Note: Descriptions are shown in the official language in which they were submitted.
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Back~round o~ t:he In~ention
In wide use and possessing various deslgns in power
tran5111i~5iOn m~chinery i5 the univer~al joint (,sometimes termed :
knuc~l~ joint~ which.has been long exi~;tant. in the art of
S machine build.ing. In its earliest and simple6~ forms such i~
familiarly known as the ~ooke's ox Cardan kype, consisting
es.sentially of two forks connected throucJh an .i.ntermediate
: ~ block straddled by the forks and journalled thexeto at right
angles to each other. The Hoo~a'.s-Cardan type of jo.int has the
~: 10 ~ixtue of simplicity combined with capab.ility o~ accommodatin~
high torques at very large angles oE deflection be~ween input
and output shafts. However, the dynamic pexformance :is not
, ideal, since there is a variation between the ancJular motions
of the connectad shafts, which ha~ ullaGceptably severe inert~al .'
,load consequences at:hi~h speeds, large deflection angles, or.
with substantial masses connected by the cooperatincJ sha~t.s.
E~lem0ntary analysis shows that the angular vari.ation
:~ characteri~tic of the Hooke's or Cardan jo.int is deqcribed by
: the relat.ionship
: tan ~ = tan ~, sea,~ ,
where c~ and ~ are the angle~ of rotat.ion o the input and
: output ~hafts and ~ i~. the d~flection angle ~etween the
, shafts. ~ typical value of the ~ariation i~ seen at ~ = 36
. which i~ at the upper limit reached in most high speed
dpplications su~h a~ automatic ront wheel dr.ives, axles and
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transmi3sio1l sha~ts. At this deflection the output shaft will
alternately ~e a1lgularly advanced or retarded, each twice per
shaft revolution, to peak values of ~ 6. The corresponding
variation in velocity is + 24% referred to the input speed.
It is clear that such variations in relative position
would be awkward in preision control applications where it is
important to maintain accuracy of motion from one point in a
machine drive to another point via a univer~al joint used to
't ~ ~ change shaft direction. The substantîally large cyclical
variations in velocity create :intolerable vibr~tions and
acceleration loads where massive loads are being driven.
De3pite the undesirable characteristlc of non-
uniformity in the Hooke's or Cardan joint, lts simplicity has
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led to extensiv application whexe speeds and in~rtial loads
are low and demands for precision of position ~re minirnal. But
increasing areas of application are leveloping where the
consequences of motion irregularity are unacceptab]e for reasons
of noise, excessive vibration, imprecision of control, or
consequent wear. Numerous examples can be cited such as the
2~ automotive front wheel drives, helicop~er rotor drives, marine
inboard-outboard propellor ~hats, hydxaulic pump swash plate
drives, etc.
Beginning with the era of abruptly higher machine and
transportation speQds cixca the end of World War I, intensive
efforts were made to find substitute ~or the Hooke's~Cardan
type of jo ~ whlch would have uniform ve10cit~ performance.
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small numbex o~ successul joi.nts ha~e been ~ouncl, notably
the Rzeppa and Weiss rolling ball ~oints, and a variety o
sliding block types of which the Tracta appears to be the most
frequently used. In addition a nun~er of "kinematic'~ models
have appeared which are theoretical laboratory solutions o~
: the problem of providing con~tant velocity ancJular transmission.
However, regarding the latter, thair practical value is
questionable for reasons of complexity, non~compactness, or
low-load caxxying a~ility.
iO ~ : A fairly frequent solution is the use oE two Hooke' 5-
Cardan joints in series~ so phased that the irregularities o~
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one are cancelled out by oppositel~ ~irected variations o~ t:he
other. This is less than ideal becauc3e of:the added space
:required by the second joint, the irregular motion o~ the
15~ intermediate membex between the twol i.nability in many
applications to ensure that each joint operates at the same
angle as the other;as well as other obvious disadvantagec3 such
~;: as cost, noise vibration and wear to menti.on a ~ew.
