Note: Descriptions are shown in the official language in which they were submitted.
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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.
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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
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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
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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
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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 .
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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
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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 . - ~ -
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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
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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~
:
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
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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
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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. ;~
.
-15-
.
~ ~ ~3~
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 --
. .
-16-
: - . - . - . .
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
-17-
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
-18-
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 .
19-
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
-20-
, . . , -
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
-21-
~ . . .
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
.
-22-
,, ' ', ,
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,
-24-
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.
-25-
,
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
.
-26-
~ . .
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
-27
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
-28-
;. ,,
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
-29-
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'`,
.
- 30 -
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
-31-
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,
-32-
,
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. -~
~ '.
34 -
, ~
