Note: Descriptions are shown in the official language in which they were submitted.
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~
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,, , " , , ., , . ~ .:, .. . .
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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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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
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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
~ .
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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.
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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
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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 -
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' ''
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
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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.
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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.
, , :
1, ~ : ,; ' ,
,' .
's . ~
, ,
"~;
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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,
--8--
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
-10-
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,
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~ 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
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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
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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
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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
-17
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.
-18-
`: :
.~
67~ ~ ~
'~`,.
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-
74
,.':;
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
-23-
~ 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 - ~
2 -
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
-26-
~ 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
-30-
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-
:
67 4
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'~
'
:
- 33 -
" ,~-
~ ,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
-34-
.
;~ !
~;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
- 36 -
674
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
- 37 -
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 -