The rolling ball joints inventcd almos~ fifty years
: 20 ~ ago have enjoyed the greatest succe~ and are used currently a~
first cho.ice where high performance is recfui.red. ~owever~ they
: have limitations in respect to high manufacturing C05t, limited
~durability, and are subject to derating at high ~peeds and lar~e
deflection angles. They are not susceptibl.e of adjustment to
~take up wear and there~ore cannot be assembled to a true zexo
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backla.sh condition (in view o~ rnanufacturing tolerances)
without preloading which detrac~s from load carrying ability
and economy of production~
It may therefore be ~airly said that much room for
improvement remains in the evolution of the universal joint
regarding economy, ~implicity, durability and other factors.
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~Summary of the Invention
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An object of the present invention is to provide a
~;~universal joint which eliminates the aforementioned disadvantages
-~ ~ 10~ and drawbacks of prior art joints and which is capable of being
mass produaed economically~ employing reasonable manu~acturing
tolerances to achieve precise uniform velocit.y operation.
AnOther object is to provide a con~tant velocity
universal joint o~ the foregoing type which i5 not only simple
~ut capable of being produced at relatively low cost compared
to existing constant velocity universal join1:s while being
~ competitivé with non-uniform velocity joints o~` the Hooke's-
'~ ~ Cardan types in general use.
` ~ 5till another object is to provide a universal joint
of the foregoing type which may be manufactured to existing
stanclards o gear tooth or cam proiles and by the use of
current gear or cam productiorl tools in manufacturing practices.
A further object is to provide a universal joint o~
the fo.egoing type which i~ cc~pable of bei.ng designed to accept
~5 torque loadings equal to or exceeding existing con~tant velocity
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universal ~oints with e~uivalerlt ~lze proportiorls and speed
rati~gs equal to or exceeding existing types as well as a joint
capable of operating at deflection anyles equal to ox greater
than existing types.
A still further object is to provide a constant
velocity univer~al joint which is adjustable as ~o operating
~clearance for the accommodation of generou~ manufacturing
tolerances to achieve minimum backlash in the original as~embly
and which is adjustable during the useul operating life to
take up wear and thereby minimize backlash throughout the term
of u~e, which fe.~ture is not avallable in existing types.
Brief Descri~tlon of the Drawin~s
; Figure 1 is an exploded i ometria view of the
constant velocity universal joint of this invention.
~ 15 ~ Figure 2 is an isometric view of the univer~al joint
; ~ in assembled form showing the drive and driven a~is a~ an angle
relative to one another~ ~
Fi~ure 3 is a top plan view of the asser~led
universal joint of Figure 2.
~ Figure 4 i9 a sectional view taken along line~ 4-4
of Figure 3~
Figuxe 5 is a sectional view taken along lines 5-5
o~ Figure 3.
Figure ~ is a sectional plan view taken Ol1 the common
c~ntral plane containing the two ~haft axe~ and parallel to the
plane of the drawing.
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Figure 7 is a view similar to E~igure 3 bu~ Wi ~h the
joint rotated 45~.
Fit~ure 8 is a view similar to Figure 3 but with the
joint rot~ted 9U.
Fi.gure 9 is a top plan view o~ the universal jOillt w:ith
the axes of the lrive and driven shafts aligned with the forks
orientated 45~ to the plane of the drawing;
Figure 10 i5 a side elevat.ional view of the universal
joint shown in Figure 9;
Figure 11 is an elevational view of a "xocking" pinion
segment;
Figure 12 is a cro~s-secti.onal view taken along
; lines 12-12 of Figure 11;
: : Figure~13 is a diagrammatic view o~ the universal joint ..
of this invention showing each o~ the drive and driven shafts
deflected /~ about axis Y~Y upwardly from the plane of the
drawings with the included angle being 180 ~ ~ with the fork3
~ oriented 45 to the plane of the clrawings;
:~ Figure 14 is a si.de elevational view of the unive.rsal
j~int of yure 13;
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Figure 15 is another diagrammatic view of the gear
segments shown in Figure 13 illustrating the rotation o the
pinion segments as a unit about axes X-X and the bevel gear
: segments as part of complete bevel gears shown in phantom
with their rotation being about axes Y-Y; and
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Figure 16 is a similar diagrammatic view showing the , ;~
movement of the pinion segments and the bevel gear segments
shown in detail in Figure 14 with axis X'-X' being perpendicular
to the rotational axes of the.drive and therefore the new position~
of the rotational axes of the bevel gears shown in phantom from
vh~ch th~ g r segmc~ts o FigU~ a-e derived.
