so that small dynamic masses can be operated at large useful amplitudes ;~
to obtain the necessary inertial reaction orces 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 e~fective spring rate, or progra~ed rate in the case of a variable tuned absorberS necessary to achieve the desired absorber natural frequency or frequencies. This procedure is ~ully outlined in Den Hartog's woxk on "Mechanical Vibrations'.' The internal steady load requirement or the absorber can be arrived at by the formula:

Fnr = (K~) (a) Where: Fnr = the internal steady load between the absorber -~32-:

masses for the various rotor speeds. ~ ;
Kx = the effective spring rate, and a is the length of the pendular arm, i~e., D - d.
~ ow 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 aircraftls impedance and excitation frequencies. This force Fnr is comparable to the centrifugal force for an ~
absorber installed in a rotating system.
The derived equations of rnotion 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:
_ i K5 = Fnr /a ( 1 - ~- ) ( 2 sin O ) ~ (~ cos ~ ) '~ .~ .
Where: Ks ~ the preferred spring rate of the physical ;
ZO spring. By considering normal operating conditions, typical values of Fnr and 4 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. ~o. 331,688 by John Madden filed on July 12, 1979 and entitled "Constant Frequency Bifilar Vibration Absorber'~

' :

" ,~-~ ,6 Using conventional methods, the steady and vibratory loads o~ the spring can be determined from previous data selected or establ-ished.
Then~ 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 checked with respect to static stability of the physical spring when placed between the : -, spring retainers o the mass means under the load conditions imposed. Again, conventional approaches can be used to establish the permissible relationships for the compression ~`
eoil spring which was chosen, for example, between the spring ~ ~`
ree 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 need of guides, since such guides are likely to result in points o contact and introduce sliding friction which will increase the absorber's damping and reduce its performance. This basic spring technology is well known and fully explained in A. M. Wahl's book entitled "Mechanical Springs". -The transverse and axial installed spring natural frequencies for the springs under consideration must be checked out to determine that neither is close to the excitation frequencies o other absorber elements so as to avoid resonance therebetween, which could bring about metal-to-metal contact and cause fretting or introduce friction damping. The final relative motions determined for the selected spring then determine the clearances .

;~ !

~;26~4 between vibration absorber components, for example, the radial clearance between the springs and the dynamic masses.
Since the connecting pins of the pendular~type connection will onLy contact the aperture bushings when the p-ns are subjected to compressive loading, it is necessary to determine all of their instantaneous applied loads from the spring and the in~rtia 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 combined applied force of the spring and the mass members inertia loads have a resultant vector with a line of action whîch at all times extends between two sets o 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 bushings - results in positive stability. Also, locating the dynamic mass c.g. c~ose to the pins/aperture 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 mo~ion. Positive reaction capability is determined by determining the pin instantaneous loading and its coefficient o friction with respect to the bushing. This absorber-was determined to have no problems in this regard and there-for one piece, 7~ ~

solid pin members were used.
Kno~ing 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 vibration absorber taught herein is adapted to be fixedly attached to a vibration-prone system to cooperaté 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 vibration absorber comprises a base member having two mass members of selected equal mass supported from the base member in opposed posltions preferably on oppoeite sides thereof, through pendular connecting means which support the mass means for allochiral pendular motion in the direction of the primary source vibrations. As used ;~ ;~
herein, allochiral means mirror-image~ Spring members extend between the mass means in preloaded condition to perform the dual function of exerting a fixed force on the ;~
mass means to thereby establish the natural 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 produces additive forces in the direction of the principal source vibrations to absorb or coact with the vibration force established by the princi-pal 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 produced by the mass means pendular motion are mutually cancelled This will be best understood by viewing FigO 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 11 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 frcm the principal excitation force, 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 direc- `~
tion Y will be cancelled by an equal force in the opposite direction generated by the opposite:Ly 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. The principles of operation just described are the same for both -the fixed frequency absorber of Fig. 12 or the variable frequency absorber disclosed in Figs. 5 and 10 and claimed herein 6~
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.

:'., .

'` '' . ;

~''' '`'~"''' ~ 38 -

Claims (40)