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Detailed Description
In the drawings, the joint main input and output
members are two massive keyed drive hubs 20 and 22 to receive
the respective input and output shafts 24 and 26 which are to
be connected by the universal joint 28. These hubs have
extending therefrom forks 30, 32, 34 and 36 which transmit
torque in either direction of rotation to the associated bevel
gear seyments 30a and 30b, 32a and 32b, 34a and 34b, and 36a
and 36b, respectively. Whera desired or necessary, the hubs
and associated forkæ with segments may be integral with each
other, that is to say they may be machined out of one solid
blank and thereby eliminate the need for fasteners. In he
~drawings, these parts are shown as separate element~ in the
accompanying illustrations only to isolate their individual
functions.
~he bevel gear ~egments 30a, 30b, 32a, 32b, 34a, 34b,
36a and 36b may be considered to be segments of a complete
circular bevel gear so situated that the axis of rotation o~
the teeth of a given segment is a line through the center of
the joint, but at 45 with the plane of the fork on which the
segment i~ located. Each bevel gear segment may be regarded as
a "slice" of its parent gear cut out on a plane at 45 to the
gear axis. The origins o the bevel gear segments and their
essential orientation are illustrated in Figures 13 to 16.
Bevel split compo~nd pinion segment~ 38a and 38b, ~;
40a and 40b, 42a and 42~, and 44a and 44b are por~ions o
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standard involute bevel pinions as mating members with the
bevel gear segments of a standard bevel reduction gear set which
in the illustrated embodiment has a tooth ratio of 2:1. It shoul
be understood that this invention contemplates utilization of
ratios. The split pinion segments are as divided at the exact
center of the parent pinion (see Figure 15) and then the split
surfaces are formed into the toothed contours 38a', 38b', 40a',
40b', 42a', ~2b', 44a' and 44b', respectively. These contours
are derived from a cylindrical spur pinion form or blank which
is revised by a slight tooth modification accomplished by sup-
plementary gear cutting of the spur gear tooth spaces to greater
depth E and at slight angle ~ to the spur gear blank axis, as
a function of the angle ~C taken around the pitch circle from th
center element designated as 0-0. The extra depth cutting pro-
duces pitch contours which depart sllghtly from a circular form
and may be described generically as paraboloid. The paraboloid
pitch contours h~ve magnitudes proportional in siæe to the dis-
tance from the joint center in the same way as the bevel pinion
teeth are proportioned in size from the center. The extra depth
E and angle ~ are as tabulated in the following table with
reference to Figures 11 and 12, as modifications to a spur pinion
blank which has a pitch diameter of .625 ac large as the pitch diamete
of khe large end of the a sociated bevel pimon; and the angle~ i9 ~ en
inwardly in the same sense as the angle of the pitch cone o the
bevel pinion teeth. Ob~iously, the described modification~ lead-
ing to the spur gear configuration will be di~ferent ~or other
than the t th ratlo of the lllustratod embodiment.
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inches degrees degrees
0.0028 1~ 0.301
0.0100 24 0.970
0.0168 36 1.678
0.01~6 48 2.328
0.0133 60 2.953
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* Applicable directly toC~ or cin~ar pi~ =
lr/q6 for bevel gear and pinio~ teeth. For larger values
of ~ decrease E propo ~ onately to-the pi ~ ~.
and ~ are unchanged.
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~ ~ Tension rods 46 and 48 have threaded ends and in
association with screw studs 14 and nuts S0 serve to tie
together palrs of pinion segments which are diametrically
opposite (and which have been found to describe theoretically
identical motions in all aspects of the join~ action). ~he two
rods 46 and 48 are trimmed near the center of their length in
such a way as to permit them to cross at about 30 in the
common plane which contains their longitudinal axes. The
; ~ trimming also provides for them to turn relative to each other
about their axes so that the "rocking" motion of the compound
pinion segments 38a-42b is unimpeded. The radially outward
thrust forces which are an inevitable consequence of 0rce3 on the
bevelpinion teeth, are balanced by the tension rods. ~he tension -;
~rods 46 and 48 are associated with only two pairs of the ~;
compound split bevel pinion segments 38b and 42b, and 38a and 42a.
respectively. The other two bevel pinion pairs are accommodated
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by two rings 52 and 54, which fit together as shown in Figure 4
each having opposed radially extending threaded studs 52a and
52b and r4a and 54b and associated nuts 50. The tension between
diametrically opposite bevel pinion segrnents is assumed in each
ring through these screw studs and nuts. The relationships among
the interlocking tension members is shown more comprehensively
in Figure 5.