The embodiments of the invention in which an exclusive property or privilege is claimed are defined as follows:-
1. A variable 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, first means-operatively connected to said mass means in preloaded condition to exert a force on said mass means to thereby establish the natural frequency thereof and of the vibration absorber, and to also cause said mass means to move in pendular motion so 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, control means responsive to the frequency of the vibrations generated by the principal vibration excitation source and operatively connected to said first means to vary the force exerted thereby and thereby the natural frequency of said mass means as a function of the vibration frequency generated by said excitation source to thereby maintain the proper relationship between the mass means natural frequency and the principal source generated frequency to control system vibration.
2. A vibration absorber according to Claim 1 wherein said first means includes spring means operatively connected between said two mass means to impose a preload on each mass means to establish the natural frequency of the mass means and hence of the vibration absorber, and wherein said control means act upon said spring means, which, in turn, acts upon each of said two mass means.
3. A vibration absorber according to Claim 2 wherein said spring means exerts a force to move said mass means relative to each other.
4. A vibration absorber according to Claim 3 and wherein said first means also includes an actuator positioned in series with said spring means between said two mass means and operable to vary the spring deflection of said spring means and hence the force exerted thereby against said two mass means to thereby vary the natural frequency of the two mass means and the vibration absorber.
5, A vibration absorber according to Claim 4 wherein said actuator is a hydraulic actuator comprising cylinder-piston means having a fixed and a movable member, with the movable member supported to be movable in response to fluid pressure to move said movable member to thereby vary the deflection of said spring means and therefore the force exerted by said spring means on said two mass means, and hence the natural frequency of the two mass means and the vibration absorber.
6. A vibration absorber according to Claim 5 wherein said control means includes a controller programmed to pro-vide hydraulic actuating fluid to said actuator at a pressure proportional to the vibration frequency being generated by the system principal vibration excitation force, and having position feedback means providing an actuator movable member position signal to said controller.
7. 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.
8. A vibration absorber according to Claim 7 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.
9, A vibration absorber according to Claim 8 wherein each of said pendular connections comprises overlapping aper-tures of selected diameters and having parallel axes perpen-dicular 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.
10. A vibration absorber according to Claim 6, wherein said controller is programmed to provide hydraulic fluid to said actuator at a pressure proportional to the square of the frequency of the system principal vibration excitation force.
11. 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 given direction, a variable frequency vibration absorber fixedly attached to said fuselage and operative to control fuselage 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, spring means operatively connected between said mass means in preloaded condition to impose a preload on each of said mass means to thereby establish the natural frequency of the vibration absorber, and to also cause said mass means to move in pendular motion so 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 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 at least +45°, means to vary the preload exerted by the spring means between the mass means and each of the base means, and thereby establish the initial desired natural frequency of the vibration absorber system, and control means responsive to rotor RPM and operatively connected to the preload means to vary the spring force exerted on each of said mass means and thereby the natural frequency of said vibration absorber as a function of rotor RPM, to thereby maintain the proper relationship between the vibration absorber natural frequency and the rotor RPM
generated frequency to control fuselage vibrations.
12. A vibration absorber according to Claim 11 wherein said two mass means are supported from said base means by said pendular connecting means and wherein said spring means is operatively connected between said two mass means to impose a preload on each mass means to establish the natural frequency of the mass means and hence of the vibration absorber, and wherein said preload varying means and said control means act upon said spring means which in turn, acts upon each of said two mass means.
13. A vibration absorber according to the preceding Claim 12 wherein said spring means exerts a force to move said mass means relative to each other.
14, A vibration absorber according to Claim 13 wherein said preload means includes an actuator positioned in series with said spring means between said two mass means and which is operable to vary the spring deflection of said spring means and hence the force exerted thereby against said two mass means to thereby vary the natural frequency of the two mass means and the vibration absorber.
15. A vibration absorber according to Claim 14 wherein said actuator is a hydraulic actuator comprising cylinder-piston means having a fixed and a movable member, with the movable member supported to be movable in response to fluid pressure to move said movable member to thereby vary the spring deflection of said spring means and therefore the force exerted by said spring means on said two mass means, and hence the natural frequency of the two mass means and the vibration absorber.
16. A vibration absorber according to Claim 15 wherein said control means includes a controller programmed to provide hydraulic actuating fluid to said actuator at a pressure proportional to rotor RPM and having position feedback means providing an actuator movable member position signal to said controller.
17, A vibration absorber according to Claim 16 wherein said connecting means comprises three pendular connections between said base means and each of said two mass means, which connections are spaced in two substantially perpen-dicular directions to provide stability in the support of the mass means from the base means.
18. A vibration absorber according to Claim 17 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.
19. A vibration absorber according to Claim 18 wherein each of said pendular connections comprises overlapping apertures of selected diameters in said base means and said mass means, and a pin member of selected diameter extending through the overlapping apertures thereby joining the mass means to the base means for pendular motion with respect thereto.
20. A vibration absorber according to Claim 16 wherein said controller is programmed to provide hydraulic fluid to said actuator at a pressure proportional to the square of rotor RPM.
21. A variable frequency vibration absorber adapted to be fixedly attached to a vibration prone system which primarily generates vibrations in a given direction so as to cooperate with the principal vibration excitation source 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, controllable spring means operatively connected between said mass means in preloaded condition to exert an initial force on each of said mass means to establish the initial desired natural frequency thereof and, to also cause said mass means to move in pendular motion so 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 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, and control means responsive to the frequency of the vibrations generated by the principal vibration excitation source and operatively connected to said spring means to vary the force exerted thereby and hence the natural frequency of said mass means as a function of the vibration frequency generated by said excitation source to thereby maintain the proper relationship between the mass means natural frequency and the principal source generated frequency to control system vibration.
22. A vibration absorber according to Claim 21 wherein said spring means exerts a force to move said mass means relative to each other.
23. A vibration absorber according to Claim 22 and wherein said spring means are controlled by an actuator positioned in series with said spring means between said two mass means and operable to vary the spring deflection of said spring means and hence the force exerted thereby against sand two mass means to thereby vary the natural frequency of the two mass means and the vibration absorber.
24. A vibration absorber according to Claim 23 wherein said actuator is a hydraulic actuator comprising cylinder-piston means having a fixed and a movable member, with the movable member supported to be movable in response to fluid pressure to move said movable member to thereby vary the spring deflection of said spring means and therefore the force exerted by said spring means on said two mass means, and hence the natural frequency of the two mass means and the vibration absorber.
25. A vibration absorber according to Claim 24 wherein said control means includes a controller programmed to provide hydraulic actuating fluid to said actuator at a pressure pro-portional to the vibration being generated by the system principal vibration excitation force, and having position feedback means providing an actuator movable member position signal to said controller.
26. A vibration absorber according to Claim 25 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.
27. A vibration absorber according to Claim 26 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.
28. A vibration absorber according to Claim 27 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 two mass means to the base means for pendular motion with respect thereto.
29. A vibration absorber according to Claim 25, wherein said controller is programmed to provide hydraulic fluid to said actuator at a pressure proportional to the square of the frequency of the system principal vibration excitation force.
30. In combination:
a vibration prone system, a second system associated with the vibration prone system and operative to provide the principal vibration excitation force to said vibration prone system primarily in a given direction, a variable frequency vibration absorber fixedly attached to said vibration prone system and operative to coact with the principal vibration excitation force to control system vibrations and comprising:
base means, two mass means of selected equal mass, pendular connecting means connecting said two mass means in opposed positions to said base means for support and pendular motion therefrom, at least one spring member operatively connected between said mass means in preloaded condition to impose a preload on said two mass means to establish the natural frequency of the vibration absorber, and to also cause said mass means to move in pendular motion so that motion of said mass means produces additive forces in said given direction to absorb the vibration force established by said principal force, and so that all other forces so produced mutually, cancel, said at least one 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 at least +45°, means to vary the preload exerted by said at least one spring member on said two mass means and thereby establish initial desired natural frequency of the vibration absorber system, and control means responsive to the frequency of the vibrations generated by the principal vibration excitation force and operatively connected to the preload means to vary the deflection of said at least one spring member and hence the spring force exerted on said two mass means and thereby the natural frequency of said vibration absorber as a func-tion of the vibration frequency generated by the principal vibration force, to thereby maintain the proper relationship between the vibration absorber natural frequency and the principal vibration excitation force generated frequency to control vibration prone system vibrations.
31. A vibration absorber according to Claim 30 wherein said at least one spring member exerts a force to move said mass means closer together.
32. A vibration absorber according to Claim 30 wherein said at least one spring member exerts a force to separate said mass means.
33. A vibration absorber according to Claim 30 and wherein said preload means includes an actuator positioned in series with said at least one spring member between said two mass means and operable to vary the spring deflection of said at least one spring member and hence the force exerted thereby against said two mass means to thereby vary the natural frequency of the two mass means and the vibration absorber.
34. A vibration absorber according to Claim 33 wherein said actuator is a hydraulic actuator comprising cylinder-piston means having a fixed and a movable member, with the movable member supported to be movable in response to fluid pressure to move said movable member to thereby vary the spring deflection of said at least one spring member and therefore the force exerted by said at least one spring member on said two mass means, and hence the natural fre-quency of the two mass means and the vibration absorber.
35. A vibration absorber according to Claim 34 wherein said cylinder-piston means are telescoping sleeve members.
36. A vibration-absorber according to Claim 35 wherein said control means includes a controller programmed to pro-vide hydraulic actuating fluid to said actuator at a pressure proportional to the vibration frequency being generated by the system principal vibration excitation force, and having position feedback means providing an actuator movable member position signal to said controller.
37. A vibration absorber according to Claim 30 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.
38. A vibration absorber according to Claim 37 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.
39. A vibration absorber according to Claim 38 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.
40. A vibration absorber according to Claim 36 wherein said controller is programmed to provide hydraulic fluid to said actuator at a pressure proportional to the square of the frequency of the system principal vibration excitation force.
CA332,863A 1978-08-04 1979-07-27 Fixed position variable frequency pendular-type vibration absorber Expired CA1112674A (en)

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US93108478A 1978-08-04 1978-08-04
US931,084 1978-08-04

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