The stabilizing links 56 serve to keep aligned those
bevel pinion segments 40b and 44b attached to the smaller tension
ring 52. These two gear segments have no other inner abutmenk
against inward displacement except the taper of the teeth. It
has been found that in dynamic operation,tooth friction forces
may cause progressive inward creeping to cause ~amming in the
absence of a positive stop against such a tendency.
Obviously, other alternative means may be adopted for
positioning the bevel pinion segments in fixed relationship with
each other. Suitable key arrangements (not shown) may be em-
ployed and shims (not shown) may be deployed so that the bevel
pinion segements may be exactly positioned inwardly for optimum
gear operating clearance (backlash)to m~miæe tooth friction and p~n~
a~ate film thickness for lubricant. Such shims pr~vlde opportunity fox
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initial fitting of the assembly in view of manufacturing
tolerances. The shims also allow take-up ~or wear in the
course of extended operation to provide essentially ideal
fitting and ~inimized ~Ibacklash~ conditions throughout a long
S useful life.
Thus, a combination (cluster) of gears transmit
tsrque by means of tangential pre~sure contacts carried around
; as near as possible to the outer periphery of the clu ter where
the load carrying surfaces have the greatest mechanical ~
advantage; and to establish contours for the gear pitch surfaces
which will provide for their intimate contact in all relative
positlons of the input and output members in rotatlon and
deflection. The compound beveI split pinions 38a, 38b,42a and 4Zb ac
as~idlers be~æ~n the bevel gear ~ff~h=~s 30b,36b, and 3Zb,34b to transmit (by
15;~ ~pressure through theml the torque forces from one fork to the other.
~The compound pinion segments 38a, 38b, ~2a and 42b may rock on
each other in such a way that the space between opposing forks
is always exactly filled when said opposing forXs assume a
relative angular attitude to each other (as opposed to a
20 ~ mutually parallel relationship). As will be appreoiated, the
necessary condition ~or constant velocity transmission is simply
that the input and outpuk shafts must not turn with respect to
~the axis of de1ection (and thereor with each other) as shafts
are de1ected from a straight line through the joint center.
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An ex~minati,on of Fi.~ures 13-16 will serve to make
clear that for the condi.ti.on o shat rotation where the shaft
forks are each.ori.ented at 45 to the axis of deflection (,in ,
thls case the axis a~out whi.ch the deflection takes place is ...
~: the YY axis) a total jo;nt defIection angle o ~ as shown in
: : Figure 14 causes the ~evel gear segments to take an . '- .
~angular attitude to each other and forces the associated ~plit . .
plnion segments to rock, and roll down in the figure movlng into
that unlque position where they just fill the space between the ~ea
~:~ 10; ~ se~ts. The spur gear pitch surface~is so ~haped that reg~ess of '~
~thè~angle taken by the gear segments there is only one position~ ~ -
that~lt can take, since any other position would not allow enough
: :room. In other words, the associated:split pinion segments roll
into:a "hollow" and stay there. It should be noted that the same ~ . .
lS ~ ~behavior is exhiblted by the matching split pinion segments;. ' .
diametrically opposlte (beneath the plane of the drawing). ...
~' The corresponding behavior of those elements 90
,removed are shown in Figure 13. Here the fork-mounted bevel
gear segments~perform a pure rotatlon about the axls o~
20~ ~' ~deflection YY, in opposite directions to each other. The
associated split pinion segments act in this mode as though it
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were solid or unsplit and simply rotates on~its axis XX, through .,
, an angle equal to ~ ~2 multiplied by the gear-to-pinion ratio ....
: ~which in this case is two).
'25:It should be obvious that under the total de~lection .
, ~ ~ th otation of the split plnion segment~ as a unit
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(no rocking necessary since the opposing gears are parallel)
the input and output shafts do not rotate about their own axes
(they simply sw;ng about YY as a pivot~. Therefore under the
conditions described above for the relationships in Figures 13
and 14, the necessary condition for constant velocity is
satisfied. In can be shown that any other position of the joint
(deflection about XX creates the same conditions as for that
about YY) creates a condition which is a vector combination o~
Figures 13 and 14. For example if the shafts are deflected by
~ about a new 45 axis WWf hal-way between X and Y, the
same effect can be obtained by flrst deflecting .707 ~ about
YY, then .707 ~ about XX. The resulting final po~ition i8 the
~same and all the splLt pinion segments have performed two
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functions as shown individually in Figures 13 and 14 but
lS ~superimposed upon each in combination. The action or the
component of deflection about YY is identical to ~hat of Fig 14 AA
(except to a lesser degreej and the component of deflection
about XX is .707 of that shown in Figure 13. This latter is
pure rotation of the associated split pinion segment~. This
20~ supplementary rotation of the pinion is not quite ideal since
it has been rocked into a slightly non-circular form by the
; other component of the de1ection. However, it has been found
mathematicall~ that in the case of 36~tot~l deflection e~ the worst
consequence~is the introduction o one thousandth of the
joint's outside radius as looseness during the rotation. Such
degree of looseness is negligible as a practical consideration
in view of necessary operating clearances.
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Ona complete revolution of the universal joint 28 will
now be described with reference to the drawings. Assuming a rela-
tive disposition of the drive shaft 24 and driven shaft 26 turned
45 from their respective disposition illustrated in Fig. 3, the
pinion segments 38a and 38b will be engaged as shown in Fig. 14.
Pinion segments 42a and 42b on the opposite side ~idden) will be
rocked to the same degree but in the opposite direction. Pinion
segments 40a and 40b (and 44a and 44b on the opposite side and
hidden) will be disposed as shown in Fig. 13. Turning the shafts
90 in a clockwise direction from Figs. 13 and 14 will reverse the
disposition of pinion segments from that illustrated in Figs. 13
~and 14. Towards this end, pinion segments 38a and 38b together
with 42a and 42b will be disposed in the manner depicted by 40a,
40b and 44a and 44b in Fig. 13. Pinion segments 40a and 40b will
~ ~then be disposed as are 38a and 38b shown in Fig. 14 with pinion
segments 44a and 44b rocked in the same manner but in the opposite
~direction. Upon fur her~rotation of the shafts through another 90
or 225 from Fig. 3, pinion seqments 42a and 42b will be disposed
in the manner depicted by 38a and 38b in ~ig. 14 with pinion seg- ;~
ments 38a and 38b then becoming hidden, rocked to the same degree~
but in the opposite direction. Pinion segments 40a and 40b and
44a and 44b will again be disposed substantially as they are shown
in Fig. 13, but with their locations reversed. When the shafts ar
turned another 90 (or 315 from Fig. 3)pinion segments 44a and
44b will be disposed substantially as 38a and 38b are shown in
; Fig. 14 with pinion segments 40a and 40b then hidden below rocked
to the same extent but in the opposite direction. At 315~ rom
Fig. 3 pinion segments 38a and 38b and 42a and 42b will be dispose
substantially as 40a and 40b are shown in Fig. 13. At 45 inter-
vals from the positions shown in Fig.s 13 and 14 the pinion seg-
ments will assume positions that would be a combination o~ rocking
and turning substantially as shown in Pig 2, 3 and 8~
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In the preferred embodiment of the invention shown in
the drawings, the rolling contact surfaces are toothed. However,
in a somewhat less preferred embodiment, the same result can be
obtained by substituting curved cam contact surfaces which dup-
licate the pitch surface contours of the illustrated gears and
pinions supplemented by "crossed-belt" metal strips interposed
between the contacting rolling surfaces and so fastened as to
prevent slippage.
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. Thus the several aforenoted objects and advantages
: are most effectively attained. Although several somewhat preferre
embodiments have been disclosed and described in detail herein,
: ~ : it -~hould be understood that this invention is in no sense limited
: thereby and its scope is to be determined by tha~ of the appended
~c.
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